PROCEEDINGS OF

THE XIth INTE RNA TIONAL CONCRESS OF REFRIGER ATION

COMPTES REN DUS

DU XIeme CONGRES INTERNA TIONAL DU FROID

MUNICH 1963

PROGRESS IN REFRIGERATION SCIENCE AND TECHNOLOGY

PROGRES DANS LA SCIENCE ET LA TECHNIQUE DU FROID

VOLUME I

PUBLISHED FOR THE

IN TERNATIONAL INSTITUTE OF REFRIGERATION

IN STITUT INTERNATI ONAL DU FROID

BY THE

PERGAMON PRESS OXFORD· LONDON· NEW YORK. PARIS

AND

VERLAG C.F.MOLLER KARLSRUHE

Pergamon Press Ltd., Headington Hill Hall, Oxford

4 & 5 Fitzroy Square, London W.1

Pergamon Press (Scotland) Ltd., 2 & 3 Teviot Place, Edinburgh 1 Pergamon Press Inc., 122 East 55th. St., New York 22, N.Y.

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Verlag C. F.Miiller, Rheinstrasse 122, Karlsruhe-West

Copyright

© 1965

PERGAMON PRESS LTD.

Fii st published 1965

Library of Congress Card No. 60-16886

Printed in Western Germany by

C. F. Muller, Buchdrudl:erei und Verlag G.m.b.H., Karlsruhe

Foreword The Eleventh International Congress of Refrigeration took place in Munich in August­

September 1963. It was held under the auspices of the International Institute of Refrige­ration, and was organized on behalf of the Government by the competent authorities of the German Federal Republic, in first place by the German Association of Refrigeration. The papers presented for discussion at this Congress by the outstanding experts of ·the world cover the whole field of production and application of low temperatures.

These Proceedings of the Congress contain the full text of all papers and the discussions which took place after the presentation of the papers.

A great number of papers was presented to the Congress in Munich. They have been devided into 3 volumes farming the Proceedings of the Eleventh International Congress of Refrigeration. They reflect our present knowledge and indicate the future developments of refrigeration in science, engineering, medicine, and food technology. Problems of education are also dealt with.

The International Institute of Refrigeration highly appreciates the excellent cooperation of the German Papers Committee under the leadership of Professor Dr. J. Kuprianoff, and the work performed by the Presidents, Vice-Presidents, and Secretaries of the Technical Board, revising and publishing these Proceedings.

RUDOLF PLANK

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Preface The papers and discussion contributions at the Congress gave a broad review of the various

scientific and technical aspects, designs, and technology of refrigeration. They have been assembled here. In doing this, the main effort was concentrated on the technical and scientific aspects of the Congress, and only summarized material from the opening and closing sessions, technical visits, etc. were presented here. Anyone interested in more details of this part of the Congress should refer to the Congress programs.

The working numbers originally assigned to the papers are given on the upper side of the pages.

The greatest part of the editorial work for the Congress papers has, in effect, been carried out by the Presidents of the various Commissions of the I. I. R. in cooperation with the Vice­Presidents and Secretaries of the Commissions. They were assisted in their work by the German Papers' Committee.

All contributions from discussions were based only on notes written by the contributors themselves.

Since the material of each of the three volumes originated from a great number of dijf erent authors and contributors, many of whom were not expressing themselves in their mother language, a certain amount of editorial changes have been made by the bureau of the German Papers' Committee. It is hoped that this has been done without undue interference with the work of the individual author. If this has not been the case, the German Papars' Committee offers its apologies.

It should be mentioned that the authors of papers and the contributors to the discussions -delivered in written form - have the full responsibility not only for the content of their papers and the discussion contributions but also for their form of presentation. To all authors an opportunity has been given twice to make any necessary corrections in the course of printing of the preprints and the Proceedings. The German Papers' Committee and its relatively small but active bureau finally took over the responsibility only for the transformation of all corrections into the final version of the text and besides has also done some minor coordinating work in the general presentation of the Proceedings.

Special thanks are due to the office of the I. I. R. for help in revising the French part of the Proceedings, etc. Thanks are due further to all authors and contributors who have cooperated very kindly with the German Papers' Committee in the final preparation of these volumes.

J. KUPRIANOFF President of the

German Papers' Committee

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Volume I Officers of Commissions 1, 2, .md 3 of the International Institute of Refrigeration Presidents et Secretaires des Commissions 1, 2, et 3 de l'Institut International du Froid

COMMISSION 1

President :

Prof. A. van Itterbeek, Universite de Louvain, Louvain, Belgique.

Vice-Presidents :

Prof. F. G. Brickwedde, Dean, College of Chemistry and Physics, Pennsylvania State University, University Park, Pa., U. S.A.

Prof. Kanda, Research Institute for Iron, Steel and Other Metals, Tohoku-Univer­sity, Sendai, Japan.

Prof. Dr. V. P. Peshkov, Directeur Adjoint de l'Institut des Problemes Physiques de S. Vavilov de l' Academie des Sciences de l'U.R. S. S., Moscou V-133, Vorobievskoie, chausse 2, U. S. S.R.

Secretary :

Dr.J. Wilks, The Clarendon Laboratory, Parks Road, Oxford, U. K.

COMMISSION 2

President:

Prof. C. F. Kayan, Mechanical Engineer­ing Department, Columbia University, New-York 27, N.Y., U. S.A.

Vice-Presidents:

Dr.-Ing. V. Ibl, Directeur VSCHP, Ostrovskeha 34, Prague XVI, Czechoslo­vakia.

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Prof. Dr.-Ing. C. Codegone, Politecnico, Istituto di Fisica Tecnica, Turin, Italy.

Mr. G. Yate Pitts, Jaquiss and Sons, Re­gal Works, Gorton Road, Manchester 12, U. K.

Secretaries :

M. M. Duminil, 87, rue Doudeauville, Paris (18 e), France.

Miss Griffith, Electrical Research Assn. Laboratory, Cleeve Rd., Leatherhead, Surrey, U. K.

COMMISSION 3

President :

Prof. G. Lorentzen, Norges Tekniske Hogskole, Institut for Kjoleteknikk, Trondheim, Norway.

Vice-Presidents :

Prof. Dr. V. Martynovsky, Direction de l'Enseignement, UNESCO, Paris, France.

M. A. Neuenschwander, Directeur des Etablissements Brissonneau-York, 8, rue Bellini, Paris (16<), France.

Prof. Dr. L. Vahl, Laboratorium voor Koeltechniek en Droogtechniek, Tech­nische Hogeschool, Delft, Netherlands.

Secretary :

Mr. H. W. Fischer, York Shipley Ltd., North Circular Road, London, N. W. 2, U.K.

Fig. r. "Deutsches Museum" in Munich. Right hand the Great Hall, where the Opening Ceremony and the Banquet took place.

Introduction OPENING OF CONGRESS

The XIth International Congress of Refrigeration was held in Munich between August 27 and September 4, 1963, under the high patronage of Dr. Heinrich Lubke, President of the Federal Republic of Germany. The opening ceremony took place on August 28, in the Congress Hall of the "Deutsches Museum", where Professor Nessel­mann as General Secretary of the Congress welcomed the participants coming from 40 countries. Professor Dr. -Ing. Balke, former Federal Minister of Atomic Energy, opened the Congress as representative of the Government of the Federal Republic of Germany, stressing that the German Government appreciated it very much to have this Congress in Germany. He transmitted greetings from the President of the Federal Republic of Germany, Dr. Heinrich Lubke, and the Honorary President of the Congress, the Mi­nister of Commerce of the Federal Republic of Germany, Professor Dr. Dr. he. Ludwig Erhard; Professor Erhard regretted not being able to open the Congress personally.

Representing the Prime Minister of the State of Bavaria A. Goppel, and the Bavarian Government, Dr. Pohner, the Secretary of State, delivered a short speech, expressing his joy that the first International Congress of Refrigeration in Germany should take place in Bavaria. Acting for the Lord Mayor of the City of Munich Dr. H. J. Vogel, who was prevented from being present because of illness, the First Mayor A. Bayerle welcomed the delegates to the International Congress of Refrigeration wishing the participants a pleasant stay in the capital of Bavaria. Further greetings to the assembly were given by Professor R. Plank as President of the General Conference of the I. I. R.

In his festival address Mr. J. Foulon, President of the Executive Committee of the I. I . R. said :

«Nous sommes reunis aujourd'hui, dans le cadre du lleme Congres International du Froid dans la ville de Munich, ville au charme mysterieux et moyenageux chante par les poetes, ville d'art, ville universitaire riche d'instituts de recherche scientifique et de tresors artistiques.

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Fig. 2. Professor Balke, former Federal Minister of Atomic Energy and representative of the German Federal Government at his opening speach.

Nous avons trouve aupres des autorites officielles allemandes l'interet que meritent nos efforts et les marques de sympathie qui rendent les contacts humains aimables et un sejour agreable. Que le Gouvernement de la Republique federale, l'Etat de la Baviere et la Ville de Munich en soient remercies. L'Institut International du Froid leur exprime sa reconnaissance.

Munich, siege de ce Congres International du Froid, c'est aussi un symbole. C'est ici que vecut et mourut, il y a 30 ans a peine, Carl von Linde un des precurseurs le plus prestigieux des disciplines frigorifiques que nous desirons honorer au cours de ce Con­gres. Carl von Linde fut professeur pendant de longues annes a l'Universite technique de cette ville. C'est au cours de son enseignement qu'il publia en 1870, son memoire sur "L'extraction de la chaleur aux basses temperatures a l'aide de moyens mecaniques." S'il prit, peu apres, la direction de la Gesellschaft fiir Linde's Eismaschinen a Wiesbaden, c'est a Munich qu'il crea une section de cette societe pour la production industrielle des gaz liquefies et qu'il mit au point la fabrication de l'air liquide et de la separation des gaz. Carl von Linde exer�a une influence profonde sur le developpement de l'industrie frigorifique en associant a sa formation scientifique, un esprit realiste et une competence technique remarquable.

Reunis a Munich aujourd'hui, nous devions une pensee a la memoire de ce grand ingenieur, de ce pionnier, a celui que l'on a si justement appele le Pere de la technique allemande du Froid.

Au debut du mois d'octobre 1908, s'ouvrait a Paris dans un grand enthousiasme le I er Congres International du Froid. Deux hommes clairvoyants, Andre Lebon et de Loverdo avaient estime le moment venu de creer un mouvement d'opinion pour la diffusion des techniques frigorifiques et d'envisager la creation d'un organisme inter­national de contact et d'etudes.

A la seance d'ouverture, le savant hollandais Kamerlingh Onnes s'associa a cette initiative et fixa en des termes qui ont ete souvent rappeles l'objet de cet organisme. II faut disait-il rassembler routes les intelligences qui s'interessent aux basses tempe­ratures.

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L' Association Internationale du Froid etait creee et elle pouvait compter, des sa cons­titution, sur le concours du laboratoire cryogenique de Leyde dont la renommee scienti­fique se propageait deja dans le monde et qui au cours de ces cinquante dernieres annees par la voie autorisee des savants qui se son succedes a sa direction: les Professeurs Kamer­lingh Ormes, Keesom, De Haas et Gorter, n'ont jamais manque d'apporter leur col­laboration active et amicale a !'effort scientifique international de l'Institut.

Les voies etaient tracees dans les dispositions d'une Convention Internationale entre Etats membres. Elles sont encore dans les grandes lignes l'objet de l'Institut - favoriser le developpement des recherches et promouvoir les etudes scientifiques techniques et economiques dans les differents secteurs du Froid.

En un demi-siecle d'activite, !'influence de l'Institut International du Froid fut profonde.

Ce sont tout d'abord les perspectives infinies des communications humaines entre specialistes de taus les pays, de toutes les opinions, de toutes les philosophies rassembles sous le signe de preoccupations scientifiques communes au sein des organes-directeurs, des commissions et des Congres.

La pensee scientifique, quelque soit son origine, en visant au meme but intellectuel n'est-elle pas internationale par sa nature meme ?

C'est egalement une documentation considerable qui constitue une source de renseig­nements precieux et varies : Etudes presentees au cours des Congres et des reunions des Commissions; Recommandations, regles, normes resultant de confrontations scienti­fiques au sein des groupes de travail; rapports rediges a la demande d'organismes inter­nationaux; statistiques, guide bibliographique, dictionnaire multilingue et enfin le Bulle­tin, trait d'union entre taus les frigoristes, qui a paru regulierement malgre toutes les vicissitudes politiques depuis 1910 et qui forme un ensemble impressionnant de donnees scientifiques, techniques et economiques sur le Froid.

Fig. 3. The stage of the Great Hall in the "Deutsches Museum" during the Opening Ceremony.

Nous soulignerons d'autre part l'interet d'initiatives qui elargissent le champ d'action de l'Institut. Ce fut en 1922, la creation d'une fondation aupres du laboratoire cryoge­nique de Leyde qui familiarisa, a l'epoque, de nombreux chercheurs de taus les pays

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aux problemes de la science des tres basses temperatures et parmi les realisations recentes, nous sommes heureux de signaler !'organisation a Lyon de cours de lyophilisation en­seignes par des specialistes eminents et qui sont suivis par un nombre toujours croissant d'auditeurs. Enfin c'est le Fonds special pour !'encouragement de la Science et de la Technique du Froid, Fonds cree en 1962 qui va favoriser les etudes originales de jeunes chercheurs et associer des pays en voie de developpement a la pensee scientifique de l'Institut.

Tel est, dans ses grandes lignes, le bilan d'un demi siecle d'activite de l'Institut Inter­national du Froid. C'est un bilan positif qui s'amplifie chaque annee et qui cree, en ce llemeCongresinternational un sentiment d'optimisme pour les destinees de cet Institut.

Ces destinees, comment peut-on les concevoir? Notre societe moderne s'embarrasse peu de vues historiques, se refere peu au passe mais a l'avenir, cet avenir qu'elle contribue a construire et qui lui parait deja superieur au present.

Comment va s'integrer l'Institut dans les perspectives de ce monde en evolution rapide, de ce monde qui eclate dans ses dimensions.

II est vrai que la facilite des echanges est favorable au rapprochement des hommes et tend a creer un sentiment profond de concorde et de comprehension qui ne manquera pas de s'epanouir et de s'affirmer au sein d'un organisme international d'etudes.

Cette comprehension, nous en sommes persuades, brisera les separations chancelantes de certains compartiments rigides des disciplines traditionnelles. La science et la techni­que, dira Louis Armand, marque le progres de nos jours, mais il n'y a plus de barrieres entre la vile technique et la noble science et une grande symbiose des connaissances est une image conforme a la grande convergence de notre epoque qui doit remplacer celle qui se font trop d'esprits en suivant les descriptions d'autrefois.

L'avenir, tend egalement a devenir sensible a tout ce qui est humain et dans le secteur des sciences humaines fait appel aux sciences economiques pour rechercher les optiques necessaires a la lutte contre les deficiences et apporter des solutions a !'inquietude des agriculteurs et a cette lutte emouvante contre la faim qui s'inscrit dans les grands des­seins de notre temps.

Dans ce domaine, le Froid qui intervient dans la production, la repartition et la con­sommation des denrees alimentaires est un element essentiel. C'est un facteur ideal de la regulation des courants economiques et il contribue a la prise de conscience technique d'un secteur capital de l'economie des pays sous-developpes.

La seule maniere de traiter les victimes de la malnutrition des etres humains declarait dernierement le Docteur Sen, Directeur General du F. A. 0., n'est pas de leur jeter

quelques vivres au fond du gouffre de la faim mais de leur lancer une corde pour les aider a en sortir.

L'Institut International du Froid est-il, dans ces concepts, a la mesure de ces grands courants qui entrainent l'humanite. Doit-il les reviser? Nous ne le croyons pas. Ses structures paraissent suffisamment fiexibles pour s'adapter aux necessites mouvantes et imperieuses du progres et les organes directeurs ont toujours eu le souci d'eviter de s'enliser dans un juridisme trop absolu et negatif.

D'ailleurs, la route semble deja ouverte vers les voies nouvelles. L'action de la Com­mission 6 s'est elargie en s'interessant aux problemes de la biologie et de la medecine. Grace a une initiative particulierement heureuse, la Commission 1 verra prochainement la technique intervenir dans les etudes des tres basses temperatures. Dans la Commission des Entrepots frigorifiques, des etudes economiques tendent a fournir des donnees essentielles pour la vie d'une industrie s'integrant dans les activites agricoles.

Le Fonds special pour !'Encouragement de la Science et de la Technique du Froid de creation recente va diriger ses efforts pour promouvoir les techniques frigorifiques dans les pays en voie de developpement, et il parait possible qu'en application des dis­positions de la Convention internationale de l'Institut, des moyens nouveaux soient etudies pour elargir la diffusion des travaux de l'Institut.

Ce mouvement qui entraine !'evolution de l'Institut International du Froid postule des ressources financieres suffisantes. Les ressources actuelles, il faut bien le recon­naitre sont modestes et elles ne sont pas adaptees a l'ampleur d'un programme base sur les imperatifs de l'avenir previsible. C'est pourquoi, au seuil de ce lleme Congres, nous

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aimerions emettre le voeu de voir cous les Beats membres et les collectivites prendre conscience des possibilites de l'Institut International du Froid dans !'evolution de son a:uvre scientifique et humaine et d'apporter a cette oeuvre le concours financier neces­saire.

Les bonnes volontes et !es competences ne manquent pas au sein de cet Institut et le devouement est inscrit dans ses traditions.

Les realisations passees et !es efforts presents sont garants de l'avenir et vous per­mettrez au President du Cornice Executif done le mandat et une longue collaboration se terminent au cours de ce Congres d'apporter a l'Institut International du Froid son souhait le plus fervent d'un avenir fecond en rappelant cette parole de Pascal: "Le passe et le present sont nos moyens. Le seul avenir est notre fin".»

Fig. 4. At the stage of the Great Hall during the Opening Ceremony. From left : Mr. Thevenot, Prof. Glansdorff, Prof. Plank, Secretary of State Dr. Bohler, Prof. Nesselmann, Mayor of Munich Bayerle, President Foulon, Dr. Fidler, Prof. S. A. Andersen.

Dr. J. C. Fidler, President of the Technical Board of the I. I. R., discussed problems of technical assistance, especially for the underdeveloped countries as follows:

"Article 1 of the "Agreement concerning the International Institute of Refrigeration" states that the aim of the member countries is to develop "the uses of refrigeration which improve the living conditions of mankind".

There is no limitation implicit in this statement; "mankind" is not defined as those countries which have signed the Agreement. The greater part of mankind lies outside our organisation, in the densely populated, desperately poor, under-developed countries.

Many agencies are already operating, rendering technical and financial assistance to these countries. Ones view of the adequacy of such assistance depends on whether one

is a donor or a recipient, but it must be said at once that the criticisms of the scale of assistance, which often occur in the African and Asian press, are not entirely fair, and such remarks as those recently attributed to the President of the Pakistan Association for the Advancement of Science are hardly calculated to encourage generosity.

It is not possible to give an estimate of the need for assistance, nor of the total already given, but as an example, my country currently supplies the equivalent of 1.5 % of our gross national product, from state and private sources. Further, we have 18,000 experts serving abroad, and 6% of all advanced educational places are occupied by students from the under-developed countries. Other industrialised states contribute likewise, in proportion to their incomes, and availability of people with suitable skills.

Much of this assistance is rendered through the Special Agencies of the United Nations (such as the Food and Agriculture Organisation, UNESCO, World Health Organisation), by loans from the World Bank, and by direct grants-in-aid, or loans, from national governments. But the total aid falls lamentably short of requirements, for two reasons.

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Firstly, there is not enough money at present available. Too much is spent on im­proving living standards of those who already enjoy prosperity, and too much on exotic and esoteric fireworks, cosmic feux d'artifice. Rocketry, inter-planetary exploration, these are exciting projects; they may or may not be necessary, but a sense of proportion is being lost. Do we seek the other planets to see if t h e y are fit to live on? How much aid to "mankind" could be given for the cost of one abortive moon-shot, or one aban­

doned space project? This is not an indictment of two nations only; the little countries also have their little expensive toys. What is needed is some slowing-down of the pace, and diversion of some of the effort to those things which matter on this planet.

The second reason for failure to achieve our targets is the shortage of skilled man­power, and the waste of those we have available. There is too little appreciation of the fact that the under-developed countries are under-developed administratively as well as technically; projects take longer to come to fruition, and expert assistance is available for too short periods. Experts are recruited on short-term contracts because of the low availability of skilled man-power, even in industrialised countries, and men with the requisite skills have to be borrowed from governmental departments or private employ­ment.

The need is for a greater output of scientists and technologists from our own colleges, coupled with an assured career for many of them, in the under-developed countries. Our own Commission 9 is very important; we must do all we can to encourage its work, and to persuade young men to take up a career in refrigeration.

U Thant, the Secretary-General of the United Nations, recently called for the establishment of a Corps of Engineers, career engineers who could be seconded for long periods to engineering projects in the under-developed countries. I hope this plan will mature; if it does, we can rely on our energetic Director to do all he can to ensure that refrigeration engineers play their part. He will need all the help we can get from governments and industry alike.

Our Institute has lagged behind in this matter. Since the war, all our meetings have been in Europe or North America, because our funds would not allow us to travel further afield. A mere handful of people from Asia and Africa have taken part. This has been a source of concern to the Technical Board, and the Executive Committee, but with the creation of the Special Fund, we are at last trying to extend our help to

more distant territories. You will know of the meeting we plan to hold in East Bengal next spring; we expect this to be only the first to take place in an under-developed country. Its success will depend, as does that of all technical assistance programmes, on the co-operation of the countries to which help is offered, as well as on our own efforts.

I make no apology for having spoken to you on a subject which only marginally touches on refrigeration. We must not take too narrow a view of our responsibilities. In this day and age, faced by such an inequality of benefit from modern technological advances, we would do well to remember the words of John Donne, who died in 1950, that 'no man is an island, entire of himself; everyman is a piece of the continent, a part of the main' ".

During the Xth International Congress of Refrigeration in Copenhagen the Danish Association of Refrigeration had instituted the Ottesen-gold medal to commemorate the work of A. J. A. Ottesen on the first practical quick freezing method. The Ottesen­gold medal, which is awarded every four years, was now presented to Professor R. Plank, Karlsruhe, for his outstanding scientific and practical work in the field of refrigeration. Professor S. A. Andersen, Copenhagen, handed over this award of honour and thanked Professor Plank in German also for the great help he had given him personally in his scientific work: ,,Ich habe sehr viel von Ihnen gelemt". Professor Plank expressed his thanks for this high award and reminded of his mutual work with A. J. A. Ottesen.

The prizes of the I. I. R. were instituted according to a decision of the Executive Committee on November 21, 1961, creating the "Special Fund of the International Institute of Refrigeration for the Encouragement of Refrigeration Science and Tech­nique"; these prizes are given in recognition of papers of outstanding and original quality.

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Fig. 5. Prof. Andersen presents to Prof. Plank the golden Ottesen-Medal.

These ptizes were first awarded on the occasion of the Xlth International Congress at Munich. At the opening session of the Congress Dr. J. C. Fidler, President of the Technical Board of the Institute announced the following awards:

- a prize of 2000 F to Mr. A. Gae (France) for his paper on "Determination of the cooling of foodstuffs by refrigeration";

- a prize of merit of 1000 F to Mr. J. Fabian (Czechoslovakia) for a thesis on "Some fundamental factors affecting the survival of microorganisms during and after freeze-drying";

- a. prize of merit of 1000 F to Messrs. R. M. Love and M. K. Elerian (United King­dom) for their study "The irreversible loosening of bound water at very low tempe­ratures in cod muscle";

a prize of merit of 1000 F to Mr. A. W. Paliwoda (Poland) for his study "The design problems of supersonic ejectors operating as booster compressors in refri­gerating systems".

The Medal of the International Institute of Refrigeration was simultaneously awarded to all the authors.

As the first speaker of the opening ceremony Professor Plank gave the following lecture entitled "Richard Mollier" :

"One hundred years ago, on November 30, 1863, RICHARD MOLLIER was born. He was the eldest son of the director of the Engineering Works and Shipyard in Triest, Eduard Mollier. Richards parents, who came from Germany, had moved there. At that time Triest belonged to the Austrian-Hungarian monarchy. In 1882 Richard Mollier passed the final examination at the German High School in Triest with distinction. He first devoted himself to the study of mathematics and physics at the university in Graz, Austria, continued his studies however in Munich in the field of mechanical engineering, where he was particularly captivated by the lectures on thermodynamics of the out­standing educator MORITZ SCHROTER. In Munich in 1888 he passed the final examination. After two years in the Engineering Works and Shipyard in Triest, Mollier returned to Munich as Moritz Schroter's assistant. Here he also met Carl von Linde and experienced the first liquefaction of air on an industrial scale. It was Linde who aroused his interest in refrigerating engineering. Mollier's academic career was somewhat unusual in that in 1892 he first acquired the right to hold lectures at the university with a paper on the heat diagram and did not earn the degree of Doctor of Philosophy (PhD) at the university in Munich until three years later, 1895. The degree of Doctor of Engineering (Dr. Eng.) did not exist at that time.

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Already in his first papers Mollier was concerned with the thermodynamics of refri­gerating machines. The behavior of carbon dioxide in the region of the critical point was puzzeling and unexplained. Mollier, in his paper "On the thermal properties of carbon dioxide and other technically important vapors" 1 calculated the first reliable steam tables for carbon dioxide on the basis of the best experimental investigations of that time. He showed that in the region of the critical point the simple Zeuner-equation of state was no longer valid and used therefore the general Clausius-equation. He suc­ceeded in evaluating the entire thermal behavior in the region of saturation. One year later, based on the measurements of the French physicist Emile-Hilaire Amagat, he extended his investigations to the superheated region and prepared the first usable temperature-entropy diagram for carbon dioxide 2•

The great recognition which was accorded these investigations is expressed by the fact that Gustav Zeuner, in the second edition of his classical work "Technical Thermo­dynamics" incorporated Mollier's results in their entirety3•

In 1897 the great Scottish engineer and scholar Sir JamesAlfred Ewing, at that time professor at Cambridge University, held the Howard Lectures on mechanical refri­geration at the Society of Arts. Several years later these lectures, with numerous additions, appeared in book-form. The main stress lay on the thermodynamic consideration of the refrigeration process. In this book, which was translated into German in 1910 by R. C. A. Banfield4, the author emphasizes that "thanks to Mollier's investigations, it has become possible to accurately determine without difficulty the theoretical efficiency of a carbon dioxide machine for given temperature conditions and for any arbitrary moisture fraction at the beginning of compression" (page 67).

In his first paper1, Mollier had already calculated a steam table for ammonia, which however was not yet very accurate because the available experimental data exhibited great contradictions.

These papers were so highly valued in the technical world that already in 1896 Mollier was appointed associate professor of applied physics and engineering in the philosophy department of Gottingen University. But shortly thereafter the most renowned expert in the field of applied thermodynamics in Germany, Gustav Zeuner, died in Dresden. In search of the best successor, the choice fell on Richard Motlier and he then held the chair of thermodynamics from 1897 until 1933; he not only retained its first place among the German Technical Universities but also spread its fame far beyond the German borders.

At the end of the last century there was complete chaos in the field of heat transfer. No consistency existed among the numerous experimental results of individual scientists, the clarification of the contradictions appeared almost hopeless. The Association of German Engineers (VDI) decided therefore in 1895 to bring about a clarification of the state of knowledge of the laws of heat transfer and assigned this task to Mollier. In 1897 Mollier presented the results of his critical investigation in his paper "On heat transfer and the applicable experimental results'"· Upon elimination of less reliable measurements, he succeeded in giving simple and sufficiently accurate formulas for the various technically important problems of heat transfer by contact as well as by radiation. The results could also be used for the calculation and design of refrigerating equipment. However he was aware that this was only the first step in the direction of order and continued to keep the task in mind. Only a few years later he appointed the very capable young scientist Willielm Nusselt to his institute and in 1909 gave him the opportunity to acquire the right to hold lectures at the university of Dresden. As lecturer (Privatdozent) and assistant in Mollier's laboratory, Nusselt created new ways by the application of dimensional analysis to the processes of heat transfer and diffusion. This method permitted the collation of individual observations under a common point of view.

r Zeitschr. fiir die ges. Kalte·Ind., Ed. 2 (1895), S. 66 und 85 2 Zeitschr. fiir die ges. Kalte-Ind., Ed. 3 (1896), S . 65 und 90 3 G. Zeuner, Technische Thermodynamik, Verlag Arthur Felix, Leipzig, 1901, Ed. II, S. 250-259 4 ]. A. Ewing, Die mechanische Kalteerzeugung, Eraunschweig, 1910, Fr. Vieweg und Sohn 5 Zeitschr. d. VDI, Bd. 41 (1897), S. 153 und 197

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Since about 1900 Mollier's ideas were directed to another task. He had recognized the importance of a thermodynamic property which he called the "heat content at constant pressure" and denoted with the letter "i". Later this property received the designation "enthalpy". It is the sum of internal energy and the product of pressure p and volume v (i = u + pv). This property, which is now usually denoted by h, was introduced into theoretical thermodynamics by the great American scientist Josiah Willard Gibbs in the years 1873 to 1878. From 1871 until 1903 Gibbs was professor of mathematical physics at Yale University. He was considered there an oddity and his work was not understood at that time. In Germany it was first made available by the translation of his papers by W. Ostwald6 and had an exceedingly stimulating effect on the further development and extention of thermodynamics.

Mollier is due the credit for introducing the enthalpy into applied thermodynamics

and recognizing its practical value. It now has the same importance in engineering that the internal energy has in physical thermodynamics. That can be explained by the fact that in engineering in addition to the temperature T, one considers the pressure p and not the volume v as second independent variable and writes the equation of state in the form v = f (p, T).

For the first time in 1902 Mollier presented the laws and formulas of thermodynamics uniformly by application of the enthalpy in the chapter on "Heat" in the 18th edition of the engineering handbook "Hiitte". He compiled this chapter in an exemplary manner until the 26th edition (1931), for the last time together with his fellow worker Friedrich Merkel.

In 1904 the basic work "New diagrams for applied thermodynamics", which made Mollier's name famous in the whole technical world, appeared on four pages7• _He de­veloped thermodynamic diagrams, in which the enthalpy, in addition to the entropy or to the pressure, was used as coordinate, and emphasized that these diagrams are also suitable for the presentation of the processes in refrigerating machines, since all deter­mining variables can be expressed by the enthalpy in the four main points of the cycle. For steam engines he gives the enthalpy-entropy-diagram preference, but believed that "for refrigerating machines the enthalpy-pressure-diagram can very well come into consideration". This view has been completely confirmed. Both diagrams have found grateful acceptance in all countries. At the suggestion of the American Bureau of Stan­dards these diagrams are generally denoted nowadays as Mollier charts. Whereas in the well known T, s-diagrams the quantities of heat appear as areas, which can only be evaluated by planimetry, one can measure them simply as a length in the Mollier chart. With these diagrams the calculation of thermal machines was extraordinarily facilitated. A. Stodola recognized their significance at once and added them to the second and third editions of his famous work on steam turbines.

For this and Mollier's following papers, the clarity and conciseness of the presen­tation is characteristic. A great abundance of ideas is presented convincingly and with­out embellishment on a few pages. Not the number of words, but the precision of the language is the measure for the value of a scientific work.

The first enthalpy-entropy-diagram for steam was based on the Regnault-and Zeuner­steam tables, which were in general use at that time. Mollier recognized that the values given in these tables had been surpassed. He chose as starting point the equation of state recommended by Callendar on the basis of throttling experiments8• The thermal properties calculated on this basis up to a pressure of 20 atmospheres are given in Mol­lier's "New Tables and Charts for Steam"9 which experienced a large number of unal­tered printings and found wide circulation. As the pressure range of industrial steam engines continually increased and new measurements of the properties of steam became known, Mollier, in 1925 in a second edition of his work, expanded the equation of state and adapted these measurements so that the steam tables could be enlarged up to the

6 ]. W. Gibbs, Collected Papers translated by W. Ostwald, Thermodynamische Studien, Leipzig, b 1892 7 Zeitschr. d. VDI, Ed. 48 (1904), S. 271 8 H. L. Callendar, Proc. Roy. Soc. Ed. 67 (1900), S. 266 9 Berlin, 1906, Verlag v. Julius Springer

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critical point, even though the values in its immediate vicinity were affected by greater uncertainties. In this range however sufficient agreement among the values of different scientists could still not be achieved today.

Another of Mollier's outstanding works was the elaboration of a chart in which the states of moist air were presented. In this chart he also made use of the enthalpy as one of the coordinates whereas for the second coordinate the quantity of moisture " contained in 1 kg of dry air was chosen. In 1923 his treatise "Anew diagram for moist air"10 appear­ed, which greatly simplified the calculations for air coolers, humidifiers, cooling towers and dryers and found wide circulation. Six years later he had extended this diagram to include variable total pressures and had also added the solid and fog zones11•

Besides this chart from Mollier, the t,x-psychrometric chart suggested by W. H. Carrier12 has found wide circulation in America, especially for air conditioning. This is not the place to consider which of these two diagrams combines the greatest advan­tages. W. Haussler13 has conducted a careful evaluation and a comparison with several other suggestions; this work is to be highly recommended. In Europe the Mollier chart is preferably used.

All of Mollier's publications are characterized by punctilious scientific accuracy and incomparable clarity. He always succeeded in expressing himself concisely and avoiding repetitions. That is the reason why his scientific legacy is not great in volume but all the more significant in content.

His accomplishments would however not be fully appreciated, if one were to measure them by his publications only. Mollier exerted the greatest influence on his students in his lectures. In addition to technical thermodynamics, Mollier's lectures also included its application for combustion engines and refrigerating machines. He also taught kinematics and dynamics which were his favorite subjects. He always lectured without any notes which enlivened his presentations extraordinarily.

Already after the first sentences, the students were fascinated. He was without doubt one of the most significant mediators of technical science. He imposed the duties of absolute clarity and consistency of presentation on a large number of professors who came from his school whereby his work was continued even after his death.

Even more than his students he stimulated his closer collaborators in his chair and laboratory with his scientific enthusiasm without inhibiting their free development. To each one, whom he considered capable, he suggested topics for their own research. He himself kept clear of all direction and organization work and resigned all consulting work in industry. He devoted all of his time to teaching and research.

It could not be avoided that his work was accorded complete recognition in public. So it came that in 1906 he became president of the Technische Hochschule in Dresden, in 1919 he received the honorary degree of a Doctor of Engineering from the Technische Hochschule in Braunschweig and in 1928 the Verein Deutscher Ingenieure (VDI) presented him with its highest distinction - the Grashof memorial award. On his 70th birthday in November 1933, all leading technical journals honored him and described his life's work. Sixteen months later, on March 13, 1935, Richard Mollier died, but his memory will remain unforgettable for us".

As the second speaker of the opening ceremony of the Congress Professor P. Glansdorff, Brussels, delivered the lecture "L'Ingenieur devant les problemes de la physique con­temporaine":

«Ludwig Boltzmann, dans la preface a ses Le�ons sur la Theorie des Gaz, nous livre cette amere reflexion:

10 Zeitschr. d. VDI, Bd. 67 (1923), S. 869 r r Zeitschr. d. VDI Bd. 73 (1929), S. 1009, Festschrift zum 70. Geburtstag von A. Stodola,

herausgeg. von E. Honegger, Ziirich, 1909, bei Orell und Fiissli 12 W. H. Carrier, Rational psychrometric formulae. Trans. Amer. Soc. Meehan. Engin. Bd. 33

(19n), S. 1005 13 W. Hiiussler, Das Mollier-i, x-Diagramm fiir feuchte Luft und seine technischen Anwendungen.

Dresden u. Leipzig 1960. Ver!. Theodor Steinkopff

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Fig. 6. At the stage of the Great Hall during the Opening Ceremony. From left: Mr. Thevenot, Prof. Glansdorff, Prof. Plank, Minister Balke, Prof. Nesselmann, Mayor Bayerle, President Foulon, Dr. Fidler.

"Ce serait a mon avis dommage pour la Science, que la theorie cinetique des gaz tombat momentanement dans l'oubli, par suite de l'hostilite generale a laquelle elle est actuellement en butte, de meme qu'autrefois, l'autorite de Newton fit succomber pendant quelque temps la theorie des ondulations".

"Je sais combien, quand on est isole, on demeure impuissant a !utter contre les cou­rants de son temps. Pourtant, j'ai voulu contribuer dans la mesure de mes forces, a ce qu'on ait pas trop de choses a redecouvrir quand on reviendra a la theorie des gaz".

Comme ces paroles resonnent etrangement aujourd'hui a nos oreilles, depuis que nous avons pris l'habitude d'etudier le comportement de la matiere a partir de ses pro­prietes atomiques ou moleculaires, avec !'aide d'une science des valeurs moyennes qui trouve precisement son origine dans l'reuvre de ce savant sur la theorie cinetique des gaz et atteint son epanouissement dans la vaste physique statistique contemporaine.

Sans doute, faut-il observer que l'isolement de Boltzmann n'etait pas absolu, puisqu'on peut compter parmi ses contemporains et meme predecesseurs des pionniers de la theorie des corpuscules aussi fameux que Kronig, Clausius, Maxwell et Willard Gibbs. Mais cette remarque n'enleve rien a la constatation d'une hostilite repandue a la fin du siecle dernier contre de telles conceptions. Contentons-nous plutot de reconnaitre combien Boltzmann avait vu juste et admirons la tranquille clairvoyance qui Jui a dicte ces paroles: "quand on reviendra a la theorie des gaz".

II en etait done intimement persuade. Et en effet, on y est revenu, si promptement meme que toutes nos connaissances actuelles sur Jes gaz rarefies, sur les gaz <lenses et condensables, sur les gaz ionises ou plasmas, sur Jes gaz d'electrons, sont largement tributaires des methodes introduites dans la science par le prestigieux physicien viennois. Les paroles que je viens de rapporter remontent a 1898, soit a 65 annees. Toutefois, c'est surtout au cours de l'essor qui a suivi la derniere guerre mondiale que !'influence boltzmannienne s'est manifestee dans toute son ampleur. Sachant avec quelle celerite le progres technique suit le progres scientifique, ii ne me parait pas temeraire d'avancer que l'heure est venue de donner a un nombre de plus en plus important de nos futurs

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ingenieurs une formation qui les familiarisera jeunes avec ces nouvelles methodes statisti­ques de pensee et qui, jointes, aux tout aussi recentes conceptions quantiques, devra leur permettre ensuite d'exploiter avec fruit les nombreuses ressources que la physique moderne offre, a present, tant sur !es proprietes des gaz, que des etats condenses et de 1' etat solide.

I1 ne faut pas chercher loin les exemples: La physique du solide rend compte des proprietes des semiconducteurs. Or les semiconducteurs se retrouvent a la base de !'ampli­fication de l'effet Peltier. Et l'effet Peltier lui-meme se presente actuellement comme un des procedes de refrigeration qui souleve le plus d'interet en raison des perspectives d'application nouvelles qu'il permet d'entrevoir. Les rapports de nos reunions et Con­gres en font foi.

Une amelioration importante de nos connaissances sur les coefficients de transport comme la viscosite ou la conductivite, sur les transitions de phase, sur !es jets gazeux, sur les effets naguere anormaux de supraconductivite et de superfluidite aux tres basses temperatures, sont autant d'autres exemples en rapport direct avec nos preoccupations actuelles.

J e me suis propose de resumer devant vous les principales raisons de cette profonde evolution, d'en situer brievement les causes et d'en tirer quelques conclusions. Ce sujet m'a semble en effet d'une portee suffisamment generale, et si en fait je l'ai oriente quelque peu vers la Thermodynamique qui est ma discipline favorite, j'espere qu'on me le pardonnera. Elle est d'ailleurs suffisamment generale.

La Thermodynamique classique a pris naissance au cours de la deuxieme moitie du dix-neuvieme siecle, sur la base des deux grands principes de conservation et d'evo­lution de l'energie. Elle a atteint rapidement un haut degre de perfection et sa fecondite a ete si importante que les savants qui ont eu la tache d'assurer cette moisson peuvent bien etre excuses a present, d'avoir neglige pendant ce temps les problemes corpuscu­laires de structure qui n'intervenaient pas directement dans leurs etudes et etaient d'ailleurs encore tres mal connus. Toute notre thermodynamique technique ainsi que la physicochimie sont des fruits de cette science qu'on appelait alors l'Energetique . . • A part quelques nuages, son ciel etait serein. On le sait, ces quelques nuages, c'etait !'incomprehensible spectre du rayonnement par incandescence et les anomalies des chaleurs specifiques aux temperatures les plus basses.

La curiosite savante s'est attaquee a ces nuages, mais au lieu de les dissiper, elle les a grossis et multiplies jusqu'a ce que, finalement, tout rentre heureusement dans l'ordre grllce a la conception geniale de Max Planck sur l'energie quantifiee. Bientot, la Mecani­que quantique qui s'en suivit venait fortifier d'une fac;:on decisive notre connaissance du comportement atomique et assurer notre confiance envers le monde des corpuscules elementaires, jusqu'a provoquer un revirement complet de toute la physique a partir de ces bases nouvelles.

Parallelement, un autre obstacle etait apparu. L'admirable edifice de la Thermodyna­mique classique presentait une faiblesse, car i1 ne permettait pas d'utiliser complerement la notion d'evolution contenue dans le principe de Carnot-Clausius. Seules les pro­prietes de l'equilibre avaient pu etre exploitees ainsi que la region voisine pour l'etude de la stabilite. En fait, l'Energetique apparaissait de plus en plus comme une thermo­statique plutot qu'une thermodynamique. Les coefficients des lois de cinetique devaient etre eux, empruntes a !'experience. Ils restaient done fatalement empiriques, sans lien entre eux et rassembles dans des tables. Telle est d'ailleurs encore la situation pour de nombreux chapitres de notre technique.

L'Histoire se repetait done, puisque la Mecanique elle-meme, science modele par excellence, et de beaucoup anterieure, s'etait developpee en passant par les memes stades: la statique d'abord deja tres complete chez les Anciens au debut de notre ere avec Archimede, la dynamique ensuite, beaucoup plus complexe et qui ne de­vait prendre forme qu'au dix-septieme siecle avec Newton. Toutefois pour la thermo­dynamique, le rythme aura ere beaucoup plus rapide, grllce surtout a !'effort pre­curseur de Boltzmann qui en elargissant la notion d'entropie, en la rattachant a une propriete de probabilite et par la a une mesure de la desorganisation d'un systeme,

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en faisant connaitre son fameux theorl:me H, et !'equation integrodifferentielle des gaz qui porte son nom, a ouvert largement Ia voie a l'etude des etats hors d'equilibre et aussi par consequent a la dynamique des transformations naturelles.

II exploitait done ainsi d'une manil:re beaucoup plus complete le contenu d'evolution enferme dans le principe de Carnot-Clausius. Cette dynamique porte aujourd'hui le nom de thermodynamique des phenoml:nes irreversibles, et son domaine s'est progres­sivement etendu bien au del:l de Ia theorie des gaz. Nombreux sont !es probll:mes d'ordre technique qui s'y rattachent des a present et parmi eux une place importante est occupee par le vaste ensemble des phenoml:nes stationnaires de non equilibre, sous l'effet des contraintes exterieures. A ce point de vue, Ia thermodiffusion, ou encore !es effets thermomecaniques et thermoelectriques comme l'effet Peltier dont nous avons deja cite I'exemple, relevent directement de cette discipline recente et en plein epanouissement.

Si, bientot, le rapide developpement de la biologie moleculaire devait permettre d'etablir un pont avec Ia physique, ce serait vraisemblablement grace a Ia thermo­dynamique des phenomenes irreversibles que ce pont pourrait etre jete. En effet, du point de vue de Ia physique, !es organismes vivants se presentent a nous comme des systemes hors d'equilibre puisqu'ils sont le siege de multiples echanges thermiques, massiques et chimiques, mais qui subsistent dans un erat stationnaire, ou tout au moins quasi stationnaire, pendant un intervalle de temps suffisamment long. L'approche de tels problemes ne pourra manquer de susciter des leur naissance I'interet de toutes nos commissions scientifiques. Defions-nous toutefois d'abonder dans Ia prophetie ou dans Ia fiction. Contentons-nous plutot d'observer que !'esprit de mode accompagne aussi Jes sciences dans Ieurs periodes de succl:s. Hier on ne voulait pas des conceptions corpus­culaires. Aujourd'hui, au contraire, une explication purement phenomenologique du type de Ia thermodynamique classique, qui ne rendrait pas compte du mecanisme micro­scopique qui Ia supporte, risquerait d'etre resolument rejete. II ne faut rien exagerer, car qu'adviendra-t-il demain ?

Le vaste et profond essor de Ia physique du vingtieme sil:cle, dont je viens d'esquisser a grands traits ce qui m'a semble !'aspect le plus voisin des travaux de notre Congres, s'est malheureusement accompagne d'une complexite des methodes de plus en plus importante. Une initiation prolongee est necessaire pour dominer le maniement des operateurs quantiques, pour se familiariser avec le formalisme mathematique du calcul des probabilites, et pour utiliser !es representations dans des espaces abstraits a un grand nombre de dimensions. Un outillage mental beaucoup plus encombrant s'impose ainsi aux jeune generations, suscitant chez Ieurs maitres une legitime anxiere, princi­palement pour tout ce qui concerne Ia formation des praticiens et des ingenieurs. Car comment alleger suffisamment les programmes d'etudes deja fort charges, pour y ajouter un pareil bagage ? J e me figure que cette question est a l'heure actuelle l'une des plus debattues dans les facultes d'ingenieurs des differents pays.

Et comme cela advient regulil:rement pour ces sortes de sujets, la classique querelle des anciens et des modernes y trouve de multiples occasions de se ranimer. Je ne crois pas qu'a cette occasion, la position trop prudente des partisans du statu quo integral puisse etre encore longtemps dffendue. Car, qui n'aper9oit qu'elle conduirait bientot notre technique vers cet erat lamentable qui n'est pas la mort mais qui n'est plus Ia vie et qu' on appelle a present le sous-developpement. Ou bien alors verrait-on par un reflexe nature! de defense, nos industries placer partout un physicien la ou ii manque un inge­nieur. De l'autre cote, les modernes opposent le plus souvent a cette prudence excessive un argument d'autorite : II faut aller de !'avant et faire confiance au pouvoir d'adaptation de la jeunesse. Voila qui est vite regle, mais non sans quelque danger ! Autant s'exprimer plus franchement et proclamer sans reserve : Ne creusons pas trop nos idees, suivons-les.

La difficulte de la mise en pratique serait d'ailleurs tout de suite considerable, parce que nous avons !'obligation democratique et sociale de maintenir la duree des etudes dans des limites raisonnables et acceptables pour toutes les couches de nos populations, parce que les programmes actuels sont deja trl:s encombres, parce que la subdivision en specialites encore plus diversifiees n'apporterait qu'une solution partielle a un pro­bteme qui touche de pres au contraire a la culture generale, parce qu'une specialisation trop hative est funeste parce qu'enfin ii y a aussi des branches concurrentes, comme !es

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sciences humaines ou les cours de langues, qui cherchent a trouver place dans les pro­grammes d'etudes pour ingenieurs.

Une idee plus seduisante, mais en apparence seulement, serait de voir assouplir la rigidite de nos reglements, afin de permettre aux mieux doues d'accomplir le cycle de leurs etudes a une cadence plus rapide que les autres. Mais qui oserait se risquer a faire un pareil choix, que l'avenir se chargerait de dementir trop souvent ? Pie de la Mirandole a ete le modele des etudiants de tous les temps et de tous les pays, mais il s'est montre dans l'age milr un auteur plutOt mediocre. Au contraire, Newton dont l'reuvre est un monument fut d'abord un mauvais eleve, pour qui l'etude n'avait aucun attrait. La premiere fois qu'il eprouva le besoin de travailler ce fut, d'apres Arago, pour conquerir la place d'un eleve turbulent qui, assis a cause de son merite sur une banquette superieure a la sienne, l'incommodait par ses coups de pied. A !'age de vingt-deux ans, il concourut pour un Fellowship de Cambridge, et fut vaincu par un certain Robert Uvedale, dont le nom sans cette circonstance, serait aujourd'hui completement oublie.

Ainsi done, le probleme qui nous preoccupe apparait tres complexe. Le developpement scientifique a tout accelere, sauf la vitesse de lecture de la page ecrite. Celle-ci a meme ete ralentie sous l'effet d'une abstraction mathematique croissante. En verite, le remede que suggere cette remarque semble surtout devoir provenir des quelques rares hommes de science, soucieux de synthetiser, de concretiser, de simplifier le cadre de nos con­naissances plutOt que d'en acquerir de nouvelles. D'eux seuls peut provenir cette eco­nomie de !'effort, qui rendra la physique nouvelle plus naturellement et plus imme­diatement accessible aux ingenieurs et a tous ceux qui auront a l'appliquer.

Ainsi eviterons-nous aussi, la naissance dans notre Institut du Froid, d'un schisme entre scientifiques et praticiens, dont le President Gorter a denonce le reel danger au cours des Congres precedents.Nous sommes tous reunis ici en effet sous la meme devise d'utilite, pourvu que le terme soit compris dans son sens le plus large, et alors seulement nos reuvres seront fecondes.

Chacun !'aura compris. Ce que nous attendons a present avec le plus d'impatience, ce que nous appelons de tous nos voeux, c'est la naissance sur le front commun de notre science et de notre technique, d'un nouveau Carl von Linde et d'un second Richard Mollier. »

OTHER ACTIVITIES OF THE CONGRESS

A number of technical visits were arranged in Munich and its surroundings during the Congress. After the Congress excursions were undertaken to various parts of Germany to visit research institutes, industrial companies and places of general interest.

In the entrance hall of the Technical University of Munich an exhibition about the history of refrigeration was organized. Besides old models - among others a model of the first refrigeration machine of Carl Linde - pictures of famous research workers and inventors in the field of thermodynamics, refrigeration, and air-conditioning techni­que of the whole world were shown.

Congress arrangements included a reception by the City of Munich in the old City Hall, a reception by the Government of the State of Bavaria in the "Alte Residenz", an opera evening in the famous Cuvillies-Theater with a performance of "Figaros Hochzeit", and a beer party in the "Lowenbrliu" on invitation of the German Association of Refrigeration, which was a general get-together for all Congress and associate members.

THE INTERNATIONAL INSTITUTE OF REFRIGERATION

During the Congress the I. I. R. arranged two meetings of the Technical Board, one meeting of the Executive Committee, and one meeting of the General Conference. In addition all Commissions of the I. I. R. held their business meetings.

The General Conference re-elected Professor Plank (Germany) as its President; it established general directions for the operation of the I. I. R. for the coming four years, i. e. until the next international congress. It was informed that the I. I. R. is still expanding and now includes 46 Member Countries. I. I. R. will in future extend its work to developing countries of the world.

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The General Conference elected Dr. B. K. Blount (United Kingdom) as President of the Executive Committee in replacement of Mr. F. Foulon (Belgium), whose mandate expired. As Vice-Presidents of the Executive Committee Messrs. David (France), Pentzer (U.S.A.), Beltran-Vivar (Spain), Eidsvik (Norway), and Ibl (Czechoslovakia) were elected.

The General Conference elected also the Presidents of the 9 Commissions: Com­mission 1 : Mr. Wilks (United Kingdom); Commission 2 : Mr. Kayan (U.S.A.); Com­mission 3 : Mr. Lorentzen (Norway) ; Commission 4: Mr. Bramsnaes (Denmark); Commission 5 : Mr. Baumgartner (Switzerland) ; Commission 6 : Mr. Ibl (Czecho­slovakia) ; Commission 7 : Mr. Fasoli (Italy) ; Commission 8 : Mr. Merlin (France); Commission 9 : Mr. van Hiele (The Netherlands).

On the Technical Board, Professor Kuprianoff (Germany) was elected President in replacement of Dr. Fidler (United Kingdom), whose mandate expired. Elected Vice­Presidents were Messrs. Glansdorff (Belgium), Jordan (U.S.A.) and Rutov (U.S.S.R.).

In gratitude to services rendered to the Institute, Messrs. Foulon and Fidler were elected i1s Honorary Presidents, and Messrs. van Itterbeek, Verlot, Kobulashvili and Hales as Honorary Members of the Institute.

CLOSING CEREMONY

The closing ceremony of the XI th International Congress of Refrigeration was held in the Great Auditorium for Physics of the Technical University of Munich. T.he Presi­dent of the General Conference, Professor Plank, opened the session and gave the floor to Mr. Thevenot, Director of the I. I. R., who reviewed the work of the Congress in the following address :

«Monsieur le President, Mesdames, Messieurs, 11 est de tradition que le Directeur de l'Institut International du Froid, a la seance

de cloture d'un Congres, rende compte brievement a !'assistance de ce qu'a ere le congres qui s'acheve.

C'est la une tache a la fois facile et malaisee. Facile parce que les impressions sont encore toutes fraiches des evenements que nous

venons de vivre ensemble pendant une semaine ; difficile parce qu'on ne peut resumer en un quart d'heure les activites qui ont occupe pendant 8 journees bien remplies plus de 1500 personnes et dont la preparation a coute des milliers d'heures de travail et d'efforts.

Essayons cependant de degager quelques points saillants dans cette importante mani­festation internationale qui se clot aujourd'hui a Munich.

Plus de 1500 participants avons-nous dit : ils provenaient de plus de 40 pays situes dans toutes les parties du monde, auxquels se sont joints les representants de IO organi­sations internationales. Des milieux professionnels tres divers, s'interessant directement OU indirectement au froid, etaient representes.

Lors du present Congres, on a sans doute accentue davantage la distinction que l'Institut International du Froid cherche a etablir entre les objectifs des congres qui se tiennent tous les quatre ans et ceux des nombreuses reunions des Commissions scienti­fiques et techniques de l'Institut qui ont lieu dans l'intervalle des congres. Alors que ces reunions de Commissions sont consacrees a des discussions approfondies, entre experts specialises, sur des sujets relativement etroits, on cherche davantage, aux congres, a apporter des informations genera/es sur des sujets d'actualite, destinees a une audience dont une part notable n'est pas specialisee dans le domaine du froid.

C'est dans cet esprit notamment que ce Xleme Congres International du Froid a inscrit a son programme 5 themes de seance pleniere (alors que 3 themes seulement avaient figure au Congres de Paris en 1955 et 3 themes aussi a Copenhague en 1959).

Vous avez sans doute presents a !'esprit les 5 themes de ces seances plenieres, que je me permets de rappeler :

applications technologiques des tres basses temperatures, et notamment de la supracon­ductivite

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refroidissement thermoelectrique ( Eff et Peltier) sources futures d' energie pour la production du froid comportement des aliments congeles en f onction du temps et de la temperature cryodessiccation.

Pour la plupart, ces themes ont ete repris au cours des sessions particulieres des Commissions competentes et y ont fait l'objet d'etudes plus approfondies. Trois questions d'actualite ont souleve particulierement l'inten�t au cours de ce congres, faisant ressortir les progres effectues depuis Copenhague : la lyophilisation, le refroidissement thermo­electrique (Peltier) et les applications technologiques des tres basses temperatures.

Les 9 Commissions scientifiques et techniques de l'Institut International du Froid ont tenu pendant ce Congres une quarantaine de sessions de travail, dont nous allons analyser tres succinctement le contenu.

275 rapports ont ete presentes et discutes lors de ce Congres, etablis par des auteurs de 25 pays. Les organisateurs allemands avaient imprime a l'avance tous ces rapports et ont pu en adresser les trois quarts aux participants pres d'un mois avant l'ouverture du Congres. La tache des Commissions s'en est trouvee facilitee et les discussions ont pu etre animees et fructueuses. 11 faut souligner aussi la qualite des interpretes attaches a chacune des Commissions, de sorte que la "barriere du langage" - un des reels obstacles rencontres dans des reunions scientifiques et techniques intemationales -a pu etre, sinon supprimee, du moins singulierement abaissee.

Resumons tres brievement les travaux des 9 Commissions :

Commission 1. Ses travaux ont porte sur trois categories de sujets : proprietes de la matiere aux tres basses temperatures; magnetisme et supraconductivite; applications des tres basses temperatures, notamment a la liquefaction des gaz et dans le domaine de la physique nucleaire; problemes de technique cryogenique, en particulier dans l' application des engins spatiaux.

Commission 2. Trois groupes de sujets ont retenu son attention: isolation thermique: comportement et efficacite des materiaux isolants de differents types, et notamment une discussion tres active sur les mousses de polymethane. transmission de chaleur dans les divers elements du circuit frigorifique thermodynamique de la production du froid.

A la Commission 3 l'activite des discussions sur les - compresseurs frigorifiques, notamment compresseurs alternatifs a grande vitesse et

compresseurs centrifuges a montre le grand interet que les frigoristes portent toujours a ce domaine traditionnel de production du froid.

Mais les etudes ont porte aussi sur les modes moins classiques : - machines frigorifiques a absorption - refroidissement thermoelectrique.

Ont ete etudies egalement les auxiliaires de production du froid : - condenseurs, evaporateurs et regulation automatique.

Commission 4. On peut classer en trois groupes les sujets discutes par cette Commission (le plus grand nombre de rapports presentes se rangeant dans le premier groupe) :

prob!emes d' actualite sur la preservation par le froid des denrees alimentaires perissables; leur comportement et I' influence des f acteurs dont on peut jouer pour ameliorer cette preservation (notamment) la viande, la volaille, le poisson, les fruits et les legumes) bases scientifiques des specifications et des normes pour produits congeles et lyophilises problemes de transmission de chaleur et rapidite du refroidissement (refrigeration et congelation) des aliments.

Commission 5. Les etudes de cette Commission ont porte sur les themes principaux ci­apres :

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les nouvelles conceptions de construction des entrepots frigorifiques et plus specialement l' examen des prix de revient des entrepots frigorifiques a un seul niveau l' exploitation des entrepots frigorifiques et notamment !es diverses methodes de congelation les particularites d' entreposage de certaines denrees.

Trois rapports de synthese ont ete presentes et discutes, resumant des enquetes effectuees sur:

1) !es chambres froides a doubles parois 2) la desodorisation des chambres froides 3) les stations fruitieres.

Commission 6. Les trois sous-Commissions specialisees de cette Commission ont etudie les sujets suivants :

- Sous-Commission 6-A: - developpements recents et tendances dans la technique du conditionnement d' air - ca/cul des besoins de froid dans !es installations de conditionnement d' air et moyens

de !es reduire. - Sous-Commission 6-B:

- certaines applications du froid dans diverses industries, notamment dans /es industries chimiques qui ont fait ressortir !es progres acquis au cours des quatre annees passees.

- Sous-Commission 6-C: Deux secteurs ont retenu l'interet : - la cryodessication: recherches sur !es methodes et sur !es produits lyophilises; developpe-

ments industriels; application dans le domaine medical, pharmaceutique et alimentaire - la preservation des tissus par le froid, notamment la conservation du sang.

En outre cette Commission a tenu une seance commune avec la Commission 3 sur la toxicite des fluides frigorigenes.

Commission 7. Les problemes de transport frigorifique terrestre etudies par cette Commission ont fait ressortir un certain nombre de conclusions : - que les methodes de construction des vehicules refrigerants sont en pleine evolution

et que les techniques nouvelles pourront sans doute permettre d'eliminer les defauts constates actuellement, tels que la corrosion des revetements, la penetration d'humi­dite a travers les isolants, etc . . • .

- que le transport en quantites toujours croissantes de denrees congelees et surgelees conduit a utiliser de plus en plus de vehicules avec equipement frigorifique; de nouveaux systemes sont realises a titre experimental, ou encore a l'etuc\e.

Commission 8. Les problemes de transport maritime traites par cette Commission ont concerne !es 4 themes ci-apres : - conditionnement d' air des navires - equipement frigorifique des bateaux de peche (secteur dont le developpement recent

a ere tres rapide) - regulation automatique des installations frigorifiques marines - transport par mer des gaz liquefies (en particulier methane).

Les rapports a la Commission 9 ont montre qu'il existe un grand interet pour les pro­blemes de /' enseignement du froid a tous /es niveaux, depuis la formation des personnels praticiens jusqu'a celle des ingenieurs et des chercheurs.

L'organisation de la recherche, notamment dans les pays en voie de developpement, a ere aussi evoquee, ainsi que certains aspects de /'utilisation domestique du froid.

D'autre part 5 films tres interessants ont illustre quelques applications nouvelles du froid.

Voila passes en revue, bien rapidement, les travaux scientifiques et techniques du Congres. 11 y a quelque chose d'un peu choquant, peut-on dire, a resumer en quelques minutes des etudes qui ont exige tant d'efforts des auteurs qui les ont preparees, des etudes sur lesquelles, pendant une semaine, 700 a 800 savants, ingenieurs, profession­nels de tous pays ont chaque jour echange leurs vues dans !es salles de cette Ecole Poly­technique. L'important ouvrage que formera le compte rendu de ce Congres - tel!.tes des rapports et des discussions - constituera une mise au point de haute valeur sur les problemes d'actualite dans le domaine du froid.

Si les etudes que nous venons d'evoquer succinctement representent evidemment l'essentiel des activites du Congres, il ne faut pas sous estimer !'importance et l'interet des visites techniques qui ont ete effectuees au cours de la semaine ecoulee, a Munich et dans ses environs. C'est en effet un des roles fort utiles de manifestations comme

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celle-ci que la possibilite qu'elles offrent a des personnes de pays et de professions tres divers de prendre contact avec !es realisations scientifiques et techniques les plus remar­quables du pays hOte.

L'expose que je viens de faire pour resumer les activites de ce Congres peut apparaitre bien severe, bien austere. Hatons nous de corriger !'impression fausse qu'il pourrait dormer. Nos amis allemands, au contraire, n'ont pas oublie les couleurs claires au tableau et ont SU agrementer et egayer ce serieux programme de travail. Reconnaissons qu'ils ont ete aides en cela par le cadre privilegie oil se deroulait ce Congres. Les richesses artistiques de la capitale bavaroise, la joie de vivre de ses habitants, !es beautes touristiques de la Baviere, tout cela a permis aux organisateurs d'offrir - notamment aux membres accompagnateurs - un programme de choix de distractions et a tous !es congressistes d'aimables occasions de detente.

Lors de chaque Congres, l'Institut International du Froid reunit ses organes de direction, et notamment sa Conference Generale qui, statutairement, tient ses assises tous !es 4 ans. La Conference Generale de l'Institut a reelu a sa presidence le Professeur Plank. Elle a procede a !'election des Presidents et Vice-Presidents des divers organes de l'Institut, et notamment aux remplacements des personnalites dont !es mandats venaient a expiration. C'est ainsi qu'elle a elu le Dr Blount a la presidence du Comite Executif, pour remplacer M. Foulon et le Professeur Kuprianoff a la presidence du Conseil Technique pour remplacer le Dr Fidler. La Conference Generale a exprime sa reconnaissance a MM. Foulon et Fidler pour les missions qu'ils ont accomplies pendant de nombreuses annees avec tant d'efficacite et de devouement en les nommant Presidents d'honneur de l'Institut International du Froid.

Les personnalites qui viennent d'etre elues ont demontre depuis longtemps deja leur attachement a l'Institut International du Froid et on peut etre assure qu'animes par une telle equipe, Jes comites et commissions de l'I. I. F. feront du bon travail pendant les annees a venir.

Avec l'achevement de ce Congres, une page est tournee dans l'histoire de l'Institut International du Froid et dans l'histoire de la grande famille internationale du froid.

Le ler Congres International du Froid s'est tenu, vous le savez, en 1908, ii y a 55 ans, (Bien peu d'Organisations internationales peuvent se prevaloir d'un aussi long passe) et si l'on regarde en arriere, on voit que !es 10 congres qui ont jalonne !'existence de l'I. I. F. ont toujours connu le succes : c'est que l'interet des savants et des techniciens, l'interet de !'opinion publique aussi, pour ce domaine du froid qui est le notre s'est continuellement renouvele, parce que !es applications scientifiques ou techniques du froid interessant la vie de tous !es jours ou presentant des aspects plus exceptionnels se renouvellent elles memes sans cesse.

Et ce XIeme Congres International du Froid l'a bien demontre. On a deja, a plusieurs reprises, felicite les organisateurs de ce Congres et on le fera

certainement encore dans les instants qui viennent: je me refuse done cette joie de dire plus longuement a nos amis allemands !es louanges qu'ils meritent. Mais je desire cepen­dant leur exprimer publiquement la reconnaissance de l'Institut International du Froid.

Comme Directeur de l'Institut, j'ai sans doute eu plus de contacts qu'un autre avec les organisateurs de ce Congres et je sais toute la peine qu'ils ont prise aussi bien a "!'echelon de Karlsruhe" - avec M. Nesselmann et M. Kuprianoff, M. Plank et M. Linge - qu'a "!'echelon de Munich" avec Mme Preuss et Melle Zwehl notamment.

Les travaux qui viennent de se derouler a Munich ont certes contribue au progres et au developpement de la science et de la technique du froid, et aussi !'esprit amical de collaboration et de cooperation qui a regne tout au long de cette grande manifestation a indiscutablement contribue a resserrer !'entente internationale.

Cette double mission, le XIeme Congres International du Froid l'a parfaitement accomplie. »

The new President of the Executive Committee, Dr. B. K. Blount, started by thank­ing the Institute for the honour which it had done him in electing him chairman of the Executive Committee. This was a duty as well as an honour; but he did not think the members would expect him to say much on this occasion ; they would have a right to information about his stewardship in four years' time. The main objective for the

xxviii

Institute during the next four years, he thought, was to bring into membership all the

remaining nations outside at present who had anything to give to the work of the In­stitute or anything to gain from it. He ended - speaking German - "Although French and English are the two languages of our Institute, I hope that I may be allowed to address a few words to our German hosts in their own language. We have all greatly enjoyed spending a week in this beautiful city of Munich. We have been very impressed with the antiquity of German history and buildings and with the youthful vigour of German technical ideas. Equally, we have been impressed by the warmth of the German welcome and the chill of the Munich beer - and that of the Munich weather, too, if I may be permitted to mention it. We all wish to thank you most heartily."

Thereupon the new President of the Technical Board, Professor Kuprianoff, spoke to the assembly :

"According to the convention of the I. I. R. a new Technical Board has been elected by the General Conference during its last meeting on Saturday, August 31, 1963, for the next period, extending from 1963 to 1967. Having in mind that the good reputation of the I. I. R. is depending on the reputation of its commissions and that on the other hand the reputation of a commission depends largely on the reputation and ability of

its president and its bureau, it is evident that the choice of personalities responsible for the commissions' work is of prime importance for the future of the Institute. Therefore, thorough consideration was given in establishing an active Technical Board, consisting of able and generally recognized personalities willing to work for the Institute. It is my honour and pleasure to present you today and now a list of the Vice-Presidents of the Technical Board and of the Presidents of the Commissions, forming an excellent team, devoted to the Institute.

As elected Vice-Presidents of Technical Board :

Professor Glansdorff (Belgium)

Professor Jordan (USA)

Mr. Rutov (U.S.S.R.).

The following have been elected or re-elected as Presidents of:

Commission 1

Commission 2

Commission 3

Commission 4

Commission 5

Commission 6

Commission 7

Commission 8

Commission 9

Dr. Wilks (United Kingdom)

Prof. Kayan (USA)

Prof. Lorentzen (Norway)

Dr. Bramsnaes (Denmark)

Dr. Baumgartner (Switzerland)

Mr. Ibl (Czechoslovakia)

Dr. Fasoli (Italy)

Mr. Merlin (France)

Mr. van Hiele (Netherlands).

As Dr. Fidler, past President of the Technical Board, has reported to the General Conference last saturday and as we all know very well, the reputation of the I. I. R. is a very good one. This has been achieved by the past "set up" of the commissions and therefore largely by the past Technical Board. Permit me, Ladies and Gentlemen, to take this occasion to express my sincerest thanks to the last Technical Board and its so active President Dr. Fidler. The Institute owes to all of them a deep gratitude for the excellent but hard work they have done, initiated by a deep devotion to the Institute. I think we may say that if this Congress was a successful one, this was also partially because the bureaus of the commissions had performed a great effort in preparing it and in conducting their many meetings. I would like to thank all these bureaus of the commissions, all authors of the papers, all chairmen and secretaries of the meetings and all participants in the discussions for their excellent cooperation. A special thank I would like to forward to the Presidents of the Commissions, who are leaving the Tech-

xxix

nical Board after 8 years of service, because - according to the rules of the I. I. R. - they could not be re-elected ; they are :

and

Professor van Itterbeek (Belgium), past President of Commission 1, Mr. Verlot (France), past President of Commission 5, Mr. Kobulashvili (U.S.S.R.), past President of Commission 6 Mr. Hales (United Kingdom), past President of Commission 8.

The future work of the Institute's Commissions will be discussed at the first meeting of the Technical Board this afternoon. Therefore, Ladies and Gentlemen, I am not able to give you an account on what is planned. Let us hope that the future program will be a good one and that the Commissions will remain as active as in the past. I think the Institute should continue and even make closer and deeper its contact with other inter­national organizations. But it seems to me to be most important that the Institute should remain also very active and elastic in the methods of its own work, for instance by in­stalling more working-parties for special problems. I am sure that this will be an efficient way to solve many problems and to perform a great deal of the work of the Institute in the future. Let me, please, wish the best to all Commissions for their future work."

Professor Plank closed the XI th International Congress of Refrigeration with the following farewell address :

"A closing ceremony has always the taste of sadness. We don't like to close, we prefer to continue or to reopen. An end should always mean a new beginning.

Each International Congress is an event of greatest importance. The intellectual forces of all countries accumulate to an imposing manifestation and show the ever increasing importance of our branch of sciences and industry.

Better than many other associations the International InstituteofRefrigerationhas trans­formed a great number of scientists, engineers, biologists and economists into members of one great family, perfectly understanding each other even speaking different languages.

New officers begin now to lead our Institute, we wish that they may earn the highest possible success and feel happy in accomplishing their responsible duties. The resting pole in our Organization remains unchanged - that is the personality of our venerated Director, M. Thevenot. We sincerely hope that he will stay with us for many years.

J e souhaite a vous tous de conserver votre elan vital et vos capacites creatrices pour continuer a developper notre art et nos connaissances.

Quand nous allons nous revoir a l'ecoulement de quatre annees dans un autre pays du globe, il y aura un grand nombre de nouvelles idees a discuter; les jeunes chercheurs seront avances et occuperont des positions plus responsables, les personnes agees auront peut-etre transforme une partie de leur savoir et de leurs connaissances en sagesse, qui leurs permettra de ceder la place a ceux qui nous suivent.

Mais il n'y aura pas de changement dans nos sympathies et nos amities, nous conser­verons notre estime l'un pour l'autre, et notre tolerance intellectuelle au dela des diffe­rents points de vue politiques qui peut-etre aussi pourront se rapprocher.

Rempla9ons done "!'Adieu" par un "Au Revoir". Notre organisation nous garanti qu'il y aura de nombreux rencontres au sein de notre Conseil Technique avant que nous allons nous rassembler en totalite pour tenir le Xllme Congres International du Froid."

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Officers of the Xlth International Congress of Refrigeration Personnalites du XI• Congres International du Froid

PATRON OF THE CONGRESS -SOUS LE HAUT PATRONAGE DE

Dr. h. c. Heinrich Labke, President of the Federal Republic of Germany

HONORARY PRESIDENT - PRESI­DENT D'HONNEUR

Prof. Dr. Dr. h. c. Ludwig Erhard, Mi­nister of Commerce of the Federal Repu­blic of Germany

SECRETARY GENERAL - SECRE­TAIRE GENERAL

K. Nesselmann

ORGANIZING COMMITTEE -COMITE D'ORGANISATION

K. Nesselmann (Chairman), E. Fink, J. Kuprianoff, K. Linge, R. Plank, H. H. Schrader

STEERING COMMITTEE -COMITE DE DIRECTION

H. Linde (Chairman), J. Kuprianoff, K. Linge, H. Loewer, K. Nesselmann, Mrs. L. Preuss, W. Simon

FINANCE COMMITTEE - COMITE DES FINANCES

K. Linge (Chairman), W. Baer, E. Metzenauer, J. Simon, H. H. Schrader

PAPERS' COMMITTEE - COMITE DES RAPPORTS

J. Kuprianoff (Chairman), L. von Cube, H. Glaser, H. Hausen, E. Schmidt, Th. E. Schmidt, W. Tamm, H. U. Thormann, 0. Wagner

PUBLICITY COMMITTEE - CO­MITE DE PUBLICITE

Th. Sexauer (Chairman), Mrs. L. Preuss, H. Schenk

VISITS AND EXCURSIONS COM­MITTEE - COMITE DES VISITES ET EXCURSIONS

H. Bach, Mrs. L. Preuss, W. Tamm

NON-TECHNICAL PROGRAM AND HOSPITALITY COMMITTEE COMITE D' ACCUEIL ET DU PROGRAMME NON-TECHNIQUE

Mrs. L. Preuss, W. Simon

ASSOCIATES COMMITTEE - CO­MITE DES MEMBRES ACCOMPAG­NATEURS

Mrs. L. Gottler-Ries, Mrs. M. Kupria­noff, Mrs. A. Linde, Mrs. V. Linge, Mrs. W. Nesselmann, Mrs. L. Plank, Mrs. L. Preuss, W. Simon

INTERPRETERS - INTERPRETES G. Peterolff (Manager), Miss I. Alenfeld,

Mrs. L. O'Brien, Mrs. M. Callon, R. Cam­bien, Miss E. Stadelmann, E. Wachtler, E. Weintraub

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International Institute of Refrigeration - Officers Institut International du Froid -Personnalites

HONORARY PRESIDENTS - PRE­SIDENTS D'HONNEUR

Dr. H. Queuille, Ancien President du Conseil des Ministres (France)

Prof. C. J. Gorter, Directeur du Labo­ratoire Kamerlingh Ormes, Universite de Leyde, Netherlands

HONORARY DIRECTOR - DIREC­TEUR HONORAIRE

M. Ch. David, Directeur General du Genie Rural et de l'Hydraulique Agricole, Ministere de !'Agriculture, Paris, France

GENERAL CONFERENCE - CON­FERENCE GENERALE President:

Prof. Dr. R. Plank, Kiiltetechnisches Institut der Technischen Hochschule, Karlsruhe, Germany

EXECUTIVE COMMITTEE - CO­MITE EXECUTIF President:

M. J. Foulon, Directeur General de la Regie des Services Frigorifiques de l'Etat Belge, Bruxelles, Belgique

Vice-Presidents:

Dr. W. H. Cook, Director, Division of Applied Biology, National Research Coun­cil, Ottawa, Canada

Dr. R. C. Jordan, Head Department of Mechanical Engineering, University of Minnesota, Minneapolis 14, U.S.A.

M. Ch. Koboulachvili, Directeur de l'Institut de Recherches Scientifiques de l'Industrie du Froid de l'U.R.S.S., Co­mite d'Etat de Planification, Moscou, U.R.S.S.

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MANAGEMENT COMMITTEE COMITE DE DIRECTION

M. J. Pou/on, President, Belgium; M. E. Baumgartner, Switzerland; M. R. Ed­monds, U.K. ; M. A. Cuttica, Italy; M. L. Vahl, Netherlands; M. 0. Wagner, Ger­many.

TECHNICAL BOARD - CONSEIL TECHNIQUE President :

Dr. J. C . Fidler, Ditton Laboratory, Maidstone, Kent, U.K.

Vice-Presidents:

Prof. P. Glansdorff, Faculte Polytech­nique de Mons et Universite de Bruxelles, Bruxelles, Belgique

M. M. Gorbounov, Chef de la Section du Comite d'Etat de !'automation et des constructions mecaniques, Moscou, U.R.S.S.

Dr. W. T. Pentzer, Chief Biological Sciences Branch, Marketing Research Division, U.S. Department of Agriculture, Washington 25, D.C., U.S.A.

DIRECTOR OF THE INSTITUTE -DIRECTEUR DE L'INSTITUT

M. R. Thevenot, Institut International du Froid, 177, Boulevard Malesherbes, Paris 17 •, France

Donor Members Membres donateurs

DENMARK Danfoss, Nordborg

FRANCE Association Fran9aise du Froid, Paris Chantiers de l' Atlantique, Saint-Denis Compagnie des Entrepots et Gares Frigorifiques, C.E.G.F., Paris Compagnie des Entrepots Frigorifiques de l'Ouest, Le Havre ISOLFEU S.A., Paris Societe d'Etudes Techniques lndu­strielles & Frigorifiques, Paris Union Syndicale Nationale des Ex­ploitations Frigorifiques, Paris

GERMANY Arbeitsgemeinschaft Kalte-Industrie, Frankfurt/Main Badische Anilin- und Sodafabriken, Ludwigshafen/Rhein Deutscher Kiiltetechnischer Verein e.V., Karlsruhe Danfoss, GmbH., Flensburg Eisfink, Carl Fink oHG., Asperg/Wttbg. Fachverband der Ktihlhiiuser und Eis­fabriken, Hamburg Farbwerke Hoechst, Frankfurt/Main Gesellschaft fi.ir Linde's Eismaschinen A.G., Hollriegelskreuth Gesellschaft fi.ir Markt- und Ktihlhallen, Hamburg

Ideal-Standard, Rheinkalte Plant, Dus­seldorf Kalle A.G., Wiesbaden KUBA, Ktihlerfabrik, Baierbrunn SUMAK, Wilh. Weckerle, Stuttgart Teves A., Maschinen- und Armaturen­fabrik K.G., Frankfurt/Main Vereinigte Deutsche Metallwerke, Wer­dohl/Westfalen

NETHERLANDS N. V. Philips Gloeilampenfabrieken, Eindhoven

SPAIN Centro Experimental del Frio, Madrid Comisario General de Abastecimientos y Transportes, Madrid Instituto Nacional de Industria, Madrid

SWEDEN Frigoscandia AB, Halsingborg STAL Refrigeration AB, Norrkoping

UNITED KINGDOM J. & E. Hall Ltd., Dartford, Kent Wm. Douglas & Sons (Eng) Ltd., Lon­don Refrigeration Press Ltd., Croydon

U.S.A. Reynolds Metals Company, Richmond, Va. USA

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Sponsor Organizations Organisations Donatrices

Bundesministerium fiir Wirtschaft, Bonn Bayerische Staatsregierung, Miinchen Landeshauptstadt Miinchen Deutscher Kaltetechnischer Verein e. V.,

Karlsruhe Arbeitsgemeinschaft Kalte-Industrie,

Frankfurt/M. AEG, Alig. Elektr. Ges . . Kassel-Bettenhaus. Alpinakalte, Kaufbeuren/ Allg. Arnold Kiihlung, Ludwig Arnold KG.,

Friedberg b. Augsburg Assmann & Stockder KG., Stuttgart-Bad

Cannstatt Atlas-Werke A.G., Bremen Bad. Anilin- & Sodafabrik A.G., Lud­

wigshafen a. Rh. Bauknecht GmbH., Stuttgart-S Bergedorfer Eisenwerke A.G., Hamburg­

Bergedorf Bitzer Kiihlmaschinenbau GmbH., Sin-

delfingen/Wttbg. Bock u. Co. K.G., Ntirtingen Borsig A.G., Berlin Robert Bosch GmbH., Stuttgart J ohannesBurmester&Co.,Geesthacht/Elbe Brown, Boveri & Cie., Mannheim Danfoss GmbH., Flensburg DE-STA-CO Metallerzeugnisse GmbH.,

Frankfurt/M. Deutsche Ranco GmbH., Hockenheim/Bd. Deutsche Waggon- u. Maschinenfabriken

GmbH., Berlin Otto Egelhof, Stuttgart-Fellbach H. D. Eichelberg & Co., GmbH., Men-

den/Sauerland Eisfink Carl Fink oHG., Asperg/Wttbg. Elektrolux GmbH., Hamburg 13 Elmore's Metall A.G., Schladern/Sieg Escher Wyss GmbH., Werk Lindau/Bo. F. A. S. Hans Jorgensen, VDI, Hamburg-

Stellingen Farbenfabriken Bayer A.G., Leverkusen Ernst Flitsch, Stuttgart-Fellbach Frigidaire-Werk der Adam Opel A.G.,

Rtisselsheim a. M. Gebr. Bratzler, Arktis-Tiefktihlkost,

Karlsruhe/Ed. Gebr. Neunert,Maschinenfabrik,Elmshorn Gebr. Plersch, Illertissen/Bayern Gebr. Ricken, Ktihlmobelindustrie,

Wattenscheid Gerling, Holz & Co, Hamburg-Altona 1 Ges. f. Linde's Eismaschinen A.G., Wies­

baden Grtinzweig & Hartmann A.G., Lud­

wigshafen a. Rh. Haenni & Cie., m. b. H., Stuttgart-Bad

Cannstatt Hansa Metallwerke A.G., Stuttgart-Moh­

ringen Heimerle & Meule KG., Pforzheim

Erich Herion, Spezialfabrik, Stuttgart Wilhelm Hilkenbach, Opladen Homann Werke, Wilhelm Homann, Wup-

pertal-Vohwinkel Ideal Standard GmbH., Bonn/Rh. M. K. Juchheim, Fulda Philipp Kirsch, Offenburg/Bd. KUBA Ktihlerfabrik H. Schmitz, Baier­

brunn b. Miinchen Leybold-Hochvakuum-Anlagen GmbH.,

Koln-Bayenthal Hans Liebherr, Ochsenhausen/Wttbg. Karl Mack, Nellingen-EBlingen Margarine-Union, Hamburg Maschinenbau A.G. Bakke, Bochum Mecano-Bundy GmbH., Heidelberg Metzenauer & Jung GmbH., Wuppertal-

Elberfeld Neff-Werke GmbH., Bretten Neumann & Reichel KG., Dtisseldorf Hans Nobis, Koln-Hohenberg Projahn-Werke KG., Waldbroel/Rhld. Prometheus GmbH., Eschwege/Werra Rheinhold & Mahla GmbH., Mannheim R. & G. Schmale, Metallwerke, Menden/

Sauerland Siegas Metallwarenfabrik Wilh. Loh KG.,

Siegen/W estf. Siemens Elektrogerate A.G., Miinchen 2 Siller & Rodenkirchen GmbH., Roden­

kirchen/Rh., SUMAK Wilh. Weckerle, Stuttgart­

Zuffenhausen Standard Kessel Ges., Gebr. Fasel,

Duisburg-Meiderich Standard Messo Ges. f. Chemietechnik

m.b.H., Duisburg Stempel-Hermetik GmbH., Frankfurt/M.-

Stid Otto Sterkel, Ravensburg Stierlen-Werke A.G., Rastatt/Bd. Alfred Teves KG., Frankfurt/M. ULEIFA Ulmer Eisschrankfabrik, A. u.

E. Hummel, Ulm/Donau Ver. Deutsche Metallwerke A.G. Zweig-

niederl. Carl Berg, W erdohl/W estf. Vorwerk & Co., Wuppertal-Barmen Robert Wahl KG., Balingen/Wttbg. Weber & Freund, Stuttgart W. Wellit GmbH., Dtisseldorf Wieland-Werke A.G. Metallwerke, Ulm/

Donau Theodor Witt, Maschinenfabrik, Aachen

Organizations Providing Gifts Organisations ayant presente des cadeaux Danfoss GmbH., Nordborg Rheinhold & Mahla, Mannheim Carl Zeiss A.G., Oberkochen/Wttbg. Arbeitsgemeinschaft Kalte-Industrie,

Frankfurt/M Farbwerke Hoechst A.G., Frankfurt/M

xxxv

Sponsor Organizations Organisations Donatrices

Bundesministerium fiir Wirtschaft, Bonn Bayerische Staatsregierung, Miinchen Landeshauptstadt Miinchen Deutscher Kaltetechnischer Verein e. V.,

Karlsruhe Arbeitsgemeinschaft Kalte-Industrie,

Frankfurt/M. AEG, Alig. Elektr. Ges . . Kassel-Bettenhaus. Alpinakalte, Kaufbeuren/ Allg. Arnold Kiihlung, Ludwig Arnold KG.,

Friedberg b. Augsburg Assmann & Stockder KG., Stuttgart-Bad

Cannstatt Atlas-Werke A.G., Bremen Bad. Anilin- & Sodafabrik A.G., Lud­

wigshafen a. Rh. Bauknecht GmbH., Stuttgart-S Bergedorfer Eisenwerke A.G., Hamburg­

Bergedorf Bitzer Kiihlmaschinenbau GmbH., Sin-

delfingen/Wttbg. Bock u. Co. K.G., Ntirtingen Borsig A.G., Berlin Robert Bosch GmbH., Stuttgart J ohannesBurmester&Co.,Geesthacht/Elbe Brown, Boveri & Cie., Mannheim Danfoss GmbH., Flensburg DE-STA-CO Metallerzeugnisse GmbH.,

Frankfurt/M. Deutsche Ranco GmbH., Hockenheim/Bd. Deutsche Waggon- u. Maschinenfabriken

GmbH., Berlin Otto Egelhof, Stuttgart-Fellbach H. D. Eichelberg & Co., GmbH., Men-

den/Sauerland Eisfink Carl Fink oHG., Asperg/Wttbg. Elektrolux GmbH., Hamburg 13 Elmore's Metall A.G., Schladern/Sieg Escher Wyss GmbH., Werk Lindau/Bo. F. A. S. Hans Jorgensen, VDI, Hamburg-

Stellingen Farbenfabriken Bayer A.G., Leverkusen Ernst Flitsch, Stuttgart-Fellbach Frigidaire-Werk der Adam Opel A.G.,

Rtisselsheim a. M. Gebr. Bratzler, Arktis-Tiefktihlkost,

Karlsruhe/Ed. Gebr. Neunert,Maschinenfabrik,Elmshorn Gebr. Plersch, Illertissen/Bayern Gebr. Ricken, Ktihlmobelindustrie,

Wattenscheid Gerling, Holz & Co, Hamburg-Altona 1 Ges. f. Linde's Eismaschinen A.G., Wies­

baden Grtinzweig & Hartmann A.G., Lud­

wigshafen a. Rh. Haenni & Cie., m. b. H., Stuttgart-Bad

Cannstatt Hansa Metallwerke A.G., Stuttgart-Moh­

ringen Heimerle & Meule KG., Pforzheim

Erich Herion, Spezialfabrik, Stuttgart Wilhelm Hilkenbach, Opladen Homann Werke, Wilhelm Homann, Wup-

pertal-Vohwinkel Ideal Standard GmbH., Bonn/Rh. M. K. Juchheim, Fulda Philipp Kirsch, Offenburg/Bd. KUBA Ktihlerfabrik H. Schmitz, Baier­

brunn b. Miinchen Leybold-Hochvakuum-Anlagen GmbH.,

Koln-Bayenthal Hans Liebherr, Ochsenhausen/Wttbg. Karl Mack, Nellingen-EBlingen Margarine-Union, Hamburg Maschinenbau A.G. Bakke, Bochum Mecano-Bundy GmbH., Heidelberg Metzenauer & Jung GmbH., Wuppertal-

Elberfeld Neff-Werke GmbH., Bretten Neumann & Reichel KG., Dtisseldorf Hans Nobis, Koln-Hohenberg Projahn-Werke KG., Waldbroel/Rhld. Prometheus GmbH., Eschwege/Werra Rheinhold & Mahla GmbH., Mannheim R. & G. Schmale, Metallwerke, Menden/

Sauerland Siegas Metallwarenfabrik Wilh. Loh KG.,

Siegen/W estf. Siemens Elektrogerate A.G., Miinchen 2 Siller & Rodenkirchen GmbH., Roden­

kirchen/Rh., SUMAK Wilh. Weckerle, Stuttgart­

Zuffenhausen Standard Kessel Ges., Gebr. Fasel,

Duisburg-Meiderich Standard Messo Ges. f. Chemietechnik

m.b.H., Duisburg Stempel-Hermetik GmbH., Frankfurt/M.-

Stid Otto Sterkel, Ravensburg Stierlen-Werke A.G., Rastatt/Bd. Alfred Teves KG., Frankfurt/M. ULEIFA Ulmer Eisschrankfabrik, A. u.

E. Hummel, Ulm/Donau Ver. Deutsche Metallwerke A.G. Zweig-

niederl. Carl Berg, W erdohl/W estf. Vorwerk & Co., Wuppertal-Barmen Robert Wahl KG., Balingen/Wttbg. Weber & Freund, Stuttgart W. Wellit GmbH., Dtisseldorf Wieland-Werke A.G. Metallwerke, Ulm/

Donau Theodor Witt, Maschinenfabrik, Aachen

Organizations Providing Gifts Organisations ayant presente des cadeaux Danfoss GmbH., Nordborg Rheinhold & Mahla, Mannheim Carl Zeiss A.G., Oberkochen/Wttbg. Arbeitsgemeinschaft Kalte-Industrie,

Frankfurt/M Farbwerke Hoechst A.G., Frankfurt/M

xxxv

I. Plenary Meetings I. Seances Plenieres

SESSIONS :

Technological Advances

Using Very Low

Temperatures

Peltier-Effect

Energy for Refrigeration

in Coming Years

Time-Temperature­

Tolerance for Frozen Foods

Freeze-Drying

Applications technolo­

giques des tres basses

temperatures

Effet Peltier

Energie pour la produc­

tion du froid a l'avenir

Comportement des ali­

ments congeles, en

fonction du temps et de

la temperature

Cryo-dessiccation

Technolog ical Advances U s i n g Very Low Tem peratu res

Appl i cati o n s tech nologiques des t res basses t e m peratu res

P-3

Technological Applications of Very Low Temperatures

Applications technologiques des tres basses temperatures

J. WILKS Clarendon Laboratory, Oxford, England

SOMMAIRE. Jusqu'a ces temps derniers, on n'utilisait l'hydrogene et /'helium liquides au laboratoire qu' en petites quantites. Actuellement, ii existe de nombreuses applications techno­logiques, souvent a assez grande echelle. L'hydrogene liquide est produit et manipute engrandes quantites comme combustible pour /es fusees spatiales. L'experience cryogenique acquise ainsi permet la construction de tres grands systemes a basse temperature. On est en train de mettre au point des groupes cryogeniques pour obtenir des vides pousses dans des chambres reproduisant !es conditions spatiales et pour fairefonctionner des tunnels a air a basse pression. On a construit des chambres a bulles de grandes dimensions pour la recherche en physique nucleaire. L'hilium Liquide en particulier est maintenant utilise couramment dans les laboratoires de physique de l' etat solide et tres utilise par exemple pour l' etude des metaux. Les appareils fonctionnant a la temperature de !'helium liquide deviennent maintenant importants dans la pratique. !ls comprennent /es detecteurs infrarouge et le recepteur de «maser» utilise dans les systemes de communication de radio des satellites et, ce qui est le plus important que tout, des circuits supraconducteurs. On decrit brievement les diverses applications et l'on donne des references bibliographiques.

Liquid oxygen (boiling point 90° K) has long been a well known feature of the industrial scene and world production is now of the order of 70,000 tons per day. It is used primarily as a convenient source of gaseous oxygen, although it has many specifically low tempera­ture applications. Liquid nitrogen (b. p. 80° K) is also now used for a wide variety of pur­poses, and can be purchased in bulk. These liquids are probably fairly well known to you, so this paper is essentially concerned with two colder fluids : liquid hydrogen (b. p. 20°K) and liquid helium (b. p. 4.2°K).

Twenty years ago both these liquids were produced only in quantities of a few litres in a small number of physics laboratories. Although they were produced in similar ways to other cryogenic fluids they had no technological applications. They were however of great interest to the physicist, enabling him to reach much lower temperatures where strange new phenomena occurred. By 1950 the advent of the Collins cryostat permitted any phys­ics laboratory to produce liquid helium or hydrogen at a relatively modest cost, and a big expansion in the laboratories was under way. Since then the scale of use has steadily increased, and today hydrogen is liquefied primarily for technological purposes. In this article we briefly review some of the new applications of these liquids.

ROCKET FUELS

The great expansion in the use of liquid hydrogen has arisen because of its merits as a propellant for space rockets [I,2]. Most propellants consist of a 'fuel' and an 'oxidiser', and a combination of liquid hydrogen and liquid oxygen gives almost the highest thrust per

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second per lb. of propellants. (Fluorine gives even better results than hydrogen but is extremely corrosive and toxic.) In addition to the vital question of the thrust produced per lb. of fuel, hydrogen has other advantages. Its low molecular weight results in the burning gases leaving with a high velocity, and one can fairly easily obtain a high efficiency of combustion. Although the low temperature of the hydrogen is irrelevant to its use as a fuel, the liquid can be used as a coolant in the rocket prior to combustion. On the debit side, hydrogen has the reputation of being a hazardous fluid to handle, but care and ex­perience have shown that safety can be achieved [3].

Rockets are often very large, so that great quantities of liquid hydrogen are required for them. Plants have been built to produce of the order of 30,000 gallons/day, and quan­tities of this order can be transported by road and rail tankers [ 4]. Spectacular progress has also been made in methods of insulation, particularly by the introduction of the so called 'superinsulants'. Thus the loss by evaporation in tanks such as those just mentioned is less than }-2 % per day.

CRYOPUMPING : SPACE SIMULATION

We now describe three other applications that depend on the large scale availability of liquid hydrogen. The first is also related to rockets : the development of space simulation chambers. The pressure of the atmosphere falls off very rapidly as we move out into space. Until recently a pressure of 10-6 mm of mercury was usually regarded as a very good vac­uum. Yet 450 miles up the pressure is only 10-9 mm, 1200 miles up only 10-11 mm, and in interplanetary space about 10-16 mm. Now high vacua affect materials in rather unex­pected ways [5]. For example graphite, which is usually a good lubricant, acts as an abrasive at pressures below l0-6 mm! (It seems that its lubricant properties depend on the pres­ence of absorbed gases which are stripped off in a high vacuum.)

Hence there is a need for very large high-vacuum chambers to test out full size compo­nents for space vehicles. One way of producing vacua is by cryopumping [5, 6]. The air in the chamber is first pumped out in the normal way, and the remainder is then frozen out by circulating liquid hydrogen through panels in the chamber. Using this technique, the freezing panels can be placed very close to the equipment being tested ; hence if any gas is desorbed it is quickly frozen and the vacuum maintained. (In other pumping systems, such gas probably has to flow an appreciable distance before it reaches the pump.)

Space is also very cold and the liquid hydrogen (or helium) may also be used to simulate its low temperature as well as the high vacuum (although for many purposes it is adequate to use liquid nitrogen [6].)

CRYOGENIC PUMPING : WIND TUNNELS

Another large scale application connected with the space programme, but which is still in the design stage, is the use of cryogenic wind tunnels. There is obviously a need to study the aerodynamics of rockets in very rarefied atmospheres, and this involves pump­ing gases which occupy large volumes but have little mass. The conventional approach is to use very large oil diffusion pumps, but at very low pressures it should be economi­cally advantageous to pump the air through by freezing it out at one end [7].

BUBBLE CHAMBERS

Bubble chambers [8] are devices used by the physicist to study the very strange particles such as mesons and hyperons which are formed in high energy reactions between nuclei. A liquid is heated under pressure to above its normal boiling point and the pressure relax­ed. Boiling should then occur but cannot begin unless some impurities are present; in this case the liquid is superheated and unstable. Should a nuclear particle pass through, it initiates boiling and its track is marked by a line of bubbles. This nuclear particle may also collide with the nuclei of the atoms forming the liquid and give rise to a nuclear reaction which yields new particles each making a characteristic track. By photographing such tracks a great deal has been learnt about these particles which present one of the

4

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chief problems in nuclear physics. Although several liquids have been used, liquid hydro­gen and helium have particular advantages, as they contain only one sort of nuclei and those are very simple ones.

To observe a reasonable number of nuclear events in the tracks, the chamber should be as large as possible. The first large hydrogen chamber at the University of California at Berkeley is about 6 ft. long and 2 ft. wide, and contains about 500 litres of liquid [9]. The cost was about $ 2,000,000. The fact that hydrogen is very cold is again quite irrelevant to the mode of operation of the chamber, although of course great care had to be paid to the low temperature behaviour of the materials used in its construction.

HEAT AND VIBRATION

We now consider some applications of liquid hydrogen and helium which depend on their low temperature. First let us ask what happens to a body when we supply or remove heat. At the lowest temperatures, the atoms in a crystalline solid form a regular array or lattice, while at higher temperatures they vibrate irregularly about their mean positions. As heat is supplied the atoms gain energy and vibrate more strongly. This heat motion greatly influences many properties of materials, as for example the flow of electricity. When a voltage is applied, the electrons in a metal can move easily if the atoms form a regular array as in a pure material. (We do not stop to explain why.) The metal is then said to be a good conductor. If the metal contains impurities these break up the regularity of the lattice, scatter the electrons, and reduce the flow of current. Hence alloys are less good conductors than pure metals. Similarly, if we cool a pure metal, the vibrational motion is reduced, the atoms form a more regular array, the electrons can move more freely and the conductivity increases. This effect can be very large, for example the conductivity of very pure copper may increase several hundred times on cooling down to helium temperatures.

Although we have mentioned electrical resistance primarily to illustrate the general effect of temperature, its reduction leads to an interesting possibility regarding the design of electro-magnets. In very high power magnets much waste heat is produced in the wind­ings and has to be removed. Under certain circumstances it may become economic to cool with liquid hydrogen or helium in spite of their high cost [IO]. The very considerable reduction in resistance greatly reduces the heating, so that less power is wasted as heat and less heat has to be removed. (In fact such magnets may be superceded by other types to be described in the next talk.)

INFRA-RED DETECTORS

There is currently much interest in infra-red radiation. All bodies radiate electromag­netic waves : very hot bodies radiate visible light, while bodies at room temperature radiate longer waves in the invisible infra-red region. If we can detect these waves, then we can 'see in the dark'. Already detectors, rather like television cameras can take pictures of a man in a darkened room entirely by his own radiation [ 1 1] . Great possibilities exist for navi­gational and detectional devices, while in medicine infra-red photography has been used to diagnose tumours.

There is a great need to increase the sensitivity of the detectors especially at the longer wavelengths. Like all electromagnetic waves, infra-red radiation travels in small units of energy called photons. The energy of a photon of visible light is very much greater than the thermal energy of the atoms in the working material (usually a 'semi-conductor') of the detector. However the energy of a photon decreases at longer wavelengths, and becomes comparable to the vibrational energy of the atoms. The photon then becomes 'lost' in the thermal motion of the detector. By cooling the semi-conductor (which is quite small and generates little heat) to hydrogen or helium temperanires we reduce the thermal motion and obtain a more sensitive instrument. Devices such as the above are clearly going to create a growing need for small, simple and reliable refrigeration units. It is almost certainly only a matter of time before helium and hydrogen temperatures are as readily available as the temperatures provided by the domestic refrigerator.

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COMM UNI CA TIO NS

Until recently radio communication over long distances generally relied on reflecting radio waves from the ionosphere, whose variation with atmospheric conditions rendered reception rather uncertain. It is now possible to reflect the waves from artificial satellites and thus obtain steady and reliable signals [ 12). In order to reflect well from the satellite, and to carry the maximum information, very high frequencies are used. Hence it is necessary to have receivers which can detect very small signals at these frequencies.

One such receiver is the 'maser' [l3]. We cannot explain here how it works, but note that it must be run at liquid helium temperatures. (The low temperature has the double effect of increasing the intrinsic sensitivity of the device, and of reducing the random thermal motion as discussed for semi-conductors.) The figure shows the G. P. O. receiver for the Telstar satellite at Goonhilly in Cornwall. The maser, cooled by liquid helium, is situated high up on the back of the 'dish' to avoid losses in long runs of cable.

The G. P. 0. Receiving Station for Telstar satellite communication at Goonhilly in Cornwall. The maser is mounted high up behind the 'dish', in a position indicated approximately by the arrow

STUDY OF METALS

Liquid helium and hydrogen are now common facilities in all laboratories which study the solid state. For example one of the standard techniques in the study of metals is to observe the properties of the specimen over a wide temperature range. Often very marked changes are observed [ 14]. To give justtwo examples : the number of cycles permitted before fatigue failure occurs may be hundreds of times greater at low temperatures, and the inter­nal friction (or damping capacity) of many metals often varies strongly as the temperature is decreased. Studies of many experiments of this type help the metallurgist to understand how metals behave under ordinary working conditions.

QUANTUM EFFECTS

We now mention some quite different properties of matter observed at about the tem­perature of liquid helium. These phenomena, unlike the progressive reduction of the thermal motion, set in suddenly below some critical temperature, and are quite unpredict­able from the behaviour at higher temperatures. They are manifestations of the quantum mechanical nature of matter. Perhaps the most striking example concerns the behaviour of liquid helium itself [ 15]. At temperatures near the boiling point liquid helium is not very different from many other liquids. Its rate of flow through a tube follows the ordinary laws of viscosity. We may think of the liquid as a collection of helium atoms each like a rathe-r small billind ball . A� the liquid flows !\'rough a narrow tube the atoms (billiard

6

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balls) collide with the walls and with each other, and with this picture in our minds we can correctly estimate the viscous resistance to flow. Below 2.17° absolute, however, the behaviour becomes quite different. No matter how narrow we make the tube, some helium always flows through very easily. The viscosity appears to have vanished completely. It is as if the atoms no longer collide with the walls of the tube. Yet such behaviour is incom­prehensible if the helium atoms really resemble billiard balls. In fact we can only under­stand those experiments in terms of the concepts of wave mechanics.

Another analogous phenomena is that of superconductivity in metals. As a pure metal is cooled, its electrical resistance steadily decreases. In some metals however, below a cer­tain critical temperature towards the helium range, the electrical resistance vanishes completely. It is as if there were no resistance at all to the flow of electrons. There are as yet no technological applications of these quantum effects in helium, but as we shall hear in the next talk, superconductivity is already of great technological importance.

THE LOWEST TEMPERA TURES

So far as technology is concerned we need at present only consider temperatures down to 1°K which can be reached by evaporating liquid 4He (the common isotope of helium). However an indefinitely large temperature region stretches away to absolute zero. The rare isotope 3He is now commonly used in quantities of about 1 cc. to reach temperatures down to 0.3°K. Magnet methods of cooling will readily take us down to .01 ° or even .001°K, and by certain special techniques temperatures of 10"6°K have been achieved [16]. So far these are used only by the physicists, but it may well be that more widespread appli­cations will be found.

REFERENCES

r. R. B. Scott, Discovery 2r , 74 (r960). 2. R. F. Blanks and K. D. Timmerhaus, Advances in Cryogenic Engineering 5, 3, Plenum Press

(1960). 3. L. H. Cassutt, F. F. Maddocks and W. A. Sawyer, Advances in Cryogenic Engineering, 5, 55,

Plenum Press (1960) . 4. M. A. Dubs and L. I. Dana, Bull. I. I. F. Annexe r96I-5, 7r , (r96r). 5 . H. A. Steinherz and P. A. Redhead, Scientific American 206, No. 3, 78, March r962. 6. E. L. Garwin, Cryogenic Pumping and Space Simulation. To be published in Advances in

Cryogenic Engineering, Volume 8. 7 . D. W. Holder and L. Bernstein, Vacuum Pumping Requirements in Aerodynamic Research,

Proc. Symp. User Experience of Large Scale Industrial Vacuum Plant, Inst. Mech. Engineers, r96r.

8. D. A . Glaser, Scientific American, r92, No. 2, 46, February r955. 9. H. P. Hernandez, Advances in Cryogenic Engineering, 5, 38, Plenum Press (r960).

ro. C. E. Taylor and R. F. Post, Advances in Cryogenic Engineering 5, 3r, Plenum Press (r960). r r . C. 111. Cade, Discovery 22, 432, r96r. 12. ]. R. Pierce, Scientific American, 205, No. 4, 90, October l96r. 13. ]. P. Gordon, Scientific American, 199, No. 6, 42, December 1958. 14. H. M. Rosenberg, Progress in Metal Physics, 7, 339, Pergamon Press (1958). r5. K. R. Atkins, Liquid Helium, Chapter l, C. U. P. (1959). 16. N. Kurti, Low Temperature Physics (Four Lectures), Chapter 2, Pergamon (1952).

N. Kurti, Science Progress, No. r79, July r957.

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Applications of Superconductivity *)

Application de la superconductibilite

Dr. P. G. KLEMENS Westinghouse Research Laboratories, Pittsburgh 35, Pa., U. S.A.

SOMMAIRE : Les deux applications principales de la superconductibilite sont aujourd'hui les elements des calculateurs et les electroaimants. L'economie des aimants superconducteurs, concernant la puisance, est discutee, les proprietes des superconducteurs de la deuzieme espece sont presentees et quelques-uns des problemes, qu'on rencontre dans le developpement des aimants, sont enumeres.

Parmi ceux il y a la bobinage, l' effet de «degradation » et le probleme de la protection contre un changement brusque dans l' etat normal.

Les applications possibles des aimants superconducteurs dans l'avenir sont mentionees.

Although the phenomenon of superconductivity holds the promise of numerous tech­nological applications, so far it has been applied mainly to computer elements and to electromagnets. The property of completely loss-less conduction is, of course, potentially attractive for such uses as power lines and transformers. Mere energy economy is, how­ever, in these cases unlikely to be the deciding factor, because there is a severe refrigera­tion requirement. Whenever there is a small residual energy loss in the superconductor, one requires about 300 times as much energy to remove this dissipated energy from the low to the ambient temperature. It is thus necessary to reduce losses by more than this factor before an advantage in energy economy is obtained. Residual losses are inevitable : the new high-field superconductors, to be described below, have appreciable A. C. losses, even though they are without loss under D. C. conditions, and furthermore there is an irreducible amount of thermal conduction plus Joule heat whereever current is led from the outside into the low- temperature enclosure.

Since a superconductor can be driven into the "normal state", i. e. one with electrical resistivity, by the application of a magnetic field, and can thus be switched from a resistive to a non-resistive state and vice versa, it can be made into a computer element (cryotron). In essence, computers are made up of an array of circuit elements, each of which can exist in either of two states. A pair of superconductors can be in two states : the current flowing in one produces a magnetic field, which drives the other conductor normal, prevents current flow through it and ensures that the magnetic field at the first conductor is below the value needed to destroy superconductivity. The second state is one with the roles reversed. The advantage of superconducting computer elements lies not only in their low power dissipation, though this in itself is very advantageous in a large compu­ter system, where removal of unwanted heat can be a problem, but also in their small size. Speed of operation is the essence of a modern computer, where complicated mathemati­cal operations are built up of many steps in repetition. In the last resort, the maximum speed of response of a computer element to a signal is governed by the time taken by that signal, moving at a speed close to that of light, to traverse one element and move to the next one. Thus small physical size is important.

At the present time the most promising application of superconductivity seems to be that of high-field electromagnets. Since even quite moderate a magnetic field is sufficient to drive a superconductor into the normal state, so that the magnetic field produced by i superconducting coil would choke off the current in this coil, the possibility of super-

*) This paper is a condensation of a more extensive review ,,Superconducting Magnets", Westing­house Scientific Paper 63-1 28-280-P2 (1963) by ]. K. Hulm, B. S. Chandrasekhar and H. Riemersma. Dr. Hulm was prevented from presenting this paper.

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conducting magnets had been disregarded for a long time. However, it was discovered by Kunzler and colleagues at the Bell Laboratories [ 1] in 1961 that in a niobium-tin compound some superconducting properties persist even in very high magnetic fields, in particular that a loss-less current with a density of 1 00,000 amps/cm2 could be maintained in a field of 88 kilogauss. This discovery, and those of similar high-field superconductors found by other workers, makes it possible to construct superconducting magnets.

In a conventional electromagnet, the current through the coils is maintained in the face of considerable resistance, and a great deal of power is dissipated. Not only must this power be supplied at considerable cost, but as it is dissipated in the magnet windings, it must be removed from there by a cooling system, which is not an easy task in high field magnets. In a superconducting magnet no heat is dissipated, and this twin difficulty is thus removed. Against this, we have the requirement of keeping the coil at liquid helium temperature. The power cost of doing so depends only weakly on the magnetic field, but the power requirements of conventional magnets (and even of a liquid-hydrogen cooled aluminium-wound magnet) increase sharply with magnetic field. Thus the superconduct­ing magnet shows superior power economy above about 25 kilogauss, and this relative advantage increases sharply with magnetic field. In some special cases superconducting magnets can be advantageous alternatives even at lower fields.

We recognize nowadays two classes of superconductors, but at first it was thought that all superconductors approached one kind of ideal behaviour, and that deviations there­from were due to inhomogeneities, imperfections and irregularities of shape. An ideal superconductor would expel all magnetic flux, until the external field exceeded some critical value, when it would become normal in its entire volume. Today we know that only in some materials are deviations from this ideal behaviour due exclusively to the above causes ; these are superconductors of the first kind. There is, however, a second class of superconductors, which behave differently : as the external magnetic field is increased, the flux penetrates gradually, and is not distributed uniformly through the material. These superconductors tend to form an intimate mixture of normal and super­conducting regions, with the magnetic flux concentrated in the normal regions. The reason for this behaviour is related to the intrinsic electronic structure, and does not depend on imperfections. The theory of these superconductors of the second kind was developed by Ginzburg and Landau [2] and Abrikosov and Gor'kov [3].

Superconductors of the second kind retain some superconducting regions even at fields well above the field where flux penetration first occurs (this defines the critical field in the case of superconductors of the first kind), and these residual superconducting regions provide non-resistive current paths. The high-field superconductors are all superconduc­tors of the second kind, but of course not all superconductors of the second kind are suitable magnet materials. As was pointed out by Gorter [ 4] it is not sufficient that super­conducting regions exist, they must also be capable of carrying a current. The magnetic field exerts a force on each current element and tends to push it sideways. Unless the superconducting elements are pinned to specific places in the material, this motion will result in energy dissipation. The pinning points are provided by various imperfections. Soft pure metals, even though superconductors of the second kind, are usually unable to carry large currents ; heavily cold-worked materials or alloys can often do so better.

In addition to the Nb3Sn compound of Kunzler et al., a number of other materials have been found to have suitable properties, particularly niobium-zirconium alloys, other alloys of niobium with titanium and hafnium, and other intermetallic compounds of the beta-tungsten group. An intensive search for better materials is being pushed by several groups. At the present time the main magnet materials are Nb3Sn and the niobium-zirconium alloys. The latter have been extensively studied by Hulm and co­workers.

Of course, having a high-field superconductor is only half the answer to the problem of constructing a magnet. One must be able to wind it into a coil, but unfortunately these materials have generally little or no ductility. The ability to carry high current densities is crucial, otherwise the magnet windings must be impossibly thick. It turns out, further­more, that if one tests the current-carrying capacity of a straight-wire specimen in a magnetic field and obtains a good result, one then finds that if the same material is wound

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into a coil, it can carry only a fraction of the current it could before. This degradation effect, and other equally baffling difficulties, blocked progress for a long time. Some of these problems have been solved, but others are still with us.

The fabrication problem is particularly severe for Nb3Sn, which is a brittle intermetallic compound. Kunzler et al. overcame it by making a fine tube of niobium, filling it with niobium and tin, winding it into the coil, and then heating it to form the niobium-tin compound inside the tube. This is effective, but has the disadvantage that a coil can never be unwound for inspection or repair of damage, and that the niobium tube occupies a large volume fraction of the winding.

Niobium-zirconium is slightly inferior in its high-field properties, and is limited to fields of up to 70-100 kilogauss, but it has some ductility. It is far from easy to work with, but with some effort long lengths of wire can be drawn. This material can be handled quite roughly without affecting its superconducting properties, so that coils made from it can be safely subjected to the mechanical shocks of ordinary usage.

The degradation effect, i. e. the decrease in the current carrying capacity of a wire when wound into a coil, proved a serious obstacle and makes magnet design very difficult. For example, a cold-worked straight wire may be able to carry 50 amps in a field of20 kilo­gauss ; the same wire in a coil can carry only 20 amps. If this wire is annealed, a straight section can carry 200 amps, but this improvement is illusory : in a coil the wire now carries only 4 amps ! We now believe that this effect is due to discontinuous rearrangements of the flux lines in the material as the magnetic field is increased (flux jumps). The instantaneous release of heat drives a section of the wire normal, and the resulting Joule heat produced by the current can, under unfavourable circumstances, make the normal region grow, leading to failure. By winding copper wire into the coil, or by copper-cladding the wire, one can provide spare heat capacity and provide alternative paths for the magnet current around the instantaneous defect. One can thus reduce the degradation effect somewhat, but not eliminate it altogether.

As the current and field in a coil is increased and the critical-current threshold is passed, the coil acquires resistance. The magnetic field energy stored in the coil is released in the form of a transient voltage pulse. This can cause damage to the coil due to arcing between the windings, and the creation of permanent shorts. As magnets become larger, this problem becomes increasingly serious, and steps must be taken to slow down the current decay, such as the introduction of shunts or the use of copper-coated wires.

Some of the many problems in magnet construction have now been mentioned : they have been overcome to varying degree. Coils have been built successfully, yielding steady fields of almost 70 kilogauss[5]. With the large current effort, improvements in this figure will come soon. The working space in the present coils is still quite small - of the order of cubic inches, and these magnets will be used mainly in research laboratories, where liquid helium is today widely used for many purposes, so that the refrigeration requirement is not a serious obstacle.

Thus a rather natural application of the new magnet is for the solid state maser, where a cryogenic enviroment already exists. For example, at the Westinghouse Baltimore laboratories a superconducting Helmholtz pair of quite low field strength (8 kilogauss) but high field uniformity was employed to tune a maser. This magnet is not only lighter than the permanent magnet it replaces, but also superior in performance.

Looking some way into the future, it is probable that superconducting magnets will have wide technological application. Numerous potential processes requiring magnetic fields are today impractical or uneconomical because of the large power cost of maintaining the field - power which is wasted and plays no part in the essential energy balance of the process. A superconducting magnet consumes no power. It is thus like a permanent mag­net, but can deliver stronger fields. Applications in nuclear accelerators, energy storage devices, magnetohydrodynamic generators and in other fields are within the bounds of possibility, even if they must await a great deal of further development.

1 1

P-1 1

ACKNOWLEDGMENT :

Many of the developments described here were carried out at the Westinghouse Re­search Laboratories. I am particularly indebted to B. S. Chandrasekhar, J. K. Hulm and H. Riemersma for making available to me much of the information upon which this paper is based.

REFERENCES :

J. ]. E. Kunzler, E. Buehler, F. S. Hsu and ]. H. Werneck, Phys. Rev. Letters 6, 89 (1961).

2. V. L. Ginzburg and L. D. Landau, J. exp. theor. phys. 12, 1064 (1950).

3. A. A. Abrikosov, J. exp. theor. Phys. 32, 1442 (1957) ; L. P. Gor'kov, ibid 36, 1918 (1 959).

4. C. ]. Gorter, Physics Letters 2, 26 (1962).

5. H. Riemersma, ]. K. Hulm, A. ]. Venturino and B. S. Chandrasekhar, J. Appl. Phys. 33, 3499 (1962).

SUMMARY OF THE DISCUSSION (Papers P-3 + P-1 1)

L. Bewilogua, Germany : Mr. Wilks in his lecture did not mention neon, which is of increasing interest not alone for low temperature techniques but also for low tempera­ture physics.

In regard to applications, it seems sufficient to cite the papers of Graham in U.S.A. and the papers of myself. I have shown also, that the cost of neon gas may be relatively modest.

More interesting for the physicist is the behaviour of transport phenomena in liquid neon. We measured the viscosity and the thermal conductivity as a function of tempera­ture along the boiling line. In both cases we observed an anomaly in the region of 27-29° K.

In connection with the observations on hydrogen, reported by Mr. Hammel, may be interesting, how the reduced thermal conductivity depends on the reduced temperature. Whereas with increasing reduced temperature the reduced conductivity rises for helium and hydrogen and diminishes for argon, neon shows an intermediate behaviour. (E. Lochtermann: Cryogenics 3 (1963) 44).

Probably, also in the case of neon there is an influence of quantum parameter.

A. van Itterbeek, Belgium : I agree with Prof. Bewilogua that neon as a cooling agent is very interesting.

We have constructed in Louvain a cryostat which works continuously between room temperature and liquid helium temperatures. In this cryostat temperatures of liquid neon can be obtained. A description of the cryostat will be reported in Cryogenics. The temperatures can be stabilized within one thousandth of a degree.

J. W. L. Kohler, Netherlands : (Comment on a remark of Prof. Bewilogua on the possibilities of liquid neon) : The temperature range between the freezing point and the boiling point is only a few degrees, which makes the use of Neon rather unattractive.

With the advent of hard superconductors, is there still some future for normal, cryogenically cooled magnets ?

R. B. Scott, U.S.A. : At the NBS Cryogenic Laboratory our best working magnet is an aluminium magnet cooled with liquid hydrogen. This produces fields of about 70 k-gauss. Increasing the current (and field) to higher values produces enough Joule heating to overcome the cooling provided by the circulating liquid hydrogen, and as a result the resistance of the coils increases until the magnet effectively turns itself off. We believe that with minor modifications we may reach fields exceeding 100 k-gauss.

N. Kurti, U.K. : I should like to make a few remarks about ,,cryogenic" magnets as distinct from superconducting magnets. The fundamental factors are : 1 ) decrease of

12

P-1 1

resistance with temperature and the proportional decrease in power dissipation; 2) ma­gneto resistance, which becomes important at low temperatures; 3) the efficiency of refrigeration; 4) the relative capital cost of large power supplies (for room temperature magnets) and of refrigerating plants (for cryogenic magnets).

With the greatly improved efficiencies of large scale refrigerators, e. g. 20 % Carnot efficiency at 20° K, one can achieve an important saving in overall power consumption. The main difficulty however is the capital cost of the refrigeration plant, and I think that, at present at least, a room temperature high field coil installation is still cheaper and more advantageous than a low temperature one.

In some cases however, a liquid hydrogen cooled magnet has its advantages. This is the case for instance if one wants to produce occasionally, and for relatively short times, fields in the 50-100 k-gauss range in coils of 5-10 ems. without installing a costly 3 megawatt D. C. power supply and cooling plant.

If liquid hydrogen is available in quantity, one can use it as a coolant, and although the thermodynamic efficiency is small since the enthalpy of the vapour is wasted, the overall cost of the operation may still be favourable. There are at present a number of cryogenic magnets in use in the U.S., notably at Los Alamos, at Boulder, at the NASA Lewis Research Centre in Cleveland, and at Livermore.

J. R. v. Geuns, Netherlands : Might not the Joule and conduction heat losses in the leads be reduced by cooling them at different levels, e. g. 60° K and 15°K with a suitable refrigerator, and using Nb3Sn leads between 15° K and 4° K.

J. Wilks, U.K. : Further details of the use of infra-red detectors in medicine were given in an article by Williams, Cade and Goodwin in the J oumal of the British Institute of Radio Engineers 25, 241 (1963).

13

Pelti er Effect Effet Pelt ier

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G. Lorentzen, Norway : opened the second session of the First Plenary Meeting by stating that the discovery of new methods of producing refrigeration was very rare. As a matter of fact all commonly used principles go back well into the last century and although they have been greatly developed since then and still are, of course, in a state of transition. Recently there has been much talk about the possibilities of thermoelectric refrigeration, the principle of which was discovered by the Frenchman - J. C. Peltier as early as 1834. But it is only as a result of the modern developments in the semi-con­ductor field that technical application on a large scale of Peltier refrigeration and heat pumping has become a reality. Again the refrigeration industry has profited by a break­through of an apparently remote science, in the field of solid state physics. Again the poor refrigerationist has to add a new theory to his already formidable arsenal of fun­damental sciences.

At the time of the Copenhagen Congress, four years ago, there was an immense optimism concerning the possibilities of further development of thermoelectric materials. Curves were plotted, showing figure of merit versus time and it was intimated that within a very few years thermoelectric systems would be competitive and even superior to the conventional vapour compression cycle from a power consumption viewpoint. Time has not, so far, supported this optimistic prediction.

The last couple of years seem to have brought little improvement to the efficiency of materials, the maximum figure of merit stays at about 3 X 10-3 per °K. Progress has taken other courses. For one thing the price of thermoelectric materials has been dra­stically reduced and at the same time the dimensions of the individual couples have become much smaller. Secondly, immense progress has been made in the manufacturing of modules, in soldering, electric insulation and heat transfer. These developments have opened the way to a large and rapidly increasing number of applicativ1l.S and there is no doubt now that thermoelectric refrigeration has come to stay.

This situation makes it necessary to expand the fundamental scienses for the refrig­eration engineer into a field, which is in many ways foreign to his traditional way of thinking. It has also been found appropriate to concentrate one of the plenary sessions on this very important subject, not to prt;!.ie the mysteries of semiconductor theory, but to give an overall picture of the situation from the viewpoint of our industry.

Concepts of Thermoelectric Refrigeration

Concepts du refroidissement thermoelectrique

Prof. E. B. PENROD Department of Mechanical Engineering, University of Illinois, Urbana, Illinois, U.S.A.

SOMMA/RE. Ce rapport f ait rapidement l'historique du froid pour montrer que l'idee de THOMSON d'utiliser une pompe a chaleur etait en avance de 75 ans sur les realisations de la mecanique et de la technique electrique et qu'il s'est ecouli 125 ans avant que les physi­ciens de l' etat solide produisissent des materiaux semi-conducteurs qui rendraient le refroidisse­ment Peltier possible. On indique le refrigerateur Electrolux de MUNTERS et VON PLATEN qui, comme le refrigerateur thermoelectrique, ne presente pas de parties mobiles.

Le thermocouple est defini et plusieurs eff ets qui lui sont associes sont presentes et expli­ques. On accorde une certaine attention a la thiorie et au fonctionnement d'un refrigerateur thermoelectrique consistant en un simple thermocouple. On presente une breve etude du bis­muth tellurite pour montrer qu'il peut etre utilise pour les deux thermoeliments dans la fabri­cation d'un refroidisseur thermoelectrique. On evalue le cout def abrication des thermocouples et des modules.

1 5

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PROGRESS IN REFRIGERATION

Every one in this audience, I am sure, is familiar with the concept of mechanical refrig­eration. First of all, for this discussion, it is appropriate to point out that there are three classes of mechanical refrigerating machines, namely, the air cycle, vapor compression cycle, and absorption cycle machines. It is only natural to raise the question : Are the above systems adequate for meeting current demands for air conditioning and theproduc­tion of cold and ice? For over a century, it has been known that water can be frozen by sending electricity through a thermocouple. It appears logical to ask the question : why hadn't thermoelectric ice machines been manufactured and put on the market before the middle of this century? Perhaps these and other questions will be discussed during this plenary session.

Before presenting the theory of thermoelectric cooling, it may be well to give a very brief account of the history of mechanical refrigeration. Trends in the past may give the engineer clues of what to expect in the future. The first recorded patent for a refrigerating device was granted in 1790. A patent was issued in Great Britain to Jacob Perkins in 1834, and only twenty years later, William Thomson proposed to heat and cool buildings with an air type refrigerating machine. His machine was never built because costly changes had to be made when the system was converted from a heating to a cooling machine. In 1870, however, his machine was built in a slightly different form for space cooling and is known as the Bell-Coleman cold air machine. Thomson's air heat pump for providing year 'round air conditioning was not rejected because of high initial cost or low efficiency.

Vapor compression machines have reached a high state of perfection, and many resi­dential refrigerators of this type have been in operation for over fifteen years without maintenance. The absorption system has wide application for commercial purposes. Heat pump installations of this type provide year 'round air conditioning in large buildings. Carl Munters and Baltzer von Platen developed a household absorption system with no moving parts at the Royal Institute of Technology in Stockholm in 1925. Will the thermo­electric refrigerating system displace conventional equipment since it has no moving parts unless fans are used for providing forced convection? The answer is : Probably not.

A thermocouple may be defined as an electric circuit consisting of two dissimilar conducting materials. In 1821, Thomas Seebeck discovered that an electric current is produced in a thermocouple when its junctions are maintained at different temperatures, Fig. 1. A phenomenon complementary to the Seebeck effect was discovered in 1834 by J ean Peltier. He sent an electric current through a thermocouple and subsequently observ­ed that one junction became warm and the other cold. According to Florian Cajori, the first thermoelectric ice machine was made in 1838 by H. F. Emil Lenz. He placed a drop­let of water in a small recess at the junction of bismuth and antimony rods, and connected the system to a battery so that the current passed through the junction from bismuth to antimony. Shortly thereafter, the droplet turned into ice.When he reversed the direction of the current, the ice turned back into water. The experiment of Lenz is significant. It clearly showed that the junction was cooled by the electric current when it passed from bismuth to antimony ; conversely, when the current passed through the junction from antimony to bismuth, heat was transferred from the junction to the ice, showing that the junction was heated by the current.

In 1889, a patent was issued for a thermoelectric refrigerator. Edmund Altenkirch renewed the interest in thermoelectric refrigeration by publishing two informative papers in 1909 and 1912. A thermoelectric refrigerator constructed of metals and alloys has an exceedingly small cooling capacity and a very low cooling energy ratio.

THE THERMOCOUPLE

When an electric circuit is formed by joining two dissimilar conducting materials and their junctions are kept at different temperatures, five phenomena take place simultane­ously. These phenomena are called the Seebeck effect, Peltier effect, Thomson effect, Joule effect, and the conduction of heat. When an electric current passes through a conducting material the corresponding electrical power developed is dissipated as inter-

16

Co Co

1 -Fe

Jl. N.P. 0 f-1 --�27"'5---+-5�5�0-- t, cc

- E IRON-COPPER THERMOCOUPLE

Fig. 1

P-8

�--______,+ f------� 1 I f--- ------,

Fig. 2

I p

Fig. 3

Fig. r . Iron-copper thermocouple is used for illustrating the Seebeck effect and the complemen· tary phenomenon discovered by Peltier.

Fig. 2. Three p-n thermocouples are connected in electrical series with a battery. When conven­tional current enters the n- arm and leaves the p· arm, there is parallel heat flow from the cold to the hot plates through the thermoelements p and n. If the direction of the current is reversed the direction of parallel heat flow is reversed also, that is the system pumps heat from the upper to the lower plates.

Fig. 3. A thermoelectric or Peltier refrigerator consisting of a single thermocouple in which the thermocouple arms are separated at their ends by suitable metals for conducting heat to or from the system. The total e. m. f. Eab of the thermocouple is given by

Eab = :7l:abh - :7l:abc ± ET where ET is the total Thomson e. m. f. in the two arms. (ET is not shown in the diagram.) The heat power terms are given by the three equations

Qh = Qabh - Qxa = o - Qxb = o'Qc= Qabc -Qxa = L -Qxb = L' and P = Qh-Qc

THERMOELECTRIC HEAT TRANSFER

SYSTEM FOR ICE FREEZING

AMBIENT(32')

Fig. 4. Cross sectional view of one of severa modules in a thermoelectric ice making system. D. C. current enters the p-arms and leaves from the n-arms, thereby pumping heat down· ward through the p-n thermoelements in parallel heat flow. which is translerred to a transport fluid by the fins. Courtesy of the York Corporation. York, PennsylMnia.

nal heat power and every part of the conducting material is heated - this is the Joule effect (Fig. 1, Fig. 2, Fig. 3). It is essential to keep in mind that the Joule heat power appears in all parts of the circuit and is quadratic in current but independent ot its direction.

When a thermocouple operates as a thermoelectric refrigerator, under highly idealized prescribed conditions, the Joule heat power in general does not split equally half way between the junction with one half going to the hot junction and the other half to the

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cold junction. The only condition under which the Joule heat power splits equally with half going to each junction is when the two junctions are at the same temperature. Under these conditions, the system is a trivial refrigerator or Peltier heat pump that transfers heat from one junction at a given temperature to the other junction at the same temperature ; the temperature gradient is zero half way between the junctions ; the temperature in the arm is maximum at this point ; and the temperature profile is a symmetrical parabola with its major axis parallel to the temperature axis. Furthermore, when the Joule heat power is twice the Fourier heat power all of the Joule heat power goes to the cold junction and the maximum temperature in the thermocouple arm occurs at the hot junction.

William Thomson showed theoretically and experimentally that if the ends of a current­carrying conductor were at different temperatures an e. m. f. is produced in it due to the absorption or evolution of heat - this is the Thomson effect. The Thomson e. m. f. is linear in current and depends on its direction and on the direction of the temperature gradient. Heat which passes from higher to lower temperatures in the thermocouple arms is called conduction heat if the transfer is due solely to the temperature difference. In this discussion the Fourier heat power is that which is conducted from the hot to the cold junction due to the total temperature difference, if no other effect is present.

In building thermoelectric refrigerators, it is necessary to insert metallic connectors in the electric circuit to join the p- and n- semiconductor thermoelements as shown in Figs. 2, 3, and 4. Similarly, when a thermocouple is used as a pyrometer for measuring temperatures, it is necessary to insert a galvanometer, voltmeter, or potentiometer in the circuit as shown in Fig. 1 . The addition of connecting plates or links, Fig. 2 and Fig. 4, e. g., involves the presence of more than the two original junctions, and hence it is important to establish the law according to which the e. m. f.'s produced by the additional junctions may be added. According to the law of intermediate metals, the insertion of an additional metal into any circuit does not alter the whole e. m. f. in the circuit, provided that the additional metal is entirely at the temperature of the point of the circuit at which it is inserted. Thus, the total e. m. f. Bab of a thermocouple when connected in the circuits, Figs. 2, 3, and 4 is equal to the algebraic sum of the Peltier e. m. f. 's at the junctions and the total Thomson e. m. f. in the thermocouple arms.

The total e. m. f. of a thermocouple Bab depends on the physical properties of its thermoelements, the temperature of the cold junction, and the temperature difference between the hot and cold junctions. The Seebeck coefficient Sabe depends on the total e. m. f. of the thermocouple since it is the temperature derivative of Bab· For high operating performance, thermoelements having a high figure of merit Z m should be chosen since it depends on the value of the Seebeck coefficient. The following data listed below apply to iron-copper, antimony-bismuth, and bismuth telluride p-n thermocouples

Thermoelements Bab, µv Sabe, µv 0 c-1 Zm, 0 c-1

Fe - Cu 497 13.7 0.005 x 10-a Sb - Bi 5 039 109.4 0.363 x 10-a Bi2 Tea (p-n) 19230 423 2.687 x 10-a

when their hot and cold junctions are at 40° C and -5° C, respectively. The low values of Sabe andZ m for iron-copper and antimony-bismuth thermocouples indicate clearly why 125 years elapsed since Lenz froze a drop of water using an antimony-bismuth couple, until a thermoelectric ice-making machine was manufactured (Fig. 4). Results of research in solid state physics during the last 30 years indicate that pn-type semiconductors of bismuth telluride and lead telluride are the best materials available at the present time for manufacturing thermoelectric refrigerators and thermoelectric generators.

Bismuth telluride has a stoichiometric composition of 52.2 per cent bismuth and 47.8 per cent tellurium. At this composition its Seebeck coefficient is about -218 microvolts per degree C. Increasing the bismuth content to 52.4 per cent results in a compound having a positive Seebeck coefficient of about 205 microvolts per degree C. Thus, with slight change of composition, the same material may be used for both positive and nega­tive semiconductors. Values of the figure of merit of about 2.8 x 1 o-a per degree C have been obtained with Bi2 Tea thermoelements.

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The net cooling capacity Qc of an individual thermocouple varies from about 0.6 to 3 watts (2 to 10 Btu/hr), depending on the design and operating conditions. Therefore, many thermocouples are assembled in units or modules. The following cost for building modu­les consisting of ten thermocouples is an estimate by the Philco Corporation.

Thermoelectric System Cost* 10 - Couple Module for 30 Btu Per Hour Application

Number of couples [Sample 10 000 180 000 4 000 000 Number of modules Basis] 1 000 18 000 400 000 Cost per couple $ 7.00 1 .60 0.60 0.25 Module $ 70.00 16.00 6.00 2.50 Finning 3.00 2.00 2.00 Air Handling (Fan and Motor) 3.00 2.00 2.00 Power supply- controls and wiring 14.00 9.00 9.00

$ 36.00 $ 19.00 $ 15.50

(Equivalent mechanical refrigeration system) $ 25.00)

* Courtesy Philco Corporation

THEORY OF A THERMOELECTRIC REFRIGERATOR

Fig. 3 is a schematic diagram of a thermocouple in a module of a thermoelectric refrigerator whose arms consist of p- and n-type semiconductors. When the module is connected to a D. C. power supply unit, an electric current passes around the circuit from arm "b" through one junction to arm "a", then from arm "a" through the other junction to arm "b", and on back through the power supply unit. The electromotive force of the supply unit must be greater than Bab, the total e. m. f. of the thermocouple. As mentioned above, when an electric current passes around a circuit consisting of two dissim­ilar materials, one junction becomes cold and the other hot, thereby creating a difference in temperature between the two junctions. The power supplied to the couple is repre-sented by P, while Qabh is the Peltier heat power developed at the hot junction and Qabc is the Peltier cooling power produced at the cold junctions.

The net refrigeration Qc is the time rate at which heat is transferred from the space to be cooled to the cold junction. It is equal to Qabc minus the heat power that is conducted through the thermocouple arms to the cold junction. The cooling energy ratio CER is

equal to the quotient obtained by dividing Qc by P, and is a measure of the efficiency of the thermocouple.

Qh is the heat power conducted from the hot junction to the ambient. It is equal to Qabh minus the heat power conducted from the hot junction through the thermocouple arms when f3 is equal to or less than 2, where f3 is the ratio of the Joule heat power to the

Fourier heat power. If f3 is greater than 2, Qh is equal to Qabh plus the heat power conduc­ted to the hot junction from the thermocouple arms. The calculated results listed below illustrate the point in question. These results are for : Zm = 1.932 x 10-3 ° C-1, Aa = Ab = 1 cm•, La = Lb = 2 cm, th =5°C and tc = 0° C. If the system is used for space heating, the system is a Peltier heat pump, and the heating energy ratio HER is equal to the

quotient obtained by dividing Qh by P.

Current Qx - o)a Qx = o)b Qc Amperes /3a /3b watts watts watts CER

50.00 18.95 42.52 -0.403 -0.939 1.203 0.412 16.24 2.00 4.50 0.000 -0.058 0.642 2.013 10.83 0.89 2.00 +0.026 0.000 0.431 2.958

4.00 0.12 0.27 +0.045 +0.040 0.1 16 5.123

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The above results show clearly that as the supply current is decreased from 50 to 4 amperes the ratio of the Joule heat power to the Fourier heat power f3 decreases in both

arms, and the net refrigerating capacity Qc decreases also ; however, the cooling energy ra­tio increases with decrease in current.

':: v -,� r--- CARNOT CER v�/' v MA•O•UM CER -ACTVAL

Fig. 5. Variation ot supply current I and cooling energy ratio CER versus difference in tempera­ture Lit between the hot and cold junctions. I' 0is the supply current that produces maximum

refrigeration Qcm for all Lit, and 10" is the supply current that produces maximum CERm for all Lit. The graphical results are for material properties yielding a figure of merit Zm of r.932 x 10·• °C·1 for the case where the Thomson e. m. f. is negligibly small and there is no heat leakage through the lateral surfaces of the thermocouple arms. The cold junction temperature is zero.

The variation of supply current I and the thermoelectric cooling energy ratio CER versus the temperature difference between the hot and cold junctions is shown graphi­cally in Fig. 5 for a geometry that gives the optimum cross-sectional area ratio. Io" is the value of the supply current that gives maxima thermoelectric CER for the couple. If the system operates with a current lo' = 61.69 amperes and a Li t = 5°C, a maximum cool-

ing capacity Qcm of 1 .568 watts obtains for a corresponding CER of only 0.457. On the other hand, if it operates with a current Io" = 4.742 amperes and a Lit of 5°C, a maxi­mum CERm of 5.330 obtains for a corresponding cooling capacity of only 0.132 watts.

The performance of a thermoelectric refrigerator is shown graphically for maintaining maximum CER as Li t is increased, Fig. 6. Maximum cooling of 0.463 watts obtains for a LI t of about 33 ° C, and the corresponding values of the heat power that is conducted to

Fig. 6. Effect produced on the performance of a thermoelectric refrigerator consisting of a single thermocouple. The conditions are the same as in Fig. 5. The cooling capacity is nearly constant as Lit varies from 20 to 45°C.

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the hot and cold junctions is 0.368 and 1 .174 watts, respectively. Both the power supplied and the heat power transferred to the ambient increase rapidly with increase in LI t. Increasing the power supplied to the system above 1.6 watts results in a decrease in the net cooling capacity and an increase in the heat power that must be transferred from the hot junction to the ambient.

APPLICATION

In selecting a refrigerating system for a given application, a decision should not be based on efficiency or initial cost only. Thermodynamics teaches us that a reversible refriger­ating cycle machine is more efficient than an irreversible cycle system. Nevertheless, the irreversible system is the one used in practice because it is more economical. Thermo­electric refrigerators using thermoelements having a figure of merit of about 2.8 x 10-30 c-1 will have an efficiency lower than conventional systems oflarge capacity. For low capacity requirements, the efficiency of the thermoelectric system will probably equal that of the absorption cycle and compare favorably with the vapor compression type. It will also occupy less space than the conventional systems.

Thermoelectric systems can be made very compact by using very short thermoelements, which in turn will reduce the thermocouple cooling energy ratio and increase the oper­ating cost. In reading Paper No. 3-24 on the Performance Characteristic of a Peltier Refrigerator at the Tenth International Congress of Refrigeration at Copenhagen, I made the following statement which applies to an antimony-bismuth thermocouple : *"How­ever, when I and Lit are 100 amp and 4.95° C respectively, the net refrigeration, 1 .2 watts, produced by the thermocouple 2 cm long is greater than 0.69 watts produced by a ther­mocouple 1 .0 cm long".

lHUMOt:UctllC Al• CONOIUONU

CUTAWAY VllW (COlO SIDI)

Fig. 7. Schematic shows the general configuration and conditioned air circulation for a pre­engineering model air to air thermoelectric air conditioner. This air conditioner was deve­loped for the purpose of evaluating the use of thermoelectric cooling and heating for U.S. Army Military vans and shelters. Commander USAERDL, Fort Belvoir, Virginia. Photo­graph used with permission.

Fig. 7 is a photograph of a compact thermoelectric air conditioner. It consists of four modules, each of which is composed of 100 thermocouples. Thus, the capacity can be accurately matched to requirements, a distinct advantage when compared to large in­crementation of vapor compression or absorption systems. The system is easily changed from a cooling to a heating device by changing the direction of the supply current. For marine application there is a reduction in volume and weight of the thermoelectric air

* Theoretical Analysis and Performance Characteristics of a Peltier Refrigerator by E. B. Penrod, Institut International Du Froid Bulletin, Vol. XL, No. 2, April, 1960.

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conditioner over the vapor compression type of about 65 and 35 per cent, respectively, When the ship is laid up for an appreciable length of time all that need be done is to disconnect the unit from the source of electricity by opening a switch.

Many believe that there will be a big market for the thermoelectric ice-maker and air conditioner in the low capacity bracket. Each of these should have wide acceptance in hospitals, clinics, pharmaceutical plants, and research laboratories.

As we all know considerable progress has been made in recent years in developing thermoelectric devices, refrigerators, ice-makers, and air conditioners. This is the result of research, experimentation, and intense interest in this very timely subject.

22

The Practical Use of Thermoelectric Refrigeration

Utilisation pratique du refroidissement thermoelectrique

THORE M. ELFVING San Mateo, Calif., U.S.A.

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SOMMA/RE. Meme avec le facteur Z actuel de 3 seulement et en ne tenant pas compte des possibilites de mise au point de meilleurs materiaux de base, les perspectives du refroidisse­ment thermoelectrique sont prometteuses.

Les caracteristiques originales des pompes a chaleur thermoelectriques permettent dans de nombreux cas d' ameliorer le coefficient global de rendement par des moyens dont on ne dispose pas avec les procedes classiques.

Le cout du materiau semiconducteur n'est plus un facteur decisij. L'atout des dispositifs thermoelectriques peut etre un bas prix, le facteur limitatif: leur rendement.

En utilisant des groupes autonomes construits avec des thermocouples a branches courtes et avec des chutes de temperatures pratiquement eliminees par des contacts soudes sur toute la chaine de transmission de chaleur, on peut construire des dispositifs pratiques d'un rendement voisin du maximum theorique.

On concentre actuellement de grands efforts sur le perf ectionnement de group es autonomes normalises de ce type pour la fabrication en serie. Combines avec des dispositifs de trans­mission de chaleur pour un systeme en cascade commode et un refroidissement d' air sans ven­tilateur, on pense que des refrigerateurs a deux temperatures rentables et d'autres dispositifs menagers ou applicables dans d'autres domaines seront mis au point d'ici peu. Pour une jaib/e puissance et des applications speciales au domaine e/ectronique OU a d'autres domaines industriels, /es possibilites du marche sont illimitees. Pour les grandes puissances, la simplicite des pompes a chaleur thermoelectriques les feront preferer partout OU le cout de l'energie electrique joue un r/Jle secondaire. La ou le probleme de l' encombrement et autres caracteristi­ques seduisantes, tel/es que I' exploitation sans entretien, le silence, etc. jouent un r/Jle decisif, le refroidissement thermoelectrique remplacera sans aucun doute le materiel existant dans une plus grande mesure qu'il n'a ete prevu jusqu'ici.

The technical-economical evaluation of thermoelectric refrigeration has been largely dominated by the discussion about the so called "factor of merit", the T-factor. This important material constant determines the efficiency and the limits of the thermo­electric heat pumps and has, for good reasons, been looked upon as a quantity represen­tative of the development and progress of thermoelectric refrigeration in relation to our conventional refrigeration processes, absorption and vapor compression systems.

When we met four years ago in Copenhagen we were told about the rapid progress made with regard to the Z-factor and although the figure obtained at that date, in Au­gust 1959, Z = 3 X 10- • l/°C, was barely comparable with the smallest absorption units and far behind the smallest compression units, it was generally believed that the progress would continue and that higher Z-factors of say, 5-10, would be reached in a matter of a few years. In other words, the efficiency of thermoelectric refrigeration was not in doubt, the optimism in this respect was general. As to the cost of thermoelectric refrigeration there were different opinions. Figures from Europe and USA regarding the cost of the basic semiconducting materials indicated rather prohibitive first costs, our Russian friends were more optimistic.

Fig. 1 illustrates what actually has happened to the Z-factor since 1959, judging from what is publicly known and available on the market. Instead of the expected prog­ress the Z-factor has remained at a value of 3 and in the last period of time we have even downgraded the effective combined Z-factor for commercially available thermo­couple material to 2.8-2.9. Investigations have confirmed that there is a definite ceiling

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YEAR

Fig. r. Improvement of the Z-factor during the last decade.

on the theoretically obtainable Z-factors for the bismuth telluride system and a sub­stantially higher Z-factor than 3 does not seem within reach with this system. The hope that we will soon reach higher Z-values is therefore fading.

Great progress has been made regarding manufacturing processes of u�iform semi­conducting materials and the assembling of thermocouple arrays and sealed package modules for use in practical application. Bismuth telluride is still the system most widely used. The highest Z-values have been obtained with crystalline materials although sin­tered materials are being rapidly improved. Sintered materials are simpler to manufac­ture, are stronger and lend themselves better to cutting in small lengths. One manufac­turer uses sintered P-type material in "hybrid" thermocouple arrays where the N-type elements are still made from crystals. The life length of thermoelectric materials is high. The degradation over a period of time seems to be insignificant although some reports indicate that repeated reversal from heating to cooling does reduce the heat pumping efficiency.

The price of the best semiconducting material in rods of various diameters is at pres­ent U.S. $ 45.00 per pound. Material can be bought in cut lengths and at additional cost also with soldered ends for assembly in thermocouple arrays.

The amount of semiconducting material per watt heat pumping capacity increases with the square of the element length. The first cost for thermoelectric heat pump arrays can therefore be greatly reduced by using thermocouples with short leg length. This leads to comparatively low-cost high capacity modules with high maximum input cur­rents. There are several limiting factors for the current development in this direction. The mechanical strength of the crystalline material is weak and it is difficult to cut in smaller length than, say, 2,5 - 3 mm without considerable loss of material at the cutting process. A high input current means high current density with risk for development of heat at the soldered joints where the electric resistance is a sensitive factor. A suitable mode of operation can eliminate this risk. Short-legged high capacity thermocouples need a larger spread by the use of larger junction plates in order to reduce heat transfer problems in connection with the increased heat load and this also increases the direct heat losses through the insulation between the hot and cold junction side of the modules. The technique of building modules and packages of short thermocouple is of increasing importance and it is here that the major improvements of thermoelectric refrigeration applications has taken place in the last period of time. The use of a leg length of only Ys " or approx. 3 mm is now standard. With a rod diameter of 7 mm and with the price mentioned above this means a net worth of semiconducting material per 100 watts of refrigeration in air cooled devices (Ll t = 40°C) of approx. U. S. $ 25.00. With still shorter leg length this figure can be considerably reduced and is, of course, basically lower for water cooled applications. We are, therefore, rapidly approaching a situation where the cost of the thermoelectric semiconducting material no longer is a decisive factor. Thus, contrary to the predictions only a few years ago, the competitive strength of thermoelectric refrigeration may be the low first costs, the limiting factor its low effi­ciency.

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EFFICIENCY CONSIDERATIONS

In an air cooled application where heat is pumped from a level of -15°C to say + 45°C, which means a temperature difference or Lit of 60°C, a thermoelectric heat pump based on material with a Z-factor of 3 X 10-3 1 /C has a theoretical max. effi­ciency equal to the efficiency of the small hermetic absorption units. Under the same conditions it would take a Z-factor of 10 to match the smallest hermetic compressors and a Z-factor of not less than 20 to match larger compressors with a capacity of 1000 watts. However, this refers only to a Lit of 60°C. As seen from Fig. 2 the efficiency

�01 " TH-TO U06 l•-IO°c • 263°1<:

05 Cl�•Cornol elf. E-,,,'. 006 (,c

oo t

Fig. 2 . Comparative efficiency curves a t varying LI t.

of thermoelectric heat pumps is highly dependent on the Lit, much more so than con­ventional systems, and at lower LI t' s the thermoelectric heat pump is decidedly more efficient than absorption units. This sensitiveness for temperature differences combined with the flexible nature of thermoelectric heat pumps makes it possible in many impor­tant cases to improve the overall coefficient of performance by means which are not available at conventional processes. A thermoelectric two-temperature application can, with no extra cost, be operated at the theoretical maximum efficiency by using separate thermocouple assemblies for each temperature. Conventional processes operating at two different temperatures would for maximum efficiency require a costly duplication of the equipment as, with a single unit, all the refrigeration capacity would be delivered with an efficiency or coefficient of performance corresponding to the lowest temperature.

In terms of an air cooled household refrigerator operating with a freezer compartment at-15° C and a refrigerator compartment at + 5° C where the relative refrigeration capac­ity of the two compartments is 1 : 2, a thermoelectric heat pump can operate in two stages with an overall efficiency theoretically twice as good as the absorption unit and more than half as good as the small compressor unit.

Continuous cooling of liquids in pipes can be carried out stepwise in several sections, each cooled by a separate thermoelectric heat pump operating at max. COP with a successively mounting Lit. The resulting overall coefficient of performance can even for air cooled devices of this type approach the value of 1 . For larger water cooled air conditioning and refrigeration devices coefficients of performance greater than 1 and for certain ship applications even greater than 2 are possible. Therefore, even with a Z-factor of only three and discounting the possibilities of developing better materials for our thermocouples the theoretical outlook for thermoelectric refrigeration is not discouraging.

CHARACTERISTIC FEATURES

For the correct design and operation of thermoelectric heat pump devices there are several characteristic features to be kept in mind. The relatively low efficiency and the sensitiveness for the temperature difference between the levels of heat pumping, the LI t,

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have already been mentioned. Another outstanding quality is, of course, the indepen­dence of the efficiency on the capacity. Different from absorption and vapor compression unit there is no lower limit for the good functioning of the electronic heat pump.

The convenient way in which thermoelectric heat pumps can be arranged in stages or in so called cascades is another important feature of thermoelectric refrigeration dif­ferent from conventional processes. Cascades ordinarily only slightly improve the effi­ciency over a single stage, since the increase in coefficient of performance even for per­fectly balanced cascades using the most favorable temperature split is seldom more than 10%. However, cascades are important for many special applications and offer simple solutions to many design problems, especially for two-temperature applications like household refrigerators with a freezer compartment.

AIR COOLED THERMOELECTRIC DEVICES

Mode of Operation The basically low efficiency and the sensitiveness for L1 t makes it mandatory that air

cooled thermoelectric heat pumps are operated with maximum coefficient of performance during maximum design conditions. A deviation from max. coefficient of performance increases the necessary energy input which in turn increases the L1 t for given heat dissi­pating surfaces, which again lowers the efficiency. In other words pressing the heat pump by a higher input current than corresponding to max. efficiency has a snowballing effect and very little is gained thereby. This is especially the case when air cooling with­out forced air circulation is used. Operation at maximum efficiency means minimum power supply and heat dissipating surfaces. By choosing an appropriate couple geometry the semiconducting material can be fully utilized. For air cooled applications maximum efficiency, therefore, also leads to minimum manufacturing costs.

CONTROLS

Operation at maximum efficiency calls for intermittent control of the temperature. Interruption of the current leads to an equalization of the temperature between the hot and cold junction side but losses can be avoided by using a one-way heat transfer system for cooling the refrigerated space instead of fin radiators directly on the cold junctions. Temperature control by modulation is complicated and still rather costly. Stepwise ener­gization is possible but is somewhat more complicated than intermittent operation and less economical.

HEAT DISSIPATION

Air cooling of thermoelectric refrigerators is not an easy problem. With a heat load of 1 watt per cm2 of hot junction plate area, which corresponds to average optimum con­ditions, we find that air cooling with a heat transmission coefficient corresponding to natural convection (no fan) would require a hundred times larger surface to keep the temperature difference down to 10°C. Enormous fins and forced air circulation is one solution to this problem but a less objectionable solution is, no doubt, the use of boiling and condensing media for higher heat transfer rates at the limited surfaces. A series of experiments have shown that double-walled panels of the "rollbond" type provide ideal hermetic heat transfer systems for this purpose. Special rollbond patterns provide a multitude of passages for gas and liquid with a minimum of pressure drop and the total temperature drop in a secondary heat transfer system of this type can in practical appli­cations be limited to 1 - 2° C.

HEAT TRANSFER CHAIN, TEMPERATURE DROPS

The high heat load on the hot junction side of shortlegged high capacity thermo­couple arrays creates problems in connection with the electric insulation which has to be included in the sealed moisture proof package units used in applications such as refrig­erators, ice-freezers, water coolers etc. The electric insulation between the junction

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plates and the cover plates in a package means a thermal resistance both in the insulation itself and in the interface region on both sides of the insulation. The hitherto most com­mon solutions to the electric insulation problem have been the use of lacquer films serv­ing as bonding agents or the use of thin films of mica, mylar, epoxy etc. usually applied together with greases or oils (silicone) in small amounts for better thermal contact. Insulation of this kind usually requires bolting between the hot and cold junction side causing extra heat losses. Anodizing or hard coating is another simple and relatively inexpensive process for providing electric insulation. The drawbacks common for all the above mentioned solutions are excessive built-in temperature drops and usually also vulnerability for pressure unless applied between very carefully machined and smooth surfaces which is costly. With a beat load of 1 watt per cm • of junction plate area the inside temperature drops usually amount to 5° C .

CERAMIC PACKAGE UNITS

A promising and more rational way of solving the problems of a thermoelectric mois­ture proof package unit suitable for household refrigerator applications as well as larger installations seems to be the use of ceramic wafer materials as cover plates for thermo­couple modules. Such wafers made from Al208 or alumina are perfect electric insulators, have a high heat conductivity in class with the stainless steel group and can be sealed to form a completely vapor and moisture tight enclosure around the thermocouples. With a heat load of 1 watt/cm • junction area the temperature drop in a 0,5 mm thick wafer of this type amounts to only 0,3 C. Interface temperature drops can be almost completely eliminated by metallizing the wafer in a pattern corresponding to the junction plates and soldering. The metallized package unit can also be soldered to heat transferring bodies to eliminate interface temperature drops outside the package. In this way it seems pos­sible to build thermoelectrically refrigerated devices without significant temperature drops and combined with heat transferring panels using boiling and condensing media we can expect a performance of the practical application very close to the theoretical performance indicated by the Z-factor .The industry is rapidly learning how to use these ans similar application techniques and it can be expected that before long standardiza­tion on high capacity low cost packages of the type indicated will take place. Only then will a correct evaluation of thermoelectric refrigeration be possible and chances are that it will find a much greater use than we at present are able to anticipate.

APPLICATIONS

With the above viewpoints and possibilities in mind let us look at a few typical appli­cations of thermoelectric refrigeration.

A great variety of thermoelectric prototypes for household refrigerators and similar applications as well as different kind of gadgets like baby bottle coolers etc. have over the last few years been shown to the public. Some of these were already shown at our meet­ing in Copenhagen four years ago. All of these devices had in common that they were either not available for sale at all or, if actually sold, the prices were exorbitant. Time has now come to disregard objects of this kind, which were shown in order to prove that thermoelectric refrigeration can be used and instead concentrate on devices which are and will be economically used.

Fig. 3. Bonded thermoelectric package unit with fin radiator on hot junction side.

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First let us look at the basic thermoelectric tools used in applications of various kinds : Fig. 3 shows a package unit in which several modules using a leg length of Ys • are

included. The modules are electrically insulated from the cover plates in both sides by a film causing inside temperature drops.

The bolted package in Fig. 4 is composed of 48 thermocouples and provided not only with cooling fins on the hot junction side but also with a fan for forced air circula­tion and effective heat removal. Here anodizing is used as electric insulation.

Fig. 4. Bolted package unit with fin radiator and fan.

In Fig. 5 is shown an expanded view of an entirely different type of package, namely, a ceramic wafer package. Here the very short semiconducting cores of the thermo­couples are joined in series by relatively large copper junction plates, which are mounted with a larger spread than in ordinary modules. On each side of the junction plates are soldered ceramic wafers which are metallized in a certain way to match the pattern of the junction plates. The package is closed by a frame to form a moisture proof or even vacuum tight package. By metallizing also the outside of the ceramic wafers the package can be soldered to heat transfering members to avoid temperature drops.

Starting with applications illustrating the remarkable independency on capacity we see in Fig. 6 a three-stage cascade unit capable of maintaining -65° C in a vacuum with a heat load of 15 milliwatts and a heat sink temperature of +27°C. This device is

,.

Fig. 5. Expanded view of all soldered package unit using metallized ceramic wafers as electric insulation.

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designed for an infrared detector but can also be used for other electronic devices. The unit is furnished with a metal enclosure containing a sapphire window. Input energy 1 watt.

Fig. 7 shows another Infrared Detector Cooler in complete mounting with cooling fins and fan. The thermoelectric section is a four stage cascade capable of cooling a 15 milliwatt load to -78°C. Entire unit occupies 20 cubic inches, weight 25 ounces. Energy input 12 watts.

6 7

Fig. 6. Three stage cascade unit for infrared eye. Reaching -65°C in vacuum with I watt energy input.

Fig. 7. Infrared Detector Cooler with 4-stage cascade ; -78°C in vacuum with 12 watts input.

There are a great number of so called spot coolers for electronic components, diodes, transistors, etc. in the market. Such spot coolers are designed for maintaining tempera­ture from -80 to + 50 depending upon the use and span over a capacity range from a few milliwatt to 50 watts and above.

There are also all kinds of thermoelectric devices for controlled temperatures, such as, standard temperature reference cells, Beaker Coolers, Environmental Chambers, Biolog­ical Controlled Temperature Chambers etc. Other applications are dehumidifiers, Dew Point Hygrometers, etc.

Comparatively little has hitherto been done in developing thermoelectric devices in the household application field. I am thinking of ordinary two-temperature refrigerators and also of ice freezers and water coolers for offices etc. This is due to several factors ; price and economy competition from conventional refrigeration system, absorption units and hermetic compressors, is the main reason. Also lack of suitable thermoelectric basic components such as moisture proof packages has delayed development in this field to which comes the need for a small low-cost power supply for DC current which not until now is being developed for this new market.

Fig. 8 shows a schematic view of a thermoelectric refrigerator using a package unit comprising three 8-couple modules, Ys • leg length, cooled by a fin radiator and fan. This unit is being installed in large numbers in hotel rooms, sick rooms, etc. in USA. Electric consumption 150 watts of which 100 watts is DC energy input, the rest is losses in the power supply.

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i.•.,. ''" IU!C' t

-----,,,.,

Fig. 8. Schematic view of heat pump arrangement commonly used in ice-freezers for hotel rooms etc. Air cooling by large fins and forced air.

Fig. 9 shows a schematic view of a similar refrigerator ice-freezer when the air cooling is solved by heat dissipating panels partly filled with a refrigerant instead of forced air circulation. The compactness of thermoelectric heat pumps, no moving parts, silent operation and elimination of maintenance are features which are well preserved

Fig. 9. Air cooled device for ice-freezing etc. enclosed in a lining of double-walled freon panels serving as heat dissipating surfaces. No fans.

in this type of devices. Operation of 32 Ys • couples at maximum coefficient of performance reduces energy input to 50 watts.

As to household and other refrigerators there are still only one-temperature designs on the market. As already mentioned conventional household refrigerators are generally two-temperature applications with a freezer compartment or at least ice freezing facilities beside the food storage space above freezing.

Fig. 10 shows a typical small thermoelectric refrigerator with an input DC energy of 54 watts. It can be attached either to a 12 volt battery or a 1 15 volt AC line. The unit uses only 18 couples, is cooled by forced air circulation and has a built-in power supply in one model, the inside lining serves as radiator.

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Fig. 1 0. Small one temperature refrigerator with built-in power supply and fan for air cooling. AC input II5 watts.

Fig. 1 1 shows a 3 cu. ft. refrigerator with two temperatures. Ice freezing or a freezer compartment is arranged at the bottom of the cabinet. This is achieved by using a two stage heat pump of which the main heat pump is cooling both the food compartment

II r2

Fig. rr. Air cooled 3 cu. ft. two stage refrigerator with ice freezing using freon panels but no fan. Fig. 1 2 . Schematic view of two stage refrigerator with freezer compartment using hermetic heat

transfer panels for cooling and heat dissipating.

and the hot junction side of the low temperature stage over a hermetic double-walled panel which serves both as a radiator and as a thermal connection between the two stages. The schematic outlay of such two-temperature refrigerator is shown in Fig. 12, where the freezer compartment is located below the heat transferring radiator panel. In both cases the final heat dissipation takes place by a boiling refrigerant enclosed in a similar heat transferring panel which can be given very large dimensions, say 100 times larger than the net area of the hot junction plates in the final stage.

So far we have dealt with air cooled applications. For water cooled applications with small LI t' s and high coefficients of performance there is hardly any limit as to the capac­ity of thermoelectric devices.

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Fig. 13 shows a water cooled thermoelectric air conditioner for commercial buildings. The capacity of this unit is 8 / 4 tons of refrigeration or 2600 watts. A number of thermoelec­tric thermocouple arrays or packages, water cooled on the hot junction side and provided with a fin radiator on the cold junction side are arranged in a battery through which air is pulled by fan equipment located above the assembly. (In the picture the front cover and fans are removed.) This design is a basic air conditioner on which further designs are based. Capacities up to 35 000 watts capacity are now being installed.

Fig. 13. Water cooled thermoelectric air conditioner for commercial buildings. Refrigeration capacity up to 35 ooo watts.

Fig. 14 shows a module used in a water to water system sponsored by the U. S. Bureau of Ships. The heat pumping system is designed to produce 3500 watts of refrig­eration per cubic foot of space by stacking these modules together. The thermocouple array is placed between bolted cover plates in contact with water pipes for heat absorp­tion (cooling) with heat dissipation on the hot junction side through Cu-Ni tubing for sea water. The junction plates are in later types insulated from the cover plates by alu­minum oxide sprayed directly onto the heat exchanger surface.

Fig. 14. Water to water thermoelectric heat pump for use on board ships.

In Fig. 15 is shown a close-up of water cooled thermoelectric air conditioner for the direct cooling, dehumidifying and heating of spaces occupied by people. The unit is designed for the U. S. Navy for mounting in a conventional air supply duct serving the living or working quarters on a nuclear powered submarine or other vessel. The unit, composed of 6 identical modules, has a capacity of 3500 watts, energy input is 1500 watts which indicates a coefficient of performance well over 2. Some 3,120 thermocouples are. used. Also here sprayed aluminum oxide is used on top of the copper straps. By spraying a layer of copper over the aluminum oxide the modules can be soldered directly to the heat exchangers.

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Fig. 15. Water cooled thermoelectric air conditioner for duct mounting.

Finally, the world's largest thermoelectric refrigeration system built so far is shown in Fig. 16. The unit is designed for installation aboard an atomic submarine under a contract from the U. S. Bureau of Ships and provides food chilling and freezing for crews under long periods of submersion. It has capacity for preserving more than 6 tons of food and comprises one smaller thermoelectric unit for chilling and a larger unit for freezing shown in the picture. The freezing unit contains 360 modules, each with 24 thermocouples and pumps a refrigeration effect of 2800 watts at an air temperature of -0° F (-18°C) to cooling water of 55° F or 13° C from the ships air conditioning system at a coefficient of performance of 0.31. The LI t under these conditions should amount to approximately 45° C and the relatively low efficiency indicates that certain temperature drops are present. The electrical insulation in the modules is hard coat anodizing on top of an aluminum plate, which is glued to the thermoelectric module by a mixture of Epoxy and Beryllium Oxide. This package is then in pressure contact

Fig. 16. World's largest thermoelectric freezer unit for frozen food storage on board. Refrigeration capacity 2 800 watts at -r8°C storage temperature.

with the heat exchange surfaces using silicon grease for better heat transfer. The figure shows the freezer unit consisting of 10 sections with cooling water inlets and DC current supply from a control central. Air is circulated through aluminum fin radiators or heat exchangers on the cold junction side as the unit is secured to the outer wall of the freezer compartment. A similar unit with only 4 sections chills the food compartment above freezing.

As already indicated, the companies working in the thermoelectric field were until about a year ago fairly optimistic regarding the possibilities of improving the fundamen­tal materials, the Z-factor. As long as this optimism prevailed more emphasis was put on that phase of the development than on the actual construction of competitive refrig­eration devices based on thermoelectricity. The gradual understanding that a further

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improvement of the Z-factor may take a long time has led to a reevaluation of the pos­sibilities of designing competitive refrigeration devices based on available materials. Great efforts are therefore finally being concentrated on the problems of perfecting the thermoelectric heat pumps, which, if working at their theoretical maximum efficiency, are more competitive than we first thought. The simplicity and maybe also the low cost of electronic heat pumps will give them preference wherever the cost of electric energy plays a secondary role. Thermoelectric refrigeration will also be the solution where space requirements are decisive and for small capacity applications the field is unlimited. What the refrigeration industry is waiting for today is an efficient, loss free, moisture proof and simple thermoelectric package unit which can be standardized for mass-manu­facture at a price not significantly higher than the price of the semiconducting material included. Seemingly nothing stands in the way of modern technique for accomplishing this relatively simple achievement; it is a fair assumption that the marketing of a per­fected low cost thermoelectric heat pump is only a matter of time, maybe months. It is probable that this refrigeration tool will be put on the market independently at several places, if it will be first here in Europe, in U. S. A. or in Japan, nobody can tell. But that it will come is sure and then, only then, can we expect the real break-through of ther­moelectric refrigeration.

ACKNOWLEDGMENT

The author is very grateful to the many individuals and companies who contributed pictures and data for the thermoelectric applications shown in the paper.

SUMMARY OF THE DISCUSSION (Papers P-8 + P-7)

A. B. Newton, U.S.A. : opened the discussion by firstly stating that he found the combination of the first paper which was a theoretical approach with the second giving the practical application most interesting. He did however want to raise a number of points which he thought might require amplification. The first point was in connection with the degradation of materials which Mr. Elfving had mentioned in his paper due to the great number of reversals which may occur. Mr. Newton stated that they now knew how to prevent this matter and also other people were familiar with the methods of preventing this. To test the material reversals have been applied and thereby alternat­ing temperature differences of 100 degrees every 3 minutes. Thermo-couples could now be made by them and also bought from others which show no measurable degradation after 60.000 to 100.000 cycles. When you consider that this is a matter of many hours of operation we therefore do not feel that this is a problem any more. The second point which he wished to raise was that one must always consider the ,,whole system". The power system is very closely tied up with the characteristics of the materials but it had been found necessary to create computor programmes on which one could optimize both the system itself and also the power system which supplied it. This, of course, applies more to the larger applications as they usually operate at partial load at much of the time. At the moment commercial air conditioning systems are installed which give capacities of 15 to 20 tons and which have been selected for installation on a purely commercial basis. They had been selected as thermoelectric plants because of the control which can be obtained from this type of system and which other systems cannot do and the precise control of individual controls in the system which other systems cannot give, such as the precise control of temperatures and rooms and conditioned spaces. It is therefore most important to realize that one is not only interested in the full load per­formance but very much in the partial load performance of such systems.

He felt that one must display the information in a way which will make it easy to apply thermoelectric principles in the field. A chart had therefore been devised with a four quadrant approach which can be understood by people in the field who have to apply these systems without knowing the basic mathematics underlying them. These charts are in fact a plot of heat rejected or received by the couple against temperature difference, i. e., Lit and show the current and C.O.P. curves.

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T

JV

-At ti

When the Li t is positive the heat rejection side is warmer than the heat receiving side; when it is negative the opposite occurs. In quadrant one which is the one which is usually published we have the heat received by the couple for a variety of currents and there is a maximum current which is normally designed for the couple. This, of course, will depend on the cross section of the couple. For application we like the C.O.P. curves because a man applying it can see how broad the curve is. The man has to know this in order to judge the economics of his system. In the next quadrant which is called quadrant IV, the same lines extend in that the Lit is now negative. They now represent the conditions which we get in air conditioning where we reject heat at a lower temperat­ure than the temperature at which we pick it up, i. e., when there is a low temperature out of doors we still require air conditioning indoors because of load. The C.O.P. lines are fairly high but it is not unusual to operate at 20 C.O.P. even at average outdoor temperatures. We then come to the zero voltage line, in other words when the applied voltage is equal to the voltage of the couple. We then come to another line where I, the current, is equal to zero but some voltage must be applied. Beyond that we have to reverse the current and that again is very useful to us in air conditioning applications where you have high temperature differences of 0° F to -10° F. The next quadrant now shows in effect a heat pump because the heat is actually rejected by the couple and again the C.O.P. lines and the current lines are shown. This quadrant is interesting because the C.O.P. lines are very fiat and we can choose the amount of heat delivered to the building for a given Lit over a very wide range of Lit with the thermoelectric system, which is something we cannot do with the vapour cycle. The upper quadrant again shows the various C.O.P. and current lines but so far no air conditioning use has been found and there are very few instances in which this is required.

The authors had been quoting the diametres of the couples but he would prefer to refer to cross sectional areas since the couples themselves were in fact not round. They had used 38.2 square millimeter couples with different shapes but equal performance. The shape is at the option of the manufacturer and will have no effect on performance.

V. Martinovsky. U.S.S.R. : stated that the Chairman had painted slightly too dark a picture of this kind of refrigeration. In fact this was not really a question of refrigeration

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of the future, but had already arrived. In very many instances for such applications as instrumentation, computor cooling etc., only a very small amount of local cooling was necessary, perhaps in the order of between 5 and 10 calories per hour and that there was only one solution, namely by cooling through the method of the Peltier effect. Obviously it was quite impossible to put a compressor into such very very small devices. These are not questions of the future but these are problems of today. Secondly, he wanted to stress the importance of efficiency, he felt that with the thermoelectric princ­iple the efficiency was equal to that of absorption systems or compression systems. In compression systems of course the efficiency went down very quickly with low tem­peratures and in this field there already existed a demand for thermoelectric cooling. It was not possible to give more than a certain number of stages with a compression system, but there was no reason why - to take it to the limit - one could not use more than a hundred stages in thermoelectrics. It was purely a matter of electrical connections. He felt that the only limit of efficiency with thermoelectric refrigeration was in the end the Carnot Cycle which did not apply to either the absorption or the compression system.

G. Lorentzen, Norway : replied that perhaps he must have expressed himself rather badly in his introduction of the session but he was well aware of the small systems which do exist and also the applications which already show a great need for this type of refrig­eration system.

T. M. Elfving, U.S.A. : then replied to some of the points raised by Mr. Newton. He was particularly interested in the fact that particularly low Llt's had been used and even negative Llt's. He considered that it was his job to develop rather higher Llt's and he was of the opinion that for mass production they may well be more useful. He also felt that Mr. Newton was quite right that there was no fatigue or degradation of the material after many reversals and he was glad to have confirmation now on this point.

With regard to Prof. Martinovsky's remarks the area in the United States from which he came - California - is considered to be the centre of the electronic industry and everyone is very well aware of this problem. He was of the opinion that a certain part of this requirement could be met by thermoelectric refrigeration. He also considered this progressive cooling in steps of Lit by thermoelectric refrigeration and he was of course aware that Prof. Martinovsky was the first to have published work on this part­icular aspect.

C. Hocking, Sweden : He was aware that there was an established field of application for thermoelectric cooling of small volumes or point cooling for electronic devices and there was also an established field for large air conditioners; but was there a good case for thermoelectric cooling in the domestic refrigeration field say for household cabinets of 10 to 30 litres capacity ? One would need a rectifier for convertingA.C. to D.C. and that should be taken into consideration. The refrigerators using thermoelectric refrigeration which he had seen showed a very small internal volume against external volume and the fact that a rectifier was necessary may well have something to do with this problem.

T. M. Eljving, U.S.A. : replied that what his paper presented today was not theoret­ical but in fact based on three years of actual practical work. One could not calculate and thereby predict the inner volume to the outer volume from theoretical equations. It would require actual experimentation and development work of refrigerators and he pointed out that he was in fact a domestic refrigeration engineer first and foremost. Trying to look into the future he would say that eventually the proportion of inside volume to outside volume, even taking into account the question of accommodating a rectifier, would still be very much better than the present day refrigerator. But in fact he thought that it was not really a certainty that a D. C. power supply would have to be included in the unit because there were new gadgets which could quite easily be used in a house using D. C. supply and it was quite possible that houses of the future might be provided with a D. C. supply of their own. Nevertheless, it was more interesting to talk about the cost of the power supply, and already today such units are obtainable in the United States at 8 dollars for a small refrigerator.

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E. P. Penrod, U.S.A. : replied to Mr. Newton, that in 1960 a paper was given before the American Society of Mechanical Engineers on the applications of thermoelectric refrigerators for space applications. He pointed out that the impedence of both the generator and the refrigerator had to be taken into consideration, which confirms Mr. Newtons' remark on this point.

W. G. Kogel, Sweden : suggested that the efficiency curve for absorption units shown by Mr. Elfving in his paper was at least 15 years old. The efficiency of small absorption systems has been improved from approximately 22 % in 1945 to 40% in 1957. This was due to the introduction of a submerged analyser concentration vessel and combined pressure vessel. Consumption in 1945 for a 1.5 cubic foot cabinet was 2.1. KW. hour per 24 hours and only 0.9 KW. hours per 24 hours for a 2.3 cubic foot cabinet in 1957. He therefore considered that the figure of 23 % given by Mr. Elfving was rather pessi­mistic and much lower than the present day figure ought to be.

V. Martinovsky, U.S.S.R. : said, that in the absorption system which is electrically powered one uses first of all electricity for the purpuse of heating and then the heat for the purpose of cooling, and therefore it is for this reason that it is impossible to reach a good efficiency figure in small absorption units. With large absorption systems one can have boilers etc., but the difficulty with small systems is mainly that very great losses occur in converting electricity to heat. This is a pure matter of the second law of ther­modynamics.

T. M. Elfving, U.S.A. : then replied that he was talking about efficiencies at certain Llt's and he was quite ready to believe that one can improve the efficiency of an ab­sorption type system by simply increasing the installation of the boiler. But for a Lit of 60° F the efficiency as he read it from his curve would be somewhere in the region of 26 %. His main point was, however, the relation of the Lit. With a thermoelectric unit one can readily see that this unit was much more sensitive to the Lit than an ab­sorption unit. The efficiency curves for the absorption unit were actually calculated according to the formula :

e absorption unit = 0 • 0 6 X e camot

whereas the maximum efficiency for a thermoelectric system would be calculated ac­cording to the formula:

VT+Tm - Th/T e max thermoelectric = V X e camot

1 + Tm + 1 .

H. Stierlin, Switzerland : As comparison is made in this paper with absorbers he would like to correct the C.O.P. curve which is given. In practice today the C.O.P. of a 120 litre domestic absorption refrigerator and a temperature difference of 85° C has been found to be over 40%. At 60° Lit it is approximately 50%. Therefore the curves for absorption and thermoelectric refrigerators intersect at a Lit of approximately 40°C and not at 60°C. This entirely changes the comparison made later in the paper.

G. Lorentzen, Norway : then stated, that the discussion was getting lost in absorption refrigeration rather than the subject of the discussion namely the Peltier effect and there would be further opportunities to discuss absorption systems at a later date.

A. B Newton, U.S.A. : felt that it was necessary to make one further point as far as cooling by thermoelectric systems was concerned. In air cooling one does not have the same problem which might exist in liquid cooling etc. In certain applications they had taken key - material and put it between two conducting bus bars. This type of thermo­couple could of course hardly be used for any application where insulation was required. These are then connected with heat conducting fins usually copper. The couple is really then in two pieces since the P is separated from the N (see diagram below).

Electric rfF====fl Cons tant ____... 1_F==lj

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If we then pass our electrons through in the direction shown, one of these, of course, will carry the electrical contact, but the next one just carries on the change. The reason why this matter was being brought up by him was that on some applications where one has to reject heat to the air one has to eliminate the other losses from temperature difference across electrical instruments. In this particular way the electricity flows in series right one after the other through the whole unit and there are therefore no therma barriers within the thermoelectric system itself. He felt that the audience may be in-1 terested to know that by choosing the right fins (by obtaining the information from a computor) it is possible to pack as much as 5 tons of heat rejection within a cubic foot. This does not mean that they had so far packed a whole cubic foot of it but that is the rate of packing. It is for this reason that he held the belief that the lower temperature difference applications and high sink temperature applications the region of 180° to 200° F was where the main problem will exist in the immediate future.

38

E n ergy for Refrig erat i o n i n C o m i n g Years

E n er g i e p o u r la p ro d u ct i o n d u fro i d a l 'aven i r

P-10

Energy for Refrigeration, Present and Future

Energie pour l'industrie frigorifique, actuellement et a l'avenir

Prof. BURGESS H. JENNINGS The Technological Institute, Northwestern University, Evanston, Illinois, U.S.A.

SOMMA/RE. Ce rapport analyse nos sources actuelles d'energie en attirant d'abord l' attention sur !es types et !es approvisionnements de nos combustibles mineraux. On note que tous ceux-ci existent en quantite suffisante pour durer pendant des siecles. On note aussi que !'utilisation de courant electrique double a peu pres tous !es 10 ans et qu'avec cet epuisement supplementaire des combustibles mineraux, ii semble evident que l' energie nucleaire trouvera plus d'application dans la production d'energie. En fait, meme maintenant on prevoir un grand nombre d'installations plus importantes de 400000 a 500000 kW pour l'energie nucleaire. On mentionne l' energie solaire comme complement du trio de sources d' energie disponibles pour les besoins de l'homme dans tous !es domaines, y compris le froid.

A I' exception de l'energie nucleaire, on fait ressortir qu'aucun changement fondamental de nos sources et de nos types d'energie ne s'est produit ces dernieres annees. Ce fait nous amene a douter que !es sources d'energie pour l'industrie frigorifique changeront dans une plus grande mesure que la generation d' energie. Cependant, on peut s' attendre a un certain nombre de variations de ]'utilisation de l' energie. L'une de celles-ci est une plus grande utili­sation des moteurs a combustion interne, y compris la turbine a gaz, comme source d' energie pour /es grandes installations frigorifiques.

Ce rapport etudie ega/ement ]es progres en COUTS de realisation dans /es elements thermo­e/ectriques et !es dispositijs photovoltaiques pour la production d' energie et de froid. Ceux-ci presentent beaucoup d'avantages, mais il est peu probable que du point de vue economique ces dispositijs soient deja prets pour l' exploitation commerciale. Avec !es recherches et !es progres intenses en cours, on espere que cette application se developpera rapidement avec le temps.

Before the future can be brought into focus, our understanding of the present must be clearly delineated, since by this means we are able to see at least the obvious paths of future progress. There may be some who say that completely new energy sources are on the horizon, yet an analysis of the future, based on the past, shows that with the exception of nuclear energy, all of our sources have been known and used for centuries. This does not mean that improvements, innovations and devices stemming from develop­ment research and our broadening knowledge of solid-state physics will not continuously appear, but the probability of a revolutionary break-through to create a new energy source is remote.

Let us consider the present situation for energy usage. Most of our present-day power is produced from fossil fuels, solid, liquid or gaseous. In populous communities such as towns and cities, electric energy largely produced from the combustion of fuel, in turn, provides the energy employed for refrigeration.

SOLID FUELS

Solid fuels constitute the major energy source for the central power stations that produce electricity. There is reason to belive that these fuels will continue to be the chief

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power plant energy source, since the coals and lignites are very widespread over the face of the globe, and varying estimates have indicated that there are sufficient reserves of them to provide the needs of man for 3 000 to 4000 years hence.

LIQUID AND GASEOUS FUELS These are luxury or convenience-type fuels and as such have been used primarily

for power applications in such areas as automotive, marine, transportation, isolated plants and for heating purposes. Their cost, in general, is higher per unit of energy than for solid fuels. However, this statement is not always true, even for a major power plant, when all of the factors such as lower transportation costs, reduced labor and simplicity of control are considered. There has, moreover, been over exploration and over production of the petroleum products to the extent that in consumer pricing these products on occasion have been less expensive than the solid fuels, particularly the residual oils. How long this pattern will continue cannot be predicted, and the economic situation is very complex since it is to some extent predicated on the type of actions the oil-rich nations might pursue relative to the other nations of the world.

Some twenty-five years ago it was commonly believed that the petroleum reserves of the world were limited to the extent that in a century or so these fuels would be es­sentially depleted. With modern exploration methods in use, it has become apparent that the reserves of oil and gas are much greater than formerly anticipated and with reasonable utilization by world populations should last for three or more centuries.

Gas, in particular, has been found to be more prevalent to the point that a whole pipeline matrix has been built on the North American continent to make gaseous fuel available in regions far remote from the gas production fields. In this connection, it should be mentioned that large amounts of gas are being located underwater, trapped under the coastal shelves of the continent. It would not be surprizing to find similar widespread natural gas deposits on the European continent, in fact some European dis­coveries give serious credence to this viewpoint. Gas, if properly distributed, would then be an effective source not only for the petrochemical industries and for heating but in a basic gas-turbine drive for refrigeration usage.

NUCLEAR ENERGY Nuclear energy has been moving toward widespread utilization as the heating source

in a number of large central stations. This trend is particularly significant in new plants in the range of 300000 to 500000 kilowatts in areas where the cost of fossil fuels is high. The Atomic Energy Commission of the U. S. A. has been conducting studies and found that nuclear-energy power plants are competitive with fossil-fuel plants when the cost of the fossil fuel ranges between 0.31 and 0.36 dollars or more per 1,000,000 Btu (253,000 kg cal) of thermal energy. It may be mentioned that in the U. S. A., coal is available in many areas of the nation at energy costs of less than $ 0.30 per 1,000,000 Btu so that not many nuclear plants are at present under construction.

Looking toward the future, it is felt that this condition will change both in the U. S. A. and Europe, since the price of coal and residual oils will in all probability rise at a more rapid rate than will the cost of nuclear fuel. One of the factors in the cost of nuclear fuel is that of its reprocessing and continuing progress can be expected relative to low­ering costs in this regard.

The demand for electrical energy has been essentially doubling every ten years. If this demand continues as expected, it will in turn produce greater and greater demands on the fossil-fuel supply and tend to increase its cost so that the trend must almost certainly move toward a greater usage of nuclear energy in all large power plants. A final factor in this pattern concerns the breeding of additional fuel during the operation of thermal nuclear plants. For example, uranium fuel for thermal purposes can interact with thorium in a breeder cycle and produce additional nuclear fuel in very large quan­tities so that a somewhat self-perpetuating energy system results. This type of proces­sing has not been developed commercially to the point that breeding is in intensive usage. However, it is obvious that central station usage of nuclear energy is facing an enormous increase and it is probable that the increase will occur at an even more rapid rate in Europe than in North America.

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SOLAR ENERGY

When we look beyond fossil fuels and nuclear energy, solar energy in its various manifestations represents the remaining energy source whether this be as hydroelectric power, or as wind power, or as tide or wave action, or photovoltaic cell, or merely as heating. Commercially hydroelectric power of the central station in connection with fuel-created power is most significant for refrigeration usages, but in several other con­nections solar energy for refrigeration is considered to have possibilities of importance. These will be mentioned at a later point in this presentation.

COMPRESSION REFRIGERATION

This type of refrigeration can be considered the most basic of all since it is flexible; can be served by reciprocating or centrifugal machinery, and can by multistaging serve from the lowest commercial temperatures to the highest temperatures at which refrig­erants and lubricants are stable.

The first obvious power source for compression refrigeration is the electric motor obtaining its electric energy from a central source such as a steam or hydroelectric power station. In terms of energy, we may well ask the question, how much energy or power is required? The answer here revolves around the temperature operating range and equipment performance. The latter is well known, for example in the air-conditioning range of 41° F to 95° F (5° C to 35° C), it is customary to think that one ton ofrefrigera­tion, 12,000 Btu per hour, requires one supplied horsepower, 2545 Btu per hour, or considering motor inefficiencies, etc., requires one kw from the power lines, 3413 Btu

12, 000 12, 000 per hour. For this the ratios are thus either

2, 545 = 4.7 or

3,413 = 3.5. These ratios

are called the coefficients of performance. In terms of purchased power we must recog­nize that an energy factor of refrigeration to power of 3.5 is thus realistic, even for equip­ment of limited size, say 5 horsepower and less. As the temperature range required for the refrigeration is lowered, the power required increases by significant amounts. For example, in the rapid-food-freeezing range of -40°F (-40° C) with multistage equip­ment condensing at 95° F (35° C), the power required per ton may increase to 3.4 to 3.8 horsepower, and the kilowatt electric power from the lines is almost of the same magni­tude, let us say 3.4. Thus the overall coefficient of performance for electric power is

12, 000

(3.4) (3413) = 1 .03. If electric power were always cheap and readily available, there

would be no point in carrying this analysis further, but since this is far from true in many parts of the world, we must consider other alternatives.

The first types of drive to consider are the internal combustion engines and these, in turn, might well be classified as reciprocating engines or turbines; more specifically, Diesel engines, spark-ignition engines (gasoline or gas), and gas turbines. Of these drives, the Diesel engine has inherently the highest economy and can use a range of liquid fuels at economies in the area of 0.4 pound (180 grams) per horsepower hour. The gasoline engine necessarily requires a more refined (often more expensive) fuel and has a poorer economy, about 0.5 pound (226 grams) per horsepower hour.

The gas turbine, at the present time, offers by far the most challenging promise for the refrigeration industry and there are several reasons why this is the case. First it can be stated that gas turbines can now be built in smaller-sized units, namely as low as 100 to 200 horsepower. This is in contrast to the viewpoint held until recently that gas turbines of less than 500 or so horsepower were not economically feasible. Second, the advances being made in metallurgy are such that now it is possible to operate gas tur­bines at gas inlet temperatures as high as l,450° F (794°C). If, as seems to be possible, new materials in a few years permit an increase of inlet temperature to some 1,700° F (928° C), the efficiency of the turbine will approach that of the Diesel engine and defi­nitely exceed the efficiency of small steam plants. Third, the gas turbine is very versatile in the type of fuel it cm use, from the broad range of liquid hydrocarbons to gaseous fuels of nearly all types. It is only with solid fuels that the open-cycle gas turbine is still

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ineffective since to date no simple and effective method of separating the ash from combustion products has been thoroughly demonstrated. Particulate ash unfortunately leads to erosion of the turbine blades and to the building up of ash deposits in the blade passages. In fact, this same condition can even occur with some of the residual (heavy) fuel oils, and care must be exercised to see that certain types of residual fuels are not employed with gas turbines.

All of the internal combustion devices, but more particularly the gas turbine, have another inherent advantage, namely the fact that the exhaust gases, having served their power function, still are hot and carry energy which can be economically recovered for other purposes. Most frequently, to use this energy, the exhaust gases are led into a waste-heat boiler in which steam is generated. For example, consider a representative type of gas turbine which develops 1,000 horsepower. When this turbine is provided with 68° F (20° C) inlet air using a hydrocarbon fuel to heat the turbine inlet combustion products to 1,450°F (794° C), it develops its power and exhausts gas at 800°F (427° C). This gas, when cooled to 350°F (188° C) in a waste-heat boiler, can generate steam at 15 pounds per square inch gage (1 .05 kg per sq cm gage) or even at higher pressures. Steam at such pressures is at 250°F (127°C) or even slightly higher and in this range is useful for many industrial processes, or for building heating, or when the need exists for operating absorption refrigeration equipment. Under the power and heat conser­vation conditions indicated, some 5,600 lb (2540 kg) of steam per hour could be pro­duced.

Expressed in another way, we can say that with effective, reasonably balanced power production with waste-heat utilization, from 50% to 80 % of the calorific value of the fuel can be conserved for useful purposes. In few instances can a perfect heat balance be reached so that performances appreciably poorer than the values indicated may be expected. Even in the case where a reasonable balance between waste heat and power required for refrigeration is at optimum conditions, the costs of fuel relative to the cost of power do not necessarily indicate a decision favorable to a gas-turbine and waste­heat system. This is particularly obvious when the initial cost of the gas-turbine, waste­heat system is so very much higher than the cost of the simple motor driven compressor system that the thermal saving is completely obscured by the higher fixed (amortization) costs. A number of studies of this problem have been made which analyze in detail all of the variables which enter the picture of composite systems of this type. In a general way, it can be stated that whenever liquid or gaseous fuel can be bought at 60 cents or less per 1,000,000 Btu (253,000 kg cal or 293 kw hr) for use in a gas turbine or other combustion engine, it is competitive with electric power for motor drive at 1 .5 cents per kwhr, even with poor waste-heat utilization. To generalize this statement on an inter­national currency basis, it can be stated that when the ratio of the cost of a kilowatt hour for electric power to the cost of a kilowatt hour of energy in gaseous or liquid fuel is greater than 7.4, serious thought should be given to the possible usage of a combustion engine for the creation of refrigeration. This is even more true when there exists the possibility of continuous effective use of waste heat from the gas turbine or combustion engine.

The problem becomes less clear when a design is being planned with the power-to­fuel cost ratio on an energy basis not well above 7.4. Part-load-operation problems must also be considered and the waste-heat usage balance can become a critical factor in making a design decision. The problem of different patterns of operation in winter and summer can also complicate the picture. The author of this paper was recently involved in the planning of a food processing plant where although the food processing load was fairly constant throughout the year, except for minimal usage on weekends, there was a great load difference between summer and winter. It was obvious that if gas turbines were used to power the whole low-temperature (-40° F, -40°C) refrigeration load for product freezing, there would be some periods when only trivial usage could be made of the waste heat from the power turbines. This led to making a design decision that somewhat more than half of the total refrigeration capacity in centrifugal units was to be provided by gas turbine power while the remainder employed conventional electric motors. By this balancing of load, it happened that under every condition, all of the

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steam generated in the waste-heat boilers was usable in the plant for processing and heating. Stand-by boilers furnished other steam as needed. This arrangement appeared to make the best economic balance considering fixed costs, simplicity of operation, overall costs for fuel (in this case gaseous) and electric power.

ABSORPTION REFRIGERATION

From a viewpoint of the direct use of a fuel or surplus thermal energy, absorption refrigeration is very important. It can be mentioned that this type of refrigeration is also not new, since intensive investigations and applications of it were made during the second and third decades of this century. However, in recent years, absorption refrigeration has had a rebirth of interest with the development of the units using salt solutions and water as a refrigerant. Such units are not usable below 32° F (0°C) but are very effective in the air-conditioning range. Moreover because they can be built in a variety of sizes from 5 to 1 ,200 tons in capacity and use thermal energy as their energy source, they are becoming widespread in use. The energy-supply temperature need not be very high, 248°F (120°C) to 230°F (110°C) being considered a satisfactory range for a number of units. Such temperatures are easily reached in many types of waste-heat boilers and surplus exhaust steam from processes or power usage is serviceable. In fact, where the latter condition applies for year around usage, the same steam source used for heating in winter can be used for cooling in summer.

In rough terms it is customary to think that 20,000 Btu per hour of energy input are required to maintain one ton of refrigeration (12,000 Btu per hour) in the air-condi­tioning range. This indicates a coefficient of performance of 12,000/20,000 = 0.6. Such a coefficient of less than unity, although poor for a power-driven refrigeration system is very satisfactory for a heat-energy-operated system. Absorption systems can be steam operated, hot-water operated or direct fired using a fuel. The steam or hot water in turn can be provided not only from industrial sources or boilers, but solar energy or nuclear sources must be considered as serious possibilities. In fact, absorption refrigeration units hold high promise of being able to utilize solar energy for the production of refrigeration, since the energy source, with some capacity reduction, can be at 212°F (100°C) or even a somewhat lower temperature can be used with some refrigeration-machine designs and produce satisfactory cooling output. It should be mentioned, however, that the capacity of absorption-refrigeration equipment decreases as the temperature of the supply medium falls below an optimum value.

Unfortunately the sun as an energy source is not as satisfactory as might be hoped because it is available for only a fraction of each day and not always for the whole fraction because of the presence of clouds or generally inclement weather. Thus energy storage [1] is required if continuous refrigeration is needed. This, in turn, means that excessive areas are required for collecting the solar energy, and because minimal collection tem­peratures are ineffectual, it is necessary to design the systems with real care. These requirements increase the initial or fixed cost of the collecting system to such a point that they are competitive with fuel-operated systems oruy in areas where fuel is exor­bitantly high in cost. The suggestion has often been made that such units are ideal for tropical and semi-tropical areas with high insolation but no solution to the high initial cost of equipment has yet been reached.

Active research, however, is being carried on and it is not impossible that a design could be developed in which the initial investment while not inexpensive is minimal. Chung, Liif and Duffie recently reported [2] their experiences on the performance of a lithium bromide-water absorption unit heated by a fiatplate solar collector. This used a fixed plate facing south, set at an angle of 35 degrees to the horizontal and having an effective heat-absorbing area of about 9.3 square meters. Water was heated in the ab­sorbing area of the plate to temperatures ranging from 179° F (82°C) to 201 ° F (94°C) and then supplied to the generator. Approximately 6 hours per day were available for operation by solar energy and the unit's output of cooling ranged from approximately one ton (12,000 Btu per hr, i. e. 3.5 kw of cooling) to over 2 tons. The absorption unit itself is rated at 3 tons for 212°F (100°C) generator operation, but output falls off

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drastically as the temperature of the heated water drops below 175° F (79° C). Cooling temperatures in the range of 50° F to 55°F (10° C to l3° C) were easily obtained and this range is very serviceable for space cooling (air-conditioning). The overall coefficient of performance, based on the cooling produced to the incident solar energy falling on the collector plate, is in the range of 0.1 to 0.2, although the coefficient relating cooling t<> energy provided by the hot water is much higher, being better than 0.6. The unit de­scribed in this paper was an experimental unit and design improvements could definitely be incorporated to improve performance. The results of this investigation and a number of similar attempts have conclusively confirmed the fact that solar energy can effectively produce cooling for air conditioning, and the more intense is the sun loading, the greater will be the cooling which can be produced. However, unless a unit is greatly oversized to provide for storage of cooling medium, air conditioning is not available except for daylight hours, primarily those 3 hours before noon and 3 hours following noon.

Major comment has here been made on salt solution-water refrigerant units for air­conditioning service. However, it is also possible to produce absorption refrigeration units, usually employing ammonia, which will operate in the subfreezing range for longer­period storage of foodstuffs. Unfortunately, for this purpose the lack of solar energy for a very great fraction of the day creates a serious difficulty, since the refrigeration is needed without interruption to maintain the quality of the food being preserved. This problem, provided other energy sources are not available, poses so much difficulty as. to be decidedly discouraging. Nevertheless, research and development effort is being carried out to produce a versatile unit which at least would have enough reserve capacity to operate food-refrigerators in the short-term-storage range at about 45° F (7° C).

THERMOELECTRIC PROCESSES

Thermoelectric phenomena, [3, 4] like many of the other topics discussed in this paper, are not new, since in fact T. J. Seebeck showed in 1823 that an electromotive force could be produced in a circuit of two dissimilar metals if their junctions were maintained at different temperatures. Contributions to our knowledge in this field were made by many following experimenters, one of whom was the Frenchman Jean Peltier, who in 1834 discovered the phenomenon that when a current was passed through a thermoelectric circuit of two metals, heat evolved and had to be discarded at one of the junctions while heat was absorbed at the other. From this particular phenomenon of thermoelectricity arises the possibility of producing refrigeration and the reverse of this shows that an electric generator with no moving parts can be made with a thermocouple circuit. For the latter no energy is required beyond the addition of heat to one junction while the other junction is cooled (rejects heat).

Thermoelectricity for power production or for refrigeration was not thought to be of commercial value until the last decade or so when knowledge of new semi-conductor materials began to filter out of research carried on in solid-state physics. Prior to the availability of such materials, with their strikingly better performance, thermoelectric­ity was used in thermocouples for the measurement of temperature or for the operation of simple temperature-responsive controls.

The semi-conductors [3, 4] are important because metals are not at all effective thermoelectric materials for refrigeration or power production. Metals possess electrons that can freely move under the action of a temperature gradient or of an electric field, for example, when heat is applied to one end of a metal bar, it not only causes heat to fl.ow along the bar to the cooler end, but an electron fl.ow is also set up. This electron fl.ow is an electric current, and the electrons are involved both in heat transfer and the electrical process. Semi-conductors lack the plentiful supply of free electrons found in a metal and in character lie somewhere between a metal and an insulating material. Semi-conductors possess another unusual attribute, namely by a slight change in their structure, two types of semi-conductors can be prepared. These are known as N-type, and P-type materials, and a junction of these materials has the fortunate effect of pro­ducing a good thermoelectric power in a material of poor thermal conductivity. Differ-

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ences between the N and P type relate to the manner in which electrons flow in the material as free agents or through the action of the positively charged holes which de­velop in the material.

To illustrate what is involved, we could note that bismuth telluride (Bi2 T 3), which is one of the thermoelectric semi-conductor materials in current usage, has a small excess of tellurium added to produce the free electrons, which are the negatively-charged carriers. The P material has added to it a small addition of bismuth to produce positively­charged carriers. Thus, by treatment of a base material having suitable semi-conductor characteristics, the two required materials for a thermoelectric circuit can be produced. A successful thermoelectric material should have a high melting point, possess fair electrical conductivity, have low thermal conductivity and a high electromotive force per degree of temperature. It might be mentioned that bismuth telluride has an electro­motive force of 200 microvolts per degree centigrade, and thus a 500 degree temperature difference between the junctions would produce a 0.1 volt. By arranging many junctions in series, it is possible then to produce any desired voltage by simply increasing the number involved. Other semi-conductors which have been investigated are lead telluride, titanium dioxide, silver antimony telluride and many others.

A very interesting utilization of thermoelectric effects occurs in producing electrical energy from a heat source. Solar energy is one of the potential sources that could be employed. Unfortunately, a large temperature difference is desirable if a reasonable efficiency in the use of heat energy to electricity is expected, but some power can be produced even with small temperature differences. Efficiencies in the utilizaion of energy input to electrical output at the present time are in the range of 1 % to 4 %, which is not a particularly creditable performance for a heat engine. Nevertheless it must be recognized that the thermoelectric generator is a potential source of electric power, which could be used for thermoelectric refrigerators or for other refrigeration systems. Additional work is required to produce materials with higher figures of merit, available at lower cost. At the present time the care with which semi-conductor materials must be made and the cost savings not yet realized from mass production are hampering progress.

Thermoelectric units for the creation of refrigeration up to this time have been developed primarily for small units [5] in the 40 to 100 watt range. Cost of materials has somewhat hampered their production but their poor performance relative to com­pression refrigeration has also been a factor. In fact, the overall coefficient of performance of a thermoelectric unit is about one tenth that of a corresponding compression system so that coefficients of performance in the range of 0.3 to 0.6 must be recognized. It is moreover true that it is difficult to operate through ranges of more than 90° on the Fahrenheit scale (50° on the Centigrade scale). The future of thermoelectric refrigera­tion is thus at a crossroads. Unquestionably there will be certain installations for which it is admirably suited but widespread usage is still very uncertain.

Other energy devices in various stages of development exist and two of these might be mentioned, the first being the silicon photovoltaic cell [6] and the second the photo­galvanic cell. The silicon cell has been found to be very effective but unfortunately its cost is extremely high. It responds to visible light and the shorter infra-red bands and can work with an efficiency of about 10 per cent in converting impinging solar energy into electricity. No material is consumed and maintenance costs are very low even when these units are put in isolated spaces and use sunlight as an energy source. If the cost of manufacturing the silicon crystals could be reduced by an extremely large factor, this type of unit would have a great potential for refrigeration and many other industrial uses. Difficulties also exist with the galvanic type to the point that both devices can be envisaged only for limited operation.

FUEL CELLS

When we think of future energy sources for refrigeration, it is difficult to escape the conclusion that the best we can hope for is to find more effective methods of using either our presently known fuels or the sun. One power device, which could come to the fore in this connection, is the fuel cell. Thermodynamically a fuel cell is merely a device in

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which a fuel, usually gaseous, combines with oxygen in such a manner that the energy released by this combustion process is almost completely converted directly into electric energy. The principle of the fuel cell is not new since the British physicist Sir W. Grove­made a so-called "gas-battery" of this nature early in the last century and computations using the potential functions of thermodynamics readily indicate the amount of electrical energy that could be produced. Such computations show that some 80 per cent of the thermal energy of the combustion process can be converted into electrical energy.

In Grove's original experiments, he showed that when electrically conductive water was decomposed into its constituents, hydrogen and oxygen, if the current were stopped and a circuit closed, the hydrogen and oxygen in their separate storage containers would react at the electrodes to combine. At the same time an electric current would flow in a reverse direction to that for decomposition of the original water and water would again be re-formed as en end product.

Unfortunately the gas-storage battery was slow in response because the reaction occurred only at the platinum electrode and water surface. Thus it required too much space per unit of output, and interest in it became almost dormant. However, the fuel cell in many modifications has recently reached a high degree of research interest and apparently is at the point of commercial breakthrough in more than one design. The most promising designs have involved hydrogen-oxygen, hydrazine-oxygen and methane­oxygen. Some cells operate at atmospheric or only slightly elevated temperatures, but certain other designs have had to use very elevated temperatures to maintain the reaction.

One such cell design for spacecraft operation develops 2 kilowatts and makes use of a so-called ion-exchange membrane. In the cell, hydrogen and oxygen are fed to their respective electrodes on the two sides of the solid membrane which acts as an electrolyte. This solid electrolyte has the ability to permit the flow of hydrogen ions which pass through, combine with the oxygen at the other electrode and produce water as an end product. The ion flow constitutes an electric current between the electrodes and the external circuit just as it would in a liquid electrolyte.

In general the performance of fuel cells has been in the range of 60 to 70 per cent of the heating value of the constituent fuel and its oxygen being available for use as electric energy. This performance is not in disagreement with the Carnot principle since it does not involve a temperature-response pattern. Fuel cells of varying designs have not been in commercial use sufficiently long to evaluate even qualitatively their perform­ance.

The life of the electrodes, membranes (where used), the removal of the water, all in­volve problems which while not insurmountable are nevertheless present. The voltage produced can be varied to suit the situation by external switching to put the cells in series-parallel operation as may be required.

The potential interest of the fuel cell to refrigeration is threefold : (1) It represents an extremely efficient generator for production of electric power for refrigerator motors, (2) It is capable of using hydrogen and oxygen directly for power, and (3) Solar energy can be used to produce hydrogen and oxygen by electrolysis and fairly simple storage facilities can be constructed.

Such an application of solar energy is very intriguing since a simple thermoelectric­generator can make use of solar energy at its hot junction and with atmospheric cooling of the cold junction produce sufficient current at any desired voltage to electrolyze water into hydrogen and oxygen. Storage of the gas poses a problem but low-pressure gas holders over water, or plastic bags which inflate or collapse under filling, and subse­quent gas withdrawal appear to be simple possibilities. For major operations the hydrogen and oxygen, or perhaps merely the hydrogen, could be compressed and stored under pressure in tanks or pipe cylinders. The oxygen produced could be discarded since oxygen is also available in the air in unlimited quantities and will work with somewhat lowered efficiency in some of the fuel cells. It is also obvious that hydrogen and the oxygen from air can also be burned in an ordinary combustion process for the purpose of generating steam or serving as direct combustion gases to operate absorption-refrig­eration units.

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The purpose of these discussions is to indicate ways and means of simply using solar energy, and storing during the daylight hours enough surplus fuel to operate refrigeration equipment continuously. In some of the less-industrially developed countries, serious consideration should perhaps be directed into these possibilities. However, we cannot avoid the almost inescapable fact that at present the initial cost of systems to utilize solar energy for refrigeration on anything but a small scale is almost prohibitively high. However, we are thinking of the future and with the advances from research and develop­ment, which are now taking place, it is not unreasonable to predict that cost figures under mass production techniques may reduce to bring to realization many things which now appear economically unrealistic.

REFERENCES

r. F. Daniels, Energy Storage Problems, Solar Energy, p. 78-83, Vol. 6, Sept. 1962. 2. R. Chung, G. 0. G. Liif, ]. A. Duffie, Experimental Study of a LiBr-H20 Absorption Air Con·

ditioner for Solar Operation, ASME paper 62-WA-347, Nov. 1962. 3. A. F. Joffe, Semiconductor Thermoelements and Thermoelectric Cooling, Infosearch Ltd.,

London 1957· 4. L. A. Stabler, Thermoelectric Refrigeration, ASHRAE Journal, p. 60-65, Vol. l , Sept. 1959· 5. W. R. Danielson, Temperature Controlled Chamber, Thermoelectric, ASHRAE Journal, p. 30-33,

Vol. l, Feb. 1959· 6. M. B. Prince, The Silicon P-N Junction Solar Energy Converter, Trans. Conference on Use of

Solar Energy, Vol. 5, University of Arizona Press, Tucson, Arizona 1958.

DISCUSSION

N. Kurti, U.K. : said that the figures presented by Prof. Jennings for the coefficients of peformance of air conditioning plant correspond to a Carnot cycle efficient of about 50%. Was there any likelihood of these efficiencies being increased to about 80 % for refrigerators running between say -70°C or -80°C and room temperature ? If this could be achieved such degree of refrigeration could be used profitably in processes involving dual heating.

It is known that the heat evolution in underground electric power transmission in built-up areas was becoming a serious inconvenience. With Carnot cycle efficiencies of 80 % the overall loss in power transmission (dual heating and power for refrigeration) for a refrigerated power line would be about the same as for an uncooled power line; there might even be a small saving. The important thing however was that the heat would not be dissipated over a wide area underground but rather at selected points.

Since cryogenic techniques using temperatures of -100°C, -200 °C or even lower, are being applied on an increasing scale in technology, even small improvements in the Carnot cycle efficiency may have a profound effect on the economics of some processes.

B. H. Jennings, U.S.A. : replied, that in his presentation he may have given rise to a misleading conclusion by the questioner. What he meant to say was that in the air­conditioning range we are able to obtain about three and a half times as much energy in terms of cooling power as we have actually supplied. Whereas as we go down to somewhere in the region of -40°F our electric power ratio i. e., the overall coefficient of performance for electric power goes down to somewhere in the region of 1 . Therefore it is at this point that our reversal occurs and if we go down to temperature of below -40° F or C more electrical power would be required to put in to the plant than is obtained in terms of refrigeration. However, in terms of Carnot efficiency we are talking about the possibilities of 60, 70 and 80 % and if our energy ratio approaches those figures we are still doing relatively well even at low temperatures.

H. B. Clark, U.S.A. : wondered if the advent of the Nuclear Power Station may not in fact change the picture quite considerably in the future.

B. H. Jennings, U.S.A. : replied that he had not the exact figures at his finger tips at the moment but he was sure that there were at least twelve nuclear power stations under construction in the U.S.A., so that from the electrical view-point the competition is quite serious or will be even more serious in the near future.

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A. B. Newton, U.S.A. : stated that he would suggest that one might consider the efficiency of the power generating plant as part of the fuel-to-cooling C.O.P. in the electrically driven refrigeration system. If this is done the overall C.O.P. for air-cond­itioning using electrically driven compressors may be between 0.7 and 1 .05 for steam power plant generation. This then compares with the 0.6 C.O.P. of salt solution type absorption systems.

It should also be noted that many cells generate D. C. which is directly useful in thermoelectric cooling, thus eliminating some problems in control and of lost efficiency experienced in use of A. C. supply.

B. H. Jennings, U.S.A. : replied, that he quite agreed with the speaker and that ob­viously the conversion from one type of power supply to another would show up in the overall cost of running a plant.

W. B. Gosney, U.K. : stated that in comparing the use of nuclear and fossil fuels it should be remembered that at present the fossil fuels are being used by only a small fraction of the world's population. As the use of these spreads to the underdeveloped countries and continuously increases, which has been the rule in the developed count­ries, the comparison will change considerably. In 40 or 50 years time the situation will be entirely different.

Furthermore fossil fuels are not only reserves of energy but of valuable chemicals also, without which our present technology would be impossible. If we consider the needs of future generations which is a mark of our civilization these reserves of chemicals must be conserved as long as possible.

B. H. Jennings, U.S.A. : stated that he agreed with everything that had been said by Mr. Gosney and in fact thought that this process of which he was speaking had already been started.

L. C. Constant, U.K. : said that his question concerned the driving of refrigeration compressors by gas turbines. It was surely a fact that the life of a turbine was very much shorter compared with a conventional drive. A turbine operating at 30.000 r. p. m. generating 300 h. p. would probably operate no longer than 2.000 hours before requiring attention. Anything under 10.000 hours for a major overhaul would be completely unacceptable under present day standards of refrigeration practice.

B. H. Jennings, U.S.A. : replied that he thought it was not correct any more to state that the blades of a gas turbine had a very limited life. In fact during the last four or five years with improved materials great strides had been made and even with tem­peratures in excess of 400°F examination after five years of running had shown that in many instances the blades do not in fact even need replacement. This was due entirely to the advances in metallurgy of the blades and this was, as he has previously said, a more or less recent development.

V. Martinovsky, U.S.S.R. : said that in the interest of just the economy which Prof. Jennings had in mind, in the U.S.S.R. they had adopted the central system and district system in order to heat whole towns in winter from one system only and provide hot water in summer. He felt that perhaps in this direction lay the solution of the relative inefficiency of our present day systems.

48

T i m e-Te m p erat u re-To l e rance for Frozen Foods

C o m po rt e m ent des a l i ments congeles, e n fo ncti o n du temps et d e la tem perature

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Frozen Foods - Recent Advances in Science and Technology

Produits congeles. Recents progres de la science et de la technologie.

A. A. KLOSE Western Regional Research Laboratory*, Albany, California, U. S. A.

SOMMAIRE. Dans une etude generale des progres recents realises dans le domaine des produits congeles au Canada et aux Etats-Unis, on decrit les besoins d'une qualite excellente en f onction des matieres premieres, de la transformation, de l' emballage et de l' atmosphere, du temps et de la temperature d' entreposage. Le meilleur contrtJle des matieres premieres et aonc de la qualite initiale se f ait par des operations integrees dans lesquelles l' exploitant etablit la variete, la maturite et les conditions precedant le traitement. L'influences du troilemant, de l' emballage et des conditions temperature duree sur la qualite sont variees et complexes, mais certains principes elt!mentaires de trans/ ormations chimiques et physiques sont generalement applicables. Tous les produits sont des systemes tres instables, continuellement soumis a diverses reactions de degradation irrlversibles se produisant a des vitesses dependant de la concentration effective des reacteurs et de la temperature. Un grand nombre de ces reactions avant des coefficients de temperature eleves, les heures passees a temperature ambiante OU a temperature de refrigeration peuvent etre equivalentes a l' eff et de deterioration de nombreux mois passes a des temperatures de congelation. On peut estimer defafon sare les pertes de qualite en integrant la vitesse de deterioration en fonction de la temperature pour toute la periode d'exposition. L' oxygene de l' air est souvent implique dans la deterioration, la reduction de la presence d' oxy­gene dans l' emballage des produits ou l' emballage avec un gaz peut done etre aussi favorable qu'une reduction nette de la temperature.

INTRODUCTION

Our ability to adapt is somehow linked with our ability to forget. In our rapid adaptation to frozen foods as a commonplace marketing form, we sometimes forget how young the frozen food industry is. Twenty years ago the average grocery store had a small closed six­foot freezing compartment in one corner. Total frozen food sales for the United States were about 150 million dollars for 500 million pounds of product. Now our large markets have long open frozen food display cases stretching the entire length of the store. Present frozen food sales approximate three billion dollars representing eight billion pounds of frozen foods. This growth does not mean that the average consumer is eating more food. Rather, the frozen food has replaced a fresh, refrigerated, or canned form. Inadequacies of the old marketing forms encourage development of the new. To the extent that the new form excels, it will replace the old ; to the extent that it is deficient, it will be sup­planted. What are the requirements for excellence in frozen foods ? And what are the limi­tations of this processing form in competing with others ? We hope to reveal some answers to these questions in our general discussion of recent advances in the science and technology of frozen foods in the United States and Canada.

" A iaboratory of the Western Utilization Research and Development Division, Agricultural Re­search Service, U. S. Department of Agriculture.

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In establishing the requirements for excellence in frozen foods it has been convenient to refer to the studies as Time-Temperature Tolerance or TTT studies. Translations of this term into other languages are not as euphonious as the English combination. This is probably fortunate, because there are several factors in addition to temperature and time of storage that are critical for insuring excellence in frozen foods. For discussion we shall divide these factors into five segments-raw material characteristics, processing conditions, packaging, temperature, and time. Their importance will be illustrated by reference to specific commodities, and if I mention poultry more often than other commodities, please excuse it as a predilection built up by thirty years experience in the poultry field. If not otherwise referred to, data used in the following discussions were derived from publications from the Western Regional Research Laboratory of the U. S. Department of Agriculture.

The term "storage life" as used in the following discussion will refer to the time required to develop the first organoleptically detectable loss of quality when com­parison is made with a fresh product or one stored at -30° F (-34.4°C).

RAW MATERIAL CHARACTERISTICS

The quality of frozen food as the consumer receives it can never be better than the potential quality present in the raw material. This is a time-worn precept, but it is still worth repeating. How is this concept being implemented ? How can further attention be directed to this area ? Through careful evaluation followed by rigid specifications, selection of the best varieties and proper degrees of maturity for frozen fruits and vege­tables has contributed greatly to the excellence of the product. In the field of meat and poultry, close attention to feed, age of the animal, and conformation and finish of the carcass will generally establish initial quality. These controls over quality have been made possible and will be extended by the trend toward integrated operations in which the proces­sor or distributor has control over the growth of the commodity. In this way, processing can be scheduled to provide an optimum variety (strain or breed) and maturity for the pro­duct. For example, an integrated turkey business of one million birds annual capacity controls breeding flocks, hatchery, growing operations, processing and distribuLion. Processing schedules are synchronized with dates eggs are set to be hatched. The same close scheduling is a part of large frozen fruit and vegetable operations. Uniformly high quality of frozen products is incompatible with haphazard methods of procurement of raw material. I suspect that the poor economic position of stewing chickens (old hens), which are frozen in large quantities, is due partly to the fact that they are not a standard product, but merely represent what the egg producer is trying to dispose of when the hens quit laying eggs at a profitable rate.

PROCESSING

Processing, which we shall define as those steps required to convert the raw material to a ready-to-eat or ready-to-cook form suitable for freezing and frozen storage, can influence quality in many ways. Some of these ways are highly specific for a commodity or type of commodity. However, there is one principle of general application. All vegetable and animal foodstuffs are highly unstable chemical systems, with no reversible Carnot­like cycles operating, so that the only direction that quality can go is down. Since our standard of high quality is the fresh product, all organoleptically detectable reaction products and physical changes represent a loss in quality. These quality losses accumu­late in accordance with the laws of physical chemistry at rates dependent on various fac­tors, such as the nature of the reaction, physical state of the system, availability of reac­tants such as oxygen, and the temperature. Rate of quality loss can be reduced by lowering the effective concentration of reactants or catalysts, or by lowering the temperature. The temperature coefficient, Q10, (the factor by which the rate multiplies for each 10°C rise in temperature), varies with the foodstuff and ranges from about 2 to 16 . It is under­standable, then, that short exposures to ambient temperatures and readily available oxygen during the processing period may result in quality losses equivalent to the loss developed over much longer periods in subsequent frozen storage. An extreme example of this

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has been reported for sliced strawberries, for which 1 1 hours at 70° F (21 ° C) is equivalent in deteriorative effect to one year at 0°F (-18° C).

However many factors other than ambient temperature can influence quality. For example, in vegetables such as green beans, the nature of blanching used to inactivate enzymes has a profound effect on initial quality and retention of quality during frozen sto­rage. Here, too little time and temperature result in quality loss. Inadequately blanched beans are inferior in retention of color, texture, and flavor. Dietrich, in the U. S. Depart­ment of Agriculture, found a high temperature (200-212° F, 93-100° C), short time (1-2 minutes) blanching to be optimum.

In poultry, several steps in the sequence of processing are critical for subsequent excel­lent quality in the frozen product. Scalding in hot water to facilitate feather removal, if carried out at too high a temperature or for too long a time, will result in less tender meat. Tenderness is also decreased by excessive beating of the bird with power driven rubber­fingered feather-picking equipment. Thus gentle care is required in processing to insure tenderness in the frozen product. In addition, a minimum aging period of 6 to 12 hours at chill temperatures is required before freezing in order to provide optimum tenderness in a bird that may be cooked from the frozen state.

In the case of fish, Peters and others at the Bureau of Commercial Fisheries have shown that holding whiting fish in ice prior to freezing results in marked reduction in frozen storage life. Fish held two days in ice had a 12 months' storage life at 0°F (-18° C), whereas those held seven days in ice had only two months' storage life at 0°F. Such mar­ked effects of prefreezing experiences provide a possible explanation for large differences in reported values for storage life, and emphasize the importance of strict control in the pro­cessing steps prior to freezing.

In some cases, additives need to be introduced during processing to preserve quality. Examples are sugar and ascorbic acid for fruits and antioxidants for fat-containing precooked foods such as poultry. Recently there has been great interest in the addition of polyphosphates to meat, fish, and poultry to control loss of natural moisture and pre­serve quality during storage. While polyphosphates have been used in red meats for some time, only within the past year have some fish processors in North America added poly­phosphates to their products in order to control loss of liquids from the tissue after tha­wing and during holding at refrigerated but non-freezing temperatures. Incorporation of polyphosphates into poultry carcasses by addition to the chilling water has been proposed but not yet approved in the United States.

The freezing process itself represents a critical step for many commodities. For foods that have a large temperature coefficient of deterioration, such as strawberries, slow rates of freezing may permit deterioration equivalent to many months of storage at 0° F (-18°C). Eighteen hours at 40° F ( 4 ° C) will result in just as much deterioration as 6 months at 0° F (-18°C) for sliced strawberries packed in composite cartons. In addition to this general effect of time-temperature, types or rates of freezing may have quite specific effects in cer­tain commodities. The appearance or degree of lightness of the skin surface of frozen poult­ry varies widely as a function of the freezing rate. Slow freezing rates, such as would be experienced in still-air freezers, produce large ice crystals in the surface layer and a corres­ponding dark appearance ; rapid freezing by air blast and very rapid freezing by liquid immersion produce progressively smaller ice crystal size and increasingly whiter appearan­ce. If the bird is to temperatures in the 20°-30° F (-7 to -1 ° C) range during storage, darkening of the rapidly frozen surface occurs, presumably due to transformation of smaller ice crystals into larger ones. We see from this that we can establish and retain the most acceptable appearance by suitably adjusting the freezing rate for the outer layer of the carcass and then holding at a sufficiently low storage temperature (0° F, -18°C, or be­low). Freezing rate has become of particular interest in view of the recent development of freezing methods utilizing liquid nitrogen or carbon dioxide.

Liquid immersion-freezing of poultry and other commodities in glycol solutions or other low freezing point liquids was first advanced on a commercially successful basis by Canadian workers. This development, which is growing in volume, has permitted very rapid freezing rates and a controlled degree of lightness or color on the surface of the product. Lentz and van den Berg, at Ottawa, Canada, established optimum liquid

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temperatures and immersion times for ready-to-cook poultry. With liquid immersion­freezing systems, a frozen outer shell can be developed, at a temperature somewhat below the planned ultimate storage temperature of the product, and the temperatures of the shell and the interior allowed to equilibrate in storage. A natural extension of this principle, first developed in Canada and destined to assume increasing importance, is the use of a liquid nitrogen shower to freeze foods. Other gases that offer possibilities in this applica­tion are nitrous oxide and carbon dioxide. Liquid-nitrogen freezing provides a significant impetus to the use of continuous freezing operations on conveyor belts, in place of conven­tional batch type freezing operations.

PACKAGING

Obviously important factors in processing are packaging and the atmosphere existing within the package. Packaging is so influential in determining the frozen storage life that we should never specify time-temperature requirements without defining the pack­aging condition. I strongly suspect that the wide range in frozen storage life reported for a product at a particular temperature is often merely a reflection of variations in packaging. Let us review a few examples that support this view. One of the major quality losses in frozen peaches in sirup is the browning or oxidation of slices exposed to the atmosphere of the head space in the package. Slices submerged in the sirup to which ascorbic acid has been added are protected from oxidation, while exposed slices are quite susceptible. Reducing or eliminating the head space, that is, the package atmosphere, will prevent seri­ous browning. Also, the less permeable the package wall is to oxygen, the less browning will develop in frozen storage. Here, then, oxygen availability, or conversely package permeability, is a critical factor.

Let us consider now an entirely different product, raw chicken meat. Packaging serves two purposes, to keep natural moisture in, and to keep the oxygen of the air out. Packaging may be just as important as temperature and time considerations. Frozen storage studies of commercially packaged, cut-up, raw chicken fryers have shown that substituting a poor package for an excellent one results in just as much additional rate of quality loss as increasing the storage temperature from 0° F (-18°C) to 20°F (-7° C). While high per­meability to oxygen (air) and high permeability to water vapor generally go together in a poor package, it is the oxygen that we must watch particularly in controlling flavor deterio­ration. Moisture loss is often reflected in deteriorated appearance (freezer burn), but it apparently has little direct effect on flavor. The relative importance to odor and flavor deterioration in frozen cut-up raw chicken of oxygen-dependent reactions, oxygen-inde­pendent reactions, and moisture loss has been demonstrated at the Western Regional Research Laboratory in storage studies in controlled atmospheres. In Table 1 are shown

Table 1 . Rates of Moisture Loss and Oxygen Consumption in Sealed Cans of Cut-up Raw Chicken

------------ ---

Storage Temperature 0°F 10°F 20° F -18°C -12° C -7° C

Type of Pack Air Air Air Air Air Air Desiccant Desiccant Desiccant

Rate of Moisture Loss < 2 27 < 2 51 < 2 100

(% of maximum)

Rate of Oxygen Consumption 12 12 38 38 100 100

(% of maximum)

the experimental conditions to which raw cut-up chicken was subjected in sealed cans, and the resulting rates of moisture loss and oxygen consumption. Excessive moisture loss had no effect on rate of oxygen consumption. Table 2 shows, for the same experiment, the

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Table 2. Rate of Off-odor Development in Raw Chicken in Relation to Oxygen Availabi­lity and Moisture Loss

Condition in Can

< 0.5% Ou No Moisture Loss 20% Ou No Moisture Loss 20% 02, Excessive Moisture Loss

Reciprocal of Months to Detect Off-odor at 0° F at 10°F at 20°F

(-18°C) (-12°C) (-7 ° C)

< 0.06 0.14 0.14

0.14 0.3 0.3

0.3 2 2

resulting rates of off-odor development. Here again oxygen availability and temperature are the important factors, and moisture loss has no effect. The same situation has been demonstrated by Steinberg and others at the University of Illinois for beef stored at 0° F. Steps to insure minimum oxygen availability are simple but nevertheless are often neglec­ted. Tight compact packaging to eliminate head space or entrapped air is a minimum re­quirement. Vacuum packing or packing in nitrogen is a further step toward elimination of oxidative deterioration. Packaging material should be selected and designed to give a continuous barrier having very low permeability to oxygen. Several of the synthetic plastic films are adequate in this respect.

As our commercial packages for frozen foods change, and they have experienced major changes over the past several years, we must be alert to possible resulting changes in storage stability.

TIME AND TEMPERATURE

The last two factors important to quality, time and temperature of storage, are naturally considered together, because it is the rate of deterioration, established by the temperature, multiplied by the time that gives us the total quality loss. Since the temperature may, and almost always does, vary over the life of the product, it is the integrated expression of rate multiplied by time that is meaningful. If rate of deterioration increased linearly with increase in temperature, then the total quality loss could be obtained by multiplying the rate at the overall average temperature by the time of exposure. But we all know that this is not the case ; rather the rate increases exponentially with temperature so that we have approximately a constant ratio Q10 between the rate at any temperature and the rate 10° below that temperature. It follows that the deteriorative effect of a series of temperature fluctuations is not equivalent to the same overall period of time at a steady temperature equal to the arithmetic mean of the fluctuations, but to the same time at a temperature somewhat higher than that. This effective steady temperature equivalent to the effect of a fluctuating condition, has been determined experimentally for several products and for several precisely known histories of fluctuation, and the values agree in general with values calculated from the temperature coefficients by use of the Arrhenius equation. Table 3 gives some illustrative values, from studies at the Western Regional Research Laboratory, of effective steady temperatures for rather large fluctuations of storage temperature in commodities that have a wide range of temperature coefficients of deterioration. It is apparent that for excessive fluctuations of temperature or for products with extremely lar­ge temperature coefficient, the effect of a series of temperature fluctuations is likely to be somewhat greater than the effect of a steady temperature equal to the arithmetic mean. This is not easy to demonstrate where we have to rely on subjective methods of evaluation ha­ving rather large standard errors of estimate. For example, the absence of a demonstrable difference between the effects of fluctuating temperature and mean temperature for the last, three items in Table 3 is probably due to the taste panel methods which are incapable of detecting the effect of a difference due to only one or two degrees difference in tempe­rature.

The experimentally determined effective steady temperatures in Table 3 established by interpolation from values obtained for 4 or 5 steady temperatures, agree in general with

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Table 3. Deterioration at Temperature Fluctuating Continuously between 0°F (-l8°C) and + 20°F (-7°C) in a 24-Hour Period, Sine Wave Cycl. Mean Temperature of Fluctuations- + 10°F (-12.2°C)

Temperature Effective Steady EST-Mean Product Quality Loss Coefficient Temperature o c

Qio0c (EST)°C

Objective Methods Green Beans Chlorophyll loss 18 -9.8 2.4 Peas Chlorophyll loss 1 1 -10.2 2.0 Strawberries Ascorbic acid loss 7 -10.7 1.5 Chicken Moisture loss 3 -10.6 1.6

Subjective Methods Cauliflower Flavor 16 -10.l 2.1 Peas " 10 -10.6 1 .6 Peaches " 14 (-12.2)* 0 Precooked chicken " 3 (-12.2)* 0 Raw chicken " 3 (-12.2)* 0

* No significant difference between fluctuating condition and mean.

values calculated from the Q10's and rate theory. Thus it would appear that for at least most of the deteriorative reactions encountered in the storage of frozen foods, the total loss of quality may be obtained by integrating the deteriorative rate, experimentally known as a function of temperature, over the period of known temperature history. Van Arsdel and Guadagni of the U. S. Department of Agriculture devised a useful graphical integration method for estimating the total deterioration produced by any temperature pattern with­in the range for which the relation between rate and temperature is accurately known. It was pointed out that the validity of this procedure depends on the additivity and commu­tativity of the changes involved, and these properties have been well established for most of the commodities and important quality changes. Possible exceptions to additivity of effects lie in physical changes, such as sublimation of ice from product to inner package wall, which depend on transient temperature differentials within the package. By the method of Van Arsdel and Guadagni, any known temperature experience, whether in processing, storage, or distribution, and however irregular, can be equated to a certain period of exposure at a selected steady temperature. If further, we know the permissible storage life of the product at the selected steady temperature, we can subtract the calculated exposure time and thus obtain an estimate of the remaining storage life possessed by the product.

If we now were to present to you a ready reference table for the storage life of standard frozen food items at several commercially available temperatures, we would be con­tradicting much of what we have said up to this time. For we would be ignoring the varia­tion in storage life introduced by differences in processing and packaging. And a table that would include data for all current processing and packaging combinations would be cumbersome, if not practically impossible, and could be obsolete in a year's time. How­ever, with these reservations in mind, we can offer some generalizations to serve as guide lines for future research and development. Many commercially prepared and packaged frozen animal and plant products remain highly acceptable at a storage temperature of 0°F (-18°C) for at least one year. Storage life at higher or lower temperatures will natu­rally depend on the specific temperature coefficient of deterioration for each product. For example, while strawberries and chicken both have a storage life of about one year at 0°F (-18° C), at a higher temperature of +20 ° F (-7° C) strawberries with a large temperature coefficient of deterioration have a storage life of only 2 weeks, but chicken with a small temperature coefficient 12 weeks. The at-least-one-year storage life at 0° F (-18° C) can apply to fruits such as strawberries and peaches, to vegetables such as peas

54

P-2

and green beans, and to raw meats such as beef, turkeys, and chickens. But there are important exceptions, which we would like to emphasize. These exceptions include products with inherent instability, such as fish with a high content of unstable fat ; products with instability introduced by processing, such as precooked poultry dishes ; and formulated products having unstable texture such as emulsions and gels in dressing, sauces, and desserts.

From a survey of the literature on the storage life of frozen fishery products made by Lane of the Bureau of Commercial Fisheries in Gloucester, Massachusetts, it is apparent that a wide range of stabilities has been reported for a great variety of fish products. Most of the values for storage life at 0° F fall around six months, which is considerably lower than for many other animal products and apparently reflects the more highly unsaturated and unstable fat of the fish. Extensive studies on the frozen storage of fish products have been made by Dyer, Tarr and others in Canada and by Slavin and others in the United States, but there is still a great need for consistent data on rates of deterioration and tem­perature coefficients for the many diverse species and types of fishery products. In order to fill this gap in knowledge, the Bureau of Commercial Fisheries at Gloucester, Massa­chusetts, is presently conducting an extensive study on frozen fishery products. Emphasis is on development of objective tests, and evaluation of quality losses in relation to pre­freezing holding conditions and time and temperature experience in storage and distribu­tion.

Precooked frozen products, which generally require only warming up to serving tempe­rature, have more stringent requirements for processing, packaging, storage and distri­bution to assure optimum quality. Rancidification during frozen storage proceeds more rapidly in cooked than in raw frozen poultry, and we can expect this to be true in many meats or fat-containing foods. Oxidation of the fat begins during cooking, so it often may be advantageous to add antioxidants prior to cooking. In storage at 0°F (-18° C), prefried packaged chicken develops off flavors within four to six months, compared with one year for raw chicken. Exclusion of the oxygen of the air from precooked frozen pro­ducts has been shown to result in appreciable increase in storage life. Tight-fitting im­permeable packaging is essential. Additional benefits can be derived from replacing air in the package by nitrogen. Precooked foods formulated with a sauce or gravy, such as meat pies, possess a built-in medium to exclude air, and correspondingly have been found to have storage life of about one year at 0°F, which is definitely in excess of that for precooked products without surrounding liquid.

Special problems in acceptability occur with frozen foods that contain emulsion or gel structures, such as sauces, gravies, fillings, and dressings. Liquid separation often appears after thawing, with a curdled appearance and a general loss of eye appeal. Texture deterio­ration in storage generally has a high temperature coefficient so that temperatures of 10°F (-12° C) and higher are particularly damaging. Alleviation of this texture deterioration has been accomplished by use of special thickening agents such as waxy rice flour. For oil emulsions, such as salad dressings, use of special oils and waxy rice flour have proved beneficial.

FUTURE TRENDS

In this volatile food industry, where trends are so pronounced and their motivation so basic, predictions are freely offered and easily forgotten. It is reasonable to expect that frozen foods will continue to be developed and marketed from a greater variety of food­stuffs, in a greater multiplicity of formulations, and with more and more convenience built into the products. The trend carried to its ultimate would make an international cook out of every housewife, and a gourmet out of every husband.

Will these trends make frozen foods the predominant marketing form ? Not necessarily, only if the frozen food industry offers more economy, quality, and convenience than alter­native competitive forms of processing and marketing. The present small proportion of perishable foods such as meats that are marketed in frozen form are an inverse measure of potential growth. Continuing research and development in frozen food quality and preser­vation is an essential step toward the realization of this potential.

55

P-9

Factors Affecting the Keeping Quality of Frozen Foods

Facteurs agissant sur la conservation des produits congeles

EMIL DALHOFF and MOGENS JUL Danish Meat Research Institute, Roskilde, Denmark

SOMMAIRE. On etudie les resultats des etudes sur le comportement des aliments congeles en fonction du temps et de la temperature. On montre que di ff er en ts types de reactions et de produits ont une influence sur la f orme - ligne droite ou courbe - des caracteristiques de conservation. L'influence des variations des methodes et des jurys de degustation sur la deter­mination des pertes de qualite est considerable et il f aut disposer d' analyses statistiques des resultats. On indique Les resultats de recherche sur les vitesses de perte de la qualite, lorsqu'on entrepose les produits pendant des periodes plus longues que celles correspondant a la premiere difference decelable. On ne peut pas utiliser def acteur constant reliant le temps d' entreposage jusqu' a la premiere dijf erence decelable et le temps d' entreposage jusqu' a la possibilite d' accep­tation par le consommateur. It semble que des periodes souvent nettement plus tongues que celles trouvees avec !'utilisation d'un f acteur de deux permettront de maintenir la qualite a un niveau ou le consommateur ne pourra pas deceler de difference avec un produit venant d'etre congele. L'influence des variations de qualite des produits crus, les methodes de transformation et en particulier l' emballage ont la plus grande importance sur la conservation des produits congeles et doivent toujours etre a !'esprit lorsqu'on considere le comportement d'un produit en fonction du temps et de la temperature. On conclut qu'il f audrait plus de donnees et de carac­teristiques de la qualite et cela pourrait aider pour commencer a stimuler I' inter et de produire des denrees plus appropriees, comme les recherches sur le temps et la temperatures nous ont deja aides a ameliorer notre equipement d'entreposage et de distribution.

INTRODUCTION

It has been known for a long time that lowering the storage temperature results in increased keeping times for frozen products. Actually, as may easily be seen from most of the recommendations, laws, and regulations concerning frozen foods, one was in the past so interested in the problems of temperature that very often the equally important factor of storage time was overlooked.

The appearance within the last six years of reports on one of the most comprehensive research programmes undertaken in the field of food preservation, the so called TTT programme, carried out by the Western Regional Research Laboratory of the US Depart­ment of Agriculture has been of the utmost importance in helping to alter this unfortunate approach to the problems of frozen food storage and distribution.

THE TIME - TEMPERATURE RELATIONSHIP

Besides many important data on a great variety of products the TTT investigations gave us a new approach to the problems. It was shown that a frozen food product has quite a regular relationship between the storage temperature and the time it takes at this tem­perature to experience a certain loss of quality. For each temperature was found a corre­sponding daily loss of quality. Also it was found that the influence of various time-tem­perature conditions during storage of a product was cumulative and that the sequence of the time-temperature experiences were without influence on the size of the accumulated total quality loss.

On the basis of these relations many similar investigations on all types of foodstuffs are now carried out in several countries. The valuable data collected in this way will broaden our knowledge concerning many aspects of the processing and distribution of frozen foods. Much more need to be done, however, before we shall be able clearly to understand,

5 7

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calculate, and predict the influence of all the different factors influencing the quality and storage life of frozen products. In the following we shall discuss some of the questions opened by the investigations of recent years.

THE SLOPE OF KEEPING QUALITY CHARACTERISTICS

The time-temperature relationship of a frozen product may be presented by drawing a so-called "keeping quality characteristic". This is done by using a semi-logarithmical diagram with the temperature on the linear axis of abscissa and days of storage life on the logarithmical axis of ordinate.

30 -JO -25 -20 -15 -10 -f -r--- -

T 20 Raspberries

10 10

� 7

V) 5 r: � ;§ 3 Pork � Chops

·2 '"� Chicken

0.7 07

05 Q5

QJ - + QJ - JO - 25 ·20 - 15 . 10 -5 0 TEMPERA TURE 'c

Fig. 1 : Keeping quality characteristics of selected types of products. Some are straight, some curved lines over the temperature intervals examined.

1000 -30 -25 -20 -15 -10 -5

1000

700 PEACt/£5 700

500 500

300 300

200 200

V) IOO 1 100 ,_

� 70 70

50

30

20

10 - JO - 25 -20 - 15 - 10 TEMPERA TURE r·c )

50

30

20

10 - 5 0

Fig. 2 : For peaches is "first detectable dif­ference" at higher freezer temperatu­res connected with color defects, at lower temperatures with flavor defects.

Testing within the temperature range from a ppr. -20° to -5 ° C the Western Regional Laboratory found straight line keeping quality characteristics for many vegetables, fruits, and for some poultry products. This seemed to be contrary to what was generally believed to be the case for other types of products as for instance fish and meat. In fact, even the most recent TTT-relation studies carried out on these products seem to indicate that the keeping quality characteristics mostly are curves that flatten out as the temperature goes down. In other words, that the influence of a certain temperature decrease is smaller the lower the temperature.

In some extreme cases, as bacon, one may even find that the keeping quality characteris­tic has a minimum, e.g. that within a certain temperature range the effect of getting a more and more concentrated solution of salts in the product may be more important than the effect in the opposite direction of lowering the temperature. Also with fruits the straight line keeping characteristics will not seem to be true always. For peaches for instance it has been found that the reaction governing the development of off-colour is much more tem­perature dependent - has a much higher Q10 - than that governing the development of off-flavour. By a temperature of -10° C it is the off-colour that is first detectable, whereas at a temperature of -20° C the off-flavour is first detected. Thus, if first detectable acceptability loss is plotted for this product, one would not get a straight line.

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Straight lines can b e expected if one reaction alone is of importance in determining the first detectable loss of quality. Of course, this may often be the case, but even one reaction may not give a straight line characteristic in the temperature range concerning frozen food processing and handling. In this range considerable physical and chemical changes take place as we go from the liquid to the solid state and this may effect the size of the temperature quotient, Q10, of the reaction.

When we have two or more reactions of importance to loss of quality, as seen in the example above, then the reaction with the smaller Q10 will always become dominant at low temperatures, as it is inactivated less when the temperature goes down. So even with constant temperature quotients of the reactions - an assumption which as mentioned is questionable - the keeping quality characteristics will flatten out at low temperatures towards a straight line with a slope corresponding to the smallest Q10•

Therefore, if in the course of some of our investigations we do not get straight line characteristics, this is only to be expected. Also, it should be remembered that as long as we do have the keeping quality characteristics - whether straight line or not - for the temperature range in which we are interested, we may still use our knowledge of time­temperature relationship since the methods of calculations are not just applicable for straight lines.

METHODS OF DETERMINATION OF LOSS IN QUALITY

The keeping quality characteristics are the basis for all our calculations on the in­fluence of various time-temperature fates during the storage and distribution of frozen foods. Therefore, it is appropriate to consider the ways these are obtained and how we shall be able to correlate results from different investigations.

First there is the question of raw materials. These, of course, should vary as little as possible, so it seems better to use products kept at low temperatures as reference samples than using fresh products when the testing is to be done. In our investigations at The Danish Meat Research Institute we have used -40° C for reference samples. As we were concerned with the temperature range -5 to -20° C, this was thought to be sufficient, especially as Q10's often are very high for frozen products. Also, if extreme low tempera­tures were to be used, some doubt may exist, whether these in themselves in some cases may introduce changes in the reference samples.

Secondly, one need consider what criteria are to be used as a measure for loss in quality. It is, of course, much easier to correlate the various investigations where chemical or physical changes can be used for the determination than when one has to rely on organo­leptic testing. Unfortunately - though much research has been done in this field - such measurements very often are not in correlation with those found by organoleptic testing. As foods are meant to be eaten, the evaluation of taste, appearance, and consistency still seem to be what one must use to determine loss of quality.

DETERMINATION OF "FIRST DETECTABLE DIFFERENCE"

In most of the TTT-investigations the keeping quality characteristics have been established by organoleptic determination of the so-called "first detectable difference". When using the term "first detectable difference" it should be kept in mind that this does not refer to some specific change in the product. The first detectable difference depends on the panel - the people who do the tasting. These, of course, differ a lot -some use people selected as especially apt for this kind of work, some have been trained in objective testing or in the evaluating of a specific product. One only has to think of wine­or coffee tasters as compared to ordinary consumers to realize the very considerable differ­ence that there may be in the term "first detectable difference".

Some testing is done by the use of scoring systems. A certain score alone has no real meaning not only because of differences in panels and scoring systems but also because of the wellknown fact that the level differs from day to day. Very often it may be noticed that the scoring level goes down the more the tasters get used to a product.

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Testing by the simple triangle method of just picking the odd one out may also involve differences in what is understood by the term "first detectable". It might in one place be used when % in another when % of the panel has picked the right one.

In order to get the term "the first detectable" more comparable than at present, it seems that all investigators should make use of statistical analyses and the term only be used on the basis thereof.

Within the last couple of years we have carried out time-temperature tolerance investi­gations on different Danish commercial frozen meat products. In these investigations our taste panels use a scoring system. We find our first detectable quality loss by statistical analyses of the results. We defined it as the time when a trained panel of 9 members, at a probability level of 0.95, would find the stored product, with a significance of 0.95, as being less acceptable than the reference sample. In our definition we use less acceptable. Our investigations indicate that for some products, as raw meat, cuts, the score for consistency may go up significantly during the first part of storage.

RELATING "FIRST DETECT ABLE DIFFERENCE" TO CONSUMER ACCEPTANCE

A question of importance to the frozen industry is: How do these first detectable differ­ences in quality compare with the practical storage life of the various products. In other words, the correlation between days of storage given by the keeping quality characteristics and a storage life still rendering a product acceptable for the consumer. We shall, of course, never be able to give a definite answer to this question. As the times for first detectable loss in quality are connected with a rather big variance, this is even more so the case when one speaks about loss of quality acceptable to the consumer.

Investigations on the first detectable loss in quality have now appeared for many prod­ucts and many more will appear within the coming years. Knowing that no complete an­swer can be given, but considering the importance of the use of these data by food pro­cessors and legislators we shall try to mention a few considerations regarding the question of consumer acceptability.

Our purpose in using a scoring system in the above-mentioned investigations of Danish commercial frozen meat products was to see how the scoring went down when products were stored for longer periods than those where we actually had detected loss in quality. Agreeing that the scoring levels do vary some, we still ventured to use the score of -2 (minor defects) as the level of consumer acceptability.

Products scored at this level were still just acceptable to the panel members who, however, must be considered very critical.

In using a method of storing the products 10-20-40-80 etc. days at the temperature we tested and keeping these samples at -40° C together with the reference samples until organoleptical testing, we were able to hold the testing sessions closely together. We, actually, found that in this way we decreased the variance in the scoring level - by comparing the score of our reference samples - to less than % point.

In table 1 is shown for 13 commercial products the time (days) for loss of flavour corresponding to a decrease in scoring of 1 point. Statistical analyses of the results for these products showed that the first significant (O. 95) loss in flavour varied from a decrease in scoring of 0.8 to 1 . 1 points dependent on the variance within the product (reference samples) itself. The data given in table 1 may thus give some indication of the relation between first detectable loss and consumer acceptability when compared with table 2 showing time (days) for loss of flavour to a score of -2 points. As will be seen from the results this relation varies from about 2 to 4. The products starting at a high level of scoring, i. e. top quality products, will, of course, have the bigger differences between the time when loss in quality was just detectable and the time when the development of off­flavour will no longer be acceptable to a critical taste panel.

60

tu Q: 0 (.} V) Q: 0 ;::,. -q: __, 4.

PORK CHOPS 2

0 - -9 27days. First detectable difference.

- I - 2

- 3 IO 20 40 80

2

0 - I

- 2 10 20 40 80 160

0 - I

320

320

P-9

+ 5°C

days

+ 10°C (34 7days)

days

+ 2 0°C - 2 _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ _ !!_61.!!<!_ar_s _

10 20 40 80 160 320 days

Fig. 3 : The basis for determining "first detectable difference" and "consumer acceptance limit" for pork drops (compare product no. 6 in tables r and 2).

Table 1. Time (days) for a decrease in flavour scoring of I point*)

Product Temperature -5° C -I0°C -20° c

I. Steakburgers I6 210 250 2. Rumpsteaks 20 230 > 320 3. Calf liver I 24 52 I70 4. Hamburgers I7 64 I35 5. Ground beef SI 120 > 320 6. Pork chops 27 IOS I80 7. Ground pork 41 79 230 8. Pork liver 1 1 6 100 I70 9. Calf liver II 60 I65 > 320

10. Pork cooked with rice and curry sauce. The meat > 80 > I60 > 320 IO. Pork cooked with rice and curry sauce. The Sauce > 80 > I60 > 320 I 1. Pork liver cooked with cream sauce. The meat 56 75 230 I 1. Pork liver cooked with cream sauce. The sauce 65 I30 (420) I2. Hamburger cooked with onions and sauce. The

meat 70 > I60 > 320 I2. Hamburgers cooked with onions and sauce. The

sauce > 80 > I60 > 320 I3. Pork sausage 22 110 240

*) The first detectable loss in flavour varied from a decrease in scoring of 0.8 to I .I points for the different products.

Maximum storage time tested was 160 days at -5° C and 320 days at -I0° and -20°, except for products 10, 1 1, and 12 : 80 days at -5°, 160 days at -I0° and 320 days at -20°C.

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Packaging : Product 1 : double-coated carton + heat-sealed double-coated paper overwrap; products 2, 3, 4, 6, 7, 8, 9, and 13 : innerwrap or bag + coated sealed carton; product 5 : vacuum-packed in plastic tube with low water vapour and oxygen permeability, tube shrunk; products 10, 11, and 12 : coated bag + sauce + coated sealed carton.

As indicated in Table 2 this state of development of really important off-flavours was never reached for many products during our tests at the lower temperatures of -10 and -20° C. In this case we used straight line extension on a diagram plotting flavour scores against days of storage. Actually, indications are that at the lower temperature levels the development of off-flavour is slower than stipulated in this way. An explanation may be that the reaction responsible for first detectable difference (for instance loss of moisture on the surface due to evaporation and the forming of ice crystals inside the package) is not the same as that being responsible for the further developing ofloss in quality (for instance development of rancidity).

Table 2. Time (days) for a decrease in flavour scoring to a score of -2*)

Product Temperature

-5° C -10° C -20° C

1 . Steak burgers 57 ((650)) ((750)) 2. Rumpsteaks 72 ((690)) no decrease 3. Calf liver I 60 135 (340) 4. Hamburgers 56 (242) (391) 5. Ground beef (163) ((396)) no decrease 6. Pork chops 86 (347) ((610)) 7. Ground pork (106) 253 ((645)) 8. Pork liver (197) 220 (418) 9. Calf liver II 96 297 no decrease

10. Pork cooked with rice and curry sauce. The meat no decrease no decrease no decrease

10. Pork cooked with rice and curry sauce. The sauce no decrease no decrease no decrease

1 1 . Pork liver cooked with cream sauce. The meat ((185)) (180) (460) 1 1 . Pork liver cooked with cream sauce. The sauce ((253)) ((442)) ((1300)) 12. Hamburgers cooked with onions and sauce

The meat ((252)) no decrease no decrease 12. Hamburgers cooked with onions and sauce.

The sauce no decrease no decrease no decrease

13. Pork sausage 57 275 (480)

*)A figure in brackets means that the decrease to a score of -2 was not actually found during the testing but was estimated by linear extrapolation on a diagram plotting fla­vour scores against days of storage. Double bracket means that not even a decrease to a score of -1 has been established. One must assume that the actual keeping times are longer than those stated in the brackets and may be much longer than the figures in double bracket indicate.

"No decrease" means no decrease found after 80 days at -5, after 160 days at -10, and after 320 days at -20° C.

In our investigations we also scored appearance in the frozen state and appearance and consistency of the prepared product. In some cases (raw meat cuts) during storage at -5 ° C, the appearance of freezer bum became significant at about the same time as the loss of flavour. For a few other products the growth of moulds stopped the tests altogether at-5° C, so this temperature cannot be recommended for accelerated testing. Tempera­tures of -7 to -9° C, where such developments seem to be stopped, are most appropriate for this purpose. Apart from this, the detection ofloss in flavour always became significant

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first. For consistence - as mentioned before - we even at times found statistical signi­ficance in that the quality got better, but for most products the loss in consistency followed the development of off-flavour, only at a slower rate. The scoring of appearance (off­colour, etc.) of the prepared products usually did not change significantly during the investigations.

It will be seen that no fixed factor can be used in relating consumer acceptability to first detectable loss in quality. In using a factor of 2 the producers should be on the safe side in assuring that their products still are of.a good quality. But, especially at a low tempera­ture as -20° C indications are that even much longer storage periods than that will keep the quality at a level where the ordinary consumer will not be able to detect any difference from the just frozen products.

DIFFERENCES IN STORAGE LIFE AS RELATED TO PRODUCTS, PROCESS­

ING, AND PACKAGING

Having now mentioned some problems concerning how to obtain and relate values of keeping times let us consider how such results apply to all the many different types of frozen foods that are offered the consumer. The big variance in the raw products them­selves and in the ways these are processed and packed are of the utmost importance. Without specifying these factors of product, processing, and packaging we shall not be able to give any estimate of the storage life.

A few examples will illustrate this :

Products : variations in the fatty tissues of meat are of influence on the rate of deterio­ration. A big variance of for instance ascorbic acid content has been found not only be­tween, but also within the same variety of pea:s.

These investigations also showed an almost fourfold variance in the rates of chlorophyll loss.

Processing : the variance mentioned in the last example mentioned above might also be because of difference in the blanching process. A storage life of about 250 days at -10°C for cut up raw chicken as compared to 25 days for fried cut up chicken show the enormous effect that heat treatments may have. Even differences in hygienic conditions during the slicing of raw meat may alter the storage life considerably. In table 1 one will notice the keeping times of raw calf liver. For the product marked II the time of first detectable quality loss is about 3 times longer than of that marked I. The only significant difference between these products seems to be in the bacterial contamination, the one marked I having a total population 300 times bigger than that No. II. After freezing, the relation of viable microorganisms was 20 to 1. Some of the processes leading to deterioration may have been accelerated by enzymes produced by the microorganisms.

Packaging : This may be the most important factor in determining storage life. Studies made by the Western Regional Laboratories in America of commercially packed raw chicken showed that by simply improving the package one was able to increase the storage life just as much as by lowering storage temperature from -7 to -18° C. The influence of packaging can also be seen from the results in table I. The ground beef (5) is actually the same meat as that used for the hamburgers ( 4). During production some of the ham­burger meat was taken from the packing machine and packed under vacuum into a tube of a plastic film with a low permeability to water vapour, and oxygen. At -5 and 10° C this resulted in a storage life 2 to 3 times longer than that of the hamburgers packed in the ordinary coated carton. Indications are that the difference in storage life will be even bigger (5 to 8 times) at storage at -20° C, and that by excluding the possibility of loss of moisture and penetration of oxygen one may altogether eliminate the process that before was the one determining the first detectable deterioration. Also when comparing the results in table I it will be seen that even though, as mentioned previously, many investigations show that meat that has been heated is more susceptible to quality loss during frozen storage than raw, the cooked hamburgers (12), "packed" tight in the frozen sauce kept 4 or more times

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better than the raw, but not tightly packed, Hamburgers (4) and Steakburgers (1). An even more striking example may be seen when comparing storage life of pork chops (6) and ground pork (7) to that of pork cooked with rice and gravy (10). Here, as seen from table 1, no loss in the quality of the latter product could be detected during the investiga­tions.

CONCLUSION

The many problems and variances mentioned here may tempt somebody to think that the use of systematic calculations will not help in solving the questions about quality and storage life of frozen foods. This is not so; more facts are still needed, but our knowledge is growing just as the frozen food industry.

In the future, we still will not be able to tell exactly how the quality of a new product will be, or how long it will keep, but the use of accelerated storage tests together with the possession of more systematically determined keeping quality characteristics will enable us to give close estimates of the influence of any changes in product, processing, packaging, storage, or distribution.

Most important, these new investigations may help in stimulating the interest of producing more suitable products to begin with ; just as the time-temperature investiga­tions have already helped us to better our storage and distribution facilities.

SUMMARY OF THE DISCUSSION (Papers P-2 + P-9)

J. Gutschmidt, Germany : I agree with Mr. Jul that it is advantageous to accelerate storage changes in research work by using a higher storage temperature as usually chosen by the freezing industry and the results of the TTT-studies demonstrate the possiblity of working so.

But I just like to know if -S° C is not too high for getting comparable results. Ex­periments in our institute proved, in accordance with the research work carried out at the Research Laboratory in Albany, a heavy growth of psychrophilic strains of yeast and fungi at this temperature. Of course, they have a long lag period, but if a storage time of half a year or longer is used the quality loss of vegetable, e. g. string beans, may be fixed by yeast colonies developed. Even at -7,S°C quality of vegetable and fruit may be influenced by yeast and fungi growing during long storage time.

Because the time of storage experiments is shortened by using the high storage tem­perature a spoilage of the frozen food will not occur. But it may be possible that the flavor of the food is influenced by the growing of microorganisms especially by the excretion of high active enzymes. It is known that fungi in a small amount, microscopic­ally not detectable, can influence the flavor of food, e. g. milk and bread. That may be true also with animal food that mainly was used in the Danish storage experiments, although the growing of bacteria is prevented at the temperature of -S°C.

M. Jul, Denmark : I agree that -S°C was a pretty high temperature. It seemed that microbiological growth did not interfere with the test for "first detectable decrease in quality", but it had at times interfered with determination of "no longer completely acceptable". I recommend -7° C as the highest (warmest) test temperature. In the Danish experiments attempts had been made to find the absolute highest test temperature for the accelerated test in order to speed them up.

J. A. Peters, U.S.A. : At the Bureau of Commercial Fisheries Technological Laborat­ory in Gloucester, Mass., we have been working on the Time-Temperature-Tolerance of frozen fish for about three years. Part of this work was Dr. Lane's literature survey mentioned earlier by Mr. Klose. Large discrepancies in the reported storage life for various species were found. Very often the authors do not report the pre-freezing handling history of the fish, the packaging materials used, or the relative humidity in the cold storage room.

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The importance of pre-freezing handling conditions was strikingly illustrated in the results of our work on whiting ( Merluccius bilinearis) . During storage in ice or refrig­erated sea water the fish did not show rapid quality loss, but in frozen storage quality loss was very rapid for fish held about seven days in ice or about eleven days in refrig­erated sea water. Therefore, since fish which appear to be of acceptable quality initially may have poor storage characteristics it is very important in Time-Temperature-Tol­erance work that the pre-freezing history of the fish be known and controlled.

In other phases of this project we have just completed a survey of times and temperat­ures in the distribution cycle of frozen fish, are studying various dehydrogenase enzyme systems in fish flesh for use as objective indices of quality, and will begin shortly a comprehensive Time-Temperature-study on Ocean perch (Sebastes marinus), in which fish about 24 hours out of the water will be held well iced for 2 to 3 weeks. Samples will be frozen at three day intervals and stored at five different temperatures. This will be the first of a series of tests covering species of major importance to the United States fisheries.

J. Moreno Calvo, Spain : Since integration function of quality losses in frozen foods storage must be really straight connected with the products' nature, and if quality losses may be dimensionated as an energy time product, then it will be most comprehensive if we consider, when it is possible, the deterioration rate not only as a function of tem­perature but also of specific heat of the product too in the frozen state. In such a way the comparison between quality losses integrations of the different products may be recommended specially as the most appropriated.

M. Jul, Denmark : I suggest that an attempt be made in the meeting to list in summary form the requirements in various countries to freezing, storage, transport, and retail storage conditions for frozen foods.

J. Gutschmidt, Germany : In Germany "Recommendations for deep-frozen food" have been published in 1961 under promotion of the Federal Government by the "Bund fuer Lebensmittelrecht und -kunde e. V.", Bonn. The Recommendations delivered by this confederation are fundamentally for forming council's opinion. According to these Recommendations food for sale under the label ,,deep-frozen food" has to be

1. frozen at good commercial practice up to a temperature in heat centre of -15°C (average temperature -18°C or lower) ;

2. transported not above -18° C (a temporary temperature of -15°C is permitted) ; 3. stored at -18°C or lower (in retail storage the temperature may rise 3°C, but

not higher than -15°C for 24 h); 4. wrapped in packaging materials that are able to protect the food during usual

storage time against influence of atmosphere, microorganisms, desiccation, and transfer of flavor;

5. labeled according to given directions. In these Recommendations ice cream is not included.

M. Kondrup, Denmark : In Denmark official regulations on the handling of frozen foods, in general, are in the draft stage. A committee of government officials has pro­duced a draft based more or less on a document published by a special export committee under the Danish Academy of the Technical Sciences.

The present draft demands -18°C (or below) in producers' warehouse, -25°C in other warehouses, and -18° C in retail sales cabinets. It also asks for -18°C during transportation. All temperatures are in principle temperatures of the produce.

The draft contains many details on packing, labeling, and among others also on pro­cessing for some foods, and on retail back-storage rooms.

Several of the proposed rules have provoked much discussion, among others, because of the natural wish of the industry for flexibility which not always goes well together with the wish of the administration for regulations that are manageable administratively and inspectionwise.

However, these discussions have proved very sound in respect of clarifying the whole subject.

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H. Y. Boersma, Netherlands : There are no government regulation on freezing rates and storage temperatures.

There is a regulation made up by the "Poultry Marketing Board". Freezing time for poultry :::;; 16 hours.

Storage temperature ;::::: -15°C

Transport temperature ;::::: -15° C

C. F. E. von Sydow, Sweden : Frozen foods are treated as a general food item in the Swedish food law of 1952. A new law is under preparation.

Recommendations are issued by an officially controlled, industrial office of frozen foods and are as follows :

Factory storage -30° C Wholesale storage -25°C Transportation < -20° C Cabinets -18 to -20° C

Some of these recommendations are also issued by some governmental offices.

R. Ulrich, France : II n'y a pas a ma connaissance de reglement en France concemant les temperatures de congelation et d'entreposage des denrees surgelees, mais nous devons travailler sur ce sujet. Est-il exact que pour certaines denrees des temperatures sensiblement plus elevees que -18°C sont recommandees ?

V. !bl, Czechoslovakia : Les temperatures recommendees en Tchecoslovaquie : entre­posage : -18°C; transport : -18°C; detail : -18°C.

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Freeze- D ry i n g C ryo-d essiccat i o n

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Basic Principles of Lyophilization (Freeze-Drying)

Principes fondamentaux de la lyophilisation (Cryodessiccation)

Prof. LOUIS REY Universite de Dijon, Faculte des Sciences, 2, Boulevard Gabriel, Dijon, France

SOMMA/RE. La plupart des produits naturels ou d'origine biologique peuvent etre con­serves par lyophilisation ( cryodessiccation) . Ce procede comprend essentiellement deux phases successives: une congelation prealable du produit a traiter, puis une deshydratation par subli­mation de la glace sous vide. Le produit ainsi desseche peut ensuite etre conserve sans precau­tions speciales a la temperature ambiante, a condition toutefois qu'il soit place dans un reci­pient etanche sous vide OU sous atmosphere inerte. Beaucoup de produits pharmaceutiques, les tissus d'origine animale ou humaine, de meme que les vaccins vivants ou les serums thera­peutiques sont prepares ainsi.

Recemment, une large gamme de debouches nouveaux s'est offerte pour la lyophilisation. Les denrees alimentaires, les dechets nucteaires et les produits chimiques peuvent etre traites avec succes par cette methode.

Apres une breve revue generate de l' ensemble du procede, l' auteur etudie chacune de ces dijf erentes etapes; preparation du produit, conge/ation, sublimation OU dessiccation primaire, desorption ou dessiccation secondaire, conditionnement final et rehydratation. Dans chaque cas, les principaux aspects Jondamentaux et les differentes realisations techniques sont etudies et discutes.

INTRODUCTION

It is well known to all those who are dealing with biological or natural products that, thanks to their high water content, they are susceptible to alter during harvesting, transportation, storage and marketing. For that reason, it is necessary to inactivate water in order to ensure long-term preservation.

Dehydration obtained by heat and vacuum drying has proved very useful indeed in many circumstances. However, most products cannot stand direct evaporation of their water content as they are highly sensitive to a progressive increase of the concentration of solutes and mainly mineral salts.

Refrigeration and deep freezing have been resorted to for many years to preserve alterable substances and essentially foodstuffs. By decreasing the metabolic rate inside the products or separating out water in the form of ice, those processes contribute to keeping the material in an adequate and stable condition. Despite their great interest they, however, present some difficulties - one of them being the obligation to maintain the low temperature all along the commercial life of the substance, which can prove difficult under certain circumstances.

Lyophilization, or freeze-drying, is another technique which is designed to present the advantages of both freezing and dehydration. Lyophilization is essentially a two-stage process in which the material to be preserved is

- first quickly frozen to a low temperature level, - then dried by sublimation of ice directly from its solid state in a vacuum chamber. The substance can then be stored indefinitely under its dry form at room temperature,

without any special requirement, provided it is correctly processed and packed free from watervapour and atmospheric oxygen. In that state it is easy to reconstitute the product

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by addition of the original water content, the water being readily adsorbed by the "lyo­philic" product. It is that property which gave the name to the process itself.

I - THE RANGE OF THE MAIN APPLICATIONS OF FREEZE-DRYING

Before going into the details of the different stages of the process, let us first have a glance at the main fields in which freeze-drying did develop in the past twenty years.

1. - Biological products Historically this has been the first achievement of freeze-drying which was applied to

the preservation of bacterial cultures [Flosdorf and Mudd, 3] and human plasma [Greaves, 6]. Three main groups of products are to be considered :

= biological extracts and biochemical derivatives These are highly complex products in which some specific biological functions must be

preserved : tissular and cellular extracts - hormones - vitamines - enzymes - anti­biotics - therapeutic sera - human plasma . . • Most of them are prepared by the pharmaceutic industries and have been in common use on the market for more than 15 years. Their stability is very high : indeed, freeze-drying is the official technique for the preparation of international standards of biological materials [2.].

= Human and animal tissues for orthopaedic and reconstructive surgery and, more gene­rally, for tissue-banking purposes

All those tissues in which the mechanical function plays a prominent part and which can be used in the human body for grafting purposes, although they are not viable, can be preserved by freeze-drying. Indeed, most tissue banks prepare arteries, bones, cartilages, corneas, fasciae, dura and skin in frozen and dried form [Hyatt, 10]. They are easy to store, ship and handle.

= Living biological preparations Under certain circumstances, as far as lower organisms are concerned, life can be

preserved by lyophilization and the technique is commonly used to store bacterial sus­pensions [Fry and Greaves, 4], viruses [Lepine, 1 1], yeasts [Mazur, 12] and most living vaccines. So far, it has not been possible to keep more organised cells and especially mammalian tissues in the viable state [Greaves, 8].

2. - Foodstuffs Recently applied to the foodstuff industry, freeze-drying has shown that many an

alimentary product can be easily prepared and stored by that technique [Rey, 19]. When compared to conventional methods, the main advantages of freeze-drying are the follow­ing :

- stability of the product which can be stored without special care, - perfect preservation of flavour, taste and the original nutritional value, - an important reduction in weight and volume due to high dehydration. Frozen and dried foods are high quality products and very convenient to use. They help

in widening the market, both in space and time, of seasonal products coming from restrict­ed geographical areas.

Three kinds of products can be found [Rey, 20] : = Structural foods : In this group are gathered those products in which a special texture has to be kept

in order to maintain their original organoleptic properties : meat, fish, fruit, vegetables • . . = Powdered foods : The essential point here is the correct preservation of aroma and taste. These types

of products are : coffee, tea, fruit juices, fruit pulps or extracts, mashed potatoes . . .

= Special foods, Dietary products and Baby foods : In this third group, besides an adequate preservation of texture, aroma and taste, well

defined biological properties must be kept intact if we wish them to fulfil their nutritional and physiological requirements. Baby milk is a good example of it. In this case, freeze­drying is quite an unmatchable technique and, as we saw previously, it is in fact the official method for the preservation of the standards of biologically active compounds.

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3. - Nuclear wastes A very interesting application of freeze-drying is the treatment of nuclear wastes and

mainly of medium and low-activity waters and muds rejected by atomic centers. In recent experiments made in collaboration with the French Commissariat for Atomic Energy and the Leybold Society, we have shown that it is possible to achieve very high decontamina­tion factors and transform active waters into dry powders which can be incorporated in permanent glass or embedded in concrete blocks.

4. - Chemical products It is quite certain that new developments of the freeze-drying industry will take place

in the chemical field. Isolation, concentration and final preparation of highly thermolabile compounds can be easily foreseen in view of the recent technological advances.

Despite its taxonomic aspect I have found it useful briefly to summarize the main develop­ments of freeze-drying essentially to show that it is nowadays a fully developped technique which has been greatly diversified in the industrial field.

II - THE DIFFERENT STAGES OF THE FREEZE-DRYING PROCESS

In this general outlook on the freeze-drying process, it is impossible to get into all the fundamental and technological details of the method, so I shall restrict myself to general considerations.

At first sight, we ought to consider the process from its very beginning, that is to say the preparation of the product to be dried, up to its ultimate stage, the rehydration of the preserved material. Six different stages are then to be studied. They are outlined in Fig. 1.

under vocvum

oddit/v,u S u b l i m o t i o n of ice ls.othermic desorption ft pN!c1p1tar1on � ir::t:' intH'sliliol / E£il 'Zormalion cryst. woltr vapour flow

b:J -� - �-� _ �LA_A _ f,f/r-olio11_cot1Cclllrorion ffi ffl ffl

dry intrt gas waler original liquid 1. srtaling A t1-8 - 8 - t:l -� � � / cr11u·9y. input

J PN!:Jl?'''°" -:..__ +f IIFr::.t19! 111Prtmory dryirig I N St!Condory drylng I 'Jl Packaging I

-· ��-�'\/\/\/j \ ' Storage j Jl1 Rt'hydration

20 J t h 11olcr

-� ,; :::.""�""="<::...' ____ ....., -<o A'� \1 i

Fig. 1 6-stages of freeze-drying and rehydration

1. - Preparation of the product = First, and exactly as in the case of cold storage, the product has to be a choice one

and requires to be processed immediately, without the least delay. Thus it will be possible for the product to keep all its original qualities.

= In the second place, the product has to be prepared so as to fit the operational con­ditions of the working plant. Essentially, it must present a large surface and a reduced thickness in order to allow an easy drying. Moreover it is advisable to distribute it in homogeneous batches of the same shape, volume and weight.

= Thirdly, it is sometimes necessary to pretreat the raw material before freezing, in order to facilitate the drying. Different methods may be used :

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mechanical pretreatment : bulky material must be ground or sliced into a regular size juices have to be prepared and, if necessary, filtrated or clarified

- physical pretreatment : diluted solutions can be preconcentrated before freezing some foodstuffs have to be steamed, bleached or precooked beforehand

- chemical pretreatment : chemical precipitation or floculation of the solutions is advisable in the chemical and

nuclear fields additives have to be included in certain media to increase the dry weight contents of the solution and help in producing an homo­

geneous stiff cake in the course of sublimation (hormones - vitamines - pharmaceuti­cals . . . )

to protect cells against the deleterious effects of freezing ( dextran, polyvinylpyrrolidone or albumin for bacterias and viruses), to buffer their residual moisture after drying (glu­cose, saccharose) or neutralize some carbonyl groups (sodium glutamate) as it has been shown by Professor Greaves [7].

2. - Freezing Freezing must be performed as correctly as possible so as to avoid any denaturation

of the product. Indeed, it is quite easy to understand that there would be no point in drying with great care a product which has been partially or totally destroyed by previous inadequate freezing.

(a)

A

_ Jo : su p e r c o oled eutectic fluid

c Log R

inlerslilial . . b - - - - -/- cryslol/1zot1on - - -Q b I , ..,

' a I I I

_ 40 ' Tes

I \ 1 i I '

Ice format/on

- - - - - < T�C

1-20 0 -

7'im= -21,'6=eutectic thawing

Variation of th e electric resistance in the course of fre e z ing

B

_ 4() : (aft e r direct cooling )

or _ 22:

{ b}

(after cooling lo _ ,m :and rewarming) c ryslollized e utectic mixture

Fig. 2 . Crystallization of an aqueous solution of sodium chloride (microscopical examination)

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= General considerations : - structure : freezing must be effected quickly and at low temperature in order to

achieve a regular crystallization with minute ice crystals. However, one must take care not to go too fast for this could disrupt its structure when the central part freezes out.

- final temperature : freezing ought to be complete. As we have shown previously [Rey, 15], it is necessary not only to induce ice formation but also to get to the point where all the interstitial fluids which lie in between the ice crystals have crystallized out or at least hardened (Fig. 2 a and b). Electric measurements as well as thermodynamic observations by differential thermal analysis [Rey, 16] have emphasized the necessity to go down to sufficiently low temperatures in order to rupture the metastable equilibria which might occur in the course of freezing. We have called [16] that temperature : maximum temperature of complete solidification (Tes). It will vary from one product to another and can lie between -30 and -80° C. Sometimes, to get rid of vitreous forma­tions which are subsequently apt to melt and foam under vacuum, it is necessary to apply thermal treatment to the product [16]. The substance has to be cooled to a low temperatu­re to achieve the glass formation, then it is rewarmed slowly to induce devitrification and cooled back to the right temperature. In some instances, thermal treatment can be applied by postfreezing the material by means of liquid gases such as liquid nitrogen after previous conventional freezing.

- thermal stability : when frozen the product has to be kept at a sufficiently low level so as to avoid any partial melting. Actually, differential thermal analysis shows that, above a certain range of temperatures, interstitial melting can occur. This will be eutectic mel­ting if we are dealing with pure simple chemicals or, more simply, incipient melting for c omplex systems. We have called [16] that temperature : minimum temperature of

incipient melting (Tim) and, again, it must be predetermined accurately for each kind of product.

= Technical aspects - Many methods can be resorted to for freezing and we shall just list them :

vacuum or snap freezing contact freezing on cooling plates freezing by a blast of cold air combined freezing : contact and ventilation freezing by immersion in a cooling bath

- Post-treatment mechanical post-treatment : products which are too thick have to be sliced or ground

when frozen. For meats it is advisable to slice them perpendicularly to the axis of the fi­bres so that sublimation may proceed more readily. When dealing with liquid products, they can be frozen first in regular layers or in blocks which are then ground and handled as solid material.

,thermal post-treatment : we have already mentioned thermal treatment which has to be used when glassy structures are likely to appear (fruit juices [22], vitamines B, sugar solutions . . . ).

3. - Sublimation (Primary drying) = General considerations : When it is thoroughly frozen and processed, the material to be dried is placed under

vacuum. Then heating is applied so that sublimation of ice may take place. The water vapour escaping from the ice crystals is extracted from the apparatus either by conden­sation on a cold trap (Fig. 3) or by adsorption on a chemical adsorbant, or else by direct removal through the pumping device. Sublimation being essentially an endothermic phenomenon, heat must be supplied during the whole period. Then drying proceeds at a regular rate and, as time goes on, the drying boundary sinks in the frozen material and moves inwards. If the material is a liquid, frozen in a tray, the sublimation interface will go down right to the very bottom of the tray and will leave a porous, rigid cake. If now the material is a solid and has been frozen in a block, the drying boundary will spread all around its surface and progressively move to the central zone as the ice nucleus recedes.

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D ry in g cham ber

heating/ (and cooling) _

shelves

� ---\--r::::::j;:::�·(JJ:::::i: . ·

C o n de n s e r

J (eventually, closing valve)

ma terial to be dri e d (sol-i_,ds, liquids i n vials o r in bulk)

Fig. 3 Scheme of an apparatus for freeze-drying

- sublimation temperature :

ice coate d cold trap

refrigerant

In consequence of what has been said previously in relation to freezing, it is quite obvious that the temperature of the frozen material has to be controlled accurately in the course of sublimation. Actually, if one wants to avoid interstitial melting and the risk of direct evaporation or even foaming of the interstitial fluids under vacuum, the ice has to be kept constantly at a temperature equal or inferior to the minimum temperature of incipient melting. This has to be strictly observed for delicate biologicals. In the case of foodstuffs, a certain degree of interstitial melting can be allowed and thus the product can be processed at a higher temperature.

All the operations must then be carried out so as to get and maintain a given tempera­ture, or better a given structure inside the frozen mass during sublimation, and there lies the main difficulty.

- mass and heat trans! ers : Actually it can be understood that, for economical and practical reasons, the drying

must be operated at the optimum velocity and that the temperature of the ice has to be as close as possible to the safety limit to enable the sublimation to proceed at maximum speed. The problem is then essentially to drag the water vapour out of the sublimation interface as quickly as possible and to supply heat to the same spot in an efficient way to allow sublimation to go on at a constant rate. In other words, we have to solve two different problems, a water vapour transfer (mass transfer) and an energy input (heat transfer) _

mass transfer : If we look at Figs. 1 and 4, we can see that, as the sublimation interface sinks in the product, water vapour has to go through an increasing thickness of dry mate­rial and the resistance to gas flow of that plug is enormous. It is quite evident, then, that there will be a limit to the permeability of the dry shell and that, depending upon its structure and thickness, it will allow a maximum flow rate of water vapour. Very elegant experiments, realized in that field by Doctor Oetjen and Doctor Hackenberg [13.], have shown that, in order to allow a given water vapour transfer, a certain pressure drop is necessary between the ice layer and the surface of the dry cake. That pressure drop must be as large as possible ; however, it is restricted to a given value because of both the maximum temperature allowed for sublimation and the design and price of the pumping equipment.

heat transfer : If the mass transfer has to be restricted and enclosed in certain limits, it is clear that there will be no need to provide more heat than can be used on the spot of sublimation. Heat transfer will then have to be adapted to fit constantly the possibili­ties of water vapour removal. However, the problem is more difficult than it appears because it is not only sufficient to supply to the sublimation interface the required amount of energy, but also, to avoid the destruction of the underlying or overlapping material. Actually, we have to provide heat to a moving surface which is enclosed between a dry and a frozen layer (Fig. 4).

72

dry material

frozen mass

inlerslilial me/ling

(a ) L i q u i d

interface

heal lrans�7 (energy inpu t)

surface charring

I ( b ) S olid

Fig. 4 Heat and mass transfer during freeze-drying

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-in the first case, (Fig. 4 (a)), liquid frozen in a tray or in a glass vial shows the situation. Heat has to be transferred to the boundary throughout the frozen mass. As heat has to flow along a thermal gradient, although the temperature could be correct at the subli­mation interphase, it might be too high in the frozen mass and especially near its bottom. Overheating and interstitial melting must then be avoided. That is why an adequate control of the structure of the frozen material has to be used during sublimation. This can be done by following and monitoring the temperature of the ice at a constant low level ( < T1 m), either by thermal determination or by water vapour pressure measurements. However, temperature measurements do not always prove to be an adequate reflexion of the struc­tural state. This is why we have been using electric measurements for that purpose [17). A high electric resistance indicates a state of utter rigidity while an abrupt decrease of the resistivity shows that interstitial melting is taking place. It is then easy to monitor the heating system according to the indications of an electric probe inserted in the product and maintain the structure of the frozen mass at its optimum state [14, 18].

- another case is the one of solid material (Fig. 4 (b )). Here, as we have already seen, the drying boundary encloses a nucleus of frozen material and the only way to provide heat is through the dry shell. Thus, the heat input will be limited by two factors.

- the heat conductivity of the dry mass in the course of sublimation ; - the maximum tolerable surface temperature of the dry substance. Indeed, if we

want to avoid denaturation and even charring, it is necessary to restrict the surface tem­perature of the dry substance to a given low level (Te) ( from +40 to + 100°C depending upon the product).

On that factor there is really nothing important we can do except keep within the given limits.

On heat conductivity, on the contrary, it is possible to act. Actually, heat transfer through the dry shell is done both by the dry material itself and by the enclosed gas phase. On the first factor, we can just act by varying its thickness and structural orientation but, here again, for a given material, nothing else can be done. However, the gas phase is susceptible to be modified. It has been shown by Dr. Oetjen and Dr. Hackenberg [13] that by increasing the water vapour pressure inside the dry material, it is possible to increase the heat transfer enormously. A similar observation made in other circumstances, has been reported by M. Rieutord [25] who showed that by increasing the partial pressure of non-condensable gases in a drying chamber it is possible to increase the heat transfer between the heaters and the product to be dried. In both cases, it is quite evident that, at a higher pressure, heat transfer is more efficient.

However, turning back to Dr. Oetjen's experiments we have to reach a compromise. Indeed, increasing the water vapour pressure will definitely improve the heat transfer but, as we have seen before, will also decrease the pressure-drop and, consequently,

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diminish the mass transfer. Here again, an intermediate solution has to be worked out in order to adapt the heating in the optimal way in the proposed case. Of course, auto­matic regulation of the heat input has to be done in any case by one of the already men­tioned techniques.

= Technological aspects : We shall not examine here the numerous designs for freeze-dryers and their different

organs but just mention how heating can be achieved. From conventional sources (electricity - circulation of hot liquids - vacuum and / or pressure steam) energy can be supplied by the following means :

- contact heating : There is a close and intimate connection between the heater and the material to be

dried.

- radiant heating : Energy transfer is done in high vacuum by pure radiation without any contact between

the source of heat and the substance.

- mixed heating : convection and radiation (and even contact) : This is the most common case for the medium pressure range, heat being carried both

by the residual gas atmosphere and by radiation on short distances.

- high frequency heating : That new technique is still in the development state but looks interesting despite some

intrinsic difficulties (flashing over - uneven distribution of energy in non homogeneous material . . . ).

4. - Desorption (Secondary drying)

When the last fragment of ice has disappeared, the first drying period is finished. Sublimation being ended, if the same heat input is maintained, the dry material will warm up and progressively rise to above zero temperatures. It is then necessary to decrease heating to prevent the product from overcrossing the safety temperature for the dry state.

At that time, drying is far from being completed although all the water which had crystallized out under the form of ice has been extracted. There remains on the enormous internal surface of the porous dried layer a large amount of adsorbed water which would be widely sufficient to prevent any kind of storage at room temperature. The fourth stage of the freeze-drying process is then a desorption period during which the remaining water molecules are sucked out under high vacuum and a constant temperature : it is the isothermic desorption period or secondary drying (sometimes, this operation is effected in a special apparatus).

It is quite understandable that this water is more strictly bound and that, on the other hand, the dry material plays the role of a huge filter-plug which tends to prevent adsorbed gases and volatile components from escaping. This is fortunate as far as binding of volatile aromas or trapping of radioactive dust or virus particles are concerned but unwelcome when dealing with water vapour and adsorbed oxygen. However, if vacuum is sufficiently good (0,01 Torr or below), water is slowly and regularly extracted and after several hours the remaining amount is very low. At the same time oxygen is pumped off too. Neverthe­less, neither of them will ever get to zero as desorption is a continuous indefinite process and does not end suddenly as sublimation does. It will then be necessary to stop the process at a given point when it is felt that the residual moisture is sufficiently low to ensure good preservation.

Actually the exact determination of that end-point, as well as the evaluation of the residual moisture [21] are not easy things and are mainly a question of convention.

- Drying end-point can be checked by water-vapour pressure determinations or by measuring the water-vapour flow or even by thermal or electric measurements.

- Residual moisture can be estimated by the constant weight method or chemical titration according to Karl Fisher's technique [22] or by equilibrium water vapour

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measurements [23] or else by electric determinations including most elaborate techniques such as Debye Dipolar Absorption [26] or Nuclear Paramagnetic Resonance [24].

In any case, for a given product, and for a given time of storage, a well defined residual moisture has to be reached. Biochemicals, antibiotics, bank tissues require very low water contents and can be kept for decades. Living bacteria or viruses must not be too dry and it is necessary to buffer their final water contents with sugars. It is not advisable either to dry foodstuffs too much for both economic and organoleptic reasons. Besides, they are generally sold very quickly and it is not necessary to get to very low moisture con­tents that would make it possible to keep them for many years if they are given to the consumers within a few months.

5. - Final conditioning : Opening and Packaging

A priori, it could seem that these are elementary stages where no special care is needed. However, it must be remembered that the structure of the frozen and dried product is that of a highly porous and active filter, generally very hygroscopic. Indeed, if vacuum is broken with ambient air, the dried material will immediately suck water vapour and oxygen and all the work done during desorption will be ruined. If we do not want to impair long term storage, great care has to be taken in order to break vacuum with a dry inert gas if direct sealing of the material under vacuum cannot be achieved. Carbon di­oxide and essentially nitrogen are used for that purpose and give excellent results. For biochemicals, biological products and mainly pharmaceutics, argon has proved to be more efficient [I].

At any rate, when vacuum is broken in that way, it is advisable to perform stopping, sealing and even packaging in the same chamber or, at least, in the same atmosphere in an air-tight and closed area filled with dry inert gas. Manipulations are then effected by remote control [Hackenberg, 9]. Of course, and for the same reasons, vials, stoppers, boxes, bags and so on, used for filling and packaging must be made of a non porous material and sealed hermetically. Plain metal, glass, synthetic rubber, silicon, poly­ethylene, plastic-coated aluminium foils are commonly used for that purpose. They must be dried beforehand and even desorbed under vacuum. In those conditions, frozen and dried products can be stored almost indefinitely, and it is possible to avoid denaturation in the course of storage : oxydative reactions, enzymatic evolution, Maillard's browning reaction . . . However, fatty products are very difficult to store for a long time and very often develop a rancid taste after a few months.

6. - Rehydration

The last stage is the reconstitution of the product which is operated by adding water to the dried material. Because of its highly porous structure, rehydration is very fast and must be complete. In fact it has been shown by Professor Goldblith [5] that the rate of rehydration of frozen and dried material is a good measurement of the quality of the freeze-drying process.

Many different techniques can be used to add the water - for pharmaceuticals, the simplest way is to pour the right amount of water directly

in the vial containing the dried product. Dissolution must be instantaneous without shaking or, at most, with gentle agitation.

- bank tissues are generally soaked in physiological saline or balanced salt solutions. Rehydration may be achieved within a few minutes (arteries, skin, cornea) or take several hours (bones).

- foodstuffs may require a few minutes to get reconstituted. Those that have been pre­cooked can even be rehydrated in warm or boiling water. Meat to be grilled will be soaked in water first, then thoroughly wiped and grilled normally. If it has been processed correctly and has retained its water holding capacity it will be cooked easily and will not shrink on the grill.

- all powders are rehydrated by direct addition of water and thus can be made more or less diluted or concentrated. This can be of great interest for biologicals and even for foodstuffs such as coffee or tea. In the medical field a very elegant application of this has

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been discovered by the ophthalmologistProfessor Paufique. Ina case of detachment of reti­na, a concentrated solution of frozen and dried vitreous is injected in the posterior cham­ber of the eye. The viscous vitreous will swell to reach its normal concentration and thus will push back the retina in position.

Those are, briefly outlined, Ladies and Gentlemen, the main features of freeze-drying. It is an interesting new development of refrigeration which has already found many a new field of application and is likely to expand quickly in the years to come. It is not supposed to be a challenging technique but can reasonably make its own original way in close connection with the well-tested conventional methods. It is my great pleasure and privi­lege now to leave this chair to Professor Greaves, Professor Goldblith and Doctor Oetjen who will treat more specific problems related to freeze-drying.

REFERENCES

1. Air Liquide and Roussel-Vela/, 1962, Brevets d'invention, Paris.

2. D. R. Bangham, 1963, International Assay for the proposed new international standard of seric gonadotropin, Department of Biological Standards of the World Health Organization, London and personal commnnication, Paris, l96r.

3. E. W. Flosdorf and S. Mudd, 1935, J. Immnno. 25, 239. 4. R. M. Fry and R. I. N. Greaves, 1951, The survival of bacteria during and after drying, J.

Hyg, Cambridge, 49, 220. 5. S. A. Goldblith, 1963, The role of food science and technology in the freeze dehydration of

foods, III emes Cours Internationanx de Lyophilisation, Lyon France, I. N. S. A. 1962, in press at Hermann, ea. Paris.

6. R. I. N. Greaves, 1946, The preservation of proteins by drying, Medical Research Council, ed., London.

7. R. I. N. Greaves, 1960, La lyophilisation des bacteries et le probleme des agents protecteurs, in Traite de Lyophilisation by REY, L-R. et collaborateurs, Hermann ea. Paris, 207-218.

8. R. I. N. Greaves, 1962, Preservation of living cells by freezing and drying, in Progres Recents en lyophilisation, by REY, L-R., et coll., Hermann ed., Paris, 167-179.

9. U. Hackenberg, 1962, Manipulation of freeze-dried foods in inert gas, Congres A. V. I. F. I. A., Dijon, 1962, Le Vide, ea. Paris, 102, 560-564.

10. G. W. Hyatt, 1960, La banque des tissus d'origine humaine, in Traite deLyophilisation by REY, L-R. et coll., Hermann ea. Paris, 279-335

u. P. Lepine, 1960, Lyophilisation des virus-vaccins, in Traite de Lyophilisation, ibid., 227-238. 1 2. P. Mazur, 1960, Physical factors implicated in the death of microrganisms at subzero tem­

peratures, Ann. N. Y. Acad. Sciences, 85, 610-629. 13. G. W. Detjen, H. Ehlers, U. Hackenberg, ]. Moll and K. H. Neumann, 1962, Temperature

measurements and control of freeze-drying processes, in Freeze-Drying of Foods, Nat. Acad. of Sciences, Nat. Res. Council, ed. Washington, D. C., 25-42.

14. L-R. Rey, 1959 Brevets d'invention du Centre National de la Recherche Scientifique, Paris. 15· L-R. Rey, 1960 a, Theorie de la lyophilisation, in Traite de Lyophilisation by Rey, L-R. et

coll. Hermann ea. Paris, 19-53. 16. L-R. Rey, 1960 b, Thermal Analysis of eutectics in freezing solutions, Ann. N. Y. Acad.

Sciences, 85, 510-534. 17 . L-R. Rey, 1961, Automatic regulation of the freeze-drying of complex systems, Biodynamica

ed., Madison, Wisc. USA, 8, 241-260. 18. L-R. Rey, 1962, Systemes automatiques et mecanismes de regulation en lyophilisation, in

Progres recents en lyophilisation, Hermann ed., Paris, 81-97. 19. L-R. Rey, 1962, Principes generaux de la lyophilisation alimentaire, Congres A. V. I. F. I. A.,

Dijon, 1962, Le Vide, ed. Paris, 101, 428-444. 20. L-R. Rey, 1963, L'Avenir economique des produits lyophilises, Vemes Joumees Internatio­

nales de Lyophilisation Leybold, Koln, W. Germany, in press. 21 . L-R. Rey, 1963, L'Humidite residuelle des produits lyophilises, III emes Cours Internationaux

de Lyophilisation, Lyon, Fr. Insa, 1962, in press at Hermann ed. Paris, and Galencia Acta, Barcelona, Spain, 1963, in press.

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22. L-R. Rey and M-C. Bastien, 1962, Biophysical aspects of freeze-drying, in Freeze-Drying of Foods, Nat. Acad. of Sciences, Nat. Res. Council, ed., Washington D. C., 25-42.

23. L-R. Rey, M-C. Bastien, L. Rieutord and ]. Mosnier, 1963, Mesure de l'humidite residuelle des produits desseches et en particulier des produits lyophilises par la methode de la tension de vapeur d'eau a l'equilibre (facteur E R H), XI th International Congress of Refrigeration, Munich 1963, Sub-Commission 6 C.

24. L-R. Rey, D. Simatos and M-C. Bastien, 1963, Unpublished research work. 25. L. Rieutord, 1960, Brevets d'invention Usifroid, Paris. 26. D. Simatos, 1963, Constante dielectrique et teneur en eau des produits lyophilises, XI th

International Congress of Refrigeration, Munich 1963, Sub-Commission 6 C.

77

Freeze-Drying in Biology

La cryo-dessiccation en biologie Prof. R. I. N. GREAVES, M. D.

Department of Pathology, University of Cambridge, Tennis Court Road, Cambridge, England

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SOMMAIRE. La technique de cryo-dessiccation est largement utilisee dans les laboratoires de biologie. La demande de plasma deshydrate pour la transfusion et de penicilline deshydratee au cours de la deuxieme guerre mondiale a entraine la mise au point d' appareils pouvant deshydrater de grandes quantites de produits.

En biologie, les principales applications de la cryo-dessiccation sont la conservation de solutions biologiques fragiles, la preparation de tissus pour la microscopie et la conservation de cultures vivantes de bacteries et de virus. Malgre la destruction des cellules de mammiferes par la cryo-dessiccation, cette technique s' est revelee utile pour la conservation des greff est en vue de la transplantation la ou l'on n'a pas besoin de greffes vivantes, par exemple les greffes de la peau et des arteres, les greffes osseuses et les greff es de cornee.

Ce rapport decrit les progres recents de la cryo-dessiccation en biologie. En particulier, la deshydratation des solutions biologiques a ete placee sur un plan scientifique par l' analyse thermique des solutions biologiques par REY. On etudie l' application pratique de ces resultats.

Les problemes sou/eves par l'essai de conservation de cellules vivantes d'une complexite croissante par la congelation et la deshydratation est l'une des plus grandes gagneures de la bio­logie cellulaire actuelle. On decrit un systeme prototype qui pourrait aider a resoudre quelques uns de ces problemes.

FREEZE-DRYING IN BIOLOGY

"The phenomenon of low-temperature evaporation of water under vacuum to pro­duce freezing, followed by sublimation of the ice is old - so old, in fact, that William Hyde Woollaston was apologetic in exhibiting it before the Royal Society of London in 1813". This sentence is a quotation from the book by the late Earl Flosdorf, 1949, who goes on to state that "the first clearly recorded use of sublimation for preservation is that of Shackell in 1909, who applied it to biological substances".

From these early small beginnings, the technique has found wider and wider appli­cations in the biological laboratory and, although originally it was only applicable for the processing of small quantities of material, it may now be used for processing commercial quantities and even for the preservation of foods.

The technique emerged from the laboratory during World War II, when it was used to produce dried blood plasma for transfusion purposes. These plants provided the data for the drying of penicillin when it became available towards the end of the war.

Originally it was an expensive technique, both in capital and running costs and was, consequently, only useful for the preservation of expensive products. Progressively, since the war, production costs have been reduced so that the technique has become useful in wider and wider fields.

Although the low temperatures used are important in slowing down chemical changes, probably the greatest benefit in freeze-drying results from the fact that sublimation from the solid state prevents the concentration and aggregation of molecules which occur when drying from the liquid state. When dry, the proteins are found to be highly stable, even at high temperatures at which, in the liquid state, they would have been completely denatured.

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But perhaps the most fascinating use of freeze-drying is for the preservation of living organisms. In the laboratory freeze-drying has become the standard method for preser­

ving stock cultures of viruses, bacteria and some fungi. In the dry state these cultures retain their full virulence for long periods of time and, consequently, the methods is of great value for maintaining the full protective power of living vaccines such as B. C. G. and Vaccinia. But as the cell becomes more complex, the freeze-drying process fails and the cells are killed, and so far no one has succeeded in drying mammalian cells without killing them. This is the great challenge to workers in this field.

In spite of the fact that mammalian cells are killed by freeze-drying, the technique has proved useful for preparing cells and tissues for histological examination and also for preserving grafts for transplantation in those circumstances when a living graft is not required, such as grafts of skin, arteries, bone and cornea. There is some evidence that the transplantation antigen is partially destroyed by freeze-drying so that the dried grafts may last longer than fresh grafts before being destroyed by the homograft reaction.

THE FREEZE-DRYING OF BIOLOGICAL SOLUTIONS

Blood serum and plasma, antisera, antibodies, hormones, amino-acids, vitamins and pharmaceutical chemicals are the biological solutions most frequently preserved by freeze-drying.

In the early days the physical conditions for drying were largely empirical, but it was recognised that usually the lower the temperature of drying the better the product. It was also recognised that the speed of prefreezing affected the appearances of the dried product and that the faster the prefreezing the smaller the crystal size and the more pleasing the look of the final product. Nevertheless, there were some products, notably the early samples of penicillin, which, although they appeared to be solidly frozen, emitted bubbles under vacuum, as Flosdorf so admirably expressed it, as drying by "puffing".

Since the war, Rey (1960), has submitted a number of biological solutions to Thermal Analysis. As the result of analysis, it is now possible to define the optimal drying condi­tions for any particular product.

When salt solutions are frozen, at first pure water crystalises out as ice, leading to a concentration of the salt in the remaining water. Eventually a concentrated salt solution known as the eutectic concentration remains, which freezes at the eutectic temperature. The eutectic temperature of Na Cl is -21.6°C, and since Na Cl is usually present in biological solutions, if they are dried at a higher temperature than -21.6° C there will be some drying from the liquid phase. Thermal analysis of a Na Cl solution shows that on freezing there is considerable supercooling of the eutectic mixture which may not solidify until temperatures of the order of -40°C are reached. On warming, however, there is no thawing until the eutectic temperature of -21.6°C is reached. This experi­ment demonstrates the importance of prefreezing to a much lower temperature than the temperature of the lowest eutectic of the solution.

Certain substances, particularly sugars, when rapidly frozen, form metastable glasses. On warming, these glasses become more and more plastic and, on drying, may give rise to "puffing". Rey (1960) has shown that if these glasses are held at a higher temperature they will crystalise in time and then, on drying, "puffing" is avoided.

With biological materials, like serum and plasma, the high protein content may mask the eutectics on differential thermal analysis, though the point of incipient melting may be observed. This temperature with horse serum has been shown by Rey (1960) to be -35°C, a temperature which is much lower than had previously been suspected.

Thermal analysis will indicate the necessary prefreezing procedure, and also the tem­perature below which drying should take place. Drying could thus be regulated by a ther­mal element placed in the material to be dried, which controlled the heat input so that this temperature was never exceeded. But such a method would be excessively slow becau­se the heat input must be progressively reduced and it would be impossible to tell when the material was dry enough for it to be safe to raise it to its final dry temperature.

Fortunately it has been shown by Rey (1961) that there is another solution to this pro­blem. When an electrolyte is in solution its electric resistance is very low. When such a

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solution is completely frozen, its resistance is very high. This resistance change takes place very rapidly at the eutectic temperature. If the resistance is determined at which all the eutectic mixture is frozen, this value may be used in an A. C. resistance bridge to control the heat input, any tendency to reduce the resistance below this determined level leading to a reduction in the heat input. In this way it is possible to dry at the fastest possib­le rate consistent with never melting the eutectic mixture. Moreover, when dry, the resistance will rise so that the dried material is quickly raised to its final temperature, at which temperature the heat is cut off by a second control device working on temperature.

The control working on resistance of the frozen material must be such as to have very little or no overswing, otherwise there might be some thawing of the eutectic mixture which would then supercool and not refreeze. It is for this reason that I have found the saturated reactor such an ideal form of control. The circuit used is that described by Roberts (1951) in which the output from an A. C. bridge is fed to an amplifying valve, through a phase discriminating valve to an output pentode in the primary winding of the saturated reactor. With such a device the control of heat is not by cutting on and off, but a slow and continuous reduction as drying proceeds. Moreover, by suitably adjusting the sensitivity of the amplifier, overshoot can be completely avoided.

Since the output windings of saturated reactors can be run in series, it is possible to use two such amplifiers, one fed from the resistance bridge and the second fed from a resistan­ce thermometer or thermistor bridge which will control the final drying temperature.

With such a method of heat control, it is not really necessary to do a preliminary analysis of the material. If the balancing resistance is set at 10 megohms, one knows that the mate­rial has been well frozen throughout. But probably at this resistance the drying temperatu­re would be unattainably low, and, in practice, 1 Mw gives an excellent end product. Sometimes an even lower balancing resistance may be justified.

In assessing a dried product, I would rather know the resistance at which it was dried, than the temperature.

FREEZE-DRYING OF TISSUES

As already stated, no mammalian cell has retained its viability after freeze-drying. This is probably due to the lethal effect of intracellular ice crystal formation. Unfortunately protective substances, which have proved so useful for the freezing and thawing of living cells, such as glycerol and dimethylsulphoxide, cannot be used since they concen­trate on drying and reach a toxic level.

However, the technique has proved useful for the preparation of specimens for histo­logical examination, particularly for the examination of enzyme systems in cells, as the enzymes are not destroyed by freeze-drying. But if the morphology of the cell is to be pre­served for electronmicroscopy, then it is important that the ice should be amorphous. Fernandez-Moran (1960), has shown that this may be achieved by very rapid freezing of small pieces of tissue in Helium II, followed by the substitution of the ice in solvents at a temperature below -100°C.

The freeze-drying process has also been used for preserving tissues for grafting, in particular arteries and veins, bone, skin and cornea. These grafts are not required to be viable and grow, but to act as scaffolding on which new tissue can be laid down. There is some slight evidence that the process of freeze-drying may destroy the transplantation antigen and claims have been made of the successful use of heterologous tissue. This cer­tainly seems to be so with corneal grafts, and Henaff (1960) claims successful grafting of heterologous corneas. Lamminar corneal grafting is, however, rather a special case, for the successful graft is isolated from the blood stream and hence the antibody forming apparatus.

When freeze-drying cornea,�it is essential that on resolution the cornea should be transparent. Henaff (1960) has,shown that this is possible if prefreezing is very rapid in liquid N, and if drying is carried out at a low temperature which must be below -50°C.

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FREEZE-DRYING OF LIVING ORGANISMS

Provided an organism is not killed by freezing and thawing, it may be preserved by freeze-drying. As a method for preserving stock cultures of viruses and bacteria, freeze­drying has been used for many years. However, the survival of all species is not uni­formally good and it has been found that the medium used for suspending the organisms in for drying greatly affects the percentage survival immediately after drying and also the long term survival at high temperatures.

A simple medium of 5 % peptone suits viruses well, but bacteria survive far better if a sugar such as glucose or, preferably, sucrose is added at about 7 Y:! % . The addition of Na glutamate to this mixture improves survival at high temperatures considerably (Greaves, 1960). Scott (1960), has shown that for good survival it is important to include in the drying medium a substance such as an amino-acid which will neutralise carbonyl groups on the surface of bacteria, which probably explains the effectiveness of peptone. The sugars probably exert their effect by buffering the residual moisture content. This is not necessary for viruses, but for the more complex bacteria complete dryness seems to be incompatible with life.

Both freezing and drying are potentially lethal to living cells. Annear (1956 a & b) has shown that by very rapid drying from the supercooled but liquid state Leptospiria and the protozoon Strigomonas oncopelti may be preserved, whereas both would be killed by freeze-drying. I have tried to carry Annear's experiments a stage further by attempting to dry the more complex organism Euglena gracilis, but without success. There is a continuous fall off in viability as drying proceeds till, at 85 % dryness, all are dead.

1 00

9 0

B O

7 0

6 0

5 0

40

3 0

20

10 ' \

0 '---0±-----,_ 1"-o-�- 2-"0,------'-. �30,----=-""•o�---"'so!,----Tcmptr.:iturc °C

Fig. r. Solid line - Percentage survival of yeasts when dried at temperatures ranging from +2 to -40°C. Dotted line - Percentage survival of yeasts on freezing and thawing at different temperatures as observed by Mazur. (See text.)

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The problems raised in attempting to preserve living cells of increasing complexity by freezing and drying is one of the great challenges of cellular biology today. As already stated, great success in preserving cells by freezing has followed the discovery of the use of glycerol and dimethyl sulphoxide as additives. The search for a suitable additive for freeze-drying first requires the discovery of a model system.

Such a system may be available using the yeast saccharomyces cerevisiae. Mazur (1961), has studied the effect of exposure to subzero temperatures of this yeast, in great detail. He found that the cells were injured only when the external medium was frozen, and then only when the temperature was -10° or below. Survival dropped abruptly between -10° and -30°C. He considers that death is the result of intracellular ice formation which occurs when external ice crystals grow through aqueous channels in the cell wall and seed supercooled water in the cell interior.

Fig. I shows the result of drying this yeast in a medium of 1 % bovine albumin at temperatures from 0° C to -40° C. At 0° C the survival immediately after drying was 18 % . This rose progressively as the drying temperature was reduced, reaching 50 % at -15°C, then falling rapidly at lower temperatures. Superimposed on this figure is the survival curve that Mazur obtained on freezing and thawing alone.

This experiment shows that drying from the liquid phase is a fairly lethal process for these organisms, that as the temperature falls survival improves, but that a phase change in the extracellular fluid is not reflected one way or the other in the percentage survival. Intracellular ice crystal damage does not become apparent on drying until -15°C is reached instead of -10° C on freezing alone. This increased protection on drying may be because damage on thawing is avoided.

It is my hope that here I have a model system for investigating the effects of intra and extra cellular additives.

REFERENCES

E. W. Flosdorf (1949), Freeze-drying. Reinhold Publishing Corporation, New York. L-R. Rey (1960), Thermal analysis of eutectics in freezing solution. Ann. N. Y. Acad. Sciences, 85, 510. L-R. Rey (1961), Automatic regulation of the freeze-drying of complex systems. Biodynamica, 8, 24I . M. H. Roberts (1951), Electric controllers for laboratory furnaces. Electronic Eng. 23, SL H. Fernandez-Moran (1960), Low temperature preparation techniques for electron-microscopy of biological specimens based on rapid freezing with Helium II. Ann. N. Y. Acad. Sciences, 85, 689. F. Henaff (1960), Preparation of grafts of freeze-dried cornea. p. 295 in Recent research in freezing and drying. Blackwell, Oxford. R. I. N. Greaves (1960), Some factors which influence the stability of freeze-dried cultures. p. 203 in Recent research in freezing and drying. Blackwell, Oxford. W. ]. Scott (1960), A mechanism causing death during storage of dried micro-organisms. p. 188 in Recent research in freezing and drying. Blackwell, Oxford. D. I. Annear (1956a), Preservation of leptospirae by drying. J. Path. Bact. 72, 322. D. I. Annear (1956b), Preservation of Strigomonas onoopelti in the dried state. Nature. 1 78, 413. P. Mazur (1961), Physical and temporal factors involved in the death of yeast at subzero temperatu­res. Biophysical Journ. l, 247.

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Freeze Drying of Foodstuffs

La cryo-dessiccation des aliments

SAMUEL A. GOLDBLITH and MARCUS KAREL Department of Nutrition and Food Science, Massachusetts Institute of Technol­ogy, Cambridge 39, Massachusetts, U. S.A.

SOMMA/RE. Ce rapport presente une etude de la cryo-dessiccation en fonction (a) des matieres premieres, (b) des parametres de transformation, ( c) des conditions exigees duproduit fini, ( d) de l' emballage et des conditions d' entreposage, et ( e) de la possibilite de rehydratation.

On presente des exemples et des donnees experimentales sur ies crevettes et les tranches de saumon pour illustrer chacun des facteurs ci-dessus.

Dans le cas des parametres de transfnrmation, on presence des donnees sur !'influence de la temperature a laquelle la surf ace d'un produit est exposee, sur le temps de deshydratation et sur la qualite du produit fini.

On presente des donnees en fonction du taux de deshydratation, de la capacite de retention d'eau, de l'activite de l'ATPase et de la retention de la couleur rose des tranches de saumon et des crevettes.

INTRODUCTION

Freeze dehydration of foods is now beginning to emerge as an important primary method of food preservation, complementing the two more important of our present methods, thermal processing and freezing.

Historically, as with thermal processing and with freezing, the basic scientific founda­tions for freeze dehydration lag well behind the technologic accomplishments.

This, however, in terms of producing products which are uniformly optimal in flavor, color and texture is not without its attendant risks.

Let us examine the freeze dehydration of foods in relation to the various factors one normally considers in any method of food processing. These are :

a) raw materials b) processing parameters c) finished product specifications d) packaging and storage conditions e) rehydration

In this instance, a comparison with preservation using refrigeration is, of course, logical and significant inasmuch as freezing represents one of the processes used in freeze dehy­dration.

Due to time limitations, brief mention and illustrations only of each can be made.

RAW MATERIALS

Any processed food product can only be as good as the raw material from which it is made. Processing decreases rather than increases the quality of a foodstuff. However, processing is a means of extending the usefulness of a food product for man.

The concepts and precepts followed with respect to raw materials for frozen foods are just as applicable for foods to be lyophilized, inasmuch as the finished frozen product becomes the material for the first stage in freeze dehydration. The influence of the freezing procedure on the quality of freeze dehydrated shrimp may be demonstrated by comparing the rehydration ratios of freeze dried shrimp frozen initially at different temperatures. These data are presented in Fig. 1. It is evident that the amount of water picked up by shrimp is markedly increased with higher freezing temperatures (slower freezing rates).

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Obviously, a partial disruption of structure by large ice crystals at the higher temperatu­res promoted higher absorption of water during rehydration. However, the amount of water picked up is not the sole objective criterion of quality upon which judgment of opti­mal process parameters can be based, since organoleptic evaluation of texture indicated that the product frozen at 15°F was judged as too "soggy." The shrimp frozen at 0°F and -40° F were considered the most acceptable.

Since the frozen foodstuff must now be processed even further by exposure to heat energy in quantities sufficient to vaporize the ice under a high vacuum, it is self-evident that, in order to obtain a high quality freeze dried food material, the raw materials must be of a quality better than those selected for freezing preservation. In the case of vegeta­bles and other foodstuffs, which contain enzyme systems which normally must be inactiva­ted for successful storage in the frozen state, the same principles apply in the case of freeze dehydration as in frozen foods, since these enzymes are active at low moisture contents, or will become active on rehydration.

PROCESSING PARAMETERS

Obviously, this is as important in the case of lyophilization as with any other type of processing. A comparable situation exists in the case of thermal processing where over­processing can easily cause damage to the product.

In the case of freeze dehydration, each product has a maximum temperature level which the outermost dry layer can tolerate. As the ice phase recedes, and there is more mass (and thickness) of dried material, there is increased resistance to both energy (heat) transfer into the ice core and mass (vapor) transfer through the dry matrix to the outside.

Since each product has a maximum temperature which the outside layer (which of course is dry) can tolerate, at the time when it becomes increasingly difficult (and increa­singly necessary) to direct energy into the ice phase to increase the vaporization rate, one must become exceedingly careful not to exceed the maximum temperature level which the dry phase can tolerate.

The optimum conditions of dehydration must therefore be developed on the basis of a compromise between the desire (and the economic need) to have the shortest feasible drying cycle, and an equally important need to maintain the highest possible quality. That these two aspects are often contradictory may be demonstrated by the following :

86

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5.0

� 4.0 .... ... �

5 i== C2 30 0 >­I w a::

0

FREEZING TEMP. (°F)

--------· -------· 16 .------· ------�• 0

--· -·--------- · -----· --------· -40

----------. -• _- • -321

--------------·

�· •

5 10 15 20 RE HYDRATION TIME (MINI

25 30

Fig. r. Rate of rehydration of freeze-dried shrimp as temperature of freezing medium.

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Shrimp dehydrated under conditions in which the maximum surface temperature is allowed to rise to 175°F may be dried to a moisture level of 2.0 % in approximately six hours, whereas shrimp dried under identical conditions except that the maximum allow­able surface temperature is 125 ° F, require eight hours to dry to 2.0 % moisture. This reduction in drying time is obtained at the cost of some loss in quality. For instance, the rehydration ratio for shrimp dried at a temperature of 125°F is 4.29, whereas for shrimp dried at 175 ° F it is 4.01. This indicates that the dehydration at the higher temper­ature has resulted in some protein changes which lower the waterholding ability.

The same considerations apply to an even greater extent to "water holding capacity" or the ability of the muscle to retain water after being subjected to pressure. This is illus­trated in Table 1, which presents data on the water holding capacity of haddock fillets subjected to various treatments. It is evident that, as the temperature of drying increases, the extent of rehydration decreases, and that such water as is being absorbed by the dried fillets is held more "loosely" than in the frozen control.

Table 1. Effect of Maximum Temperature of Surface During Freeze Drying on Rehy­dration Ratio and Water Holding Capacity of Haddock Fillets [1]

Frozen control Freeze dried at 125°F Freeze dried at 175°F

Rehydration Ratio* gm water/gm solids

3.9 3.6 3.47

Water Holding Capacity** gm water/gm solids

1.25 1.00 0.87

The same phenomenon may be demonstrated by the retention of A TPase activity in muscles of various marine species after freeze drying at various temperatures. Since A TPase activity is associated with the muscle protein, actomyosin, the loss of this activity is a measure of protein denaturation in muscle. Table 2 presents data on retention of ATPase activity in salmon, mackerel, and shrimp after freeze drying at different maxi­mum surface temperatures.

Table 2. Retention of ATPase Activity in Freeze Dried Marine Products [2]

Percent of Original Activity Salmon Mackerel Shrimp

Dried at 125°F 72 62 80 Dried at 175°F 69 55 62 Dried at 125°F and stored 59 50 80 1 week at 98° F Dried at 175° F and stored 50 36 56.5 1 week at 98° F

I t is evident that higher surface temperatures during drying result in denaturation of protein. Further, it is evident that the denaturation during drying and additional dena­turation during subsequent storage are additive. Thus, it is important to minimize these changes throughout the various operations in order to have high quality freeze dried foods with long shelf life.

*Fillets weighed before and after soaking in cool water. **After being subjected to a uniform pressure (approximately roo lbs/inch") between two sheets

of filter paper. Moisture content of rehydrated material actually determined by analysis.

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Another aspect of processing parameter relates to the temperature of the ice layer which, of course, is a function of the pressure of the dehydration chamber. If the ice temperature is too high, especially with certain types of high sugar foods, e. g., bananas, concentrates, etc., recrystallization and interstitial melting may occur with resultant deleterious effects on the product quality. This has been well illustrated in the researches of Professor Rey and his associates.

FINISHED PRODUCT SPECIFICATIONS

This is extremely important in the case of freeze dried foodstuffs, particularly because of the sensitive nature of the finished products.

These specifications obviously fall into several categories, including : a) chemical b) microbiological c) physical

Included in the chemical properties are those such as color. In the case of shrimp, for instance, improperly handled shrimp with pink astacene pigments may become yellow­ish on dehydration or on subsequent storage, due to oxidation. Physical properties include those such as texture.

Both of these are greatly influenced by the raw materials as well as the progressing storage parameters.

Table 3 shows the retention of pink color in salmon and in shrimp as a function of conditions of dehydration and of conditions of storage. These data show again the additive effect of storage and processing conditions on deterioration of freeze dried foods. The role of oxygen in promoting the loss of the pink color is particularly evident.

Table 3. Effect of Maximum Temperature of Surface on the Retention of Pink Color (as determined by 0. D. of extracted astacene) in Shrimp and Salmon [3]

Immediately after drying After 1 week's storage at 68°F in less than 1 % oxygen After 1 week's storage at 68° F in air

Percent Color Retention Salmon Shrimp

Dried at 125°F

100 %

86 % 65 %

Dried at 175°F

78 %

67 % 59 %

Dried at 125°F

100 %

100 % 23 %

Dried at 175°F

100%

73 % 23%

·-------··----

Microbiological specifications of freeze dried foods are extremely important inasmuch as one is dealing with completely unknown entities, in respect to the ability of freeze dried foods to support the growth of bacteria on rehydration and storage at refrigerated temper­atures. This is an area deserving of a great deal of study. Until such time as sufficient data are at hand, however, it is necessary to maintain maximum microbiological control via good sanitation and manufacturing practices and thus achieve minimal counts.

Specification and standards must be tailored to fit each specific food product. No single set of any type of standard can be made to fit each foodstuff. Suffice it to say that if rigid specifications and standards, based on actual data and judiciously applied, are prop­erly drawn up, these offer a means of insuring good final quality of finished products. These final product specifications are a reflection of the degree of care used in raw mate­rial selections and manufacturing procedures.

PACKAGING AND STORAGE CONDITIONS

Freeze dehydrated foods, by their very nature, are extremely sensmve materials . Whereas the chemical constituents of frozen foods are protected from oxidation by the ice layer, freeze dried foods do not have any such protective layer.

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Since freeze dehydration occurs with foods which are in the frozen state, there is no loss in size or shrinkage. As a result, however, these foods are extremely porous with a great deal of surface area in the dried matrix. This offers a large surface for oxidation and other types of reactions to occur.

As a result, such simple routine procedures as breaking the vacuum in the chamber with nitrogen and packaging under nitrogen with a maximum arbitrary oxygen content of the headspace gas of less than 2 % has proven efficacious.

Obviously, the package is extremely important as it must maintain the nitrogen at­mosphere and be impervious to moisture.

Experience has shown that freeze dehydrated foods, properly manufactured from high quality raw materials, packaged in impervious containers under nitrogen and under 2 % moisture are of good quality and equivalent to their frozen counterparts.

In certain instances the quality of some freeze dried foods may be better than their frozen counterparts. This would include such products as cooked lobster and crabmeat. Frozen cooked crabmeat undergoes protein denaturation and becomes tough in frozen storage. This does not occur in the case of freeze dried pre-cooked crabmeat or lobster, however, which, if properly freeze dried to under 2 % moisture and stored under nitrogen, remains succulent and without such deleterious changes even after 18 months storage at room temperature.

REHYDRATION

To date, most of the research in the field of freeze dehydration has concerned itself with the mechanism, techniques, equipment, etc. for dehydration.

Of equal importance, but scarcely studied at all, is the area of rehydration. Such simple questions as optimal temperature and pH of rehydrating medium are of tremendous im­portance and have received relatively little research effort. This area is worthy of concerted effort on the part of all manufacturers of freeze dried foods inasmuch as the marketing potential for such products is to a large measure dependent on making available to the potential user as complete a set of facts as possible pertinent to their uses.

THE FUTURE

Freeze dried foods have arrived on the marketplace. At the present time they have been largely sold for institutional uses and to manufacturers as components for such diverse products as dehydrated soups, casseroles and other main course dishes.

Due to their convenience characteristics these offer tremendous potential in Europe, the U. S. A. and, in fact, all over the world.

Food science and technology (which includes refrigeration engineering) holds one of the keys to the success of this new method of processing.

Fundamental data on the five characteristics outlined above can do much to break open the market potential for the product inasmuch as market tests have shown that the con­sumer is able and willing to pay for convenience-providing she also gets top quality.

Freeze dehydrated foods will not replace frozen foods nor canned foods. Lyophilization will find its rightful niche and will complement thermal processing and freezing to help feed tomorrow's world.

BIBLIOGRAPH Y

r . H. Pimental, S. M. Thesis, Massachusetts Institute of Technology, 1 963. 2. G. Lusk, Ph. D. Thesis, Massachusetts Institute of Technology, 1963. 3. S. A. Goldblith, M. Karel and G. Lusk, The role of food science and technology in the freeze dehydration of foods. Food Technol., in press.

P-1 Engineering Problems and Economical Aspects of Freeze-Drying

Problemes d'engineering et points de vue economiques de la lyophilisation

Dr. GEORG-WILHELM OETJEN Firma Leybold Nachf., Koln-Bayental, Germany

SOMMA/RE. La conference essaie d'eclaircir les problemes d'engineering, le necessaire know-how pour le processus meme et les points de vue economiques de la lyophilisation. On donne une solution, pour laquelle on dispose deja des experiences d'operation, pour la mani­pulation des larges denrees alimentaires dans des installations de lyophilisation et on montre a l'aide d'un exemple la fafon d'effectuer un parfait emballage sous gaz protecteur a la suite de la lyophilisation. En ce qui concerne lesfrais, on partit d'une installation d' environ 10 tonnes de materiel brut par jour et ensuite on indique les frais totaux y compris la congelation et l' emballage en prenant pour base !es conditions normales en Europe.

The engineering of a freeze-drying facility for food presents a complex problem to the designer. Economic handling of large volumes through a multistage process is one re­quirement, considerable process know-how concerning material behaviour, heat and mass transfer, water vapour _isotherms and vacuum technology are others. Practically all freeze-drying installations for food prior to 1962 are of the "cabinet" type of either rectangular or round cross-section (Fig. 1). The process is purely a batch operation : the material to be dried is placed between heated plates either on trays or as a sandwich

Fig. 1 . Cabinet type freeze-dryer for food (designed in 1957)

stack (Fig. 2). The water vapour from the ice is removed by multistage steam ejectors or by cooled condensers, inside or connected with, the drying chamber. The "cabinet" principle is limited in capacity : this is because one must keep the loading and unloading times reasonably short, have structural elements such as doors, condenser plates etc., of an economic size and retain heating and cooling periods which are short compared with

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Fig. 2 . Diagrammatic representation of the assembly of the heater plates.

the drying cycle. Such limits are, of course, quite arbitrary but generally speaking (and accepting that this figure may vary for different designs by a substantial factor) two tons of food to be processed per day could be an average figure for any single cabinet. Larger production therefore had previously to be obtained by installing the necessary number of units in parallel.

Besides these technical limitations, a multi-cabinet plant creates difficulties when dry atmosphere or inert gas has to be used for unloading and packing. Certainly each cabinet can be filled with the necessary atmosphere at the end of drying to protect the food against absorption of moisture and oxygen during transportation, as far as possible. Then the material can be subjected to vacua and filled with inert gas again in a conventional packing machine. It is one of the great advantages of freeze-dried material that it has an extremely extensive openwork structure for easy reconstitution. On the other hand, this feature means that very large quantities of gas are absorbed per gram of dry substance and unit time. Bearing this latter fact in mind (and knowing that the dry product is evacuated to the best possible extent actually during the drying process itself), it seems a paradox that the food material should, even temporarily, be brought back into normal at­mosphere after freeze-drying only to be once more evacuated in order to provide it with its final protective atmosphere. The requirements of increased capacity and the desire to exclude moisture and oxygen have led to an "in-line freeze-drying system" whereby the material passes through the various steps of the process in a sequence of operations (freez­ing, freeze-drying, unloading, packing) without contacting the normal atmosphere from the point when the vacuum chamber has been closed until the dried product is sealed in a

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Fig. 3. 2-tunnel-freeze-dryer with freezing-device and inert gas unloading and packaging sy­stem

1 loading of carriers 2 cooling tunnel 3 freezing tunnel 4 monorail for carriers 5 freeze drying tunnel 6 ice-condensers 7 vacuum pump sets 8 refrigerating machines

9 control panels ro unloading of carriers under protective gas 1 1 unloading of trays under protective gas 1 2 balances 13 packaging machine 14 washing room l 5 drying room for carriers and trays

Fig. 4. Introducing of carriers into CQC-plant

suitable envelope. Fig. 3 shows the layout for such an in-line system. The material is placed on carriers (Fig. 4) which pass through the freeze-drying system. Unloading, weigh­ing and packing in the system is carried out on items 1 1 to 13 of Fig. 3. The layout of these is shown in Fig. 5. Fig. 6 shows a view of a fully continuous in-line system with an air-lock at entry and an inert gas lock at exit. One such plant has been in operation since 1962 ; another one has been commissioned recently (1963). The following short film shows this unloading and packing technique in operation.

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ill.li!!gMl.d..Jowering deyice for I s.usRended carriage

Fig. 5. Unloading and packaging of freeze-dried food m inert gas atmosphere.

Einfriertu�I u Eintrierband

Beladeraum

Einfahrsdilieuse

Vakuumdlchler Schieber

Trockenlunnel

AuslahrscNeuse

Kondensaloren

Vakuumpumps:itze

14

Schattsctvank 10 W�gel"Kln�lion (Sch.Jlzgasraum)

Scha)Qn•nttecrung

12 MUhtc-13 Oosie-r·u �rpaclrungsmaschme

\4 AJ.Jlom.alisdle Sc�wasdimaschine 15 K51tean!aqe --- Schalentransporl

_ _ Wagenlransport

·····-·- Pnxtukfunsftuss (Trod:enprodukt) \

Fig. 6. Continuous 4-tunnel-freeze-dryer with airlocks, freezing-device and inert gas unloading and packaging system.

l freezing tunnel and freezing belt 2 loading of carriers 3 entrance lock 4 vacuum tight gate 5 freeze drying tunnel 6 exit lock 7 condensers 8 vacuum pump sets 9 control panel

94

lo unloading of carriers under protective gas II unloading of trays under protective gas 12 mill 13 dosing and packaging under protective gas 14 washing machine for trays 15 refrigeration machine - - - - conveyer for trays

monorail for carriers . . . . . . . . conveying dry product under protective gas

P-1

Heat and mass transportation during the freeze-drying process as well as permissible moisture contents of dried material have been studied by different authors e. g. Chiche­ster and Harper, Ehlers, Hackenberg and Oetjen, Salwin, Decareau and others, while the significance of pre-freezing has been investigated by Rey and Greaves. Fig. 7 shows how widely the thermal conductivity of freeze-dried material varies from substance to substance and how much it depends on the water vapour pressure in the system. Apart from actual heat transfer, transportation of water vapour from the ice through the already dried material is one other important time factor. From measurement of the "permeabili­ty" of freeze-dried products, the pressure drop between the ice surface and the outer sur­face of the dried product can be calculated to obtain the achievable mass transfer for any product. If, for example, a pressure drop of 0.4 Torr is necessary to transport a

A Sr--�-----..,--,...--------,--r---,.--, +

10·• 10 ·• tO --

p H2D (Torr J

Fig. 7. Thermal conductivity of freeze-dried materials I cod slices IVb spinach, pulverized II beef slices V apple powder III cauliflower pieces VI banana powder IV a spinach leaves

corresponding amount of vapour through the dried layer of material, and if an ice tem­perature of -l8 ° C cannot be exceeded, then the operating pressure in the chamber should be kept below 0.5 Torr. For an increase of pressure drop, operating pressures have to be lower, which means more energy is demanded from the extraction equipment whe­ther it be steam ejectors or refrigerators.

The technical limit to increased rate of drying is reached when the required pressure difference approaches the absolute value of the vapour pressure at the ice surface. But the economical limit will be reached in many cases much earlier because of the increase in investment and energy demand when the operating pressure has to be low. The ambition of shorter and shorter drying cycles works in many cases against the necessary economy. The optimum drying conditions for a given product should therefore be established by the following steps : determination of the maximum tolerable ice - and dried material -temperatures which just yield the desired quality of product ; measurement of the heat -and mass transfer data for the material ; calculation of the respective pressure drops du­ring the drying ; establishment of optimum operating pressure to give a maximum heat transfer, to avoid thawing and yet remain within economical limits. This sounds very

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theoretical, but measurements with an automatic balance built into a freeze-drier as shown in Fig. 8 and 9 proved that theory and practice are well in agreement. A freeze­drying cycle (Fig. 10) of prawns bears this out.

meas.uring currl'nl electric -balance

brine -- ===:!J .. --

system for hlting balance automatically

electric balance inside fil(>rmocons11nt ct-iamber

producl inside tray

waler vapour rernova'

Fig. 8. Electric balance inside experimental drying chamber

...,, I ·.;,,I ·_.; Fig. 9. Experimental freeze-dryer with automatic balance

The culmination of the engineering effort towards easier handling and transportation of large volumes and the research underlying better understanding of the mechanics of the freezedrying process itself would remain incomplete if the behaviour of the freeze­dried product during mixing or blending and during storage were not to be carefully considered as well.

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100 90 80 70 � 60 � � t 30 J 20 10

0

--

•• ,1

� °lo solids -- - --�-

/ /

, / / ,

2 3

> °le water

t' (units)

Heating plate temperature 80 ('C) Water wpour pressire 1 (Tarr) Loading af trays 30 ( kgfm')

Fig. ro. Freeze-drying curve for prawns

The influence of water vapour on freeze-dried products is the subject of many publi­cations and Salwin has given a general view of this problem. Fig. 1 1 is a resume of his conclusions. These figures show how important it is that moisture should be excluded during packing and storage and also indicate the order of magnitude to which water vapour can be tolerated for four categories of products.

Fig. 11 . Categories of freeze-dried products as suggested by H. Salwin

Monolayer moisture

value (%) of solids

Relative Water vapour

Starchy foods, p. e. potato, beans, com, rice

Protein foods, p. e. meat, fish, egg

High sugar and high molecular weight constituents, p. e. green pepper, carrots, peas, cabbage, onions, milk

High sugar food, p. e. peaches

6

3,5 (on fat free

basis)

2

humidity pressure (approx.22°C) (Torr)

15

8 ,..,1,6

6 ,..,1,2

as low as possible

The economical evaluation of the freeze-drying process is difficult and influenced by many factors which are hard to measure and therefore to reproduce exactly. Quality of food the world over involves some data which can be measured but also assessment depending on individual taste. Furthermore, certain temperature levels are necessary to obtain flavours of food which do not pertain to the natural unchanged product, but which occur only by decomposition, caramelizing or a certain degree of burning. The fi­gures shown in the last diagram therefore relate to freeze-drying of products retaining their natural qualities as near as is possible. The basis for these calculations are as follows :

Amortisation: 10 years Interest : 5 % Fuel oil: 100 DM/to Electricity: 0,1 DM/kWh Steam : 15 DM/to Cooling water: 0,10 DM/m• Labour: 5 DM per manhour

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1 02

t i � 6

.... E 5 11.

� 4 � :::J _g- 3 CJl :::J 0 L..

.t::

L.. QI a. -;;;· 0 u 1 0'

9 8 7 6

I"'

I' 11 0 � [', � I//'

! 11-l I � I I I

! I

r--,.__ I//; fie/ -

I I b t, /I I I/ I /

ru V7 I I/ /I/ I 11 I!+ LL I/ I I/ I

� k IV I/ I ill I 11

I �w I I/ 11 11 N I

I

I 5 1 o• 3 4 5 6 7 8 9 10� 2 3 4 5 6 7 8 9 104

capa c i t y of i n s t alla t i o n �on/year] •

Fig. 12. Production cost for one kg of throughput for the freeze-drying process.

The diagram (Fig. 12) shows the order of magnitude of production cost for one kg of throughput for the freeze-drying process only, that means withoutprefreezing and without packing. Depending on the requirements of each product up to 30 % for freezing and pretreatment and up to 40 % for unloading and packing have to be added to these costs to provide the total direct cost per kg of throughput.

SUMMARY OF THE DISCUSSION (Papers P-6, P-5, P-4, P-1)

Technical and economical considerations about freeze-drying

A. P. Longmore, U.K. : Dr. Oetjen in his paper explained the advantage of the semi­continuous tunnel system and the disadvantages of the batch chamber system of freeze­drying but did not examine the advantages of the latter such as simplicity of control and operation, flexibility and versatility in handling several products at one time in a multiple system. There are also advantages in the reduction of the danger of losing a large quantity of product due to equipment breakdown and ease of maintenance due to processing in separate chambers.

G.-W. Oetjen, Germany : As I tried to explain in the paper, the special advantages of the discussed CQC-system are just the simplicity of control, operation and handling. Furthermore, this system is flexible to be used as batch, semi-continuous or continuous, depending on product and throughput: A single CQC-tunnel is a batch unit but with all the advantages of handling and control; two tunnels in series form a semi-continuous plant with only one small room needed with controlled atmosphere or inert gas to unload difficult products. Three, four or more units give a continuous plant for a production above 8 t. for example.

A. P. Longmore, U.K. : Fig. 3 shows a two-section plant which can only be charged at infrequent periods. Can Dr. Oetjen explain what the advantages are of the system shown as compared with two separate batch chambers ?

G.-W. Oetjen, Germany : As shown, the two-tunnel unit needs only one loading area and only one unloading area but operates otherwise like a batch unit.

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A. P. Longmore, U.K. : Dr Oetjen said, that there is no special merit in a short drying cycle as compared with a long cycle, due to the high proportion of the cost of freeze­drying being involved in the capital cost and amortization of the equipment. Therefore, the higher the throughput per unit time the lower will be the cost of freeze-drying and this can be appreciable.

G.-W. Oetjen, Germany : I did not make a statement like this. I tried to explain only, that the advertised shorter and shorter cycles become very often uneconomical: If a load of 15 kg/m2 can be dried e. g. in 8 hrs, but a load of 45 kg/m2 in 16 hrs (even at a lower heating temperature), it is more economical to take 45 kg/m2 in 16 hrs than to use the 8 hrs cycle. In addition to a better use of the equipment the longer cycle needs less handling per kg and makes better use of inert or dry atmosphere etc.

A. P. Longmore, U.K. : Can Dr. Oetjen give us some indication of what he considers to be a reasonable cycle time for some typical products ?

G.-W. Oetjen, Germany : The figures used in 3) could be taken for meat under certain conditions.

G. Seffinga, Netherlands : I fully agree with Professor Goldblith and Dr. Oetjen that both equipment manufacturers and food processors must not go after short drying cycles. Nowadays freeze-drying equipment can be built where the energy costs amount to about 30 % of the total drying costs. The investment costs of the vacuum chamber amount to about 30% of the total plant and the rest of the equipment is not seriously affected by the drying-time for a certain constant daily output. So the higher investment costs for bigger drying-chambers will only have an influence of some per cent on the rentability of the plant.

Batch type freeze-drying plants can keep up very favourably with continuous processes when the batch type drying plants have the facility to freeze the products inside the drying-cabinet. In this case it is possible to build bigger and more economic drying cabinets because the time for loading and evacuation of the chamber is not longer a limiting factor for the size of the cabinets.

G.-W. Oetjen, Germany : I cannot agree with your second statement : In a continuous plant either a simple pre-freezer or the air-lock itself can be used as the freezing facility (as well for blast- or for contact-freezing). If you use the freeze-dryer itself for this purpose, it is less economical, since your investment for the freeze-dryer stands idle during the freezing operation.

W. Bedert, Switzerland: Discussion concernant l'economie d'energie a realiser par un procede lent (de lyophilisation) : On utilise surtout le procede de lyophilisation pour obtenir un produit de qualite. Mais ii y a des cas ou celle-ci ne peut etre obtenue qu'avec un procede de lyophilisation tres rapide. La vitesse peut done jouer un r6le essentiel. Avec un procede lent on peut obtenir ainsi des produits d'une qualite meme bien inferieure a celle obtenue par un procede conventionnel de sechage et ainsi bien moins cher.

Vue les frais fixes tres eleves du procede de lyophilisation le petit gain realise semble en outre tres illusoire ! Ceux-ci demandent avant tout une grande production par heure c.-a-d. une bonne utilisation de !'installation.

Pour le moment done et surtout dans le cadre de ces conferences une discussion a ce sujet semble prematuree.

G.-W. Oetjen, Germany : I feel personally, that there are facts throughout the world to justify a different opinion.

H. F. Th. Meffert, Netherlands : Dr. Oetjen, do your statements about the energy costs of 2,5-3,5 Pf/kg and the distribution of costs over pretreatment, handling and packaging, and actual freeze-drying, correlate with each other, taking energy cost in freeze-drying cost as 1 / 3 ?

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G.-W. Oetjen, Germany : Energy cost in freeze-drying is about the same as amortiza­tion and interest for the freeze-drying equipment or 50% of energy, amortization and interest.

The 1/3 figure includes some other cost and is therefore not directly convertible.

PACKAGING AND STORAGE OF THE FREEZE-DRIED MATERIAL

Microbiological aspects of freeze-drying

M. Kondrup, Denmark : Referring to Dr. Goldblith's statement on the importance of microbiological specifications for freeze-dried foods, I believe this point should be enlightened. Following the preparation stage, the freeze-drying process roughly speaking has two stages : freezing and drying by sublimation. During the freezing stage micro­organisms cannot multiply, and after drying life of microbes also will be quite difficult. Isn't so, that microbiological specifications actually will tell how the raw material was just before freezing, and that the freeze-drying process itself cannot make the micro­biological state of the food any worse. Of course, microbial growth may again start during and especially after rehydration, but until then nothing will happen to the freeze-dried foods, and plate counts will just disclose which microbiological state had the food before the freeze-drying process proper began.

(The replies received from the auditory seemed to contain the above point of view).

G. Lusk, U.S.A. : Bacteriological specifications are important to indicate the quality of the raw material which is so important in determining the storage life of the freeze­dried product. Bacteriological specifications are also necessary for precooked freeze­dried products because they are ready-to-eat.

W. Bedert, Switzerland : Quelles sont les exigences a observer du point de vue bacterio­logiques pour un produit lyophilise? (Question posee par M. Kondrup).

Les memes precautions doivent etre observees qu'ont doit prendre egalement pour un produit fabrique par un autre procede de sechage (pulverisation etc.). Ceci avant, durant et apres le procede c.-a.d. le produit final. (Par exemple poudre de lait).

J. F. Hearne, U.K. : There are various procedures for breaking the vacuum and trans­ferring the product to the package, e. g.

1) the vacuum in the chamber could be broken with dry air and the product separately packed in nitrogen or other inert gas ;

2) the vacuum could be broken with nitrogen, and the product then packed in nitrogen;

3) the product could be handled throughout in nitrogen as in the procedure outlined by Dr. Oetjen.

I would like to ask Dr. Oetjen whether he could give some specific examples of the extension in storage life of foods, or improvement in their quality arising from his method.

U. Hackenberg, Germany, answers for Dr. Oetjen : "Direct controlled atmosphere handling and packaging" of freeze-dried products does not mean that in any case the oxygen and/or water content of that controlled atmosphere was to be at the lowest possible level. For instance, the controlled atmosphere may consist of dry air, depend­ing on the product and the storage life wanted. Direct controlled atmosphere handling and packaging only means breaking the vacuum in the freeze-drier at the end of the drying process with that gas or gas vapor mixture which shall be in the package and handling the product in that controlled atmosphere without interruption.

W. Spiess, Germany : Mr. Hearne has asked Dr. Oetjen for some specific examples about the extension of the storage life of freeze-dried foods, when they are stored under different atmospheres. Salwin (Quartermaster Food and Container Institute, Chicago)

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has done a lot of work which covers to some extent this field, and also in the Federal Research Institute for Food Preservation in Karlsruhe we did a couple of trials with freeze-dried carrots, meat and other foods. The investigations have not been finished up to now, but they show that there is a real relation between the oxygen content of the surrounding atmosphere and the deterioration of the material.

The following table gives a rough survey on the changes during the first 8 months :

Storage time days 02-concentration in the storage atmosphere (Vol. %) 21 % 2% 1 % 0,1 %

0 6.6 6,6 6,6 6,6 30 1,5 2,8 5,3 5,9 60 0,6 2 4,8 5,5 90 0,2 1,8 4,4 5,2

120 0,1 1,5 3,9 5,0 150 0,1 1,5 3,9 4,8 180 0 1,5 3,8 4,8 210 0 1,45 3,6 4,6 240 0 1,45 3,6 4,5

W. Bedert, Switzerland: M. Spiess repond a M. Hearne qu'il n'existe que peu de litterature concernant le sujet en question (gazage/conservation).

J'aimerais quand meme faire allusion qu'il existe entre autre un travail, dans lequel i1 est demontre combien i1 est difficile d'enlever les restes de gaz contenus dans les poudres (poudre de lait, oxygene : gazage lait en poudre). Meme en repetant le vide et le gazage avec des gaz inertes on n'y parvient pas facilement.

C'est de Ia d'ailleurs que resulte la necessite de couper le vide avec des gaz inertes. Un autre probleme est la correlation entre la teneur en eau des produits alimentaires

lyophilises et l'oxygene de }'atmosphere d'entreposage. Un grand nombre de procedes d'autoxydation ne peuvent se derouler qu'en presence

d'eau. Ce fait c.-a.d. la teneur en eau joue done un role important dans la conservation des

aliments (lait en poudre, extraits de cafe, ect.). II ne vaut done pas la peine de douter ces faits bien connus et d'etendre la discussion

a ce sujet.

J. Strasser, Germany : Closer investigations on the necessity of packaging dried vegetables under nitrogen are given in the publication of P. Gorling : Die industrielle Obst- und Gemiiseverwertung 47 (1962), No. 23, p. 703-709.

THERMODYNAMIC PROBLEMS CONCERNING THE FREEZE-DRYING PROCESS

J. Strasser, Germany : Question to Professor Rey about the influence of noncondens­able gases on the heat transfer and on the drying rate.

L. Rieutord, France : L'accroissement de la pression des gaz incondensables dans la cuve de dessiccation n'avait pas pour objet la reduction des temps de dessiccation, mais elle avait pour but de faciliter le transfert des calories depuis les plaques chauffantes jusqu'au front de sublimation dans les produits, et de permettre d'abaisser la temperature de ces plaques chauffantes. II ne semble pas que ce procede ait apporte, dans la plupart des cas, une reduction appreciable de la duree de dessiccation, sauf cependant, dans le cas de produits faciles a lyophiliser qui ne pouvaient etre portes a la temperature des­irable par le simple effet du rayonnement des plaques, et qui par !'injection de gaz incondensables, peuvent sans difficultes etre portes a cette temperature, sans elevation excessive de la temperature des plaques chauffantes.

J. Fabian, Czecholovakia: To the connection of mass transfer and heat transfer as mentioned in Professor Rey's paper (observations of M. Rieutord with increasing the partial pressure of noncondensable gases) I feel that the input of noncondensable gases

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into drying-chamber must decrease the rate of the mass transfer. In my experiments with the regulation of drying temperature by input of pure nitrogen the time of the regulation at -5° C was longer than of the regulation at -20°C, e. g. the evaporation rate of water at -5° C was probably lower than that at -20°C.

H. F. Th. Meffert, Netherlands: In the Russian literature an equation is given in terms of:

Nu = Const · Re · Gu0 Er By the flow of gas or water vapor the value of Re is increased, so we have a larger heat transfer coefficient.

J. Strasser, Germany: We have measured in several experiments that there is a slow­ing-up of the drying rate if a greater amount of air or other non-condensable gases is admitted to the drying chamber.

During the freeze-drying process the partial pressure of non-condensable gases is much lower in the vicinity of the product than in the vicinity of the condenser.

L. Rieutord, France : J'ai dit, tout a l'heure, que le procede d'injection de gaz incond­ensables dans Ia cuve de lyophilisation, ne semblait pas apporter une reduction apprec­iable de la duree des operations, mais je dois souligner, maintenant, qu'une experience de trois ans, a l'echelle industrielle, sur de nombreuses installations fonctionnant avec ce procede, a demontre que l'accroissement de pression, raisonnablement limite, n'ap­porte pas, non plus, d'augmentation de la dun�e des dessiccations malgre la crainte que !'on pouvait avoir a ce sujet, !ors des premieres experiences.

L'accroissement de conductibilite des gaz dans la cuve s'explique par le fait que leur accroissement de pression place ces gaz en regime de transition entre le regime mole­culaire et le regime de l'ecoulement visqueux.

La courbe montre que le regime de transition entre le regime moleculaire et le regime visqueux est celui qui correspond a l'accroissement le plus rapide du coefficient de conductibilite des gaz en fonction de l'accroissement de leur pression.

Dans les conditions habituelles des appareils de lyophilisation, les intervalles entre les plaques chauffantes et les fonds des boites, d'une part, et ceux entre les fonds des boites et les fonds des flacons d'autre part, sont de l'ordre de quelques dixiemes de millimetres. Les libres parcours moyens des molecules gazeuses aux pressions habitu­elles de !?oP a !?ooP sont egalement de cet ordre.

On se trouve done bien, dans les intervalles consideres, et dans ces intervalles seule­ment, dans les conditions avantageuses d'un regime de transition.

H. F. Th. Meffert, Netherlands: The equation fitting the curve given by Mr. Rieutord is:

1 + 2 �· (�- y-_)1) A = Ao ----

given by Kessler (Diss. Darmstadt, 1961); L1 represents the mean fee path of the mole­cules depending on the pressure, d = the distance of the surfaces. So you get a couple of curves as Mr. Ehlers has shown (5. Gefriertrocknungstagung, Leybold, Koln, 1962).

L. Rey, France: Le principal resultat de !'injection de gaz incondensables est de permettre un transfert de chaleur plus homogene et plus intense sans elever de fa�on excessive la temperature des plaques chauffantes. L'acceleration de !'operation qui peut en resulter est due, evidemment, a une elevation de la temperature de sublimation.

R. I. N. Greaves, U.K. : I agree with Professor Rey's explanation. In a simple ex­periment of drying a frozen block of milk between radiant heater plates at + 70°C, by injecting air I could put far more heat into the heater plants without raising their tem­perature above + 70°C. But the temperature of the milk rose from -35°C to -20° C. By this means the speed of drying was increased without danger of overheating the dried surface.

102

II. Commission 1

Scientific problems of low temperature physics and thermodynamics. Cryogenic engineering.

Problemes scientifiques relatifs a la physique et a la thermodynamique des basses temperatures. Techniques cryogeniques.

SESSIONS :

Liquefaction

Applications to Nuclear

Physics

Thermodynamical

Properties

Magnetism

Liquefaction

Application a la physique

nucleaire

Proprietes thermodyna­

miques

Magnetisme

L iq u efacti o n Liqu efact i o n

Cryogenics and Space Technology *

Cryogenie et technologie spatiale

R. B. SCOTT Cryogenic Engineering Laboratory, National Bureau of Standards, Boulder, Colorado, U.S.A.

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SOMMAIRE. Le Laboratoire de Cryogenie du »National Bureau of Standards« pro­jette plusieurs experiences relatives au programme spatial des Etats-Unis. La plupart de celles-ci resultent de la decision d'utiliser de l'hydrogene liquide comme agent de propulsion des fusees. Parmi les plus interessantes de ces experiences, on peut citer:

Mesure des proprietes du parahydrogene a l'etat liquide et a l'etat gazeux entre 15° et 100° K a des pressions atteignant 340 atm. Ces mesures comprennent les relations p-v-t, la chaleur specifique, la conductibilite thermique, la viscosite, la vitesse du son et la constante dielectrique. Des tableaux et des diagrammes des proprietes thermodynamiques derivees sont en cours de preparation.

Un type de projet tres different est l' etude des difficultes rencontrees lors de l' essai de refou­lement d'une vapeur seche provenant d'un reservoir d'hydrogene liquide de propulsion en !'absence d'un champ de gravitation pour effectuer la separation de phase. Il est apparu qu'un separateur centrifuge conviendrait, mais cela a entrains un autre probleme: celui des paliers qui fonctionneront dans /'hydrogene liquide OU gazeux a des temperatures voisines du point d'ebullition de l'hydrogene.

Comme il sera necessaire de refouler l'hydrogene dans le vide spatial on craignait que les gouttelettes de liquide entrainees ne vinssent a se congeler et a boucher la conduite. Une etude de laboratoire a montre que cette crainte etait justifiee.

When liquid parahydrogen was selected as a propellant for some of the more advanced space vehicles and rocket stages, it became apparent that much of the information needed by the design engineers was either inadequate or non-existent. The many properties of parahydrogen were not precisely known, or known only in very limited regions of pres­sure. Experimental work was needed to predict the behavior of liquid and cold gaseous hydrogen in the tanks, pumps, valves and instruments of the rocket system. New special purpose instruments were needed. The properties of many materials of construction were not known in this temperature region. The behavior of moving mechanisms could not be predicted.

Because of these circumstances, the National Bureau of Standards Cryogenic Engineering Laboratory, with support from the National Aeronautics and Space Administration, undertook several experimental projects specifically designed to provide some of the more urgently needed information.

PROPERTIES OF PARAHYDROGEN

A. Pressure-Volume-Temperature Relations Fig. 1 is a cross section of the piezometer region of the PVT cryostat [1]. The sample

is contained in the thick-walled, 25 ml copper piper, which can be cooled by admitting

* Supported in part by the National Aeronautics and Space Administration

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hydrogen into the reflux tube so that condensation on the upper part and evaporation at the pipet transfers heat from the pipet to the H2 tank. The temperature of the sample is measured with a strain-free, capsule-type platinum resistance thermometer. The pres­sure is transmitted through the capillary tube to pressure measuring equipment con­sisting of a precision piston-type pressure balance. The hydrogen is isolated from the oil in the pressure balance by a sensitive diaphragm fitted with electrical contacts to in­dicate the null position. The experimental procedure consists of filling the pipet with parahydrogen at the desired starting temperature and pressure, adjusting the tempera-

- CAPILLARY

,E:�l -i-- __ LIQUID HYDROGEN

· _ ,.. = 1 - ' --- fl -

• ,.-

COLD RING

-PIPET

,COLD \VALL

DEWAR JACKET

Fig. r. The P-V-T Cryostat

P-V-T CRYOSTAT ture to an exact integral Kelvin temperature, and reading the pressure. Next the tem­perature is raised to a higher integral temperature and the pressure again read. This is continued until a pressure of about 320 atm or a temperature of 100°K is reached. After completing such a set of measurements, the amount of hydrogen in the pipet is measured by releasing it into a gasometer system consisting of a set of calibrated spherical glass bulbs and a precision manometer. Both the bulbs and manometer are maintained at constant, uniform temperatures.

Next, the pipet is filled to a different density and another set of measurements made. Each such set of measurements constitutes a pseudo-isochore since the density remains constant except for the small decrease caused by the slight amount of fluid that flows into the capillary and diaphragm cell as the pressure rises.

To obtain true PVT values from the observed data, adjustments must be made for the following: (1) elastic stretching of the pipet, (2) thermal contraction of the pipet, and (3) the fraction of the total sample in the capillary and diaphragm cell. Details of these adjustments are given in [1]. Final results are reported by Goodwin, Diller, Weber and Roder [2].

B. Specific Heat at Constant Volume Fig. 2 is a cross section of the lower part of the calorimeter cryostat [l]. It is very

similar to the PVT cryostat except for the sample container itself, which is a sphere of type 316 stainless steel having a normal volume of 72.35 cm 3 and 1.5 mm walls. Both the inside and outside of the sphere are copper plated to a thickness of 0.2 mm to increase heat conduction. A 100-ohm constantan heater is varnished directly onto the sphere and shielded by the lightweight calorimeter case. During measurements of C v, the shield, calorimeter, and guard ring are maintained at the same temperature by automatic con­trollers responding to signals from differential thermocouples.

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The sample holder is cooled during loading by means of helium exchange gas in the spaces inside the cold wall and shield. (These two spaces communicate.) This cools the sample holder so that hydrogen may be condensed into it. The sequence of measure­ments is similar to that of the PVT determinations. For each filling of the calorimeter, the heat capacity as a function of temperature is determined in the usual way, termina­ting a run at a temperature of about 100° K or a pressure of 300 atmospheres, whichever is reached first. The calorimetric data require adjustment for the same imperfections of apparatus as did the PVT measurements. For the C v data, these adjustments also in­clude compensating for the work done by the fluid as it expands. Results are given by Younglove and Diller [3, 4] and Goodwin [5].

_.... COLD RING

SHIELD

,....CAPILLA RY

--.CASE

�-���---'-THERMOMETER Fig. 2 . The Cv Cryostat

CALORIMETER CRYOSTAT C. The Velocity and Absorption of Sound in Liquid Parahydrogen

The velocity and absorption of sound of liquid parahydrogen will be measured from the triple point temperature, approximately 14°K, to 100°K and for pressures from about 10 atmospheres to 340 atmospheres. The method uses ultrasonic pulse methods, originally developed by Pellam in 1946 at MIT and Pinkerton at Cambridge at the same time. A fixed path length is used and the pulse train is long enough so that the signal frequency can be adjusted for constructive interference by comparing the relative phase of the signals of the first and third reflections of the sound wave as registered at the re­ceiving crystal. This is the method of McSkimin of Bell Labs., who works mainly on solids (Fig. 3). Attenuation is determined by comparing the amplitudes of the signals received at the first and third reflections.

Ampl i f i e r

Gate

Osci l l ator 0 - 30 Mc

Ampl i f i e r H i g h Goin

Decode Atten uator

Osc i l loscope

Fig. 3. Diagram of Apparatus to Measure the Velocity of Sound in Fluids

The emitter and receiver of the sound wave are x-cut quartz crystals and the path length is set by an extremely well-made spacer of fused quartz. This is a cylinder with optically flat ends. The length, nominally 6 cm., was measured to 1 part in 150,000 (an

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average of four measurements) and the ends are parallel to within less than one-half micron. The variation with temperature and pressure will be accounted for so that the length should be known to 0.001 %. Special precautions were taken in design of dimen­sion of the cell and in choice of frequencies so that the velocity of sound values would be accurate to somewhere in the range 0.05 % to 0.01 %, with greater inaccuracy near the critical point. Accuracy of attenuation will be within a few percent.

D. Viscosity The viscosity of hydrogen is being measured by a method proposed by W. P. Mason

[6] in which a cylindrical quartz crystal is caused to undergo torsional oscillations near its resonant frequency. The damping of the oscillations by the fiuid surrounding the

crystal may be measured in terms of the width of the resonance curve of the crystal. The viscosity-density product of the fluid is related to the bandwidth of the resonance curve by the following expression [7] :

[7) (.!

] = 1: (�)2 (LJ/) ', where f o is the crystal's resonant frequency, M its weight, S its surface area and LJ/ the

bandwidth due to the fluid damping. The measured LJ/ is adjusted for the damping in vacuum caused by mounting losses and the internal friction of the crystal.

Fig. 4 shows a cut-away sketch of the quartz oscillator with a schematic diagram of the measuring circuit. The crystal is suspended at a nodal plane by nylon cords so

35 - 45 kc Sinusoidal 1 -3000 mv; Stability • 10-e

Nyfon <. Scope Support

Thread

XCElectric) Axis

Grooves for Support Thread

Teflon Supporting

Cylinder

QUARTZ CRYSTAL OSCILLATOR

Fig. 4. Apparatus for Measurement of Viscosity

that it can vibrate as freely as possible. The measuring circuit consists essentially of an a. c. bridge for the determination of the crystal's electrical properties and an electronic counter for frequency measurements.

The crystal is shunted by a capacitor to keep the reactance capacitative at frequencies greater than the resonant frequency. The measuring arm of the bridge consists of an accurate decade resistance box in parallel with a variable capacitor. The precision of the measurements depends greatly on the quality of the bridge oscillator and tuned null detector.

The crystal is contained in a sample holder very similar to the PVT piezometer and the same instrumentation is used for the measurement of temperature and pressure. The experimental procedure is also similar to that used for the PVT work except that the previous PVT data are used to compute the density of the hydrogen.

E. Dielectric Constant of Fluid Hydrogen Apparatus is being constructed for the measurement of the dielectric constant of pure

parahydrogen in the liquid, gaseous, and super-critical regions. It is planned to measure the capacitance of a cylindrical capacitor containing hydrogen at a known pressure and temperature. A capacitance bridge, newly on the market, capable of six-figure relative precision, will be used. Since the capacitor contains no dielectric material between its

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measuring plates, the dielectric constant, e, of the hydrogen can be obtained as the ratio of capacitor with sample to capacitor evacuated. It is hoped that the quantity e - 1 can be determined to within a few units in the fourth decimal place.

The previous data on hydrogen can be represented to the precision of the measure­ments [8] by the Clausius-Mossotti equation,

e - 1 1 4 :n:

B + 2 • e = 3 M a,

where e is dielectric constant, e density, M molecular weight, and a the molecular pola­rizability, assumed constant. It is expected that the present measurements will provide a more sensitive test for departures from the equation that are to be expected.

F. Engineering Studies

In addition to measurements of the physical properties of cryogenic fluids such as those just described and many properties of solids which will not be discussed, the Cryo­genic Engineering Laboratory has undertaken a number of purely engineering tasks in support of the space program.

Descriptions of two of these will illustrate some of the peculiar problems encountered. The designers of space vehicles early recognized the possibility that difficulty might be expected in venting vapor (without liquid) from a cryogenic propellant tank in the ab­sence of any gravitational field to effect phase separation. To solve the problem they designed centrifugal separators. These required bearings that would operate reliably at low temperatures in either liquid or gas. The Cryogenic Engineering Laboratory had previously made extensive studies of ball bearings running at speeds as high as 9200 RPM submerged in liquid nitrogen [9]. Additional tests in dry gaseous hydrogen showed that these rearings would operate satisfactorily if kept cool by a stream of hy­drogen gas [10]. The minimum cooling gas requirement was found to be 1/4 to Y2 liter per second at 20° to 90° K for a bearing 26 mm OD with a IO mm bore, supporting a thrust load of about 20 kg.

Another problem associated with the use of cryogenic liquids in space vehicles is the formation of solid when the triple point pressure is reached. It is possible for the solid to block passages or cause malfunction of pumps or valves. The Cryogenic Engineering Laboratory is conducting laboratory studies on such solid formation.

REFERENCES

r . R. D. Goodwin, J. Research, Nat. Bur. Stand. 65 C, 23r (r96r).

2. R. D. Goodwin, D. E. Diller, L. A. Weber and H. M. Roder, J. Research, Nat. Bur. Stand. (Spring 1963).

3. B. A. Younglove and D. E. Diller, Cryogenics 2, 283 (r962).

4. B. A. Younglove and D. E. Diller, Cryogenics 2, 348 (r962).

5. R. D. Goodwin, Cryogenics 2, 353 (1962).

6. W. P. Mason, Trans. Am. Soc. Mech. Engrs. 69, 359 r947.

7. B. A. Welber, Phys. Rev., r r9, r8r6 (r960).

8. R. ]. Corruccini, Nat. Bur. Stand. Technical Note No. 144 (1962).

9. ]. A. Brennan, W. A. Wilson, R. Radebaugh and B. W. Birmingham, Advances in Cryogenic Engineering 7, 262 (1962).

ro. L. E. Scott, D. B. Chelton and ]. A. Brennan, Advances in Cryogenic Engineering 7, 273 (1962).

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The NBS * Cryogenic Data Center **

V. J. JOHNSON Chief Cryogenic Technical Services U. S. Department of Commerce, National Bureau of Standards, Boulder Labora­tories, Boulder, Colorado, U.S.A.

The two principal activities of the Cryogenic Data Center are :

1) The evaluation and compilation of data on low temperature properties of mate­rials, and

2) The acquisition, handling and dissemination of technical literature.

We started the program of compilation of thermophysical property data about 5 years ago. A modest but capable staff has now been organized and developed. Their efforts have been concentrated primarily on the thermodynamic properties of the cryogenic fluids but transport properties and solids have been covered to a limited extent as well. Tables and charts for helium, hydrogen and nitrogen have been prepared. Tables and charts for oxygen, argon, neon, and carbon monoxide are nearing completion. Most probable values are being compiled for several of the transport porperties of these same fluids. Values of electrical resistivity for a number of selected solids are also being assem­bled.

The compilation staff have acquired considerable competence in the use of many techniques of data evaluation. The theory of corresponding states of course is a valuable means of estimating property values of a fluid for which very little experimental data exist by comparing it with a similar well characterized fluid. For example, nitrogen and argon P-V-T surfaces are quite well established and these are serving as models for cal­culating neon, oxygen and carbon monoxide properties. A number of modifications to the conventional theory have been made as well as the use of the multiterm type of equation of state in this work. These same techniques are also very useful in evaluating experi­mental data.

The literature service activities which were started primarily for the data compilation program were soon expanded to serve all the projects of the Cryogenic Engineering La­boratory as well as much of the cryogenic industry. The principal functions are:

Fig. 1 is of the data evaluation staff discussing the multiterm equation of state. R. B. Stewart (at the blackboard) is in charge of the data compilation activity.

Fig. 1, Staff discussing the multiterm equation of state

* U. S. National Bureau of Standards ** This paper was presented as a supplement to Mr. Scott's paper 1 - 3 110

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1) The acquisition of publications, reprints and reports from worldwide sources. 2) Thorough coding of this material for subject matter and data content for storage,

retrieval and use. 3) The preparation of bibliographies on specific subjects. 4) The distribution of CEL literature of the cryogenic industry.

Only a limited awareness service is now being provided to our own staff and to the industry. Considerable expansion of the current awareness effort is planned to obtain maximum value from our literature program.

The following pictures show some of the Data Center staff and the operation. Fig. 2 shows some of the forms in which the data are disseminated; such as charts,

compendia, conference proceedings, technical reports, and in primary publications.

Fig. 2. Forms in which the data are disseminated : charts, reports etc.

Fig. 3 is a view in the· literature procurement and records room. The input supervisor is checking one of the more than 16,000 master catalog cards on file. Last year some 2,500 reports and reprints were received. These plus about 1,500 other articles in the current periodical literature were entered into the system. The Visifile directly in front of her contains authority lists essential for maintaining consistent citations. The World Literature File of about 7,000 documents and the Microfilm File of nearly 5,000 articles are also located in this room.

Fig. 3. Literature procurement and records room

1 1 1

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Fig. 4 is a view in the "machine room" where Flexowriter tapes are prepared for each accession. Also the coding is punched onto IBM cards for transfer to the magnetic search tape.

Fig. 4. Preparation of tapes for each accession

Fig. 5 shows the Flexowriter tape storage in the back. The tapes are used for print-out of bibliographies and preparation of lists of documents including a catalog of references.

Fig. 5. Tape storage

Fig. 6 is a picture of the Boulder Laboratories Computer Facility where our storage tapes are searched and accession numbers of references for specific queries are compiled.

Fig. 7 is a view of the CEL document supply room. The lady in charge is checking one of the nearly 4,000 address cards of persons and companies who regularly receive announcements of the CEL literature. Over 300 separate documents are now available. We distribute some 25,000 items annually in response to nearly 2,000 re­quests. We have a stock of 45,000 documents and 20,000 charts on hand. Our objec­tive is to facilitate and expand the dissemination and exchange of information of cryogenic interest as much as possible.

1 12

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Fig. 6. Computer room

Fig. 7. Document supply room

DISCUSSION

N. Booth, U. K.: Would Mr. Scott care to comment on any work which the Bureau of Standards has done or is doing on the safe handling of liquid hydrogen?

R. B. Scott, U. S. A . : We are very safety-conscious and have collected all the informa­tion that we can find on accidents and hazards in the use of liquid hydrogen. From this information we have prepared a document on safety which is available to the public.

A. v. Itterbeek, Belgium: You mentioned research on waveguides at low temperatures. Is that concerned with superconducting waveguides ?

R. B. Scott, U. S. A . : Yes.

A . v. Itterbeek, Belgium: What was the method used for the measurements on the dielectric constant of liquid hydrogen? Would it be affected by changes in the dimensions of the apparatus when pressure is applied to the hydrogen ?

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R. B. Scott, U. S. A.: The only dimensional changes that can occur should be very small, because the pressure is applied to both sides of the condenser plates. Thus the only change will be that resulting from the compressibility of the metal electrodes.

J. Wilks, U. K.: Have you experienced any difficulty with your quartz crystal visco­meter due to spurious damping, such as been observed in liquid 3He and 4 He ?

R. B. Scott, U. S. A.: We found spurious results which seem to arise from standing waves around the circumference of the crystal cylinder. We are avoiding these resonant frequencies, and will cover the missing areas by using cylinders of different dimensions.

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Liquefaction of Hydrogen and Helium for Nuclear Applications

Liquefaction de l'hydrogene et de l'helium pour les applications a la physique nucleaire

DR. W. BALDUS Gesellschaft fii.r Linde's Eismaschinen Aktiengesellschaft, Hollriegelskreuth b. Miinchen, Germany

SOMMA/RE. La physique nucleaire et les techniques nucleaires s'occupent de plus en plus de probtemes d'une importance particuliere en relation avec les tres basses temperatures. Ces problemes sont, entre autres, l' etude des reactions nucleaires, des coupes, des alterations aux radiations, des neutrons lents et des proprietes des materiaux. Les installations frigorifiques necessaires a ces etudes doivent repondre a des besoins specifiques, tels que puissance frigorifique variable en f onction de la temperature et du temps, observation des conditions recommandees du point de vue de l'agent refrigerant, des besoins de place et de la securite. Les problemes sont resolus surtout par des liquef acteurs speciaux d' hydrogene et d' helium. A l' aide d' exemp­les, on etudie les divers types d'installations, y compris ceux fonctionnant avec un gaz froid comme frigorigene.

(1) INTRODUCTION

There has always been a mutual relationship between the physics of low temperatures and the technology of generating such low temperatures. In recent times, also the devel­opments in the field of nuclear physics have given rise to a number of questions which are closely related to low temperature processes. For applications in nuclear physics it is mandatory that specific conditions be fulfilled in producing the low temperatures. In­vestigations of nuclear problems involving low temperatures are still directed to a great extent towards arriving at new findings in basic research. They include, among other things, measuring of effective cross-sections, angular distributions, energy conditions, and the life of the constituents of the nucleus and, moreover, investigations into all those fields in which nuclear physics borders on solidstate physics. Over and above that, however, there are numerous examples of practical applications of low temperatures in nuclear engineering, such as the functions of low-temperature purification of reactor cooling and shielding gases, e. g. He, C02, Ar, or cleaning of gases for spark chambers. Another though less closely related field is the production of raw materials for nuclear engineering, for example the production of deuterium by distillation of liquid hydrogen or of helium from natural gas by means of a low-temperature process.

In connection with nuclear problems, special interest has centered round the produc­tion of very low temperatures such as can be obtained with the aid of liquid hydrogen or liquid helium. It is the object of this paper to discuss a few examples in this respect. The physical aspects of the problems will be discussed only to the extent that they are necessary for a better understanding of the requirements made of low-temperature equipment.

(2) BUBBLE CHAMBER

The term "bubble chamber" is generally applied to a container filled with a liquid through which charged elementary particles are injected at regular intervals, e. g. every 0.3 seconds. These particles usually come from a large accelerator for protons or elec­trons, from which they are emitted in the form of pulsed packets. Prior to the arrival of each new particle packet, the liquid in the chamber is in a superheated state. This is caused by periodic expansion, which is again followed by compression, of a piston pro­vided in a cylinder on the upper side of the chamber. As a result of this arrangement,

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the chamber is continuously supplied with heat. When the charged particles enter the superheated liquid, they form along their path bubbles which are photographed by a stereoscopic camera. The object of such a bubble chamber is to obtain information on the interaction of the charged particles among themselves, with their products and with those particles of which the liquid consists. Consequently, it is obvious that the results will be different if different charged particles and different liquids are used, and selection of both the type of charged particles and the liquid in the chamber will thus depend on the kind of processes to be investigated. The chamber is, in addition, usually located within a powerful magnetic field which causes deflection of the charged particles and hence of the paths of the bubbles they produce. The degree of such deflection permits further physical conclusions.

(a) Low temperature plant for hydrogen bubble chamber The first bubble chambers were filled with liquids having a relatively high boiling

point, e. g. propane. For a number of physical reasons which will not be discussed in this context, it was later found that liquid hydrogen, if used as a refrigerant, yields partic­ularly important results. The following description given by way of example deals with a hydrogen liquefaction plant ordered from the German Linde company for the specific requirements of a bubble chamber. As regards the dimensions of the low temperature plant, a number of characteristic requirements with respect to process and design had to be met.

Procedure The vessel of the bubble chamber consists of several tons of steel and, as mentioned

earlier, is provided with windows permitting photographs to be taken. This makes it necessary to cool the bubble chamber in accordance with a fixed time schedule prior to filling, in order to avoid undesirable material stresses. Cooling is effected by means of special heat exchangers provided inside the bubble chamber, through which increa­singly colder hydrogen is passed. It is thus the first function of the refrigeration plant to make this cooling process possible. The second step consists in filling the chamber with liquid hydrogen, thus preparing the plant for operation. In stationary operation the plant must further, as a third function, compensate the cold losses caused by the piston move­ment, the latter being required for periodic procurement of a superheated state. Finally, the plant has the fourth function of compensating for the insulation losses occurring in the process.

These four functions are performed by a refrigeration plant according to Fig. 1, which comprises 2 hydrogen liquefiers I and II and their ancillary equipment. Liquefier I first supplies liquid hydrogen to a storage vessel. Subsequently, liquefier II starts

r---- - - - - - - - -, - - . - -i I I I I I I I I

i--1- -1 I

lBI I I

fi- ,,,,,.,,.,, - Ll<>--i ILL _ _ _ _ _ _ _ _ _ -�I\ I / Ll 5torag"

t1quid 1>-Hi odtq"td p·H1 f55") abou.t 60l1fr1!>/hr

or liquid p•H1 (95"1 about 30/drl!s/hr

Ta t� slack.

Chambt1r

aboulttrJOW al 11•K

Fig. r . Flow sheet of hydrogen liquefaction and refrigeration plant for H2 bubble chamber

1 16

I-1

supplying increasingly colder hydrogen, in a closed circuit, through the heat exchangert inside the bubble chamber. Its refrigerating capacity is approx. 1 100 W at 21°K in a stationary state. To obtain the required cooling rate in the bubble chamber, liquid H2 from the storage vessel may either be passed through the heat exchanger in the bubble chamber or filled into the bubble chamber itself. Liquefier I may optionally supply either liquid n-H2 (i. e., H2 in high-temperature orthopara equilibrium) or liquid H2 with 55 % para-H2 (p-H2 (55 %) or p-H2 (95 %). In stationary operation the entire plant is fully automatic. Suddenly vaporizing quantities of liquid H2 are automatically dis­charged into bottles. The hydrogen required for liquefaction is first cleaned in two purifiers arranged for alternate use and thus permitting uninterrupted operation even if liquefier I is set for continuous operation.

Design Each of the two liquefiers and the two purifiers is located in an enclosed metal vacuum­

jacket vessel. The liquid hydrogen is fed to the bubble chamber through vacuum-insu­lated piping. Liquefier I is mounted on a mobile trailer, together with the compressors. The connecting lines to liquefier II are detachable so as to permit the latter, together with the bubble chamber, to be connected at different points of the accelerator. The entire plant is designed in conformity with general explosion-proof standards.

(b) Refrigeration plant for helium bubble chamber As far as equipment is concerned, a helium bubble chamber must fulfill the same 4

requirements that have been enumerated above in connection with hydrogen refrige­ration plants. In addition, there are a number of further items which will now be descri­bed in some detail on the basis of a process developed by the German Linde company.

In the case of hydrogen liquefaction plants it is still relatively simple to provide a sufficiently large temperature difference between the liquid hydrogen in the chamber and the hydrogen flowing through the heat exchangers in the chamber, whereas this causes considerably more difficulty in the case of helium. Over and above that, the very small heat of vaporization of helium (which is only approx. 1 /11 of that of hydrogen per standard cubic meter at atmospheric pressure) and a number of other factors make it appear expedient that the He to be filled into the chamber be available at the lowest possible pressure. The refrigerant flowing through the heat exchangers in the chamber must, consequently, possess sub-atmospheric pressure. Since, on the other hand, a comparatively high refrigerating capacity is required for the chamber, it is necessary to exhaust large amounts of vaporizing He by means of a vacuum pump and discharge them from the plant in separate suction lines.

Fig. 2 shows a flow diagram of the plant. The process employed is characterized by the fact that helium is the sole refrigerant used, which in this case is expanded by doing external work in 2 turbines. The plant thus includes no precooling system employing

(ompttJSMS

Cylmders

!iquid He

•---- &bbl.e 014mbtr ----- about 500W al eo�K

Fig. 2. Flow sheet of helium refrigeration plant for 3°K with helium as refrigerant only.

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liquid nitrogen or hydrogen. The turbines rotate at a speed of the order of 100 000 rpm. From the compressors the helium is fed to a system of heat exchangers arranged within a cylindrical evacuated tank, which supports the two turbines encased in separate hous­ings of their own. The cold tank is mounted on a framework structure above the bubble chamber. The plant ordinarily operates as a refrigerator, but may in case of need also be used as a He liquefier. The refrigerating capacity required for the bubble chamber in this case is approx. 90 W at 3°K and, in addition, 500 W at 80° K, and is automatically controlled. Special importance attaches to absolute tightness of the plant, since any major losses of He would be costly. A further important requirement of any such He refrigeration plant is that insulation losses be kept within minimum limits. This necessi­tates, among other things, special-design mechanical mounts for all parts and units pro­vided inside the tank. Losses due to radiated heat must likewise be minimized as far as possible.

(3) REFRIGERATION PLANT FOR COLD NEUTRON SOURCE

Certain experiments in nuclear physics require as large as possible a flux of very slow neutrons, for they permit scientists, by measuring effective cross-sections, to draw con­clusions as to the structure of matter. There are various ways of producing very slow neutrons. For reasons which cannot be discussed within the scope of this paper, the meth­od of cooling a small moderation volume to as low a temperature as possible has pro­ved most expedient. The moderator chiefly used for such a process is liquid hydrogen. A plant for producing very slow neutrons is called a "cold neutron source" and, in prin­ciple, comprises an enclosed vessel of approx. 1 litre (2 pints) capacity located in the immediate vicinity of the core of a reactor and filled with liquid hydrogen. The neutrons from the core enter the vessel and leave it again at very slow speed. These neutrons are now available for the experiments to be conducted.

In view of the supply of energy caused by insulation losses and by radiation emitted from the reactor, the hydrogen in the vessel must be constantly cooled by suitable means to keep it in a liquid state. This is effected by a cooling coil. It would be possible to pass through this cooling coil liquid hydrogen having a lower temperature than the hydrogen in the vessel. However, this would require the installation of a hydrogen liquefier within the reactor building, which is undesirable for safety reasons (installation outside the reactor building is not feasible). Therefore, a continual flow of cold gaseous helium is provided, which warms up as it passes through the coil.

The volume of the cooling coil is necessarily rather small. Hence, the area available for heat exchange is small as well. In addition, the coil must be of such geometric form that the scattering conditions for the neutrons are more or less uniform. Provision must therefore be made for a certain pressure drop of the cold helium gas in the coil which, in turn, requires the helium to enter the coil at an adequately high pressure. As a result, the specific object of the refrigeration plant consists in providing pressurized cold helium gas. For thermodynamic reasons, the absolute pressure should be as low as possible, e. g. 2 atm. (29.4 psi). Moreover, it is desirable to keep the quantity of gas flowing through the coil as small as possible, because then (in a given geometric form of coil) the pressure drop will also reach a minimum value. A small amount of gas, on the other hand, means that the cold helium gas (for a given refrigerating capacity) is heated more than would be the case if a large amount of gas was used. It will be obvious from the foregoing that a compromise must be found between two extremes :

large quantity of gas, high inlet pressure, high inlet temperature, little increased outlet temperature; and small quantity of gas, low inlet pressure, low inlet temperature, greatly increased outlet temperature at the end of the coil.

The choice among the various possible processes for producing the cold helium gas is governed by the cost factor, i. e., the designer must choose the process involving the least expenditure per watt of refrigerating capacity.

The process developed by the German Linde Company provides for a gas refrigera­tion plant which again employs helium as sole refrigerant, similar to the system already described in connection with the plant for the helium bubble chamber. Here, too, the

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equipment includes two turbines for expansion. The evacuated tank containing the essential parts of the plant is attached to the reactor. The refrigerating capacity is approx. 400 W at l6° K.

(4) FURTHER OUTLOOK

The types of plant discussed here constitute a few characteristic examples of specific new requirements for hydrogen and helium liquefaction plants in nuclear physics. Numerous further examples could be found without difficulty. The type and number of problems to be solved are constantly increasing. It will now be necessary, for instance, to expose targets in reactors to radiation at extremely low temperatures. In cases in which the target is to consist of solid hydrogen, the refrigeration plant will further have to provide for a solution of the problem of heat transfer from the refrigerant to the hydro­gen. Another important field worth mentioning is that of cryo-pump plants, which play a decisive role in plasma investigations. The problem in designing refrigeration plants appears to be a very simple one : provision of sufficiently large refrigerating capacity, if possible to obtain temperatures below 3 ° K. Practical realization of such a project, however, which has been one of the objectives of the German Linde Company for a considerable period of time, presupposes and entails the development of new principles in procedure, design and manufacture. They often deviate substantially from the time­tested principles applied in the construction of conventional low-temperature equipment and are thus a constant challenge for all who work in the field of low-temperature en­gineering.

DISCUSSION

J. B. Gardner, U. K. : What continuous running time may be expected from the hydro­gen refrigerator for the liquid hydrogen bubble chamber?

W. Baldus, Germany: 1 to 2 months or longer : two interchangeable purifiers are pro­vided.

J. B. Gardner, U. K. : Are the expansion turbines for the helium refrigerator for the liquid helium bubble chamber oil or gas lubricated ? Also what is their expected effi­ciency ?

W. Baldus, Germany: At present the cold end has gas bearings and the warm end oil bearings.

M. J. Doulat, France: Would not a vacuum-protected liquid hydrogen circuit in the reactor building and a liquefier outside be feasible ? I know three such installations at Harwell, Saclay and Grenoble, and they have worked satisfactorily for several years.

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A Simple Cooling System with a Cryogenic Pump

Un systeme frigorifique simple avec une pompe cryogenique

Dr. H. YONEMITSU, H. MAEDA and M. OHKA WA Central Research Laboratory, Tokyo Shibaura Electric Co., Ltd., Kawasaki, Japan

SUMMARY. On indique une application d'un systeme frigorifique simple avec unepompe cryogenique.

The following methods were employed in order to keep the temperature of super­conductor electric machines at 4.2°K or below:

(1) A large metal dewar contains both the superconductor devices and the liquid helium. This is called a dip-type cooling system. In this method, that dewar is neces­sary to have a large opening hollow, and therefore its containing ability is prevented. On the other hand it is very difficult to keep the maintenance of devices, if the inlet hollow is made small.

(2) As the ideal cooling system, it is very profitable to use a small helium refrigerator which contains the superconductor devices in the working cooling space.

But this helium refrigerator is very delicate and expensive. So it was examined how to obtain the simple cyclic cooling method by using a liquid helium storage dewar and

t

Fig. r . Cryogenic cyclic cooling system

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transfer tubes. Then the new cryogenic cyclic cooling system was succeeded by adopting a cryogenic pump as a working pump, as shown in Fig. 1. The cryogenic pump is of a plunger-type which is hanging with the three long and small supporting bars into the liquid helium bath and operated in the coolant.

A few of small opening ports are punched at the end stroke of its cylinder, and the coolant is transferred through the ports as the gravity flow.

It is unnecessary to use the usual suction value to operate as the vacuum suction of inlet part, on the other hand the ports themselves play a role of cut off valves. As the crank mechanism is kept outside the dewar, it is very easy to make remote control of the exhausting pressure and flow rate. The revolution speed is 60-100 rpm. Piston and cylinder are constructed by the fine finishing techniques in order to use the liquid as lubricant.

Between the end of delivery side of its pump and the heat exchanger enclosing the superconductor devices a double walled transfer tube is connected, and the return transfer tube is also connected between the heat exchanger and beginning dewar.

So the cyclic motion is able to drive the plunger pump. Temperature control is very easy, by combining the sensing part of temperature and the small driving motor con­nected to the crank mechanism.

The characteristics of this cryogenic pump are as follows :

(1) It is able to be operated having no connection with the vapor pressure of its con­tainer.

(2) It is very easy to control the exhausting flow rate, by keeping the linear relationship between flow rate and crank revolution.

(3) The succeeding speed is instantly controlled for maintaining the constant exhaus­tion.

(4) The temperature change is not caused by the non-compressibility.

(5) The usual syphon method by pressurizing the dewar vessel has such defects as follows :

(a) It is very difficult to control mechanically the flow rate by the pressurizing method, and even if it is possible, it is very expensive.

(b) It needs the gas cylinder or electric heater as a pressurizing source, but has a large time lag.

(c) Even if the pressurizing action were removed, the inner pressure of container can not immediately return to the initial condition. so the liquid flow cannot instantly be stopped.

(d) By this method it is not possible to make circulation of liquid helium.

(e) If the container is unsuitable for pressurizing, the syphon method is not used.

On the other hand, our cryogenic pumping method has many superior charac-teristics as compared with the simple cyclic refrigerator.

ACKNOWLEDGEMENT

The authors are grateful to Mr. Jiroo Okada, the manager of the Liquid Oxygen Chemical Engineering Co., Ltd. Tokyo for the construction of the cryogenic pump.

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Helium Refrigerator for the Production of Cold at Temperatures down to 2.5°K

Refrigerateur a helium pour la production de froid a des temperatures atteignant 2,5° K

A . SPIES Gesellschaft filr Linde's Eismaschinen Aktiengesellschaft, Hollriegelskreuth b. Mtinchen

SOMMA/RE. On decrit une installation utilisant le cycle d'hilium f enne. Elle a une puissance jrigorifique de 25 watts environ entre 2,5 et 4,5° K et elle est confUe de fafon qu'environ 7 l/h d'hilium liquide puissent etre retires si elle fonctionne uniquement sur la liquefaction.

Des modifications appropriees a un liquejacteur de MEISSNER a 3 l. d'helium donnent une augmentation de la capacite de liquefaction et de la puissance Jrigorifique en fonctionne­ment continu si les vapeurs d'helium sortant d'un cryostat retournent au liquefacteur.

Si les echangeurs de chaleur sont de dimensions appropriees et s'il est prevu dans le cycle d'hilium une pompe a vide pour abaisser la pression de vapeur de /'helium liquefie, on peut obtenir au COUTS du fonctionnement continu des temperatures d'ebullition de 2,5 a 4,5°K environ.

On indique des details sur la construction et des valeurs de fonctionnement.

INTRODUCTION

The purpose of this research work was to build a helium refrigerator capable of generating about 25 watts within the temperature range from 3.5° K to 4.5° K

The installation is primarily intended for cryo-pumping of air and hydrogen in an effort to obtain a vacuum of at least 10-6 Torr. Recent measurements have shown that for the condensation of larger quantities of hydrogen with vacua of about 10-6 Torr the normal boiling temperature of the liquid helium is not sufficient so that necessarily lower temperatures have to be used. From known vapour-pressure measurements of hydrogen it can be concluded that a temperature of 3.5° K is sufficiently low to satisfy the requirements specified.

Experiences gained with the MeiBner helium liquefier have been very helpful in devel­oping this apparatus. It appeared convenient to retain the principles of this liquefier consisting in the precooling of the pressure helium by means of liquid nitrogen, in the expansion process inside the expansion machine and the Joule-Thomson valve from 30 at. abs. to 1 at. abs., and in the compactness of the heat exchangers. While making use of these features it should be noted that the required low temperatures below 4.2°K are obtained by making the Joule-Thomson expansion into subatmospheric pressure instead of down to 1 at. abs. This subatmospheric pressure and consequent low boiling temperature of the liquid helium are maintained by a vacuum pump to be fitted at the outlet of the apparatus in the helium return pipe leading to the compressor.

Because of the low pressure of the expanded helium care should be taken that the pressure drop in the outer spaces of the heat exchangers is sufficiently small, so that large flow cross sections have to be provided. This means that the heat convection constant in the outer space also decreases and consequently larger exchanging surfaces are required than in circulatory operation at 4.2°K.

Moreover, operation in closed helium cycle entails the necessity of providing for intermediate expansion of the highpressure helium in the last exchanger prior to the Joule-Thomson expansion to ensure heat transfer.

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The end temperature attainable strongly depends on the pressure drop in the heat exchangers, the suction capacity of the vacuum pump and the circulating amount of helium.

With the aid of the apparatus, which is designed for a refrigeration capacity of 25 watts at 3.5°K, even lower temperatures can be obtained by reducing the volume of the circulating helium. Needless to say, the refrigeration capacity of the apparatus decreases. Employers of such installations mostly express the desire of being able to draw from the apparatus also liquid helium for experiments. This installation lends itself for this pur­pose as it is capable of liquefying about 7 litres of helium an hour by means of the compressor designed for recycling.

FUNDAMENTALS OF THE PROCEDURE

The thermodynamic calculations are based on the Leiden S-log T diagram of helium and on the recent values established by Zelmanoff.

The best way of explaining this procedure seems to be to illustrate it by the diagram of the heat exchanger system (Fig. 1). In the heat exchangers above the Linde exchangers

Fig. r . Diagram of heat exchanger system A, A' B, C Countercurrent heat exchanger D r, D 2

.

E Expansion machine F Collector for liquid helium G Evaporator vessel for liquid nitrogen V r Joule-Thomson expansion valve for liquefaction V 2 Intermediate expansion of pressure helium V 3 Joule-Thomson expansion valve for refrigerator operation V 4 Valve for feeding liquid helium to refrigerator feed line V 5 Valve in pipe for withdrawing liquid helium from F V 6 By-pass valve for cooling-down period II (down to r4°K) V 7 By-pass valve for cooling-down period I (down to 8o°K)

8 Refrigerator feed pipe, 2.5 -4.5°K 9 Refrigerator return pipe r o Siphon pipe for withdrawing liquid helium

D1 and D2 - the reason for this division into 2 exchangers can be gathered from the fol­lowing - part of the helium compressed to 30 at. abs. is precooled, with the aid of liquid nitrogen and of the expansion machine from room temperature to a temperature level

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of 14.5° K in two stages. With this precooling temperature and a pressure of 30 at. abs . a maximum thermodynamic efficiency is obtained which corresponds to the enthalpie difference i2 - i1 of the entering and leaving gases at the warm end of D1•

In the following it appears to be expedient to deal with the refrigerator operation and the liquefaction separately.

A) REFRIGERATOR OPERATION

The helium gas purified and compressed to 30 at. abs. enters the apparatus at room temperature.

In exchanger A heat is extracted from the high-pressure helium by the returning expanded helium gas of equal volume in countercurrent; it cools already down close to the boiling temperature of the liquid nitrogen.

The final precooling of the pressure gas current to 80° K takes place in a cooling coil inside evaporator vessel G. The nitrogen evaporated in G is used for the precooling of pressure helium in exchanger A'. The nitrogen, almost heated up to room temperature, escapes into the open.

In stationary refrigerator operation liquid nitrogen is merely required to make up for losses caused by insulation and heat exchange above a temperature of 80° K. The amount of liquid nitrogen required is about 1-2 l/h including the losses incurred while filling the evaporator.

In exchanger B the temperature of the high-pressure helium drops from 80° K to about 28°K.

Between B and C part of the high-pressure helium (in refrigerator operation about 35 % of the incoming gas) is branched off to the expansion machine. This partial current does external work during expansion in the machine and keeps on cooling down to about 14° K.

The gas expanded in the machine precools in heat exchanger C the other part (about 65%) of the high-pressure helium to 14.5° K. This precooling is assisted by the low­pressure helium leaving the warm end of D1•

The energy balance of the expansion machine and of all parts below 80° K shows that only part of the refrigeration capacity produced by the expansion machine is available as effective thermodynamic efficiency of the plant. About double the amount of the re­frigeration capacity is used to cover losses caused by heat exchange and insulation and to compensate for the enthalpic difference between entering and leaving gas, a difference which with the temperature prevailing at the warm end of B (80° K) is still positive.

The transformation of the negative enthalpic difference between entering and leaving gas at the warm end of D1 into lower temperatures occurs in heat exchangers D1 and D2 with the aid of expansion val.ves V 2 and Vi and V 3 respectively.

Since in refrigerator operation the gas current to be cooled and that to be heated up are of equal quantity and since the specific heat of the high-pressure helium rises closely to saturation, intermediate expansion in V 2 must be carried out between exchangers D1 and D2 for the heat transfer. For this it appears convenient to select a reference point of the pressure helium at which expansion is possible isothermically and isenthalpically from an outlet pressure of 30 at. abs. to a suitable lower pressure level.

These interrelations are best illustrated by an Enthalpy-Temperature diagram or a Q-T diagram of helium. (Fig. 2). Intermediate expansion from 30 at. abs. to 7 at. abs. resulted to be the optimum pressure stage.

Fig. 2 shows the isobars for the pressure gas (30 at. abs., line 2-4) and the gas ex­panded to 0.4 at. abs. (line 3-1) in exchanger D1 as well as the isobars for the gas reduced to 7 at. abs. through intermediate expansion. In exchanger D2 the cooling down from condition 4 to 6 and the heating up of the returning gas expanded to 0.4 at. abs. from condition 5 to 3 are effected. The location of the isobars at 30 at. abs. and 7 at. abs. respectively results from the energy balance of the exchangers and from the requirement of isothermic, isenthalpic intermediate expansion, shown in Q--T diagram point 4.

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Fig. 2

Fig. 2. Isobars for the pressure gas and isobars for the gas reduced to 7 at. abs. through inter• mediate expansion.

In refrigerator operation the expanded medium (Joule-Thomson expansion by valve V 3) is fed into the refrigerator feed pipe. Depending on the desired temperature be­tween 4.5°K and 2.5 ° K this expansion has to occur either against atmospheric pressure or into subatmospheric pressure. As already mentioned, the required subpressure has to be maintained by connecting a vacuum pump.

B) LIQUEFIER OPERATION

In liquefier operation the expanded helium gas quantity returning from below is smaller than the inflowing pressure gas quantity by the portion liquefied.

In the heat exchangers only a smaller amount of expanded helium is available for the precooling of the high-pressure helium than is the case with refrigerator operation. From this it results that during the first precooling stage an increased heat transformation takes place in nitrogen evaporator G. The consumption of liquid nitrogen increases to about 7 l/h in stationary operation including filling losses.

During the second precooling stage the thermodynamic efficiency of the expansion machine must be increased by boosting the throughput. For this purpose about 60% of the incoming high-pressure helium are branched off to the machine from the gas distribu­tion between B and C. During liquefaction the heat transfer in exchanger D is war­ranted without intermediate expansion in V2, if exchangers D1 and D2 are connected in series, with valve V2 completely opened.

In this case the Joule-Thomson expansion occurs via valve V1 into collector vessel F and liquid helium is obtained at about 4.3 °K. From this collector vessel liquid helium can be withdrawn as the requirements may be :

1. Via valve V4 to feed liquid helium additionally to the refrigerator feed pipe, e. g. in order to obtain a higher cryogenic efficiency of short duration.

2. Via valve V5 to fill transport vessels with liquid helium. The apparatus under discussion is primarily designed for the refrigerator operation

discussed under A). This leads necessarily to the conclusion that the heat exchangers are overdimensioned for liquefier operation.

DESIGN PRINCIPLES

As already stated earlier, the design of the apparatus reflects in principle the con­struction features of the 3-litre MeiBner helium liquefier.

Fig. 3 shows a schematic cross section of the apparatus. Note the arrangement of the heat exchangers and of the expansion engine as well as the pipe system with the valves from V 1 to V 8. The whole assembly is suspended in the vacuum-insulated vessel 9 which serves simultaneously as collector for liquid helium. For the heat exchangers

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VS u

15 16

18

- /1

12 V6

Fig. 3. Cross section of apparatus (3-l MeiBner helim liquefier)

A, A', C, D l, D 2 V I, V 3 V 2 v 4, v 5 V 6, V 7 v s

9 IO I I 1 2 13 14 15 16 I7 18

Countercurrent heat exchanger

Joule-Thomson expansion valve Intermediate expansion Valve in the liquid helium pipe Valves for cooling down the apparatus Three-way valve in refrigerator return and off-gas pipe from helium collector tank Vacuum jacket Radiation protection Expansion engine Activated charcoal Expansion chamber Refrigerator feed pipe Refrigerator return pipe Vacuum jacket siphon for liquid helium Collector for liquid helium Evaporator vessel for liquid nitrogen

copper pipes were used and wound around the supporting tube in staggered arrange­

ment. The supporting tube itself and the jacket pipe of the heat exchangers consist of chrome-nickel steel to reduce heat convection in longitudinal direction.

The expansion machine is the same as that used with the MeiBner helium liquefier. The machine tube was provided with an outer jacket up to the level of the liquid nitrogen temperature range to prevent heat being transferred from the enshrouding exchangers.

When the machine is recooled, a vacuum builds up in the sealed space owing to con­densation of the gases to insulate the machine tube.

Special attention was given to the reduction of the heat quantities fed from outside and therefore chrome-nickel steel of poor thermal conductivity was used wherever necessary. For this reason the pipes 14, 15 and 16 conducting cold helium consist of this material. Apart from this the pipes are vacuum-insulated against radiant heat and gas convection within the vessel.

Likewise the long valve stems and suction sleeves of chromenickel steel act as high thermal conduction resistances on the cold valves against heat input from components at room temperature. The valves are externally sealed by metal bellows without the use of stuffing boxes.

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Fig. 5

Fig. 4

Fig. 4. Replaceable inset of refrigerator Fig. 5. General view of housing with control fittings

B--Fig. 6. Circuit diagram of the plant

1 Helium compressor

2 Motor for compressor

3 High-pressure oil separator

4 High-pressure dry cylinder

5 Oil mist fine filter

6 Low-pressure oil separator

7 High-pressure purifiers

8 Helium refrigerator

9 Gearbox

1 o Electric motor and generator

II Vacuum pump

128

12 Electric motor for pump

13 Gas holder change-over valve

14 Helium gas holder

15 High-pressure helium gas cylinder battery

16 Transport can for liquid helium

1 7 Refrigerator feed pipe

18 Refrigerator return pipe

19 Centralised supply for liquid nitrogen

20 Nitrogen offgas pipe for discharge into

atmosphere

21 Helium return gas pipe from trial apparatus

I-21

The operating results are compiled in the following table 1.

Table 1 . Operating results

Cooling down period from starting the operating of the warm apparatus

until commencement of liquefaction : 3.5 h nitrogen consumption : 55 I

Through- Pressure Electric Consumption Liquid Thermo- Liquefying put without with nitrogen dynamic output

of helium vacuum vacuum in efficiency at pump pump stationary a temperature

operation of Nm3/h at. abs. kW kW l/h OK watts l/h

4.5 25 Refrigera-tor So 30 rS 23 I -2 3.5 26 operation

2.5 abt. 6*

Liquefier operation So 30 rS � 1 7

* with vacuum pump designed for an end temperature of 3.5 °K.

The described plant represents a prototype which operated already satisfactorily down to temperatures corresponding to subatmospheric pressures of Helium (see Figs. 4, 5, 6).

DISCUSSION

N. Kurti, U. K.: Why is it that a higher refrigeration value is obtained for an end tem­perature of 3.5° K than for 4.5°K?

A. Spies, Germany: At 4.5° K the plant i s operating at 1 .3 atm. abs. discharge pressure, i. e. without vacuum pump.

With the help of the vacuum pump in the expansion machine as well as in the Joule Thomson valve V3 expansion to lower pressures is possible corresponding to the suction capacity of the vacuum pump.

With increasing pressure difference we get a higher cold production from the expan­sion machine and that is why the refrigeration capacity in the consumer also increases.

In the heat exchanger D1 heat transfer quantity increases because with decreasing pressure in the low pressure region the enthalpic difference at the warm end increases.

J. W. L. Kohler, Netherlands: In the introduction is stated : Recent measurements have shown that for the condensation of larger quantities of hydrogen with vacua of about 10- 6 Torr, the normal boiling temperature of liquid helium is not sufficient . . . Question : Is 4.2°K sufficient for small quantities, or is the vapour pressure of hydrogen at 4.2°K higher than 10 - 6 Torr ?

A. Spies, Germany: Measurements of Borovik, Grishin and Grishina, Mark and Hen­derson have shown a vapour-pressure of 10 - 6 Torr at a temperature of 4.4°K. Tempe­ratures of 4.2°K to 4.4°K have resulted to be not sufficiently low to satisfy the require­ment of a vacuum of 10 - 6 Torr by condensation of larger quantities of hydrogen.

With decreasing cooling temperature the capacity of the condenser is increasing. Naturally by lower cooling temperature and lower gas flow quantity a higher vacuum

can be achieved.

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A Miniature Helium Turbo-Expander for Cryogenic Refrigeration Systems

Un turbo-derendeur miniature a helium pour les systl:mes cryogeniques

M. T. NORTON Cryogenic Engineering Laboratory, National Bureau of Standards, Boulder, Colorado, U.S.A.

SOMMA/RE. Le Laboratoire de Cryogenie etudie actuellement l'utilisation de petits turbo-detendeurs, appuyes par des paliers a gaz, qui donnent de petites quantites de froid a 4,2 ° et 30 ° K. On a deja traite de la turbine et des paliers a gaz pressurise a l' exterieur utilises avec le systeme a 4,2° K.

Ce rapport decrit un detendeur a turbine a helium destine a un systeme a circuit f erme. La turbine a ete confue pour extraire environ 200 watts a 30° K du courant d'helium. La turbine a 7,9 mm de diametre, tourne a 9800 t/s, et a ete COnfUe pour Une pression d'entree de 4 atm et a un rapport de pressions de 4/1. Les essais portant sur des rapports de pression de 4/1 a 6,4/1 indiquent un rendement de pointe de 65%. On indique la conception de la turbine et les donnees de fonctionnement de cette turbine.

INDRODUCTION The Cryogenic Engineering Laboratory of the National Bureau of Standards has

investigated the use of small, high speed turbines, supported by gas lubricated journal bearings, in small capacity refrigeration system. Two systems have been studied, one that operates at 4.2°K, and another at 30°K.

As a result of optimization studies, a 20 : 1 pressure ratio was used for the 4.2° K system. Although the final low temperature is obtained with an expansion valve, work

Fig. r . Turbine Test Apparatus

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extraction is necessary for system operation. A previously reported [4] single stage im­pulse turbine was used for this purpose in a test apparatus illustrated in Fig. 1 . Across this large pressure ratio, high efficiency is not normally anticipated in a single stage turbine. However, the 45 percent obtained is acceptable where the simplicity of a single stage is more important than high efficiency.

Refrigeration at 30° K is accomplished by shutting the expansion valve and leaving the low temperature loop of the system unused. Refrigeration is obtained downstream of the turbine. For this system, optimization studies indicated best performance at 4 : 1 pressure ratio. A program was initiated to investigate the performance of a 4 : 1 pressure ratio turbine for this application. A design goal of 70 percent was selected from the qua­litative reasoning that in larger sizes 80 to 90 percent was common, but in small machines friction and leakage losses tend to be relatively high.

DESIGN PHILOSOPHY

For simplicity, minimum modifications were made to the test apparatus of Fig. 1 . The same general type of radial flow nozzle and axial flow rotor was maintained. The same 7.94 mm (5/16 inch) shaft was used, except for minor modifications made to increase the critical speed to above 9800 rps. Since oil lubrication is impractical at these low tem­peratures, externally pressurized gas bearings [1], [2], [3], [4] were used.

The design conditions were 5 grams/second of helium gas entering at 30° K and 4 atmospheres, exhausting at one atmosphere. A 50 percent reaction design was selected. A turbine of this type has equal enthalpy change in nozzle and rotor, and so the nozzle exit velocity is equal to the rotor exit relative velocity. Vector diagrams of 50 percent reaction turbines are shown in standard texts [5]. The design tip speed was taken to be equal to the nozzle discharge velocity. Furthermore, the relative velocity of the rotor inlet becomes axial; the rotor blades become axially oriented in the region of the nozzle outlet; and a comparatively smooth transition results from the radial inflow nozzle to the axial flow rotor. For the given inlet and outlet conditions, isentropic flow would result in nozzle exit velocity, tip speed, and rotor outlet relative velocity equal to 257 meters/second. It was estimated that losses would reduce these velocities to 240 meters/second. For the 7.94 mm (5/16 inch) diameter rotor, the corres­ponding rotational speed is 9600 rps. With an assumed nozzle flow coefficient of 0. 90 the computed nozzle throat area becomes 4.68 mm2 (0.00726 in 2) which is the actual area of the nozzle used.

The rotor exit flow area required to obtain 240 meters/second would be 7.39 mm2

(0.01 145 in 2), if the flow were isentropic. Another calculation, based on estimated in­efficiencies within the nozzle and rotor, indicated that it should be 8.68 mm2 (0.01345 in'). Three preliminary rotors were made, whose exit flow areas were 7 .48 mm 2 (0.01 1 6 in 2), 8.00 mm 2 (0.0124 in 2), and 9.92 mm 2 (0.0154 in 2) . Since the 8.00 mm 2 rotor had the best efficiency, the final turbine was built to have the same area.

MECHANICAL DETAILS

The arrangement of the nozzle and rotor is shown in Fig. 2. Helium flows from right

(

Section A-A

132

I ' ' . . "' 1 r-&35mm 1 +-A

I E E � I I -- ' -

Fig. 2. Turbine nozzle and rotor

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to left along the outer wall and turns radially inward. As the gas approaches the rotor in 16 converging nozzle passages, it accelerates to design velocity. At the nozzle exit, flow is essentially tangential. With rotor tip speed equal to gas velocity, flow should enter the rotor with a minimum of disturbance. In the vicinity of the nozzle exit, the rotor blades were scraped to a point to further facilitate smooth transition. The 12 rotor blades had a constant depth of 1 .27 mm (0.050 in). A gradually curving, convergent passage accele­rated flow to the outlet.

The flow angle at the rotor exit is difficult to determine from the rotor geometry, be­cause of the continuous curvature of the blades. If the flow velocity does not change be­tween the rotor throat and the annular area just downstream the exit flow, angle, fl, may be found from [6],

Sin fl = A1/A2

where A1 is the rotor exit flow area and A is the annular area downstream of the rotor. For the final rotor used, angle fl becomes 17 - 1 12 degrees. At the design flow rate of 5 grams/second and a tip speed of 240 meters/second, this results in an axial leaving veloc­ity of 72 meters/second or a loss of 13 watts. Since the design output was to be 230 watts this leaving loss was considered small. As may be seen in Fig. I there is a diffuser in the flow passage downstream from the rotor. Because its efficiency was unknown, and because it was convenient to remove its central core, no benefit from the diffuser was anticipated.

During the tests an attempt was made to measure diffuser efficiency. Because of gas swirl at the turbine exit, centrifugal force caused the wall static pressure measurement to be higher than the average flow pressure. As a result, the diffuser inlet pressure read­ing was too high, and the computed diffuser efficiency was too low. At high rotor speeds this effect is minimized and the computed diffuser efficiency is more nearly correct. At one test point the computed diffuser efficiency was 80 percent even with the core re­moved. As a result, the indications are that a good part of the leaving loss was recovered.

TESTS

First tests were on a turbine type slightly different from the one described. In these first turbines the rotor blades were, in effect, cut away in the region of the nozzle exit. Thus a rotating annular plenum chamber existed, which collected the nozzle discharge and fed it into the rotor blades. This arrangement did not perform very well, but the test results were used for determination of optimum flow area in the final turbine rotor.

Poor performance of the turbines with the annular plenum is attributed to the for­mation of a free vortex in the plenum. With flow from the nozzles nearly circular and almost isentropic, the preservation of angular momentum requires that a free vortex form. Thus the tangential gas velocity is inversely proportional to the radius. For the turbine of this report the tip radius is 3.97 mm (0.156 in) and the hub radius is 2.69 mm (0.106 in). The hub to tip gas velocity ratio, which is the tip to hub radius ratio becomes 1 .472. If the nozzle exit velocity is 240 meters/second at the tip, the tangential gas veloc­ity entering the rotor at the hub is 353 meters/second. At the same time, when the blade tip speed is 240 meters/second, the blade speed at the hub is only 163 meters/second. Because of these velocity mismatches the efficiency is reduced.

The final turbine of this report was tested over a range of pressure ratios from 4 : I to 6.4 : I. Figure 3 shows the efficiency obtained plotted as a function of the turbine parameter, U/C o, where U is the turbine tip speed, and C o is the isentropic gas veloc­ity. C o is computed from the equation :

C o2 = 2 LJH isentropic

where LJH isentropic is the isentropic enthalpy drop from the inlet conditions to the outlet pressure.

Efficiencies obtained never quite reached the design goal of 70 percent, but 60 percent was exceeded at all pressure ratios. At 4.8 : I, 65 percent was exceeded at one test point. Extrapolation of the 5.7 : I and 6.4 : I data indicate that 65 percent would have been exceeded if the shaft critical speed had not limited the tests to a maximum of 9800 rps.

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At 4 : 1, the best efficiency obtained was about 60 percent. Here again the efficiency would have improved at a higher shaft speed. The limit here was power. At 4 : 1 the turbine power was too small to drive the power absorbing brake faster than 8320 rps.

"'Ro1or-No. 107 ,.050in Deep, Exit Flow Area 0.0124in 1 , 12 Blades , No Inlet Plenum -

70 1-+++++++-H-+-lf-+-+-+-++++-H

f-+-t-+-t-+-t-+-+-+- f.,0\\0 H-jc+-Jc+-J--+-i �1e��\)1e

5��.s'-'-+-+--"-'--'-'�o�.s.,.,.....'"-'-'-'-+-+-_._,_.,.o.1 U/C, Fig. 3. Performance of Low Pressure Ratio Turbine

Increasing the U /C o ratio by increasing the operating speed above 9800 rps or by reducing the inlet temperature would improve the performance somewhat, but it is expected that the amount of improvement would be small.

At all pressure ratios, over the speed range tested and at 30° K inlet temperature, the flow rate was almost independent of rotational speed. All inlet temperatures, except one, were 30 ± 0.1 degrees Kelvin. The exception was at 25.4 degrees Kelvin and 4.8 : 1 pressure ratio. The low temperature was a means of increasing the U /C 0 ratio to obtain the one data point above 65 percent. The following table presents typical data.

Pressure Ratio 4.05 4.06 4.83 4.75 4.87 5.70 5.71 6.44 6.39

Turbine Speed rps 7220 8820 7800 9600 9450 8450 9600 8850 9400

Inlet Pressure Ats. 3.82 3.83 4.70 4.62 4.70 5.74 5.71 6.61 6.61

Outlet Pressure Ats. .944 .944 .972 .972 .966 1.006 1.000 1.027 1.034

Inlet Temp. ° K 29.9 30.0 29.9 30.l 25.4 30.l 30.0 30.1 30.l

Outlet Temp. ° K 22.8 21.9 21.7 20.9 17.5 20.8 20.2 20.0 19.8

Flow Rate g/sec 4.45 4.43 5.50 5.46 5.89 6.86 6.72 7.41 7.40

Efficiency .542 .621 .581 .644 .652 .609 .644 .622 .642

U/Co .495 .603 .506 .622 .668 .527 .600 .538 .571

The efficiencies reported include expander shaft and other heat leak. Considering the turbine as an expander in a refrigeration system, this procedure is correct. To evaluate efficiency as a turbine a heat transfer correction is helpful. For the computed heat leak of 13 watts, the improvement is not large. At 4 : 1 the increase is 1.6 percent; at 6.4 : 1, it is 0. 7 percent.

CONCLUDING REMARKS

Improvements are to be expected in turbine performance. Areas of improvement would include, improved inlet nozzle efficiency, reduced transfer losses between nozzle and rotor, redesign to utilize the good diffuser efficiency. Future work should include the evaluation of externally pressurized gas lubricated bearings to operate at about the same pressure as the turbine. Until recently the bearings had utilized 20 atmospheres gas pressure. The externally pressurized bearings have been proven to be stable and appear capable of indefinite operation. Tilting pad or other self acting bearings that do not need external pressurization should also be investigated.

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The research program described here indicates that acceptable efficiency can be ob­tained in turbine gas expanders for use in refrigeration systems as small as 200 watts capacity. Furthermore, in our experience, the operation of these turbines in properly designed and fabricated gas bearings will result in good reliability.

REFERENCES

r. B. W. Birmingham, H. Sixsmith and W. A. Wilson, The Application of Gas· Lubricated Bear­ings to a Miniature Helium Expansion Turbine, Advances in Cryogenic Engineering 7, K. D. Timmerhaus, Editor, 30-42 (Plenum Press, New York, N. Y., 1962).

2. H. Sixsmith, W. A. Wilson and B. W. Birmingham, Load-Carrying Capacity of Gas-Lubricated Bearings with Inherent Orifice Compensation Using Nitrogen and Helium Gas, NBS Technical Note u5, PB 16 16 16 (August, 1961).

3 . H. Sixsmith, The Theory and Design of a Gas-Lubricated Bearing of High Stability, Procee· dings of First International Symposium on Gas-Lubricated Bearings, ACR-49, Office of Naval Research, Washington, D. C. (October, 1959).

4. D. B. Mann, H. Sixsmith, W. A. Wilson and B. W. Birmingham, A Refrigeration System Incorporating a Low Capacity High Speed Gas Bearing Supported Expansion Turbine, Ad­vances in Cryogenic Engineering 8, K. D. Timmerhaus, Editor, in Press.

5. ]. F. Lee, Theory and Design of Steam and Ga� Turbines, page 243, McGraw-Hill Book Co., Inc.

6 . H. A. Sorens�n, Gas Turbines, The Ronald Press Co., New York.

DISCUSSION

J. B. Gardner, U. K.: In the preprint of Mr. Norton's paper it is mentioned that future work should include the evaluation of externally pressurised gas lubricated bear­ings operating at the same pressure as the turbine inlet pressure, i. e. 4 to 6 ats. Does this mean that in Mr. Norton's work the bearings operated with a higher gas pressure than that at the turbine inlet, and if so what was the pressure ?

M. T. Norton, U. S. A.: At the time that the paper was written, the bearing inlet pressure used had been 20 atm. abs. Since that time we have made a bearing redesign und have operated the same and similar turbines in bearings whose inlet pressure was 6 atm. abs.

J. B. Gardner, U. K.: What was the bearing gas consumption as a percentage of the turbine throughput ?

M. T. Norton, U. S. A.: The gas bearings consume a quantity of helium gas between 5 % and 10% of the turbine throughput.

V. Chlumsky, Czechoslovakia: What is the superficial finish of such a rotor ? I suppose that the superficial finish has a great influence on the efficiency of that small machine.

M. T. Norton, U. S. A.: One of the rotors tested was hand polished under a microscope that magnified 10 times. It had poor efficiency. The final turbine of the report, and the later turbine mentioned in the talk, appeared rough under the microscope.

The efficiencies of the rough turbines were better than that of the smooth one. There were other differences between the turbines which made the final ones good, but it seems that surface finish is not of primary importance.

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Utilization of a Combined Expansion Cycle in Liquid Air Separating Installations

Le cycle a detente combine dans les installations de separation d'air par lique­faction

S. S. BUDNEVICH, I. K. KONDRYAKOV, L. A. AKULOV and G. A. GOLOVKO Leningrad Technological Institute of the Refrigerating Industry, Leningrad, USSR

SOMMAIRE. Le cycle a detente combine de liquefaction representant la combinaison du cycle a detente a haute pression et du cycle regenerateur a detente a basse pression, se carac­terise par l' efficacite energetique et ii est raisonnable de l' employer pour les grandesinstallations de separation d'air par liquefaction. Dans le schema de principe de /'installation pour la production de l'oxygene liquide avec le cycle a detente combine de liquefaction, le cycle a detente a haute pression est realise a l' aide de l' ecoulement de l' azote en recirculation. On donne dans le rapport la relation entre la consommation specifique de l' energie pour la pro­duction de l'oxygene liquide dans les conditions de regime optimal et des dif.ferentes pressions de l'azote en circulation (la consommation minimale de calcul de 1'02 liquide au regime optimal est de 0,97 kwh/kg) .

On donne ensuite le schema de !'installation pour la production de l'oxygene liquide en utilisant !es turbomachines pour la compression et I' expansion des gaz. Parmi les avantages de ce schema il est a noter !'absence de l'huile dans les gaz liquefies. Enfin, on donne les schemas de principe de /'installation a grandes dimensions pour la production de l'azote liquide avec le cycle a detente combine de liquefaction ainsi que pour la production simul­tanee de l'oxygene et de l'azote liquides provenant de l'air.

The efficiency of a combined liquefying expansion cycle, which is a combination of a high pressure expansion cycle with a low pressure regenerative expansion cycle, is due to the fact that the ratio of compressor work to the work of the expansion engine is considerable in both cycles owing to which the worsening of the data of the actual cycle with reference to the theoretical one is not sharp [1, 2].

The high efficiency of the combined liquefying expansion cycle makes it expedient to apply it in large liquid air separating installations.

Fig. r. Diagram of installation for the production of liquid oxygen with a combined expansion cycle. r - centrifugal air compressor, 2 - regenerators. 3 - expansion heat exchanger, 4 - expansion turbine, 5 - nitrogen preheater, 6 - rectifying column, 7 - reflux subcooler, 8 -nitrogen com­pressor, 9 - preliminary heat exchanger, ro -nitrogen expansion engine, rr - main heat exchanger.

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Fig. 1 presents the diagram of an installation for the production of liquid oxygen with a combined liquefying expansion cycle. A detailed description is given in [3]. The high pressure expansion cycle is effected here by means of a circulating flow of nitrogen. The low pressure regenerative expansion cycle effects the flow of the air to be separated. The peculiarities of the given circuit include also the utilization of regenerators for coo­ling the air and purifying the latter of C02 and H20. The presence of circulating nitro­gen makes it possible to perform complete air separation in a single column apparatus. The peculiarities of the given circuit include also the existence of optimum energy conditions, which are characterized by a quite definite pressure of compressed nitrogen. The presence of optimum conditions is explained as follows. The share of liquid oxygen, separated from the air, is quite definite and stipulated by the quantitative content of 02 in the air and conditions of rectification. As a result the waste nitrogen flow has a given value owing to which a quite definite flow of nitrogen may be supplied from the circulating nitrogen flow for cooling the air to be separated. The high pressure expansion cycle can develop a refrigerating effect which is required for separating liquid oxygen at different pressures. Each pressure is characterized by a certain optimum temperature before the expansion engine with a minimum consumption of power.

The flow of circulating nitrogen is quite definite under these conditions. However, as mentioned earlier, the flow of nitrogen, directed for cooling the air, is given, therefore the value of the return flow of nitrogen in the high pressure expansion cycle becomes quite definite. In this case the value of the return flow of nitrogen is imposed also in the optimum possible duty at the given pressure. Imposed, therefore, is also the part of nitrogen which is supplied to the expansion engine. The optimum operation of the high pressure expansion cycle coincides with the optimum possible duty (bound with the mentioned value of the return flow of nitrogen) at a quite definite pressure of the nitrogen in the circulation cycle.

The analysis points out that this pressure in the installations producing liquid oxygen is approximately 160 atm. abs. Fig. 2 presents the dependence of the specific consump­tion of power for the production of liquid 02 under optimum possible conditions at different pressures of circulating nitrogen.

t01 � .. 1,00 c:.. '$. /199 � /198

O,'J'l ........

........... ......_

J ,,/

,,,,,,. / -�

120 130 140 150 160 170 180 190 200 Patm. aos.

Fig. 2. Dependence of specific power consumption for the production of liquid 02 on the final pressure of nitrogen under optimum conditions.

The minimum rated consumption of power under optimum conditions is approxi­mately 0.97 kwh/kg of liquid 02•

Fig. 3 illustrates the diagram of the installation for the production of liquid oxygen when using only centrifugal machines for compressing and expanding the flows. The pressure of the circulating nitrogen in this case reaches 30 atm. abs. The advantage of this circuit lies in the possibility of producing liquid oxygen without any admixture of oil. Similar to the previous circuit, complete separation of the air is effected in a single column apparatus with concentrating and skimming sections. The given circuit is inferior to the previous one from the energy point of view because the value of the nitrogen pressure differs from the optimum value.

The specific rated consumption of power for the liquefaction of oxygen is 1 .2 kwh/kg of liquid 02•

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Fig. 3. Diagram of installation for the production of liquid oxygen. x - air compressor, 2 -regenerators, 3 -expansion heat exchanger, 4 - air expansion turbine, 5 - rectifying column, 6 - reflux subcooler, 7 - first nitrogen expansion turbine, 8 -second nitrogen expansion turbine, 9 -nitrogen heat exchanger, 10-nitrogen centrifugal compressor.

The diagram of a large installation for the production of liquid nitrogen with a com­bined liquefying expansion cycle is given in Fig. 4. The amount of liquid product, produced by the installation, is not limited in this case by its content in the air, owing to which the optimum pressure of circulating nitrogen will be the maximum possible value. The analysis of the given circuit points out that the output of liquefied nitrogen is 0.28 n. m3/n. m3 of air or 0.35 kg/n. m• of air. The specific rated consumption of power is about 0.82 kwh/kg of liquid N 2•

2

Fig. 4. Diagram of installation for the production of liquid nitrogen with the co:nbined expansion cycle. r - expansion turbine, 2 -regenerators, 3 -separating column, 4 -subcooler, 5 - nitrogen com­pressor, P=3 ata; P= 200 ata, 6 - 7 -heat exchanger, 8 - reciprocating expansion engine.

Fig. 5 presents the diagram for the simultaneous separation of liquid oxygen and liquid nitrogen from air with the combined liquefying expansion cycle. The application of this type of installation is expedient when it can replace two liquefying installations in case of simultaneous need in liquid oxygen and nitrogen.

This type of installation provides for simultaneous complete separation of oxygen in the liquid form from the air, as well as different portions of liquid nitrogen.

This paper deals only with the principal problems of developing liquid air separating installations, based on the utilization of a combined liquefying expansion cycle.

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Fig. 5. Diagram of installation for the simultaneous production of liquid oxygen and nitrogen with a combined expansion cycle. l - air centrifugal compressor, 2 - regenerators, 3 -expansion heat exchanger, 4 -expansion tur­bine, 5 - one cycle rectification column, 6 - subcooler, 7 - heat exchanger, 8 -nitrogen com­pressor, 9 -nitrogen expansion engine, IO- collector.

It is quite obvious that the solution of problems of nonclogging of the regenerators with carbon dioxide is highly essential. The problem of providing self-purging of the regenerators in installations, operating according to the latter circuit, in which the waste nitrogen flow is considerably smaller than the direct one, will encounter many difficulties.

However, these problems are of independent character and will not be discussed herein.

An analysis of the circuit of liquid air separating installations with a combined lique­fying expansion cycle points out that this type of installation is highly effective if consi­dering power consumption. This type of installation inherently combines the utilization of an effective liquefying cycle with a simplest separatory unit, which provides in a one column apparatus for complete air separation. The latter is explained by the pres­ence of a circulating flow of nitrogen which makes it possible to add a concentrating section to the separatory unit. Concentrated liquid nitrogen is supplied as reflux to the upper part of the concentrating section.

Liquid oxygen without any admixture of oil can be produced when applying only centrifugal machines for the compression and expansion of the flows but with some increase in power consumption.

REFERENCES

r. S. S. Budnevich, I. K. Kondryakov, Combined expansion cycle for liquefying air. Works of LTIRI, 1958, vol. XV.

2. S. S. Budnevich, I. K. Kondryakov, Improvement of refrigerating cycle for air liquefaction. Scientific Papers of Higher School, Energetika, 1958, No. 2.

3 . S. S. Budnevich, I. K. Kondryakov, Circuits of large capacity liquid oxygen installation. Works of LTIRI, 1956, vol. XIV.

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1-15

Purification Method for Obtaining Very Pure Hydrogen at High Pressure

Methode de purification permettant d'obtenir de l'hydrogene tres pur a pression elevee

A. van ITTERBEEK, K. ST AES, 0. VERBEKE and F. THEEUWES

Instituut voor Lage Temperaturen en Technische Fysika, Leuven, Belgique.

SOMMAIRE. Pour obtenir des mesures tres precises de l'hydrogene a pression elevee, il a ete necessaire de le purifier et de le stocker a des pressions atteignant au mains 100 atm.

Nous partons d' hydrogene a 1 % d'impuretes. Get hydrogene est comprime dans un recipient d'acier inoxydable a 150 atm et a 20° K. La methode de purification s'appuie sur le principe que toutes les impuretes presentes (0,2 N2 • • • ), sauf !'helium, se solidifient a 20° K. Un filtre metallique separe ces impuretes de l'hydrogene liquide. Un courant continu d'hydrogene a haute pression s' ecou[e a travers [' apparei[ et i[ est fina/ement Stocke dans Un cy/indre a haute pression.

Le degre de purete est determine par la mesure de la vitesse du son dans le gaz et par comparaison de cette valeur avec la valeur theorique.

INTRODUCTION :

For the purpose of carrying out comparative experiments on the thermodynamical properties of ortho- and para-hydrogen, and their mixtures, it was necessary to purify the hydrogen gas used.

The hydrogen gas that we normally use in the liquifaction process is prepared for industrial purposes by dissociation of water on steel. The degree of purity given by the manufacturer is of the order of 99,5 % to 99o/o.

Since we normally use the thermal compression method to obtain pressures up to 3000 atm., it was necessary to compress the gas, after purification, again to pressures of 120 to 150 atm.

We must also be able to control the degree of purity and tried to solve this problt>m by measuring the velocity of sound.

I . PURIFICATION OF NORMAL HYDROGEN

a) The purification method is based on the principle that 02, N2, CO, A and hydro­carbons, cooled down to 20° K, are solid for partial pressures greater than 10 -5 mm Hg. It then follows that if the mixture of H2 and the impurities at 100 atm is cooled down to 20°K, the relative pressure of the impurities would be less than 10 - 10• By using a filter to remove all the solid particles, it will be possible to obtain very pure hydrogen.

b) Apparatus (Fig. 1) and description of the purification process.

Before starting the purification process, the apparatus is evacuated by a rotation pump (P). The impure hydrogen, stored in a steel vessel of 25 or 40 litres at 150 atm (H21), streams through a spiral (S1) cooled down by liquid nitrogen. In this way, im­purities like H20, 02 and C02 can be removed. The main reason however, is to cool the gas to about 80° K.

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Fig. 1. Apparatus used for the purification of Hydrogen.

The gas flows through thermally insulated tubes to a second spiral (S2) cooled down to 20°K by liquid hydrogen. In the second spiral, impurities like N2, A and some hydrocarbons, formed by the preparation, can be solidified.

A filter (F) takes care that the impurities remain in the spiral. The liquid hydrogen will evaporate and flow in a thoroughly cleaned high pressure vessel (H211).

During the process, the velocity of sound is measured to control continuously the purity in the resonator (R).

With this apparatus we are able to purify 1200 litres atm. of hydrogen in one hour, and therefore we used about 6 litres of liquid hydrogen and 10 litres of liquid nitrogen.

c) Some remarks about the filters used.

In order to economise on the consumption of liquid hydrogen we introduced a pre­cooling arrangement with liquid nitrogen. Since at lower stream velocities this effect is annulled, we use an adequate amount of filter for at higher stream velocities there exists a danger that the impurities are transported with the liquid hydrogen.

A Fig. 2. Two types of filters used.

B

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The first filter we used (Fig. 2 A) consists of a set of small capillaries (10-2 cm diametre) arranged in a high pressure bomb.

This type was given up because we needed too much hydrogen to cool it down. The second filter (Fig. 2 B) consists of a tube filled with compressed glass-wool. With

this one better results are obtained.

2. DETERMINATION OF THE DEGREE OF PURITY

Different ways are possible for the determination of the degree of purity of a gas. The best of them is the analysis by spectroscopic way.

We could not apply this way, since we wanted to control the purity during the pro­cess.

It was also necessary to control frequently the variation of the ortho-para mixture during the experiments, mentioned in the introduction.

The measurement of the thermal conductivity and the velocity of sound are the two possible ways that we could use for determining the degree of purity. The first method requires an extremely good temperature stability and therefore we selected the second method.

a) The velocity of sound in a gas is given by

and since we measured the velocity at about 300°K and 1 atm. we can also write :

W = vy� For a gas mixture the velocity of sound is given by

WM = x W1 + (1 - x) WH2

where WM, W1, WH2 and x are respectively: the measured velocity of sound for the mixture; the velocity of sound of the impurity; the velocity of sound in pure hydrogen and the concentration of the impurity.

If a greater accuracy is wanted we should also consider the compressibility coefficient for the calculation of W1 and WH2• Since the most important part of the impurity is air we can easily take W1 as the velocity of sound in air.

For C p, C v and the compressibility coefficient Z (p) many data are given in literature [l , 2, 3, 4] for hydrogen and air. They are results of theoretical and spectroscopic research.

b) The apparatus for measuring the velocity of sound consists of a resonator filled with the gas, where in by means of a generator standing waves are produced. At re­sonance, when the effective length of the resonator is n times, the half wave length; there is a maximum energy transport.

The outcoming signal is amplified and recorded. By measuring different resonant frequencies, we can put :

W = 2 L (vn - Vn-1) where L is the effective length of the resonator and Vn and Vn-1 are two succeeding resonant frequencies.

The effective length of the resonator is calibrated by measuring the velocity of sound in pure helium at the same temperature.

The resonator is thermal stabilized at 0.02°K and a relative frequency stability of 10-5 is obtained.

With this condition we can detect impurities to 0,1 %. In Fig. 3 a diagram of this apparatus is given.

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6

Fig. 3. Diagram of the apparatus for measuring the velocity of sound.

3. SOME RESULTS OF PURIFICATION OBTAINED WITH THIS METHOD

AND APPARATUS.

In the table the results of different processes are shown. In the different columns the type of filter; the quantity of purified hydrogen; the

quantity of consumed liquid hydrogen; the time of the process and the purity are noted down.

Type Quantity of Consumed Time Purity

of purif. H2 liquid H2

filter (lit. atm) (lit.) (hours) %

A 1500 9 2 99,975 A 1200 8 2 99,981 A 1200 6,5 1,5 99,978

B 1500 8,5 1,5 99,997 B 1300 7,8 1,5 99,995 B 1400 8,1 1,5 99,993 B 1200 6 1 99,996

- -- - - -·- ---·- -

From the above table we can conclude that only an adequate amount of filter gives the possibility of a fast process without risk of decreasing the degree of purity.

We should remark that it is absolutely impossible to remove He if it is present.

REFERENCES

r. H. W. Woolley, R. B. Scott, F. G. Brickwedde, "Compilation of thermal properties of Hydrogen and its various Isotopic Ortho-Para Modifications", Journ. Res. of N.B.S. (nov. 1948).

2. A. Farkas, "Ortho-hydrogen, Para-hydrogen and Heavy Hydrogen", Cambridge University Press (1955).

3. H. L. Johnston and D. White, "Viral Coefficients", Phys. Rev. 79.236 (1950).

4. ]. Hilsenrath et al, "Tables of Thermal Properties of Gases", N.B.S. circ. 546 (1955).

144

Appl icat i o n s to N u clear P hysics

Appl i cati o n a l a p hysiq u e n u c l e a i r e

Cryogenic Technology i n the Nuclear Rocket Program*

1-18

Technologie cryogenique dans un programme de mise au point de fusees nucle­aires

E. F. HAMMEL Los Alamos Scientific Laboratory, University of California, Los Alamos, New Mexico, U. S. A.

SOMMA/RE. On passe en revue la technologie cryogenique dans le programme de mise au point des fusees nucleaires aux Etats-Unis. On souligne d'abord le materiel cryogenique pour l' essai des moteurs de fusees nucleaires au Centre d' Essai du Nevada, dependant de la Commission Americaine pour l'Energie Atomique. On decrit specialement le stockage, le transport et les systemes de pomage et de reglage de l'ecoulement de l'hydrogene liquide utilise comme agent propulseur. On presente un rapport sur le comportement general de ces appareils dans les essais reels. Outre /es renseignements obtenus dans les essais a grande echelle reels, d'autres prob/emes de cryogenie comprenant le comportement de l'hydrogene liquide dans l' ajutage du moteur de fusee, le refiecteur de neutrons et le coeur du reacteur au cours de la mise en marche, du fonctionnement et du ralentissement pour !'interruption du reacteur ont ete etudies au Centre d' Essais et !ors d' experiences simulees a plus petite echelle au Laboratoire Scientifique de Los Alamos. On etudie ces experiences ainsi que leur relation avec le fonctionnement d'un moteur de fusee nucleaire.

Since several different methods of utilizing nuclear energy for space vehicle propulsion are currently under investigation, it is important at the outset to state that the parti­cular nuclear rocket program which will be discussed in this paper is the so-called Project Rover. This undertaking is being conducted jointly by two United States govern­ment agencies, the Atomic Energy Commission and the National Aeronautics and Space Administration. The objective of the program is to develop a satisfactory upper stage rocket powered by a nuclear reactor. The Los Alamos Scientific Laboratory has been assigned responsibility for the development of the nuclear reactor. When the develop­ment of the reactor is completed it will be incorporated into a rocket engine system by the Aerojet General Corporation and the Westinghouse Electric Corporation. Following that, Lockheed Missiles & Space Co. will incorporate the engine system into a suitable upper stage vehicle.

The test reactors under development at the Los Alamos Scientific Laboratory have been called Kiwi reactors, not only because of their relationship (operational) to the New Zealand bird which does not fly, but also because, for reasons of convenience : the reactors to date have all been tested upside down, and hence cannot fly. In principle a Kiwi reactor and the operation thereof is quite simple. The device consists of a high power density nuclear heat exchanger reactor in which energy generated by fission is imparted to a suitable propellant passing through the reactor core. The heated propellant then exhausts through a de Laval type nozzle by means of which its thermal energy is converted into kinetic energy, and a thrust is imparted to the rocket system in the usual manner. A schematic diagram of such a system is shown in Fig. 1 .

*) Work performed under the auspices of the U.S . Atomic Energy Commission.

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Since the specific impulse obtainable from a nuclear rocket system is inversely pro­portional to the square root of the molecular weight of the exhaust gas, hydrogen is by far the most efficient of the various propellants available. For roughly comparable con­ditions of chamber temperature and pressure the specific impulse of a nuclear rocket using hydrogen as the propellant can be shown to be approximately twice that obtainable from one using any of the high energy chemical fuels presently available. As a result

Fig. I . Nuclear Rocket Propulsion System

of this increase in specific impulse it can also be shown that, for a given mission (for example, earth escape), it will be possible to increase the payload by a factor of at least four when nuclear propulsion systems become available.

146

11411 "' 'UI [D c

Fig. 2. Schematic Diagram of a Kiwi Reactor

1-18

In Fig. 2 a slightly more detailed schematic diagram of one of the reactors currently under development is shown. The reactor core consists of an uranium-graphite matrix. The hydrogen propellant enters the regeneratively cooled nozzle at about 20° K and exhausts from the reactor core at a temperature of the order of a few thousand degrees Kelvin. To date the development problems have resulted primarily from the stringent operational requirements of designing a reactor in which the following reactor para­meters a) core temperature, b) rapid start-up capability, c) ability to withstand extreme temperature gradients, and d) minimum weight are all optimized.

In order to make significant progress in the development of a satisfactory high power density nuclear reactor required for this application, full scale tests have been required to check advances in reactor design. These te�ts�have been carried out at the Nuclear Rocket Development Station in Nevada. The overall facility contains several so-called test cells, some of them operational and others still under construction. Test Cell A, in which all of the tests to date have been performed, is shown in Fig. 3. In Fig. 4

l· ig. 3. Test Cell A from rear, showing liquid hydrogen dewars, piping, and maintenance ; assembly and disassembly building in the background

Fig. 4. Kiwi B 4-A at the face of Test Cell A being readied for testing

a reactor is shown being installed in its test position at the face of the test cell. During an actual test, shown in Fig. 5, about a thousand data channels transmit to the control station information relating to core and propellant temperatures and pressures, pro­pellant flow, reactor power, vibration, radiation, etc. This information is subsequently analyzed to provide new design criteria. In addition to full scale reactor tests at Nevada, fundamental studies relating to the reactor operation as well as extensive component

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Fig. 5. Kiwi n r-B Reactor Test at NRDS September r, 1962, using liquid hydrogen as the propel· !ant.

testing programs are being carried out simultaneously at the Los Alamos Scientific Laboratory in New Mexico and at National Aeronautics and Space Administration laboratories in order to provide additional technical support for the Rover program.

Prior to September 1, 1962, there had been much concern about possible disastrous consequences if, during the start-up of the reactor, liquid hydrogen should happen to enter the reactor core in appreciable amounts. If this should occur, the resultant increase in reactivity caused by the excellent neutron moderating property of high density hydrogen could be expected to cause an uncontrollable reactor power excursion. This in turn might either melt the reactor or otherwise damage it. It was therefore decided to attempt to program the start-up in such a way as to increase the reactor power simul­taneously with the flow rate in such a way as to prevent two-phase hydrogen from entering the reactor core. This was to be accomplished by adjusting the heat input into the propellant as it progressed sequentially through the various portions of the reactor shown in Fig. 2 (i.e., the nozzle and the neutron reflector) such that the hydrogen entering the reactor core was always maintained well above its critical temperature (33 ° K) during the period that the core inlet pressure was less than the critical pressure of hydrogen (12.9 atm). Given the necessary transport prcperties and heat transfer data for fluid hydrogen a complicated but straightforward set of hydrodynamic and heat transfer equations were developed to produce the necessary start-up program for the reactor. There were many uncertainties, however, and of these the most troublesome was the possibility of liquid entrainment by the gas stream and the possibility that thermal equilibrium would not be maintained as the propellant passed through the various passages in the nozzle and reactor.

In order to obtain some information on the validity of these assumptions prior to full scale testing of a reactor, a reflector flow passage was constructed and in­stalled in a suitable liquid hydrogen test flow system capable of simulating start-up conditions in the reactor itself. Both wall and fluid temperatures were recorded during typical reactor start-up programs and compared with the corresponding theoretical in­formation produced by the calculation. These results are being published [1] but since this information is not yet generally available, it is appropriate to note at this time that, in general, good agreement between the calculation and the experiment was obtained only after the fluid had progressed a considerable distance along the flow passage. While the initial transient flow conditions obtained, whenever and wherever the state of the fluid corresponded to either two phase or near critical, very poor agreement between the experimental and the calculated behavior was observed. In view of the inadequate experimental or theoretical transport property data available for this region of the state diagram, it was not possible at that time to determine how much of the disagreement was attributable to the use of erroneous values for the transport properties in the cal-

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culation and how much was attributable to the failure to obtain thermodynamic equi­librium in the flow passages. Experiments are currently underway to investigate in more detail the pressure drop and heat transfer phenomena occurring under transient condi­tions, such as will be experienced in nuclear reactor start-ups.

In order to obtain information on the degree to which thermal equilibrium is achieved in the propellant as it progresses along its flow path, a density gauge has also been developed in which the sensing element is a capacitor located in the flow channel. Since the densities of liquid and cold gaseous hydrogen differ considerably, any variation of the density of the fluid flowing between the plates of the capacitor will correspond, according to the Clausius-Mosotti relation, to a change in the dielectric constant and hence the capacitance of the condenser. If this capacitance is made part of a resonant LC circuit, density changes can then be easily detected by measuring the resulting frequency shift. After preliminary tests several of these gauges were installed in the flow passage immediately upstream of the reactor core inlet in a recent cold-flow*) run in Nevada in order to determine the existent to which the attainment of thermal equi­librium produced single phase uniform density hydrogen by the time the propellant reached the reactor core inlet.

Some typical results are shown in Fig. 6. It should be noted that marked fluctuations in the density of the fluid entering the reactor core began when the mean temperature of this fluid was as high as 200° K. Thermodynamic equilibrium obviously was not achieved as the propellant was forced through the warm nozzle and reflector passages despite expectations to the contrary. These expectations were based upon the fact that the design of the flow passages was such that a well-developed turbulence should have been produced in the fluid stream over most of its path.

" " ..

PROPELLANT PRESSURE (p1l14) 40 AT REACTOR CORE INLET

PROPELLANT TEMPERATURE (•JO AT REACTOR CORE INLET

'" , ..

... ..

or--������������---<

PROPELLANT MASS FLOW RATE (ARBITRARY SCALE)

'·'

pl'!i AT REACTOR CORE INLET 0 (I

__ .• L_____________l_. _ _ _ .l_ 2.0 30 40

TIME (SECONDS)

Fig. 6. The relative density, flow rate, temperature and pressure of the propellant vs time just before entering the reactor core during a cold-flow reactor test.

For unavoidable reasons these data were not available at the time the first hot reactor was ready to be tested using liquid hydrogen. Since other, but far less specific, data did indicate that such a test was feasible, two reactors were actually tested before the information shown in Fig. 6 became available. Although other problems related to the mechanical design of the reactor were uncovered in these hot runs, it was quite con­clusively proved on September 1, 1962, that it is possible to start up a reactor on liquid hydrogen. It thus appears that any reactivity fluctuations due to changes in propellant density are either self-compensating or at least easily controllable. Alternatively, it may

*) Several cold-flow runs have been performed using unloaded, i . e. no U230 in the core, reactors to investigate transient start-up conditions.

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be that in a hot run thermodynamic equilibrium is more nearly achieved and that the density of the fluid entering the core is indeed quite uniform. In future reactor tests these questions will be exhaustively investigated. Simultaneously, fundamental studies relating to two-phase flow and heat transfer into two-phase fluids under nonequilibrium conditions are being undertaken with considerable interest.

Of the many cryogenic engineering problems which have arisen in the Rover program thus far, the one described above is but an example. Many other difficult ones relating to the design, construction, and operation of the facility have been met and resolved -at least sufficiently to carry out reactor tests. As the development progresses more will be encountered and hopefully solved as successfully.

ACKNOWLEDGEMENT

The work reported in this article is obviously that of many individuals both of the Atomic Energy Commission, and its contractors. Particular acknowledgement should be made however to Mr. Walter Willis and Dr. Kenneth Williamson of the Los Alamos Scientific Laboratory and to Mr. Jess R. Smith of The Aerojet General Corp. for their work in connection with the development of the density gauge and the cold flow measure­ments shown in Fig. 6.

REFERENCES

I. G. P. Watts, A. R. Lyle, and J. D. Balcomb, Proceedings, Nuclear Propulsion Conference, Monterey, California (1962).

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Les problemes techniques souleves par les irradiations neutroniques au dessous de 30°K

Technical Problems raised by Neutronic Irradiations below 30° K

L. BOCHIROL, J. DOULAT, A. LACAZE et L. WEIL Centre d'Etudes Nucleaires de Grenoble, Section de Physico-Chimie et Basses Temperatures, Chemin des Martyrs, Grenoble (!sere), France

SUMMARY. The main problem raised by neutronic irradiation in high fluxes at very low temperatures is to provide sufficient refrigerating capacity to dissipate the relatively large amount of energy (which may amount to 1 watt/gram) resulting from the nuclear heating of the sample itself and of the materials making up the irradiation space proper. The most advantageous solution is to plunge the samples in a suitable cryogenic liquid (neon, hydro­gen or helium) .

The problem is then to achieve compact and manageable enough irradiation cryostats so that they can, on one hand, fit in to the small space available near the core of the reactor and, on the other hand, be placed near other testing equipment located there; finally, when it is not possible to make measurements "in situ", it should be possible to recover the irradiated samples without warming them.

The solutions to these various technological problems are set out, in a description of the cryogenic plants recently installed in the "Melusine" cell at the Centre of Nuclear Research, Grenoble. Available refrigerating capacities are higher than 5 watts at 4.2° K and about one hundred and fifty watts at 28° K.

Les physiciens du solide portent un interet croissant aux irradiations a basses tem­peratures, en particulier en pile. Les dispositifs d'irradiation en pile a basses tempera­tures posent un certain nombre de problemes difficiles, provenant :

1 . de la puissance quelquefois importante qu'il faut retirer de l'enceinte d'irradiation, 2. des reactions chimiques qui peuvent etre induites par les rayonnements sur le fluide

de refroidissement, et enfin, 3. des sujerions diverses imposees a toute experience en pile : protection, encombre­

ment, reactivite, choix des materiaux, etc . . . . II y a quelques annees [1, 2] nous avons pu resoudre ces problemes pour les irra­

diations a 77°K, et les boucles a azote installees dans la pile Melusine comptent au­jourd'hui plus de 15 000 heures d'irradiation, qui ont permis un certain nombre d'etudes sur !'action des rayonnements sur les solides. Cependant, certains defauts se guerissent au-dessous de 77° K, voire meme au-dessous de 20° K, et leur etude necessite des irra­diations a des temperatures inferieures. Deux dispositifs d'irradiation en pile a helium liquide [3, 4, 5] et un a hydrogene liquide [6] ont jusqu'a present ere realises par ailleurs.

Nous avons construit pour la pile Melusine deux dispositifs d'irradiation, aussi simples et aussi stirs que possible, l'un a 28°K, l'autre a 5°K, comportant le maximum d'elements communs.

1 - IRRADIATIONS A 28°K

Les dimensions de l'enceinte d'irradiation, imposees par certains echantillons, sont : 3 cm de diametre et 15 cm de hauteur environ. Le taux d'echauffement gamma a !'­emplacement de cette enceinte est de 0,2 Watt/gramme environ. II en resulte une puissance necessaire de refroidissement approchant 100 Watts, ce qui exige un debit d'hydrogene liquide de 10 litres/heure, plus les pertes en ligne.

a) Principe de l'appareil Cette consommation importante eliminait pour nous la possibilite d'alimenter la

boucle a partir d'un reservoir de stockage a cause des difficultes d'approvisionnement

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en grosses quantites d'hydrogene liquide, et pour des raisons de securite. Nous avons done choisi de faire circuler l'hydrogene en circuit ferme entre un liquefacteur et la boucle. Ce procede a, de plus, l'avantage d'utiliser toujours la meme quantite limitee de gaz, ce qui diminue le cout de fonctionnement et elimine le risque d'explosion chimi­que. Si, en effet, on fait en sorte qu'il regne toujours dans l'ensemble du circuit une surpression par rapport a l'atmosphere, on est assure que l'hydrogene reste parfaitement pur, meme s'il existe de petites fuites. Ce procede a d'ailleurs ete deja utilise pour une boucle a refroidir les neutrons [7] et s'est revele tout a fait stir (4 annees de fonctionne­ment dans EL 3).

La Fig. 1 donne le schema du circuit. Un reservoir-ballast (1) de 4 m3 amortit les variations de pression. Deux compresseurs Corblin (2), a membrane - pour eviter la pollution du gaz - delivrent ensemble un debit de 50 m3/h (TPH) d'hydrogene a la pression de 150 Atm. Le liquefacteur T. B. T. (3) assure une production maxima de l'ordre de 20 l/h d'hydrogene liquide. II a ete installe en dehors du hall de la pile. La conduite de transfert d'hydrogene liquide (4), longue de plus de 20 metres, isolee par le vide et a ecran d'azote liquide, est du type decrit par ailleurs (8.). La circulation d'­azote liquide dans l'ecran est utilisee pour l'alimentation du liquefacteur. Un sysreme de by-pass automatiques (5) entre compresseurs et liquefacteurs assure une limitation superieure a la haute pression et une limitation inferieure, fixee un peu au-dessus de la pression atmospherique, a la basse pression.

La boucle devait etre con9ue de fa9on a permettre la recuperation sans rechauffage des echantillons irradies. Le circuit principal d'hydrogene devant essentiellement rester ferme, cela imposait que l'irradiation soit faite dans un bain separe, refrigere par l'hy­drogene du circuit principal. Le principe (Fig. 1) est identique a celui de nos boucles a azote liquide (1).

Fig. I . Schema du Circuit d'hydrogene I. Reservoir-Ballast 2. Compresseurs 3. Liquefacteur 4. Conduite d'hydrogene liquide 5. By-pass automatiques

152

® -

6. 7. Articulations permettant une pseudo­translation (6) et une rotation (7) du crvostat

8. Rese.rvoir d'azote liquide

9. Reserve d'hydrogene comprime

1-10

b) Choix du Neon

Nous avions le choix, pour le fiuide de ce second bain, entre l'hydrogene et le neon. Nous avons, pour !'instant, choisi ce demier qui offre les avantages suivants :

1. On peut ouvrir l'enceinte d'irradiation pour sortir les echantillons sans qu'il y ait de degagement d'hydrogene dans le hall de la pile.

2. Le coefficient d'echanges thermiques entre le liquide et l'echantillon est meilleur pour le neon que pour l'hydrogene.

3. La chaleur de vaporisation du neon par unite de volume est 3 fois superieure a celle de l'hydrogene, ce qui laisse plus de temps pour la sortie des echantillons dans un bain de liquide contenu dans un recipient non isole.

4. L'utilisation d'hydrogene demanderait une surface de condensation beaucoup plus grande, a moins de consentir a travailler a des pressions elevees pour obtenir, a la con­densation, le meme ecart de temperature qu'avec le neon.

En contrepartie, !'utilisation de neon presente, en plus de son prix eleve, !'incon­venient d'exiger des precautions speciales pour eviter sa solidification sur le condenseur. Le bas de boucle realise pour assurer cette condition est represente schematiquement sur la Fig. 2. La surface de condensation est un tube de cuivre (1) sur lequel est soudee, le long d'une generatrice, une boite cylindrique (2) egalement en cuivre, parcourue par l'hydrogene liquide a debit constant. La surface de condensation a une valeur telle (400 cm2) que, lorsque la puissance normale (100 Watts) est degagee dans le bain (3),

f j

2

i Fig. 2. Bas de boucle a Neon

r . Tube condenseur 2. Bolte a hydrogene liquide 3. Bain de Neon liquide 4. Ecran chauffe de regulation 5. Coeur du reacteur 6. Tube amovible de defournement 7. Tete du bas de boucle 8. Bouchon 9. Recipient a echantillons

ro. Echantillons r r. Resistance chauffante 1 2 . Vide d'isolement

l'ecart de temperature a la condensation est superieur a 7°K. Ainsi, non seulement le neon est liquide, mais encore il regne dans l'enceinte de neon une pression superieure a !'atmosphere, ce qui y evite toute entree d'air OU d'eau. Lorsque la pile est arretee OU a faible puissance, OU lorsque la boucle est hors pile, Un ecran de cuivre (4) place a l'interieur et a proximite de la surface de condensation, est chauffe par une resistance electrique, et apporte ainsi la puissance d'appoint necessaire pour realiser la meme condition. Cette puissance peut fare reglee automatiquement en fonction de la pression qui regne dans l'enceinte d'irradiation.

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c) Quelques details de realisation La partie immergee, de 7 cm de diametre, possede, comme pour les boucles ii azote

liquide (2), un decrochement horizontal (Fig. 1) qui assure la protection biologique et le degagement de la piscine au-dessus du coeur. I1 est possible de sortir la boucle en fonctionnement de la position d'irradiation, par exemple pour permettre la desactiva­tion d'echantillons irradies tout en les maintenant froids avant defournement. Pour cela, un systeme de double articulation (6 et 7, Fig. 1) de la conduite d'hydrogene liquide, utilisant soufflets metalliques et joints toriques tournants, rend possible un mouvement horizontal, ii deux degres de liberte, du bas de boucle.

Le defournement des echantillons est possible aussi, sans deconnexion de la boucle ni arret de l'alimentation en hydrogene. La boucle se trouvant en position de desacti­vation, hors flux, un tube (6, Fig. 2) est assujetti sur la tete du bas de boucle (7) qui se trouve ii 4 m environ au-dessous de la surface de l'eau, et vide par l'air comprime. Le bouchon (8) est alors devisse au moyen d'une cle speciale, et remonte rapidement ii la surface de l'eau. L'ecran (4) et le recipient (9) contenant du neon et Jes echantillons (10) remontent aussi et le recipient est rapidement plonge dans un vase Dewar contenant du neon. Naturellement, il faut laisser la boucle se rechauffer completement avant d'y placer de nouveaux echantillons.

d) Resultats Des essais avec chauffage electrique ont permis d'evaluer h puissance maxima de

refrigeration (ii 28°K) ii 150 Watts. La boucle a ete placee contre le coeur de la pile fonctionnant ii 2 Megawatts, ii un

emplacement ou les flux sont approximativement les suivants : Neutrons rapides (E > 1 Mev) : 2,1 X 101 2 n/cm2 sec. Neutrons thermiques : 1,7 x 1013 n/cm2 sec. Rayonnement y : 1 ,5 X 108 r/h.

Le debit maximum (20 l/h) du liquefacteur etant utilise, la pression du bain de Neon se maintient au voisinage de 1,2 Atm, ce qui correspond ii une temperature d'irradiation de 28°K.

2 - IRRADIATIONS A 5°K

On utilise cette fois l'helium liquide obtenu dans la cascade Joule-Thomson Azote­Hydrogene-Helium.

Pour obtenir la puissance de refrigeration maximum et annuler les pertes au transfert de l'helium, la liquefaction de !'helium est faite dans le bas de boucle lui-meme, ii proximite de l'enceinte d'irradiation. L'appareil fonctionne done non en liquefacteur mais en refrigerateur, dont le rendement est ii peu pres le double.

L'hydrogene liquide est envoye dans le bas de boucle par le circuit decrit plus haut. Le circuit d'helium est schematise sur la Fig. 3. Le reservoir-ballast est ici de 1 m3, le compresseur, ii membrane lui aussi, donne un debit de 25 m3/h TPN, ii une pression de 30 Atmospheres.

Deux serpentins S1 et S2, refroidissent l'helium haute pression, le premier par l'azote liquide, le second par l'hydrogene liquide, et deux echangeurs tubulaires ii contre­courant E1 et E2 refroidissent l'helium haute pression par !'helium en retour et econo­misent ainsi, le premier l'azote liquide, et le second l'hydrogene. L'echangeur final ii contrecourant E3, permet la liquefaction apres la vanne de detente reglable D.

Les echangeurs E1 et S1 se trouvent au-dessus de la surface de l'eau, E2 dans le «haut de boucle » forme du trorn;-on horizontal et du trorn;:on vertical superieur de la partie immergee, S2, S3 et D sont dans le «bas de boucle».

Le bas de boucle contient done les organes essentiels d'un liquefacteur d'helium. Son diametre exterieur a du etre limite pour des raisons d'encombrement en pile ii 12 cm et sa fabrication a, de ce fait, pose des problemes delicats. La Fig. 4 represente la dis­position des organes dans ce bas de boucle. On remarque que leur ensemble est protege thermiquement par un ecran ii trois parois parcouru par l'hydrogene vaporise, et que l'echantillon est accessible par le haut. Naturellement, ii ne saurait etre question ici

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Fig. 3. Schema du Circuit d 'helium

r . Reservoir-Ballast

I ' r---E==:1..1 : r - - - - - - -'

Fig. 4. Bas de Boucle a helium

r . SerpentinS2 dans la boite hydrogene 2. Echangeur final E,

3. Vanne de detente 4. Commande de la vanne de dctente

(par pression d'helium gazeux). Le capillaire de commande n'est pas represente

5. Vase reserve d'he!ium 6. Emplacement de mesure 7. Emplacement d'irradiation 8. Col d'acces a l'echantillon 9. Ecran refroidi par l'hydrogene

vaporise ro. Arrivee d'hydrogene r r . Depart d'hydrogene r 2 . Arrivee d'helium r3. Depart d'helium r4. Ecran de plomb I 5· Coeur de reacteur

2. Compresseurs E1, E2, E3 Echangeurs helium-helium

51 Serpentin dans l'azote 52 Serpentin dans l'hydrogene D Vanne de detente

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de recuperer l'echantillon froid apres irradiation, mais on peut faire in situ toutes les experiences qui ne necessitent que des mesures electriques et, eventuellement, un deplacement vertical limite, comme c'est le cas pour des mesures magnetiques par extraction. On peut aussi effectuer des recuits contrOles, hors flux, en diminuant le debit OU la pression de !'helium de fa�on a arreter la liquefaction, et en reglant a !'aide de la vanne de detente ces parametres a une valeur appropriee, pour obtenir la tempera­ture desiree.

Comme la boucle a 28° K, la boucle a 5° K peut etre deplacee horizontalement dans l'eau sans arret du refroidissement.

La puissance de refrigeration, mesuree par chauffage electrique, est de l'ordre de 5 a 6 Watts a 5 ° K. Il suffira d'ailleurs d'introduire un pompage sur le bain d'hydrogene liquide pour la porter a une dizaine de Watts.

Ces puissances, compte tenu de la possibilite d'adjonction d'ecrans de plomb reduisant le rayonnement y, suffisent pour de nombreuses experiences.

Nous voudrions, pour terminer, signaler les grands avantages que nous avons trouves a l'emploi d'une pile piscine pour les irradiations a basses temperatures. En dehors meme de l'avantage fondamental du a la grande proportion de neutrons rapides dans le flux, la pile piscine permet, par rapport a d'autres types de piles, une simplification de la conception des appareils et une remarquable souplesse dans leur emploi.

REFERENCES

r . L. Bochirol, ]. Doulat, L. Weil, A cryogenic device for irradiation in liquid nitrogen. Cryogenics r, 44 (r960).

2. L. Bochirol, ]. Doulat, L. Weil, Principle of a liquid nitrogen irradiation device and its reali­zation for use in a swimming-pool type reactor. Advances in Cryogenic Engineering, vol. 6, p . 130 (1960).

3. R. R. Coltman, T. H. Blewitt, T. S. Noggle, Techniques and Equipments used for reactor irradia­tions at low temperatures. Rev. Sci. Instrum. 28, 375 (r957).

4. R. R. Coltman, Reactor irradiation studies at 4 °K. Symposium on Radiation Damage in Solids and Reactor Materials, Venise, Mai 1962.

5 . H. Riehl, W. Schilling, H. Mei/Jner, Design and installation of a low temperature irradiation facility at the Munich Research Reactor FRM. Research reactor journal 3, 9 (1962).

6. ]. F. Watson, ]. L. Christian, ]. W. Allen, A study of the effects of nuclear radiation on high strength aerospace vehicle materials at the boiling point of hydrogen. ERR - All - 085.

7. B. jacrot, A. Lacaze, L. Weil, Cellule it hydroglme liquide dans la pile EL3 de Saclay. Comptes rendus du Xeme Congres International du Froid, Copenhague r959, Vol. r, p. 2r4.

8. A. Lacaze, L. Weil, Pipe-line pour hydrogene liquide. Comptes rendus du Xeme Congres Inter­national du Froid, Copenhague r959, Vol. r, p. 2r8.

SUMMARY OF THE DISCUSSION (Papers I-18 and I-10)

J. Wilks, U. K. : Since there is a finite probability that the initial launching of any rocket system may fail, what provisions are being made to insure against a nuclear acci­dent during the launching of a nuclear rocket ?

E. F. Hammel, U. S. A.: With respect to safety in general, it is contemplated at the present time that the nuclear rocket will be utilized as the third stage of a Saturn system. The nuclear system and its initial firing will be done from that orbit. In the event of a failure of the Saturn system during launch the nuclear rocket will be dropped more or less intact into the ocean. Tests are currently being conducted to insure that any such immersion of the reactor core will not create either a dangerous nuclear or radioactive situation.

In reply to a question about when the hydrogen exhausted from the rocket nozzle ignites, Dr. Hammel replied :

A lit propane torch is located immediately above and slightly to the side of the rocket nozzle. This serves to ignite the effluent hydrogen as soon as the propellant flow is turn­ed on and has replaced the purge gas in the lines.

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J. R. v. Geuns, Netherlands: I should be interested in the running times of your helium loop, as there are no means in it to adsorbe impurities which even in a closed circuit may come into the helium gas through small leakages and by desadsorption from the walls of the tubing in a long run.

J. Doulat, France: We did not as yet operate the loop for periods longer than 10 hours, and we observed no blocking due to impurities. Small leakages cannot result into pollu­tion since the pressure in the circuit is everywhere greater than atmosphere. We hope troubles due to desadsorption should be avoided, since the whole circuit was carefully vacuum-degassed before it was filled with helium once for all. If it is not the case we can add a cleaning device somewhere in the circuit.

Ph. Akar, France: Le liquefacteur a hydrogene liquide qui sert a alimenter le circuit decrit sur le schema I est-ii bien situe a l'exterieur du hall de la pile, comme indique sur le texte ecrit de votre communication ? Cette solution vous parait-elle difficile a realiser (feasible) ? y a-t-il d'autres exemples de telles realisations ?

J. Doulat, France: Le liquefacteur d'hydrogene est, en effet, en dehors du hall de la pile, et la conduite d'hydrogene liquide est longue de plus de 20 m. Les pertes dans cette conduite ne sont cependant pas superieures a 10% du debit (2 l/h). Cette solution n'est pas difficile a realiser. Elle est d'ailleurs aussi utilisee pour des sources de neutrons froids a Harwell et a Saclay.

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A Metallic Helium Cryostat for Double Resonance Experiments

Un cryostat metallique a helium pour des experiences de double resonance

L. van GERVEN*), P. GROBET**), A. van ITTERBEEK, Y. H. TCHAO**), and G. van DAM**)

Instituut vor Lage Temperaturen en Technische Fysika, Leuven, Belgique.

SOMMA/RE. On decrit la partie basse temperature d'un appareil destine a la polarisation nucleaire dynamique (double resonance), comprenant un vase Dewar metallique a helium liquide.

On indique les premiers resultats d'une etude, a l' aide de cet appareil, sur le spectre de RMN du DPPH a la temperature de !'helium liquide.

We designed and built a cryostat destined for magnetic resonance experiments and especially for double resonance experiments at very low temperatures. We preferred a completely metallic construction because of the requirement of disposing of as much liquid helium as possible in the smallest possible magnet gap.

1. THE DEWAR

A detailed drawing of the dewar is given in Fig. I .

p�

I /invid =r n11ro9en

.,mL 2cm

- - - - - - - - I" bd =

lJ== � G � r

f D

(

8

A

Fig. r . Cross-section of the dewar.

The lower. part A of the helium vessel is made of I st quality brass. Brass indeed seems to be the best material as to its magnetic properties at liquid helium temperatures : its shielding coefficient is low as well for DC as for AC (50 Hz) magnetic fields [I].

The radiation shield B is cooled at the top by the liquid nitrogen bath. A horizontal cross-section of the shield is represented in Fig. 2. We rejected the use of a shield made completely of copper, for such a shield would screen off the AC field for about 44%. On a brass tube with a wall thickness of 0.5 mm we electrolytically deposited a copper

*) Scientific Collaborator of the Belgian Interuniversity Institute for Nuclear Sciences. **) Research Student of the same Institute.

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Fig. 2 . Cross-section of the radiation shield.

layer 0.5 mm thick. In this layer 25 grooves of 3 mm wide were fraised, with 25 copper strips of 4.5 mm wide remaining. Now the AC shielding coefficient at 77° K is only 14%, while the heat conductivity from the top to the bottom is still sufficiently high. The temperature of the lower part of the shield is estimated to reach a steady value of about 85° K after Y:i h. Later on we shall try to determine this temperature accurately by measuring the actual shielding coefficient.

The lower part C of the outer vessel and the upper part D of the helium vessel are made of stainless steel F. D. P. 321 . Tubes E, F and G are made of brass. Tube D is turned off along about half its length to a thickness of 0.3 mm.

The helium vessel is centered by means of a moving stainless steel needle, which is fixed to the bottom of the radiation shield.

All connections (to pumps, to manometers, filling tubes etc.) can be fixed at H, so that the upper side remains free for apparatus. The dewar can be fixed by means of 5 screws to any flat plate. The dewar is closed by a neoprene ring, put in the circular groove in the top plate.

The two vacuum jackets can be evacuated separately.

2. THE LOW TEMPERATURE PART OF THE DOUBLE RESONANCE

APPARATUS

In Fig. 3 a cross-section is represented of the transmission tube, which connects both the 23 GHz oscillator (high power klystron and accessories) and the 35 MHz marginal oscillator to the VHF cavity and the RF coil down in the dewar.

The cavity, a Ha11 mode cavity, is made of brass with a wall thickness of 0.3 mm. Along the central line of one of the two largest faces a vertical slit of 1 mm wide is fraised exactly there where the VHF electric current is zero. The cavity is tunable and is coupled to the wave guide in a classical way . . .The RF coil is wound around the cavity [2].

A brass nut with a rabbet is soldered in the transmission tube. The corresponding bolt is soldered on the wave guide; the RF line is fixed to the wave guide by polystyrene discs. The whole is screwed into the tube and is fastened in this way. At the top the tube is closed by a soldered disc, through which the wave guide and the RF line pass. At the bottom the tube is closed - after having introduced the sample and tuned the cavity-by a cap, soldered onto the tube. During the measurements the tube is filled with helium gas. As the whole system (RF line, wave guide, cavity, coil) is mounted in the transmission tube, the wave guide and cavity are always at the same pressure as their

160

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Fig. 3. Cross-section of the transmission tube.

surroundings, and there is no need for closing hermetically the microwave system. Only a pressure window in the wave guide above the transmission tube is necessary. Since there is no pressure difference between the outside and the inside, the upper part of the wave guide in the transmission tube can be very thin. It is made of German silver with a wall thickness of 0.3 mm. The RF line and the transmission tube itself are also German silver.

The NMR spectrometer is finished. The central part of the spectrometer is a Gabillard­Wang frequency scanned marginal oscillator. We are awaiting now the electronic micro­wave equipment.

In the meantime we did some preliminary measurements on simple proton magnetic resonance in diphenylpicrylhydrazyl (DPPH).

The sample consists of powdered DPPH, contained in a small glass vessel. This vessel has about the inner dimensions of the cavity, and is sealed off. It is put in the cavity. The Q factor of the NMR coil remains sufficiently high, even at low temperatures.

A typical record of the NMR spectrum of DPPH is shown in Fig. 4. The struc­ture in this spectrum is due to a paramagnetic electron shift, i. e. to a different inter­action between the strongly polarized unpaired electron and the 48 different groups of protons in the DPPH molecule (possibly the large central line is a common line) [3]. We also made a record at l .4°K. From these first measurements we can conclude already now that lowering temperature below 4 ° K does not increase much the shift of the NMR lines as it should be according to Curie's law, because, at these temperatures and in a magnetic field of the order of 10' 0, the electron spin system approaches complete pola­rization (static saturation). On the other hand it has to be pointed out that the intensity of the satellite lines decreases from 4.2°K to l.4°K, which is probably due to broadening.

It is our intention to study, by means of the equipment described in this paper, the NMR spectrum of DPPH :

- in natural conditions, i. e. with the electron spin system being polarized;

- after depolarizing more or less the electron spin system by saturating the para-magnetic electron resonance transitions at 23 GHz (dynamic saturation).

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I Jl,4

I I �

Fig. 4. The proton magnetic resonance spectrum of DPPH at 4.22°K. Constant field Ho � 7,420 0 . Magnetic sweep amplitude Hr � 2.5 0.

ACKNOWLEDGMENTS

The authors are much indebted to the Belgian Interuniversity Institute for Nuclear Sciences, which supported this work, as a part of a dynamic nuclear polarization project.

REFERENCES

r. L. Van Gerven, A. Van ltterbeek, L. Stals, Bull. Inst. Inst. Froid, Ann. 1960-1, 203 (1960) .

2. Y. H. Tchao, C.R.Ac.Sc. Paris, 250, 700 (1960).

3. R. Reimann, These, Paris (1961).

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Analyse thermique en dessous de 300°K de graphites irradies aux neutrons a basse temperature

The Thermal Analysis of Neutron Irradiated Graphites below 300°K

E. BONJOUR et L. BOCHIROL Centre d'Etudes Nucleaires de Grenoble, Section de Physico-Chimie et Basses Temperatures, Chemin des Martyrs, Grenoble (!sere), France

SUMMARY. Defects caused by neutronic irradiation in the crystalline latticework of graphites are variable after a temperature of 100° K.

For the purpose of the study, on samples irradiated with neutrons at 78°K (in the pile "Melusine", at the Nuclear Research Centre at Grenoble), the evolution, between 100 and 500°K of the energy associated with the processes of the recombination of these defects, an extremely accurate thermal differential analyser was utilised (some millicaloriesf°/g).

It is briefly described, as well as some spectra of energy freeing and measurements of the evolution of electrical resistivity made at the same time, which evidence the effect or the dose on the complexity of defects appearing at low temperature.

Si, d'un point de vue purement technologique, les effets d'irradiation sur le graphite sont aujourd'hui mieux connus, les mecanismes elementaires d'endommagement sont encore mal discemes.

On sait l'interet qu'il y a, dans cette optique plus fondamentale, d'effectuer des irra­diations a basse temperature, dans des conditions telles que les processus de recuit sous irradiation etant minimises, on peut esperer ne conserver que des defauts primaires relativement simples.

Les installations cryogeniques de la pile «Melusine » [l] nous ont deja pennis d'ir­radier a 78°K des graphites de provenances diverses, et d'effectuer sur ceux-ci des mesures, soit en cours d'irradiation, soit apres, les echantillons etant bien entendu recuperes dans ce demier cas sans rechauffage.

Au moyen d'un premier dispositif d'analyse thermique [2] nous avons en particulier etudie, depuis 80°K jusque vers 500° K, le degagement de l'energie emmagasinee (energie «Wigner») qui donne d'interessantes informations sur la cinetique d'evolution des defauts crees dans le graphite au cours d'un recuit progressif. C'est ainsi que nous avons situe vers 100°K le seuil de degagement d'energie, pour des graphites irradies a 78°K, ce qui confinne une mobilite notable des defauts a cette temperature, et mis en evidence !'existence de plusieurs «pies » de degagement d'energie [3, 4]. Ceux-ci, marquant differents stades dans l'evolution de defauts primaires, correspondent de fa­�on satisfaisante a des etapes dans le restauration des caracteristiques electriques de graphites irradies aux electrons a basse temperature par d'autres auteurs [5, 6].

Nous avions pu noter, cependant, que des phenomenes de saturation se manifestent pour l'energie liberee en dessous de 200°K lorsque les doses depassent 4.1018 n/cm9 (neutrons rapides d'energie superieure a 1 MeV) et que l'importance relative des diffe­rents «pies » varie en fonction de la dose.

Nous avons pense que ces phenomenes revelaient la creation de defauts primaires relativement complexes, meme a basse temperature lorsque leur concentration devient forte, et que l'etude de la cinetique d'evolution de defauts simples necessitait !'appli­cation de doses aussi faibles que possible.

Cela nous obligeait a mettre au point des moyens de mesure encore plus sensibles, en particulier pour !'analyse thermique. Nous decrirons tres brevement l'appareillage realise pour cela, et commenterons surtout quelques resultats obtenus avec des graphites irradies a faibles doses.

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1 - EXECUTION DES MESURES D'ENERGIE EMMAGASINEE DEPUIS 80° K

Nous utilisons une analyse thermique differentielle classique, comme celle pratiquee par ailleurs pour des determinations semblables sur des graphites, a haute temperature [7]. Mais nous avons du !'adapter a nos imperatifs essentiels qui sont :

- !'execution des mesures a partir de 78° K, les echantillons n'etant en aucune mani­ere rechauffes !ors de leur mise en place dans le calorimetre,

- l'obtention d'une sensibilite, qui, d'apres nos mesures anterieures, devait etre de quelques millicalories/°K/gr, pour nous permettre l'etude d'echantillons faiblement irradies.

La Fig. 1 represente une coupe schematique du calorimetre realise.

D

I I t IV' .

G

Fig. r. Schema De Principe Du Calorimetre D'analyse Thermique.

Les echantillons (A = irradie, B = temoin non irradie) sont des disques ( 0 20 mm, epaisseur 2 mm) presentant un rapport surface-volume favorable aux echanges avec !'element chauffant. Ces disques sont associes deux par deux, des microresistances chauffantes R1 et R2 etant prises en sandwich entre eux. Celles-ci sont constituees par du fil de constahtan ( 0 0,02 mm) maintenu par collage entre deux tres minces feuilles de mica, dont la surface est celle des disques : ainsi est assuree une repartition homogene du flux thermique dans les echantillons.

On mesure de fai;:on continue l'ecart de puissance a fournir: - d'une part au temoin B, pour lui faire suivre un programme preetabli de rechauffage

(temperature enregistree par le thermocouple E>a). - d'autre part a l'echantillon irradie A, pour lequel elle est moindre, son evolution

thermique etant strictement asservie a la precedente (thermocouple differentiel LI e e).

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Deux ecrans E1 et E2, un anneau de garde G permettent, a l'aide des elements chauf­fants independants qu'ils comportent, du thermocouple differentiel Lf B E c, et de dis­positifs electroniques de regulation que nous ne decrirons pas ici, d'assurer les mesures dans des conditions strictement adiabatiques, et d'atteindre ainsi la sensibilite desiree.

Un point particulierement delicat etait d'assurer la mise en place des echantillons dans le calorimetre sans aucun rechauffage. Celle-cl se fait sous azote liquide, dans un recipient separe. On constitue les «sandwich» resistance chauffante-disques en soli­darisant ces derniers au moyen de petits clips metalliques, puis on transfere !'ensemble tres rapidement, dans le calorimetre rempli au prealable d'azote liquide. Apres vidange, celui-ci est mis sous vide secondaire, et l'on peut alors appliquer les puissances necessaires a la realisation du programme de rechauffage (2°C/minute) dans les conditions precitees. On enregistre simultanement la temperature Ba, et la puissance differentielle fournie. Cette derniere est mesuree avec une precision de l'ordre de 500 µ Watts, ce qui conduit, dans les conditions de mesure, a une determination de l'energie liberee a ± 2 millicalo­ries/0gr.

2 - RESULTATS EXPERIMENTAUX

Nous avons reporte Fig. 2, les spectres d'energie de deux types de graphites irradies a plus faible dose (6.1017 n. r.).

La courbe 1 est relative a un echantillon de graphite nature! comprime de haute densite.

8

q

Fig. 2. Spectres De Liberation D'energie.

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La courbe 2 concerne un graphite industriel de qualite nucleaire. On observe d'une part, que le premier pie d'energie (A) situe a 215°K, apparait net­

tement plus aigu dans le cas du graphite nature!. Certains auteurs [8] associent a ce pie d'energie un premier processus d'association, sous forme de paires, des interstitiels dont la mobilite deviendrait effective au dela de 160°K.

11 n'est pas impossible de penser que le profil tres etroit de cette pointe d'energie, qui emerge brutalement d'un fond continu dans le cas du graphite naturel, soit dtl au fait que la taille des cristallites dans ce dernier etant plus homogene, l'effet d'annihilation des interstitiels aux joints de grains se manifeste pour une plage tres serree d'energie d'activation.

Nous representons egalement (courbe 3) pour comparaison, le spectre d'un graphite de qualite nucleaire irradie, a plus forte dose (4,3.1018) determine anterieurement dans notre premier dispositif. On remarque ainsi, que le maximum d'energie (B) voisin de 390°K, apparait tres attenue a plus faible dose.

11 nous parait particulierement interessant de rapprocher ce phenomene de mesures de resistivite induite par irradiations et de son evolution ulterieure par recuit que nous avons effectuees parallelement. Nous avons en effet suivi !'evolution sous irradiation a 78°K, de la resistivite de ce meme type de graphite naturel, suivant une direction perpendiculaire a la majorite des plans graphitiques Fig. 3 (1).

166

25

FIG 3 VAR IATIONS R E LA T IVE S

D E R ESISTIVITE

" 18 Q 1...���....1.����...,����.l-.M.11;11&11.1&.�nI�

0 2 3

75

@ au· cours des recuits 18

dose reGue 3 ,4 . 10 n r

a'

Fig. 3. Variations Relatives De Resistivite

I-23

La variation de resistivite se sature rapidement et atteint un maximum pour une dose de 6,2.1017 n. r. Au dela, on observe une legere decroissance puis un palier. Cette satura­tion est attribuee par certains a un processus competitif entre la croissance du nombre de porteurs de charges et l'effet de diffusion sur les defauts crees dans le reseau. Mais on peut aussi bien y voir une manifestation d'autres defauts plus complexes a partir d'une certaine dose, done d'une certaine concentration en defauts primaires.

En effet, lors des recuits posterieurs a !'irradiation effectues suivant la methode de «pulse annealing » classique, on observe alors deux formes bien distinctes de restaura­tion, de la resistivite, suivant que la dose est inferieure ou superieure a 6,2.1017 n. r. (Fig. 3 (2).

Pour la dose la plus forte (3,4.1018) les deux inflexions d' «antiannealing » (A', B') se situent de fai;:on precise aux temperatures des deux maxima de liberation d'energie A, B a 215°K et 398°K que nous avons indiques plus haut comme bien marques. Par con­tre, si la dose reste inferieure au maximum de variation de resistivite precite (6,1017) n. r. on voit que !'inflexion a 398°K (B') disparait en meme temps que le pie d'energie qui lui est associe (B) s'estompe.

11 est done bien clair qu'a faible dose la disparition presque complete du pie de de­gagement d'energie B, est liee la suppression d'un stade B' d'evolution de la resistivite.

Ceci fait supposer que ce stade B' est dependant de la creation, au cours meme de !'irradiation a basse temperature, de defauts relativement complexes qui n'apparaissent que lorsque la dose, done la concentration en defauts primaires, atteint une valeur suf­fisante. Dans l'hypothese ou tous les defauts resteraient de meme type, quelle que soit la dose, on ne devrait pas observer de disparition du stade B' d'evolution de la resistivite par recuit, et Jes spectres de liberation d'energie devraient garder un profil identique, avec une simple reduction des ordonnees proportionelle aux doses rei;:ues.

Nous pensons que ces resultats, outre la contribution qu'ils apportent a l'etude des comportements des graphites sous irradiation neutronique, mettent bien en evidence !es possibilites qu'offrent les dispositifs cryogeniques pour irradiations, et les mesures, calorimetriques et electriques a basse temperature qu'on peut leur associer.

BIBLIOGRAPHIE

1. L. Bochirol, ]. Doulat, L. Weil, Principe d'un dispositif d'irradiation a azote liquide et sa reali­sation pour utilisation dans une pile piscine. Rapport CEA 1827, r96r .

2. E. Bonjour, ]. Faivre, Dispositif d'analyse thermique differentielle pour la determination des spectres d'energie Wigner de graphites irradies a basse temperature. Note CEA N° 404, 1962.

3. E. Bonjour, L. Bochirol, L. Weil, «Comptes rendus Academie des Sciences», 1962, 254, 456.

4. L. Bochirol, E. Bonjour, L. Weil, Irradiations neutroniques a 78°K de graphites polycristallins. Colloque A.LE.A. Venise, Mai 1962.

5. S. B. Auslerman, ]. E. Hove, Phys. Rev. roe, 1214.

6. W. N. Reynolds, P. R. Goggin, Phil. Mag. 5, 1049, (1960).

7. R. W. Henson, ]. H. W. Simmons, An adiabatic rise calorimeter for measuring stored energy in irradiated graphite. AERE M/R 2564.

8. P. R. Goggin, 2eme Colloque de Metallurgie CEN-S, Juillet 1962.

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Paramagnetic Resonance in r -Irradiated Donetz Coal

Resonance paramagnetique dans du charbon du Donetz irradie aux rayons y

A. van ITTERBEEK, J. WITTERS, G. FORREZ and J. TALPE Instituut voor Lage Temperaturen en Technische Fysika, Leuven, Belgique.

SOMMA/RE. Lors de recherches preliminaires sur /'influence de /'irradiation y mr la resonance paramagnetique electronique du charbon du Donetz, ii est apparu un retrecissement important de la raie.

II se produit probablement aussi une elevation du nombre des centres paramagnetiques (cf. egalement J. DUCHESNE e. a., Geochimica et Cosmochimica Acta 23 (1961) 209) .

De nouvelles recherches sont en cours, avec une plus grande precision et a diverses tempe­ratures (temperature de traitement ainsi que temperature de mesure, sur la detection a /'aide de la resonance paramagnetique electronique de /'influence de /'irradiation.

1. INTRODUCTION

In 1954 a narrow line of electron paramagnetic resonance (EPR) was found in acti­vated coal by Ingram et al. [l]. Since then several studies on ERP in coal have been made. In 1956 a very narrow EPR line (32 A/m half width at half height, or 0.4 Oe) has been found by van Gerven, van ltterbeek and de Wolf [2] in an anthracite of the Donetz basin, and studied by van Gerven in low fields down to liquid helium temperatures. The resistivity of a bulk sample of this coal was measured. For currents less than 1 mA/cm2, the resistivity was 2.18 ohmmeter at room temperature and 1890 ohmmeter at 83°K. This semiconductorlike behaviour, suggested also by measurements of Schuyverand vanKrevelen [3], urged us to investigate accurately the ERP line of this coal. A publication of Duchesne et al. [8] on Belgian coals, irradiated with y-rays, has drawn our attention on radiation effects.

2. DESIGN OF THE EXPERIMENTAL ARRANGEMENT

An X-band spectrometer has been adapted for recording narrow EPR absorption lines at temperatures from 300°K to 2°K.

In general the sample can be cooled by two different methods. One can put the sample in the narrow lower end of a dewar, and then introduce the dewar in the cavity. This method has been used by Uebersfeld [4] down to liquid nitrogen temperatures. The cavity can be of the cylindrical type, having a high Q-factor. However, at liquid helium temperatures a double dewar has to be used, as one should avoid excessive turbulence in the cooling liquid. In this way a lot of glass is introduced in the cavity. Even when quartz would be used, this would result in an important reduction of the Q-factor and hence of the sensitivity. But there is a more important difficulty with this method. It is almost impossible to avoid condensation of water on the outer wall of the dewar. As a consequence the Q-factor, to which the ERP signal is proportional, will vary in an un­controllable manner.

In order to avoid these difficulties, we apply the other method: We introduce the whole cavity with the sample in the cooling liquid. The cavity and the wave guide are hermetically closed and filled with helium gas. This avoids the cooling liquid to come inside, and no condensation of water or gases will disturb the electromagnetic properties of the cavity. In this method however the cavity can not easily be tuned. We use a rigid cavity and adapt the frequency of the klystron to the frequency of the cavity. Even more : during cooling the cavity, one can follow the change in resonance frequency, resulting from the change of the dimensions of the cavity. And in this way one can be sure the whole cavity be cooled - together with the sample - when the resonance fre� quency reaches a stationary value.

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Since the inner diameter of the dewar is only 3 cm, the cavity must be a rectangular reflex cavity and consists of a piece of wave guide, one wave length long, closed at one end and coupled to the wave guide by an iris of 7 mm diameter. As a matter of fact a rectangular cavity has a lower Q-factor than a cylindrical one. However this entails only a small loss in sensitivity, and has no influence on the accuracy and linearity of the spectrometer. During the measurement this Q-factor will remain constant.

For our preliminary measurements we used a cavity with a loaded Q at 300° K of about 2000, when empty. At lower temperatures the Q-factor increases and becomes about 4000 at liquid nitrogen temperatures. In order to keep the heat leakage below a reasonable limit, the cavity is connected to the rest of the system by a stainless steel wave guide. We also constructed a new cavity now. The samples are sealed off in a glass tube of 2 mm diameter, 1 cm long and 0.2 mm wall thickness. As a result we obtain a very low distortion of the fields in the cavity and a very good homogeneity of the static as well as the UHF field over the sample. The inner surface of the cavity has been covered with about 2 µm of copper (copper being a better conductor than silver at low temperatures) and about 10-1 µm of gold (much less than the skin depth) as a protection against corrosion.

3. PRELIMINARY RESULTS

Some preliminary measurements have been made in the previous, less carefully constructed, spectrometer.

DPPH was used for calibration. It must be emphasized that there is a great incon­venience in using this substance as a reference in high fields. The line width and even the g-factor of DPPH depends too much on the preparation of the sample. For mono­crystals an anisotropy of the g-factor and the line width has been found by Berthet [5], resulting in a strong asymmetry of the line of a powdered sample. As a consequence the zero of the first Fourier component of the absorption signal, used in differential recording, does not coincide with the top of x", the deviation depending on the ampli­tude of the modulation field. In integral recording there is a similar difficulty in the de­termination of the field where x" is maximum. For powdered DPPH at 9.1 Gc/s we observed between the low field inflection point and the top of x" a distance of 51 A/m (0.64 Oe), against 102 A/m (1 .24 Oe) between the top and the high field inflection point. The accuracy was better than 5 %, and several controls were made in order to check that dispersion or other irregularities in the apparatus are unimportant. In the future we will use charred dextrose as a reference for our measurements, following a suggestion of Hoskins and Pastor [6] .

A sample of Donetz coal was outgassed during 16 h by cooled activated coal, and sealed off. Afterwards the EPR-line was recorded (Fig. 1), before and after irradiating the sample by means of y-rays from 60Co (1.1 and 1 .3 Mev) during 1 h (about 3000 r). All the resonance lines are approximately symmetric and, except in the wings, all display a Lorentz shape. We introduced corrections for the finite amplitude of the modulation field, according to calculations of van Gerven [7], and for changes in the Q-factor of the cavity. The g-factor of non irradiated Donetz coal deviates less than 0.1 % from the g-factor of DPPH, as we could estimate from the record on oscilloscope of a double sample of DPPH and Donetz coal (chemically isolated).

We can give only tentative data for the half line width at the inflection points, Cl, and for the spin-density compared to the spin-density of DPPH, N. Before irradiation the line width Cl is 10.4 ± 0.2 A/m (0.13 Oe), and after irradiation 6.0 ± 0.2 A/m (0.075 Oe). The spin-density N, before as well as after irradiation, is 0.015 ± 0.003 times that of DPPH.

4. PROSPECTS

We are now starting a series of measurements in which we will examine, using a more accurate apparatus, the line width, the g-factor, the line shape (eventually) and the spin-density of this Donetz coal (and for reference, of charred dextrose), under different conditions.

170

Fig. 1 .

6.DA/m £.P. R. - line of Donefz coal.

19 eels ; 300°K. J

1-14

First we will investigate the coal after different periods of pumping, in order to check after which period of pumping we are reaching a stationary spin-density and line width. Afterwards we will study an irradiated sample at different moments after the irradiation, in order to check how quickly the effect of irradiation disappears. After these two controls we hope to measure the line width, and the influence of irradiation on it, in function of the number of radiation photons and of temperature, all this in well known conditions. Moreover we hope to make a quantitative study of the influence of adsorbed gases.

REFERENCES

1. D. ]. E. Ingram and J. E. Bennet, Phil. Mag. 45 (1954) 545·

2. L. van Gerven, A. van ltterbeek and E. de Wolf, "Le Journal de Phys. et le Radium" 17 (1956) 140.

3. ]. Schuyer and D. W. van Krevelen, Fuel 34 (1954) 213.

4. ]. Uebersfeld, "These de doctorat" Paris (1956) (p.33).

5. G. Berthet, "Annales de Physique" 13/3 (1958).

6. R. H. Hoskins and R. C. Pastor, J. Appl. Phys. 31 (1960) 1506.

7. L. van Gerven, "Lijnvormen in Paramagnetische Resonantie" Brussel 1963 (Ed. I.I.K.W., P. 75).

8. ]. Duchesne, ]. Depireux and J. M. van der Kaa, Geochimica et Cosmochimica Acta 23 (1961) 209-218.

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Lasting Changes in Properties of Metallic Materials Caused by Low Temperatures

Modifications permanentes des proprietes des metaux provoquees par les basses temperatures

HEINZ A. BARTH Gesellschaft fiir Linde's Eismaschinen Aktiengesellschaft, Hollriegelskreuth bei Mlinchen, Germany

SOMMA/RE. On sait que !es proprietes caracteristiques des metaux se modifient sous !'influence de la temperature. La plupart des modifications des proprietes sont cependant reversibles, c'est-a-dire qu'elles restent liees a une certaine temperature et varient avec elle.

L'A. a trouve que, sous ['influence reglee des basses temperatures, il se produisait des modifications de certaines proprietes des met aux f erreux et non ferreux et que ces modifications n'etaient pas influencees par des temperatures plus elevees. Les observations et les mesures sont illustrees par des exemples et des diagrammes.

On decrit comment !'influence de temperatures reglees en alternance au-dessous de ± 0°C provoque des modifications permanentes de la structure, du volume e t de la durete qui restent inchangees a temperature ambiante. Ce fait offre des possibilites d'applications techniques.

On montre encore que des processus de conversion impossibles a regler a des temperatures superieures a ± 0°C peuvent maintenant etre reglees par application de temperatures in­ferieures a ± 0°C.

Metallic materials are today the essential elements used in the construction of plants and equipment employed in processes involving temperatures close to absolute zero.

Applications range from simple cooling and freezing equipment to space-flight projects and include processes for the liquefaction of low-boiling gases, physical experiments at temperatures near absolute zero, and investigations into the structure of matter.

A 1 Within this scope, one group of problems - which might be called "the techno­logical realization of processes at low temperatures" - requires profound knowledge of the changes in the coefficients of materials at low temperatures, in order to ensure that the given problems are solved with certainty and the materials are capable of enduring such processes at low temperatures. This means that the equipment and installations must, for example, remain sufficiently strong and tenacious and, consequently, the coefficients such as heat conductivity, thermal expansion, etc., of the materials con­cerned must be known. In brief, the materials must behave "as expected", and the previously determined behaviour data must change reproducibly in analogy with the changing temperatures.

Fig. 1 is an example to illustrate the foregoing. It shows 3 characteristic mechanical properties, at 2 characteristic temperatures each, of five metallic materials which are of significant importance in the construction of all kinds of refrigeration plant and low­temperature equipment.

Let us take two examples from this Fig. 1 :

1. At - 195°C (-319°F), the austenitic steel X 12 CrNi 18 8 has more than 2 Y2 times the strength of the same steel at + 20°C ( + 68°F). Its ductility, on the other hand, is reduced: elongation and notch impact strength are lower at -195 ° C than at +20° c.

2. The same 3 mechanical properties, at the same temperatures as in the first example, of pure, weldable technical copper show higher values in spite of the lower temperature, i. e., this material (SF-Cu) develops both greater strength and greater ductility if ex­posed to lower temperatures.

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� 12 Ni19 EEll X12 CrNi 18 8 811 Sf-Cu Em Sa M!58 g,AIMg4

KubJcNagZd!iQMit �gmtt"'�,

u.55 Fig. I. Tensile Strength ; Elongation ; Notch

Bar Strength ; D.V.M.Test

D. V. M. = German Society for Testing Materials

12 Ni 19 X 12 CrNi 18 8 SF Cu SoMs 58 A1Mg 4

Mechanical Properties of Materials at +20 and -195°C

5 % Ni steel austenitic steel pure, weldable technical copper special brass Al-alloy, similar 5154 (USA A. A.)

What has been demonstrated in the above two examples also works in the reverse direction : the data shown in Fig. 1 return to their original values when the material is heated again e. g. from -195°C to +20°C. In other words, the characteristic values shown in Fig. 1 are dependent on the temperature. The tensile strength of the austenitic steel is again reduced, returning from the value measured at -195°C to the value measured at +20°C. The same applies to the copper, which is again less strong and less ductile at +20° C than it was at -195°C.

For practical applications, these findings lead to the following conclusion: All components of equipment to be used at both high and low temperatures must

be so designed that their mechanical properties are still within a safe range at that temperature at which such properties have the lowest values. B 1 Another group of problems worth discussing is the field of "physical experi­

ments and investiagtions into the structure of matter" referred to in the introductory paragraph of this paper.

In the course of experiments conducted for the purpose of ascertaining property changes in metallic materials under the influence of - or in relation to - temperatures below ± 0° C ( + 32° F), the author found a number of non-reversible changes : several materials, at room or higher temperatures, retained certain properties originally acquired when subjected to low temperatures.

Due to time and space restrictions, the author will confine himself to discussing merely two particularly striking and novel findings.

1. If certain alloy steels, which are capable of being hardened, - e. g. steels used for bearings that must withstand considerable mechanical stresses - are subjected to a systematic and methodical alternating low-temperature treatment in addition to the conventional heat treating, hardening and tempering procedure, it is possible to cause lasting changes with respect to structure, volume and hardness of the parts so "treated".

Fig. 2 (TT 65) shows the results of two different low-temperature treatment proce­dures* on two identical parts of hardenable steel.

*) The type of treatment, or subcooling programme, is of course dependent on the type of material and its metallurgical composition and must be individually determined for each type of alloy, very much similar to the various procedures governing the hardening process and heat treatment of materials by exposure to high temperatures. Information as to specific problems will be furnished by the author on request.

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"'

t ' "' 1 I . ioo ;-� ! t · 1s { L �J I . . � I

' _,,.

Fig. 2. Increase in Hardness (Rockwell C Scale) ; Increase in Volume (in microns) ; Temperature (in °C) ; Hardness ; Vol. ; Effect of Subcooling Temperatures in Two Different Processes (acc. to German Patent No. 879,555).

By subjecting the parts to low temperatures in accordance with a fixed, systematic programme*, the structure of the steel can, for example be changed in such a way that the volume is permanently increased and, simultaneously, the hardness of the steel parts is brought to values which cannot be obtained by means of conventional heat treatment (see lines I, Fig. 2).

In practice, these subcooling processes are important for close-tolerance precision parts of high wear-resistance at high and very low temperatures.

A few examples of such parts are : heavy-duty ball, roller or needle roller bearings; liquefied-gas pumps; refrigerator parts; cutting tools or other parts of great material strength subjected to heavy wear; etc.

2. A second result from among a considerable number of lasting property changes caused by subjecting materials to low temperatures, which have been established by the author, is shown in Fig. 3 (TT 88), illustrating by way of example how the age­hardening process of an aluminium alloy can be influenced by low-temperature treatment.

·- 100 ! f

��t-----l--:t.".'.'.'S:;::;;;�:;;;;;;;:;:.,,�::::==i � �

Fig. 3. Hardness (Vickers Scale) ; Hours ; "Cold Treatment" at Various Temperatures (prior to Precipitation Hardening) ; Aging at +160°C ; Aging without Previous Low-temperature "Cold Treatment" ; Effect of "Cold Treatment" at Various Sub-zero Temperatures (Iv) on Precipitation Hardening at +r6o°C of an Aluminium Alloy.

The diagram shows how systematic exposure to low temperatures makes it possible first to prevent any metallo-physical effect and then to influence the behaviour of the material with respect to both rate and extent of property changes.

Fig. 3 shows the treatment of an aluminium alloy aging at normal, or slightly in­creased, temperature, i. e., a metallurgical process is shown which takes place "auto­matically" in accordance with a "law".

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This process can now be influenced, as has been proved by the author's investigations, by exposing the material to low temperatures for a certain time, so that the process normally taking place at a natural rate is "frozen", as it were. A very interesting point is the hitherto unknown fact that the degree of age-hardening and the rate - observed after a certain time - at which this degree of hardness is reached can apparently be in­fluenced by the temperature applied during such "cold treatment" prior to the age­hardening process.

Fig. 3 shows by way of example that, at a "cold treatment" temperature of -40° C (-40°F), the maximum degree of hardness can be obtained within the shortest period of time (approx. 3 hrs.). As compared to "normal" precipitation-hardening (see line

----, Fig. 3), there is an increase in hardness of nearly 40% if the Al-alloy parts are "cold-treated" at -40° C.

These findings and this new process are of extreme practical importance for the aluminium-working industries, opening up a wide field of new applications for the refrigeration industry and new prospects for both manufacturers and consumers.

SUMMARY:

It is possible, by systematic low-temperature treatment, to influence the properties of materials and parts so as to increase their resistance to stresses of all kinds and their service life. In other words, low-temperature treatment of materials means better quality and greater economy.

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Physical Aspects of Bubble Formation in Hydrogen and Thermo­dynamical Properties of Liquid n-Hydrogen

Aspects physiques de la formation de bulles dans l'hydrogene et proprietes thermodynamiques de l'hydrogene-n liquide

A. van ITTERBEEK, 0. VERBEKE, de BOELPAEP, Research Student of the Belgian Interuniversity, Institute for Nuclear Physics, and F. THEEUWES, Research Student of the Belgian Interuniversity, Institute for Nuclear Physics. Instituut voor Lage Temperaturen en Technische Fysica, Leuven, Belgique.

SOMMA/RE. Cette etude est destinee a comb/er le manque de donnees necessaires pour les comparaisons des resultats theoriques et experimentaux sur la formation de bulles dans /es chambres a bulles. Mesures sur !'equation d'etat pour /es gaz liquefies.

1 . Let us first consider bubble growth from critical size. The growth from critical size has been described at first by Rayleigh [I] who neglected evaporation heat. Later and nearly at the same time were published the theories of Plesset, Zwick [2] and Forster-Zuber [3].

They used equation (d2 R) 3 · (p (R) - P a) R

d t2 + 2 R2 = e

e : density of the liquid p(R) : liquid pressure boundary R : radius of the bubble P o : pressure at infinity Pv: vapour pressure

Here we have

2 a p (R) = Pv (T) - -R

[l]

[2]

They assumed that the quantity of heat needed for this evaporation process per second is :

. 4 ;n; [ d ] Q = --3 L dt (Ra e) with e : vapor density

A : heat conductivity coefficient L: heat of evaporation

The flow of heat to the boundary of the bubble needed for this process is

· (d T) Q = 4 ;n; R2 ;. TR R

[a]

[b]

From equations ([a], [b]) they obtain the temperature T as a function of R and R: T (R.R).

Introducing T (R R) in [2] we obtain:

. 2 a p (R) = Pv [T (R, R)J - --R

[3] in [l] gives R as a function of time.

[3]

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2. Let us consider secondly the theory of bubble formation. Since the discovery of nucleation (1952) by Glaser [ 4) two theories of this phenomenon have been proposed.

a) The first proposed by Glaser and later renounced by himself has been worked out by Martelli [5] and coworkers.

b) The second theory of Seitz [6] which was called "theory of the thermal spikes". a. Glaser and Martelli impute the formation of bubbles to the lowering of the critical

radius by forces of mutual repulsion.

2 a C Pb = Pr + R + R4

N < N

2 a C Pb = Pr + R - R4

N > N

N: number of ions;

N: critical value •

This theory seemed to be of interest for different liquids but not when using Xenon as a bubble chamber liquid.

Xenon appeared to have no sensitivity for bubble formation, unless hydrocarbons were added.

Also the appearance of Cl-rays on radioactive tracks seemed to prove that bubble formation was necessary due to heat production by energy loss of secondary electrons.

Later [7] this process was shown to be impossible through the extreme low value of the recombination time for electrons.

b. This second theory was worked out by Seitz and was corrected by Bugg [7). He first estimated the energy needed for the production of a critical bubble. So he

found

E = 16 n •3 (LI + }-_ _g_ L) = 4 08 e V (H ) m

Pr2 3 Pr ' 2

He compared this value with a semi-experimental one deduced from bubble number density along a radioactive track

Et = 513 eV In the same way he calculated Et for protons and found this value too small to be

responsible for the processus. Et was too high for the reason that not all the energy produces bubbles and he calcula­

ted by means of

0,58 E2

R = --Z-e x A

the maximum energy E which an electron may have in order to �dissipate all this energy in a critical bubble (2 Re)

Z = number of electrons, per molecule

4 a withR = 2Rc ""' -

Pe

we see that Et ,...., 1 J a V Pe

A = molecular weight

(Energy which can be transformed into heat in the bubble.) We know that

aa Em ,...., 3 (energy needed).

Pe So we can estimate (with E t > Em) that a must be low and Pe high.

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Thus we can conclude that only the critical region is of particular interest. In a more extensive treatment, bubbleformationincluding heat dissipation is calculat­

ed, which implies the values of the specific heats.

3. CONCLUSION

Concluding this brief review we can resume that there are many physical quantities to be determined in this critical region :

i-: surface tension constants; P (T) : vapour tension curve; e : density; L : latent heat of vaporisation; I.; heat conduction coefficient (which seems to be of no importance in the

corrected theory of Bugg) ; Gp, Cv: Specific heats.

4. THERMODYNAMIC QUANTITIES OF LIQUID N HYDROGEN

In the critical region we have started experiments on the equation of state of hydrogen. Our apparatus is nearly the same as that used in earlier measurements and is repre­

sented in Fig. 1 [8].

® trHfl 1000 KgJc,,,i

e3--, � .------"*"------. � l $

0 I I I :

--- - - 91as

1 1 I I lJ I I I I I I I I I I I I I I I I I I I I L_J

I I o7 I

I I L _ _J

M ® ttroon UO K9/cml EB 9la$Aroan

Fig, I . Apparatus for measurements on the equation of state of hydrogen in the critical region

The liquid compressed in V is expanded in several steps in an expansion volume Ve and changes in molar volume and absolute molar volume are determined in this way.

T is measured by a Tinsley-platinum-resistance-thermometer calibrated by N. P. L. P is measured by a pressure balance, using a differential manometer constructed in

the laboratory. The isotherms are analytically represented by means of the equation:

V = A + Bp + Cp2 + Dp3 + Ep' + Fps

Each of these constants is expanded by means of the least squares method as a function of temperature

A = A(1) + A(2) T + A(3) T2 + A(4) T3 + A(5) T4 + A(6) T5

The experiments are carried out from 20 to 33° K and up to 300 atmospheres.

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In order to obtain thermodynamical quantities we proceed, using the 36 constants of the equation of state,

0 0 0

and .we calculate Gp Gv, y, H7 (velocity of sound), Ll S, µ (Joule-Thomson coefficient), using experimental values of Gp on the vapour tension curve [10] .

The calculation is carried out by means of following equations :

p ;· (02 v) Gp CP) = Gp � ;rr2 P dp

Pct

Gp +

(o v) 2 T � p

(� :) r fl

We measured the vapour tension curve of liquid normal hydrogen and compare it with existing data [11] .

The experimental points (about 50) are calculated by means of least squares in the form

B log P = A log T + T + G

It has been shown that this equation seems to fit very well for liquid argon and methane to represent the vapour tension curves.

We take the opportunity to express our truly thanks to the Belgian Interuniversitair Centre for Nuclear sciences for the financial support during these measurements.

REFERENCES

r . Rayleigh, Phil. Mag. 34, 94 (1917).

2. M. S. Plesset, S. A. Zwick, Journal Appl. Phys. vol. 25, 493. 3 . H. H. Foster, N. Zuber, Journal Appl. Phys. vol. 25, 474. 4. D. A. Glaser, Phys. Rev. 91, 712 (1953). 5. L. Bertanza, G. Martelli, A. Zacutti, Nuo Cim. 2, 487. L. Bertanza, G. Martelli, Nuo Cirn . r, 324. 6. F. Seitz, Phys. of Fluids, vol. l, 2 . 7 . D . V . Bugg, Progress in Nuclear Physics, vol. 7 , p. r (1959). 8. A. Van Itterbeek, 0. Verbeke, Physica 26, 931-938 (1960). 9. to be published.

r o. F. Simon, F. Lange, H. Gutschi, Z. Phys. 1 5 (1923) 312 . r r . D. White, A. S . Friedman and H. L. Johnston, ]. Amer. Chem. Soc. 72, 3927 (1 950).

180

Thermodyn a m i ca l Propert i es Propri etes thermodyna m iq u es

1-5

The Logarithmic Temperature Scale

Echelle logarithmique de temperature

PROF. DR.-ING. R. PLANK Technische Hochschule, Karlsruhe, Germany

SOMMAIRE. L'echelle lineaire habituelle de temperature ne montre pas clairement qu'il existe une tres grande region de temperatures au-dessous de 1 ° K. Une temperature de 0,00001° K n'exprime pas la difference physique de la matiere a cet etat par comparaison avec l'etat a 1 ° K. Si l'on tient compte de ce que, dans la region des temperatures extreme­ment basses, les valeurs essentielles ne sont plus des differences de temperature mais plutot des rapports de temperature, il semble bon de remplacer l'echelle lineaire de temperature par une echelle logarithmique. Dans cette ichelle le zero absolu de temperature se trouve a moins l'infini. Il peut etre recommande de disposer l'echelle logarithmique de telle sorte que la tem­perature de 1° sur cette echelle coincide avec 1° K. La valeur du coefficient l' expansion thermique peut encore etre choisi librement; avec un choix approprie de cette valeur une coin­cidence des temperatures sur les deux echelles peut etre atteinte aussi a 0° C. Mais il n' est pas necessaire d' abandonner l' echelle lineaire de temperature au-dessus de 1° K.

1. THE STRUCTURE OF THE LOGARITHMIC TEMPERATURE SCALE

In the temperature scale based on an ideal gas the temperature at constant pressure is proportional to the volume of the gas. This scale was internationally adopted in 1927. The increase of volume between the melting point of ice and the normal boiling point of water is divided in one hundred equal parts and each part is allotted a temperature change of one degree. The proportionality between volume and temperature is expressed by GAY-LUSSAC'S law

V = V o (1 + OG t) (1)

where v o is the volume of the gas at the melting point of ice, which is t = 0° centigrade. The coefficient of expansion of the ideal gas is OG = 1/273,15. Cooling the gas down to t = -273,15° C leads to zero volume; therefore, this temperature is the lowest possible limit or the absolute zero. Temperatures which are counted from this zero-point are denoted by T and expressed in Kelvin-degrees. Thus T (°K) = t (0C) + 273,15.

Equation (1) can be written

v/v o = 1 + OG t = OGT (2)

and we find

v-v o 1 (av) QG - ··� - - --

V o t - V o Ot p (3)

The increase in temperature of 1 ° C (or 1 ° K) at constant pressure corresponds to a constant increase OGv o in volume along the wohle temperature scale. This increase is related to the volume v o at 0° C. Such a constant increase means very little in the range of high temperatures, where a high volume is already reached; but it means very much

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at low temperatures. It is seen that a temperature difference of 1 ° in the linear scale expressed by equation (1) has not the same meaning within the different ranges of tem­perature. That is the reason why this scale was subjected to serious criticism. Other proposals have first been made by J. P. Dalton; Lord Kelvin also expressed some ideas in this field. Later on K. Schreber [1]. insisted on a logarithmic temperature scale without success; this may be explained by the fact that at this time (1898) research in physics was still far away from reaching absolute zero. But as soon as the method of adiabatic demagnetization permitted us to reach temperatures of a few thousandths of a degree Kelvin (1933), (and even "much lower" temperatures have been in prospect), I decided to draw once more attention to the logarithmic temperature scale [2]. At the opening session of the ninth International Congress of Refrigeration in Paris (1955) C. G. Gorter clearly expressed that in the range of extremely low temperatures the temperature differences should be replaced by temperature ratios [3] .

I have proposed to define one degree as the temperature increase corresponding to a volume increase of a certain fraction f3 of the actual volume v and not of the volume v o at 0°C. This would mean a shrinkage of the temperature scale in the high region and an expansion in the low region. If we designate the temperatures in such a scale by iJ, we must replace equation (3) by

1 ( av ) /3 = -v aD P (4)

By integration we find

ln (v/v o) = {3 ({} - {} o) (5)

where the values of f3 and {} o can be freely selected1• I proposed to put {} o = t o = 0° for the melting point of ice. Retaining also the value of the coefficient of expansion f3 = CG = 1/273,15, equation (5) can be written

ln (v/v o) = CG {} (6)

Introducing in this equation v/v o from equation (2) we find

ln (CGT) = CG {}

Thus, the temperatures T and {} are related by the equations

1 1 CG1J {} = - ln (CG T) or T = - e CG 'X

(7)

It becomes evident that the absolute zero T = 0 of the Kelvin scale corresponds to {} = - oo in the logarithmic scale, which certainly can never be reached. We shall designate logarithmic temperatures {} by 0 L. Table 1 shows corresponding values of T, t and {},

Table 1

T° K 00 100000 10000 1000 273,15 100,0 t°C 00 99726,85 9 726,85 726,85 0,00 -173,15 {}oL 00 1 612,5 983,5 354,5 0,00 -274,5

T° K 10,0 1,0 0,10 0,01 0,001 0 t° C -263,15 -272,15 -273,05 -273,14 -273,149 -273,15 {}oL -903,5 -1532,5 -2161,5 -2790,5 -3419,5 - 00

1 ( av) 1 I Retaining the centigrade or the Kelvin-scale we find f3 = v a T p = T which is no longer

constant. This definition of the coefficient of expansion is used in the theory of heat transport and in the physics of solids.

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2. LIMITATIONS IN THE USE OF THE LOGARITHMIC TEMPERATURE SCALE

For a wide range of temperatures there is no need to change anything in the linear temperature scale of an ideal gas. This scale is generally recognized and approved. Be­sides, it coincides with the thermodynamic temperature scale, derived by Lord Kelvin from the Carnot-cycle2• But at extremely low temperatures the practical disadvantages are so great and the valuation is so misleading that another scale should be used. I pro­pose to adopt the logarithmic scale for temperatures below 1 ° K. Below this limit there is - as expressed by F. E. Simon- an enormous temperature region [4]. With purely thermodynamic processes only temperatures slightly below 1 ° K could be reached 3• Lower temperatures could only be reached by application of other methods, e. g. mag­netic cooling. Therefore, the use of the thermodynamic temperature scale in this region is somewhat strange. But also the use of the ideal gas-scale can not be recommended, because even liquid helium at 0,1 ° K has only a pressure of about 10-32 Torr [5] ; at 0,03°K within a volume embracing our whole galactic system scarcely one helium-atom could be found. It is evident that in this temperature range new concepts of temperature should be used.

There are several possible ways to introduce the logarithmic scale for temperatures below 1°K.

a) We can make a completely new beginning by putting {} = 0° L at T = 1 ° K. Follow­ing equation (2) the volume of an ideal gas at T = 1 ° K is v = rxv o· With this value we find from equation (5) for {} = 0

and

!XV o Zn - = Znrx = -fl{} o V o

lnrx ln 273,15

f> o = - 73 = fl (8)

v o is always the volume of the gas at the melting point of ice and therefore {} o is the lo­garithmic temperature at this point. It is evident that the coefficient of expansion fl defined by equation (4) can be freely selected in the new scale. Putting fl = rx = 1/273,15 we find

{} 1 0 o = - - ln rx = 273,15 Zn 273,15 = 1532,5 L rx

and with equation (5)

1 ln (v/v o) = rx ({} + - Znrx) = rx {} + lnrx !X

Combining with equation (2)

lnrx T = ln rx+ lnT = rx {} + Zn rx

or ZnT = rx f>.

Thus, the temperatures T and {} are now related by the equations

1 r:t,{} f> = - Zn T or T = e !X

This should be compared with equation (7).

Table 2 shows corresponding values of T and {} following equation (11).

(9)

(10)

(11)

2 Unfortunately this scale can not be used practically for direct temperature measurements. 3 W. H. KEESOM reached with He4 a temperature of o,71°K at a vapour pressure of 0.0036

Torr. Nowadays o.35°K could be reached with He•.

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Table 2

T°K 100000 10000 1000 273,15 100 10 f}oL 3 145 2516 1887 1532,5 1258 629

T° K 1 0,1 0,01 0,001 0,0001 0,00001 fr0L 0 -629 -1258 -1887 -2516 -3145

We find T = f} at 2085° and at 1,0037°. b) We also can make both scales coincide at 1 ° putting f} = 1 ° L at T = 1 ° K. With

v = av a and f} = 1 we find from equation (5)

and In a = f3 (1 - f} a)

In a f} o = 1 - /J

Assuming again f3 = a = 1/273,15 we find

f} o = 1 - 2_ In a = 1533,5°L a and with equation (5) and (2)

1 In (v/v a) = lnaT = a (f} - 1 + - Ina) = af} -a + Ina a

or ln T = a (f} - 1)

Both scales are now related by the equations

1 a('&-1) f} = - ln T + l or T = e a

(12)

(13)

That gives nearly the same corresponding values of T and f} as in Table 2. For f} = 0° L we find T = 0,981°K.

c) If we select different values for f3 and a then equation (12) leads to the relating equations

1 (3({}-1) f} = ff ln T + l and T = e (14)

Now for f3 we could select such a value that both scales coincide at 1 ° K and at 0° C = 273,15° K. The condition would be, following equation (14)

therefore

and

1 273,15 = 7f In 273,15 + 1

In 273,15 1 f3 = 272,15 =0,020614 or 7f = 48,511

f} = 48,511 ln T + 1 = 11 1,701 log10T + 1 Corresponding values of f} and T are given in Table 3.

Table 3

T°K 100000 10000 1000 273,15 19'0L 559,5 447,8 336,1 273,15

T°K 0,1 0,01 0,001 f}oL -1 10,7 -222,4 -334,1

184

(15)

100 10 224,4 112,7

0,0001 - 0,00001 -445,8 -557,5

I- 5

For {} = 0° L we find T = 0,9979° K.

Schreber [1] proposed a logarithmic scale for which t and {} coincide at 0° (melting 1

point of ice) and at 100° (normal boiling point of water). In this case {} = 7f Ina T; with

{} = 100 and T = 373,15 we find fl = 0,003120 = 1/320,51 . The temperature of T = 1 °K corresponds now to {} = -1796,8°L.

REFERENCES

r. K. Schreber, Wied. Ann. 64 (1 898), p. 163.

2. R. Plank, Forschung a. d. Geb. d. Ing.-Wesens (VDI), 4 (1 933), Nr. 6, p. 262 and Annexe 1933-2 Bull. Intern. Inst. of Refrigeration, Paris.

3. C. G. Gorter, Rep. IX Internat. Congress of Refrigeration, Paris, 1955, Vol. I, p. 95.

4. Low Temperature Physics, Four Lectures, p. 19, London, Pergamon Press Ltd., 1952.

5. ibid. p. 15 and 16.

DISCUSSION

H. Hausen, Germany: One should keep in mind that by solution of the thermodynamic differential equations, as Max Planck for instance has shown, one gets the absolute temperature in the form

T = const · e- - - - -

which can also be written as follows

lnT = - - - - -

(same expression as in the power of the 1st equation). An example of this is the equation

(�;) = T (�;) - p

T v

which can be solved in the form mentioned. So the solution of many thermodynamic equations gives a logarithmic scale of temperature quite naturally.

F. G. Brickwedde, U. S. A. : Professor Plank's suggestion is interesting because it is different from other suggestions that have been made to abandon the present thermo­dynamic (Kelvin) scale altogether and replace it with a logerithmic thermodynamic scale. Professor Plank, realizing that both thermodynamic scales have advantages, has recommended that both scales be used, one above 1 ° K and the other below 1 ° K. The advantages of a logarithmic scale, which Professor Plank has discussed, will appeal to many low temperature physicists. Professor Plank's suggestion, therefore, merits the consideration of Commission I. Lord Kelvin originally proposed for the definition of the thermodynamic scale the logarithmic scale, and only afterwards proposed the Kelvin thermodynamic scale that is now in use. It is interesting to note that Kelvin proposed also the present number scale for the Kelvin scale, which was not adopted until 1954. This is the number to the triple-point ofH20. Before 1954, the number scale was defined by assigning 100 degrees to the difference between the normal boiling-point, and the melting-point in air of H20 - the centigrade definition. That after so long a time, we should turn to an original proposal of Kelvin, concerning the number scale, should cause us to give consideration to Kelvin's other proposal - the logarithmic scale.

I know that Professor Plank realizes that the forms of the equations of thermodyna­mics are dependent upon the definition of the thermodynamic scale, but it may not be obvious to all. Thus, for entropy, dS = iJQ/T, for a reversible change, would on Pro­fessor Plank's logarithmic scale become, if the present definition for entropy in statistical

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mechanics were retained, dS = .dQ/ ( exp [OG (EJ - 1)] } . A divided two-fold scale would, therefore, involve two sets of thermodynamic equations : one, for the T (Kelvin) Scale, and the other, for the EJ (Plank) logarithmic scale.

N. Kurti, U. K. : The rapidly increasing accessibility and importance of the tempera­ture range below 1 ° K makes this an opportune moment to discuss the question of the logarithmic temperature scale and we are indebted to Professor Plank for raising it. Before taking a decision we must weigh up the relative advantages of the present scale and the logarithmic scale. Of the two advantages cited by Professor Plank I would regard the second, namely the apparent unattainability of absolute zero (- oo in the loga­rithmic scale) as unimportant. As F. E. Simon used to say very emphatically, one cannot replace a fundamental law of Physics (the 3rd law of thermodynamics) by a trivial mathe­matical artifice.

Turning now to the advantages of the Kelvin scale, I want to emphasize that it is more than a "gas-scale". Temperatures in the Kelvin scale are proportional to the quan­tities of heat exchanged in a Carnot cycle, or, mathematically, they are the integrating quotients of the quantity of heat, making it into a total differential, the entropy.

This definition is particularly useful for temperatures below 1 ° K, which are reached chiefly by isentropic demagnetisation of electronic or nuclear spin systems. These proc­esses are best described with the help of particle distributions in an energy spectrum and are thus governed by Boltzmann factors exp (- s/kT) ; the proportionality between T and the energy for a given value of the Boltzmann factor is of great convenience. The Kelvin scale i s also preferable to the logarithmic scale in experiments to establish the thermodynamic temperature scale below 1 ° K. These experiments consist in the meas­urements of heat contents and entropies of well-defined states of the system and the slope of the resulting Q - S graph gives immediately the Kelvin temperature T = dB/dS. I think that giving up these advantages and having to get used to a new set of thermo­dynamic formulae for the temperature range below 1 ° K would be too high a price to pay for making it plausible to the layman that the temperature range below 1 ° K is limitless.

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The Correlation of Experimental Pressure-Density-Temperature and Specific Heat Data for Parahydrogen

Correlation entre les donnees experimentales de P-e-T et la chaleur specifique pour le parahydrogene

H. M. RODER, L. A. WEBER AND R. D. GOODWIN Cryogenic Engineering Laboratory, National Bureau of Standards, Boulder, Colorado, U.S.A.

SOMMA/RE. L'utilisation d'hydrogene dans les engins spatiaux et les appareils atomi­ques exige la per! ection de ses proprietes mecaniques et thermiques.

Les proprietes mecaniques du parahydrogene sont definies par pres de 1200 points P-e-T experimentaux tres pres les uns des autres. Les limites de temperature sont 14 et 100° K, tandis que la pression experimentale varie de 2 a 350 atmospheres. La surface P-e-T etait approchee par un grand nombre de polynomes associes a des combinaisons d'interpolation appropriees. L'utilisation de la detente virielle permet I' extrapolation a des pressions inje­rieures a 2 atm.

Les proprietes thermiques peuvent etre obtenues apres rapprochement de l' experience ci-dessus et d'une seconde experience dans laquelle on a determine la puissance calorifique a volume constant pour divers es conditions experimental es. A des temperatures inf erieures a la temperature critique, ces puissances calorifiques sont utilisees comme donnees primaires, tandis qu'a des temperatures superieures a la temperature critique, elles servent de veri­fication pour la correlation entre !es chaleurs specifiques calcutees statistiquement et les don­nees P-e-T. 11 apparait de nouveaux essais de correlation lorsqu'on calcule les fonctions thermodynamiques.

Par suite de cette correlation, on peut maintenant calculer en fonction de la pression et de la temperature des grandeurs tel/es que le volume specifique, l' enthalpie, l' entropie, la cha­leur specifique a volume constant, la chaleur specifique a pression constante, la vitesse du son /'inversion Joule-Thomson et les chaleurs de vaporisation.

INTRODUCTION

Experimental programs at this laboratory have yielded values of Pressure-Density­Temperature (P-e-T), and of heat capacities at constant volume, in the temperature range from 1 4 to 100 ° K and at pressures from 2 to 350 atm. The two types of data have been correlated to yield a self-consistent set of thermodynamic functions.

REPRESENTATION OF THE P- e-T DATA

An accurate wide-range equation of state for parahydrogen is not available. For this correlation the P-e-T surface was approximated by a large number of polynomials along lines of constant temperature and constant density, and along the two-phase boundaries.

The experimental P-e-T data provide 39 isotherms, which, as described in [1], have been represented by

P = RTe + E1A1e (i + 1l where i = 1, 2, 3 . . . 15 (1)

The maximum value ofi is 1 5 for the 33° K isotherm, and is smaller for all other isotherms ranging down to 5 for the 100 ° K isotherm. Additional isotherms at 13.8, 14, 15, and 16° K were established from the limited number of experimental points available, and from the saturation boundaries.

The smoothed isotherm polynomials permit calculation of pressures at even incre­ments of density in gas and fluid phases. The pressure-temperature pairs so obtained for

1 87

1-6

a given density, including the intersections at the appropriate lines of saturation, were fitted to

(2)

A total of 90 sets of coefficients for (2) describe the lines of constant density between 0.0005 and 0.0450 g mole/cm3•

At densities of 0.0070 g mole/cm 3 and less the isochore polynomials (2) above were replaced by the truncated virial expansion

p = RTe + RTB e ' + RTC e 3 (3)

Values of RTB and RTC were obtained from the low density data on all isotherms from 24 to 100°K as described in [1], and have been approximated as power series in temperature.

The densities of the saturated liquid and vapor have been published elsewhere [2]. However, the saturated vapor densities below the boiling point are presently calculated from (3) and the vapor pressure rather than by the equation given in [2] . Expressions for the saturated liquid density along the liquid-solid boundary and the melting pressures have been given by Goodwin [3, 4].

The vapor pressure equation by Weber, et al. [5], the heats of vaporization and critical parameters as given by [2], and the properties of the ideal gas state at 1 atm [6, 7] have been used either directly or as supplemental information for comparisons and tests.

CALCULATION OF THERMODYNAMIC FUNCTIONS

A computer program has been developed which will find a value (or values) of density corresponding to an input temperature and pressure. The interpolation scheme utilizes the isotherm equations, the isochore equations, the virial expansion, and the saturation boundaries. After the point on the P- e-T surface is defined, the program calculates such properties as entropy, enthalpy, and specific heat at constant volume.

Thermodynamic functions are calculated in regions I and II of Fig. 1 . If the P- e-T point selected falls in region I, the computations proceed from the properties of the ideal gas at 1 atm. [6, 7] as the line ofreference. If this point falls in region II, or on the liquid-

0 >-1- :i Cf) 0 z <J) W I

0 a:

0

0 Q_ <( >

. . ��:��.i��·�� .. . .. :::: : < . . �::� '. ::�.::::: ::: :. ­

.. ··!A����·--:�·��.����:::.�:-':�·�: TEMPERATURE 100

Fig. r. Regions for thermodynamic computations, density vs. temperature

solid boundary, an auxiliary table of smoothed specific heats at a particular constant density ( e1 = 0.037821 g mole/cm 3) is utilized as the line of reference. Should the selec­ted point fall on the vapor pressure curve, the appropriate densities for the saturated liquid and saturated vapor are found, and thermodynamic functions are computed for both densities.

188

1-6

The integrations called for by the standard thermodynamic equations are handled in closed form in the virial region, that is, up to a density of 0.0070 g mole/cm 3• At higher densities the appropriate derivatives of the isochores are calculated at all intermediate integral densities, and the integration is performed numerically. In the region near the critical point, at temperatures from 32 to 37°K, and at densities from 0.007 to 0.0210 g mole/cm3 the isochore derivatives are found by direct differences from a special table in pressures rather than from the isochore equations.

For region II the values of the thermodynamic functions at T = 46°K and lh serve as the starting point. Contributions are first calculated along the path of constant density to the temperature in question, and then along an isothermal path to the desired density.

The specific heat at constant pressure is computed from temperature, density, the spe­cific heat at constant volume, and the appropriate isotherm and isochore derivatives, while the velocity of sound can be derived from the two types of specific heats and the isotherm derivative.

COMPARISON AND TESTS

A comparison of experimental vs. calculated C v is one of the more stringent tests that may be applied. Error estimates in C v allow direct computation of errors in entropy, and will indicate the quality of the calculated thermodynamic functions. Values for C v were computed at the 121 experimental points published by Younglove [8] for temper­atures greater than Tc. Deviations have been plotted in Fig. 2 as a function of tern-

, 4 % '" c�

� 0 .150 -

MEAN DENSITY

mole lee

0 0109 � 8 .; ;, 0133 " . 0162 .:- :gk�� • . 0255 � g�� 3 100 -

.J I � .050

,., 0378 o .0379 .. . 0394 0.0418

i •;. Error in Cv talc.

-050 LL-D=:t::1:I:::=t==:t==:t==d 15 20 50 60

TEMPERATURE, °K

Fig. 2. Difference between experimental and calculated heat capacities at constant volumes vs. temperature

perature for 12 densities. Mean deviation for all 121 experimental points is 0.017 cal/g mole ° K or about twice the error estimated by the experimenters. The 20 points nearest the critical temperature contribute heavily to the over-all mean deviation, ranging up to 4 % error in Cv. Errors of this magnitude illustrate the difficulty of obtaining accu­rate second derivatives from the P-e-T data near the critical point. For the remaining 101 points, all of which lie at temperatures above 36.1°K, the maximum errors corres­pond to 1 % in Cv, while now the mean deviation of 0.010 cal/g mole °K is equivalent to the estimated error in the experimental values of C v. The statistically calculated con­tribution (Cv0) is approximately 90% of the total, the contribution from the P-e-T integral about 10%. The P-e-T contribution involving second derivatives of the smoo­thed P- e-T data can be computed for the most part within 7 %, but the error in this con­tribution becomes as large as 15% near the critical point.

At temperatures below Tc in region II we have used a smoothed table of C v as the line of reference. The values of C v computed at experimental conditions merely show the internal consistency among the various calorimetric runs, as the P-e-T contribution is small. The differences, experimental minus calculated C v, are plotted in Fig. 2. Maximum deviation in this range is twice that expected experimentally, while the mean deviation for the 42 points in this range is less than 0.004 cal/g mole °K.

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1-6

The heats of vaporization have been derived from the saturated liquid and vapor den­sities and the derivative of the vapor pressure using the Clapeyron equation [2]. Values so calculated are listed in column 2 of Table I. The heats of vaporization obtained in the present method are listed in column 3 of Table I as enthalpy differences between sat-

Table I. The heats of vaporization, and entropy differences along the saturated liquid line,

Heats of Vaporization, cal/a mole Entropie s, cal/g mole K (S

T. p

."' Z. 394)+

T C ) Temp. , Clapeyron t> H T · l> S Maximum Saturated s. ( ;:it

dT Difference •K eqn. Loop Loop Difference Liquid T . P.

1 3 . 803 Z l 6. 8 Z 1 6. 44 Z l 6. 47 • 3 z. 390 z . 394 -0. 004 14. Z I 7. I Z l 6. 68 Z 1 6. 73 . 4 z. 435 2. 438 -o. 003 15 Z I B. 3 ZI 7. 68 2 1 7. 7 1 . 4 z . 66Z z. 659 o . 003 1 6 Z I B . 5 Z I B . 24 2 1 8 . 24 • 3 z. 8 8 7 z. 8 8 1 o . 006 17 2 1 8 . 4 2 1 8 . 30 218. 31 . I 3. 1 1 0 3 . 104 0 . 006 1 8 Z I 7 . 9 217. 8 1 2 1 7 . 8 4 . I 3. 333 3 . 327 0 . 006 1 9 2 1 6. 8 Z 1 6. 73 Z 1 6. 77 0 3. 557 3. 552 o . 005 zo Z l 5. Z Zl4. 98 2 1 5 : 0 4 . z 3. 783 3 . 779 o. 004 zo. Z68 214. 8 Zl4. 55 Z14. 60 • 2 3 . 843 3. 840 o. 003 ZI 212. 5 Z l 2 . 68 ZIZ. 77 • 3 4. 0 1 0 4. 0 0 8 o. 002 zz Z09. 5 Z09. 43 Z09. 57 0 4. Z40 4. Z39 0 . 001 Z3 zos. 6 205. 3 4 Z 0 5 . 48 . I 4. 474 4. 474 0 , 000 Z4 zoo. 8 zoo. 3Z zoo. 47 • 3 4. 7 1 3 4. 7 1 3 o . 000 ZS 195. 0 194. Z7 194. 42 • 6 4. 957 4. 957 0 . 000 Z 6 1 8 7. 8 187. 0 1 1 8 7. 1 5 • 7 5. Z09 5 . 209 0. 000 Z7 179. 2 178. 33 1 78. 44 . 8 s. 471 5 . 47 1 o . 000 ZS 1 68 . 7 1 67 , 93 1 68 . 00 • 7 5. 746 5. 746 o. 000 Z9 1 55. 8 1 55 . 3 2 1 55. 4 1 . 4 6. 039 6. 039 0 . 000 30 140. I 1 3 9 . 7 1 1 39. 77 • 3 6. 360 6. 358 o. 002 31 1 1 9. 8 1 1 9 . 54 1 19 . 60 • 2 6. 727 6. 7 1 9 o . 0 0 8 3 2 9 0 . 8 90. 7 6 90. 7 8 0 7. 190 7 . 162 o . OZ8

urated vapor and liquid, and in column 4 as entropy differences multiplied by the ap­propriate temperatures. A comparison of columns 3 and 4 indicates that the present method is thermodynamically consistent, while a comparison of columns 2 and 3 shows that the heats of vaporization derived by the two different methods agree within 0.5 per­cent or less.

A comparison of the entropies of the saturated liquid calculated by the present method

T

with the integral JC;: t dT shows the internal consistency of two distinct calo-

Triple point rimetric experiments (references 8 and 9). The comparison suggests a value of 2.394 cal/ g mole ° K as entropy at the triple point. Values and differences are given in Table I as function of temperature. It is seen that the agreement is indeed excellent, with deviations ranging up to 0.006 cal/g mole ° K except near the critical point. The error near the criti­cal point was anticipated by Younglove [9], who represented Cs at by

C s at = (T 0A_:._TT) n + A2 + AaT + A4T2 + AsT3 + A6T4 + A7T5 ( 4)

Younglove used n = 0.1 to give the best average representation of the experimental data, whereas n = 0.6 is required to yield the proper value of the integrated function near the critical point.

In a preliminary tabulation of thermodynamic functions the P-e-T surface was de­scribed by an equation of state [10]. Assuming the equation to be thermodynamically consistent the present method was compared to earlier results. No gross inconsistencies were detected as all deviations corresponded closely to the errors anticipated in the preliminary computation.

Thermodynamic properties consistent with this correlation will be submitted to the J. Research, NBS.

190

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REFERENCES

1. R. D. Goodwin, D. E. Diller, H. M. Roder and L. A. Weber, Pressure-density-temperature relations of fluid parahydrogen from 15° to 100°K at pressures to 350 atin, submitted to J. Research NBS.

2. H. M. Roder, D. E. Diller, L. A. Weber and R. D. Goodwin, The orthobaric densities of para­hydrogen, derived heats of vaporization, and critical constants, Cryogenics (Mar. 1963).

3. R. D. Goodwin, Melting pressure equation for the hydrogens, Cryogenics 2, No. 6, 353-355 (Dec. 1962).

4. R. D. Goodwin and H. M. Roder, Pressure-density-temperature relations of freezing liquid para-hydrogen to 350 atinospheres, Cryogenics, (Mar. 1963).

5. L. A. Weber, D. E. Diller, H. M. Roder and R. D. Goodwin, The vapour pressure of 20°K equilibrium hydrogen, Cryogenics 2, No. 4, 236-238 (June 1962).

6. H. W. Woolley, R. B. Scott and F. G. Brickwedde, Compilation of thermal properties of hydro­gen in its various isotopic and ortho-para modifications, J. Research NBS 41, No. 5, 379-475 (Nov., 1948).

7. H. W. Woolley, private communication.

8. B. A. Younglove and D. E. Diller, The specific heat at constant volume of para-hydrogen at temperatures from 15° to 90°K and pressures to 340 atinospheres, Cryogenics 2, No. 6, 348-352 (Dec. 1962).

9. B. A. Younglove and D. E. Diller, The specific heat of saturated liquid para-hydrogen from 15 to 32°K, Cryogenics 2, No. 5, 283-287 (Sept. 1962).

10. H. M. Roder and R. D. Goodwin, Provisional thermodynamic functions for para-hydrogen NBS Tech. Note 130 (PB l6I 631), December 1961.

19!

1-19

The Velocity of Ultrasonic Pulses in Hydrogen Between 60 and 90°K as a Function of Pressure

Vitesse des pulsations ultrasoniques dans l'hydrogene entre 60 et 90° K

A. VAN ITTERBEEK, W. VAN DAEL and A. COPS Instituut voor Lage Temperaturen en Technische Fysika, Leuven, Belgique.

SOMMAIRE. Nous avons publie recemment des valeurs experimentales sur la vitesse du son dans de l'hydrogene normal et para, liquide, a la courbe de pression de vapeur saturee. La vitesse des ondes ultrasoniques dans l'hydrogene liquide depend de la concentration ortho­para dans le liquide. Entre 14 et 20° K, la vitesse dans un melange para a 99,8% est plus faible d'environ 8 m/s que dans le melange para a 25% le long de la courbe de pression de vapeur. Nous avons mesure alors cette difference a des pressions plus elevees.

En utilisant laformule de Tait, nous avons calcule le volume molaire de l'hydrogene normal. En combinant les durees avec les valeurs experimentales de la vitesse du son, nous pouvons calculer d'autres proprietes thermodynamiques, par exemple les coefficients de compressibilite

/3s et /3T et le derive ( ;: ) T·

SYNOPSIS Measurements are reported on the velocity of sound in gaseous hydrogen at 5 tem­

peratures between 63 and 9 1 ° K for pressures up to 250 kg/cm-2• These results are compared with the thermodynamical data tabulated by Woolley Scott and Brickwedde.

INTRODUCTION Using the same experimental method and apparatus as already described in previous

papers [l, 2] we have measured the velocity of ultrasonic pulses of 1 MHz in gaseous nor­mal hydrogen. The main purpose was to compare these data with values deduced from direct PVT measurements and to judge in that way on the reliability of these latters.

RESULTS In table 1 are collected the experimental data at 63.6, 70.0, 77.4, 83.2 and 90.2°K.

Table 1

T = 63,6° K T = 70,0° K P (kg/cm2) W (m/sec.) P (kg/cm2) W (m/sec.)

2 3 4

246 1401,5 248 1362,0 237,5 1381,0 239,5 1340,2 229 1359,4 229,5 1313,6 219,5 1333,3 220 1288,2 208 1299,2 210,5 1263,0 199,5 1273,8 200 1233,4 190 1243,5 192,5 1210,7 174 1 1 93,6 180,5 1 1 73,7 167 1 167,3 172 1 147,9 1 57,5 1 135,8 160,5 1 1 1 1,7 1 50,5 1 1 10,5 149 1073,3

193

1-19

2 3 4

138 1067,6 140 1042,9 131,5 1042,4 132,5 1017,4 120,5 998,7 1 1 9 970,2 1 1 1 959,0 1 1 1,5 941,6

97 902,0 101 904,9 82 839,9 92 873,2 72 798,3 81 835,9 61 757,4 69,5 798,9 49,5 721,9 59,5 770,7 39 694,8 50 746,2 27,5 674,7 38,5 722,9 20 667,3 28,5 707,6

18,5 697,1

T = 77,4° K T = 77,4°K P (kg/cm2) W (m/sec.) P (kg/cm2) W (m/sec.)

248,5 1349,8 127,5 995,8 238 1310,6 1 1 9 969,1 230 1289,1 108 935,1 220,5 1265,1 98 906,2 210 1235,3 88,5 877,2 199,5 1207,6 78,5 848,9 188 1 175,3 69 825,0 179 1 149,4 59 800,5 169 1 121,0 48 777,1 158 1088,7 36,5 757,6 1 49,5 1063,2 26,5 744,6 139 1030,1

T = 83,2°K T = 90,2°K P (kg/cm2) W (m/sec.) P (kg/cm2) W (m/sec.)

241 1294,9 239 1267,5 231 1270,9 228 1242,5 221 1245,9 220 1223,5 207 1210,8 210 1200,0 191,5 1 1 71,0 200,5 1 176,9 180,5 1 140,7 190 1 150,8 1 66,5 1 104,1 176 1 1 1 6,5 153 1067,4 161 1 079,7 139,5 1028,6 151 1054,6 128,5 997,4 141,5 1030,1 1 1 6 962,2 125 988,9 101 922,5 1 10 953,4

89,5 892,6 99 927,5 76 858,5 84,5 894,3 60 823,9 68 860,5 49,5 804,0 57,5 841,2 40,5 788,3 47,5 824,5 28 771,4 37 808,7 19 760,8 28 797,1

1 8 784,8

194

I-19

The graphical representation in Fig. 1 reveals a quite special relation between the veloc­ity W and the parameters p and T.

Fig. I . The velocity of sound in gaseous normal hydrogen as a function of pressure and temperature

m sec-'

I I 1200 I

w

l

6000:,----������---020�00---�������4�0�0�-s�-A�m

Fig. 2. The velocity of sound in gaseous normal hydrogen as a function of the density (dotted line - results derived by Michels e. a.).

195

I-19

The velocity isotherms between 70 and 90° K intercept all at about 120 kg/cm·2 : at

higher pressures the slope ( ��) p becomes negative as in the case of a liquid. An ana­

logous inversion of the sign occurs also in other gases e.g. in argon at a pressure of 060 kg cm·2 and a temperature between 300 and 450 ° K [3].

In order to show in what extent the equation of state itself is responsible for this inversion we have graphically drawn in Fig. 2 the velocity as a function of the density at different temperatures.

The densities, expressed in amagat-units, are taken by interpolation from the tables of Woolley, Scott and Brickwedde [4]. The velocities as represented in Fig. 2 have a more regular aspect : all these curves are quite parallel and the velocity increases both with temperature and density.

The agreement with the data W ( e) derived by Michels, De Graaff and Wolkers [5] for -175° C, seems also to be good. It has however to be taken into account that there is an initial small discrepancy between both sets of PVT data so that the dotted line in Fig. 2 is plotted in reality with another scale factor.

The velocity of sound can be calculated by means of

W2 = y ( �p) �(! T

Using the W. S. B. tables this formula gives values which correspond completely with the experimental data up to about 50 kg/cm·2 : at higher pressures a systematic deviation occurs being about 1 % at 250 kg/cm·2• The same discrepancy appears in the ratio of the specific heats calculated on one hand directly from the PVT data and on the other hand from the experimental values of W (Fig. 3).

10.0"K

90,Z"K --------.,,,. ................

........

Fig. 3. The ratio of the specific heats from VI'. S. B. tables (dotted line) and from velocity data transformed by the same PVT data (full line).

196

I-19

The deviation Wexp - W calc is caused simultaneously by the imprecision on ( �) T

and

on y. It seems rather difficult and tedious to separate the influence of both factors and to evaluate the corresponding corrected thermodynamical data.

We take the opportunity to express sincere thanks to the Belgian Ministry of Education for its financial aid during these experiments.

REFERENCES

J. A. van ltterbeek and W. van Dael, Cryogenics 2, 2069 (1961).

2. A. van ltterbeek and W. van Dael, Physica 28, 86! (1962).

3 . A. Lacam, Journ. Rech. C.N.R.S. 34, 25 (1956).

4. H. W. Wooley, R. B. Scott and F. G. Brickwedde, Journ. Res. Bur. Stand. 41, 379 (1948).

5. A. Michels, W. De Graaf/, G. ]. Wolkers, Physica 25, 1097 (1959).

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Experimental Determination of HE for the System N2-H2 in the Gaseous State

Determination experimentale de HE par le systeme N2-H2 a l'etat gazeux

M. KNOESTER, K. W. TACONIS and J. J. M. BEENAKKER Kamerlingh Onnes Laboratorium, Leiden, Netherlands

SOMMAIRE. Pour remedier a !'absence de donnees thermodynamiques sur !es melanges gazeux, on a construit un calorimetre a ecoulement, qui permet de mesurer la chaleur de melange. Jusqu'ici les mesures Ont ete effectuees pour trois concentrations du melange N2-H2 a plusieurs temperatures ( entre 145° et 293° K) et a des pressions de 30 a 130 atm.

La description de ces mesures sera suivie d'une breve discussion du comportement de la chaleur de melange en fonction de la pression.

As was pointed out in a panel discussion at the meeting of the Commission I of the International Institute of Refrigeration in 1960, there is a great lack of thermodynamic data on gaseous mixtures. Even for the mixtures of more simple molecules one does not often know the value of properties as volume excess or heat of mixing.

Especially the latter property, which makes it possible to calculate rather accurately the total enthalpy functions of the mixtures, is of special importance for the refrigeration problems connected with gas separation.

To fill the lack of these data we developed a method to measure directly the heat of mixing in a flow type calorimeter.

The two gases, in this case N 2 and H2, are mixed in the main part of the calorimeter, the mixing chamber, that is isolated by vacuum.

Around this chamber a resistance wire has been wound by means of which one can apply heat to the mixed gases. Since the excess enthalpy HE has a positive sign for the system N 2-H 2, which means a cooling effect, it is in this way possible to carry out the ex­periment isothermally.

Temperatures are read on special low temperature N. T. C. resistances before and after the mixing chamber. From the amount of heating and the measured gas flow one obtains directly the value of HE.

Measurements have been performed for three concentrations : 24.5, 52 and 79 % N 2• For the mixture 52 % N 2 - 48 % H2 the results are represented in Fig. 1 as HE vs P curves. The 145°K curve shows a maximum. This appears at this temperature also in the case of the other concentrations.

The temperature dependence is rather large. Since no density data for the mixture are available, it is not possible to plot HE vs d.

Plots of HE/ 4 x ( 1 - x) vs x, where x is the molar concentration of one of the compo­nents, show for the higher temperatures that the relation HE = C x (1 - x), where C is a constant in x, is rather well obeyed. For the lower temperatures the behaviour is better represented by an expression of the form : HE = C x (1 - x) (1 + f3 x).

Coming back to the maximum in the 145°K curve we have calculated the total con· figurational enthalpy values of the mixtures, using the formula H = x H1 + (1 - x) H2 + HE.

The dotted curves in Fig. 2 and 3 represent the H-values in case of ideal mixing (without HE). The differences between full drawn curves and dotted line curves are the HE values (divided by R T). The N 2 enthalpies have been taken from Leiden diagrams [1] in Fig. 2 and from Din [2] in Fig. 3 and the H2 enthalpies from work of the National Bureau of Standards [3]. From these curves it is clear that the maximum in the heat of

199

1-4

800

200

-1

1 4 5 ° K '-... -2

.._

� -

'-....

.- - 24.5%N2 52/. N2

- -s2%N2 79°/0 N2

\ '--

H-H0 RT

� �

-- - °79°/. N2

--.::::::: 100°/0 N2

-30 p 50 100 atm

Fig. 2. H-H0 as a function of P. N2 enthalpies taken from Leiden diagrams.

-I

RT Pure components. Ideal mixtures. Probable corrections.

- - - - 52 °/o N2

145 °K -2 - - - 79 a/o N2

!00°/o N1 -3o��--,P,--�����so,,_������--,'c-=-�����-' Fig. 3. H-H• as a function of P. N2 enthalpies taken from Din.

RT Pure components. Ideal mixtures. Probable corrections.

I-4

201

1-4

mixing does not arise from the properties of the mixture itself, but is caused by the behaviour of the N 2 in the region of the critical point that does not occur anymore in the mixture.

In the enthalpy curve for the mixture there is still a small irregularity. However, a slight change in the N 2-enthalpy values, as shown by the dot-dash lines, will take this away.

Such a change seems well inside the uncertainty on the enthalpy data. Further work on other systems is in progress. This work is a part of the research program of the group for Molecuiar Physics of the

"Stichting voor Fundamenteel Onderzoek der Materie (F. 0. M.)" and was suggested by Dr. P. J. Haringhuizen from the Stikstofbindingsbedrijf der Staatsmijnen in Limburg. It has been made possible by financial support from the "Nederlandse Organisatie voor Zuiver Wetenschappelijk Onderzoek (Z. W. O.)".

REFERENCES

1. W. H. Keesom, A. Bijl and L. A. ]. Monte, Commun. Kamerlingh Onnes Lab., Leiden, Suppl. No. rnS a ; Bulletin de l'Institut International du Froid 23 (1942) Annexe No. 4.

2. F. Din, Thermodynamic functions of gases, Volume 3 (Butterworths, London, 1961). 3. H. W. Woolley, R. B. Scott and F. G. Brickwedde, J. Res. Nat. Bur. Stand. 4• (1948), 440.

DISCUSSION

R. Ayber, Germany : 1) Which combination rule was used in the calculation of the constants of the Beattie-Bridgeman equation of state for using it in the case of N2-H2 mixtures ?

2) How is the agreement between the enthalpies of N2-H2 mixtures when : a) computed from pVT-data of Michels for N2-H2 mixture ? b) computed by means of the Beattie-Bridgeman equation of state as applied to

mixtures ?

3) What are the expressions used for the reduced properties p* and T* ?

Z. Dokoupil, Netherlands: 1) We have calculated the enthalpy by the B-B-equation only for a 3 : 1 H2-N2-mixture of 100°C. We used the square root combination rule for the constant A and c, Lorentz combination for B and linear combination for a) and b).

2) The agreement between the pressure dependent part of the enthalpy of the mixture up till 100 atm. for the methods a) and b) turned out to be better than 5 o/o.

M. Knoester, Netherlands : (to 3) p* = pa3/E, T* = kT/E, the enthalpy was reduced by E. On this point a brief explanation : By means of a law of corresponding states, cal­culations have been carried out for HE. It has been tried to apply a cell model on gases of sufficient high density. The interactions of the molecules are described by the Lennard­Jones potential (with parameter of energy s and of length a). Then the interactions in a mixture have, in the case of the cell model, and "average" potential with for every mixture own parameters [e] and [a], which are functions of the concentration and of the e's and a's of the three kinds of interaction. Using these values of the mixture and the e and a of a pure gas (of which enthalpy H is known) one can calculate H M (and so HE) :

HM (p[a] 3/ [e], kT/ [ej) Le]

H (pa3/t:, kT/e) e

A. van ltterbeek, Belgium : Looking at the results, obtained for N2-A which do not agree very with theory, one may suppose that corrections have to be considered for heat of adsorption and thermal diffusion effects.

M. Knoester, Netherlands: Since the measuring process is a continuous one, no diffi­culties are to be expected from the heat of adsorption. The temperature of the mixture is measured at the outlet side of the chamber of mixing where no concentration gradients occur.

202

1-8

Freezing Pressures of 3He-4He-Mixtures

Pressions de congelation des melanges 3He-4He

C. LE PAIR, K. W. TACONIS, P. DAS and R. DE BRUYN OUBOTER Kamerlingh Onnes Laboratorium, Leiden, Netherlands

SOMMA/RE. Dans ce rapport, on donne les resultats de mesures de p (v) t ejfectuees sur des melanges de 3He-4He, pour sur toute la gamme des concentrations a des temperatures de 0,5 a 1,4° K. Lagamme despressions est de 0 a 27 kgcm-2• On a etudie la congelation du liquide, la separation de phase et la transition A. La plus basse pression de congelation est de 23,58 kgcm-2 a 0,725° K et 60,7% de "He. On presente des diagrammes des deux phases adijferentes pressions. A une pression de 24,0 kgcm-2 la region non homogene de liquide et de solide n'est pas liee a la region non homogene des deux phases liquides stratifiees.

Continuing our measurements published in Physica [l] and reported at The 8th Conference on Low Temperature Physics [2], we determined the freezing pressures of 3He-4He-mixtures down to 0.5°K. This time we used an apparatus in which the meas­uring vessel (a bourdon tube) was attached to a 3He-cryostat. Here for measuring the pres­sure we did not use our mirror detection system [2]. Instead we attached a piece of ferrox­cube to the free end of the bourdon tube, the displacement of which within a supercon­ductive coil was measured with an Anderson bridge method. This permitted relative pressure measurements with an accuracy of 0.006 kg cm-2. Unfortunately the absolute measurements were much less precise, the instrument we used as a standard (a Heise manometer) did not reproduce exactly.

We were able to detect minima in the freezing pressures such as were observed first by Lee and coworkers [3, 4, 5].

THE RESULTS

The results of our work are given in Fig. 1. The drawn lines represent the freezing pressures of mixtures of different concentrations as a function of temperature. The dashed lines represent the A-pressures as a function of temperature of the 48.2% and the 60.7% mixture. The -- . -- line marks the onset of phase separation in the fluid of the 60.7 % mixture.

An explanation for the occurrence of a minimum in the freezing pressures must be as a v

sought in the influence of the fix and ax terms in the Clausius-Clapeyron equation for

binary mixtures :

( :�) X, freezing = � s - (!�) �x

� v - (��) �x (6.)

As none of these terms are known at high pressures it is difficult to make predictions about the slope of the freezing curves.

PHASE-DIAGRAM

From Fig. 1 the solid-liquid phase diagram of these mixtures at pressures between 25.9 and 23.6 kg cm-• can be constructed. It is rather peculiar. Our results are given in Figs . 2 and 3. In Fig. 2 the triangles represent Lee's data and the circles represent

203

1-8

.!.i c m' 27

25

24

23 iP.s. I I I

,, I I I I I

I

I I I I I I • I I I I \ I I I

1.5

Fig. 1. Drawn lines represent freezing pressures.

204

Dashed lines represent A-pressures. -- · -- represents the onset of phase separation of the 60,7 % mixture curve A �orres­ponds to pure He. curve B corresponds to a mixture with a molar fraction •He = 0.085 curve C corresponds to a mixture with a molar fraction 3He = 0.228 curve D corresponds to a mixture with a molar fraction •He = 0.482 curve E corresponds to a mixture with a molar fraction •He = 0.607 curve F corresponds to a mixture with a molar fraction •He = 0.801

T 0 gives the accuracy for a complete line. _I_ 2.0 ..------..,.------,

1.5

1.0

0.5

Tl ' I ' '

0 ___!...

P: 24.8 k9 c�

0.5

Fig. 2. phase diagram of a 3He-4He mixture at a pressure of 24.8 kg cm-2.

v represent measured points of the A-line. o represent measured points of the

freezing line. I!>. represent measured points of the freez­

ing line measured by Lee a. o. o representsextrapolatedpointsof the phase

separation line in the liquid, see fig. r . The dashed line represents the phase dia-1.0 gram of Lee a. o.

2.0 -----�-----,

P : 24.0 kgc�

1 .5 He I.

1.0

C2)., \ \ \ 0.5

0 __.!__ 0.5 Fig. 3. Phase diagram of 3He-4He mixtures at 24.0 kg cm·2•

v represent measured points of the A-line.

'

1.0

o represent measured points of the freezing line. o represent extrapolated points of the phase separation line in the liquid.

1-8

ours. Inside the upper (inhomogeneous) region a region of solid must exist. Up till now, we have no evidence whether the melting lines are close to the freezing lines or not. In Figs. 2 and 3 we give two points (squares) which are points of the phase separation line in the liquid at these pressures, which could also be detected with our apparatus. These points have been found by extrapolation of a curve in which the phase-separation onset temperature in the liquid of the mixtures ( 48.2 % and 60. 7 %) was plotted against pressure. We found a decrease in onset temperature of the phase separation of the 48.2 % mixture from 0.75°K to 0.49°K and for the 60.7% mixture from 0.84°K to 0.54°K. According to these observations we expect the line of the three coexisting phases to be much shorter than that found by other investigators [5] at a pressure of 24.8 kg cm·•. At 24 kg cm·• we suppose the freezing region to become free from the phase separation in the fluid.

The measurements will be extended to cover the melting lines of the solids.

NEW DEVELOPMENTS

An explanation for the occurrence of the minima has been given [7] . Due to the same arguments it can be understood why the freezing pressures of mixtures are lower than those of pure 4He. With the same apparatus we used for the measurements on the mix­tures we investigated the freezing curve of pure 4He [8]. A minimum was found. The pressure difference between melting at the minimum at 0,75° and at 0,5°K is ca 0,008 kg. cm - 2.

REFERENCES

1. C. le Pair, K. W. Taconis, R. de Bruyn Ouboter, P. Das, Physica 28 (1962), 305. 2. C. le Pair, K. W. Taconis, R. de Bruyn Ouboter, P. Das, Progress of the 8 th Conference on Low

Temperature physics (to be published). 3. H. Weinstock, F. P. Lipschultz, C. F. Kellers, P. M. Tedrow, D. M. Lee, Progress of the 8 th

Conference on Low Temperature Physics (to be published).

205

I-8

4. H. Weinstock, F. P. Lipschultz, C. F. Kellers, P. M. Tedrow, D. M. Lee, Phys. Rev. Letters 9 (1962), 193·

5. D. M. Lee e. o. private communication, reporting their communication to the American Physical Society on Thanksgiving meeting r962.

6. R. de Bruyn Ouboter, ]. ]. M. Beenakker, Physica 27 (1961), ro74.

7. C. le Pair, K. W. Taconis, P. Das, R. de Bruyn Ouboter, Cryogenics 3 (1963) I I 2 K . W. Taconis, R. de Bruyn Ouboter, Progress i n Low Temperature Physics IV, edited by C. J. Gorter, to be published.

8. C. le Pair, K. W. Taconis, P, Das, R. de Bruyn Ouboter, Physica 29 (r963) 755.

DISCUSSION

E. F. Hammel, U. S. A . : How did you determine the locus of the A-line ? Was it based on the position of the minimum in the specific volume temperature curve ?

C. le Pair, Netherlands: We did not take the minimum in the specific volume for A-transition. We took the point where the change in volume with temperature had its maximum.

E. F. Hammel, U. S. A.: 1 . In connection with the location of the locus of the A-lines for 3He-4He mixtures, it may be of interest for you to know that Dr. Eugene Kerr of the Los Alamos Scientific Laboratory has recently completed a series of measurements on the expansion coefficients of 3He- 4He mixtures over a range of concentration and in the vicinity of the A-line. His results show that the locus of the zero in the expansion coefficient does not coincide with the locus of the A-point as determined from specific heat or thermal conductivity measurements. Instead the A-point locus always lies below that of the expansion coefficient zero. The former begins with pure He 4 about 6 milli­degrees above the A-point and decreases steadily as the He 3 concentration increases, meeting the pure He3 axis at about 0,5 ° K.

206

Mag n et i s m Mag n eti s m e

1-2

Flux Trapping and Flux Pumping with Solenoidal Superconductors *

Piegeage du flux et pompage du flux avec des supraconducteurs soleno'ides

HENRY L. LAQUER Los Alamos Scientific Laboratory, University of California, Los Alamos, New Mexico, U.S.A.

SOMMAIRE. On decrit les techniques permettant de determiner les proprietes des supra­conducteurs de grande section transversale ainsi que les divers es methodes d'utilisation de ces materiaux pour la production de champs magnetiques durables.

1 . INTRODUCTION

It is generally agreed [1] that hard, high current superconductors conduct throughout their entire volume, possibly by some mechanism involving dislocation chains [2]. The bulk conduction mechanisms must be superimposed on the more generally studied sur­face conduction effects. The present report describes some of the cryogenic and magnetic techniques developed to study the properties of superconductors of large cross section and presents preliminary results on these properties for a number of Nb-Sn samples. It also describes the utilization of these materials for the construction of persistent magnets.

2. APPARATUS AND TECHNIQUES

Before any particular superconducting material can be used in a magnet or other practical structure one should know the variation of the maximum permissible, or criti­cal, supercurrent density, I c/A, with the magnetic field H at various temperatures T. This curve usually shows a hyperbolic drop from high values at negligible fields to a "knee" or inflection at the Kunzler field HK [3] after which it drops more rapidly to immeasur­ably small values at the critical field He.

Conventional methods [1] for determining critical currents use external current sour­ces and determine the current at which a measurable voltage, often as small as 10-9 volts, appears across the sample which is in a cryostat and in a magnetic field. Current den­sities as high as 104 and 105 amperes/cm2 have been reported [3] and since it is difficult to conduct currents much greater than 100 amperes into a cryostat, the cross-sectional area of a typical sample is limited to 10-2 or 10- 3 cm 2, which represents a very fine wire. With many alloy superconductors such fine wires are the material with the greatest work hardening and hence highest critical current densities ; however, with other super­conductors such as intermetallic compounds and particulary Nb-Sn and similar compo­sitions made by powder metallurgical techniques, there is no advantage in restricting oneself to the very small cross sections. As a matter of fact, there are many mechanical and electrical advantages in using conductors of the largest possible cross sections for the construction of magnets.

The present method [4] simply involves the placement of a short circuited coil of heavy wire or ribbon, or of a cylindrical shell fabricated from an appropriate super-

* Work performed under the auspices of the U.S. Atomic Energy Commission.

207

1-2

conductor inside the cryogenic region of a high field magnet. Induced currents are set up by changing the applied magnetic field. The magnitude of these shielding or trapping currents is calculated by observing the field inside the superconducting shape using a calibrated bismuth magnetoresistance probe [5]. There is no need for high current leads and the only limitations on specimen size is given by the diameter of the cryogenic high field region, which is 4.9 cm in our particular 80 kOe installation [5]. The magnetore­sistance probe has a diameter of 0.7 cm and can slide up and down inside the 1.25 cm diameter specimen support tube, so as to allow detailed field plots.

3. FLUX JUMPS AND FLUX CREEP

The measurement simply consists in recording (on an XY recorder) the internal field Hin seen by the magnetoresistance probe as a function of the applied or external field Hex as indicated by a current shunt. In the absence of a superconductor or above Tc a single reproducible calibration line is obtained. With a superconducting sample in place, shielding currents will be generated when the applied field is increased gradually. As soon as these currents exceed I c the sample may suddenly and temporarily go normal permitting all or most of the field to penetrate in a "flux jump" as reported by Swartz and Rosner [6] and by Autler [4], or the field may leak or creep in as discussed by Kim, Hempstead, and Strnad [7]. On decreasing the applied field from a high value after some penetration has taken place, the superconductor will again attempt to keep the enclosed flux constant by setting up trapping currents. Once these exceed permissible critical values, flux jumps or flux creep will again occur. As pointed out by Kim et al. [7] the distance between the envelope traced by the creeping curve or drawn between the jum­ping points and the calibration line represents the maximum magnetization the specimen can produce. This critical state (for cylindrical shells) or limiting critical current (for a short circuited coil) is a reproducible property of the sample.

I (Komp)

50

10

0

-10

-so

I (amp)

1 5 00

1000

500

0

·500

-1000·

- 1 500

208

6H (Koe)

20

10

-10

-20

t. H (Koe)

· 1 0

· 5

-10

10 20 30 40

1 0

N b S n C - 3

5 0 so "u ( Koe)

NbSn JZ: m

Fig. r . Magnetization o f cylin­drical shell C-1

Fig. 2. Magnetization of coil V m'.

1-2

Figs. 1 and 2 show some typical magnetization curves obtained in the present study. The ordinates are marked in both magnetization and current units. The abscissa is the applied field Hex· After making a correction to average the field on the inside and out­side of the specimen the resulting critical current curves obtained from the shielding and trapping regions should be identical. Kim et al. [7], who used long thin cylindrical shells or tubes in order to avoid end effects, still found slight discrepancies. These are even more pronounced for the larger and shorter cylindrical shells used in the present work. On the other hand, the agreement between the shielding and trapping regions appears to be better with the short circuited coils where only a single central turn will be at the critical state.

Both jumping and creeping may be observed with the same specimen even at the same temperature. The jumping is more likely to occur at low fields and high current densities whereas creeping is more likely where current densities are low, i. e., with poorly super­conducting specimens, and at high fields and at high temperatures (17 and 16° K for Nb3Sn specimens). Creeping is also more likely to be found in liquid helium below the A-point where heat transfer is improved. On the other hand, flux jumps can always be forced by rapid changes of the applied field.

All these observations agree quite well with Anderson's [8] theory of hard supercon­ductors, in which the onset of normal resistivity is not a step function of the magnetic field but merely a logarithmically increasing function. Thus, the temperature of the specimen will rise above that of the bath by an amount which is governed by the balance between heat input from the moving flux lines and heat transfer to the bath (divided by the heat capacity of the specimen and closely associated parts). The variation of the super­conducting properties with temperature will in turn determine whether the rate of change of the temperature accelerates, leading to a flux jump or decays, leading to creep.

4. PROPERTIES OF SOME Nb-Sn PREPARATIONS

The present study has so far been limited to the properties of a number of NbSn preparations. Large, but reproducible variations in the critical current densities can be obtained from relatively small variations in fabrication techniques. All samples were made from mixed powders of approximately Nb3 Sn stoichiometry, fabricated and sub­sequently heat treated in the neighborhood of 1000 ° C. Table I lists preparation and

Table I. Cylindrical Shells

1-Ieat Treatment Dimensions

Designation Time Temp. Atmosphere I. D. O. D. (hours) (O C)

C-2 9 1000 vacuum 1.70 2.30 C-3 14 950 argon 1 .70 2.30

4 1000 C-5 8 1000 vacuum 1.78 2.28 C-6 2 1000 vacuum 1 .7 2.25 C-7* 14 1000 argon 1 .65 2.2

8 1 150 C-8* 14 950 argon 1 .67 2.20

4 1000 C-13A 14 950 argon 1.71 2.32

4 1000

* pre-reacted powders 14 hours at 950°C in argon then re-ground and re-pressed * * plus exterior in con el shell

(cm) Length

3.90 4.07

3.8 * * 3.92 4.0

4.1

4.00

209

I-2

dimensions for a number of cylindrical shells pressed from -325 mesh screened pow­ders. Vacuum when used was at least 8 x 10-6 torr, argon atmosphere, where indicated, was gettered with zirconuim hydride. Table II gives similar information for a number

Table II. Short Circuited Coils

Heat Treatment Dimensions (cm) Turns Designation Time Temp. Wire Dia.Core Dia. I. D. O. D. Length N

(hour) o c

V i 2 1010 0.25 0.125 2.1 2.6 2.8 10 V m 2 1000 0.25 0.125 2.0 3.1 2.5 18 V m' 2 + 9 1000 0.25 0.125 2.0 3.1 2.5 18

IX d 8 1000 0.25 0.20 2.0 2.55 2.9 10 !X e 8 1000 0.25 0.20 2.0 2.5 2.7 10 IX f 8 1000 0.25 0.20 2.76 3.3 2.2 8

of short circuited coils made from 0.25 cm 0. D. composite (Nb shell and Nb8 Sn core) wire prepared according to Kunzler's [3] technique. All coil samples were heat treated in a vacuum. All samples were slowly cooled in the furnace taking about 12 hours to reach room temperature.

The results are summarized in Table III which lists critical current densities (in

Table III. Critical Current Densities at 3.95°K

H (kOe)

Specimen

C-2 C-3 C-5 C-6 C-7 C-8 C-13A Vi Vm Vm' !Xe IXf

5

34 150 51 390

124 700 91 750

102 030

51 710

10 20 30

62 500 49 150 39 120 81 130 61 310 48 320 62 410 38 830 25 050 66 730 50 210 39 230 25 450 16 940 14 190 38 730 23 140 17 570 57 420 42 030 32 040 82 230 49 200 35 720 67 900 37 800 27 530 75 970 48 810 35 230 83 490 36 180 22 260 35 830 19 940 1 1 830

40 50 60

34 610 32 150 30 610 42 170 38 120 35 080 19 590 16 400 34 700 13 460 14 640 12 850 11 710

16 850 23 120 22 020 26 420 22 020 19 080

3 790 8 450 3 380

amperes/cm •) for various field values (in kOe) at 3.95°K. The current densities for the wire samples are referred to the cross sectional area of the core only.

Typical current densities are between 20000 and 100000 amperes/cm• and are gener­ally only 20 to 30 per cent of the results reported by Kunzler [3], Kim [7], Wernick [9], and others at the Bell Telephone Laboratories.

Our samples thus represent what can fairly readily be obtained in any laboratory with commercially available materials. However, there is one areain which ordinary commercial procedures are insufficient, namely the purity requirements of the atmosphere while the sample is at elevated temperatures. A drop in vacuum from 10-6 torr to 10-4 torr, or a non gettered argon atmosphere will produce completely useless samples. It may well be

210

I-2

that the quality of our samples is still limited by the quality of the atmospheres used during heat treatment. There is no doubt that all efforts to improve these atmospheres and to protect the samples by other means will be very worthwhile indeed. It is also evident that the wire coil samples appear to drop off more steeply and irregularly with increasing field than do the cylindrical shells. This is probably caused by some weakness at the contacts. Various procedures for improving the contacts are presently under investigation.

5. PERSISTENT MAGNETS

The results of the previous section indicate that it should be possible to build fairly large superconducting magnets of at least 60 kOe using the 0.25 cm diameter composite wire. There are a number of advantages in using this wire rather than still heavier ma­terial such as the cylindrical shells, the main one being a more exactly calculable field profile since there is no possibility of an axial or radial variation of the current density. On the other hand, the cylindrical shells are more readily fabricated, require no con­tacts, and are more efficient since they have the maximum possible space factor, namely 100 per cent.

With all of these materials of large cross section, procedures must be developed which create the required large currents of kilo ampere magnitude within the cryostat and dis­pense with heavy leads from room temperature to 4 ° K.

5.1 Trapped Fields Careful manipulation of the flux jumps described in section 3 has given permanently

trapped fields of 24 to 30 kOe with the better samples (C-2, C-3, C-6) all of which had similar dimensions (1.7 cm I. D., 2.2 cm 0. D., 4 cm length). The field profiles were usually somewhat distorted indicating nonuniform current densities. These result from the tendency of the cylinders to approximate the flatter field applied by the 12.7 cm long high field magnet coil. Once the trapped fields had persisted for a minute they would re­main persistent for hours (so long as the cryostat stayed cold).

In order to prove the practicality of using high field magnets as central "charging stations," a small portable helium dewar of 3.6 cm inner diameter has been built which fits inside the high field region and permits removal of the trapped fields from the high field magnet. With this arrangement, fields of 10 to 12 kOe were moved freely around the laboratory (using one of the poorer samples, C-5). These fields would collapse with rather pronounced sound effects once all the liquid helium was removed either by eva­poration or by pouring out of the dewar.

The short circuited coils generally contained less material and hence trapped lower fields. However, except for the calculable field profiles, there appear no other qualita­tive differences in their use as persistent magnets.

5.2 Pumped Fields All the trapping procedures demand external fields larger in magnitude and in volume

than the maximum trappable fields. Because of the heating and rate effects discussed in section 3 it appears likely that pulsed fields will not be suitable for charging persistent magnets. On the other hand, the great capital cost of D. C. high field magnet installations limits their availability.

There are a number of ways of pumping up the field strength in a region which is surrounded by superconductors. Only a small initial field is required and the pumping can be done either mechanically [10] or electrically [11]. Mechanical flux pumps require the motion of superconducting pistons within the cryostat and hence involve the trans­mission of rather formidable forces. The electrical method described by the author [11] is apparently free of some of these problems and since it uses wire coils it allows con­siderable freedom in field configuration. On the other hand, there is the previously men­tioned problem of making reliable contacts. It is interesting to note that the current of 1060 amperes created in a 21 turn Nb-Sn composite wire coil (of 4.1 cm I. D., 5.1 cm O. D., and 2.5 cm length) and generating a field of 5.4 kOe falls on the critical current curve of similar wire reported in Table III. Hence it should be mainly a matter of

211

J-2

winding a coil from a larger piece of wire than the 3.8 meters available at that time in order to create larger fields. This effort is now in progress.

The author would like to thank Messrs. R. W. Keil and H. Sheinberg for preparing the various samples used in this work.

REFERENCES

r. ]. E. Kunzler and M. Tanenbaum, Scientific American 206, No. 6, op. 6 (r962).

2. R. W Shaw and D. E. Mapother, Phys. Rev. n8, 1474 (1960).

3. Kunzler, Buehler, Hsu and Wernick, Phys. Rev. Letters 6, 89 (1961).

4. Similar methods have been used previously in many laboratories, e. g. Autler, S. H., High Magnetic Fields, edited by Kolm, Lax. Bitter, and Mills. M. I. T. Press 1962, chapter 34, p. 324.

5. H. L. Laguer, ibid. chapter 13, p. 1 56.

6. P. S. Swartz and C. H. Rosner, J. Appl. Phys. 33, 2292 (1962).

7. Y. B. Kim, C. F. Hempstead and A. R. Strnad, Phys. Rev. Letters 9, 306 (1962).

8. P. W. Anderson, ibid. 9, 309 (1962).

9. ]. H. Wernick, Superconductors, edited by Tanenbaum and Wright, lnterscience Publishers 1962, p. 42.

ro. D. E. Elleman and A. F. Hildebrandt, Eighth International Congress on Low Temperature Physics, London, 1962, paper C G.

r J . H. l .. Laguer, Cryogenics 3, March 1963.

DISCUSSION

A. H. Cooke, U. K. : Can the author say what determines whether flux jumps or flux creeping will occur ?

H. L. Laquer, U. S. A.: I believe that is primarily determined by thermal conditions. Small local flux jumps are probably unavoidable in any hard superconductor. All one can try to do is to slow these flux motions down inductively so that the rate of heat input does not exceed the rate with which the heat can be carried away (without excessive temperature rise in the specimen and consequent degradation of current carrying capac­ity). Big or catastrophic flux jumps can always be forced by fast changes of the applied magnetic field Hex· Flux creep or non-accelerating flux motion can be made more likely by the incorporation of good normally conducting material within the superconducting structure.

J. R. v. Geuns, Netherlands: Are the currents not uniformly divided over the thickness of your cylindrical shell ? A uniform current distribution would give a linear change from the inside field to the outside field, whereas the magnetic field to be shielded (or trapped) is constant over this thickness.

H. L. Laquer, U. S. A.: This is quite probable. The field and current distributions will certainly depend on how any given magnetic state has been reached, in particular whether it has been by stable creep or by abrupt jumps. Only in the case of creep is there a possibility of really predicting current patterns. Jumps do not always heat the entire specimen into the normal state so that extra trapped flux may remain near one end. Thus we have observed some trapped field profiles with pronounced axial asymmetry.

P. M. Roubeau, France: Did you observe a decrease of the collapse times with de­creasing temperature ?

H. L. Laquer, U. S. A. : Photographs of the voltage induced in a pick-up coil inside the cylindrical shells show that flux jumps have three stages : (1) an exponentially in­creasing signal lasting 50 to 100 µsec, (2) a further slower increase for another 100 µsec, (3) a long decaying tail of about 800 µsec. Each of these regions can be understood in

212

1-2

terms of the temperature of the specimen and the instantaneous difference in flux pressure, i. e. the force driving the flux lines. However, we have not as yet taken sufficient pictures at different temperatures to analyze this in detail.

L. v. Gerven, Belgium: I would like to mention that we have done similar experiments in Leuven for metals with low critical fields. The field in the cylinder was measured by means of EPR. The critical field could be determined rather accurately, if the cylinder is sufficiently long and if the material is homogeneous enough.

P. Grassmann, Switzerland: The mentioned difficulty caused by the mutual inductance between field coil and sample may be overcome by a pick-up coil within the magnetic field and compensating both inductances. This method is used with good success by Olsen under about the same conditions as in this paper.

J. Wilks, U. K. : These measurements of the change in heat flow when the metal becomes superconducting are very informative because exactly the same surface is used in both conditions. Hence the uncertainties involved in comparing the behaviour of different specimens are avoided. The only other types of experiment satisfying this condition are those of Oransfeld, Newell and Wilks (Proc. Roy. Soc. A 260, 31, 1961), in which the Kapitza resistance was measured as a function of the pressure on the helium. It is clearly necessary that any satisfactory account of the change in resistance at the superconducting transition must also account for this pressure dependence. It would therefore be valuable to make measurements of the pressure effect at the same time as observing the effect of the magnetic field.

213

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A Flux Pump for Generation of High Currents in a Superconducting Foil Magnet

Une pompe a flux magnetique pour la production de courants eleves dans un aimant a feuille de metal supraconductrice

H. VAN BEELEN, A. J. P. T. ARNOLD, R. DE BRUYN OUBOTER, J. J. M. BEENAKKER and K. W. TACONIS Kamerlingh Ormes Laboratorium, Leiden, Netherland�

SOMMA IRE. On decrit un nouveau type de pompe a flux magnetique dans laquelle le «robinet supraconducteur» est actionne par l'aimant de pompage lui- meme. On donne quel­ques resultats d' essais, utilisant la pompe a flux magnetique en combinaison avec un aimant a feuille de plomb. On a produit des courants atteignant 80 A, ca/cutes a partir du champ mesure. Le rapport de la temperature a la valeur de la saturation correspond au rapport de la tempe­rature au champ magnetique critique de Pb.

The conservation of fluxoid in multiply connected superconducting bodies [1] makes possible the construction ofmagnetic"flux pumps and supercurrent generators [2, 3, 4].

The development of high field superconducting magnets opens a very useful applica­tion for such devices. Instead of feeding these magnets from outside, which makes small currents and consequently many windings necessary to avoid excessive heat input into the low temperature bath, one can use even single-turn magnets in which large persistent currents are generated by means of a flux pump [5, 6, 7, 8].

We have developed a new type of flux pump which was tested in combination with a lead foil magnet as shown in Fig. 1. For clarity the outer winding has been drawn unwound

A

n

B

Pb

Fig. 1 . Schematic diagram of the apparatus ; for clarity the outer winding is drawn unwound. The pumping magnet is moved up and down between A and B.

in the diagram. A Pb foil is wound on a 2 cm diameter tube. The inner and outer win­dings are connected to each other as indicated, by means of Woods-metal solder (Tc = 8.3°K). A small Nb-coil magnet, the coils of which are wound on two bars connected to

215

I-7

an Armco-U-frame, is moved up and down between A and B, disturbing superconductivi­ty over a small area of the Pb foil. As the magnet moves from A to B, the flux, embraced by the edge of the normal spot, is brought into the hole, without causing a current to compensate for the flux change. Subsequently in B the direction of the field is reversed, generating a current in the circuit to maintain the total flux embraced originally. In mov­ing back to A, with the opposite field, a normal area is again formed so that the current through the circuit is left unchanged. By changing the field direction in position A and repeating the cycle, flux is built up in the foil magnet. The circuit can be emptied by reversing the field and repeating the cycle.

The field in the foil magnet was measured by a pick-up coil, moving up and down with the Nb-magnet, into and out of the foil magnet. The current through the foil was calcu­lated from the measured field. The flux changes during the strokes were measured by a second coil, wound around the foil magnet. The measuring coils were connected to ballistic galvanometers.

Some of our results are shown in Fig. 2, where the current through the foil is plotted

A

80

60

40

20 1 20mA

: ----- T = l.5°K ' T: 24°K

Fig. 2. Current in the foil plotted against the number of strokes for different fields of the Nb-mag­net. Saturation levels at 3 different temperatures are indicated. The amount of flux pumped at every stroke is given by the current through the coils of the Nb-magnet.

against the number of strokes. The saturation value is independent of the field strength of the Nb-magnet (indicated by the current through the coils at each curve) as long as this field is larger than the critical field of the Pb. The area of the normal spot is dependent on field strength, decreasing with decreasing field strength.

The temperature dependence of the saturation value is indicated in the diagram and corresponds to the critical field dependence of Pb. Currents up to 80 A were reached.

Extension to hard superconductor foils is in progress, as there are obvious advantages in foil magnets in comparison with wire magnets.

One of the features of our flux pump is the possibility of changing the field in the foil magnet practically at will by regulating the field of the pump magnet. By reversing the field in B slowly we can continuously adjust the rate of increase of current in the foil magnet.

216

I-7

In diminishing the number of turns of the foil magnet the saturation field kept constant, so that larger currents were achieved, indicating that the field and not the current deter­mines the saturation value.

It is a very great pleasure for us to thank Mr. E. S. Prins for his prompt and inventive technical assistance in building and discussing the apparatus.

REFERENCES

r . F. London, Superfluids I, Wiley New York.

2. D. D. Elleman and A. F. Hildebrandt, Preprints of LT 8 conference, London 1962, pag. 295.

3. ]. F. Marchand and ]. Volger, Phys. Lett. 2 (1962), u8.

4. ]. Volger and P. S. Admiraal, Phys. Lett. 2 (1962), 257.

5. F. Rothwarf, R. C. Thiel and S. H. Autler, Bull. Am. Phys. Soc. 7 (1962), 189.

6. D. D. Ellemann, A . F. Hildebrandt, R. Simpkins and F. C. Whitmore, Bull. Am. Phys. Soc. 7 (1962), 309.

7. A. F. Hildebrandt, H. Wahlquist and D. D. Ellemann, J. Appl. Phys. 33 (1962), 1 798.

8. A. F. Hildebrandt, D. D. Ellemann, F. C. Whitmore and R. Simpkins, J. Appl. Phys. 33 (1962) . 2375.

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Equipment for Producing Pulsed Magnetic Fields of High Intensity and Magneto-Resistance Measurements on Germanium

Materiel de production de champs magnetiques pulses de grande intensite et mesures de la magneto-resistance sur du germanium

H. HENDRICKX, H. MYNCKE and A. VAN ITTERBEEK Instituut voor Lage Temperaturen en Technische Fysica, Leuven, Belgique.

SOMMA/RE. On decrit une methode de production de champs magnetiques d'une duree de 50 µs a 1 msec .. On utilise une batterie de condenseur de 1000 µ F, chargee jusqu' a 3000 volts. On etudie !es problemes de construction d'un serpentin resistant du point de vue me­canique. On a atteint des champs de 400000 Oe dans un volume de 1,5 cm3•

Description de l'effet de magnetoresistance sur du Ge de type n.

1. THE APPARATUS

A very simple method to obtain magnetic fields of high intensity is to produce pulsed magnetic fields. One of the most economical ways is to use condensors as the energy source and an air-cored magnet coil which forms with the storage condensors an os­cillatory circuit.

With the circuit as is represented in Fig. 1 the switch is closed when a pulse is desired and opened after the first half period. In practice only an electronic switch is useful.

3000 v

Fig. r . Scheme of apparatus for pulsed magnetic field production.

In the first equipment we employed an ignitron which has to be ignited by means of a discharge of a capacitor through a thyratron. The thyratron itself is triggered with the photoapparatus by means of an amplifier and a relay (Fig. 2). Several switches in the apparatus prevent wrong manipulations.

2. THE COILS

If it would be possible to convert all the energy stored in the condensors into a uni­form magnetic field H o within the cylinder enclosed by the inner diameter of the coil, then we should have

H2o - x n a2 2 I = 1 /2 CV2 • 10 S n 1 (1)

We remark that we charge up the condensorbattery of 1000 µ F up to 3000 volts. This is an energy of 4,5 kJ.

2 1 9

1-13

Fig. 2 . Trigger-system for the thyratron

From (1) we deduce that

H o = (2� �r� : H o is given in Oersted when C is given in µ F and V in volts. The symbols representing the coil dimensions are defined in Fig_ 3. In pratice we have to introduce two corrections_ In the first place the field is not uniform and secondly, there are resistive losses. Thus we can express the field with :

H = (20 C) Yz _J;' SJ 1 a1

Fig. 3. The dimension of the coil

S depends only on the shape of the coil and results from the fact that the field is not uniform and not completely contained in the cylinder limited by the inner diameter of the coil.

We computed these values for one hundred coils of different shapes. The formula is rather complicated :

n ( i) Yz a2 + ( � + �) Yz

S = 10 X 5 Yz ( ::- 1) A

Yz ln

a1 + (}1:

+ :) Yz

J depends on the qualities of the LC-circuit. Its expression is more complicated. In practice we measure the magnetic field with a pick-up coil placed in the core of the coil. The signal of the pick-up coil is integrated by a RC circuit and gives us the magnetic field which is compared with the current in the coil (Fig. 4). These results we compare with the following formula which gives us the field in function of the current-density for a given shape of the coil in a given place of the core.

220

Fig. 4. Magnetic field compared with the current in the coil.

The unit of length is the inner diameter.

H o Ja1 [ IX + ( ({J + !;)2 + oi:2 }

o,6283 <fJ + t;) in !+l<.B-+-�T+ 1 1 Yz

+ Yz a. + I <fJ - t;)2 + a.2 I Yz

+ ({J - l;) In 1 + ( ((J - !;)2 + 1) ) Y2

1-13

We calculated this formula also for a one-hundred of shapes with an electric computer. Now we will discuss some coils in detail. We started with very simple coils (Fig. 5). They were wire-wound and soaked in Araldite and strengthened by a kind of cement

brass

araldite Fig. 5. Construction of a simple coil.

and brass bolts. All these coils exploded at fields of about 200.000 Oersted. Then we made Bittercoils (Figs. 6 and 7). They were machined from a solid bar of copper or Be-Cu. The insulation was Araldite and strengthened by a kind of cement or a stainless steel cylinder. At first these coils were not successful. The forces try to make the coil shorter and the contacts at the ends of the Bittercoil were pulled loose at fields of about 300.000 Oersted. The coils themselves strengthened with a stainless steel cylinder (Fig. 7) showed no damage at all. Secondly, the self-inductance of these coils is very low and so the pulsetime is very short. In this case the use of an ignitron is no more successful because of the fact that the deionisation time of the ignitron is very long and the circuit will not be opened after the first half period. All the energy then is transformed into heat in the coil and this gives a larger damage. Secondly, the parasitic self-inductance

221

1-13

ce/eron

araldite

stainless steel Be cu coil

brass

Fig. 6

Fig. 7 Figs. 6, 7. Construction of an employed Bittercoil .

of the circuit is considerable und the energy transfer is not good. Therefore we con­structed a second equipment, discussed later, more suitable for short pulses.

Thirdly, we made wire-wound coils strengthened with Araldite and a brass cylinder. The current induced in the cylinder will reduce the field but the gain in field by the reinforcement is more considerable. With such coils we reached fields up of 400.000 Oe. without explosion. But we expect better results of Bitter-coils in the second equipment. Further results will be discussed at the conference.

3. MEASUREMENTS OF THE MAGNETORESISTANCE OF N-TYPE GER­MANIUM

The method of measuring the magneto-resistance is very simple. We use the bridge method (Fig. 8). We keep a constant current in the sample and measure the voltage drop in function of the magnetic field-strength by means of an oscilloscope.

222

500 JI'

� L___?-osc

Fig. 8. Bridge method for measuring the magneto-resistance.

1-13

Fig. 9. Voltage drop as function of the magnetic field-strength.

But we do not have a constant magnetic field during the pulse and there is a strong induced voltage in the measuring circuit. Only at the top of the field the induced voltage will be zero. So we have only one significant point in the curve giving V as a function of H. Fig. 9 represents such an enregistration.

More detailed results of these measurements will also be given at the conference itself.

4. SECOND EQUIPMENT

We discussed already the disadvantages of the first equipment using Bittercoils : the ignitron was no more suitable and the parasitic self-inductance was too large, for a good transfer of energy. Fig. 10 represents the block diagram of the second equipment (Fig. 1 1). Instead of an ignitron two spark-gaps are used. The 1000 µ F condensorbattery is charged up to 3000 volts. Over the capacity Cs we have 6000 volts. So we have a voltage of 3000 volts between A and B because the polarity of C1 and C2 is opposite and spark-gap 2 does not start. The trigger system starts spark-gap 1 and C1 discharges through the coil. After the first half period the polarity of C1 and Cs is no more opposite. The voltage between A und B exceeds 6000 volts and spark-gap 2 crowbars the circuit.

The trigger circuit (Fig. 12) consists in principal of a capacity which is discharged in a pulsetransformer by a thyratron PL2ID2. The output is a 50 µ sec. pulse of about 10.000 volts. This trigger system itself is started by the photoapparatus.

3000V y

6000V

+ c A

Trigger System

Fig. 1 0. Block diagram of the second equipment.

Coil

223

I- 13

-] - �

Fig . I r . Second equipment.

6L6 PL2 ID2

Ph.A. Fig. I 2. The trigger circuit.

The resistance of 10-4 ohm used for measuring the current in the coil consists of a German silver plate with axial symmetry. Fig. 13 gives a section of this resistance.

Cu

Fig. ' 3· German silver plate resistance for measuring the current in the coils.

FM FG

1000V

KV

Fig. 14. Measuring system for the magnetic field.

224

1-13

In this equipment we measure the magnetic field by the magneto-optical Faraday­effect in flint-glass. A parallel beam of polarized light is sent through a 2 mm plate of flint-glass and the rotation of the polarisationplane is recorded by an analysator and a photomultiplier. (Fig. 14).

CONCLUSION

The interest of low temperatures in all these experiments is that the high magnetic field coils have to be cooled with liquid nitrogen to lower the heating by the very high currents of about 10.000 Amp.

Indeed we can write

TE j i 2 pdt = j Cv v d T c TB p resistivity

C v specific heat

v density

Using still lower temperatures will diminish the resistivity very much. Other investi­gators use also either helium or hydrogen in order to lower still the resistivity.

ACKNOWLEDGEMENTS

We take the opportunity to express our truly thanks to the Belgian Interuniversity Institute of Nuclear Science for the financial support during these measurements.

225

I-24

The Kapitza Conductance of Lead

Conductance Kapitza du plomb

L. J. CHALLIS and J. D. N. CHEEKE Department of Physics, University of Nottingham, U. K.

SOMMA/RE. On sait que la conductance Kapitza du plomb peut varier de pres de 10 fois entre les diff erents echantillons. Il est prouve que ces dijf erences proviennent de la presence de couches d' oxyde et d' etendues variables d' alteration superjicielle. Une etude a done ete faite en vue d'obtenir des resultats reproductibles sur des surfaces propres sans traces. On rend compte de mesures a l' etat normal et a l' etat supraconducteur pour trois echantillons polycristallins purs apres divers traitements superficiels. Les traitements comprenaient le recuit, l' electropolissage et le bombardement par ions d' argon. A l' etat normal, les trois echantillons presentaient a peu pres la meme conductance apres un traitement semblable. On notait un effet semblable a l'etat supraconducteur. Dans les deux etats, la relation avec la temperature etait T3'5 environ. Le rapport des conductances des deux etats ( � 1,5) etait d'environ la moitie de celui indique precedemment (Challis 1962, Gittleman et Bozowski 1962, Barnes et Dillinger 1963).

1. INTRODUCTION

Measurements of the thermal boundary conductance between lead and liquid helium II [1, 2], have shown that the observed value can vary by a factor of nearly ten between different specimens and depends greatly on the condition of a given specimen. The measurements made by Wey-Yen on superconducting lead showed that an increase in conductance was obtained when a machined surface was heavily electropolished and he correlated this increase with the removal of the damaged layer produced by machining. He also observed a large decrease in conductance if the surface was scraped under liquid helium (at which temperature defects produced would be frozen in) but that when the specimen was annealed at room temperature for a week the conductance rose greatly and after four months was equal to that of an electropolished surface. A similar increase has been obtained by etching a normal metal (copper) [3] and a similar decrease on scraping under liquid helium [4]. However, the interpretation here is complicated by the possible existence of mechanisms of heat transfer which would be very sensitive to the removal of oxide on cleaning [5, 6, 7]. No previous experiments of this kind have been reported on normal lead, whose conductance can be three times larger than that of superconducting lead [1, 8, 9].

The conductance in both states of lead is also known to be decreased by the presence of an oxide layer [1, 2.] The strong dependence of the conductance on these para­meters confuses any comparison of different materials and the aim of the present in­vestigation is to develop a method of treatment after which the same conductance is obtained for different specimens of the same bulk purity irrespective of their original physical condition. The present work is concerned with measurements made on lead.

2. SURFACE TREATMENTS AND RESULTS

Two methods of surface treatment have been used (i) Electropolishing using an Ellopol tampon [10] (ii) Argon ion bombardment immediately before the experiment with the specimen positioned in the cryostat. In addition, some of the specimens were annealed in vacuo for 15 hours at 270° C before surface was machined and treated. Full details of these processes will be given in a later publication. Electropolishing could result in the formation of a thin chemically different layer, and oxide growth between the time of polishing the surface and assembling the specimen is possible ; it is hoped

227

I-24

that ion bombardment removes either of these layers and hence produces a surface which is chemically clean or at worst is covered by one or two atomic layers of oxide.

The normal and superconducting conductances for all specimens are shown in Figs. 1 and 2. Specimen 1 was not annealed, and five separate experiments were made on it. The curve labelled 1 A shows the conductance of this specimen after about a day in air following machining and with no surface treatment other than a light rinse in methanol.

5

·3 ·3

/ / /

/ /

1 ·8

2 3

0 2·2 T K

Fig. r . The Kapitza conductance of lead in the normal state as a function of temperature. (The curve labelled r A, B contains points for r A, .A and r B, [;., .)

1 B, an almost identical result, was obtained after the specimen had been left in the cryostat in air for six days and in oxygen for two days. For 1 C the specimen was lightly bombarded, while for 1 D at least 200 microns of lead was removed from the surface by electropolishing and the surface was then heavily bombarded. For 1 E the specimen was heavily electropolished only, and immediately placed in vacuo (,-.,10-5 mm).

The conductance of the specimens 2 and 3 were measured after they had been an­nealed (before machining) lightly electropolished and heavily bombarded. The values in the superconducting state have been corrected for the resistance between the ther­mometer and the surface (the correction is typically of the order of 10 %). For comparison, the results of an earlier experiment [l] in which the specimen had been annealed in vacuo after machining and then lllOunted immediately are shown (Cl). The results for the same specimen after eighteen months in air (C 3) are also shown. (No values were obtained for Cl in the superconducting state.) The value obtained by Wey-Yen for electropolished lead is also included.

228

·S

·25

-2 - I h W cm d•g

, '

1·3 1·4 I· /

/

I-24

IA,B

, W E Y Y E N

,( 3

Fig. 2. The Kapitza conductance of lead in the superconducting state as a function of tempera­ture. (The curves have been corrected as described in the text.) The curve obtained by Wey·Y en for clectropolished lead is also shown.

In the normal state the differences between 1 D, 2 and 3, whose surfaces should have been reasonably free from strain and oxide, are quite small (�40%). It also seems that the small oxide layers which may have formed before the specimen was mounted have not significantly reduced the values of Cl and 1 E. However, the oxide layer present even on surfaces which have been bombarded may be sufficient to inhibit an electron tunnelling process of the sort proposed by Bloch [6]. In the superconducting state the conductances are again high but the variation of the treated specimens is greater. Since

2·5 R

1 ·5

l ·O

0

-I 2 W cm d cg

- - - - - - - - - - - - - - - - - - - - - - - - ·

2 3 4 s 6 7 Koc Fig. 3. The thermal resistance of specimen 3 as a function of magnetic field at r .33 ° K.

229

I-24

the difference in conductance between the two states is about the same as that for sur­faces which have been machined only but are relatively free from oxide, the ratio of

the conductances is less than that previously observed [1]. (For 1 E and 2 �: ,..._, 1 .35,

for 3 �: ,..._, 2.) There is no significant difference between the temperature dependences

( ,..._, T3'5) in the two states after surface treatment. In Fig. 3 the Kapitza resistance is shown as a function of magnetic field parallel

to the surface for specimen 3 at T = 1 .33 ° K. The variation is representative of the fully treated specimens. In the normal state the thermal conductivity of lead decreases steadily with field and a correction to the Kapitza resistance is necessary at the higher fields. There is again a small negative magneto-resistance as in earlier work [1] but the relatively large decrease at fields between He and 1.2 KOe that occurred in the earlier work (specimen C 3) has disappeared.

3. DISCUSSION

The results on the fully treated specimen in the normal state suggest that it is possible to obtain conductance values reasonably representative of a material by surface treat­ments of the sort described above. However, even after the surface has been cleaned by ion bombardment, the background pressure of oxygen (,..._, 10-5) mm is probably sufficient to cause the growth of at least a monolayer of oxide which would inhibit an electron tunelling process such as that suggested by Bloch. In the superconducting state the problem is complicated by the correction. Nevertheless it appears that there is still a difference in the conductance between the two states which suggests that the electrons contribute directly to the energy transfer as suggested by Little [5], Bloch [6] and Andreev [7] and the existence of a magneto-resistance in the normal state supports this conclusion. Since the measurements are made at low reduced temperatures (T/Tc < 0.3) it may be reasonable to assume that hs is due to processes involving the phonons of the metal only, but that hN-h s involves the electrons directly. There is some evidence from the present and earlier measurements [1] that suggests that hN-h s is less affected by the presence of a damaged layer than h s, but more experimental work is required to confirm this. Both hN-h s and h s are affected by oxidation [1].

The reason for the disagreement between our values for electropolished lead in the superconducting state and those of Wey-Yen has not been determined.

Finally it is of interest to note that the tail to the transition observed in a previous specimen (C 3) is missing in the fully treated specimens. It is suggested that this may have. been due to the existence of superconductivity of the second kind in a surface layer of low electron mean free path and there is similar evidence from measurements of cyclotron resonance at 36 Gc/s [1 1]. Since C 3 was annealed and the cyclotron reso­nance was on electropolished specimens it suggests that the reduction in mean free path resulting in the phenomenon is due to chemical impurity (oxide) rather than to strain.

REFERENCES

I . Challis, L. ]. 1962, Proc. Phys. Soc., 80, 759 2. Wey-Yen, K. 1962, J. E. T. P., IS, 635 3. Dransfeld, K. and Wilks, ]. 1957, Proc. 5th Intern. Conf. Low Temp. Phys. and Chem. p . 39 .

Madison, Wisconsin Challis, L. ]., Dransfeld, K. and Wilks, ]. I96I, Proc. Roy. Soc. A, 260, JI

4. Johnson, R. C. and Little, W. A. I963, Phys. Rev. 130, 596 5. Little, W. A. 196I, Phys. Rev. I23, 435 6. Bloch, F. (reported by Little, W. A. 1962, I. B. M. ]. Research and Devel. 6. 3I) 7. Andreev, A. F. I963, J. E. T. P., I6, 1084 8. Gittleman, ]. I., and Bozowski, S. 1962, Phys. Rev. 1 28, 646 9. Barnes, L. ]. and Dillinger, ]. R. I963, Phys. Rev. Letters, ro, 287

IO. 'Ellopol', Model A. P. M., Metallurgical Services, Reliant Works, Betchworth, Surrey n. Young, R. C., 1962, Phil. Mag., 7, 2065

230

1-22

Basic Aspects for Superconducting Electric Machines

Aspects fondamentaux des machines electriques a superconducteurs

Dr. H. YONEMITSU, H. MAEDA and M. OHKA WA Central Research Laboratory, Tokyo Shibaura Electric Co., Ltd., Kawasaki, Japan

SOMMA/RE. Les courants supraconducteurs sont definis par les champs electriques et magnetiques locaux et /'interaction entre les courants est dif.f erente des interactions entre des materiaux conducteurs normaux.

En appliquant les supraconducteurs aux machines electriques, il est necessaire d'etablir la theorie de conception tenant compte de ces effets. Voici quelques- unes des conclusions caracteristiques:

( 1) au lieu de diviser un courant supraconducteur en de nombreuses branches pour conserver le champ magnetique du courant au-dessous du point critique, ii vaut mieux diviser une paire de courant vers l' avant et de courant vers l' arriere en plusieurs plaques paralleles en series alternees.

De cette f afon, les courants voisins /es uns des au tr es et s' ecoulant en sens inverse, agissent /es uns sur les autres pour se soutenir mutuellement et augmenter le champ critique.

(2) Il peut etre possible de realiser un transformateur supraconducteur de construction tres simple en utilisant des plaques supraconductrices para/le/es construites comme dans le cas precedent.

En particulier dans le cas de courant continu, on peut obtenir un trans! ormateur de courant pour produire des courants intenses et des champs magnetiques eleves a basse temperature.

1 . INTRODUCTION

Electric machines using superconductor have recently been used widely to find superconductor alloys with higher critical field He and to investigate them from the material point of view. They are also necessary to find such forms as wires, plates or thin films most suitable for electric machines.

Nb-Zr and Nb3Sn are known as superconductors of high He, and they are usually used at the temperature 4.2° K or below to increase critical current le.

One of the problems about feeding large current into a superconductor electric machine is how to connect the normal metal current lead and the superconductor one. To avoid Joule heating, thick wires of highly conductive copper must be used, but they are also good heat conductors, which causes greater consumption of liquid helium.

Direct current transformer proposed by J. B. Ferran [1] (G. E.) is a solution to the problem.

It was proved that a direct current transformer using superconductor foil is very convenient as an input current transformer of superconductor magnets. This paper treats the theoretical analysis of the direct current transformer.

2. THE INTERACTIONS BETWEEN SUPER-CURRENTS

Superconducting currents are decided by the local electric and magnetic fields follow­ing the London equations, and the interactions between currents are different from those in normal metals.

For quasi stationary conditions, that is, as long as the angular frequency w satisfies the inequality

w � 1/a A ( ,._, 1012 sec _1)

231

1-22

the normal current and the displacement current are negligibly small in comparison with the super-current, and the London-Maxwell equations are reduced as follows :

...... ......

172 j = f32 j 1 ......

curL 1 js = - - h

a . - A Js = a r e

{32 4 :rc/c2 A

c

(1)

(2)

(3)

(4) where e, h, and j are the electric field strength, the magnetic field strength and the current density, respectively, and A is the material constant.

In order to obtain qualitative aspects, we consider two semi-infinite parallel plates made of superconductors. In the xy plane chosen parallel to these plates, x is tak.:n as the current direction andy as the magnetic field direction. z ;:::: d is the supermetal (1), z � -d is the supermetal (2) and -d < z < d corresponds to insulator or vacuum. The solution of eq. (1) is :

j (Z) j u = f31 J1 e -{J, (Z - d) Z 2:: d { = f32 J2 e + {32 (Z + d) Z � - d (5) = 0 -d < z < d

Here J1, J2 are parameters representing the total currents to be determined from the boundary conditions.

According to eq. (2) and the Maxwell equations in insulators, the magnetic field strength is given as

h (Z) -- hy { = (4 :re J1/c) e - fli (Z - d) Z ;::: d = - (4 :re J2/c) e + f32 (Z + d) Z � -d (6) = const. = ho -d < z < d

The continuity of the tangential component of the magnetic fields at the boundaries requires :

then

4 :rl J1/C = ho -4 :rc J2/c = h o

J1 = -J2 .

z = d (7)

z = -d (8)

From these simple conditions, several conclusions are obtained. Some of them are as follows :

(1) In case of the heavy current transfer, it is of little use to divide a super-current only into many wires in order to keep the magnetic fields of the current below the critical field.

Instead, it is better to divide a pair of the forward and the backward currents into many parallel plates (or coaxial cylinders) in such as alternate way as a forward comes next to backward.

In this way, the currents adjoining each other and flowing in the opposite direction, interact to sustain each other, and increase the critical field.

(2) It may be possible to make a superconducting transformer of very simple structure, by using superconducting parallel plates constructed similarly to the preceding case.

Especially in the D. C. case, the current transformer is obtained which can be used to generate heavy currents and strong magnetic field at low temperature.

We are now under experiment to choose the best way of current transformation among various transformer designs by pure Nb sheet. And the preliminary experi­mental data on these effects will be reported.

REFERENCE l. ]. B. Mc Ferran, "The Direct Current Transformer", Advances in Cryogenic Engineering,

Vol. 6, K. D. Timmerhaus, (ed.), Plenum Press, Inc., New York (1961).

232

Studies on Some Sulphide Phosphors in the Temperature Range from 4.2° to 77.4° K

Etude de quelques sulfures de phosphore entre 4,2° et 77,4°K

Dr. PETER THOMA

1-1 1

Laboratorium fUr Technische Physik der Technischen Hochschule, Mtinchen, Germany

SOMMAIRE. Pour etudier les proprietes luminescentes de 20 echantillons de poudre de phosphore de ZnS et de ZnSCdS entre 4,2 et 77,4° K, on a mis au point un cryostat pour /'helium liquide, qui consistait en deux vases Dewar concentriques pour l'helium liquide et pour le stockage de l' azote liquide pour la protection contre le rayonnement et le prerefroidissement. On utilisait un appareil de chauff age reg le pour obtenir une vitesse de chauff age arbitraire et constante pour la mesure de la phosphorescence bleu-vert et infra-rouge en f onction de la temperature ( courbes d'incandescence des echantillons). On mesurait la temperature avec des resistors a Constante etalonnes (du type Allen-Bradley) et avec des thermocouples avec de I' azote jluide pur a milieu de reference. Etude des avantages et inconvenients des resistors a carbone comme thermometres aux temperatures ci-dessus. Taus les echantillons etaient excites par le rayonnement ultra-violet (0,365 µ). On a mesure la luminescence avec des photomulti­plicateurs en relation avec des combinaisons de filtres pour obtenir une mesure approchee de !'emission phosphorescente. En general, les couches d'incandescence mesurees presentaient 3 pointes. La pointe d'incandescence dominant se trouve au voisinage de 50° K a une vitesse de chauffage de 3°/mn. Le piege a electrons correspondant a une profondeur de 0,05 eV au­dessous de la bande de conduction du modele a bandes d'energie. Ce piege a ere mis en relation avec des interstices de Zn++. Les ions interstitiels de Zn+ + existent dans taus les echantillons. Pour certains echantillons il existe un autre piege a une profondeur de 0,026 eV au-dessous de la bande de conduction, correspondant a l' emplacement de 25 ° K sur l' axe des temperatures a une vitesse de chauff age de 3 ° /mn. Ce piege peut probablement etre explique par la double occupation d'un emplacement interstitiel dans le reseau ZnS par deux ions de Zn+ +. Pour les deux echantillons ii existe une pointe d'incandescence au voisinage de 4,2° K, correspondant a une profondeur de 0,0056 eV au-dessous de la bande de conduction. Ce/a ne peut s'expliquer par un piege physiquement reel mais probablement par une transformation de I' exciton. On a aussi mesure la stimulation infra-rouge et l' on a etudie la reponse des pieges au rayonnement infra-rouge.

1) THE APPARATUS AND THE MEASURING DEVICES

For the investigation of the luminescent properties of 20 powder samples of ZnS- and ZnSCdS-Phophors in the temperature range from 4.2° to 77.4° K a cryostat was used, which consisted of two concentric dewars for liquid helium and for storage of liquid nitrogen for radiation shielding and precooling. A controlled heating device was used to achieve an arbitrary and constant heating velocity for the measurement of the green, blue and infrared phosphorescence as a function of temperature (glow curves) of the samples. The temperature measurements were made with calibrated carbon resistors and with thermocouples with pure liquid nitrogen as a reference medium. All samples were excited by UV-radiation (0.365 µm). Luminescence measurements were made by photomultipliers in connection with filter sets for the separation of emission and ab­sorption bands.

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2) MEASUREMENTS AND RES UL TS

After a 15 min. UV-excitation at 4.2°K the glow curves were measured with two multipliers simultaneously (for visible and infrared emission) with a heating rate of 3 deg/min.

Sample 1 : ZnS(Cu) with 10-4 w% Cu. with ZnO concentration less than 0.1 %. It is prepared with chloridic melting agent. There is a dominating glow peak at 52°K and a weak peak at 32° K in the green range. There is no infrared emission. See Fig. 1 .

Fig. r . Glow curve o f sample 1 (green phosphorescence).

Sample 2 : ZnS(Cu) with 10-4 w% Cu. It is very similar to sample 1, but it is pre­pared completely free of oxygen. Here an infrared phosphorescent emission is existing with a glow peak near 40° K. The glow peak of the green thermoluminescence is near 60° K. See Fig. 2.

Y .-10 z.. o !tO Y o 50 01< Fig. 2 . Glow curves of sample 2 (green and infrared thermoluminescence).

Sample 3: ZnS(Cu) with 10-4 w% Cu. Prepared by a 1 hour heat treatment at 1000°C in a HCl-gas-stream under oxygen free conditions. No solid chloridic melting agent was used. The dominant glow peak of green and infrared thermoluminescence is located near 27° K here. See Fig. 3.

,, .

60 'O -,o Tf- o J(

o,S

" Ao L<> �o '-to � 'o ?o H 0 /,( Fig. 3. Glow curves of sample 3 (green and infrared thermoluminescence).

234

I-1 1

Sample 4 : ZnS(Cu) with 10-4 m % Cu. The same preparation method as for sample 3 was used with the exception that the heat treatment temperature was 1 100° C and that the sample lattice is not free of oxygen. There is no infrared phosphorescent emission but only a green phosphorescence with glow peaks at 45.5°K (dominant peak) and at 4.2° K. The glow curve was measured beginning at 2.5° K here. See Fig. 4.

Fig. 4. Glow curve of sample 4 (infrared).

Sample 5 : ZnS(Cu) with 10-4 m % Cu. The sample is prepared by the same method as used for sample 4, but under a completely oxygen free atmosphere. There is a green and an infrared fluorescence during the UV-excitation. After switching off the excitation the green fluoroscence decays very rapidly (10-6 sec) and only a relatively strong infra­red phosphorescence remains. Only a infrared glow curve with a strong peak at 49°K can be measured but no green thermoluminescence is existing.

It is essential that the samples 3, 4 and 5 are very similiar and that the glow curves above 77° K are nearly identical. See Fig. 5.

Fig. 5. Glow curve of sample 5 (infrared).

Sample 6 : This is a pure ZnS-phosphor, prepared without any doping. It is activator free (self activated) and it has a blue thermoluminescence with a glow peak at 50° K. The fact of the existence of the traps corresponding to the glow peak near 50° K also in an activator free sample shows that these traps are not connected with the activator itself. They must be connected with a defect of the ZnS-lattice alone. See Fig. 6. :"] x ' " "' " " ' " I

4 ..(O 'ZO "lO

Fig. 6. Glow curve of sample 6 (blue thermoluminescence).

Sample 7 : ZnS(Cu, Ga) with 10-4 m % Cu and 10-4 m % Ga. The sample is pre­pared by the conventional method under an oxygen free atmosphere. The thermo­luminescence shows a green and a strong infrared emission with glow peaks near 50° K. See Fig. 7.

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o,s

4 A 0 "2.0 "';) O 40 >e> 'o 10 H 0 J.<

Fig. 7. Glow curve of sample 7 (green and infrared emission).

With this phosphor stimulation experiments in the near infrared region have been performed. The sample was first excitated at 4.2° K with UV light (0.365 µm) for 10 min. and the temperature was kept constant. After an arbitrary long time an infrared light beam was directed onto the sample and the resulting stimulation pulses were measured with an oscilloscope. The maximum intensity, which was generated by the tungsten lamp, had a value of 170 µW/cm2 (measured on the phosphor surface). As a source of minimal intensity a glowing cigarette (in complete darkness without filters) was used. In all cases stimulation pulses were observed, which rose rapidly (0.2 µ sec) and decayed slowly. See Fig. 8.

I

Fig. 8. Stimulation pulse of sample 7 after irradiation of an infrared light beam.

The maximum amplitudes of the pulses were nearly proportional to the input inten­sity of IR-light.

If this phosphor, which consists of relatively large crystals, is completely de­excitated by heating and than cooled down to 4.2°K an apparently stimulated fluores­cent emission has been observed, if the sample is excitated by IR light with a wavelength of 0.9 ± 0. 1 mµ by the tungsten lamp in connection with the filters. The wavelength of this emission, the lines of which could not be measured, lies in the region of the in­coming IR light. Probably this effect is a resonance fluorescence phenomenon and it does not exist at 77° K or room temperature.

Sample 8, 9, 10 and 1 1 : These samples are ZnCdS phosphors with various CdS concentrations, activated with Ag and in one case with Cu (sample 8).

236

Sample 8: ZnCdS (Cu) with 30% CdS

Sample 9 : ZnCdS (Ag) with 20% CdS

Sample 10: ZnCdS (Ag) with 40% CdS

Sample 1 1 : ZnCdS (Ag) with 60% CdS

1-11

They are industrial products, prepared by the conventional method with solid chlo­ridic melting agent. There is no infrared emission. The glow peaks near 50° K are broader than the corresponding peaks of the other samples. See Fig. 9.

� o so 60

I

Fig. 9. Glow curves of samples 8, 9, ro und r r .

3) DISCUSSION O F THE RESULTS

Supposing a heating rate of 3 deg/ min the results can be summarized as follows :

The existence of the dominant glow peak near 50° K is independent of the spectral composition of the emitted phosphorescent radiation. It is indifferent if the phosphor is a pure ZnS or a ZnCdS-sample. There is no connection between the traps correspond­ing to the 50° K peak and the kind of activator or coactivator because this glow maximum exists also in selfactivated phosphors.

At some samples there is another weaker glow peak near 25° K. With the exception of sample 3 this peak is located in the rising region to the dominant glow peak near 50°K. Probably the traps corresponding to this peak are associated with the traps cor­responding to the dominant glow peak near 50° K. The emptying of twofold occupied traps is a possible explanation At one sample there is a glow peak at 4.2°K. This peak is probably not caused by a real trap but by an exciton transformation for which Samel­son and Lempicki have given an energy-value of 0.008 eV.

Some samples have also an infrared thermoluminescence. The most of these "infrared" glow peaks have nearly the same position on the temperature axis as the green glow peaks. This fact indicated, that the process of emptying of traps with succeeding green and infrared recombination radiation is a multiple stage process, where the green- and infrared glow emission is associated. An infrared emission exists only in copper activated phosphors and not in silver activated ones.

The essential condition for the infrared thermoluminescence is the preparation of the phosphor under a completely oxygen free atmosphere.

The energetic depth EH of the trap levels in the energy band model can be determined by the approximated expression of Randall and Wilkins EH = 25 kT max or by the method of variation of the heating velocity given by Schon. If two different heating velocities q1 and q2 (q2 > q1) are used, then the glow peak has two different temperature values T1 and T2•

For a monomolecular reaction-characteristic (most samples have a monomolecular reaction characteristic) the relation for the trap depth is:

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By this method the depth values of the traps corresponding to the 3 glow peaks below the conduction band have been calculated after some experiments with various heating rates. The values are :

Glow peak near S0° K: EH = 0.048 to O.OS eV

Glow peak near 2S° K : EH = 0.018 to 0.026 eV

Glow peak near 4.2°K : EH = 0.0048 eV

Corresponding to the Franck Condon principle a radiation with a wavelength equi­valent to these energies is not sufficient for emptying the traps.

The concentration of trapped electrons have been calculated by a method developed by Kallmann.

The results are :

for Cu-activated phosphors

for Ag-activated phosphors valid at 4.2°K.

4. THE NATURE OF THE TRAPS

13.8 X 1017 cm-3

1.2 x 1017 cm-a

The traps in the lowest temperature range cannot be explained by the activator or coactivator, because they are existing also in activator free phosphors. Often experiments have shown, that the defects which are responsible for these traps must have a rather high mobility. The traps at S0° K vanish, if further copper is diffused into a copper­activated green fluorescing phosphor and so a blue fluorescing phosphor is made.

For charge compensation the Zn+ +interstitial ions return to lattice places and copper interstitial ions are associated to Cu+-lattice ions to form the blue emission center. So the concentration of the Zn + + interstitial ions must be reduced intensively and the glow peak vanishes. In this way it can be shown, that Zn+ +-interstitial ions are the physical cause of the S0° K glow peaks.

238

III. Commission 2

Transfer of heat. Thermal properties of materials. Instrumentation. Insulating materials.

Transmission de chaleur. Proprietes thermiques des materiaux. Instruments de mesures. Materiaux isolants.

SESSIONS :

Problems of Insulation

Thermodynamics

Heat Transfer

Problemes d'isolation

Thermodynamique

Transmission de chaleur

Problems of I n su l at i o n Problemes d ' i s o l at i o n

II-2

Heat Transfer by Natural Convection in Porous Insulants

Transmission de chaleur par convection naturelle dans les isolants poreux

G. MARTIN and G. G. HASELDEN Department of Chemical Engineering, University of Leeds, Leeds, England

SOMMAIRE. On decrit un appareil dans lequel on peut mesurer la conductibilite thermique renforcee d'un isolant poreux, provenant des courants de convection naturelle qui s'y produisent. L'influence de la convection est augmentee par !'elevation de la pression. On indique des re­sultats pour de la laine de laitier tassee a une densite de 145 kg/m3 a une temperature moyenne de -81° C et a 3 pressions.

On decrit une methode de prevision de /'influence de la convection. Une section a travers !'isolation est divisee en un reseau dans chapue carre duquel !es bi/ans thermiques et massignes sont resolus successivement par une calcultatrice digitalejusqu'a ce qu'on obtienne le point de convergence. L'influence de la convection determinee experimentalement est plus elevee que celle prevue, ce qui montre qu'il peut exister des courants multiples.

INTRODUCTION

In most cases in which a temperature gradient exists in a granular or fibrous insulating material there will be a tendency for bulk convection currents to occur within the insulant leading to enhanced heat transmission. Whether the effect is significant will depend on the porosity of the insulant and size factors. The presence of these currents has been demon­strated by a number of workers [I, 2, 3] . Lorentzen and Brendeng [3] have shown that the effect of convection may be appreciable for the wall of a cold store when insulated with a loose-fill material. It has also been claimed that these effects may be significant within the "cold-boxes" of large low-temperature plants, in the insulation of large storage tanks for liquefied gases, and in the insulation applied within the shells of gas-cooled nuclear reactors. The authors are unaware of any attempt either to measure or predict these effects with even moderate precision.

THE NATURE OF THE ENHANCED HEAT TRANSMISSION PROCESS

It is possible to postulate two ways by which additional heat transmission could result from convection currents. Firstly there is the direct transport of heat in the moving gas, secondly there could be an increase of the conductivity of the insulant itself arising from the presence of micro-turbulence within individual pores. Preliminary tests were made in which air was caused to flow through different insulating materials which had a tem­perature gradient across them. When correction was made for the heat carried away by the flowing air it was found that the effective thermal conductivity was independent of the air flow-rate ; thus the second mechanism, described above, was discounted.

EXPERIMENTAL APPARATUS FOR CONVECTION STUDIES.

To separate the effect of convection from the other variables it was decided to deter­mine the influence of pressure on the thermal conductivity of a wall of porous insulant. The thermal conductivity of permanent gases is almost independent of pressure at mode­rate pressures (except if mean-free-path effects are present) but convection effects are pressure-dependent.

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The test cell comprised a vertical, electrically heated, aluminium plate 52 in. high x 24 in. wide of which the central test section was 48 in. high x 12 in. wide. The test section was separated by a 0.125 in. gap from the surrounding guard section, and was heated by three separate resistance windings attached to the upper, middle and lower areas of it. The complete hot plate was backed by a 2 in. thick slab of foamed glass, to the remote surface of which was attached a further heated plate maintained at the same temperature. Stray heat losses from the test section were thus minimised.

The hot plate was maintained at a spacing of 1 ft. from the cold plate by an open box structure of 2 in. thick foamed-glass slabs extending around the edge of the plates. The cold plate was made of copper and had vertical copper tubes soldered to its rear surface at a spacing of 3 in. from each other. These tubes were joined into a common manifold at the bottom and terminated at their upper ends in a header-tank. Provision was made to circulate either water or liquid oxygen in the tubes.

The test cell was held together by stainless steel tension wires and was supported by 4 wires of the same material within a mild-steel pressure shell allowing pressures up to 10 atmospheres.

Fig. r. Test cell in position inside pressure vessel

The photograph in Fig. 1 shows the empty test cell within the pressure shell. The heated plate, backed by a guard plate is in the foreground, and the cold plate, cooled by tubes leading to the header tank, is behind it. The test cell was carefully packed with mineral wool, tamping with a light but uniform pressure, to a density of 9.1 lb/ft3• The top of the cell was sealed with a lid made from 2 in. thick foamed glass. The pressure shell was then filled with expanded pearlite insulation up to the level of the flange, thus submerging the test cell. Sufficient mineral wool insulation was then placed above the pearlite to fill the dished-head when this was bolted in place.

The main guard ring was provided with eight independently controlled heaters dis­posed on different areas of its surface, and there were sufficient thermocouples attached to it, and on the test section to ensure that the whole surface of the heated plate was at a uni­form temperature.

The apparatus was designed primarily for operation at low temperatures. The guard ring provision was inadequate when the heated plate was operated much above room temperature. Tests were made with air pressures both below and above atmospheric. When convection was present it was necessary to readjust the heating currents so that the lower section of the heated plate received much more heat than the upper sections.

The results of two test runs are shown in Fig. 2. A separate apparatus was set up to determine the permeability of mineral wool at

various densities. It was found that at 9.1 lb/ft3 the permeability was 2.5 ft3/hr.

THEORETICAL CALCULATION OF HEAT TRANSFER BY NATURAL CONVECTION.

A complete mathematical solution of the problem is difficult because of the large num­ber of variables to be considered. At any particular point in the insulation the variables can

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be reduced to gas and solid temperatures, gas pressure, velocity and direction of flow. If it is considered that the gas and solid at any point are at essentially the same temperature (this may be justified, since the rate of flow of gas will be small) the problem is very simi­lar to that of the free convection of fluid between parallel plates at different temperatures, except that a resistance to gas flow is introduced due to the presence of the insulant. Attempts have been made to solve the problem of the free convection of fluids mathemati­cally [ 4, 5] but gross approximations have been necessary in order to arrive at a solution. For example, in the work of Batchelor [ 4] the assumption was made that the temperature differences were small compared with the absolute temperature. This may be justified in the case of small temperature differences near room temperature but is not justifiable in the case of temperature differences between, say, liquid methane and athmospheric temperature.

A study of the mechanism of convective heat transfer shows, however, how an approxi­mate finite difference method could be used to find a solution.

A I!.

Fig. 3. Distribution of iso­therms-conduction.

Fig. 4. Distribution of iso­bars.

Fig. 5. Distortion of isotherms due to convection.

Considering again the case of heat transfer between vertical parallel plates at diffe­rent temperatures separated by insulation and supposing that initially there is no con­vection occurring, heat transfer will take place entirely by conduction (the insulation is assumed opaque to radiation). The temperature distribution will then be as shown in Fig. 3 with the isotherms being vertical and parallel (equally spaced if conductivity is assumed constant with temperature). Assuming (to maintain symmetry) that the pressure across AB (mid-way down the insulation) is constant, the pressure at other points may be

243

I!-2

calculated from the gas density and distance above or below AB. Fig. 4 shows the distri­bution of the isobars so calculated. It may be seen that there are now pressure gradients in the horizontal direction which will cause the gas to flow, carrying heat and distorting the isotherms to the form shown in Fig. 5.

If the insulation is considered to be divided into a number of small elements, the pressure and temperature in each element may be calculated, as a first approximation, to be as in Figs. 3 and 4.

The mass flow into each element from the surrounding elements may then be calculated from the pressure gradients, and by carrying out a mass balance on the element a new approximation to its pressure may be found. Similarly a heat balance, including the heat flow by conduction and the heat carried as sensible heat by the gas, enables a new appro­ximation to the temperature to be found.

Repeating the calculation several times will produce a reasonably accurate approxima­tion to the final temperature and pressure distribution.

The rate of heat transfer may then be calculated by summing the amounts of heat transferred both by conduction and by gas flow for all elements along any vertical plane through the lagging.

The calculation was programmed for a Pegasus Computer, and the calculation over the network was continued until it converged in a stable solution. Calculated results for the conditions corresponding to those of the two reported experiments are shown in Fig. 2.

DISCUSSION OF RESULTS

Considering the results of the low temperature run shown in Fig. 2 it is seen that the experimentally determined pressure dependence of thermal conductivity of mineral wool is greater than that predicted theoretically. The simplifying assumption made in the cal­culation, such as the neglecting of the temperature difference between the fibres and gas at any point in the insulant, would have led to the reverse effect.

The most probable explanation appears to be that multiple convection currents are occurring. The present calculation method shows that the effect of convection is greatest for a height to thickness ratio of I, and therefore with the present test cell, having a height of 4 ft. and a thickness of I ft., the maximum heat transmission would occur if 4 convection pockets became established within the insulation.

Further experiments are now being made to determine more accurately the effect of convection for a cell in which the h/x ratio is I .

REFERENCES l. E. A. Allcut, "An analysis of heat transfer through thermal insulating materials". General

Discussion on Heat Transfer, Inst. Mech. Eng. and A.S.M.E. Sept. 195r . 2. E. A. Al/cut and Ewens. School of Eng. Res., Univ. of Toronto, Bull. No . 149 (1937) and No . 158

(1939). 3. G. Lorentzen and E. Brendeng, Proc. roth Int. Congress of Refrig. Copenhagen 1959, l, 294 . . . + G. K. Batchelor, Quarterly App. Maths r2, (No. 3) 209-33 (1954). 5 . R. A. Wooding, ]our. Fluid Mechanics 1956.

DISCUSSION

C. Hocking, Sweden, asked: 1 . The data presented in the paper applied to a density of 9 lbs/cu. ft. In refrigeration practice one of ten uses only 3 lbs/cu. ft. What effect does change in density have ?

2. What proportion of heat conductivity is due to convection at normal temperatures and pressures ?

M. H. Myncke, Belgium, asked : How do you explain the maximum in the curve giving the apparent thermal conductivity as a function of the thickness ?

G. Walker, U. K., asked : Is it possible to give any actual values to the increase in heat transfer through the insulation in the case of very large storage vessels for liquid oxygen or liquid methane ?

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G. Laing, U. K., asked : Do you consider that the maximum effect of convection currents at a 1/1 ratio of thickness to height is generally valid ? It has been reported that on curtain­wall insulation for buildings up to a height of 80'-0", considerable improvement was made by introducing horizontal spacers. It would appear from this paper that this modi­fication had no theoretical basis and I would like the authors' comments.

G. G. Haselden, U. K., replied to the above : The contribution to heat transfer arising from natural convection is determined by the permeability of the insulating material, and the composition and density of the gas content within it.

In general, the use of insulants with a lower density than the one used in our experiments, will result in a larger contribution of convection due to higher permeability.

The proportion of the total heat transfer arising from bulk convection is influenced also, as shown in the paper, by the geometry of the system. For the insulant which we tested, and at normal temperature and pressure, the additional convection contribution was almost negligible for a 1 ft. cube, but became about 30 % of the whole for a 3 ft. cube.

The cube shape gives maximum convection heat transfer due to the interaction of two factors. For a wall of insulation of a given thickness, increasing the height to a value in excess of its thickness will lead to a slight increase in the rate of circulation of gas, since the ratio of the gas path length to the pressure driving force will increase. However, as the height of the wall increases and the rate of circulation increases there is still only one circulation current conveying heat from the warm surface to the cold surface. If the same wall of insulation is divided by horizontal partitions into a number of cubes the rate of circulation within each cube will be somewhat less than for the overall situation, but there will be many more circulating currents to contribute to the heat transfer.

This theory is based on the assumption that the insulating material fills all the available space. If in a large structure contraction causes voids to open up, then the use of horizontal partitions may be desirable.

For the 60 ft. diameter liquid methane tanks at Canvey Island it appears possible that the extra contribution of convection within the powder insulation may be of the order of 25 %. R. W. Powell ( U. K.), (Comment) : Regarding the suggested use of cubical specimens, I would like to draw attention to a method that has been used at the National Physical Laboratory for thermal conductivity measurements on thermal insulating materi­als. The latter, in the· form of a thick walled hollow cube is built around a cubical con­tainer of liquid nitrogen. The heat flow to the liquid nitrogen is derived from obser­vation of the loss of weight with time.

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11-4

Investigation of the Influence of Free Thermal Convection on Heat Transfer through Granular Material

Etude de !'influence de la convection thermique libre sur la transmission de cha­leur par des materiaux granules

K.-J. SCHNEIDER Technische Hochschule Karlsruhe, Germany

SOMMAIRE. L'infiuence de la convection libre sur la transmission de chaleur par des materiaux granutes a un inter et particulier pour !'isolation frigorifique. II n'y a pas encore eu beaucoup d'etudes sur ce sujet. La plupart d'entre el/es donnent des resultats pour des isolants tres speciaux.

Les experiences decrites dans ce rapport ont ete entreprises pour verifier le rapport entre les phinomenes de convection et des variables particulieres. Les echantillons utilises dans les experiences etaient des couches des materiaux granutes deforme sphirique. Dans les experiences, on f aisait varier !es f acteurs suivants : direction de l' ecoulement de chaleur, grosseur et sub­stance du grain, le substance dans les interstices difference de temperature et temperature moyenne.

De meme que pour la convection dans !es fluides purs, on peut presenter les resultats en f onction du nombre de Grashof, Gr, et du nombre de Prandtl, Pr, qui sont cependant un peu modifies pour les materiaux granulis. Les resultats indiquent que le debut de la convection et la convection bien etablie ne dependent pas des memes f acteurs.

En s'appuyant sur ces resultats, on peut evaluer l'ordre de grandeur de !'influence de la convection libre sur la transmission de chaleur a travers les couches d'un materiel granuli quelconque.

Nomenclature.

The following nomenclature is used in this paper : C constant, dimensionless ; CK for Kozeny constant c specific heat of the void fluid, J/kg °C ; cp for the specific heat of the void fluid at

constant pressure d diameter, m ; da for outer diameter, dh for hydraulic void diameter, d1 for inner dia­

meter, dk for diameter of spheres and for the equivalent diameter of nonspherical grains

f fraction void, dimensionless Gr Grashof number, dimensionless ; Grk for free convection in granular material g acceleration due to gravity, 9,81 m/sec2 h height of a vertical layer, m k fluid-flow permeability, m2 l length, m Nu Nusselt number, dimensionless ; Nuc for free convection in enclosed spaces 0 surface, m2 ; 0 k for grain surface, 0 z for void surface Pr Prandtl number, dimensionless ; Prw = 'f/Cp/ ). 0, Pr z = 'f/Cp/ Jc z p pressure, N/m2 Ra Rayleigh number, dimensionless Re Reynolds number, dimensionless ; ReB for flow in granular material t temperature, °C ; tm for mean temperature V volume, m3 ; V k for grain volume, V z for void volume w macroscopic velocity, m/sec ; wx for velocity in the x-direction x horizontal coordinate axis, m

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f3 coefficient of volumetric expansion for the void fluid, 1 / °C LI t temperature difference, ° C (J thickness of a layer, m

'fJ dynamic viscosity of the void fluid, kg/m sec ). thermal conductivity, W/m ° C ; A err for effective thermal conductivity of granular

material, Ac for contribution of free convection to A err, A a for contribution of con­duction to A err, As for contribution of radiation to A err, A z for thermal conductivity of the void fluid

v kinematic viscosity of the void fluid, m2/sec

!! density of the void fluid, kg/m3

INTRODUCTION

For many areas of technology the conduction of heat through loose materials is of significance. Such materials are granular materials, fibrous materials and powders. Heat transfer through these materials can be influenced by free convection, i. e. thereby, that the void fluid comes into circulation as a result of density differences caused by temperatu­re. This can be the case in materials for insulation in refrigeration, since the properties of the gas contained in the insulating material permit the development of convection espe­cially at low temperatures [8, 1 1, 12]. If the voids of the material are filled with a liquid convection occurs more easily at high temperatures. This can be of interest in geophysical problems [7, 10, 14]. The studies published on the effect of convection are not numerous. Sometimes it is mentioned as an undesired side effect during the measurement of thermal conductivity [2, 6] ; sometimes the influence of convection in certain insulation materials is examined [8, 1 1, 12] ; several papers present theoretical and experimental studies mainly about the point of the onset of convection [7, 10, 14].

EXPERIMENTAL EQUIPMENT AND TEST PROGRAMME.

Experiments were conducted to determine the dependence of free convection in granular materials on the different variables. A plate apparatus, which essentially corres­ponds to the American and German standard specifications, served as the experimental equipment [1, 5]. Fig. 1 shows the schematic diagram of the experimental equipment. The effective thermal conductivity A err results from the measurements. It is composed of contributions by heat conduction, radiation and free convection :

A err = A o + As + Ac (1)

By vertical downward directed heat flow, no convection can occur since the layers are stable. The convection part therefore can be determined as the difference between two measurements in different directions of heat flow. The radiation part is only important at

Fig. L

248

2 2

A � Schematic diagram of the plate-apparatus. 1 Heating plate, 2 Cooling plate, 3 Sample layer, 4 Guard heating plate, 5 Guard heating ring, 6 Guard layer, 7 Frame, 8 Insulation, 9 Housing, 10 Thermocouples, I I Thermostats for hot and cold water.

11-4

Table 1 .

grain grain dimensions frac- permeability k void fluid material shape m ti on m2

void/

glass spheres dk = 1,1 · 10-3 0,385 0,93 . 10-9 distilled water

" " 2,3 • 10-3 0,387 4,17 . 10-9 " "

" " 4,0 · 10-3 0,399 13,8 10-9 " "

" " 7,1 · 10-3 0,409 43,5 10-9 " "

" " 9,9 • 10-3 0,434 99 10-9 "

" rings = 4,5 · 10-3 0,514 15,1 10-9 " " da = 4,1 • 10-3

di = 2,0 · 10-3

" spheres dk = 1,1 · 10-3 0,390 0,99 10-9 turpentine oil " " 9,9 · 10-3 0,426 92 10-9 " "

steel " 4,8 • lo-a 0,393 17,8 10-9 " "

" " 15,1 • 10-3 0,390 126 10-9 " " glass " 9,9 · 10-3 0,435 100 10-9 air steel " 15,1 · 10-3 0,417 165 10-9 air

higher temperatures. For gaseous void fluids it can approximately be determined by ex­periment or calculation ; for liquids in the voids it is practically zero. Table 1 is a summary of the granular materials examined. Fibrous and other materials, the properties of which are partly dependent on direction, were not studied. The experiments were conducted on horizontal layers with vertical upward and downward heat flow as well as on vertical layers with horizontal heat flow. The temperature difference L1 t and the mean temperature t m were varied.

TEST RES UL TS

Fig. 2 shows the determined values of the effective thermal conductivity of plain horizontal layers with upward directed heat flow. The dependence on the temperature

�-

•

'

0 0

�

I j I I / / 'j/ �

"60· 1 y J/ I •

I /

/ 20.5> iJ

20"-· /' � . � 17-v--

/ _ __. ..... / •

'°

/ ;

/ S--

---

--- 0---,,/ � __.a-

---

• _,, -

20 JO

I 6 g l a s s s ph e r e s 1 mm/water a " " 2 " " ,_ 0 " " ' " " " " " 7 " " <> " " 1 0 " " '1 " r i ng s ' " " ... " spheres 1 " / t u r pe n t .,_ ,, " " 10 " " • s te e l " 5 " " • " " 1 5 " I-+ g l a s s " 1 0 .. /a i r x s t eel " 1 5 " ..

--�- -i

--' I

I - - -- ---- - - - t----- ---

'

--+-----,_Id=- )(--

"' so 60

Fig. 2. Variation of the effective thermal conductivity Aerr of granular materials with temperature difference and mean temperature. Horizontal layers with heat flow upwards. Thickness of the layers 15 = 40 mm. Mean temperature Im ""' 40°(' unless otherwise indicated.

249

ll-4

difference and the mean temperature is to be attributed mainly to the influence of free convection. For vertical layers a similar diagram results. As for layers of pure fluids, the occurance of convection in layers of granular material can be represented by the use of the dimensionless groups Nusselt number Nu, Grashof number Gr and Prandtl number Pr. The product Gr · Pr is also denoted as the Rayleigh number Ra. The Nusselt number for free convection in enclosed spaces [9] is given by

A o + Ac Nuc = -- ---­A o According to Wooding [14] free convection in granular materials is described by

{J gLJt {jk 1) C p Grk · Prw = -----(11/ 12)2 A o

The permeability k is defined by the equation

k ap Wx = - --1) ax

6 I I 4 3 � 6 g l a s s s ph e r e s 1 mm/woter

a " " 2 " " 0 " " 4 " "

2 '--- 'l " " 7 " " <> " " 1 0 " "

" l

ii " r i ng s 4 " " "' " sp heres 1 · · / t u r p e n t . o'>----- ,, " " 10 .. "

8 � • s t e e l " 5 " " • .. " 1 5 ..

6 � + g l a s s .. 1 0 .. / a i r x s t e el .. 1 5 .. .. v

4 i I ,-3 _ Q

I I •o -2 I . ! i I • . �� ! i I i I I i I'' i ,,,.

.. v ' �

'" ' -v . 0

. . ••

.. �·

I I I

(2)

(3)

(4)

---

ID° 2 3 4 6 8 101 2 3 4 6 8 1o2 2 3 4 6 8 10 J 2 3

""' A g l a s s "' 6 >-- a � 0 �""' ' � 6 -< 3 � ii "'

2 '-- : s te e l •

s pn e r e s

r i ng s spheres

1mm/water 2 " 4 .. 7 ..

10 .. 4 .. 1 " / t u r pe n t .

.. 1 0 .. �-l--l--+--1----��·+.-<-�-1---+-+----l----

5 .. 1 5 ..

+ g l a s s 1 0 " /a i r , 0v• 701·1--�x�·�t�•·�·�l--....:!�-..U.5�"-..::...-..---1--J-�J>'..-+--l---l---+--+--t--+-+---t---I

8 1-----l--+--t--l--+-+---l--,�o�.'+----11---•,-t----t--t--+---J-t-t-----J--j 6'-----l.--l--1---l---+-l------l-<>l--+--l---+-l------+-+--l--t--t--t-----t---\

2 3 6 102 2 3 4 6 8 103 2 3 4 6 8 10" 2 3

Fig. 3 a and b. Heat transfer by free convection in granular materials. Horizontal layers with heat flow upwards. Thickness of the layers o = 40 mm.

250

II-4

This equation is valid for laminar flow only. In Fig. 3a Nuc is plotted against Grk · Prw for horizontal layers. It can be seen that convection for all the layers examined in this range begins at Gr k • Prw ""' 35. At higher values however the curves for different layers diverge. This can be explained by the fact that by increasing convection, the influence of the ther­mal conductivity A o of the layer decreases, while that of the thermal conductivity A z of the void fluid increases. If one calculates the Nusselt and Prandtl numbers with A z instead of with A o and plots (A z + Ac)/ A z against Gr k • Pr z agreement among all layers examined actually results for fully developed convection, while no agreement is found in the vi­cinity of convection onset. Fig. 3 b shows the corresponding diagram for horizontal layers. The examination of vertical layers gives very similar results, as Fig. 4a and b show. The reference temperature for the material properties for calculation of the Grashof and

Prandtl numbers is the mean temperature tm of the layer.

�-

�---

10 8 6

'r----�

--·-

3 2

6 g l a s s a " 0 " v "

() "

'" " .a. " ,. "

• s t e e l • "

+ g l a s s x s t e el

I s ph e r e s

" " " "

r i ng s spheres

" " " " "

I I

I I

1 mm/wa ter 2 .. " 4 .. " I 7 " " ! 1 0 " " I 4 " "

1 " / t u r p e n t . 1 0 " "

5 " "

1 5 " " 1 O " / a i r 1 5 " "

• I.;> o• 0

�'" I

->t- + ID° 10° I� ,_ ' ... ... � �"li It. : k

2 3 4 6 8 101

102 8

� 6 6 g l a s s s p h e r e s - a � 0 .< v . ' �

" () -:::. 3 � '" r i ng s .a. sp h e r e s

2 � ,. • s t e e l • + Q la S S

10' x d , .1

2 3 4 6 8 102

1 mm/woter 2 " 4 " 7 "

1 0 " 4 " 1 " / t u r pe n t .

1 0 " 5 "

1 5 " 1 O " / a i r 1 5 ..

I I

-.

8 .� 0

6

3 2

'

100 Lo---i.cL.fL .6 101 2 3 "'o •

4

'o ' <r

6 8 702

i ' "'

o"° i d' 'b",

4J �u· i .

I I

2 3 ' 6 8 103

-

!

'

v � 0 & _ 0vo • '

, o o '

' . .

. . ..

.

I 2 3 4 6 8 10]

I �-

' ' '

I

.. . - ' � ,. • ' •

. ' '.

.

2 3 4 6 8 10' 2 Grk· Prz 3

Fig. 4a and b. Heat transfer by free convection in granular materials. Vertical layers with hori­zontal heat flow. Thickness of the layers .5 = 40 mm, height h = 300 mm.

251

11-4

As soon as the convective current starts to become turbulent, the permeability k is no longer constant and a calculation of the Grashof number is no longer meaningful. The onset ofturbulence is to be expected at a Reynolds number ofReB ""' 10 [3] . Re B is given by

ReB = 0,9 (5)

with

(6)

Since the definition of the permeability is based only on a macroscopic consideration of the flow, the onset of turbulence for various layers can occur at different values of the Grashof number. A rough calculation of ReB from the quantity of heat transferred by convection shows, that laminar flow existed for all measurements conducted.

DETERMINATION OF PERMEABILITY

The permeability of granular materials is dependent on the shape, size and arrange­ment of the individual grains :

(7)

The constant C, which is influenced by grain shape and arrangement, can be deter­mined only experimentally. The introduction of the Kozeny constant CK, which is usually applied for the determination of pressure drop by flow through granular materials, results in

(8)

Values from about 3 to more than 5 are given for CK [4, 13]. For the calculation of the permeability of the layers studied, the value CK = 4.94 was used. It results from tests on layers of spheres of equal size and is, no doubt, also valid for non-spherical, but smooth grains [3] . The permeability of the layer of glass rings was also calculated with CK = 4.94 with the additional assumption that flow does not take place in about Ya of the holes inside the rings since some of the holes lie perpendicular to the direction of flow or are covered by neighboring rings. This corresponds to a decrease off and dh. Near the boundary surfaces of a layer of granular material, the fraction void f is somewhat larger than inside the layer, which changes dh and k. For an infinite layer in agreement with equation (6) :

(9) with

(10)

For a finite layer only equation (6) is valid for dh, in which the sum of the grain surfaces and the bounding walls is to be substituted for 0 z· Thereby the effect of the boundary on the permeability can be approximately taken into account. This approximation is probab­ly quite good for horizontal layers, since for these the direction of flow is mainly transverse to the large boundary surfaces. In vertical layers the direction of flow is mainly parallel to the large wall surfaces. The greater permeability of the loose surface layer has therefore greater influence, especially at the onset of convection. This explains why for vertical layers the onset of convection is earlier but less sharply defined than for horizontal layers. This can most easily be seen from the layer consisting of 5 mm steel spheres. Only

252

II -4

by more fully developed convection is the effect less than in horizontal layers. For fibrous and other materials with strongly direction-dependent permeability, remarkably greater differences in the convection effect for different orientations of the layer are to be expected.

CONCLUSIONS

The measurements conducted show that free thermal convection in granular materials can be of considerable significance. The results indicate that the onset of convection and fully developed convection are not dependent of the same factors. On the basis of the results, it is possible to estimate the order of magnitude of the effect of free con -vection on heat transfer through plain layers of any granular material when the permeability of the granular material is known. This can be determined from pressure drop measurements by laminar flow.

ACKNOWLEDGMENT

The study conducted in the Maschinenlaboratorium of the Technische Hochschule Karlsruhe was suggested by Prof. Dr.-Ing. K. Nesselmann, who, as well as Prof. Dr.-lng. K. Linge, supported the work with valuable advice. The study was made possible by the financial aid of the Deutsche Forschungsgemeinschaft in Bad Godesberg.

REFERENCES

1 . ASTM designation C 177-45. 2. K. Bartens, Warmeleitfiihigkeit eines Gemisches von Metallkugeln und 01. Forschg. Ing.-Wes. 7

(1936), 1 74· 3. W. Barth, Der Druckverlust bei der Durchstromung von Fullkorpersaulen und Schuttgut mit

und ohne Berieselung. Chem. Ing. Techn. 23 (1951), 289/293. 4. H. Brauer, Eigenschaften der Zweiphasen-Stromung bei der Rektifikation in Fullkorpersaulen.

Dechema-Monogr. 37 (1960), 7/87. 5. DIN 52612 Blatt I. 6. G. Kling, Das Warmeleitvermogen eines Kugelhaufwerkes in ruhendem Gas. Forschg. Ing.-Wes.

9 (1938), 28/34. 7. E. R. Lapwood, Convection of a fluid in a porous medium. Proc. Cambr. Phil. Soc. 44 (1948),

508/52! . 8. G. Lorentzen, E. Brendeng, On the influence of free convection in insulated, vertical walls. Proc.

of the Xth intern. congr. of refr., Copenhagen 1959, Vol. I, 294/303. 9. W. H. McAdams, Heat transmission, 3rd Ed., McGraw-Hill Book Co., New York 1954·

ro. F. T. Rogers, C. W. Horton, H. L. Morrison, L. E. Schilberg, Convection currents in porous media, part I-V. J. Appl. Phys. r6 (1945), 367/370 ; 20 (1949), 1027/1029 ; 21 (1950), r r77/rr80 ; 22 (1951), 1476/1479 ; 24 (1953), 877/880.

IL N. C. Toftegaard, D. Ahlquist, P. 0. Persson, Temperaturfalt och varmeflode i pa olika satt isolerade vaggar. Kylteknisk Tidskrift lJ (1954), 13/19.

12. G. B. Wilkes, Heat Insulation, John \Viley & Sons, Inc., New York 1950. r3. M. R. ]. Wyllie, A. R. Gregory, Fluid flow through unconsolidated porous aggregates. Ind.

Engng. Chem. 47 (1955), 1379/1388. 14. R. A. Woodin�, Steady state free thermal convection of liquid in a saturated permeable medium.

J. Fluid Mech. 2 (1957), 273/285 ; 3 (1958), 582/600.

DISCUSSION

J. A. Clark, U.S.A. : How do the physical properties of the grains affect the results of your tests ?

K.-J. Schneider, Germany : From the physical properties of the grains only the thermal conductivity Ak influences the convection part Ac of the effective thermal conductivity Aerr. This influence decreases with increasing Gr-number.

G. G. Haselden, U. K. : In the dimensional analysis reported by the author the Nusselt number is given as a function of the Prandtl number and a modified Grashof number. Martin and I have found that the free convection process can only be described by a more complex equation which involves geometric factors and other conductivity effects :

253

II-4

where :

( ki Nu = f Pr · Gr · kg . � )

k1 and kg are the conductivities of the evacuated insulant and gas respectively ex is the permeability L is the thickness of insulation H is the height of insulation. Does the author think that the simple equation he reports is adequate ?

K.-J. Schneider, Germany : In the paper it is not stated that Nu depends on Gr and Pr only. There is no contradiction between the results of the experiments and your equation. The modified Grashof-number Grk used in this paper (equation 3) ist the product of the usual Grashof-number Gr and kf o2 (it is cxfL2 in your equation). The experiments indicate also an influence of the geometric proportions : The results for horizontal and vertical layers (Figs. 3 and 4) are different. Experiments (not mentioned in this short paper) on vertical layers with another value of hf o (it is HfL in your equation) indicate that Nu decreases somewhat with increasing hf o. The parameter of the curves in Figs. 3 and 4 is the ratio }. 0fAz (corresponding to k1fkg in your equation). The values of Ao/Az are :

254

Ao/}. , ""' 1,3 for the upper curve, glass in water ). al). z ""' 4 for the middle curve, glass in turpentine ). of). z ""' 13,5 for the lower curve, steel in turpentine.

II-8

On the Thermal Conductivity of Powder Insulations

Conductibilite thermique des poudres isolantes

A. EVEREST, BS, research engineer; P. E. GLASER, PhD, group leader; and A. E. WECHSLER, ScD, senior research engineer. Arthur D. Little, Inc., Cambridge, Mass., U.S.A.

SOMMA/RE. On presente les resultats d'etudes theoriques et experimentales a basse temperature concemant l'influence des variables physiques sur la conductibilite thermique de deux poudres isolantes choisies. Les variables etudiees comprenaient la pression du gaz, le type de gaz, les temperatures limites, la grosseur des particules et la densite de la masse. On a essaye d'isoler et de mesurer individuellement l'influence des divers mecanismes d'echange thermique sur la transmission de chaleur totale.

Les materiaux etudies etaient de la silice colloidale et de la perlite. On a mesure la conduc­tibilite thermique de ces poudres dans un appareils a plaque froide de garde simple.

Les echantillons etaient exposes a des pressions de 1 x 10·5 mm de Hg a 760 mm de Hg avec de /'helium et de l'azote comme gaz residuels.

La conductibilite thermique des echantillons etait determinee a des densites de 0,06 a 0,12 gm/cm3• L'influence des temperatures limites sur la conductibilite a ete determinee pour des temperatures de la plaque chaude variant de 27 3 a 388° K et des temperatures de la plaque froide de 77 a 232° K.

On a etabli une correlation theorique entre les donnees existantes en vue de determiner /'influence des variables etudiees sur la conductibilite thermique.

1 . INTRODUCTION

The thermal behavior of insulating materials has long been the subject of theoretical and experimental investigations. This interest points both to the need for more effective thermal insulators to meet the requirements of applications extending over an ever widen­ing temperature range and to the difficulties which are encountered in these investigations. The wide use of insulating powders requires that a better understanding of the effect of physical variables on thermal conductivity be obtained, particularly at cryogenic tempera­tures to aid in the selection of effective insulators. The variables of particular interest are gas pressure and type of gas within the voids between the particles, average particle size, bulk density and boundary temperatures. Expanded perlite and colloidal silica were selected to provide a wide range of particle sizes suitable for this investigation. The following discussion summarizes the approach chosen in this investigation and the

results obtained.

2. THERMAL CONDUCTIVITY APPARATUS

Thermal conductivity tests were carried out in a modified single-guarded cold-plate apparatus [1] (See Fig. 1). A sample holder, covered by a pressure-tight polyester film (0.1 mm thick) or stainless steel cover (.025 mm thick) is mounted between the hot and cold plates and in good thermal contact with them. Vacuum-line connections allow the achievement of various pressures between about 1 x 10-6 Torr and atmospheric pressure in the container independent of the bell-jar pressure. Because the temperature in the measuring and guard dewars can be closely controlled

the effects of thermal shorts caused by the presence of the covers are small. In addition, the covers are used only when the powder insulating effectiveness is decreased by in­creasing gas pressure, resulting in higher boil-off rates.

255

II-8

Hot-plate temperatures in the vicinity of 277°K were obtained by passing tap water through the channels in the hot plate ; temperatures in the vicinity of 383 ° K were ob­tained by circulating hot oil pumped from a thermostatically controlled oil bath.

GUARD VESSEL

SAMPLE ZONE (SEE ABOVE

DETAIL)

Fig. J. Thermal Conductivity Apparatus

22"

Sample and bell-jar pressures were determined with ion and thermocouple gages, and mercury manometers, depending on the pressure range. All temperatures were measured with calibrated copper-constantan thermocouples read out on a Type K-3 potentiometer. The boil-off rate of the cryogenic fluid in the measuring vessel of the thermal conductivity apparatus was measured by volume displacement of oil in a graduate for small boil-off rates, and on a wet test meter for large rates.

3. SAMPLE PREPARATION

Colloidal silica is a white, fluffy, submicroscopic particulate silica prepared in a hot gase­ous environment (1400°K) by the vapor phase hydrolysis of silicon tetrachloride. The average size of primary particles of colloidal silica ranges between 150-200 A.

Expanded perlite is produced from a volcanic glassy rock which expands when heated above 1088°K. This expansion causes microscopic voids to be formed in the heat­softened glass, resulting in the porous structure which accounts for the low density of expanded perlite.

Perlite is available only in grades containing particles ranging in size from 100 to 1600 µ. Therefore, sieves were used to obtain particles of the desired size (297 to 354 µ).

The unsettled density of a powder can be obtained by carefully pouring the powder into a vessel of known volume. For colloidal silica and perlite, this procedure gave the following results :

Material

Colloidal Silica Per lite

Particle Size

150-200 .A

297-354 µ

Unsettled Density gm/cm a

0.034 0.043

Although an even lower density could be achieved, densities above 0.045 gm/cm3 were used to ensure that the top surface of the sample would be touching the cold plate of the thermal-conductivity apparatus.

256

II-8

For any given particle shape, a specific packing configuration will yield the maximum density. For spheres, closest packing occurs with a rhombohedral arrangement, where the sphere centers form a lattice such that each sphere has 12 points of contact. The porosity for this arrangement is 26 per cent. However, in experimental packings obtained by prolonged shaking of spherical particles, the voids account for about 39.5 per cent of the volume, which is characteristic of orthorhombic or eight-point-contact packing. Since an orthorhombic arrangement of the particles yields the highest density that can be prac­tically achieved, the maximum density of powders can be calculated if the size and weight of each particle are known. The methods by which colloidal silica and per lite are produced do not yield particles of uniform size and weight, and the very minuteness of the par­ticles complicates any attempt to physically weigh a sampling of particles to determine an average weight.

Expanded perlite is very sensitive to impact forces and the prolonged shaking necessary to obtain a maximum density would cause breakdown of the perlite particles. However, perlite with particles approximately 300 µ in diameter could be prepared in densities up to about 0.1 1 gm/cm3 with an insignificant amount of particle breakdown.

The fluffiness of colloidal silica makes high-density packing of the sample holder in the thermal-conductivity apparatus very difficult. Because there was more interest in the thermal-conductivity value in the optimum density range, the maximum density of the colloidal silica was not considered to be of experimental significance.

4. EXPERIMENTAL RESULTS

To minimize the number of tests, we established the basic behavior of several particle size distributions for the variables under study. For example, a sample of colloidal silica at a density of 0.080 gm/cm3 was tested between liquid-nitrogen and hot-oil temperatures at several sub-atmospheric pressures. Once this data for the pressure-thermal conductivity relationship was established, only a limited number of tests were made on samples at other densities and temperature boundaries.

The results obtained are summarized in Table I and presented graphically in Figs. 2, 3 and 4.

Fig. 2 shows gas pressure increases the thermal conductivity of colloidal silica after a pressure of about 1 x 1 o- ' mm Hg is reached. The rate of increase of thermal conductivity of colloidal silica with pressure above 10- 1 mm Hg is less than that for per lite because of the small size of the colloidal silica particles.

Fig. 2 shows also the effect of residual gas on the thermal conductivity of perlite. Below about 1 x 10-4 mm Hg, the contribution of gaseous conduction to the thermal conductivity ofperlite is negligible. Therefore, the data taken at 1 x 10-4 mm Hg with air as the residual gas is a common point for both the nitrogen and helium curves. All of the data points fall approximately as predicted by calculations, i. e., at any given pressure above about 1 x 10-4 mm Hg, the thermal conductivity of perlite is higher with helium as a residual gas than with nitrogen.

�600 � �500 e �400-> (:; �300 �200 � i=' JOO

,8°PE:,.L���dLE SIZE� 300µ

DENSITY � 0.080 gms/cm3 Teald plate : 77•K \/ Thal plate"t 38Q•K

El COLLOIDAL SILICA: I

29-32 5

PARTICLE SIZE";' 150-200 A I �:;:'::":' �·�76.: gms/cm3 / 0.3 m Thor p!ate "t zao•K '-----l-----'-----+0-----' -;:: RESIDUAL GAS : NITROGEN >

NOTE: DOTTED PORTION OF SILICA CURVE PERLtTE: _,A,,. � �:�AB�i�u�xi:o�:��A;;g ,�s��6 1-----1------1----�, RES. GASi"' 35 0.2 �<.> �������1�::0���11:�!�NWl!�H / MT��EN ��s���

Eo��AD������ �E

UF��

E�� .____ __ _,_ ___ _,_ __ ,,�-','----' ; RUN No's IN TABLE I / 11 _ ... ...... 0.1 �

I I f----+----,,,_+t::.]6�� ' - - - -

� �3 � I I I 1 1 1 1 1 I �14 I l l

-4 l x lO

-3 l x tO

-2 -r t x l O l x l O I JO 100 1000

GASEOUS PRESSURE, mm Hg

Fig. 2. Effects of Pressure and Residual Gas on Thermal Conductivity of Colloidal Silica and Perlite

257

Il-8

Table I Test Results

c = Colloidal Silica p = Perlite Tcold plate = 77°K for runs 1-24, 29-40

= 232°K for runs 2S-28

Thermal Run Sample Density Gas Pressure Residual Thot plate Conductivity No. (% By Weight) (gms/cm3) (mm Hg) Gas (o K) Microwatts

cm - °K

1 100 c 0.048 3.2 x 10-7 Air 281 20 2 100 c 0.048 1.7 x 10-6 Air 379 44 3 100 c O.OS6 s.7 x lo-• Air 28S 16 4 100 c O.OS6 i.o x l0-6 Air 388 38 s 100 c 0.064 l .S x 10-4 Air 277 17 6 100 c 0.064 l .S x 10-1 N2 277 22 7 100 c 0.064 9.3 N2 27S SS 8 100 c 0.064 26.4 N2 273 8S 9 100 c 0.064 s.o x lo-s N, 372 36

10 100 c 0.064 4.0 x 10-6 N. 382 39 11 100 c 0.064 9.0 x 10- 2 N2 379 40 12 100 c O.D7S 2.4 x 10-5 Air 3S9 23 13 100 c O.D7S 2.4 x 10-5 Air 361 24 14 100 c O.D78 i.s x 10-4 Air 376 43 lS 100 c 0.080 4.S x 10-4 Air 279 20 16 100 c 0.080 1.4 x 10- 1 N2 279 27 17 100 c 0.092 7S3 N2 372 176 18 100 c 0.092 760 N. 372 181 19 SOC/SOP 0.080 1 .4 x 10-s Air 379 28 20 SOC/SOP 0.080 1.8 x 10- 1 N2 379 S7 21 SOC/SOP 0.092 1.1 x 10-5 Air 383 26 22 2SC/7SP 0.080 2.9 x 10-4 N2 38S 23 23 100 p 0.064 4.8 x lo-6 Air 382 2S 24 lOO P 0.080 S.3 x 10-5 Air 284 7.5 2S 100 p 0.080 2.3 x l0-6 N, 383 26 26 lOO P 0.080 1.6 x 10- 1 N2 382 4S 27 lOO P 0.080 20 N2 383 92 28 lOO P 0.080 760 N2 383 360 29 lOO P 0.080 s.o x 10-6 Air 376 23 30 lOO P 0.080 l.S x 10-s Air 379 22 31 lOO P 0.080 8.4 x 10-5 Air 381 23 32 lOO P 0.080 1.0 x 10-5 Air 382 23 33 lOO P 0.080 2.4 x 10-2 N, 38S 26 34 lOO P 0.080 22.5 N, 380 63 3S lOO P 0.080 762 N. 380 320 36 lOO P 0.080 i.s x lo-• He 382 33 37 lOO P 0.080 20 He 382 88 38 lOO P 0.080 764 He 382 660 39 lOO P 0.092 1.4 x 10-5 Air 37S 21 40 lOO P 0.114 S.S x 10-5 Air 383 36

Fig. 3 shows the effect of change in cold-plate temperature on the thermal conductiv-ity of per lite. The thermal conductivity of perlite, as predicted by theory, is higher at a mean temperature of 300°K than at a mean temperature of 220°K.

2S8

360

PERLITE : 320-

28

l' � 24�

:; i 2--

e �160 > ti ;s 120

8 J " 80 � ,_ � 25

29-32 -·

l x l O

I PARTICLE SIZE� 300JL DENSITY • 0.080 9m•/cm3 Tllo! p1011"i :!IBO"K RESIDUAL GAS: NITROGEN

26

c 33 . . . . . . . -3 - 2 - 1

l x lO lxlO titlO

I I I I I I !

I

5/ - · "';..?

"Tcold 1 1 - ,,, 77 •I(. Tcold p\ol•

. . . . . I 10

I I I ,

GASEOUS PRESSURE, mm Hg

II-8

�,'28 0.24 , , " I ,� 022

I , , I I 0.20 I I / ...

I 0.IB ci: I I i-0.1• ,i; I ... I

0.14 I I "' I 0.12 ,: I ... > 0.10 � 0.08 z 8 0.06 _J .. "' 0.04 �

,_ 0.02

. . . . 100 1000

Fig. 3. Effect of Change in Cold Plate Temperature on Thermal Conductivity of Perlite

3 DENSITY, lbs/ft

5 0 PERUTE:

Tcold plole = 77"K Thotplote='380°K c::J COLLOIDAL SILICA: Tcold plot• = 77"K 0.03 Thot plote�280°K

£. COLLOIDAL SILICA: 40,+--'._-1----/-� Tcold plote = 77"K ::.:: Tho! plait �380°K

NOTE: P<103mmHIJ

0.04 0.06 0.08 0.10

DENSITY, gms/cm3

0.12

... N�

l 0.02 .,: ... > ;::

I 0.01

Fig. 4. Effect of Density on Thermal Conductivity of Evacuated Perlite and Colloidal Silica

Fig. 4 shows the effect of density on the thermal conductivity of evacuated perlite and colloidal silica. The curve for perlite shows that the optimum density for lowest thermal conductivity is in the vicinity of 0.09 gm/cm3• The optimum density for colloidal silica is about 0.06 gm/cma.

5. THEORETICAL AND EMPIRICAL CORRELATION OF DATA The following techniques were used to correlate the experimental data, using theoreti­

cal and empirical considerations : Simple analytical models were chosen to represent the insulation materials. Parameters

specified in the models were particle size (assumed uniform), density, packing arrangement, chemical composition of particles, and thermal properties of the pure solid components.

259

Il-8

The thermal conductivity of the insulation materials was assumed to be equal to the sum of the contributions of the mechanisms of solid conduction, gas conduction, and thermal radiation. The contributions of these mechanisms were estimated by correlations presented in the literature and derived in this study.

Calculated and estimated values of thermal conductivity were compared with experi­mental measurements under several conditions of temperature, gas type, and gas pressure.

a. Perlite

For purposes of calculation perlite was characterized as consisting of hollow, uniform spheres, 300 µ in diameter, composed of a glassy material with a true density of 2.5 g/cm3• True porosities of 93 to 99 per cent and "Bulk" porosities of 26 to 48 per cent were chosen to correspond to densities of 0.032 to ·o.1 6 g/cm3•

The thermal conductivity of the pure material was assumed to be independent of tem­perature in the range considered and was given a value, based on the chemical composition of the perlite, of 12.1 x 103 microwatt/cm-°K.

Estimates of the contribution of solid conduction to thermal conductivity were made according to the theoretical and empirical correlations of Russell [2], Riemann [3], Wilhelm [4], and according to an empirical method (developed in this study) which accounts for hollow particles. Combined contributions of solid and gas conduction were estimated by relations developed by Rayleigh [5], Russell [2], Woodside [6]. Gorring and Churchill [7], and Kunii and Smith [8]. At low gas pressures, the estimated values for the contribution of these mechanisms were very high or very low (except for the correlation of Riemann) compared with the experimental data.

Values for the radiation contribution to thermal conductivity, estimated by correlations of Damkohler [9], Laubitz [10], Schotte [ 1 1], and Russell [2], were very similar.

For a typical perlite insulation of 0.064 gm/cm3 at an average temperature of 228°K, with a nitrogen gas pressure of 10-5 mm Hg, the contributions of solid conduction, gas conduction, and radiation are estimated to be 6.0, 0.01 and 6.3 microwatt/cm-°K, respectively, or an over-all thermal conductivity of 12.3 microwatt/cm-° K. For the same conditions, experimental values are in the range of 20-24 microwatt/cm-° K. The disagree­ment is not unreasonable in view of the assumptions made in evaluating the individual contribution to thermal conductivity and in view of the fact that we were not completely able to account for the interactions between the individual components.

The effects of gas pressure on thermal conductivity can best be correlated by using Schottes' method for computing the gas conductivity and estimating the interaction be­tween gas and solid by the method of Deissler and Eian [12] . Fig. 5 shows typical

260

l'

PARTICLE SIZE = 300,U DENSITY • 0.08 gms/cm3 Tcold p!alt • 77"K T1io1p101e • 380"K

NOTE! NUMBERS BESIDE DATA POINTS REFER TO RUN

1-4o's IN TABLE I

POROSITY• 0.37 NO SOLID-GAS INTERACTION

0.18

0.t6N�

0.14 � 0.12 � 0.10 � > ;:: 0.08 �

0.08 8 -' � 0.04 �

0.02 I-

o -l-����������.-��--.���,-��-+ l ll l0

-3 l x lO

-I I 10

GASEOUS PRESSURE, mm Hg 100 1000

Fig. 5. Calculated and Experimental Values of Thermal Conductivity of Perlite

II-8

agreement between experimental and calculated values of thermal conductivity of per lite as a function of gas pressure. The contribution of radiation and solid conduction was assumed to be 20 microwatt/cm-°K in the calculation.

b. Silica Because of the extremely small particle size and lack of data on optical properties of

colloidal silica, computation of the solid conduction and radiation contributions to thermal conductivity could not be made. Estimates of the contribution of gas conduction as a function of gas pressure were made according to the ADL method [13]. Fig. 6 shows calculated and experimental values of the conductivity of colloidal silica. Effective inter­particle distance was chosen as a parameter ; comparison of the experimental data with calculations indicate effective interparticle distances of 50 to 100 particle diameters. This indicates agglomeration of the colloidal silica into units 150-250 times the ultimate particle size.

160�-----------------�-�

140 f20 1100 .. � ao � � 60 8 ;;i 40 �

DENSITY•0.064 gm$1Cm3 Tcold plate = 71 " K T �ol plate = 280"!<

op� PARTICLE SI Z E ' 150 ,\ d = EFFECTIVE INTERPARTICL.E SPACING

NUMBERS BESIDE DATii. POINTS REFER TO RUN No's IN TABLE I GAS CONDUCTIYIT'I' CALCULATED

USING AOL METHOQ(13J

"' 20 cj,;'�-....i'.::.:"i.:":11:1"�'"•�;:;;:::::::::.__:::::;::::;;. _ _.::::::::�

10 100 GASEOUS PRESSURE, mm Hg

0.10

0.08 � •

<D 006 � > i 0.048

-' .. � 0.02 I-

1000 Fig. 6. Calculated and Experimental Values of Thermal Conductivity of Colloidal Silica

6. CONCLUSIONS The results of the combined experimental and theoretical study on the effects of phys­

ical variables on thermal conductivity of insulating powders have indicated the marked effects that do occur. This would underline the necessity of closely controlling the condi­tions under which insulating powders are to be used so as not to suffer a deterioration in their insulating effectiveness. Perlite and colloidal silica exhibit excellent insulating properties. Their thermal

insulation is approximately 10 times that of multi-foil radiation shield insulations. Their insulating effectiveness does not appreciably decrease even when gas pressures rise to a few millimeters of mercury. This behavior has been observed for both nitrogen and helium as a residual gas and can be assumed to apply to hydrogen. This would imply that a pow­der insulation purged with helium or infiltrated by hydrogen, although exhibiting a slightly higher thermal conductivity than a powder containing nitrogen, would not show a marked decrease in insulation effectiveness. At higher pressures, the thermal conduc­tivity of powders with helium residual gas increases about fivefold. For highest insulating effectiveness, it is desirable to use an optimum density of the pow­

der insulation. For evacuated per lite between 77 and 380° K, the optimum density was found to be approximately 0.09 gm/cm3• For evacuated colloidal silica between 77 and 280°K, and between 77 and 380°K, the optimum densities were about 0.056 gm/cm3 and 0.064 gm/cm3, respectively. No general theory exists which can precisely predict the combined effects of the

mechanisms of conduction and radiation through powders and take into account the

261

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influence of physical variables. Specifically, data on the optical properties such as ab­sorption and scattering cross sections are lacking, and further work is required to charac­terize the influence of contact area on solid conduction.

5. ACKNOWLEDGMENTS

The authors gratefully acknowledge the support of the George C. Marshall Space Flight Center, NASA in this investigation. Their thanks go to Dr. K. Schocken, whose interest and suggestions during this work were greatly appreciated.

REFERENCES

r. I. A. Black, A. A. Fowle and P. E. Glaser, Proc. of 1 0 th International Congress on Refrige­ration, Denmark, r959.

2. H. W. Russell, J. Amer. Ceramic Society, r8, r (r935). 3. G. M. H. W. Riemann, Die Partiellen Differential·Gleichungen der Mathematischen Fhysik,

Band r, p. 474, F. Vieweg and Sohn, Braunschweig (r9r9). 4. R. H. Wilhelm, W. C. Johnson, R. Wyncoop and D. W. Collier, Chem. Eng. Prog., 44, ro5 (1948). 5. L. Rayleigh, Phil. Mag., 34, 481 (1892). 6. W. Woodside, Can. J. Physics, 36, 815 (1958). 7. R. L. Garring and S. W. Churchill, Chem. Eng. Prog., 57, 53 (1961). 8. D. Kunii and ]. M. Smith, A. I. Ch. E. Journal, 6, 7I (r960). 9. G. Damkohler, Der Chemie-Ingenieur, Eucken-Jakob, Vol. III, Part r, p. 445, Akadernische

Verlagsgesellschaft M. B. H., Leipzig, Germany (r937). ro. M. ]. Laubitz, Can. J. Physics, 37, 798 (1959). r r . W. Schotte, A. I. Ch. E. Journal, 6, 63 (r960). 12 . R. G. Deiss/er and C. S. Eian, NACA Research Memorandum R M E 52 C o5 (1952). r3. A. E. Wechsler, Aerodynamically Heated Structures, Edited by P. E. Glaser, p. 250, Prentice

Hall, Englewood, New Jersey, r962.

DISCUSSION

C. Hocking, Sweden : Are the results in agreement with Smoluchowski's observation of 1901-191 1 ?

The authors : The results presented in our paper are in general agreement with the observations made by Smoluchowski as reported at the 1911 International Congress of Refrigeration. However, the data that Smoluchowski has reported is in error and the corrections required have been discussed by N. C. Liu and W. I. Dobar (The Nature of the Lunar Surface : Thermal Conductivity of Dust and Pumice) in the Proceedings of the Lunar Surface Materials Conference, Lunar Surface Layers -Materials and Characteristics, to be published by the Academic Press, December, 1963, and edited by J. W. Salisbury and P. E. Glaser.

J. Menard, France : Avez-vous fait quelques experiences confirmant les calculs presentes au sujet du partage a peu pres egal du flux thermique total entre le rayonne­ment et la conduction solide, aux tr es basses pressions residuelles ?

The authors : We have not carried out any confirming experiments on the contributions of the mechanisms of radiation and conduction at very low gas pressures. Experiments are in progress which we believe will provide an indication of the contributions of the indivi­dual heat transfer mechanisms.

M. Griffith, U. K. : Can we rely on the maintenance of vacuum insulation in practice if these effects are so critical?

262

II-8

The authors : The maintenance of the high vacuum required for the adequate function­ing of an evacuated powder insulation can be reliably obtained in practice. Examples of devices which utilize evacuated powders are over-the-road transport dewars which are capable of retaining a vacuum for several years. In addition a large number of labora­tory size dewars have been produced commercially without suffering from unexpected leaks. Additional references on this subject may be found in the review paper on "Evac­uated Powder Insulations" by P. E. Glaser, to be published in the Bulletin of the Inter­national Institute of Refrigeration.

263

11-19

The Influence of Gas-Filled Cells on Thermal Conductivity of Rigid Polyurethane Foam

Influence des cellules remplies de gaz sur la conductibilite thermique de la mousse de polyurethane rigide

Dr. F. KAHLENBERG Robert Bosch G.m.b.H., Giengen/Brenz, Germany

SOMMAJRE. Si des isolants courants, tels que la fibre de verre, le liege et meme quelques materiaux plastiques, par exemple le polystyrene mousse avec des cellules remplies d' air, presentent une diminution presque lineaire de la conductibilite thermique en fonction de I' abaissement de temperature, cette relation lineaire ne se maintient pas bien pour la mousse rigide de polyurethane mousse au R 11.

Au-dessous de 10 a 15° C, il se produit une legere elevation de la conductibilite thermique au-dela du point d'elevation maximal a -30° C, la conductibilite diminue en fonction de l' abaissement de la temperature.

En s' appuyant sur !es considerations theoriques, les experiences montrent qu' en ajoutant de faibles quantites d'anhydride carbonique au melange de gaz dans les cellules, le maximum de conductibilite par rapport a la courbe de temperature moyenne a presque disparu.

Kinetic theory indicates the known relationship that the thermal conductivity of a gas decreases with increasing molecular weight. As the heat transfer of cellular materials is influenced mainly by the gas filling the cells, it may be expected that thermal conductivity will be lowered by encapsulating high molecular weight gas. This could be confirmed by the thermal properties of rigid polyurethane foam.

Generally this foam type is produced by simultaneous formation of a gas and a poly­urethane polymer from a hydroxyl-containing resin and a polyisocyanate under such conditions that the polymer is expanded by the gas. Catalyst and stabilisers must be added to promote the reaction at the wished rate and to create the proper conditions for the polymer to form.

In the early systems carbon dioxide was used as blowing agent, produced by the reaction of the isocyanate with water. If the foam is not sealed with vapour-tight metal foil, C02 diffuses very quickly out of the cells and is replaced by air at a lower rate.

The thermal conductivity of these air-filled foams is found to be 0.032 Kcal/mh° C at a mean temperature of 25° C.

In the last years fluorochlorohydrocarbons such as fluorotrichloromethane (CF Cl3) proved to be most satisfactory as blowing agent in the preparation of these foams. CF Cl3, known as R 1 1 in the refrigerating industry, is a volatile liquid which is volatilized by the heat of the chemical reaction between isocyanate and polyol. It is dissolved into either the polyol or isocyanate.

The thermal conductivity of R 1 1-expanded urethane foam is found to be 0,015 -0,016 kcal/mh ° C at a mean temperature of 25 ° C.

This low heat transmission rate is due to the high molecular weight of R 1 1 which is nearly four times higher than that of air.

R 1 1-expanded urethane foams show an anomal relationship between thermal con­ductivity and mean temperature [1, 2].

While the common insulating materials such as glass fiber, cork and some plastic foams too, for example air-filled polystyrene or polyurethane foam, show a nearly linear decrease of thermal conductivity with decreasing temperature down to -80 to -100° C, this linear relationship does not hold for R 1 1-expanded rigid polyurethane foam.

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II-19

Below 10-15°C a slight increase in thermal conductivity occurs. Beyond the point of maximum rise at -30° C conductivity decreases with decreasing temperature (Fig. 1 curve 1).

.034 � - 032 --.... Q! .030

0 �.028 ... � . 026 .... t . 024 � o . 022 0 al . 020 � ;J . 0 1 6 "' "c: - 01 6 'i: . 0 1 4 " ·n � . 0 1 2 ... ... " 0

2

0 -30 -20 -10 0 +10 +20 +30 Mean Temperature ( oc )

Fig. r . Thermal Conductivity Versus Mean Temperature Curve I : R I I-expanded polyurethane foam ; density 32,5 kg/m3 Curve 2 : Theoretical curve for air-filled cellular material ; density 40 kg/m3 Curve 3 : Theoretical curve for fluorocarbon-expanded polyurethane foam ; boiling point of the

fluorocarbon below -30°C ; density 40 kg/m3 Curve 4 : R n-C02-expanded polyurethane foam ; aluminium clad panel, initial data ; density

51,7 kg/m3 Curve 5 : Panel from curve 4, aged 3 months at 25°C

This effect results from the condensation of R 11 in the cells at lower temperatures. Simultaneously the molar concentration of the gas content of the cells is changed towards a higher percentage of non-condensible air which has a higher heat transmission rate than gaseous R 1 1 .

Below -30° C this foam type behaves like air-filled insulating material, due to a certain amount of air which is unavoidable from foam preparation.

For practical use it is of great interest how to avoid the increase of heat transfer at lower temperatures.

Resulting from theoretical consideration there are two possibilites. 1 . Application of a blowing agent with low heat conductivity and a low boiling point

at atmospheric pressure. For application in home freezers, where service temperatures are between + 30 and

-30°C (hot and cold), the boiling point must be below -30° C. In this case R 12 or R 22 will be advantageous (see table), and conductivity versus mean

temperature-relationship will correspond to curve 3. Because of bad handleability of these low boiling liquids at room temperature, they are not yet employed in industrial preparation of urethane foam.

2. A second possibility consists in applying a blowing agent with low thermal conducti­vity beside R 1 1, not condensing in the interesting temperature range.

This was confirmed by thermal conductivity measurements made on foam samples the cells of which were filled with a R 1 1 - C02 - Air - mixture.

The reacting mixture for the samples contained 1 part of water per 100 parts of polyol. Immediately the isocyanate (MDI) reacted with water with formation of carbon dioxide.

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11-19

Table

Blowing Molecular Boiling Point, Thermal

Agent Weight ° C, at 1 atm Conductivity [3]

of Vapor at 30° C

R 1 1 137,4 +23,4 0,0078

R 12 120,9 -29,8 0,0094

R 22 86,5 -40,8 0,0101

Still air 28 -196 to -183 0,0223

C02 44 -78,5 (sub!.) 0,0142

The test panel was 300 x 300 x 40 mm with 0,75 mm aluminium covering each 300 x 300 mm face. The foam was machine mixed and was poured vertically.

The thermal conductivity determinations were made by a stationary two-plate-method. The apparatus used in the experimental work was a horizontal guarded hot plate corres­ponding to DIN 52612 and ASTM-Designation C 177-45. Data obtained are shown by curve 4.

The maximum at -30° C, which is characteristic for R 1 1-expanded foam with a certain amount of air in the cells, has practically disappeared. This is due to a C02-air­mixture in the cells, which has a lower conductivity than pure air, but a higher one than pure carbon dioxide.

After the initial measurements the test panels were aged 3 months at 25° C and then measured for a second time. From these data a new curve 5 resulted, shifted towards higher conductivity values, with the known maximum at -30°C.

The reason is that C02 had diffused out of the cells through the unsealed face of the panel, so that a mixture of R 1 1 and air only filled the cells.

Because of the high permeability of the cell walls of urethane foam for carbon dioxide it is more advantageous applying a gas with an extremely low diffusion rate, for example R 12.

The so-called "frothing process" is characterised by the use of a R 1 1 -R 12-mixture as blowing agent in the preparation of urethane foam. The conductivity data for this foam type are represented by curve 3. It is the optimum of heat transfer attainable with ex­panded foams in the range between +30 and -30° C.

REFERENCES

r. R. E. Knox, Insulation Properties Of Fluorocarbon Expanded Rigid Urethane Foam. ASHRAE­Journal, Vol. 4, No. ro, p. 48, Oct. 1962.

2 . Own measurements.

3. Handbuch der Kaltetechnik, Vol. 4, (1956).

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DISCUSSION W.H. Emerson, U. K. : As Dr. Kahlenberg mentioned in his presentation, the rate of

diffusion through the cell walls of each constituent of a gas mixture is proportional to the difference of partial pressures. Now the diffusion coefficients of the constituents of air are an order of magnitude greater than that of Refrigerant 1 1 . If, therefore, the initial pressure in a R 1 1-blown foam is 1 atmosphere, the ingress of air will eventually give rise to a pressure of nearly two atmospheres. I would be interested to know whether the dimen­sional changes resulting from this change of internal pressure are appreciable.

F. Kahlenberg, Germany : At room temperature dimensional changes due to pressure build-up are not appreciable. There will be only a certain effect at higher temperatures.

H. Myncke, Belgium : I would like to point out that I could not admit the expression "a nearly linear" decrease of the thermal conductivity for common materials such as glass fibers, cork etc. In the temperature region considered the decrease is "progressive" but not "nearly linear" at all. This is quite normal for gas-filled substances.

F. Kahlenberg, Germany : As to curve 2 in the graph it is correct to say that the decrease of the thermal conductivity is "progressive" in the considered temperature-range. In the case of air-filled polyurethane foam the decrease is found to be "nearly linear".

E. Merlin, France : Que !es cellules soient pleines de C02 ou de Freon du bout d'un certain temps plus ou moins long, peut-etre plusieurs annees dans le Zeme cas les cellules doivent se trouver pleines d'air. Quelle est alors la valeur du coefficient de conductibilite et au bout de combien de mois ou d'annees cela peut-il se produire dans la pratique ?

L'auteur ayant trace la courbe representant la variation de ce coefficient en fonction du temps pur !es cellules remplies au Freon 12 et pour une temperature de 25°C. J'aimerais qu'il indique le temps au bout duquel la pente de la courbe devient foible et reguliere.

11 parait raisonnable de prendre la valeur de k la plus elevee pour le calcul d'une isolation pratique, par exemple, celle d'un navire.

F. Kahlenberg, Germany : For the practical use of R 1 1 or R 1 1/R 12-blown rigid poly­urethane foam there are only two possibilities in selecting the right K-factor for construc­tional design :

1 . If the foam is sealed with an efficient diffusion barrier such as metal plate or foil, one may take the initial K-factor which remains practically constant over a long period of time. This value will be in the range of 0,015-0,016 kcal/mh°C.

2. If diffusion takes place, it is necessary to select the K-factor corresponding to the equilibrium state i. e. the partial pressure difference for air is practically zero. The equili­brium K factor for MDI-based systems will be in the range of 0,023-0,024 kcal/mh° C at 25° C medium temperature.

For security's sake it will be better in any case to choose the first way. V. Ibl, Czechoslovakia : Nous avons mesure chez nous plusieurs echantillons du

polyurethane d'origine Anglais et aussi quelques-unes fabriques par lamethodeAllemande. Nous avons constater !es grands changements de ces materiaux quelques semaines apres !'examination.

Permettez-moi de demander quelles sont les experiences avec le polyurethane et le changement de ses mesures et puis, quel est la comportement du polyurethane dans !es appareils qui travaillent avec les temperatures plus bas que -30°C.

F. Kahlenberg, Germany : Samples of Freon 1 1-blown rigid polyurethane foam with saw-cut surface will show a rise of thermal conductivity due to diffusion of air into the closed cells. This could be established by several investigators. The rise does not take place if the foam is protected by diffusion barriers such as metal plates, metal foils or plastic sheets with very low permeability for air.

As to the behaviour of polyurethane foam at temperatures lower than -30° C, there have been developed formulations which guarantee non-shrinkage at -50°C, even in the free-blown state. For the sake of security it is recommended to apply higher densities, for example 35-40 kg/m3•

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B. A. Killner, U. K : In the graph accompanying Dr. Kahlenberg's paper, the rigid urethane foam tested is shown as having an initial K factor of 0.17, and several of the conclusions subsequently drawn are related to this figure. I would like to make two collllllents on this point -

1. In any assessment of rigid urethane foam this should surely be related to physical characteristics now easily obtainable with correct formulation and technique. In the case of the K factor the initial figure should not in fact exceed 0.1 1 -0.12, and details of both laboratory and field work in the U. K. and the U.S.A. are available to show that such initial figures can be regarded as a basic physical characteristic of correctly formulated and dispersed polyurethane foam.

2. With foams of such initial K factor, the significant differences between R 1 1 and R 1 1 /R 12 blown foams dealt with in this paper have not been found in our work with these systems; in both cases the initial figure was of the order of 0.1 1-0.12 and subse­quent behaviour appeared essentially similar.

F. Kahlenberg, Germany : It is true that K factors of the order of 0.1 1-0.12 are available for foamed samples in the initial state. The difference of conductivity at 20° C between R 1 1 (curve 1) and R 1 1/R 12 (curve 3) blown foams refers to the experimental error of measurement. The K factor off 0.17 stated by Mr. Killner in his collllllent must be read 0.017 because the only system used in the paper has the dimension of kcal/mh ° C.

J. R. Stott, U. K. : What is the effect of age on the conductivity for No. 3 on the last slide ?

F. Kahlenberg, Germany : If diffusion of air into the cells takes place, there will result a new curve corresponding to curve 1 with the heat conductivity shifted to highervalues.

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Urethane Rigid Foams: Factors Affecting their Behaviour as Thermal Insulants

Mousses d'urethane rigide : facteurs agissant sur leur comportement en tant qu'isolants thermiques

J. M. BUIST, D. J. DOHERTY, and R. HURD Imperial Chemical Industries Limited, Dyestuffs Division, Hexagon House, Blackley, Manchester 9, United Kingdom

SOMMA/RE. On passe brievement en revue les aspects theoriques de la transmission de chaleur a travers une matiere plastique cellulaire par conduction et par convection. On accorde une attention particuliere au troisieme mode de transmission de chaleur, a savoir le rayonnement. On presente des donnees experimentales montrant que le « diametre» cel­lulaire a une influence importante sur la valeur de K observee. On montre qu'il existe une anisotropie dans la valeur de K, due a l' anisotropie de la structure cellulaire et que, pour une concentration en gaz constante donnee, les differences de la valeur K observees sont proportionnelles aux differences du diametre cellulaire observees. On presente un argument theorique pour montrer que cette «contribution au rayonnement» est en partie rayonnante et en partie conductrice {par l'intermediaire du polymere solide) .

On etudie egalement /'influence de la structure cellulaire sur !'alteration de la valeur de K. On presente des donnees quantitatives montrant !es modifications de la pression et de la composition du gaz qui se produisent dans diff erentes conditions de vieillissement.

On indique quelques avantages pratiques retires de !'utilisation des mousses rigides dans les refrigerateurs menagers, les transports routiers et maritimes.

INTRODUCTION

Urethane rigid foams are now well established as thermal insulants of outstanding efficiency. In the early days of the development of these products the gas used for expansion was C02• On ageing, this gas quickly diffused out of the foam was replaced by air. In these foams, the thermal conductivity (which is largely determined by the

0.16 BTU in. ( kcal )* contained gas) rapidly drifted from an initial value of about

sq. ft. hr. op 0.020 mh oc to an equilibrium value of about 0.23 - 0.24 (.029 - .030). In recent years the blowing agents used have been halogenated hydrocarbons, either alone or in admixture with C02, the best known compound of this type being CFC13• It has been pointed out [1] that these gases, because of their higher molecular weight and consequent lower heat conductivity, confer enhanced thermal insulation properties on closed cell rigid foams, such that the heat conductivity CK-value) of a foam containing CFC13 in the cells is only half that of the C02 derived types. This means in effect that only half the thickness of insulation is required to perform the same duty.

MECHANISM OF HEAT TRANSFER

It has been shown [1, 2, 3] that heat transfer through a cellular material depends primarily on the conduction of heat through the gas phase, although additional heat is transferred by conduction through the solid phase and by radiation across the cells. Further measurements have confirmed these theoretical arguments and have shown, in particular, that cell size has a considerable effect on the K-value irrespective of the composition of the contained gas.

* BTU in. kcal Watt 1 --- - = 0.124

m h oc = 0.1442

sq. ft. hr. °F m ° C

271

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METHODS OF MEASUREMENT

The standard method for the determination of the K-value of rigid foam is the guarded hot plate [4, 5, 1l Quicker methods of test have been developed, i. e. the thermal conductivity probe [6, 7, 8, 9] and the heat flow meter [10, 1 1] . Experience with these methods in the authors laboratories has shown [12] that the heat flow meter is an instrument of high precision, the results of which correlate very well with those of the guarded hot plate. The thermal conductivity probe has been used in two forms by the authors - (i) as a needle containing within it the heater and thermocouple as described by d'Eustachio and Schreiner [6], (ii) as a double line system as described by Vos [9]. In the authors experience [12] these instruments are comparable to each other with regard to accuracy and reliability but they are not as reliable as the heat flow meter. In the data given in this paper all measurements are quoted in terms either of the guarded hot plate or the heat flow meter which are regarded as equivalent.

THE INFLUENCE OF CELL SIZE ON K-VALUE

It has been shown [1] that in some urethane rigid foams the structure of individual cells as revealed by microscopic examination approximates to that of truncated octahedra. However, further work has shown that this is not true under all circumstances, and that in many foaming systems of practical importance the cells are elongated [12] such that the apparent cell 'diameter' in one direction (the direction of flow) is greater than that in a perpendicular direction. Harding & James [3] and Guiffrea [13] have also published evidence to this effect.

Cell 'diameters' along and perpendicular to the direction of flow have been measured by the authors using a technique [12] based on an earlier method described by Harding [14]. K-values were alo measured in the two directions. In this work, large slabs of foam were prepared and the test specimens were taken from the centres of these slabs as close to each other as possible. The K-values are 'initial' values, measured within 4 hours of first cutting the foam. The foams were in all cases blown only with CFC13•

Table 1 .

Foam No.

2

3

Density (a) lb./cu.ft.

1.6

1 .5

1 .8

K-value (b) parallel to grain

. 123

. 132

. 126

(a) 1 lb./cu. ft. = 16 kg/m3•

K-value (b) Cell 'diameter' Cell 'diameter' perpendicular parallel perpendicular

to grain

. 1 12

. 120

. 108

to grain

mm

0.9

1 .3

0.8

to grain

mm

0.6

1 .0

0.6

(b) K-values quoted BTU in./sq. ft. hr. ° F.

The gas content of the samples in Table 1 remained unchanged, while the K-value measurements were carried out in the two directions at right angles to each other. The difference in K-value must therefore be attributed to either conduction through the solid phase and/or radiation. It is necessary to determine whether the difference in the number of polymer fibrils in the two directions would tend to produce the observed dif­ference in K-value. Doherty, Hurd and Lester [1] in their detailed discussion of heat transfer through a cellular structure considered a simplified geometrical model (Fig. 1 )

272

11-22

AREA OF CONTACT. GASEOUS LAYER

Fig. I. Simplified geometrical model of a cellular structure.

and showed mathematically that the ratio of conductivity of the solid phase K1 to the overall conductivity K is given by :

(1-f-p) Ko + (f + p) K1

(1-j) Ko + f K1

where K0 = the thermal conductivity of the gas.

f = the fractional area of the total surface occupied by polymer, i. e. the sum of the cross-sectional areas of all the fibrils (j1, f2 - f n in Fig. 1) as a fraction of the total cross-sectional area

. volume of gas p = porosity =

total volume

Transposing we have :

K (1 -f) (Ko/K1) + f K1 (1 -j-p) (K0/K1) -/- (f -/- P)

If we make the following substitutions:

Then

K1 = K0/K1 = constant k2 = 1 - k1 k3 = k1 + Pk2

K k1 -/- fk2 K1 = k3 + Jk;

We wish to know how a change in the value off will affect the value of K. Taking an extreme case where the ratio of the cell 'diameters' is 2, then f will change also by a factor of 2 (assuming constant area of cross-section of the fibrils in both directions). For a foam containing CFC13 only in the cells

K0 = .054 BTU K1 ,._, 1.4 BTU [12, 15] k1 ,._, 0.04 k2 ,._, 0.96 k3 = 0.97 given that p = 0.97

In the two cases under consideration f is (a) = 0.015 and (b) Then:

(a) � = 0.054 1

(b) � = 0.069 1

0.030.

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K The difference between the values of

K 1 in the two cases can amount to as much as

0.02 BTU in the thermal conductivity. Similarly, if the radiative heat transfer is consid­ered in this hypothetical case where one cell 'diameter' is twice the other, it can be shown from the radiation equation in the same paper [l] that the difference in radiated heat is again approximately 0.02 BTU. This is in broad agreement with the values calculated by Guenther [2]. The effect, therefore, of cell elongation is to give lower K-values for measurements made across the cells because (i) the radiated heat transfer is smaller, [ii] the heat conducted through the solid polymer is smaller. Clearly the rela­tive magnitudes of the conductive and radiative effects will vary from foam to foam according to the degree of anisotropy and the gas content. In the remainder of this paper the effect of cell size on K-Value is referred to as the 'radiation contribution' for brevity, although the additional effect of conduction through the polymer is always understood.

Further data showing the effect of cell size on K-value are given in Fig. 2. All measure­ments were carried out in a direction parallel to the grain.

0 ·1 30

0 0 0 · 1 20

0 0 0

· l l O .__ ___ o,...·•--'o"----,'oc,.·8----,,,c,·2-----:-;, .• CELL DIAMETER IN M llUMETRES

Fig. 2. Effect of cell size on K-value.

The theoretical equations indicate that the radiative and conductive effect should vary approximately linearly with cell size. The results given in Fig. 2 confirm this. Guenther [2] has published data showing a similar effect.

COMBINED EFFECT OF GAS CONTENT AND K-VALUE

In order therefore to produce a foam with a given (initial) K-value, or with the lowest possible K-value, attention must be given both to the gas content of the cells and the dimensions of the cells. Furthermore, if the cells are not of a regular geometrical con­figuration then this will give rise to anisotropy in K-value as well as anisotropy in strength [1].

It has been claimed that in order to obtain the lowest possible K-value from the gas phase, then CFC13 gas only should be used, undiluted by C02 (say). The K-value of CFC13 in admixture with C02 and with air has been determined [16] in the authors laboratories by a modified "thick hot-wire" method [17, 18]. The results are presented in Fig. 3 ('Arcton' = CFC13). The measurements were carried out at a mean tem­perature of 36°C. It is interesting to observe that CFC18 may be diluted by as much as 30% C02 with very little effect on K-value. This is an important technological fact since a foam may be blown with CFC13 and water simultaneously to produce desirable properties such as good flow and low pressure build-up. The authors have several years experience in the use of a rigid foam system in which the blowing agent is approxi­mately a 50/50 mixture by volume of C02 and CFC13• The initial K-value of this system is 0.125. An alternative foaming system in which the initial gas mixture is C02/CFC13 = 17 /83 by volume gives a K-value of about 0.115 at equivalent texture. This difference is roughly what would be expected from the data given in Fig. 3.

274

AGEING

K VALUE 20 AT 36"C

·09

·OS

· 07

·06 ·OS

,00

ARC TON I AIR +--

50 0 0/o ARCTON BY VOLUME

Fig. 3. Effect of Arcton content on K-value.

II-22

It is now well known [1, 2, 3, 19] that urethane foam is subject to an ageing process when exposed to air such that the K-value changes from its initial low value to a final higher equilibrium value. It has been shown that this is due to a diffusion process. Air diffuses into the cells until its partial pressure within the cells is equal to atmospheric pressure. Any C02 contained in the cells diffuses out very rapidly, while the CFC13 gas does not diffuse, even under high temperature exposure (see below). The effect of the inward diffusion of air is to increase the heat transfer through the gaseous phase. The degree of ageing is therefore determined by the change in composition of the gas in the cells, but the initial and final values are determined both by the gas composition

K-VALUE

0 - 1 0�--�--�---�--�--� too 2 oo 300 400 soo DAYS AGED (20°C)

Fig. 4. Comparison of the ageing characteristics of foams with different cell size.

and the cell size. Fig. 4 compares the ageing characteristics of three foams of identical initial composition but differing in cell size. The extrapolated K-values and cell sizes are given in Table 2.

Table 2.

Foam Density K-value (lb/cu.ft) (initial)

A 2.0 . 140 B 1.9 .128 c 2.0 . 115

K-values quoted in BTU in./sq. ft. hr . °F.

K-value (estimated CG

500 days)

.165

.161

.151

Cell diameter (mm)

1.5 1.1 0.5

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11-22

Table 2 again illustrates the dependence of K-value on cell size, and shows that the difference in K-value is maintained throughout the ageing process. This is to be ex­pected since the radiation contribution does not change.

The rate of ageing is determined by several factors. Assuming that a sample is exposed to air on all sides, the rate of ageing will depend on the geometrical configuration of the sample, the surface to volume ratio and the temperature of storage. Clearly a com­plex system of air (and C02) concentration gradients will exist at any one time while the ageing process is taking place. A flat thin sample will age at a much faster rate than a cubical sample of the same volume. Doherty, Hurd and Lester [ l ] have shown that

0 a DT the rate of diffusion of gas is determined by ----r/:-

Table 3

Where cp = fraction of open cells

a = Solubility of gas in the polymer

D = Diffusion constant

T = Absolute temperature

A = Cell wall thickness

h = thickness of foam sample

Pressure Measured Storage Temp. Age at 20° C. (cm. Hg).

-15°C 0 days 62.4 2 " 55.9 6 " 58.8

14 " 59.9 30 " 68.2 77 " 82.2

189 " 86.5

20°c 0 days 62.4 2 " 52.4 6 " 62.8

14 " 72.5 30 " 81 . 1 77 " 96.3

189 " 99.6

50°C O days 62.4 1 " 57.5 3 63.4 (16° C) 7 " 78.7

17 " 88.1 189 " 91.5

100° c 0 days 62.4 1 " 72.9 3 " 79.5 7 " 77.9

16 " 82.4 30 " 81 .1 77 " 81.6

276

Pressure Cale. at Storage Temp.

(cm. Hg).

54.9 49.2 56.8 52.8 60.0 72.5 75.2

62.4 52.4 62.8 72.5 8 1 . 1 96.3 99.6

68.8 63.4 69.9 86.8 97.2

101.0

81.6 92.8

101.3 99.3

104.9 103.2 104.0

11-22

It is seen that the diffusion rate is proportional to the absolute temperature. In order to obtain direct experimental evidence of this theoretical prediction, some measurements were made of the actual gas pressures within the cells of samples which were maintained at various ageing temperatures. The particular foam used in this investigation was blown by a mixed C02/CFC13 system, the actual volume of each gas component being roughly the same. Samples 6" x 6" x Y:i" were cut from the foam and stored at the following temperatures : -15° C, 20° C, 50° C, and 100°C. The pressure of the gases within the cells was measured at intervals in a specially developed apparatus [12]. The results are shown in Table 3 and Figs. 5 + 6.

All the measurements were carried out at 20° C and the figures given in column 3 are the pressures actually recorded at this temperature. In column 4 these figures are con­verted (assuming a simple Boyles Law relationship) to pressures at the different ageing temperatures. These calculated pressures are plotted against time in Figs. 5 + 6. The effect of the very rapid outward diffusion of C02 giving an early minimum is seen for the curves at 50° C, 20°C, and -15°C. It also appears that the sample at l00°C attains close to equilibrium within a few days. Equilibrium at 50°C was attained in 70-80 days, whereas more than 200 days were required for the sample at 20° C (see Fig. 4). Storage at -15° C greatly reduced the rate of diffusion, such that the cell pressure was still slightly below atmospheric, even after 200 days. It is clear from these measurements that diffusion rates of C02 and air through the cell walls are highly temperature dependent. During the course of this work, an attempt was made to break down the observed total pressure into the partial pressures of the three constituents, CFC13, C02 and air, using an infra-red analytical technique [20]. A certain amount of experimental difficulty was experienced in doing this, and for that reason the results are not quoted here in full, but it is of interest to quote the figures obtained on the samples aged for 7 days at 100°C. These are given in Table 4.

110

� , 90 . 3

ro g: 70 � 0 � 60

50

2 4 -.�.-10-.21416 0AY5'

120 CMS H�,

1 1 0 1 00

80

70 60 50

DAYS

Fig. 5, 6. Gas pressures within the cells of samples maintained at various ageing temperatures as a function of time.

Table 4.

Age (days)

0 1 3 7

Total pressure observed at 20° C

62.4 cm Hg 72.9 " 79.5 " 77.9

Pressure of CFC13 observed at 20°c

19.7 cm Hg 18.7 " 18.9 " 18.4 "

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II-22

It is interesting to observe that within the limits of experimental error, no change occurred in the amount of CFC13 over the period of 7 days. The C02 had almost van­ished in this time and the air pressure had reached atmospheric. It should be noted that the actual mass of air present in the cells at equilibrium under these ageing conditions (100° C) is less than that which would be present at equilibrium if the ageing temperature were lower. From this it may be deduced that the final K-value measured at 0°C of the sample aged at 100°C would be lower than that of the sample aged at 20° C (say). However, if the latter sample were then moved to an environment at 100°C it would presumably lose air and lower its K-value. These speculations serve to emphasize that in determining K-values, great care must be taken in the storage and ageing of the sam­ples and the thermal history of the samples must be clearly known.

SOME PRACTICAL IMPLICATIONS

The data given in this paper shows that when the foam is contained within a gas impermeable membrane no change in K-value due to ageing occurs. Experimental confirmation of this view has been obtained where foam-cored sandwich panels (lOcm thick) in the authors' laboratories with rigid P. V. C. facings retained their initial K­value of 0.125 BTU after 2 Yz years' storage at 20°C. Cut exposed blocks of the same foam would have a K-value of 0.16 BTU after this ageing period. Further evidence has been obtained from heat balance tests on refrigerators which were insulated by foaming-in-place techniques and after one year's storage the same results were obtained as with the original heat balance test.

Manufacturers of insulated equipment, designers of refrigerated vehicles etc. have therefore a part to play in ensuring that their design is such that the very low initial K-values are maintained. The foaming-in-place technique gives the best opportunity of achieving this objective.

At the Copenhagen Congress, Lorentzen [21] produced evidence to show that even when slab insulation was fitted with care, heat loss occurred by convection currents. This evidence appeared to support and explain other papers [22, 23] at the Congress which argued that carbon dioxide blown urethane rigid foam applied by foaming-in­place gave a greater reduction in heat leakage than insulants applied by conventional means. Such foams were, of course, similar in K-value to conventional insulants, their improved efficiency being due to the elimination of convection losses between joints. To this foaming-in-place efficiency factor can now be added the greatly reduced heat transmission of rigid urethane foams containing CFC13 in their cells. The efficiency of foaming-in-place has also been improved in the past 2 years due to advances in the chemistry of the polyether resins used in foams and in the design of the equipment used for manufacturing foam. There is also a greater understanding of the factors affecting the dimensional stability of rigid urethane foam [1].

Laing [22] stated at Copenhagen that "the use of urethane foam for the insulation of transportable containers is worthy of fullest consideration as it would provide maximum cubic capacity within definite limiting dimensions, combined with low weight and minimum heat leakage". This contention has been supported by the data published by Heffner [24] in which he reported the results of heat balance tests designed to simulate actual operating conditions at 50 m. p. h. road speed. A traditional road wagon insulated with fibreglass and polystyrene foam had an inward leakage rate of 5340 BTU/hr. due to transmission, 3160 BTU/hr. due to latent heat of condensation, and 3000 BTU/hr. due to air leakage. A trailer insulated by foaming-in-place with CFC13 blown urethane rigid foam, using 4" thickness in the walls instead of 6", gave a transmission leakage of 5285 BTU/hr. and only 616 BTU/hr. due to patent heat and air leakage. Heffner points out that the sealing effect of the foaming-in-place accounts for a major percentage of the improvement, although the similar heat transmission figures for 4" of urethane against 6" for the conventional insulants correlates with the respective K-values. Many other instances [25] of the efficiency of urethane foams in refrigerated transport and in other applications have been reported.

278

II-22

The improved efficiency due to the use of urethane rigid foam in domestic refrigera­tors is now accepted and a substantial number of refrigerator manufacturers are begin­ning to use these materials on a production line. One manufacturer [20] has found, for instance, that he is able to reduce the kwh used in absorption models from twice those of compression models to an equivalent figure.

Contrary to the view which many manufacturers held originally, it is now being found [27] that production costs are lower using foaming-in-place techniques in refrig­erator production, despite the higher volume for volume cost of urethane compared to fibreglass.

REFERENCES

I. Doherty, Hurd and Lester. Chemistry & Industry, 30, I962, I340. 2. Guenther, SPE Transactions, July I962, p. 243. 3. Harding and James. Modern Plastics, March I962, p. I33· 4. B.S.S. 874. 5 A.S.T.M. Method of Test C-I 77-45. 6. d'Eustachio and Schreiner. A.S.H.V.E. Trans. 58, 331. 7. Blackwell. J. Appl. Phys. 25, p. I37· 8. Van der Held and Vos. Int. Inst. Refrig. Meeting Copenhagen, I959, p. 401. 9. Vos. Int. Inst. Refrig. Meeting Cambridge, Sept. I96r.

Io. Lang. A.S.T.M. Bull. Sept. I956, p. 58. II. Norris and Fitzroy. A.S.T.M. Bull. Sept. I96I, p. 727. I2 . To be published. I3 . Guiffrea. J. Appl. Pol. SCI. 60, 9I . I4 . Harding. Modern Plastics, 37 , I 56. I5· Ratcliffe. LR.I. Trans. Oct. I962, p. I8I. I6. G. R. Nicholson. LC.I. Dyestuffs Division. Private communication. I7 . Kannuluik & Martin. Proc. Phys. Soc., A. I44, 496. I8. Kannuluik & Carman. Proc. Phys. Soc., B. 65, 701. I9· Patten & Skochdopole. Modern Plastics, July I962, p . I49· 20. M. St. C. Flett. LC.I. Dyestuffs Division. Private communication. 2 1 . Lorentzen and Brendeng. Proc. Int. Inst. Refrig. (Copenhagen) r959. 22. Laing. Proc. Int. Inst. Refrig. (Copenhagen) I959· 23. Hurd, Hampton. Proc. Int. Inst. Refrig. (Copenhagen) I959· 24. Heffner. Society of The Plastics Industry, New York. Plastic Foam Conference, New York,

26/4/62. 25. Buist. Rubber and Plastics Weekly. r44, 5, r34. I44, 6, I65. 26. Sibir, Switzerland. Technical Bulletin issued I4/2/62. 27. Von Cube. Kaltetechnik, June I962, I86.

DISCUSSION

W. H. Emerson, U. K. : In the graph of total pressure against time, shown in one of the slides, equilibrium is apparently reached in the sample at 100° C, and the authors conclude that at the end of the test period all diffusion processes had ceased. But the "equilibrium" pressure shown is considerably above one atmosphere due to a residual partial pressure within the cells of Refrigerant 11 . To say therefore that all diffusion processes are at an end is to claim that the cell walls are totally impermeable to Refrigerant 1 1 . Surely, though, one has to bear in mind the practical time scale. In the tests described here time was measured in hundreds of days, but the life of an insulated installation runs into thousands of days. I wonder whether the measurements of pressure were sufficiently sensitive to detect a rate of change which would have an appreciable effect only after many years.

D. J. Doherty, U. K. : It should be noted that the tests described in the paper may be regarded as accelerated tests since they were conducted at temperatures well above those which we would obtain in practice. In fact, the measurements at 100° C were carried on for a period of 70 days. From a comparison of the rates of diffusion of air and C02 at

279

1 1-22

100°C with those at 20°C, it is clear that rates of diffusion are greatly increased at the higher temperature so that any change in the R 1 1 concentration would be more easily picked up at this temperature. In fact, no change was found after 70 days, and it can be calculated that this corresponds to a life of more than 50 years at temperatures of 0 ° C or below.

The R 1 1 concentration in aged pieces of foam has also been checked by mass spectro­metry, and the results have confirmed the views expressed above. Measurements of K-value on test specimens aged for over two years at room temperature have no change from the equilibrium value attained after 1 year.

F. Kahlenberg, Germany : I would like to know why the R 1 1 content in the cells of aged polyurethane foam remains constant. The measurements of W. Schmidt have shown that the diffusion coefficients of polyurethane film for C02, air and R 1 1 have the ratios 100 : 10 : 1 . This leads necessarily to a certain diffusion rate for R 1 1 .

D. J. Doherty, U. K. : I am not familiar with the work of Schmidt to which you refer. It may have been done on films, rather than foams. In our experience, the ratio of the diffusion coefficients for R 1 1 and air is much less than 1 : 10. In fact, we are unable to quote a ratio since we have not been able to measure the diffusion coefficient for R 1 1 since it i s s o small. I t should b e emphasized that our tests were done on foam samples, rather than on film.

L. L. Westling, U.S.A. : Much time and space have been devoted in the past to the matter of moisture-vapour migration in permeable insulation. The effort could have been more profitably spent in the study of the newly available impervious insulating materials.

"An ounce of prevention is worth a pound of cure". There is no place in industry where this old proverb has fuller meaning than when applied to ship's low temperature insu­lation.

It is my firm conviction that there is no place on shipboard where permeable fibrous low temperature insulating materials, whether blanket or block, whether coated or sealed, are acceptable. It is realized that this can not be a popular statement in some quarters but it is based upon long experience and practical demonstration. It is possible to construct ship's refrigerators that will outlast the economic life of a ship. With permeable insulation major repairs can be expected within eight years.

For well over 25 years I have conclusively demonstrated on American West Coast vessels the feasibility of installing block type insulators secured with adhesives only, without benefit of concealed wooden grounds, and with fire-resistant rat-proof linings.

In more recent years we have demonstrated the feasibility of installing styrofoam blocks in adhesives and the use of polyurethane foams poured in place. The latter material approaches the long sought ideal for marine use in terms of thermal and moisture-vapor barriers, strength, and bonds to all shipbuilding materials.

Laboratory tests have shown polyurethane foams expanded with R 1 1 to have insula­ting values (k) of 0.13 approx., but practical considerations of aging, skin densities, and imperfections in application a value of 0.20 is suggested for design (as compared to 0.28 approx. for corkboard and dry mineral wool blocks).

Early applications of urethane foams poured in place indicated some problems of pressure upon restricting surfaces and change of dimension in aging. The application of insulation between ship's frames, beams and bulkhead stiffeners call for thicknesses that exceed economic thicknesses and the above problems which are factors of volume and thickness would be amplified in such areas.

Another problem that retards acceptance of shipboard pouring in place is the toxicity and unpleasantness of the vapors released. In ship construction the isolation of compart­ments is difficult. With the presence of the many other craftsmen in the vicinity the operation can be a health and economic hazard. The magnitude of these retardants is again a function of volume.

I would like to suggest a method of application which would minimize the effects of these retardants, promote perfection in application and should markedly reduce costs. The method is illustrated in Fig. 1, in which the interframe spaces represent at least three­fourths of boundary installations.

280

II-22

p ace , Fig. r. Method of application of expanded foam in interframe spaces.

The extent and dimensional uniformity of the interframe spaces are such that a repeti­tive operation of application is favored. The operation involves the procedure of expan­ding foam in the shop under ideal conditions within non-adhering forms providing precast core blocks of suitable lengths. The illustrated extended fins (Fig. 2) may be rasped for rough fitting and spacing in the interframe spaces.

I I I I J I I 1 1

I 1 1 ! I I

Fig. 2. Expanded foam block.

The peripheral voids thus formed would be filled with foam expanded in place behind progressively installed courses of room lining. The problems as previously described are thus minimized and effective retardants erased.

D. J. Doherty, U. K. : I thank Mr. Westling for his notes on the use of rigid poly­urethane foam in ship's insulation. I would like to comment on his observation that the in situ foaming process is toxic and unpleasant. Foaming in situ using a diphenylmethane di­isocyanate composition has shown over several years that there is no toxic hazard and no need for expensive extraction equipment. Because of modern metering and dispensing machines, foaming in situ may be regarded as a clean and easy industrial process to operate.

281

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Effectiveness of Evacuated Multiple-Layer Insulations

Efficacite des isolants composes de plusieurs couches et de vide

I. BLACK, P. DOHERTY, P. GLASER and M. MELLNER Arthur D. Little, Inc., Acorn Park, Cambridge 40, Massachusetts, U.S.A.

SOMMAIRE. On rend compte des resultats de recherche experimentale sur !es effets de la charge de compression, de la densite, des temperatures limites froide et chaude, de la pression du gaz et de l' emissibilite des ecrans de rayonnement sur la conductibilite thermique des isolants composes de plusieurs couches et de vide. On a etudie 6 types d'isolants a plusieurs couches. On a essaye des combinaisons de 3 types de materiaux empechant le rayonnement et de 3 types d' espaces. Les variables des essais etaient !es suivantes :

temperature froide limite : 77° K et 20° K temperature chaude limite : 77°K a 400° K densite des echantillons ." 0,J a 0,8 g/cm3 compression mecanique : 0 atm a 1 atm nombre d' ecrans de rayonnement : 10 a 100 pression du gaz : 10-5 a 1 atm.

Tous !es essais, sauj un dans lequel on recherchait !'influence de lapression dugaz, etaient effectues a un vide de 5 x 10-5 Torr environ dans la chambre d'essai.

On presente une breve description de l'appareil utilise dans /es experiences. On etudie /es resultats de mesures et on etablit des rapports entre eux.

I. INTRODUCTION

In the last decade, many cryogenic engineers have devoted their attention to develop­ing improved insulation systems based on the radiation-shielding concept. These systems are variously known as high-efficiency insulation, multiple-radiation-shield insulation, or superinsulation. The principle of this insulation and the background of its develop­ment were presented in versa! papers [1 - 6] at the Cryogenic Engineering Conference.

Because of the increasing industrial use of cryogenic liquids, efficient insulation sy­stems were required immediately, and the newly developed insulations were applied before being fully evaluated. Consequently, a comprehensive study of the effects of such variables as temperature, gas pressure, compressive loading, and the number of radiation shields and their emissivity on the performance of the insulation was deferred. This paper presents the results of the first phase of a study of these effects.

II. MATERIAL SELECTION

The multiple-layer insulation, which consists of evacuated layers of closely spaced shields of high reflectivity, held apart by low heat-conducting spacers, minimizes the different modes of heat transfer between the warm and cold surfaces. Radiation is re­duced by placing a chosen number of high-reflectivity shields in series ; solid conduction, by reducing the conductivity of the spacers separating the shields ; and conduction and convection through the gas, by evacuating the insulation to pressures below 10-4 Torr.

The commonly available multiple-layer insulations use aluminum foil or aluminum­coated plastic films as radiation shields, even though other materials might provide better heat shielding [1]. Representative radiation-shield materials are :

1. 0.05-mm-thick, H-19 tempered aluminum foil, bright on both sides ; 2. 0.012-mm-thick, soft aluminum foil, bright on one side only;

·

3. 0.006-mm-thick, polyester film with a thin coating of aluminum, deposited by vacuum evaporation, on one side.

283

!I-13

Conduction through the spacers between the radiation shields can be decreased by: 1 . Using low-conductivity solid materials to separate neighboring radiation shields,

the spacers being arranged to make as few point contacts as possible with the shield. 2. Using mats of fibrous materials which depend for their low thermal conductivity

on contact resistance between the layers of fibers. 3. Using embossed or crinkled shields, without spacers, which permit the shields

to touch only at a few contact points. We investigated three types of spacers using identical radiation shields : 1 . Resin-covered fiberglass netting, 0.5-mm-thick, 3 mm x 3 mm mesh; 0. 16 gm/cc

density. 2. Fiberglass mat, 0.20-mm-thick, 0.27 gm/cc density, and another fabric.* 3. Crinkled polyester film, aluminum-coated on one side.

III. TEST APPARATUS

Fig. 1 is a schematic representation of the double-guarded cold plate thermal conductivity apparatus used by us. The arguments which led us to choose a flat-plate type of calorimeter were described in the paper which we presented at the Tenth Inter­national Congress of Refrigeration [7].

LIOVIO lllTl'OGCH !;IJMO V(S�ll"

von 1,.11oc

IOI.Al l"OIC.t.f()q-------(/)

'/(HTS

Fig. I. Diagram of the double guarded cold plate thermal conductivity apparatus.

The calorimeter consists of five major parts : 1. Measuring and guard vessels, which constitute the cold plate, 2. Warm plate,

* Proprietory material.

284

II-13

3. Specimen chamber, 4. Liquid nitrogen guard vessel, 5. Evacuated bell jar.

The cold side of the sample can be exposed to discrete temperatures ranging from 4.2 to 243 ° K, depending upon the boiling point of the specific fluid used. The tempera­ture of the warm side of the sample can be varied from 20 to 500° K by proper choice of fluids and semi-automatic temperature control. The test sample, in the form of a disc, can be 0 to 5 cm thick, and 15 to 30 cm in diameter; the measuring section of the sample is 15 cm in diameter. The sample can be exposed to any desired gas at pressures from 10-6 Torr up to 1 atm. The test chamber can be continuously purged with an inert gas.

During a test, the spacing between the cold and warm plates can be changed by moving the hydraulically actuated warm plate. Parallel alignment of the warm and cold plates and the distance between them can be checked, by three dial indicators, and with an accuracy of 0.0025 cm. Compressive loads from 0 to 3 atm can be applied to the sample by a hydraulically controlled and calibrated pressure unit. Use of a guard shield permits the edges of the sample to be exposed to temperatures ranging from 4.2 to 500° K.

The heat flux through the sample is calculated from the measurement of the boil-off rate of the cryogenic liquid in the measuring vessel. The cold boundary temperature of the specimen is assumed to be the same as the boiling temperature of the liquid in the measuring and guard vessels. The temperature of the warm boundary of the spec­imen is measured by thermocouples imbedded in the warm plate. From the data obtained, the mean apparent thermal conductivity of the test sample can be calculated.

The calorimeter is calibrated by measuring the electrical heat input to a heater coil mounted on the measuring vessel [8]. The schematic diagram, Fig. 2, shows the instrumentation used for measuring boil-off rates and controlling the pressures between the guard and measuring vessels.

Sample Sp:al·e

Uil Tr.ip Oil Uath for Giw.r•I Vessel Vr;>nt

{ loni'Z..ltion G:at;e 0-�I psla �ati:1.' T .C. GaI,:c

Double A1:ti� ltyJr.aulk Cyl-

Relief V3 \vcs

J\111111 ��J��:�� R:'.' l:s1;11�����=:J Pla1c I L nd 01'0.:r.n1.·d

I Jyllr. Pum1'

1)iJ Trap g. { ; r;uh�tL'

Forcpumr

Fig. 2. Diagram of instrumentation required for double-guarded cold plate thermal conduc­tivity apparatus.

285

II-13

IV. TEST PROCEDURE

After the insulation specimen is installed in the calorimeter, the bell jar and specimen chamber are evacuated simultaneously. When a vacuum of 10-" Torr is reached, the flow of liquid which controls the temperature of the warm plate and the transfer of the cryogenic liquid into the cold plate are started. All three vessels are filled; they are then refilled in one hour, to assure full vessels after the initial high boil-off subsides. The calorimeter temperatures are stabilized overnight. The data are then taken every fifteen minutes for a period of six to eight hours. The same procedure can be repeated the following day, if necessary. We find that the data obtained during the first and the second day are reproducible within experimental accuracy, showing that equilibrium conditions have been attained the first day.

V. DISCUSSION OF TEST RESULTS

A. COMPRESSIVE LOADINGS

We examined the capability of several insulation systems to withstand compressive loading. Such loads can be caused by the weight of the insulated vessels, insulation fastening techniques, or atmospheric pressure acting on a flexible outer skin of the insulation after evacuation. By applying a hydraulically controlled force to the warm plate of the calorimeter, we increased the compressive load on the insulation samples. At increments of 0.1 atmosphere, we measured the thermal conductivity over a range of zero to one atmosphere, first increasing and then decreasing the magnitude of the pressure. We returned the warm plate to the zero load position before applying the desired new force and thus reduced the influence of friction losses between the moving parts of the calorimeter.

The effect of compressive loading on the thermal con ductivity of several multiple layer insulations is shown in Figs. 3, 4 and 5. Each curve in Fig. 3 shows the effect of a different spacing material. The radiation shielding material used in all samples was aluminized polyester film. Numbers in parenthesis indicate the number of shields used during the experiment.

At "zero load" condition, all curves start between 0.26 and 1.20 microwatts/cm-° C; this indicates the effectiveness of the different material combinations that we tested. Also, only 26 crinkled aluminized polyester shields per centimeter were required to produce the same effect as 13 shields per centimeter of the same material spaced with perforated fiberglass mat or 78 shields per centimeter spaced with fabric.

Beginning at a point where the compressive load was about 0.2 kg/cm2, the slopes of all curves are approximately the same regardless of the spacing material used, but the levels of the plots differ. The location of the curves in Fig. 3 indicates the relative effectiveness of the spacing materials.

When applied to multiple-layer insulation, compressive loading increases the solid conduction component of heat transport through the system. Applied loads decrease the high heat-flow resistance of the spacer by increasing the contact area between the spacer and the neighboring radiation shields. If the contact is assumed to be between two spheres or between a sphere and a plane, the contact area (and, hence, the thermal conductivity) increases with the two-thirds power of the load [9]. Such a theoretical curve is designated with "X" in Fig. 3. Note the good correlation between theoretical and experimental data.

Fig. 4 shows the effect of different radiation-shield materials on insulation perfor­mance. In both cases, the spacer was perforated fiberglass mat with a 50 per cent open area. The effect of radiation shielding is greatest at "zero load" condition because at higher compressions the percentage of heat transferred by solid conduction is drastically increased, while the radiation component of the heat flux remains unchanged.

The curves indicate that the mean emissivity of tempered aluminum shields tends to be lower than that of polyester shields aluminized on only one side. At zero load conditions, the thermal conductivity of the multiple-layer insulation with tempered

286

100

50

1 0

1 . 0

0. 5 '

Fig. 3.

x,,tl' / . ,

� .. I ..

! .. � /"77 //'7 7 dt 7 ,-7

11-13

--�---- - TCold= . ---�- 77'K • ..£>.-- . ••.. o - ---

�--- : ..................... ....... .. ... ..... ····· ..

·········· ..... ············ ' TCold = - --- -- · -- onov . . . .. . . . -

�

\ LEGEND

·--i:P- (60) - Crinkled Aluminized Polyester Fllm}Run No. 3001

••• 0 ... {(60) - Alumtntzed Polyester Film (61) - Fiberglass Cloth

}Run No. 2005

-o--{(10) - Alumtnized Polyester Film ( 1 1) - Perforated. Fiberglass Mat }Run No. 1033

.Q-{(95) · Aluminized Polyester Film } · • (96) • Fabric Run No. 2008

X - (Mechanical Load)213 for Reference Twarm=290 � K_!" 4 °

Test Chamber Pressure: 5 X 10·5 TORR (Approximate)

0 . 2 0 . 4 0. 6 c� ) 0 . 8 1 . 0 Mechanical Load cm2.

Effect of mechanical pressure on mean apparent thermal conductivity of the multiple layer insulations.

I I I

' (

0. 1

I (10) Aluminized Polyester Film .---� �

' ,...-,, . ,,,,.__..,...

,Y"' //

// J

0.2

/ (Run No. 1033)

- · -:::;;- --,__ '

(IO) Tempered Aluminum (Run No. 1032)

TWarm "' 29Q° K ,;t 2°

TCo ld "' 200K

Spacer Material: Perforated Fiber1tlass Mat Test Chamber Pressure: 5 X 10-5 TORR (Approximate)

0 . 4 Kg 0.6 Mechanical Load ( cm2 )

0 . 8

Fig. 4. Effect of radiation shield material on mean apparent thermal conductivity of multiple layer insulation.

287

fl-13

aluminum shields is 0.37 microwatts/cm°C and of the insulation with polyester shields aluminized on one side is 0.43 microwatts/cm°C.

Fig. 5 shows test data which tend to indicate hysterisis effects. The scatter of the data may be argued to be the result of experimental error and that no significant hysteri­sis effects can be deduced. However, several samples tested, those containing aluminized polyester film, consistently indicated that the heat flux after removal of mechanical load, was reduced compared to that measured before load application. The sample containing tempered aluminum shields indicates a lower insulation effectiveness upon removing the load. For the sample containing soft aluminum foil, an improved insulation effec­tiveness was indicated after removal of the load.

>-------+-------{]--���-- ---- - __ J __ - ---1 .. l . . . . •"] T � L O ' K

1() ,, �--,.-:- ��=?· -� ···· · ····����� _'_"_'"

__ _

�1� / ...... t;L: - �--�---� p / ,...,:.�·-.··�·��·�--·--------1------+-----+-------l-------l

-----· -------+-------! I ff? 1�1/·

j !-+'�'----!------+-----

·f-----

--

f------+----_J

� l f J •

i LOI>-__,-::_!-::_-::_--�---- -�l E;;:G-;;;EN;;-D;--_,___ _

___ _j _____ l=====::::j

! __,____ -:-- ···0···· ln�· rcasinl' Lo.Ht } Run No 11132 {(lll) T<:'mpered Aluminum

z .J..- •·•O•••• Deneas1ng Load (1\) Perforall'd f'l�rgla:.s M.i1 -?----- ----- lrK r<..•usi.og Load Hun No 2005 (tlO) Alum rnizcd F i l m :;:;; . } { : � D1.«.:n.•asrng Load (ti]) F1herglass -- _.....__ Increasing Load }

Ru n No . 2000{ (20) Soft Alurninum -�- t-ie�·n•asing Load (21) Polyt'>tcr Fiiler Mai

Test Charnher Pressure: 5 x 10 5 TORR (Approximntt:)

o.l i___ _____ 1J..1.-2 ___ ....J.o.-, ----'o.-,------.,,,.'--,------,l"".,,------,-' Mechanka! Pressure (�)

Fig. 5. Effect of mechanical pressure on the mean apparent thermal conductivity of multiple layer insulation.

The difference in the behavior of the samples may be the result of differences in the strength properties of the radiation shield material. Because the solid conduction com­ponent of heat transfer through the insulation is proportional to the two-thirds power of the load, and to the number of contact points, the zero load thermal conductivity can show an improvement or not, depending on the amount of deformation of the shields.

B. EFFECT OF THE BOUNDARY TEMPERATURE

Table 1 and Fig. 6 illustrate the effect of the warm and cold boundary temperatures on the mean apparent thermal conductivity of multiple-layer insulations. Table 1 gives the thermal conductivity of various insulations, for which the cold boundary tempera­ture was changed from 77° K to 20° K. The heat flux through each sample is approxi­mately the same for both temperature levels.

Fig. 6 shows a plot of heat flux versus warm boundary temperature for an 88 mm thick sample of then tempered aluminum shields spaced with eleven 3 mm x 3 mm resin-covered, fiberglass nettings. Heat flux is a function of warm boundary, for low

288

Il-13

Table 1. The effect of cold boundary temperature on thermal conductivity

Number of

Run No. Layers

JO

1032 J I

20 1036 20

95

21107 96

--. 200 �1°' o E � 'J 2 "--" l . S

100

Thickness

Sample (cm)

Tempered Aluminum Foil 0.76 50 % Open Fiberglass Mat

Aluminized polyester film, crinkled 0.98

Aluminized polyester film 2.0:l Fabric

LEGEND Run No. 1035

0 Cold Plate Temp. 0 Cold Plate Temp. l::J. Cold Plate Temp.

0 .___.._ _ _.__ _ _.__ _ _.__...._ _ __._ _ _, 120 160 200 240 280 320 .160

Warm Plate Temperature (°K)

Test Chamber Pressure: -5 5 x 10 TORR (Approximate)

100 200

Boundary Temperature Cold Warm Density Heat Flux

(Micro· watt/cm•) (oK) (°K)

77 290

20 290

77 279

20 279 77 286

20 286

300

(gm/cc)

0,21

0.02

0.31

Pnlyes1er Film (Crinkled)

TCold

= 200K

R u n No. 1035

400

73

63

285

276 133

136

Warm Plate Temperature (nK)

Fig. 6. Effect of boundary temperature on the mean apparent thermal conductivity and on heat flux of multiple layer insulations.

cold boundary, temperatures. Consequently, even when the cold temperature is elevated to 243°K (represented by the top point on the curve), the effect on heat flux is negligible.

The two lower curves of Fig. 6 are plots of mean apparent thermal conductivity versus warm boundary temperature calculated from the heat flux points shown therein. The one which represents a cold boundary temperature of 20° K indicates a lower thermal conductivity than the one representing a cold boundary temperature of 77°K. Thus, the warm boundary temperature is held constant, a higher thermal conductivity results from increasing the cold boundary temperature as shown by the two curves and the point at cold boundary temperature of 243 ° K.

Fig. 6 also shows a plot of the thermal conductivity of an 8 mm thick sample of 20 crinkled aluminized polyester shields. The emissivity of the aluminum layer deposited on the polyester film was apparently higher in this sample, because its thermal conduc-

289

II-13

tivity was approximately three times higher than those measured in samples made up of radiation shields of different quality materials. Thermal conductivity is approximately proportional to the third power of the warm boundary temperature. The third-power relationship for the warm boundary temperature is plotted through the test point at the top right corner of the polyester film curve. The rest of the points are marked "X". The correlation between the theoretical and experimental data is very good.

100 0

IO 0

0

c I I I ,

Cf ,JV

(�

0 . I

I v , ,

I J

·2 10

I I I

/ � I° v I

I I , J

I p

I

LEGEND Run No . 3005

(10) Tempered Aluminum

Radiation Shields and

(l l ) 3 mm x 3 mm Netting TWann = 275°K

TCold = 770 K

10'

Helium Gas Pressun; (TORR)

-� -

---

·; 10

Fig. 7. Effect of helium gas pressure on the mean apparent thermal conductivity of a multiple layer insulation.

C. GAS PRESSURE

We have examined the effects of helium gas on the thermal conductivity of a multiple­layer insulation. The test results for various gas pressures between one atmosphere and 10-5 Torr are presented in Fig. 7. We have observed that the thermal conductivity of the sample increases rapidly between 5 x 10-4 and 5 Torr. It is in this pressure region that the mean-free path of the gas molecules starts to approach the distance between individual particles.

The practical significance of this result is that multiple-layer insulations reach the insulating effectiveness only at pressures below 10-4 Torr. This may be compared to a pressure of 10-2 Torr for powders or fibers.

290

II-13

This paper presents only preliminary results of a systematic study of the variables influencing the performance of multiple-layer insulations. Further work will be con­ducted to investigate thoroughly the effects discussed here, as well as the effects of such factors as number of shields used, effect of edges, and penetration of the sample by attachments and piping.

REFERENCES

I. I. A. Black, A. A. Fowle and P. E. Glaser, Development of High Effici_ency Insulation, Adv. in Cryogenic Engineering, K. D. Timmerhaus (ed.) Vol. 5, Plenum Press, Inc., New York, 1960, p. l8I.

2. R. H. Kropschot, ]. E. Schrodt, M. M. Fulk and B. ]. Hunter, Multiple-Layer Insulation, Ibid., p. 189.

3. M. P. Hnilicka, Engineering Aspects of Heat Transfer in Multi-Layer Reflective Insulation and Performance of NRC Insulation, Ibid., p. 199·

4. P. M. Reide and D. 1-J. Wang, Characteristics and Applications of Some Superinsulations, Ibid., p. 209.

5. S. T. Stoy, Cryogenic Insulation Development, Ibid., p. 2 16. 6. I. A. Black and P. E. Glaser, Progress report on Development of High Efficiency Insulation,

Adv. in Cryogenic Engineering, D. K. Timmerhaus (ed.), Vol. 6, Plenum Press, Inc., New York, 1961, p. 32.

7. I. A. Black, A. A. Fowle and P. E. Glaser, Single-Plate Apparatus for Tests of Low-Tempera­ture Thermal Conductivity, Proceedings, Tenth International Congress of Refrigeration, In­ternational Institute of Refrigeration, Paris, 1959·

8. Contract No. N AS 5-664, Liquid Propellant Losses During Space Flight, NAS, Lewis Research Laboratories, p. 43.

9. Raymond ]. Roark, Formulas for Stress and Strain, McGraw-Hill Book Co., Inc., New York, 1954, p. 288.

lo. ]. D. V erschoor, P. Greeb/er, Heat Transfer by Gas Conduction and Radiation in Fibrous Ma­terials, ASME Transactions, Vol. 74, Aug., 1952, p. 961-967.

DISCUSSION

J. Menard, France : Si on comprend bien que le coefficient moyen de conductibilite thermique K puisse decroitre lorsqu'on abaisse la temperature froide (emissivite des ecrans amelioree, conduction solide des intercalaires abaisses . . ) on comprend difficile­ment que le flux puisse decroitre dans les memes conditions. Pensez-vous que cet effet soit significatif? Si oui, vous parait-il comparable a celui rapporte par le Docteur Kropschot (C. E. L. - N. BS. 8th Cryogenic Engineering Conference, Los Angeles, Amit 1962) et en avez-vous une explication de principe ?

The authors : We observed experimentally a decrease in the heat flux though a multi­ple-layer insulation when the cold boundary temperature was lowered from 77° K to 20° K. This decrease in the heat flux may be caused by a decrease of emissivity of the radiation shields and/or by a decrease of gas pressure in the insulation layer. However, we have no definite facts to prove this hypothesis.

Dr. Kropschot, in the article to which you refer in your question, describes the same phenomenon occurring in the evacuated powder insulation. His explanation of the occurrence is basically the same as ours.

M. Windgessen, Germany : Have you already experimented with technical applications of multiple-layer insulation ? If so, what are the dimensions of the containers and how high are the values of thermal conductivity you have obtained ?

The authors : We have built several vessels using the multiple-layer insulation. In an early application (I. A. Black, P. E. Glaser, "Progress Report on Development of High­Efficiency Insulation", paper presented at Cryogenic Eng. Conference, Boulder, Colorado, August 1960) a liquid hydrogen tank, 30 cm in diameter by 60 cm long, was wrapped with a 0.6 cm thick layer of the high efficiency insulation (see Fig. 6, Run

291

Il-13

No. 1035). The mean apparent thermal conductivity of the insulation calculated from the measured boil-off rates was :

microwatts K = 2.3 ----­

cm0C

We assumed that the heat losses through the neck of the vessel and through the support­ing structure were negligible. We have also applied a 0.6 cm thick insulation layer to a cylinder 2.5 cm diameter and 2.5 cm long. The cylinder with a weight of 23 � grams was supported by the insulation. The cylinder was kept at 85°C by a heater while the outer shell of the insulation was exposed to -55° C (I. A. Black, A. Everest, T. P. Heucling, "Design of Low-Power Crystal Oven", paper presented at 15th Annual Symposium on Frequency Control, Atlantic City, New Jersey, May 1961). The mean apparent thermal conductivity was calculated to be:

microwatts 3.0

cmoc

obtained from the power supplied to the heater. The weight of the cylinder resting on the insulation, the large exposed edge to surface area of the insulation ratio, and the relatively high warm wall temperature contributed to the poor performance of the insulation in this application.

Based on more detailed studies of the performance of multiple-layer insulations and based on added experience in applying such insulations, we have recently applied 5 layers of the multiple-layer insulation to a hydrogen tank 150 cm in diameter by 60 cm high. The tank was supported by the neck. The heat leak through the neck was eliminated by a guarded section. The mean apparent thermal conductivity of the insulation calculated from the experimentally measured boil-off rate was :

292

microwatts K = 0.3 ---­

cm0 C

Il-18

Analysis of Economic Factors Affecting the Selection of Piping In­sulation Thickness

Analyse des facteurs economiques ayant une influence sur la selection de l'epaisseur de !'isolation des tuyaux

KAREL STRNADEL, M. E., C. Sc. State Research Institute of Thermal Technique, Prague, Czecho-Slovakia

SOMMA/RE. L'un des problemes qui se posent le plus frequemment dans la conception de /'isolation thermique est celui de la determination de son epaisseur economique. Son ca/cul numerique sou/eve certaines difficultes qui sont generalement palliees par l' adoption d'une methode de solution graphique. L'A. a etabli une methode analytique s'appuyant sur le concept de cas comparables choisis, donnant une expression relativement simple. Cette methode permet des analyses commodes des facteurs economiques ayant une influence sur l'epaisseur d'isolation des tuyaux.

Elle indique !'influence des principaux parametres economiques, tels que la duree de stabili­sation, l' epaisseur optimale de /'isolation et l' elevation du coiit total d' exploitation dans le cas ou l'epaisseur reelle differe de l'epaisseur optimale. L'elevation depend du diametre du tuyau et du gradient thermique et s'eleve entre 0,2 et 4% si la difference est de 2 cm et entre 1,4 et 40%, si elle est de 5 cm. En s'appuyant sur ces faits, on peut tirer certaines conclusions sur la progression necessaire de l'epaisseur de /'isolation thermique des tuyaux.

The problem of designing economic thicknesses of insulation was the object of many studies. We know several methods of solving which aim to find minimum operating costs combined with thermal insulation or to attain optimum economic factors [1 - 15 j .

Total annual operating costs for insulation have two essential parts. One part includes mainly the price of m2 insulation, is of investment character and can be expressed by equation (1) :

A = a C1 F [u. p./year],*) (1) where a

C1 = factor expressing amortization or other form of maturity %/year = m + ns - price of insulation u.p./m2

s = insulation thickness m m,n = constants corresponding to price-list of insulation materials

u.p./m2, u.p./m3 F = surface of insulation [m2].

The second part is formed especially by the price of annual thermal losses = equation (2) :

B = K (t; - to) Ch [u.p./year],

where K = coefficient of heat passage through insulation layer kcal/m2h°C t1 - to = temperature difference on both sides of insulation° C C = price of heat u.p./kcal h = annual operating time of equipment [h/year].

The required economic thickness must comply with condition (3) :

(2)

A + B = min, (A + B)' = 0. (3) An analytic solution is possible in case of a flat insulation layer with the result of a

comparatively simple equation (4) :

s = _ (-� + �) + 1 /Ch A. (tt - to), (4) az av V an

*) u . p. = unit of price, e. g. Czech crowns, Dollar, D-Mark etc. Selection of u. p. does not effect further arithmetic method.

293

II-18

where A = thermal conductivity [kcal/mh0C] iXz, iXv = coefficient of heat transfer on inner and outer surface of insulation

[kcal/m2h 0 C] Considerable mathematic difficulties arise when adopting a similar method at a cylindric

layer, which do not allow to reach a simple resultant relation. The usual graphical solu­tion is disadvantageous because it does not indicate at the first moment the dependence of economic thickness of piping on individual technical and economical parametres. Analytic method of solving economic thickness of piping insulation as derived by the State Re­search Institute of Thermal Technique clears this difficulty sufficiently.

The principle idea is the comparison of the solved problem with the original chosen case, where technical and economical conditions are known. Thus the computation is not done with absolute values but with adopted ratios of given and chosen values. The result­ing solution is not the absolute greatness of economic insulation thickness but its change caused by new order-conditions.

If solved otherwise : Economic thickness (s) is worked out for certain order-conditions by using an arbitrary method. That means that a certain insulation thickness (s) corre­sponds to certain annual operating period (h), price of insulation (m, n) etc. (see equation (5)) :

c, h, m, n, a, dv, A, (ti - to), iXz, iXv -+ s (5) Symbols in equation (5) have the same significance as in equation (1) and (2). Should

there occur a change in any of given order-conditions, e. g. C changes into Ce, then the in­sulation thickness changes from s to s e· The aim of solution is to find dimensionless

Se Ce relation s as a function C (see equation (6)) :

(6)

The whole computation is rather complicated and its abbreviated method is shown in Appendix. The result is relation (7) :

5-" =

dv [ (l + 2s_) 1 + "' - l] s 2s dv

(a)

where

w = - p + Vr + Rq

In equation (7)

(7)

(b)

Se ratio between wanted and chosen insulation thickness

p,q ,r= factors introduced into computation. Their value is nearly constant (see Fig. 1)

d, • f'tmm 263 519

t :: r 1 0,1 -����--=-=-��� 0

dr =14mm 263 519

q, • f4 mm 263 519

' ' 0,2 --9_3 � ____!!_ q!J i :1:-t --�=---0 O.f n•

294

Fig. r . Dependence of coefficients p, q, r on piping diameter and ratio mh/nh-

11-18

Ce he n a Ae (ti-to) e fe (dv, dz), R = --- - - ---- -----

C h ne a e A (ti - to) f (dv, dz) (7c)

Ce h e where etc. is the ratio between the formerly chosen and newly stated value of heat

c h

price etc. The course of dependence (7) is shown in Fig. 2. If the binominal in rounded

brackets in equation (7 a) is replaced for approximation by first two terms of binominal series the result is equation (8) :

Se = (1 + w) s = (s - sp) + s V r + Rq. (8)

1,0

q't � I

t q3 � I G) !

f,5

q2 q't

s� s 0

-qt

-q2t C.J

-03\ • I

-qlf

Fig. 2. Graphical solution of equation 7 a - = - - 1 Se dv [( + d2v

S) 1 + W _1] S 2 S

Equation (8) resembles in structure equation ( 4) for insulation of a flat wall. This anal­ogy that could be after all expected, shows that individual technical or economical para­metres influence the economic thickness of flat and cylindric surface in the same way.

Dependence (9)

s_e = f (R) (9) s

for special values p, q, r, which correspond to economic conditions in Czechoslovakia, is shown on Fig. 3. From this follows that e.g. by increasing the value R 1,2 times -which may occur for example when heat price is 20% higher - the economic thickness changes only from 3 to 6 %, i.e. for example 4 - 9 mm increase at a thickness of 15 cm. According to this, items which will not influence the original balance considerably, do not have practically any effect on the insulation thickness. As such one may consider e.g. amount of consumed cooling medium, lubricants, price of spare parts etc., as sometimes stated in literature [4, 12, 13].

With the help of described method it is relatively simple to find out the rise of total costs connected with insulation with lower thickness in comparison with computed opti­mum. Some values may be seen on Fig. 4 which were computed for chosen economic relations and for industrial piping, whose diametres and temperatures and insulation thicknesses accordingly, are moving within broader limits than at cooling plants. Com­parative increase of all operation costs is shown on vertical axis, with reduction of insu­lation thickness from 2 to 6 cm. From the diagram follows that at greater insulation

295

11-18

( t+o) ;. � 1 1, 20 r \ 15

f.10

Fig. 3. Course of function w = f (p. r, R) if q = 0,1

thicknesses the total costs rise only slightly, e.g. at a piping with diameter 175 mm and a temperature difference of 170° C between piping and neighbourhood with 2 cm reduc­tion of insulation, the rise of costs is 2,5 %· At smaller diametres of piping and smaller thicknesses the computed economic thicknesses must be observed more accurately.

-- Ll5 • 6cm -- - 4,-4cm - · - · A.• 2Cm

f.5 �-�-- �, i It 1-1---\,+--- d,. 50+'--, 1 -1 d, - 175 I

-- . -d,•ltOOmm

0 /()() 200 300 f, - t0* r·c]

Fig. 4 Comparative increase of total operation costs if the thickness of insulation is lower than the economic value.

Shown curves are not generally valid. They are assorted for special conditions which correspond to economic relations in Czechoslovakia. By using the indicated method it is however possible to reach a suitable range of thicknesses of insulation prefabs for piping, which corresponds to estimated accuracy in choice of initial economic relations as heat price, actual operating time per year, time of maturity etc.

APPENDIX

Purpose of the problem is solution of an equation

296

d d

- (A e + Be) = O, Se

(10)

11-18

where Ae = f (se) ; Be = f (se) are implicit functions of wanted value Se. In the interest of further solution it is possible to put together analogous expression As, Ba with a constant value which corresponds to certain thickness s. Dimensionless factors

A e B e A� = fA (se) ; Bs = fB (Se)

are new functions of wanted value s e and the solution of equation

(11)

(12)

must lead to the same result as the solution of equation (10). Substitution as per (1) and (2) gives [ (m e ) ( 2se ) - + Se 1 + --- --

d ae ne ne dv ------ ------------------- +

ds as ns (ms ) ( 2s ) ---- + s 1 + -ns dv

Zn (1 + 2--=) j B (1 + P1 + Pz)s d v Ce h e Ae LI e t + As (1 + Pi + Pz)e ( 2 se

_) Cs hs As Li s t In 1 + -- -dv

The coefficient of heat passage was expressed as

= 0

k = n n

-_-_1__ ___ +_1 __ 1_n_dz_+ __ _l __

=

-(1 + ���-�z) I�(-; + �_:) IXz dz 2 A dv !Xv dv dv

With further substitutions

s = ( 1 + �: )s 1 + ' , = ( 1 + ¥;•) (1 + Pi + Pz)S Ae Lf et Ce h e ns as R = ---=-----=----(1 + Pi + Pz) e As List Cs hs ne ae m 2

M = -- --- - 1 n dv

the equation (12) is formed as follows

is

-d_ [Me + S 1 + w

S<' B 1 ] dw Ms + S u + A; T-+-;; A = O.

After completed indicated derivations and after introducing the expression In S

p = Ms + S

Bs R 1 2 S 2 ,,, + 1 + Mh S 6' - --- - -- - -- = 0 As P (1 + w)2

(13)

(14)

(15)

(16)

(17)

(18)

(19)

(20)

Exponential terms of this equation were developed in a series and only absolute, linear and quadratic terms were selected. Other ones were substituted by a remainder. Should this simplification cause an error of 10% at a value w it represents an inaccuracy of few millimetres at the computed economic thickness s e. The solution of quadratic equation which thus arises, gives

w = - p ± y � + Rq (7b) 297

11-18

and with the help of (15)

� = dv [(1 +

�) I + w - 1

] s 2s dv

In the expression (7 b) is

1 P = Z J [(4 S + Me) ln S + 4 S + 2 Me]

r = p2 - J (2s + Me)

(7a)

Bs Me + S q = J As ln S (21)

y = 1,3 [ (

4 S + � Me) ln2S + (8 S + 2 Me) ln S + 2 S + Me]. Out of two possible solutions of equation (7 b ), one is important for technical problems

only, i. e. solution with positive value of second term. Expressions p, q, r, J are functions of insulation prices and of piping-diametre. It is rather complicated to express them. However it is possible to compute once and for all times their values in a certain order (Fig. 1).

Shown mathematic method of determining s e is more accurate, the more the constants A., B s are approaching the values A e, B •· It is therefore suitable to choose A., B s so as to correspond to possibly the greatest number of solved problems and to compute to them, by using the classic method, the economic thickness s and the values p, q, r, J. Then the ratio of wanted thickness se and thickness s, which is determined from chosen conditions, will not differ considerably from one. That means that w � 0 and the computation itself will be in the region of greater accuracy. A mistake in the computation w will appear as an

inaccuracy in the resulting ratio �, some values of which are shown in the following s Table :

Se 1,25 1,5 1,75 2,0

s

dv = 0,014 m + 1,6% -2,5 -7,6 -13,8

s = 0,06 m

dv = 0,030 m + 0,75 -3,6 -9,8 -14,8

s = 0,08 m

dv = 0,104 m +o,6 -1,2 -5,6 -10,9

s = 0,11 m

dv = 0,519 m + l,05 + 0,35 -1,0 - 2,6

s = 0,19 m

REFERENCES

r . Badylkes, Vlijanie tolshtschiny izoliacii kholodilnikov na usushku produktov (Cholodilnaya technika 31 , nr. 3 VII-IX 1954)·

2 . Badylkes, Vybor racionalnoy tolshtschiny izoliacii v cholodilnych soruzheniach (Cholodilnaya technika 29, nr. 4 1952).

3 . Bach, Auswahl der wirtschaftlichen Isolierstiirke for Kiihlraume (Kaltetechnik 1953).

4- Borschke, Berechnung der wirtschaftlichsten Isolierstarke (Archiv for Warmewirtschaft Nr. 9 1923)-

298

11-18

5 . Fabry, Die Bestimmung der wirtschaftlichsten Isolierstiirke bei Rohrleitungen (Wiirme 55 1932).

6. Fontanel, Les problemes economiques de l'entreposage frigorifique (Institut International du Froid I954).

7. Gerbel, Die wirtschaftliche Starke einer Isolierung (192I).

8. Grigull, Die Ermittlung der wirtschaftlichsten Isolierdicke (BWK - N r. 5 1950).

9. Koger, Die Kosten der Instandhaltung von Produktionsmitteln (Zeitschrift VDI 94 Nr. 1 , I952).

Io. Lang, Das Problem der wirtschaftlichen Isolierdicke vom Standpunkt des Betriebsingenieurs (BKW 2 Nr. I I, 1950).

I I . Pirog, Techno-ekonomitscheskie faktory v ocenke termoisoliacionnych materialov (Cholodil­naya technika nr. 3 1955).

12. Ruetz, Die wirtschaftlich giinstigen Rohrleitungen fiir HeiBwasser und Dampf (BWK 2 Nr .10, I950).

13. Seifert, Wirtschaftliche Isolierstiirke in der Kiihltechnik (Kiilte-Industrie 36, I939)·

I4. Seifert, Wirtschaftliche Isolierstiirke in Klein-Kiihlriiumen (Wiirme- und Kaltetechnik 44, Nr. 9, I942).

15· Seifert, Thermal Insulation-Calcnlations Simplified (Combustion and Boilerhouse Engineering nr. IO, 1955).

299

T h e r modynam ics Thermodyn a m i q u e

11-7

A Non-Steady-State Method for the Measurement of the Thermal Conductivities of Liquid and Gases

Une methode a etat variable pour la mesure de la conductibilite thermique du liquide et des gaz

P. GRASSMANN and W. JOBST Institute of Caloric Apparatus and Cryogenics of the Federal Institute of Tech­nology, Zi.irich, Switzerland

SOMMA/RE. La mesure de la conductibilite thermique des fiuides est souvent f aussee par la convection libre. A cet egard, les methodes a etat variable presentent un avantage car au debut l' elevation de la temperature du fiuide est f aible. Si un fil mince est chauffe electri­quement dans le fiuide a mesurer, sa temperature s'elevera en fonction logarithmique du temps apres une breve periode de debut. Avec le logarithme du temps en abscisse et /'elevation de temperature du fil - mesuree electriquement par l' augmentation de resistance - en ordonnee, on obtient une droite sur l'enregistreur X-Y. La pente de cette droite est inver­sement proportionnelle a la conductibilite thermique du liquide.

A l' aide de cette methode on peut mesurer la conductibilite thermique des liquides en mains de 10 s, a une precision de 1 a 2%. Le debut de la convection libre est observe immediatement par un ecart de la ligne droite. Jusqu' a present, on a mesure de nombreux frigorigenes liquides a dif.f erentes temperatures et dif.ferentes pressions. Les resultats seront presentes.

On a souvent utilise cette methode comme relative, c'est-a-dire que /'augmentation de la resistance du fil dans le liquide a mesurer est indiquee sur l' axe des Y de l' enregistreur et la resistance d'un autre fil dans un liquide de reference est utilisee comme echelle logarithmique sur /'axe des X. Mais ii est possible d'obtenir /es valeurs absolues de la conductibilite ther­mique par une mesure electronique du temps.

A glance at the modern literature on thermal conductivities of liquids shows that the values given by the different scientists often differ by more than 10 %. This is even true for such common liquids as C6H6 or the often used refrigerants such as CF2Cl., CHF2Cl and so on. Since the thermal conductivity is often needed for calculation of heat transfer coefficients, it is desirable to get values reliable within a few percent. Therefore Grass­mann and Straumann [l, 2, 4) have worked out a method for the fast determination of thermal conductivities of liquids and gases by an instationary method. After a short description of the apparatus used, new results on thermal conductivities will be given, and the development of an absolute method will be described.

If beginning with the time t = 0 a wire is heated in a liquid, its temperature will rise after a short initial period (mostly less than .1 s) linearly with the logarithm of time accord­ing to the equation

T 4 n A/q* = In (4 at/r2) ---0.5772 (1)

(T = temperature increase, ). = thermal conductivity of the surrounding fluid, q* = heat input per unit length of the wire per unit time, a = thermal diffusivity of the fluid = A/ e cp where e = density and cp = specific heat, t = time, r = radius of the wire.)

Using the wire simultaneously as a heater and as a resistance thermometer, a recorder writing the resistance of the wire as a function of time would plot a logarithmic curve. By differentiation one gets

301

Il-7

B T 8 ln t

q* 4 n Jc

T = 0 if 4at/r2 = exp (0.5772) = 1 .780, or solved for t :

t = 1.780 r2/4a = t o

(2)

(3)

The time is to be measured from this point on - and not from the time t = O, when the heating of the wire begins. It would, however, be inconvenient to transform the linear time scale into a logarithmic one. Therefore a second wire is heated - also beginning with the time t = 0 - in a reference liquid, and the increase of the resistance of this wire is plotted on the other axis of an X-Y-recorder. It is easily shown that as long as the loga­rithmic law holds, the recorder will write a straight line ; for the increases of the resist­ances of both wires are linear with the logarithm of time. The inclination of the line will be inversely proportional to the ratio of the thermal conductivities of the two liquids. The conductivity of the reference liquid being known, the unknown conductivity may be easily calculated. However, the results would be dependent on an accurate knowledge of the thermal conductivity of the reference liquid.

As Fig. 1 shows, the arrangement consists of two Wheatstonebridges Bx and By, the resistance of the wire Dy within the fluid F)c to be measured forming one branch of Bx, and the wire Dx within the reference fluid Fr forming one branch of the other bridge By. B is a battery, S a switch and R the X-Y-recorder.

11 7 �

B :

R, Bx

Fig. r . Simplified diagram of connections

Dx

The actual measurement lasts no longer than 10 s. If in spite of this short time free convection sets in, this is at once recognized by a deviation from the straight line. There­fore one of the most frequent errors in measuring thermal conductivities is avoided. If the surface of the wire is spoilt-let us say oxidized-this does no harm; for only the initial time is a little longer, whereas the inclination of the straight line and accordingly the thermal conductivity calculated by this inclination are not altered .

As examples Fig. 2 shows the results for CHF2Cl ( = R 22) of different investigators. The straight line No. 5 indicates the results of Widmer [3] obtained by this method, whereas No. 6 shows the results obtained by Jobst by the absolute method described be­low.

The application of the method to gases and vapors is in some respects somewhat more complicated, as the heat capacities of gases per unit volume are small compared to the capacity of the wire and free convection sets in very early. But in spite of this, the method can be successfully applied to gases as Fig. 3 demonstrates. It shows the thermal con­ductivity ofN2 at different pressures. If the measurements were to be falsified, e. g. by free convection or wall effects, there would be a marked dependence of the results on the pressure ; for all these influences are strongly dependent on pressure.

302

0.08 1--·--+----+--1mit',C =0860 mk�a!c=.5?9 r/h�F

-100 -80 -60 -40 -20 T (°C)

0 +20 +40

Fig. 2. Thermal conductivity of R 22 = CHF2Cl measured by different scientists.

Il-7

r . W. H. Markwood and A. F. Benning, Refrig. 3. R. Plank, Zeits. ges. Kiilteind. 49 (1942) 47. Engng. 45 (1943) 95. 4. L. Chemeyeva, Kiiltetechnik 6 (1954) 225.

2. R. W. Powell and A. R. Challoner, Proc. Xth 5. F. Widmer, Kaltetechnik 14 (1962) 38/4r. Intern. Congr. Refrig., Copenhagen 1959, 6. W. Jobst, absolute method. Vol. 1, p. 384, Pergamon Press 1960.

0.03

ll02

llOI

0 0

� -

! Wm •1

· -

' I 0.86 K • 6.94 B

�

al Cmh'C ru in. C l�

,_____

i' l't °Fi1

2 3 ' PRESS URE Cat a >

Fig. 3 . Thermal conductivity o f N2 at different pressures measured by Jobst.

There are some difficulties in measuring the thermal conductivities of strong electro­lytes. The electrolytic conductivity may be avoided if the difference in voltage between the two ends of the wire is less than the discharge voltage for the ions in question. But even then some difficulties caused by polarisation remain.

Recently an absolute method of measuring thermal conductivities has been developed. The wiring diagram remains the same as shown in Fig. 1, and two photoelectric switches are mounted along the X-axis of the recorder, each connected to a Philips decade time

303

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counting unit. Both counters are started at time t = 0 by the closing of switch S. As the bridge of the recorder moves along the X-axis, an attached pointer operates first one and then the other of the photoelectric switches, stopping the corresponding counting unit. Thus, two intervals of time t1 and t2 are registered, and the thermal conductivity is evaluated using the equation

A. = � In (t2 - to) - In (t1 - to) 4n T, - T1 (4)

Except for to, all of the quantities in this equation can be measured electrically with high accuracy. Fortunately the time to defined by equation (3) is very short (in the order of magnitude of 1 ms) compared to the times t1 and t2• Therefore an error in the deter­mination of r has no influence on the value of A.. Fig. 2 shows the results obtained by Jobst on liquid F 22 (= CH F2Cl) using this

absolute method. There are many refinements of this method, too lengthy to be described here, which

will be published elsewhere [5].

REFERENCES

r . P. Grassmann and W. Straumann, Int. ]. Heat and Mass Transfer Vol. l (1960) 50--54. 2. W. Straumann, Ein instationares Verfahren zur Messung der Warmeleitfahigkeit von Fliissig­

keiten und Gasen. Thesis Prom. No. 'lo78 of ETH Zurich, and Schweiz. Archiv 27 (1961) 7, 290--304.

3. F. Widmer, Kaltetechnik 14 (1962) 2, 38-4r . 4. P. Grassmann, W. Straumann, F. Widmer and W. Jobst in "Progress in International Research

on Thermodynamic and Transport Properties" 2nd Symposium on Thermophysical Prup�rt;�s, Academic Press, New Nork and London 1962, pp. 447-453.

5. W. Jobst, in preparation.

DISCUSSION

R. W. Powell, U. K. : This paper has increased the range of available values for the thermal conductivity of R 22 in the liquid form. At the meeting in Copenhagen I presented to Commission 2 the results of some meas­

urements made by an absolute guarded-plate method on several liquid refrigerants of the fluorochloro-derivative types (R. W. Powell and A. R. Challoner. Proceedings of I. I. F. Xth Congress, Copenhagen). The values at 20°C ranged from about 10 to 22% below the then accepted values of Markwood and Benning and I ventured to propose some most probable values that were considered good to about 5 %. The method described by Professor Grassmann is claimed to give thermal conductivity

values that are accurate to 1 or 2 %, and it is very disturbing to see that for R 22 the values which I had proposed as most probable exceed those presented today in curve 6 of Fig. 2 by some 12 to 13 %- Nor do the two sets of results, curves 5 and 6, from Professor Grassmann's laboratory appear to agree to within the limits which he claims. It seems necessary that further work should be undertaken to determine the cause of

these differences. Are the differences due to limitations of some of the experimental methods, or have the liquids tested as R 22 themselves differed? I must admit that we made no chemical analysis of the sample that we used. It was supplied by the Imperial Chemical Industries Company. I would like to suggest that samples should be interchanged so that thermal conductivity

determinations by different methods be obtained for the same samples. At the Belgrade meeting of Commission 2 I described (R. W. Powell, Belgrade) how a

two-ball thermal comparator method developed at the National Physical Laboratory could be used for comparative measurements of the thermal conductivities of fluids.

304

11-7

An improved direct-reading form of this method is now is use for comparative ther­mal conductivity determinations on liquids and gases at normal temperatures (in course of publication). It could be used at once for tests on various samples that are liquid at normal temperature and pressure, say for R 1 1 and R 1 13, but some modifications will be required before R 22 could be tested by this method.

I hope that at the next meeting of Commission 2 it will be possible to present the results of thermal conductivity tests made in different laboratories on the same samples.

P. Grassmann, Switzerland : I should like very much to interchange the samples, for a large part of the deviations of the results may be caused by purity differences of the liquids. The thermal conductivity of ethanol, especially, is very sensitive to traces of water. I think that the deviations of curves 5 and 6 in Fig. 2 are to be explained in the same way, curve 6 relating to a product of the highest purity from Carrier, and curve 5 relating to a refrigerant of technical purity from another firm. I may add that my co-worker Jobst has in the meantime measured over 50 liquids, each at 6 different temperatures and each point representing the mean value of 4 measurements. This has been the work of about 5 months, which illustrates how quickly the method works.

305

II-17

Thermodynamic Properties of an Azeotropic Mixture of Freon-124 and Freon-C 318

Les proprietes thermodynamiques du melange azeotrope des F-124 et F-C 318

I. I. PERELSTEIN Scientific Research Institute of the Refrigerating Industry of the U. S. S. R., Moscow, U. S. S. R.

SOMMAIRE. Le melange azeotrope de F-124 (C2HF4Cl) et F-C 318 (C4Fs), sous­produits des processus chimiques, a sa temperature d'evaporation voisine de celle du F-142. A la base des essais des compresseurs fonctionnant avec ce melange on a demontre Jes per­spectives favorables de son utilisation.

Dans cette etude on passe en revue Jes resultats des recherches experimentales de la relation P-v-T des vapeurs saturees et surchauffees ainsi que des valeurs de la chaleur specifique du fluide bouillant du melange azeotrope. On a etudie la relation P-v-T a !'aide du thermo­metre de condensation, et du piezometre a volume variable. La chaleur specifique du fluide bouillant etait mesuree avec le vacuum-calorimetre adiabatique. Les donnees experimentales sont decrites par des equations d'interpolation sur la base desquelles et suivant des equations dijf erentielles thermodynamiques connues ont ete calcutees la chaleur de vaporisation, l' ent­halpie et l'entropie dufluide bouillant, de la vapeur seche saturee et de la vapeur surchauffee a la temperature de saturation de -20° a + 80°C. Les parametres thermodynamiques du melange azeotrope sont presentes dans la table de vapeur saturee et sur le diagramme i, S.

The azeotropic mixture of freon-124 (C2HF4Cl) and freon-C 318 (C4F8), which is a by-product of chemical production, is close to freon-142 by its normal boiling tem­perature. The tests of compressors, operating with this mixture, have indicated the perspectiveness of its utilization.

The investigated mixture is not azeotropic in the precise sense of this word but an alteration of the normal boiling temperature at 95 % evaporation of the initial amount does not exceed 0.8°C.

The results of investigation are given herein for the thermal and caloric properties of this mixture of the following composition: freon-124 - 55.4%, freon-C 318 - 44.4 %, admixtures - 0.2 %-

The pressure of the saturated vapour was measured up to 2 kg/cm2 abs. by means of a condensation thermometer. The specific weight of the boiling liquid and dry satur­ated vapour was measured with a pyknometer. The dependence P - v - T was studied by means of a piezometer of alternating volume in a wide range of temperatures and pressures, including the top boundary curve. A quite original filling device was used for determining the amount of substance in the piezometer by direct weighing. The pressure was measured in this case by a piston type pressure gauge of the class of pre­cision 0.05, the temperature - by a platinum resistance thermometer precise to 0.01° C, volume - by calibration with mercury.

The specific heat of the boiling liquid was measured in a calorimetric plant with the application of an adiabatic vacuum calorimeter precise to 1.5 %.

The following equation has been derived on the base of the data resulting from the condensing pressure

B /g P = A +

T + C lg T + D T (1)

where A = -249.95805, B = 5497.5353, C = 102.7116, D = -0.073887, and P is expressed in kg/cm2 abs., T - in °K. The equation (1) with a mean degree of exactness 0.5 % characterizes the curve of condensation in the temperature range from -20 to 80° C.

307

II-17

Data on the density of the boiling liquid can be presented quite precisely by the equation

r/ = 1 .4997 - 3.574 . 10-3 • t -5.45 . 10-6 • t2 (2) where r/ is expressed in kg/l, t - in ° C.

This equation has been derived for describing the dependence P - v - T P = 'l' (e) + p ( e) . T • (3)

In the equation (3) 'l' (e) = Ai e + Bi e2 + c1 e

3

where A1 -16.2714, B1 = -1101.81, C1 = 3063.13, <p (e) = A2 e + B2 (!2 + C2 (!3

where A2 0.606236, B2 = 0.44861, C2 = 1.7759. In equation (3) P is expressed in kg/cm2 abs., (! - in kg/l, T - in ° K. The equation

of state (3) corresponds to the experimental data with a mean degree of exactness 0.3 % in the temperature range from -20 to 120° C and pressures from 0.8 to 15 kg/cm2

abs. The critical temperature and pressure, measured on the piezometric device, respec­

tively equal: tor = 113.5 ± 0.5°C, Per = 34.0 ± 0.5 kg/cm2 abs. The temperature dependence of the specific heat C' x-o of the boiling liquid along

the bottom boundary curve is presented by the equation C' x-o = 0.2761 + 6.63 . 10-4 • t + 1 .5 . 10-6 • t2 (4)

where t is expressed in ° C; C' x-o - in kcal/kg ° K. The heat of vapourization has been calculated according to the Clapeyron-Clausius

equation with account of the equations (1-3). The following values were adopted when calculating the enthalpy and entropy :

i' = 100.00 kcal/kg, S' = 1.00 kcal/kg ° K at 0° C.

The enthalpy and entropy of the boiling liquid was calculated according to the ther­modynamic equations

t p •I = 100.00 + J C'x-o dt + J V' d P i '

oo Pt-oo T

I C'x-o d T

S' = 1.00 + T 273.15

in which the temperature dependence C' x-o and V' was expressed by the equations (2) and (4).

The enthalpy and entropy of the dry saturated vapour was calculated respectively as follows :

i" = i' + r , r

S" = S' + T · Applying the obtained values i" anti S" and the derived equation of state (3), the

enthalpy and entropy of the superheated vapour was calculated by the isothermal increments according to the dependences

where

where

308

i P, T = i" T + LI i T V p, T [ ( p )]

. r2 a r LI (PV)T LI 1 T = LI UT + LI (PV)T =

426.94 / - -a T - v av + 426,94 '

s P, T = S" T + LI s T Vp, T

V"T

A sT = 42�.94 / (� �)v a v.

V"T

11-17

As the specific heat C' x�o of the boiling liquid was measured only up to 80°, the values i and S were calculated above this temperature by extrapolation of Gp, adopting the dependence C p linear from the temperature.

The thermodynamic properties of the saturated vapour of the azeotropic mixture are given in the table and the zone of superheated vapour is given in the i, S-diagram.

Table of saturated vapour of azeotropic mixture C2HF4Cl and C4F8

kg/ cn12 1 1 kg

" kg kcal kcal = kcal kcal kcal

t, ° C P , abs. V' - V", - e', 1 e , T i' - i" - r - S' -- S" --'kg kg

'kg

' kg ' kg 'kg°K

' kgoK

-20 o.8I I o.6373s r s r .9 r .s690 o.oo6r 79 94·S9 126.39 3r .80 o.979s r . r os2 -r s 0.980 0.6443 I3S·9 r .552 r 0.007361 9S·92 r27.s9 3r .67 0.9847 r . ro74 -IO r .r83 o.6sr s n4.1 I ·S349 0.008764 9p6 l28.7ss 3 1 .so 0.9898 I .I09S - s 1.424 o.6s90 95.80 l 51 74 0.01044 98.62 l29.89s 3 1 .27 0.9949 I . I I lS

0 1 . 7 1 0 o.6668 80.60 1 .4997 o.or24r I00.00 131 .00 3 r .oo I .0000 I . I I34 5 2 .04s 0.6749 68.oo r .4817 0.0147 1 101.39 l32.07s 30.68 I .ooso I . I I S3

I O 2 .439 0.6833 S?-23 r .4634 0.01747 ro2.8os 1 33.12 30.32 I .OIOO I . I I 7 1 r s 2.89s 0.6921 48.28 r .4449 0.02071 ro4.23s 134.14 29·9°S I .OlSO I . I I 88 20 3.421 0.7012 40.87 r .426os 0.02447 l OS.68 135.13 29.45 I.0200 r . 1 204 2s 4.024 0.7 108 34.65 l .40695 0.02886 ro7.15 136.09 28.94 1 .0249 I . 1 220 30 4.709 0.7207 29.48 1.3876 0.03392 ro8.64 137.03 28.39 1 .02985 r . 1 23s 3S 5-483 0.73105 25.16 1 .3679 0.03975 I I0.145 137·935 27.79 1 .0347 1 . 1 249 40 6.349 0.7418 2 r .s4 1 .3480 0.046425 I I I .68 138.82 27.14 r .0396 1 . 1263 45 7.310 o.7s31 l8.s2 1 .3278 0.05400 u3.23 139.68 26.45 1 . 0444 r .1276 50 8.370 0.7649 16.01 1 .3074 0.06242 I I4.80 140.51 2s.71 1 .0493 1 .1289 55 9.528 0.7772 13.96 1 .28665 0.07163 u6.40 141.32 24.92 1 .0542 1 .13015 60 ro.78 0.7901 1 2 .28 1 .26s65 0.08143 u8.02 142 . I I 24.09 1 .0590 1 .1313 6s 12.13 0.8036 10.91 1 .2444 0.09166 u9.66 142.87 23.21 r .0638 1 .1324 70 13.56 0.8178 9.835 1 .2228 o.ro17 1 2 1 .33 143.61 22.28 r .0686 I .l33S 7S 15.06 0.8326 8.975 I .2010 0.1 I I4 123.025 144.325 2 r.31 r .0734 1 .1346 So 16.63 0.84825 8.270 1 . 1 789 0.1209 124.73 1 45.02 20.29 1 .0782 1.1356

1. l<tal/lJ Lil:! Ull u� IJ2l) '2 VNI Khl l.S-dt.agrom

148 aztatn:pe mLxtun! Of /f?OllS Ill omt C318

146 144 142 142 J.10 l40

13& IJ6 134 1:12

130 130 128

128

128 ... l2fi Im 11.i U11 U20 lZ WD U1' l.llO I.Ml too l,m lliO � \DD S.kcol./�1t Fig. r. i , s - diagram of the azeotrope mixture of freons 124 and C 318.

309

11-9

Joule-Thomson Effect in Hydrogen-Methane Mixtures at Temperatures Between -35 and +40°C

Effet Joule-Thomson dans les melanges d'hydrogene et de methane a des tem­peratures de -35° a +40° C

R. AYBER Forschungsgruppe fiir Wiirme- und Kiiltetechnik im Max-Planck-Institut fi.ir Striimungsforschung, Giittingen, Germany

SOMMA/RE. On a determine experimentalement l'effet Joule-Thomson de 4 melanges d'hydrogene et de methane a des temperatures de -35° a +40°C et a des pressions atteig­nant 90 atm.

Les melanges sont prepares dans un recipient a gaz rendu etanche par de l' eau puis com­primes dans un compresseur utilisant aussi de l'eau comme lubrifiant pour les pistons. Le melange de gaz comprime est amene a la temperature desiree par passage a travers un echan­geur de chaleur et un bain thermostatique, puis detendu a la pression atmospherique dans un robinet specialement confu a cet eff et. Le gaz detendu est ensuite envoye a tr avers un compteur a gaz et il atteint finalement le recipient a gaz OU il est pret a repeter le cycle.

L'appareil a d'abord ete essaye avec de l'air. Les valeurs de l'effet Joule-Thomson, ob­servees de cette f afon concordaient assez bien avec celles calculi es a partir des tables publiees par Baehr et Schwier, l'ecart moyen etant de l'ordre de ± 0,63%.

Les resultats sont present es graphiquement sous f orme de courbes isothermes de l' elf et Joule-Thomson par pourcentage de molecules d'hydrogene, pour des pressions de 90, 70, 60 et 50 atm.

INTRODUCTION

Experimental data in connection with the volumetric and thermodynamic properties of gaseous mixtures are still relatively scarce. A satisfactory method for determining these properties would be the measurement of the temperature change that a gas undergoes when it is adiabatically throttled under conditions of constant enthalpy. This temperature change

is referred to as the Joule-Thomson coefficient µ = ( �;) h when the pressure drop

during the throttling process is infinitely small and as the Joule-Thomson effect (LI T)h when the pressure drop is finite and the expansion carried out until atmospheric pressure.

No adequate theoretical method to determine the Joule-Thomson effect of a gaseous mixture has been so far developed. Perry and Herrmann [1] have computed the Joule­Thomson coefficient of mixtures of methane and nitrogen using a simplified form of the Beattie-Bridgeman equation of state. Edmister [2] has derived an empirical equation that could be used in determining the Joule-Thomson coefficient of certain hydrocarbon mix­tures. Koeppe [3] has calculated the Joule-Thomson coefficient of a number of pure gases and mixtures at zero pressure as a function of the second virial coefficient. However, all these suggested methods have very limited application, thus proving the necessity for further experimental investigation.

Sage and co-workers, during their extensive investigations on hydrocarbons, have also measured the Joule-Thomson coefficient of mixtures of methane-ethane [ 4], methane -

311

Il-9

n-butane [5], and methane-propane [6]. Koeppe [7] has published certain measure­ments on the Joule-Thomson effect of nitrogen-argon, argon-oxygen, nitrogen-hydrogen and nitrogen-helium mixtures.

The present work covers the measurement of the Joule-Thomson effect of four mix­tures of hydrogen-methane at temperatures between -35 and +40°C and pressures up to 90 atm. * This combination is of special interest, since the first component of the mix­ture is an element having light molecular weight and considerable quantum effect, while the second component is an element with nearly spherical molecules having an intermolec­ular potential that obeys closely the Lennard-Jones 6-12 equation.

APPARATUS

Fig. 1 shows a general layout of the apparatus as employed for the measurements. The mixture is prepared at approximately the desired composition' in the water sealed gashold­er 1 and then drawn into the compressor 2. Considering the solubility of methane in mineral lubricating oils, a compressor was chosen in which the pistons are lubricated with water that is introduced into the incoming gases before the suction valve of the first stage cylinder. The compressed gas flows first through the water trap 3 that is provided with a let-off valve 4. The gas is then passed through the silica-gel dryer 5, whose proportions are such that the gas leaving it has a dewpoint lower than -35° C. The dried gas reaches then the thermostat bath 7 by flowing first through the heat exchanger 6. The final condi­tioning of the gas is accomplished by passing it through the coil 8 placed in the thermostat bath. The bath is filled with methyl alcohol and contains a second coil 9 which serves as the evaporator of a Freon-22 refrigerating unit. In addition, a manually adjustable heater JO and a temperature control heater 11 that is actuated through a mercury contact thermome-

70

7�

I Gasholder 2 Compressor 3 Water trap 4 Let-off valve 5 Dryer

Fig.

6 Heat exchanger 7 Thermostat bath

2 \

r. General Layout of the Apparatus 8 Cooling coil 9 Evaporator coil

ID Adjustable heater II Control heater

0--1 - I

I2 Mercury contact thermometer IJ Throttling valve I4 Gas meter I 5 By-pass valve

* The physical atmosphere (r atm = 760 mm Hg) is used as the pressure unit in the evaluation of the measurements.

312

11-9

ter 12 are placed in the thermostat bath. By flowing through the conditioning coil the gas reaches thermal equilibrium with the thermostat bath and is ready for throttling. This takes place in a specially constructed throttling valve 13 which resembles in its essential features that used by Johnston [8]. To avoid head leak to the expanded gas, material with low heat conduction is used in the construction of the valve. The valve seat is made ofplexiglas, the valve stem ofpolyamid and the valve body of V2A steel. The expanded gas flows through the heat exchanger, where it cools the incoming high pressure gas ; then passes through a gas meter 14 and returns to the gasholder, where it is ready to repeat the cycle.

The flow rate of the gas is controlled by means of the by-pass valve 15. The flow rate through the conditioning coil can be varied, by regulating the by-pass valve and simul­taneously readjusting the opening of the throttle valve, thus permitting the gas pressure before the throttling valve to remain constant irrespective of the flow rate. This possibility constitutes a considerable advantage in favour of the valve throttling method over thrott­ling through a porous plug.

The temperatures before and after the throttling valve are measured with platinum resistance thermometers that are placed in direct contact with the flowing gas. A Diessel­horst potentiometer is used to measure the resistance of the platinum thermometers.

100 kp/cm2 and 250 kp/cm2 precision gages, with 250 mm dial, are used to determine the gas pressure before the throttling valve. The gas pressures after the throttling valve do not exceed 1 .25 kp/cm2 and are recorded by a 4 kp/cm2 precision gage with 100 mm dial.

The composition of the mixture is determined with a Beckman Gas Chromatograph. Three samples are analysed for each mixture, the samples being taken at regular intervals during the runs.

PRECISION OF MEASUREMENTS The pressures at the throttling valve were controlled manually by means of the by-pass

valve 15. This method was found to be satisfactory since no noticeable fluctuations were observed on the manometers, once the equilibrium state was attained. Thermoelements placed on specific points in the apparatus were connected to a recording potentiometer and readings for the Joule-Thomson effect were taken only after these temperatures showed a constant course. The pressure at the throttling valve was set to an even value on the manometer and the error thus introduced would directly be equivalent to the instrument's accuracy, which in this case was ± 0,4%. The gas temperature before and after throttling could be measured with an accuracy of ± 0.03 ° C. Taking into consideration the systematic uncertainties that may be involved in the measurements and also allowing for the error brought in determining the composition of the gas mixture, the overall accuracy of the results reported in this paper may be estimated as being 3 % .

It is rather difficult to estimate correctly the overall accuracy of the final results in an experiment of this nature. The conditions for an adiabatic throttling at constant enthalpy are fulfilled when : a) The changes in velocity and elevation between both sides of the throttling valve are

small enough, thus causing no appreciable energy utilization that could result in a decrease of enthalpy.

b) The heat leakage between the valve body and its surroundings is negligible. c) The heat flow between the gases on both sides of the throttling valve and the "jet

kinetic effect" due to the high gas velocity attained at the restriction through the valve have no noticeable effect on the temperature drop measured.

All the points listed above are dependent on the flow rate of the gas through the thrott­ling valve. As a check, runs were made with air at five various flow rates varying between 22.80 and 10.20 m3/h. The observed values of the Joule-Thomson effect for a throttling from 107 atm and 17.48°C were 22.61 ; 22.68; 22.71 and 22.67 degrees, indicating a highest deviation of 0.44 %. The corresponding value computed from the tables published by Baehr and Schwier [9] was found to be 22.64 degrees. This result shows that the possible source of errors, pointed out above, have negligible influence on the measured temperatures.

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Further, three isotherms of the Joule-Thomson effect in air at temperatures -35. 75 ; -23.95 and + 40.22° C were measured and the results are presented in Fig. 2. The values computed from Baehr and Schwier's tables and those from the tables of U. S. Bureau of Standards [10] are also depicted on the curves in Fig. 2. 28 values of the Joule-Thomson effect at equal pressure intervals were read out from the smoothed curves and compared with the corresponding values calculated from the tables ofBaehr and Schwier. The average deviation was found to be ± 0.63 % and only 6 of the observed values showed a deviation greater than 1 %· The largest percentage deviation of -2.45 occurs by a very small temperature drop, i. e. 5.72 °C .

..... t.J � � c:

• Observed i . Baehr

o N B S.

� 2,._ _ _ �-+-��--+��---r-+-��+--��-+-�---+ g � "' :; �

Pressure p1

Fig. 2. Joule-Thomson Effect in Air

PURITY OF GASES

The gases utilized were obtained from Gesellschaft fiir Linde's Eismaschinen in Holl­riegelskreuth. The hydrogen was of very high purity containing only 0.6 ppm of nitrogen. The methane, on the other hand, contained an average of 1 mole % of ethylene and 0.3 % of ethane. To study the effect of these impurities on the Joule-Thomson effect of methane three isotherms at temperatures of 248.54 ; 268 and 311.86° K were measured. A com­parison between 24 values obtained from smoothed curves and those evaluted from log p vs. H diagram of Keesom and coworkers [11] showed an average deviation of ± 1 .12%. This figure being well below the estimated overall accuracy of the investigation, the meth­ane available was used without any purification. This result was to be expected, since according to Roebuck and Osterberg [12] the presence of impurities having a critical temperature higher than that of the main component have no appreciable effect on the Joule-Thomson coefficient of the main component.

314

11-9

RESULTS

The measured values of the Joule-Thomson effect for a hydrogen-methane mixture containing 14.70 mole % of hydrogen are presented as isotherms in Fig. 3. The increase in the Joule-Thomson effect with the pressure is very nearly linear. The slope of the isotherms passes through a slight maximum at low pressures, this being more pronounced in the case of low temperature isotherms. Mixtures containing 28.80 ; 48.15 and 61.10 mole % of hydrogen were also investigated within the same range of temperature and pres­sure. In general, the experimental results for these three mixtures are similar to those indicated in Fig. 3 except that the gradient of the isotherms decreases as the mole fraction of hydrogen in the mixture is increased. Further, it appears that the isotherms for the mixtures rich in hydrogen tend to flatten out sooner, indicating that they would pass through a maximum at relatively lower pressures than those required for mixtures rich in methane. This maximum will correspond to a point on the inversion curve, beyond which a differential throttling would cause the gas temperature to rise. Unfortunately,the present investigation could not be carried out to pressures enabling the location of this maximum.

t; � vv"---+----+- r-;r-+--,�+T----1 tj 5 § � .§�v1---+---H'-+-r-+--P----l---i -s:

Fig. 3. Joule-Thomson Effect in Hydrogen-Methane Mixture Containing 14.70 Mole % of Hydrogen

The variation in Joule-Thomson effect with the composition for a series of temperatures at pressures of 90 ; 70 ; 60 and 50 atm are presented in Fig. 4, 5 and 6. The values corres­ponding to pure hydrogen and methane were taken from [10] and [11] respectively. It is clearly seen that an increase in the concentration of hydrogen causes first a rapid decrease in the Joule-Thomson effect, this decrease becoming less pronounced as the mole fraction of hydrogen increases. It can be concluded, that the addition of small quantities of hydro­gen exerts a stronger influence upon the Joule-Thomson effect of the mixture as that caused by the addition of small quantities of methane. It is also clearly seen, that for the mixture investigated a linear combination of the Joule-Thomson effect of the pure gases

315

II-9

4

___ T _ _ _

.. ,,._� __ ,__ p1 -BOatm -�1

IJJ 0 Mole % Hydrogen Content

5

Fig. 4. Variation in Joule-Thomson Effect with Composition in Hydrogen-Methane Mixtures

Fig. 5. Variation in Joule-Thomson Effect with Composition in Hydrogen-Methane Mixtures

Fig. 6_ Variation in Joule-Thomson Effect with Composition in Hydrogen-Methane Mixtures

would, by no means, be sufficient to give the true value of the Joule-Thomson effect of the mixture. This fact can, as a first consideration, be attributed to the relatively large difference between the critical temperatures of hydrogen and methane.

316

11-9

REFERENCES

r . ]. H. Perry, C. V. Herrmann, ]our. Phys. Chem. 39 (1935) I I89. 2. W. C. Edmister, Pet. Refiner 28 (1949) 128. 3. W. Koeppe, Z. filr ang. Phys. II (1959) 432. 4. R. A. Budenholzer, B. H. Sage, W. N. Lacey, Ind. Eng. Chem. 31 (1939) 1 288. 5. R. A. Budenholzer, B. H. Sage, W. N. Lacey, Ind. Eng. Chem. 32 (1940) 384. 6. R. A. Budenholzer, D. F. Botkin, B. H. Sage, W. N. Lacey, Ind. Eng. Chem. 34 (1942) 878. 7 . W. Koeppe, Kaltetechnik II (1959) 363. 8. H. L. Johnson, ]our. Am. Ch. Soc. 68 (1946) 2362. 9. H. D. Baehr, K. Schwier, Die Thermodynamischen Eigenschaften der Luft. Springer-Verlag

Berlin (1961). 10. U. S. Department of Commerce National Bureau of Standards, Tables of Thermal Propertie,

of Gases, Circular 564. r r . W. H. Keesom, A. Bijl, L. A. ]. Monte, Comm. Kamerlingh Onnes Lab. Univ. Leiden Suppl.

r o8b (1955). 1 2 . ]. R. Roebuck, H. Osterberg, ]our. Am. Soc. 60 (1938) 340.

DISCUSSION

Mr. Morsy, Egypt : The usual dimension of the Joule-Thomson effect is deg/unit pressure and not degrees only.

Has the author attempted to plot µ against p (density) and to try to correlate the data with any equation of state ?

R. Ayber, Germany : It is true the dimension of the Joule-Thomson coefficient µ is in deg/unit pressure. However, we have measured the Joule-Thomson effect (LIT) h by carrying out the throttling until atmospheric pressure, here (LIT) h represents a tempera­ture drop and its dimension is in deg.

No attempt was made to correlate the experimental data with an equation of state, because no equation of state is known to the author that can be satisfactorily applied to a hydrogen-hydrocarbon mixture. Due to the lack of a suitable equation of state the density p could not be determined.

P. Petit, France : Avez-vous cherche des correlations faisant intervenir les pressions partielles ?

R. Ayber, Germany : Quand i1 s'agit des pressions partielles dans un melange, dont la pression totale p t corresponde a I atm, la concentration Xr est directement proportionnelle a la pression Pr suivant la relation Xr = PrlP t. Le rapport entre (LI T) h et Pr sera le meme que celui represente dans les Figs. 4--6. Si le melange est considere sous pression, alors l' evaluation de la pression partielle devient difficile sans la connaissance d'une equation d'etat ou bien des facteurs de compressibilite qui permettent a calculer la densite molaire du gaz pur et du melange.

P. J. Haringhuizen, Netherlands : The author has made an important contribution to the thermodynamics of gas mixtures.

As was stressed during the meeting of Commission I at Eindhoven in 1960, one of the outstanding questions is how far the enthalpy of a mixture of two gases can be approximat­ed by a linear combination of the enthalpies of both gases at the total pressure of the mixture, or by that of the enthalpies at their partial pressures.

Now using the data shown in the curves of Figs. 4 and 5, a rough calculation gives the impression that the Joule-Thomson effect of the mixture is not so far away from the temperature drop after expansion which would occur if the CH4and the H2 in it were expanded from their respective partial pressures.

In this connection reference should be made to the paper submitted to Commission 1 by Knoester, Taconis and Beenakker. This gives measurements made in the Leiden Laboratory on the heat of mixing for the gas system H2 and N 2•

From their results it can be deduced that in many cases (i. e. for the mixture 3 : 1 H2 and N2) the value of the enthalpy comes also rather close to that of the linear combination of the enthalpies of both gases at their partial pressures.

317

Il-9

R. Ayber, Germany : This is an important point which requires further investigation. It is possible to determine the partial pressure of the component by means of the relation x1 = p1f p t· This would, of course, not be equivalent to the exact partial pressure of the component in the mixture, for which the ideal gas law is no more applicable. Assuming that p1 from x1 = p1f P t will be used, the possibility of applying a linear combination of (LIT)cH4 and (LIT) H2 (these are the temperature drops of the components when they are throttled from their respective partial pressures) can be investigated by comparing the isotherms of CH4 and H2 with that of the CH4-H2 mixture in a (LIT) h vs.p diagram. A geometrical analysis of this diagram gives the conditions for a linear combination. They are the following :

1 . The isotherms must be straight lines, i. e. (LIT) h a linear function ofp

2. The isotherm of the mixture that lies between the isotherms of CH4 and H2 must have a slope M that is defined through :

X1 (LIT)cH4 + X2 (LIT) H2 M =

P t

In my opinion it is possible that, for a mixture whose isotherms resemble those present­ed in Fig. 3, a linear combination might give values that are in good agreement with the ex­perimental values. For mixtures containing higher Mol % of H2 the isotherms have a stronger curvature, and in this case it is doubtful if a linear combination could be used. The same applies to the isotherms in the low pressure region. In conclusion I like to add that the linear combination rule should only be used with caution.

The author is indebted to Dr. Haringhuizen for his reference to the paper of Messrs. Knoester, Taconis and Beenakker.

318

On the Thermodynamics of the Cold-Air Cycle with Throttling

Thermodynamique du cycle d'air froid avec etranglement

Prof. Dr.-Ing. HANS DIETER BAEHR Technische Hochschule, Braunschweig, Germany

11-1 1

SOMMAIRE. A l' aide de nouvelles tables tres precises des proprietes thermodynamiques de l' air, on etudie le processus cyclique de la machine a air froid avec etranglement en supposant que la compression isotherme reversible a lieu a la temperature ambiante. D' apres la puissance frigorifique, on ameliore les limites de pression du processus de sorte que son coefficient de rendement soit maximal. L'analyse exergetique du cycle montre que /es pertes par exemple, provoquees par l' etranglement, se reduisent lorsque la temperature de refroidissement est basse, de cette f afon le rendement exergetique du cycle est el eve. Le diagramme du courant d' exergie montre la quantite des pertes thermodynamiques et leur repartition sur diff erentes parties de I' installation.

The cyclic process of the cold-air refrigerator with throttling was already used by Carl v. Linde in his method for liquefaction of air as a high pressure cycle for primary cooling. A cold air cycle with throttling is assumed to be advantageous for refrigeration at low temperatures (i. e. temperatures about and below -100°C) and especially for small refrigerating capacities.

The disadvantage of the low coefficient of performance is compensated by the advan­tages of rather small expenses for construction and mainly by favourable conditions for starting and shutting off. K. Linge [2] discussed the advantages of cold-air cyclic processes with trottling firstly and with expansion and work output secondly and he compared both of them.

The cycle of the cold-air machine with throttling is based on the properties of the real gas air since the pressure dependence of the enthalpy is of fundamental importance for the production of cold. The thermodynamic properties of air have recently been investigated again up to high pressures. Exact and comprehensive tables and diagrams have been published by Baehr and Schwier [l]. With the aid of these data the cycle will be investiga­ted in detail. Hereby the exergy-function (availability-function)

e = h - hu - Tu (s - Su) (1)

is used for the thermodynamic rating of the cyclic process and parts of it (h = enthalpy, s = entropy, Tu = temperature of the surroundings, index "u" stands for the equilibrium with the surroundings). The exergy-function weights different forms of energy by stating what amount of energy can be transformed to work under the restrictions of the second law of thermodynamics. Thus the exergy of heat, for instance, is given by the product of heat and the Carnot-factor. The exergy of a streaming medium in a steady-flow process is given by equ. (1), compare also [3] for further details.

Fig. 1 schematically shows a refrigerating plant which uses the process with throttling. The counter-flow heat exchanger forms the main element of the plant. In that heat ex­changer the high pressure air, before it is throttled down to the low pressure, is cooled down by the low pressure air, which in turn warms up. The cyclic process is shown in the h, s-diagram Fig. 2 from which the specific refrigerating capacity qo and the work input Wt can be found as distances. A reversible isothermal compression of the air at ambient temperature T = Tu has been assumed here to form the ideal process.

REFRIGERATING CAPACITY AND COEFFICIENT OF PERFORMANCE

The following well known equation for the refrigerating capacity, compare [3, p. 312] f. i., is valid

(2)

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Il-11

q

Throttle ­Valve

5

Fig. r . Schematic flow diagram of a cold air refrigeration plant with throttling.

Cooler

h

Tangen t with slope (ah ), = T = Tu as p

Fig. 2. Cold-air cycle with throttling in the h, s-diagram.

Here it is used that h4 = h3 holds for throttling and that the heat exchanged in the counter-flow heat exchanger is given by

q G = h2 - h3 = h1 - h5. (3)

It can be seen that the refrigerating capacity qo is the greater the more the enthalpy of the air changes with pressure at the temperature of the isothermal compression. The maxi­mum refrigerating capacity

(4)

320

II-11

is obtained for a given initial pressure p 1 when one compresses up to that pressure p2 which is equal to the inversion pressure Pi (Tu) of the Joule-Thomson-Effect. For Tu = 288.15°K (tu = 15°C) the inversion pressure is Pi = 432 bar and one gets (qo)max = 49.0 kJ/kg for P1 = Po = 1 bar.

The work necessary for the reversible isothermal compression at T = Tu can be calculated as follows

(5)

Since the compression takes place at ambient temperatures the work input Wt equals the increase of the exergy of the compressed air. While compressing the heat

q = Tu (s2 - s1) (6)

has to be transferred to the surroundings. This energy is thermodynamically useless ; its exergy is nought. The coefficient of performance of the process

qo hi-hz e = · ·- · = -- = f (P, Po)

Wt e2 -e1 (7)

only depends on the choice of both thepressuresp1 = Po andp2 = p. It is only fixed by the properties of the isotherme T = Tu and it is independent of the temperature at which the cold is generated. It is possible to ask for those two pressures p1 = Po and p2 = p at each refrigerating capacity qo for which the work input gives a minimum i. e. the coefficient of performance shows a maximum. In Fig. 3 the calculated optimal pressures Po and p are plotted versus qo. The coefficient of performance emax which is obtainable at the most decreases with increasing refrigerating capacity q o· At the same time the ratio of the pres­sures Pf Po = p2/p1 increases very rapidly. For that reason in practice one is satisfied with smaller refrigerating capacities in order to obtain a favourable coefficient of performance s and a favourable ratio of pressures f. i. Pf Po between 4 and 5.

:�� ,-- -----r ·- -r--· -;r:: 400 I I-

0,5

0,4

0,2

0,1

10 20 30 0

40 kJ/kg 50 Qo --

Fig. 3. Optimized high pressure p and low pressure Po and coefficient of performance emax as functions of the specific refrigeration capacity q0. (Valid for isothermal compression at Tu = 288.r 5 °K.)

The calculations which have been done here, however, are only applicable to such cyclic processes for which T3 � T5 holds. Otherwise there is no positive temperature difference in the counter-flow heat exchanger for the heat transfer between high pressure and low pressure air. The lowest temperature T5 for refrigeration which results out of this restriction lies between 140°K and 150° K.

THE EXERGETIC EFFICIENCY

The exergy e as given by equation (1) is used further on to analyse the process thermo­dynamically. In order to be able to use the exergy data given in the table [1] the ambient

321

II-1 1

state (subscript u) which has an exergy of e = o is chosen to be Pu = 1 bar and Tu = 288.15°K according to [I]. We define the exergetic efficiency of the cycle as

exergy of the cold produced C = exergy of the work input

eqo

Wt (8)

Although the process generates cold within the temperature interval from T4 to T5 the scope of the process is to be seen to generate cold at the constant temperature T5• Then the exergy of the generated cold is given by

eqo = qo Tu - T5

T5 (9)

Using this one gets

C = <J_� Tu - T5

Wt T5 = B (10)

The exergetic efficiency increases with decreasing temperature T5• For that reason the cyclic process with throttling is only suitable for refrigeration at low temperatures since C becomes very small at higher temperatures.

For the following calculations a cyclic process is chosen which operates between the pressure levels Po = 70 bar and p = 280 bar. The exergetic efficiency of that process is (288.15 °K ) C = 0.250

Ts - 1 • (IO a)

It is shown as a function of Ts in Fig. 4. The lowest temperature obtainable becomes (T5)min = 148°K.

The width of the temperature interval T5 - T4 in which the refrigerating capacity of qo = 28.9 kJ /kg can be generated decreases with decreasing temperature since the specific heat c p rapidly increases as the temperature decreases.

Fig. 4.

322

0,25 25

l;' � OK qo

0,20 8 20

t 0,15 6 15

(T5 )min1 � I I �

0, 10 4 10

0,05 2 5

0 0 -'-��-'-'---��-'-���'--��-'-��-' 0 725 150 175 200 225 ° K 250

T5 Exergetic efficiency C according to equation (1oa), ratio qG/q0, and temperature inter· val T • - T,, in which cold is produced as a function of the temperature T •. (Valid for Po = 70 bar, p = 280 bar.)

11-11

With decreasing temperature T5 the heat q a increases. This heat q a in transmitted in the heat exchanger and is given by equation (3). Fig. 4 shows the ratio q afqo as a function of T5• The enlargement of the heat exchanger is the price which has to be paid in order to generate the refrigerating capacity qo at lower and lower temperatures and in order to increase the exergetic efficiency C.

THE EXERGY LOSSES

Since the exergetic efficiency C is very small, even at low temperatures T6, large losses of exergy take place within the different parts of the cycle. These are :

1. The exergy loss evn of the throttling. 2. The exergy loss e v a in the heat exchanger caused by finite temperature differences for

heat transfer. 3. The exergy loss e v K caused by the generation of cold within the temperature interval

from T4 to T5 below the desired temperature T5•

60 evo Wt

40 // /

/:'. � 20 evK

Wt eqo

0 Wt 125 150 175 200 225 °K 250

T5 Fig. 5. Exergy losses evo according to equation (r4), evn according to equation (II), ev x

according to equation (r5), and exergy of the generated cold eq0 according to equation (9). All quantities are related to the exergy input Wt· (Valid for Po = 70 bar, p = 280 bar.)

These energy losses are shown in dimensionless form in Fig. 5 as a function of the tem­perature T5• The single amounts are related to the total exergy input Wt = e2 - e1• As can be seen from Fig. 5 the increase of the exergetic efficiency with decreasing tem­peratures T5 has its main reason in the fact that the exergy loss e vn of the throttling is strongly decreased at low temperatures.

THE EXERGY LOSS OF THE THROTTLING-PROCESS

The throttling process causes the greatest exergy loss of the cycle as is shown in Fig. 5. It is calculated by

evn = ea - e4 = Tu (s4 - sa) (11) since h4 = ha holds for throttling. Fig. 6 shows the exergy loss evn together with the tem­perature decrease T3 - T4 which the air experiences during throttling. The decrease of the exergy loss with decreasing temperature T5 can be explained by considering the

323

11-1 1

shape of the isobars p and Po in the h, s-diagram, Fig. 2. Those isobars approach each other more and more the lower the temperature (approach to the wet vapour region). So the governing entropy difference s4 - s3 in equation (1 1) decreases as temperature decreases.

As shown in Fig. 6 it is not important that the air is cooled down during throttling. Throttling only is the simplest method to reduce the pressure in order to get from the isobar p to the isobar Po· Especially at low temperatures the temperature decrease due to throttling is small and expecially in that region the exergy loss of the throttling is decreased.

�� Tr

r---- - - -125 �--r-100 40 vo = er e4

OK

t 75 30

150 175 200 Ts -

225 °K 250

Fig. 6 Temperature drop T 3 - T 4 and exergy loss evo of the throttling process as a function of 7'5. (Valid for Po = 70 bar, p = 280 bar.)

THE EXERGY LOSS OF THE HEAT EXCHANGER

It is well known that any heat transfer from a temperature TH to a temperature TL < TH causes a 'local' exergy loss whose amount can be calculated, compare [3, p. 190]

TH - TL de = dq

T T Tu.

H " L (12)

This exergy loss is not only proportional to the temperature difference TH - TL it also increases with lowering the temperature level for heat transfer. In Fig. 7 the temperature difference TH - TL between high pressure and low pressure air and the loss factor

de dq -

(13)

are plotted versus the temperature TL of the low pressure air. The maximum of the loss factor accordingly lies at lower temperatures than the maximum of the temperature differ­ence TH - TL.

The whole exergy loss occurring in the heat exchanger is

e v a = e2 + e5 - (e1 + e3) = Tu [s1 - s5 - (s2 - s3)] (14) since the gain of exergy of the high pressure air is smaller than the exergy loss of the low pressure air. eva increases with decreasing temperature T5 because the temperature difference between high pressure and low pressure air is especially large at low tern-

324

11-1 1

peratures, before it becomes nought at T = (Ts)min, see Fig. 7. The increase of e v G with decreasing temperatures, however, is largely exceeded by the decrease of the exergy loss e v n of the throttling-process. The exergetic efficiency of the cycle, therefore, increases as the temperature Ts decreases.

t 20 20 eyG , OK kJ I

� kg

�, � 15 15 I i-.:; � 2=

0,2 10

0, 1 5 5

0 0 0 150 175 200 225 250 275 OK 300

T5 ' TL -Fig. 7 . Temperature difference TH - TL between high pressure and low pressure air and local

loss factor (TH - TL)· T u/(THTL)according to equation (13) as functions of the temper­ature TL of the low pressure air in the counter-flow heat exchanger. Exergy loss ev G according to equation ( 14) as a function of the temperature T5• (Valid for Po = 70 bar, p = 280 bar.)

THE EXERGY LOSS OF THE COOLER

The refrigerating capacity qo is generated at varying temperature within the interval from T4 to Ts. Thereby the exergy of the air is decreased by the amount

- LI e = e4 - es = h4 - h5 - Tu (s4 - ss) =

= - qo (1 - Tu �: �; ). Within the temperature interval from T4 to T5 the specific heat of air can be assumed to be constant as a good approximation. Then one gets

s4 - s5 Cp Zn T4/Ts Zn T5 / T4 --- = -�- - ----- = ------·-h4 - h5 c p (T4 - T5) T5 - T4 Tm

Tm is the thermodynamic average temperature at which cold is produced. Since we stated the scope of the process to generate cold at the constant temperature Ts > Tm, an exergy loss

e v K = e4 - es - e qo = qoTu (r�- - A) = qoTu(�s ��: �J (15)

is experienced in the cooler. Usually the ratio T5 - T4 is so small that the evaluation of Zn in equation (15) into a

series rapidly converges. One gets

_ qo Ts � T4 [ _ l _ !:s - T4 _l_ (Ts - T4) 2 ] e v K - 2 Tu

T T l

3 T +

6 T • • • 5 4 4 4 (16)

This exergy loss is very small compared with the exergy losses of the remaining parts of the process as can be seen from Fig. 5.

325

II-1 1

AN EXERGY FLOW DIAGRAM

It is very instructive to compare an exergyflow diagram with the well known energy flow diagram (Sankey-Diagram), compare [3, p. 127). In opposition to the energy flow diagram the exergy flow diagram shows the thermodynamic losses based on the assessment of energies by the second law of thermodynamics. Fig. 8 shows an energy and an exergy flow diagram for the cycle in which a temperature T5 = 170° K was chosen for the generation

. . . . . . :: ' :'. <�.:.-:· · · : . ·

. ·: . . . .

5

E = o, 2so

Energy-Flow

Counter-Flow Heat-Exchanger ·.<�� : ·· · . . .. .. 1l

Exergy -Flow Fig. 8. Energy-flow diagram (left) and exergy-flow diagram (right) for the cold-air cycle with

throttling according to table i .

of the refrigerating capacity. For the sake of a better illustration of the converted exergies, in the exergy flow diagram the exergy differences of the air e* = e - 300 kJ /kg have been plotted instead of the exergies themselves. The data on which Fig. 8 is based are given in table I.

It is worthwhile noticing that at temperatures below ambient temperature Tu the direc­tions of the heat flows and their accompanying exergy flows are opposite. That means f. i. that an amount of heat qo absorbed from the chilling chamber corresponds to an amount of exergy which is transferred to the chilling chamber.

The main difference between the energy and the exergy flow diagram lies in the distri­bution and in the rating of the losses. In the energy flow diagram throttling and heat transfer seem to be processes without losses. All the losses can only be found totally in the amount of heat q which is transferred to the surroundings during compression. In the

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Table 1. Thermodynamic properties of air and characteristic data of a cold-air cycle with throttling.

state p T h e

bar O K kJ/kg kJ/kg

1 70 288.15 272.3 349.8

2 280 288.15 243.4 465.2

3 280 185.7 84.9 504.5

4 70 158.3 84.9 422. 1

5 70 170.0 1 13.8 400.3

Compression Wt = 115.4 kJ/kg

q = 144.3 kJ/kg

Counter-flow heat exchanger : q a 158.5 kJ/kg

e v a = 1 1 .2 kJ/kg

Throttling : evn 82.4 kJ/kg

Refrigerating capacity : qo 28.9 kJ/kg

eqo 20.1 kJ/kg

e v x = 1 .7 kJ/kg

Coefficient of performance : e 0.250

Exergetic efficiency : c 0.174

exergy flow diagram that amount of heat q does not appear at all since it is thermodynami­cally useless energy which posesses no exergy. The losses of the irreversible parts of the cyclic process where exergy is destroyed, however, are clearly noticeable.

The exergy flow diagram which is shown here does not contain all the losses which occur while running a real cold-air cycle with throttling. Here especially the considerable exergy loss of the irreversible compression has to be mentioned. Additional exergy losses are caused by the fact that also at the "warm end" of the counter-flow heat exchanger a temperature difference between low pressure and high pressure air has to be present. Lastly the loss of refrigerating capacity due to inefficient insulation of the plant ought to be taken into account.

ACKNOWLEDGEMENT

The valuable help of Dipl.-Ing. S. Schulz M. Sc., Braunschweig, in preparing the English text is greatly acknowledged.

LIST OF IMPORTANT SYMBOLS

e exergy eqo exergy of the cold produced e v n exergy loss of the throttling process e v a exergy loss in the heat exchanger e v x exergy loss of the cooler h enthalpy p upper pressure of the cycle Po lower pressure of the cycle q heat transferred to the surroundings

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q a heat transferred in the heat exchanger qo refrigerating capacity .< entropy 1' thermodynamic temperature Tu temperature of the surroundings Wt work input e coefficient of performance l; exergetic efficiency

REFERENCES

r. H. D. Baehr and K. Schwier, Die thermodynamischen Eigenschaften der Luft im Temperatur­bereich zwischen -2ro°C und + 1250°C. Springer-Verlag, Berlin/Gottingen/Heidelberg 196r .

2. K. Linge, Kaltluftmaschinenprozesse fiir tiefe Temperaturen. Kaltetechnik 13 (1961) S. 95-98 .

3. H. D. Baehr, Thermodynamik. Eine Einfiihrung in die Grundlagen und ihre technischen An­wendungen. Springer-Verlag. Berlin/Gottingen/Heidelberg 1962.

DISCUSSION

J. B. Chaddock, U. S.A. : What is the origin of the word "Exergy" ?

H. D. Baehr, Germany : The word "Exergy" was created some years ago in Germany by Prof. Dr. Z. Rant. For further details compare : Forsch. Geb. Ing.-Wes. 22 (1956) p. 36.

V. A. Martinovsky, UNESCO (Comment) : To obtain temperatures around -80°C and lower, the cycle with throttling is not suitable and cannot in practice be justified. It results in additional losses of energy. With these temperatures it is much more rational to utilize an open air cycle without throttling. The open vacuum cycle, in which the air entering in the decanter is being cooled in the regenerator up to a temperature lower than the temperature of the air on entering the decanter, offers substantial advantage. In this cycle the compressor can compress the air up to the pressure of one atmosphere and remove it from the installation. A highly positive factor for such a cycle is the absence of an air-cooler, since a new quantity of air enters in the regenerator at the ambient temperature.

Il-20

Recovering of Cold by Evaporating Liquid Methane Employed in the Air Separation to Obtain Liquid Oxygen and Nitrogen

Recuperation du froid par le methane liquide en evaporation utilise dans la separa­tion de l'air pour obtenir de l'oxygene et de l'azote liquides

Dr.-Ing. FRANCO RIVOIRA Corso Belgio 107, Torino, Italy

SOMMAIRE. L'utilisation de grandes quantites de methane liquide en Europe permettra d' obtenir, dans un proche avenir, une quantite importante de froid correspondant a la chaleur de vaporisation et de surchauff e recuperee du methane avant son passage par les tuyaux a methane.

Parmi /es dijf erentes applications de la recuperation, une methode semble pratique: utili­sation du froid a la separation de l' air pour obtenir de l' oxygene et de l' azote liquid es. C' est un procede peu couteux consommant peu d'energie.

II est possible d'utiliser l' azote liquide obtenu, transporte sur les memes methaniers, cet azote pourrait etre utilise avec profit pour la liquefaction du methane au point d'embarque­ment.

Cette methode pourrait realiser une chaine du froid entre l'installation de liquefaction et /'installation d' evalJoration, le result at permettant d' abaisser le cout de liquefaction du methane.

The energy deficiency of many European countries has contributed to the development of utilization of natural gases, principally, disposable in large quantities in Northern Africa and Middle East, which liquefied at the source of origin are transported by sea by tankers and then evaporated at the arrival harbour before being sent to utilization.

The technical and economical importance of the problem appears in all its width looking at the great quantity of liquid methane treated, which is foreseen ranging many thousand million of Nm3/year for the next years.

Till now the evaporation of liquid methane in the processing installation operating for the major part on experimental basis was accomplished by heating up the liquefied gas with sea or river water. In this manner it was obtained on the first hand a straight cut to costs by disposing of the large quantities of necessary heat supplied by natural elements, even if at a low thermal level, renouncing however to a rational utilization of the developed cold.

For exemplification it is to be precised that the quantity of heat recoverable through one year from one plant processing 1 x 109 Nm3/year of liquid methane evaporating up to a pressure of 30 kp/cm2, admitting to utilize the obtained heat between -162° C and 0°C, is :

(295 - 155) x 1 x 109 = 140 x 109 kcal/year,

which is really an important figure. Practically the total recovery of evaporation heat on the transformation site is not

technically convenient or economically desired because it is generally impossible to establish in one restricted area a series of consumption facilities, utilizing the obtained cold at different levels up to approximate the room temperature and in total quantity.

It is evident that a procedure aiming to recover the heat developed by the evaporation and overheating of liquid methane will become more effective as more as :

- the larger is the interval of temperature between the extremes of which the cold is utilized ;

- the lower is the capital investment necessary for the cold recovery ; - the more the cold utilization is valued (i. e., substituting high priced cold).

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It is our firm conviction that the cold recovering system described in this paper answers at the first and last conditions cited above, while the necessary investment does not sur­pass that required to reach the same scope adopting a conventional system. Fig. 1 illustrates one of the realizable scheme of the depicted procedure (covered by

Italian Patent n. 18860).

0 liquid altrogen

• Liqu.ld oxygen

® Liquid metbarle

� Ev&po!'&ting methane

F I G I

L E G E N D

Z air methane heat-exch&n1el'

3 liquid methane pump

4 M P column

6 LP column

7 refiu.s a.it1'ogon compreaaor

8 reflux nitrogen beat-exchanaer

9 reflux nitrogen con.denHr

10 nltrogen-alr heat-exchana:er

11 re-llque!ted oltroa:en compreaaor

IZ re-llq1141!1ed Jdtroaen b..,t-exchanaar

Fig. r . Scheme of a cold recovering system

The air is compressed by the air compressor 1 up to a pressure compatible with the subsequent heat-exchange with methane taking place in the heat-exchanger 2 ; the scrub­bing and drying stages are here omitted since they present no particular interest as they can be performed in the conventional manner. The liquid methane is compressed up to the requested pressure (30 kp/cm2) by means

of pump 3 and is then subjected to evaporation and overheating in the heat-exchanger 2. The processing air is in this way cooled down and forced to liquefy. The liquefied air is sent to the column 4 in which it is separated in a liquid reaching 30 % of oxygen and nitrogen. The nitrogen is liquefied at the expenses of the evaporated oxygen in the con­denser 5. The liquid nitrogen obtained is in part sent as reflux to column 4 and the rest is convey­

ed to storage tanks. Owing to the scarcity of reflux liquid nitrogen in the column 6, where the rich liquid

is separated in oxygen and pure nitrogen, an auxiliary group composed by the compressor 7, the heat-exchanger 8 and the liquefier 9 has been envisaged for the purpose of produc­ing the reflux liquid. The apparatus 9 furnishes also the necessary heat required at the botton of column 6 for

closing the thermal balance. The liquid oxygen obtained in column 6 is discharged to storage tanks, the gaseous nit­

rogen is divided in two parts, one of which supplies the necessary cold required in the heat-exchanger 8 as the second one is conveyed through the heat-exchangers 10 and 12. The first part of nitrogen is used for cooling down a fraction of the processing air for balancing the cold insufficiency in 2 and the second part allows to obtain, using the amount supplied by the compressor 1 1, additional liquid nitrogen. The method described above permits to obtain from one Nm3 of methane evaporated

and overheated under pressure, more than 0.3 Nm3 of liquid oxygen and about 0.75 Nm3 of liquid nitrogen. The energy consumption varies obviously as function of the pressure of the processing

air (which is determined by the pressure of methane), and depends from the working conditions of compressor, dimensions of plant, etc. Under particular circumstances it is possible to obtain 1 Nm3 of liquid oxygen and

1.9 Nm3 of liquid nitrogen with an energy consumption of0.9 kWh (excluding the energy

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necessary for pumping the liquid methane). It is interesting to compare such energy con­sumption with that required by conventional installations producing on a parity basis of output and purity the same quantity of liquid oxygen and nitrogen.

For comparison purposes let us choose two methods :

a) Oxygen and nitrogen production realized by means of air fractionation ; the plant involved operates at low pressure and employs heat regenerators. Liquefaction of both gases occurs with the Claude cycle.

b) Oxygen and nitrogen production as item a) with the difference that liquefaction is accomplished through a cascade refrigeration cycle adopting C3H3, C2H6, CH4, 02, Na.

The involved hourly output of oxygen has been considered to be 10,000 Nm3/h in both cases.

It was calculated that for producing 1 Nm3 of liquid oxygen and 1 .9 Nm3 of liquid nitrogen with methods a) and b), it is necessary respectively 2.50 and 2.35 kWh.

It is so evidenced a minimum gain superior to 1.40 kWh and reaching therefore a 155 % of the previous cited figure of 0.9 kWh.

An evaluation, even approximated, of the difference of cost existing among the three mentioned plants stresses the fact that on a parity basis the plant exploiting the liquid methane cold is the most economical one.

Evidently on considering the very big quantity of liquid methane evaporated in such works, the output of liquid oxygen and nitrogen would reach an astonishing figure. For example, such a plant processing a thousand million of Nm3 /year of liquid methane, would produce with the system previously described something like :

875,000 Nm3 of liquid oxygen a day

2,000,000 Nm3 of liquid nitrogen a day

of course less the losses.

It is urgent that the possibility of the placement on the market of such quantities of liquid oxygen and nitrogen would not be realizable economically.

On such premise the proposed solution could not be applied for the entire recovering of the disposable cold, but only on the prospect of recuperating a part of it. The recovered cold quantity should be determined on the basis of criteria of placement of the obtained products.

The other part of the disposable cold could be utilized for the production of energy which may be employed in the same production plant ; the cost of about Lire 7 for every kWh calculated for the energy produced with the procedure that utilizes the liquid methane as source of cold, seems reasonable good and acceptable for the indicated service.

A particular problem is then accentuated by the utilization of the liquid nitrogen ; such material, contrarily to whatever favorable prospects of employ, would be disposable in such a quantity exceeding any local market absorption. For this reason it has been thought convenient to load aboard the methane tankers the disposable liquid nitrogen, either for ballasting of the transport ships or utilizing it at destination for liquefying more methane.

The advantage of the last application appears evident in the light of the following considerations :

Let us suppose to liquefy methane at the pressure of 1 ata by circulating through a heat-exchanger in countercurrent with liquid nitrogen, which consequently evaporates and overheats itself. Depending on the fact that methane is obtained as a saturated or undercooled liquid we can obtain respectively 0.80 and 0.70 Nm3 of liquid methane for Nm3 of liquid nitrogen.

Now as the energy consumption required for the methane liquefaction using modern methods amounts to 300 kWh per 1000 Nm3 approximately, it derives consequently that every 1250 Nm3 of liquid nitrogen re-employed for the liquefaction of methane corresponds to 300 kWh, not considering the energy - on the other hand very low -required for circulating the methane and nitrogen fluids through the stages of the liquefaction plant described above.

33 1

II-20

If for example, we suppose that the quantity of liquid nitrogen obtainable given above be utilized to re-liquefy methane on a 50 % basis, we could benefit by the energy recovery of 250,000 kWh a day.

Concluding the present survey, it may be stressed that the recovering of cold from liquid methane which evaporates and overheats, appears to be inviting from the per­spective of technical standpoints and additionally even very interesting from the econom­ical point of view. Its introduction would consequently bring to dispose of a great quantity of oxygen and nitrogen in the liquid form and at very low output cost. The utili­zation of liquid nitrogen as liquefying agent of the methane on the spot of production would lead to further relevant reduction of the energy requirements, which is bound to straight cost in operation outlay.

DISCUSSION

H. Hausen, Germany : I think, in the figures for power consumption, the consumption for the liquefaction of methane is not included. If you add this, how does the total value of power consumption compare with the power consumption for the liquefaction of 02 and N 2 by the usual methods ?

F. Rivoira, Italy : I did not give the calculation about the power consumption for the liquefaction of methane because I assumed that the methane arrives at the discharge harbour in liquid form, and the problem is to vaporize it recovering as large a quantity as possible of the energy stored in the same liquid methane.

I remember that Professor Hausen said that in the actual plants the liquid methane is vaporized by means of the heat of river or sea water.

332

li-24

Selection of Comparative Theoretical Cycle of Vapour Compression Refrigerating Plants

Sur le choix du cycle theorique de reference des machines frigorifiques a com­pression de vapeur

I. S. BADYLKES Scientific Research Institute of the Refrigerating Industry of the USSR, Moscow, USSR

SOMMA/RE. On utilise actuellement de nombreuses combinaisons d'hydrocarbures comprenant le fluor, le ch/ore et le brome dans diff erentes proportions. La qualite thermodyna­mique de certains «Freons » du groupe methanique dans les conditions d'application du cycle de reference theorique itabli est a peu pres egale a celui du cycle de recuperation. Cependant, en utilisant le dernier on peut gagner un effet pratique considerable du a la pre-evaporation dans l'echangeur thermique dufluidefrigorigene dissout dans l'huile.

P<Jur !es fluides frigorigenes a poids-mo!eculaires eleves caracterises par la solubilite limitee dans l'huile (R 115, C4F8, C4F10 etc.), le cycle de reference theorique avec recuperation compare au cycle de reference theorique etabli actuellement, assure les valeurs volumetriques et energetiques considerablement plus hautes.

ll ne nous semble pas possible d' etablir un cycle de reference theorique unique pour tous les fiuides frigorigenes et tous les types de machines.

En comparant des pertes reel/es dans les machines frigorifiques de differents types, fonction­nant avec dzfferents fluides frigorigenes, il est necessaire de comparer pour chaque machine le cycle reel avec le cycle theorique de reference optimal correspondant.

As is known, the highest and similar for all the refrigerants limit of thermal efficiency of a refrigerating machine is obtained with the Carnot cycle and equals

To Kc = -·--- 860

T - To kcal/kwh

In the case of a detailed evaluation of the losses in an actual refrigerating machine it is more correct, however, to proceed from the comparative theoretical cycle which approaches the actual one and not from the Carnot cycle. In this case

where : K1

Kth

K; K;

'YJ 1 = k th = K.� 'YJ g

is the indicator performance factor ;

the performance factor of the comparative theoretical cycle ;

'Y) g the degree of thermodynamic perfection of the refrigerant.

The following deviations from the Carnot comparative theoretical cycle are considered when determining the degree of thermodynamic perfection :

a) compression in the superheat zone follows the adiabatic curve (1 - 2, Fig. l ) ;

b ) the superheated vapour cools down in the condenser to the saturation point, lique­fies, and the condensed liquid subcools (2 - 3, Fig. 1 ) ;

c) adiabatic expansion is replaced by a throttle valve (3 - 4, Fig. 1).

According to the decision of Commission 4 of the International Institute of Refrigera­tion (1938), it is adopted that in the comparative theoretical cycle the compressor takes in dry saturated vapour (point 1, Fig. 1).

333

11-24

1

i -

Fig. r Comparative theoretical cycle with the deviations a, b, c from the Carnat comparative cycle

However, in connection with the utilization of different fluorochloroderivatives (freons), a regenerative heat exchange process is effected with considerable superheating of the vapour from the evaporator owing to the subcooling of the liquid freon from the conden­ser (3 - 4, Fig. 2) [1]

l

z. -Fig. 2 Refrigerating plant cycle with superheating of vapour

Different freon compressors, operating at to = -15°C and t = +30°C are compared in the USSR [2] at a vapour suction temperature of foh = + 15°C (Fig. 2).

According to the USA rules [3, 4], which are used also in the Federal Republic of Ger­many [1], the testing of single-stage freon compressors (to = +5° to -25°C, t = +40°C) is effected at foh = +20°C.

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It is evident from the data given above that the comparative theoretical cycle, adopted at present with compressor suction of dry saturated and not of superheated vapour, does not conform with the existing national rules.

The preservation of this comparative theoretical cycle is explained by the fact that the values of the theoretical performance factor and of the theoretical refrigerating effect per unit of swept volume are approximately the same in both cycles. This has been proved on a strictly scientific basis by the author [5] and then by K. Linge [6].

The author [7] derived later the following generalized dependence :

1 - <p CTr • Ts . -T T C' = idem

- 0 µ 8

where : <p = �: [ qo - refrigerating capacity per kg, kcal/kg] ; CT.- Trouton number ; Ts - normal boiling temperature, ° K ;

C's - specific heat of boiling liquid at 1 atm. ; kcal/kg° K ;

µ - molecular weight of refrigerant.

It is quite clear from this generalized dependence that atthe given operating temperatures T and To, high values of the molar heat capacity of the boiling liquid - µC' ., and at low normal temperatures T s the parameter rp decreases and the losses sharply increase owing to the replacement of the expansion cylinder by the throttle valve. Therefore, the appli­cation of a cycle with suction of superheated vapour provides under these conditions for a substantial increase of the theoretical performance factor and theoretical refrigerating effect per unit of swept volume as compared with the set cycle.

As in the case of freons [5] :

µC's 2:'. (3.215 + 0.925 m) µ0•4

where : m = the number of carbon atoms ; Cs' = the specific heat of boiling liquid,

considerable losses are due with polyatomic and high-molecular freons. As a rule, the values of the vapour specific heat of these freons become positive on the saturation curve and adiabatic compression is effected in the zone of wet vapour [7].

For instance, it has been stated [8] that with freon-1 15 (t s = -38°C) in a cycle with suction of superheated vapour (toh = + 20° C) the theoretical refrigerating effect per unit of swept volume (respectively the theoretical performance factor) increases approximately by 40% at t0 = -38° C, t = + 30°C and subcooling of the liquid down to -7°C.

A considerable improvement of the energy data when superheating vapour is due also in machines operating with perfiuorobutane [5], sulfur hexafluoride [9] and some mixtures.

Further subcooling of the freons with an external source (water, air) should be excluded from the comparative theoretical cycle because high superheating of the vapour, taken in by the reciprocating compressor by subcooling the liquid, considerably increases (with consideration of the actual losses) the volumetric efficiency and indicator coefficient [7, 10, 1 1, 12].

Thus, the adoption of a new comparative theoretical cycle with subcooling of the liquid by superheating of the vapour taken in by the reciprocating compressor meets the require­ments of practical needs.

It is quite evident, nevertheless, that it is impossible to set up a unified comparative theoretical cycle for all the types of refrigerating machines and refrigerants.

For instance, considerable actual losses when operating with suction of dry saturated vapour (instead of superheated) are due to take place also in refrigerating rotary freon compressors and, therefore, a comparative theoretical cycle with regenerative heat ex­change can also be recommended for them.

335

11-24

Quite different is the problem with refrigerating centrifugal compressors. Owing to the small influence of the superheat on the value of the actual losses, the comparative theoreti­cal cycle existing at present with suction of dry saturated vapour and subcooling of the li­quid with water can be preserved. However, the application of the comparative theoretical cycle with regenerative heat exchange is highly necessary when using high-molecular and polyatomic refrigerants.

As far as refrigerants with a low molecular weight are concerned, the main of which is ammonia, it is advised to use the generally adopted comparative cycle taking into conside­ration the theoretical losses at increased suction temperatures and resulting too high temperatures at the end of the discharge process.

We stress in conclusion that when comparing actual losses in refrigerating machines of different types and operating with different refrigerants, the actual cycle of each machi­ne should be compared with the optimum comparative theoretical cycle. Further research is required for developing unified comparative temperatures.

REFERENCES

r. Kaltemaschinen-Regeln, Verlag C. F. Muller, Karlsruhe, 1958.

2 . Refrigerating Engineering. Encyclopeadic handbook, vol. I, Moscow, Gostorgizdat, 1960.

3. ASRE-Standard, Circular No. 14-41, 1940.

4. ASRE-Standard, Circular No. 23 - R, 1949.

5. I. S. Badylkes, Working substances for refrigerating machines, Moscow, Pistschepromizdat, 1952.

6. K. Linge, Kaltetechnik, Vol. 8, 1956.

7. I. S. Badylkes, Working substances and processes in refrigerating machines, Moscow, Gostor· gizdat, 1962.

8. I. S. Badylkes, Kholodilnaya Tekhnika, No. 4, 1956.

9. I. I. Perelstein, Thermodynamic properties of sulfur hexafluoride, Moscow, Gostorgizdat, 1961

10. D. M. Joffe, V. B. Yakobson, Small refrigerating machines and commercial refrigerating equipment, Moscow, Gostorgizdat, 196r.

1 1 . G. Lorentzen, International Institute of Refrigeration, Annexe No. I, 1953.

12. R. Plank, Kaltetechnik Vol. 4, 1952, p. 11.

DISCUSSION

V. A. Martinovsky, UNESCO : I agree with the point of view of the author that the theoretical reference cycle in use at the present time should be modified, taking into consideration the various refrigerants used. Many refrigerants produce in practice a big amount of superheating. Therefore, the comparison between different refrigerants in conditions fixed by the theoretical reference cycle, is very relative.

C. Codegone, Italy : On donne quelque eclaircissement sur le cycle thermodynamique propose par M. Badylkes en faisant observer que c'est aux valeurs elevees de la masse mo!eculaire µ qu'il y a lieu de I' adopter.

336

H eat Transfer Trans m i ss i o n de chal e u r

11-14

Heat Transfer of Boiling Refrigerant 12 in Horizontal Tubes with Internal Flow Channel Guides

Transmission de chaleur d'un Frigorigene 12 en ebullition dans des tubes hori­zontaux a ailettes internes

HIDEO UCHIDA, Professor University of Tokyo, Tokyo, Japan

SYUNITI TEZUKA, Assistant Professor Tokyo University of Mercantile Marine, Tokyo, Japan

SOMMAIRE. Ce rapport presente quelques recherches experimentales sur la transmission de chaleur et la perte de charge dans le cyclefrigorifique d'un evaporateur.

Dans cette experience, les AA. utilisaient deux types d' evaporateurs; l'un etait d double-tube avec du Frigorigene 12 d l'interieur des tubes de cuivre horizontaux ( diametre exterieur : 19 mm, diametre interieur : 17,4 mm, et L 2,260 mm) qui, d leur tour, contenaient des ailettes transversales (plaque de cuivre de 2 mm d'epaisseur) ; l'autre etait un evaporateur a double­tube muni de lamelles helicoidales.

Les conditions des experiences etaint les suivantes : debit de frigorigene : 190 d 340 kgjh, pression dufluide : 2,8 a 4,4 kg/cm2 abs.,facteur de qualite dufluide : 0 a 1. (fraction du poids de vapeur dans lefluide).

Les 1 esultats de ces recherches sont les suivants : 1. variation du coefficient de transmission de chaleur ou du flux de chaleur en f one ti on du

f acteur de qualite, 2. variation du flux de chaleur en f onction de la difference de temperature entre la paroi du

tube et le frigorigene, 3. formule approchee pour la perte de charge totale.

1 . INTRODUCTION

When a refrigerant is evaporated the weight fraction of vapour phase changes. The weight fraction of vapour in the mixture (vapour kp/kp of mixture), which we call the quality of the refrigerant, is denoted by x and varies from 0 to 1 .0. For the analysis of evaporator, therefore, we should investigate the heat transfer (heat transfer phenomena in this case, being very complicated) and pressure drop of two phase flow for which vapour void varies in a wide range.

A number of studies about the boiling heat transfer of refrigerant have already pre­sented e. g. ; for boiling inside tubes by Ashley [1], Seigel [2], Yoder [4], Bryan [ 3, 5], Baker [6], Pierre [7], Hofmann [8] et al., and for boiling inside inner finned tubes by Bryan [3], Boling [9], and Seigel [5] et al. However, most of them were carried on experiments under constant heat flux. There have been no investigations regarding relation of heat flux vs. quality of fluid or the maximum heat flux vs. quality.

In recent years, the problem of pressure drop in two-phase flow has been investigated in many ways, but their results can hardly be applied to the refrigerant evaporator, except the Martinelli's experiment [10] with water.

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The investigations, presented here by the authors, illustrate several problems of heat transfer and pressure drop in horizontal tube evaporator with internal flow channel guides. In their experiment the types of fins installed in the evaporator tubes were longitudinal cross guides and the twisted tapes.

2. EXPERIMENTAL PROCEDURE

The evaporator in this experiment was a double tube heat exchanger as shown in Fig. 1. Refrigerant 12 (CF2CL2) flowed inside the inner tubes made of copper and 8 in number (O· D · l9 mm, I · D · l7 mm and length of 2,260 mm). Water flowed through the annular space between inside and outside tubes.

REFRIGERANT

t t

tw thermo-coup les points on the water

- - - - - .. - - - - - - - -·- - - - - - - -- - -- · -- - -- - - - -- - ---- -

- - - - - - - - - --e- - - - - - - - • - - - - - - - --- - - - -

t t - - - - - --- - - - - - - - -... - - - - - - - _.. _ - - - -

- -- · -- - -- - --- - -- - - -- - - - - - - - .. - - - - - - - ... - - - - - - ---- - - - - - - -

p

thermo-coup l e s point on the tube wal l s . thermo-coupl e s point o n the refr i gerant . pre s sure taps points on the refri gerant .

( 111 111. ) Fig. r. Drawing showing a Double tube evaporator.

Two types of internal flow channel guides were used. One was longitudinal cross guides, made of 2 mm thick copper plate, which is shown in Fig. 2 A and the other was twisted tapes, made of 0.3 mm thick copper plate, shown in Fig. 2 B.

'ft 0. 3,,,m t --+-" r:- T "

"" <!II 11- r:-IK 0

. "'"' _Jc__ i I r/ t

2 A 2 B

Fig. 2 A. Longitudinal cross guides. Fig. 2 B. Twisted tapes. Drawing showing construction of inner fin tubing.

Pressures of the refrigerant were measured by calibrated Bourdon gages, and pressure drops in each tube by a mercury-manometer inclined at 15 degree. Temperatures of the refrigerant, water and tube wall were measured by copper-constantan thermo­couples and flow rates of refrigerant and water by means of calibrated orifices. For meas­urement of inner tube wall temperatures, thermocouples were soldered on the tube surface at three points on the circumference (top, mid-side and bottom of tube) at three sections (inlet, middle and outlet) of each tube as shown in Fig. 1.

338

11-14

3. RESULTS AND CON SID ERA TIONS

3.1 Local heat transfer coefficient oc and Quality x

Fig. 3 A shows temperature range for water and refrigerant employed in the experi­ment. For elementary temperature difference LI tw, enthalpy change LI i was calculated and hence the quality element LI x from equations (1) and (2). This helped to under-

3 0

p �.�•. G \£-2 0 3.0 -2.3 202

...... � ;> r 0

I o 0.4 : 3t � .µ

o. 3 <l .... .., .. ...., 0 0 .2

0.1

- /Q i...c=-.J.._��2�--'-3�----'4�----'$��6"--��7��8 o s e c t i o n

0 S I 0 I S P o , ition ato�9 t .. b e . 1!l

1 i . I

1 . 0

o.�

(!�

0.1

O.b ><

o.s

0.4 0. 3

0.2

0. 1

0

Fig. 3A. Curves showing relation of temperature (IR, lr, lw), pressure drop LI Pb and quality x vs. position.

stand the quality x which was calculated for a number of points in order to plot the curve for the variation of x with position along the tube.

g = Gw LI tw = G LI i

LI i LI x r

Gw L'.ltw G r

(1)

(2)

The results for longitudinal cross guides are shown in Fig. 3 B and for twisted tapes, in Fig. 3 C.

From these results we can conclude as follows. The values of oc for longitudinal cross guides; cross guides are higher than those for twisted tapes. The points giving local heat transfer coefficient in Figs. 3 B, 3 C are distributed almost symmetrically about the line represented by x = 0.5 on the graph. The values of oc for longitudinal cross guides; cross guides are almost constant for a wide range of x (x = 0.3 to 0.7). But there are two transition points, occurring at x = 0.15 and 0.85. On the other hand, the curves of oc vs. x for twisted tapes are seen to have a rounded shape, with no distinct transition points. Also, oc value increase by a small amount with the increase in G as shown in Fig. 3 C.

339

II-14

C.Yoss 11..lil es P G- C. I A Kp/cm'o.� KP/k KP/m2k.

0 4.4-4. 0 340 2. 1 }t 106 • 3. 8 � 3.2 2 � 0 !. 5 a 3. 'l-2.q 2 2 0 1. 3'. • 3. '1-M 2 0 s 1 . 2 7 ti. 3 . '1 � 3.o I q 1, 1. 2. I • ... .3. b -2.q 1 q 3 1. 1q '•

o 0. 1 0.2 o. 3 o.4 o.s o. 6 o. 7 o.� o. q 1. 0 x:

Fig. 3 B. Local heat transfer coefficient vs. a Quality x. (Longitudinal cross guides) .

� - ....... � � � � - -

JI; ?/ � ' �

I I -� ll

I :! ' r �

P lf/cm�bs. � l'Wh it 0 4.'J - 3. '7 352 • 3. q - 3. 0 2 85 c 3. 'I - 2.q 2 3b '-· • 3. 0 - 2. 3 2 0 2.

- -

1 0 2 0 0. 1 o. 2 0. 3 o. 4. 0. s 0. 6 o. 7 0. 8 0. q /. 0

.x Fig. 3 C. Local heat transfer coefficient a vs. Quality x. (Twisted tapes).

3.2 Quality x and Local heat flux q/A The relations between local heat flux and quality are shown in Fig. 4 A for longitu­

dinal cross guides and in Fig. 4 B for twisted tapes. q/A values, used for plotting the curves Fig. 4 A and 4 B, are calculated mean values for every interval of 0.1 along x axis. From these considerations we can say that both the above mentioned results are similar and q/A attains a maximum value at around x = 0.8.

340

31-104 ..--� :l � ........

1 04

<

" �

•

�k� h 0 .f.4-4.0 3 4 0 �--+-­• 3.i-3.'2 2. 6 0 0 3.'1-?.q 220 • 3.'1 -2.q 20!i A 3.'l-3.0 I q(, ... 3.i,-2.q 1 4 3

/ 03 ..._--''---'-�-'-�--'--�-'--�'------'-�-"-��------4 o 0. 1 0. 2 o. 3 o.4 o. s o. 6 o. 7 o. 8 o.q 1. 0

x Fig. 4A. Local heat flux q/A vs. Quality x. (Longitudinal cross guides).

II-14

The increase in heat transfer with increasing quality x at constant weight flow G is mainly caused by the increasing velocity as growing x. Only when x is approaching value 1 (gas only, no liquid fluid), the heat transfer is decreasing due to the reduction of the amount of evaporated liquid.

-< " clo

0 • Cl •

p l<fk�� �/h 4.'1 -3. 'I 352 3.'i -3. 0 2.8!i 3.'J -2.'j 2 3b 3.0 -2.3 2 0 2

o 0. 1 0. 2 o. 3 o.4 o. s o. 6 0. 1 o. & o.q 1 . 0 x

Fig. 4B. Local heat flux q/A vs. Quality x. (Twisted tapes).

3.3 Local heat flux q/A and temperature difference between tube wall and refrigerant L1 t The values of q/A are plotted against temperature difference between the tube wall

and the refrigerant in Fig. 5 A and 5 B, for longitudinal cross guides and twisted tapes

341

II-14

respectively. In these experiments, LI t was found to be increasing with quality. Now, we can consider the tendency of the refrigerant evaporating inside tubes, by comparing the results shown in Fig. 5 A and 5 B- with the so-called pool boiling curves. The in­creasing tendency of q/A, for decreasing of LI t, is similar to that of pool boiling. However, the maximum heat flux is excessively smaller than that, in case of pool boiling. This difference shall depend on the effect of two phase flow inside the tubes.

2 x 1 04 I -

l L 19 /- U "' :r

I I I f

� I I j ) J, 1 Al I f

7 0 • a • A -- �

I

fl t

f � ltil� '°':I \" A

... \.

-,-- - ·-

p "'/ell' <:r i<p/h 4.4-4.0 & '·3 4 0 3.'&-3.2 2 6 0 3.1 - 2.q 2 2 0 3.7 - 2.q 2 0 5 3.'1-.�0 l 'l 6 3.f:i -2.q I 'l 3

I 5 J O (ae5 · )

-

---

2 0

Fig. 5A. Local heat flux q/A vs. Temperature difference between tube wall and refrigerant LI I (Longitudinal cross guides).

2 X / 03 1-----'�-'---'-J_j.--1-'--'-�--+---'-�--' 2 4 b 8 1 0

A t (de9. ) 3 0

Fig. 5 B. Local heat flux q/A vs. Temperature difference between tube wall and refrigerant.Lii. (Twisted tapes).

342

Il-14

3.4 Pressure drop At first, the authors carry on experiments of pressure drop through the tubes in

which refrigerant gas flows. From these experiments they obtained frictional factors for both the types of fins. The experimental formulae obtained from the results are shown as following. The frictional factor f c for longitudinal cross guides can be repre­sented by a function of Reynolds number. The factor fi, for twisted tapes, does not depend on Reynolds number but is a constant for a wide range of the same. In Fig. 6 A and 6 B are shown the ratios of LI Pb (LI Pb refers to the pressure loss of boiling

� ""' CL 00 ...: IV)

-..I) «t ,..... +

6 0 -

3 0

,_

1 0 l.1 0

�p I/

3 -;;; I/ v

3

��

0 l;l � �

I

J,., ,.,r:si/

� ;:c�

- - -·- --,_

Fig. 6A. The experimental formula of LI Pb/LI Pr Longitudinal cross guides).

/

,_

b O

fluid in this experiment) to LI P1 (LI P1 refers to that of mere liquid instead of boiling fluid and calculated with the above mentioned frictional factor fc or ft. U is the velocity of the fluid in liquid state and therefore varies with the weight flow and density only).

0. 1

P �P�i Gr "% � 0 4.1 - 3. 1 3 � 2 • 3,q -3.0 2 8 5 Cl 3. 1 -2.1 2 % • 3. 0 -2. 3 202

o. s .x

/ . 0

Fig. 6 B. Pressure drop ratio LIPb/LI P1 vs. Quality. (Twisted tapes).

343

11-14

For longitudinal cross guides ;

U2 L L1 P 1 = 4 fe 2 g Y De

f e = 0.32 Re -o·a55 L1 Pb/Ll P1 = 1 + (146 - 31.8 P) x

For twisted tapes;

U2 L 4 /1 - y 2 g De

"t = 0,052 (/1 = 0.058 by Gambill (1 1 .) )

CONCLUSIONS

1) The values of heat transfer coefficient CG for longitudinal cross guides in Fig. 3 B are higher than those for twisted tapes in Fig. 3 C and the points giving local heat transfer coefficientes in Figs. 3 B and 3 C are distributed almost symmetrically about the line represented by x = 0.5 on the graph for the experimental conditions (flow rate of refrigerant 190-350 kp/h, pressure of fluid 2.3-4.7 kp/cm2 abs.).

2) Variation of local heat flux q/A with temperature difference between tube wall and refrigerant L1 t in Figs. 5 A and 5 B is similar to that for pool boiling. However, the maximum heat flux is excessively smaller than that of pool boiling. This difference shall depend on the effect of two phase flow inside tubes.

3) The values of L1 Pb/ L1 P 1 for longitudinal cross guides in Fig. 6 A are smaller than those for twisted tapes in Fig. 6 B.

NOMENCLATURE

A : Inside surface of inner tube excluding fin area. = : D2 L (m2)

D : Inside diameter of inner tube. (m)

De : Equivalent diameter of inner tube including fin effect for pres· sure drop. = 4 X net free volume / frictional surface. (m)

:rt De = 4 -,;f D2 - 2 Ye • D + Ye2) I D (:rt + 4) - 8 Ye

- - - for longitudinal cross guides.

:rt De = 4 4 D2 - Yt · D) / D (:rt + 2) - 2 Yt

- - - for twisted tapes.

f Friction factor

g : Acceleration of gravity. (m/sec.2)

G Feed rate of refrigerant. (kp/h)

Gw Feed rate of water. (kp/h)

L1 i Calculated enthalpy difference of boiling refrigerant. (kcal/kp)

L Length of test section. (m)

P : Absolute pressure of refrigerant. (kp/cm2 abs.)

L1 Pb : Measured pressured drop of boiling refrigerant. (kp/m2)

L1 P 1 : Calculated pressure drop for isothermal flow of liquid refrigerant. (kp/m2)

r : Latent heat of refrigerant. (kcal/kp)

344

11-14

U De Re Reynolds number.

Heat flow.

µ

q q/A Local heat flux.

Temperature of refrigerant.

Temperature of tube wall.

(kcal1h)

(kcal/m2h)

coq

Temperature of water.

Temperature difference between inner tube wall and refrigerant flowing there in.

(° C)

(O C)

u Velocity of refrigerant in internal flow channel guides.

Weight fraction of vapour phase (we say quality).

(deg.)

(m/sec.)

x LI x Calculated quality difference of boiling refrigerant.

Ye, Yt : Width of fins for longitudinal cross guides and twisted tapes. (m)

(kcal/m2 h)

(m2/sec.)

(kp/m3)

x Local heat transfer coefficient.

µ

}'

Kinetic viscosity of refrigerant.

Specific weight of refrigerant. The physical properties of refrigerant was used Plank's data. (R. Plank, Handbuch der Kaltetechnik. vol. 4 1956.)

REFERENCES

r. C. M. Ashley, The heat transfer of evaporating freon. Refrig. Engng., Feb. 1942.

2. L. G. Seigel, W. L. Bryan and M. C. Huppert, Heat transfer rates for refrigerant boiling in horizontal tube evaporators. Heating, Piping & Air Cond., Jan. 1949.

3. W. L. Bryan and G. W. Quaint, Heat transfer coefficients in horizontal tube evaporators. Refrig. Engng., Jan. 195 r .

4 . R. ] . Yoder and B. F . Dodge, Heat transfer coefficients of boiling freon 12 . Refrig. Engng., Feb. 1952.

5. W. L. Bryan and J. G. Seigel, Heat transfer coefficient in horizontal tube evaporators. Refrig. Engng., May 1955.

6. M. Baker, Y. S. Touloukian and G. A. Hawkins, Heat transfer film coefficients boiling inside tubes. Refrig. Engng., Sept. 1953.

7. B. Pierre, Warmeiibergangszahl bei verdampfendem F 12 in horizontalen Rohren. Kalte­technik 7, 1955.

8. E. Hofmann, Wiirmeiibergangszahlen verdampfender Kaltemitte!. Kiiltetechnik 9, 1957.

9. C. Boling, W. ]. Donovan and A. S. Decker, Heat transfer of evaporating freon with inner-fin tubing. Refrig. Engng., Dec. 1953.

10. R. C. Martinelli and D. B. Nelson, Prediction of pressure drop during forced-circulation boiling of water. Trans. ASME., Vol. 70. 1948.

I I . W. R. Gambill, R. D. Bundy and R. W. Wansbrough. Heat transfer, burnout, and pressure drop for water in swirl flow through tubes with internal twisted tapes. ORNL-29II, !\far. I96o.

DISCUSSION

Mr. Benke, Germany : 1 . What was the magnitude of the oil component in the coolant, in per cent by weight,

in the investigations ?

2. Have measurements been made on the effect of the oil component in the coolant, and if so, with what results ?

The authors : We have not measured the oil component and that effect in the evaporator, but we thought the Refrigerant 12 was mixed with a trace of refrigerant oil.

345

11-14

S. Tauber, Netherlands : I would like to know which kind of control device was used for admitting the liquid refrigerant to the evaporator. Their experience is that automatic devices generally tend to fluctuate a great deal.

The authors : The quantity of refrigerant flow is controlled by hand to make the outlet temperature nearly constant.

J. J. Kowalczewski, Australia : The surface heat transfer coefficient ex, between the wall and the boiling refrigerant is given by:

ex, = q/A · L1 t = G · r · Llxf Ll t

using the nomenclature of the authors. What were the values of L1 x for ex, shown in Fig. 3 B, i. e. how "local" are the heat transfer coefficients in this figure ?

The values of ex, in Fig. 3 B were measured at a variable heat flux, as can be seen from Fig. 4A. The data of Yoder and Dodge, also shown in Fig. 3 B, are for a constant heat flux. This may partly explain the onset of liquid deficiency manifesting itself in a decrease of cx. This occurs in the data of Yoder and Dodge at about x = 0.45, compared with that of Uchida and Tezuka (at about x = 0.65).

The kilopound (kp) is a unit of force, used in the technical system in Europe. In this system of units, the unit of mass is a kilogram (kg). It is therefore not necessary to express mass by force, as done by the authors, although this is p�rmissible at not too high velocities ?

The authors : The local heat transfer coefficient ex, was calculated from the heat trans­fered to water as under :

ex, = qfA Llt = Gw LltwfA Llt

and the local heat transfer coefficient ex, and local heat flux q f A in our paper were calculated mean values for every interval of 0.1 along the x axis (see p. 340 in the paper).

The data of Yoder and Dodge were obtained from experiments under constant heat flux (see p. 337 in the paper).

M. Duminil, France : Would the author specify where he has located the thermocouples on the walls of the tubes ?

How has he defined the mean temperatures of the walls, and how many points are used to find the mean ?

The authors : For the measurement of the inner tube wall temperatures, thermocouples were soldered on the tube surface at three points on the circumference (top, mid-side and bottom of tube) at three sections (inlet, middle and outlet) of each tube as shown in Fig. 1 (see p. 338 in the paper).

The mid-side tube wall temperatures at three sections of each tube were used to calculate the local heat transfer coefficient cx.

C. F. Kayan, U.S.A : Is it correct that some super-heat was indicated by the refriger­ant temperature in Fig. 3 A ? I believecx, the local coefficient, was calculated progressively along the length from the corresponding wall temperature. Was the leaving quality measured on exit of the refrigerant, and was the quality actually measured along the length of the refrigerant travel ?

The authors : The refrigerant temperatures shown in the paper are measured values.

The quality was calculated from equations (1) and (2) and the meanings of Lltw and Llx are shown in Fig. 3A. The initial quality x1 in the evaporator was decided from the Mollier p-i diagram, and the quality x n at any point n along the evaporator tubes was calculated as under :

346

Xn = Xn-1 + Llxn-1 = Xi + L'L1Xi L'Llx; = L'GwL1tw1fG r

II-10

Boiling Heat Transfer to a Cryogenic Fluid in Both Low and High Gravity Fields *

Transmission de chaleur par ebullition a un fluide cryogenique de gravite faible et de gravite elevee

PROF. J. A. CLARK and PROF. H. MERTE, Jr. Heat Transfer and Thermodynamics Laboratory, Department of Mechanical Engineering, The University of Michigan, Ann Arbor, Michigan, U.S.A.

SOMMAIRE. L'un des problemes importants rencontres dans la conception du materiel destine a I' exploration spatiale est le processus de la transmission de chaleur par convection pour les substances cryogeniques ayant des gravites plus elevees et moins elevees que la terre. Ce rapport traite des donnees experimentales et de leur correlation par rapport a des spheres et des plaques planes pour l' ebullition de l' azote liquide, dans les regions du film en ebullition, nucleee et passagere, comprenant le flux de chaleur maximal et minimal, de gravites a/lent de 0 a la normale. On presente de nouvelles donnees sur !'ebullition nucleee entre la gravite normale et une gravite egale a 20 fois la normale. Les resultats sont obtenus a !'aide d'une tour de chute a contrepoids et un appareil centrifuge. En utilitant une technique transitoire, puis une reduction a l' aide d'un calculateur digital electronique. On compare ces donnees avec celles obtenues par d' autres chercheurs.

INTRODUCTION

With the advent of space exploration, certain new gravity conditions have been en­countered which require the refocusing of attention on forces normally not considered as variables in design for earth application. The range of gravities to be encountered in space flight may be expected to be from near zero to many times standard (terrestrial gravity, g = 32.2 ft/sec • or 980 cm/sec 2) gravity. Space vehicles in earth or lunar orbit or in drifting space flight will experience force fields corresponding to gravities which are very nearly zero. Such vehicles when propelled with low thrust engines or parked on the moon or similar bodies in space, will be exposed to gravity fields somewhat greater than zero but still considerably less than on earth. Intermediate, sub-terrestrial gravities, may be expected to exist on space vehicles when propelled by small engines or at partial thrust. The condition of high gravity also will be obtained in vehicles propelled by high thrust engines, during exploration of the larger planets, in space vehicles subject to ro­tation, and during lift-off and re-entry among others.

For the design ofreliable apparatus it is important to understand the response of phys­ical systems to these new kinds of environment. It is evident that equipment design for space flight is practical, at least for flights of limited duration [ 1].

An important example of these new types of application is the behavior of fluids under various gravities for both adiabatic and diabatic environments. Studies on the former have been conducted at low gravity and reported in [2, 3, 4, 5, 6, 7]. Movies showing the adiabatic behavior of liquids in zero-g are available [3, 4, 6]. The behavior of liquids under low gravity with heat transfer have been reported in [8, 9, 10, 1 1, 12, 13, 15, 20J.

• The work reported was sponsored by the NASA, George C. Marshall Space Flight Center, Hunts· ville, Alabama, U.S.A., under contract NAS-8-825.

347

11-10

Data are reported on boiling heat transfer at high-g in [16, 17, 19]. New data for the boiling of liquid nitrogen at gravities up to approximately 20-times standard are present­ed in this paper.

HEAT TRANSFER AT REDUCED AND NEAR ZERO GRAVITY

Reduced and near-zero gravities are obtained in a 32 ft. (9.75 meter) drop tower, Fig. 1, having a free fall time of 1.4 seconds. A counterweight provides for gravities in the range zero to standard gravity.

A 5-liter insulated container filled with a cryogenic liquid (nitrogen in the present tests) is mounted on the 100 pound (45.3 kg) test platform. The platform is released by passing a large AC current through a single support wire, thus causing it to melt. The test platform is brought to a smooth halt in 2-1 /2 ft (0.762 meter) by a specially designed hydraulic buffer.

True zero-g is not attainable with this system because of air resistance and possible friction with the guide wires. However, several tests with a 0-1 g accelerometer under free-fall conditions indicated that the maximum force field present on the platform was less than that corresponding to 0.01 g. An accelerometer is mounted to the test package to provide a continuous measurement of system acceleration during both free and counterweighted fall.

With the short free-fall time available, a transient technique was adopted in which the time rate of change of enthalpy of a body would provide a measure of the heat transfer rate at its surface. A 1-in. (25.4 mm) and a Y2-in. (12.7 mm) diameter copper sphere, Fig. 2, was constructed to serve this purpose as a dynamic calorimeter. Copper was selected because its heat capacity is well known as a function of temperature and because it has a high thermal diffusivity. Two thermocouples were provided, one at the center and one near the surface, of the sphere. The EMF from these thermocouples were record-

348

SECOND

FLOOR

FIRST

FLOOR

32 '

Fig. r. Drop tower facility.

THERMOCOUPLE

(SHIELDED) CABLES

GUIDE WIRES

COUNTERWEIGHT

CABLE

COUNTERWEIGHT

COUNTERWEIGHT

RELEASE

(FRACTIONAL GRAVITY)

HYDRAULIC BUFFER

ARRESTING GEAR

1 1-10

Fig. 2. View of 1-inch (25.4 mm) diameter copper test sphere assembly.

ed on an oscillographic recorder through approximately 50 ft (15.25 meter) of shielded cable which falls free with the test platform. Liquid temperatures also are measured and recorded. The spheres were highly polished and cleaned but no other surface condition­ing was performed. Details of the experimental apparatus are given in reference 15

and 20. Data were taken by immersing the copper sphere, initially at room temperature, into

the liquid nitrogen on the test package. Because of the large initial temperature differ­ence, stable film boiling is first obtained. As the sphere cools, it ultimately passes through the entire boiling regime from film boiling to non-boiling free convection. The record of the sphere temperature and internal differential temperature while it is in a zero gravity condition is shown on the oscillographic chart in Fig. 3. Highly accurate data of sphere temperatures, time rate of temperature change, surface-fluid temperature differ­

ence and corresponding real time are obtained. Experimental data similar to that shown in Fig. 3 are used in a transient analysis to

provide the instantaneous heat flux and temperature difference [20]. Experimental time­

temperature data are converted into heat flux information using a finite-difference calcu­lation performed by a digital computer [20] which computes the instantaneous, local temperature at 10 special points within the sphere and at time intervals of 0.001 seconds.

Experimental results for boiling heat transfer at both standard gravity and at low grav­ity are shown in Fig. 4 and compared with the results at 1-g of other investigators. At the point of maximum heat flux the measured values of (q/A) agree well with that predicted by Noyes [18].

The low-gravity results in Fig. 4 indicate the gravity sensitivity of film boiling, transition boiling and the maximum heat flux. This behavior appears to be a conse­quence of the buoyant force controlled character of these processes. Of particular inter­est and significant of the results in Fig. 4 is the similarity of the l·g and low-g nucleate boiling data. Similar results are reported by Sherley [12], Siegel and Usiskin [10]. On the basis of a buoyant force controlled bubble departure mechanism, such gravity insen­sitivity would not be expected, which suggests that nucleate boiling processes are govern­ed by a mechanism other than buoyancy. It is natural to examine other effects and atten­tion is readily drawn to the forces present in bubble growth.

Adelberg and Forster [11] have estimated the bubble dynamic force associated with bubble growth and formulated a comparison between this force and the buoyant force. The ratio of these two forces is of the nature of a bubble Froude number.

349

II-10

0 fT1 ::u 0 fT1 0 "Tl .,, fT1 .,, ::u fT1 "1

::u z 0 (') fT1 0 z c.. (/) c ..... Z )lo (') z -i -i - > O z 1 z

Fig. 3. Sanborn oscillographic record of temperature within r-inch (25.4 mm) diameter copper sphere during free fall.

350

II-10

1 0 ' �----------------------V-

P-RE

_S_

E-NT

_D_AT

_A_

o/-

g=�-

o-

.o-

1,_

1_1 2

_1_

N_CH

_D

_I A

�.

(q/A ) BY EQUAT I O N ( 7 ) , R E E ( 1 8) , o/o "l ••

- - a--•.-- ­.

• .. O /Q = I __..,. •v

(q/Al , EQUATJON (7 ) 1 REF. ( 18), o/g = 0.01 e V

• •

/ v · v,,.&�� ..

•

- .,r;l.. - - -o-- \:--1 0 \ • v I

v '° 0 \ '/:, v 0 /o 10 I

I / o •

I·

• PRESENT DATA o/g� 1 1 112 INCH DIA.

• PRESENT DATA o/g f:t I, I INCH CIA. 0 PRESENT DATA o/o � 0.01, I INCH DIA.

D H S U AND WESTWATER ( 2 1 ) 6 BROMLEY (22) O> RUZICKA (23) A-A EQUATION (18), REF. ( 14 ) Q PRESENT DATA o/g = 0.20 e PRESENT DA TA a/Q = 0.3 3 G PRESENT DATA o/Q = 0 6 0

"Fig. 4. Boiling heat transfer to liquid nitrogen at atmospheric pressure at various sub-terres ·

trial gravities.

The results have been extended [15] and the bubble Froude number computed for several cryogenic liquids in pool boiling as shown in Table 1 .

Table 1 . Froude Number for Nucleate Boiling Liquids

p = 15 psia (775 mm Hg), a/g = 1, Lit = 16° F (8.9°C), R (0.127 mm)

Liquid Froude Number

452 546 352

13,900

0.005 inch

As is evident from this calculation the importance of inertia forces far outweigh that of buoyant forces in nucleate boiling. Hence, it follows that the heat flux in nucleate boiling is governed by a liquid flow pattern resulting from the dynamics of bubble growth. In sub-cooled nucleate boiling these effects will be greater.

Fractional gravity film boiling data are shown in Fig. 4. These results are correlated by the analysis of Frederking, et al [14].

HEAT TRANSFER AT HIGH-GRAVITY

Nucleate boiling heat transfer data for liquid nitrogen were obtained in the range from standard to 20 times standard gravity. A centrifuge was used for this purpose, the details of which are reported in Ref. [19]. It was modified slightly to accomodate liquid nitrogen for the present measurement. Results of these measurements are given in Fig. 5. As can be seen, non-boiling free convection is gravity sensitive as is the region of incipient

351

11-10

40,000�--------------�

30,000 0 O/Q :: I + a/g = 5 o a/g = 10

20,000 • 0/'11 = 1 5 t:. a/g = 20

1 0,000

8,000

6,000

4,000

2,000

1,000

800

BOO

' I ,JAY !?11 /!1 j

lf ,, / J/d l ;/I I J NON-BOILING ' I I CONVECTION ;�1 / ""� 0.05051G, .. J0'" , /!/ 111 /1 1 1 '/ I Jj1/ '/ I I ; 4oo2!-L---'--!c---LL--J.--.,___.,,,___.,8,__,.9�1�0--�,5--,!20·

fyt - fSAT•o F

Fig. 5. High gravity nucleate boiling data for liquid nitrogen at atmospheric pressure.

boiling (Llt 6 to 10°F or 3.3 to 5.5°C). This is consistent with the authors' water data [19] and believed to be a result of an intensified, superposed free convection in this region of low bubble population. These data suggest, furthermore, that at Llt less than 10° F (5.5°C) and a/g ;;;;10, nucleate boiling may even be suppressed.

Of significance in these data is the relative insensitivity of nucleate boiling to high grav­ity fields. This also was found with high gravity water data [19] as well as with the essentially zero gravity nitrogen data shown in Fig. 4. Taken together all these data point to a nucleate boiling or bubble dynamic mechanism which is largely or completely uninfluenced by a gravity field. Such a condition is what would be expected to produce the large bubble Froude numbers in Table 1 which represents a possible explanation of this phenomenon.

ACKNOWLEDGEMENTS

The authors wish to express their appreciation to Messers. H. G. Paul and C. C. Wood of the NASA, George C. Marshall Space Flight Center for their encouragement and assistance in the course of the research.

NOMENCLATURE

a system acceleration - ft/sec2, cm/sec2 A - heat transfer surface area - ft 2, cm 2

352

11-10

g standard (terrestrial) gravity-32.2 ft/sec1, 980 cm/sec• Gr Grashof Number - (1) Nu Nusselt Number - (1) Pr Prandtl Number - (1) P system pressure-psia, mm Hg q heat transfer rate - Btu/hr, Watts/cm2 R bubble radius-ft, mm Lit t w-tsat, ° F, °C t w heat transfer surface temperature-°F, ° C tsat fluid saturation temperature corresponding to pressure at heat transfer

surface, ° F, ° C

Table 2 . Conversion Factors and Temperature Calibration

A. CONVERSION FACTORS :

I . To convert (q/A) to watts/cm 2 divide (q/A) in BTU /hr-ft 2 by 3180 2. To convert Lit to °C divide Lit in °F by 1.8 3. To convert length to meters multiply length in feet by 0.3048 4. To convert length to mm multiply length in inches by 25.4 5. ° C = 5/ 9 (°F - 32)

B. TEMPERATURE CALIBRATION FOR FIG. 3 (Ice point reference)

Millivolts Temperature ° F

-4.9 -258.0

-5.0 -266.7

-5.1 -275.6

-5.2 -284.8

-5.3 -294.5

-5.4 -304.6

-5.5 -315.0

-5.55 -320.5

REFERENCES

r . Results of the First United States Manned Orbital Space Flight - February 20, 1962 - Manned Spacecraft Center, NASA.

2. E. T. Benedikt, Editor, Weightlessness-Physical Phenomena and Biological Effects, American Astronautical Society, Plenum Press, New York, 196r .

3. June-August (1961) Progress Report for the Combined Laboratory and KC-135 Aircraft Zero-G Test Program, AE 61-0871, Convair-Astronautics, September 5, r96r.

4. R. G. Clodfelter, Fluid Studies in a Zero-Gravity Environment, ASD-Propulsion Laboratory, ASD Technical Note 6!-84, June r96r.

5. W. C. Reynolds, Behavior of Liquids in Free Fall, Journal of the Aero-Space Sciences, Vol. 26, No. 12 , December 1959. See also Ref. 2 .

6. ] . Cary Nettles, NASA Lewis Laboratory, Private Communication, June r96r . 7 . T. Li, Liquid Behavior in a Zero-G Field, Report No. AE 60-0682, Convair-Astronautics,

San Diego, September r960. See also, Advances in Cryogenic Engineering, Vol. 7, K. D. Timmerhaus, Editor, Plenum Press, 1962.

8. C. M. Usiskin and R. Siegel, An Experimental Study of Boiling in Reduced and Zero-Gravity Fields, Trans. ASME J. of Heat Transfer, August, 196r . See also, Ref. 2 .

9. H . F . Steinle, An Experimental Study o f Transition from Nucleate to Film Boiling Under Zero-Gravity Conditions, Proceedings r960 Heat Transfer and Fluid Mechanics Institute, Stanford University, June 1960. See also, Ref. 2.

ro. R. Siegel and C. M. Usiskin, A Photographic Study of Boiling in the Absence of Gravity, Trans. ASME J. of Heat Transfer, August 1959·

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r r . M. Adelberg and K. Forster, Discussion of Ref. 8. See also Ref. 2 . 12 . john E. Sherley, Nucleate Boiling Heat Transfer Data for Liquid Hydrogen at Standard and

Zero-Gravity, Advances in Cryogenic Engineering, Vol. 8, K. D. Timmerhaus, Editor, 1963. 13. H. Merle and ]. A. Clark, Boiling Heat Transfer Data for Liquid Nitrogen at Standard and

Near-Zero Gravity, Advances in Cryogenic Engineering, Vol. 7, K. D. Timmerhaus, Editor 1962.

14. T. H. K. Frederking and ]. A . Clark, Natural Convection Film Boiling on a Sphere, Advances in Cryogenic Engineering, Vol. 8, 1963 (K. D. Timmerhaus, Editor).

15 . ]. A. Clark and H. Merte, Nucleate, Transition and Film Boiling Heat Transfer at Zero Grav­ity, Paper presented at the Second Symposium on Physical and Biological Phenomena Under Zero-G conditions, Sponsored by the American Astronautical Society, Los Angeles, California, January 19, 1963.

16. H. ]. Ivey, Preliminary Results on the Effect of Acceleration on the Critical Heat Flux in Pool Boiling, Reactor Development Division Report, AEEW-R 99, AXE, Dorchester, Durset, England, September 196r.

17. R. W. Graham and R. C. Hendricks, NASA Report TN D I I96, Lewis Research Center, Cleveland, Ohio, USA (Movie Film Available from authors on request).

18. R. C. Noyes, An Experimental Study of Sodium Pool Boiling Heat Transfer, ASME Paper 62-HT-24, AIChE-ASME Heat Transfer Conference, Houston, Texas, August 1962 .

19. H. Merle and ]. A. Clark, Pool Boiling in An Accelerating System, Trans. ASME, Journal of Heat Transfer, Vol. 83, No. 3. August 1961, p. 233.

20. H. Merle and ]. A. Clark, Boiling Heat Transfer with Cryogenic Fluids at Standard, Fractional and Near-Zero Gravity, Paper in preparation, Heat Transfer and Thermodynamics Labo­ratory, Department of Mechanical Engineering, University of Michigan, 1963.

2 r . Y. Y. Hsu and ]. W. Westwater, Film Boiling from Vertical Tubes, AIChE Journal, Vol. 4, p. 58, 1958.

22. L. A. Bromley, Heat Transfer in Stable Film Boiling, Chem. Engineering Progress, Vol. 46, p. 221 , 1958.

23. ]. Ruzicka, Heat Transfer to Boiling Nitrogen, Problems of Low-Temperature Physics and Thermodynamics, Pergamon Press, p. 323, 1959.

DISCUSSION

C. F. Kayan, U.S.A. : Was the inertia effect (transient characteristic) of the tempera­ture-measurement system explored ?

The authors : Yes. The thermocouples were constructed of 30 gage wires (.010 inch diameter) by spark welding together and then attaching it to the bottom of the drilled holes with a minute amount of solder. Any lag in the response of the thermocouples to changes in temperature of the sphere at the point of attachment would be due primarily to the presence of this solder. A conservative evaluation of this lag leads to a time constant on the order of 2 millisec for the solder and thermocouple, a negligible value when compared to the maximum rate of change of temperature which actually takes place. A more complete discussion of the experimental apparatus connected with the fractional gravity work is presented in reference 20 of the paper, available as : H. Merte and J. A. Clark, "Boiling Heat Transfer with Cryogenic Fluids at Standard, Fractional and Near­Zero Gravity". ASME Paper No. 63-HT-28, presented at the ASME-AIChE Heat Transfer Conference, Boston, Mass. Aug. 1 1-14, 1963.

J. J. Kowalczewski, Australia : In the Froude number there appears a characteristic velocity and length. Could Prof. Clark state how these quantities were defined in the bubble Froude number?

The authors : The characteristic length and velocity in the Froude number are bubble radius and time rate of change of bubble radius, respectively. Each of these is, of course, time dependent. However, in our reference 15, it was shown that for bubble growth in a uniformly superheated liquid (a conservative approximation for the present case) the Froude number could be expressed in terms of a fluid property group, Lit, and the bubble radius. This was a consequence of the fact that under these conditions the product R (dR/d6J), where e is time, is a function only of LI t and fluid properties. For the calculations in table 1, the bubble size chosen was approximately that of bubbles as they exist near the surface during their growth at a/g = 1 and hence still large enough to influence the process.

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B. W. Birmingham, U. S.A. : As I recall Miss Sherley and her co-workers conducted a few experiments with liquid hydrogen which extended into the film boiling regime. Do your results agree with her data in this regime ?

The authors : We have not received Miss Sherley's film boiling data on hydrogen and hence are unable to make any comparison. However, our film boiling data at fractional gravity can be correlated in terms of the Grashof, Prandtl and Nusselt Parameters for Turbulent Free Convection, which is in agreement with other data for classical fluids.

W. H. Emerson, U.K. : In the experiments described in this paper to determine the rates of boiling heat transfer in low gravity fields, the boiling regime was initially of the film type, but eventually became nucleate as the sphere cooled. I was at a loss at first to see how, under virtually no gravitation field, the film of vapour was removed from the sur­face so that the transition to nucleate boiling could become possible. The answer is of course that boiling was taking place before the gravitational field was removed, and strong convection currents were present in the liquid. During the second or so of free fall these currents would lose little of their momentum and would serve to strip the vapour film from the sphere's surface. However, the purpose of the experiment is to throw some light on the behaviour of boiling liquids in space vehicles, in which the duration of weightlessness is not measured in seconds but in days. Under such conditions the convection currents which may have been promoted under gravity are not maintained and no natural forces are present to remove the vapour from the surface. To this extent, therefore, that part of the boiling curve in the film regime appears a little unrealistic.

The authors : During free-fall experiments (essentially zero-g) from an initial condition of film boiling the heat transfer rate decreased immediately, to near-zero, indicating that the vapour film thickened around the sphere rather than being "stripped off", as Mr. Emerson supposes. Even under this severe condition, the momentum of the liquid resulting from the initial lg boiling apparently is insufficient to influence the vapour film. We agree with Mr. Emerson that during sustained low gravity flight in space a condition of film boiling will likely prevail if the heat flux rates are high and no vapour removal means are provided. It should be remembered, however, that at low heat flux nucleate boiling will exist and the dynamics of the nucleating bubbles will create their own cir­culation field which will promote the removal of vapour from the surface. The degree to which this mechanism will be effective must await carefully executed long-time low gravity experiments.

In the nucleate boiling regime, these experiments have shown that the process is essentially uninfluenced by gravity forces. Hence, the results should be realistic even for long-term low gravity if some means can be provided (if it is shown to be necessary) to remove the vapour once the bubbles have left the region of the surface propelled by their own inertia forces.

A. P. MacKenzie, U.S.A. : In the time that the sphere takes to fall through the liquid nitrogen, how much nitrogen gas is formed around the sphere, expressed as a fraction of the volume of the sphere ?

The authors : We understand you to refer to the volume of nitrogen gas formed during the 1 .4 second free-fall. This amount depends upon the heat flux (i. e. boiling regime). For nucleate boiling such as shown in the oscillographic record in Fig. 3, the volume generated is approximately 25 times the volume of the sphere. At film boiling with high temperature difference the volume produced would be much less and is estimated to be about 1 to 10% the sphere volume during the 1 .4 second free-fall.

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Control and Economy of Air Cooled Refrigeration Condensers of Mean and High Outputs

Reglage et economie des condenseurs frigorifiques a air de rendements moyens et eleves

Dr. Ing. K. A. F. BECK Firma GEA Luftkilhlergesellschaft, Happel GmbH & Co., Bochum, Germany

SOMMAIRE. Le manque croissant, la mauvaise qualite et /'elevation du prix de l'eau de refroidissement necessitent l'utilisation de condenseurs a air meme pour /es installations frigorifiques ayant des temperatures de condensation de quelques degres seulement plus elevees que la temperature de l'air ambiant.

Les avantages des installations a air sont les suivants: economies sur /es frais d' exploitation, elimination du coat du nettoyage, securite et peu d'entretien.

Les condenseurs peuvent etre adaptes a to us /es cas en f onction des conditions locales. Un reg/age precis est necessaire pour la pression de condensation, car les distributeurs des

evaporateurs ne f onctionnent avec precision qu' a un certain rapport de pression. Pour ces raisons, /es installations a air offrent diff erentes possibilites.

Un reg/age par paliers economisant du courant est possible avec utilisation de moteurs a changement de pole OU par remp/acement des Ventilateurs simples par Un groupe a plusieurs ventilateurs.

Reglage automatique a !'aide de persiennes qui sont regUes par la pression de condensation. Le reglage le plus precis et !es plus economique se f ait par orientation des aubes de ven­

tilateur pendant la marche. Avec ce moyen, !es quantites d'air peuvent correspondre a la charge du condenseur et a la temperature de l' air ambiant, donnant ainsi une temperature de condensation constante. L'orientation des aubes de ventilateur peut se faire avec des appareils electromecaniques ou pneumatiques en relation avec la pression de condensation.

Until now water has mostly been used as cooling agent for refrigeration condensers as it was readily available and involved relatively small heat exchange surfaces because of the high heat transfer coefficient. Only the condensers for small refrigeration plants for household and commercial purposes had been already air cooled since a long time. The extraordinary quick industrial development has led during the last years to a consider­able increase of water consumption and today the available ground water is already insuf­ficient at many places to satisfy the increasing demand of the industry.

Moreover it is already prohibited by law in many cases to take water from rivers and lakes, to avoid a further sinking of the ground water level and heating of rivers water. The capital investments necessary for supply and cleaning of cooling water leads to increasing water prices, which become a decisive factor in the operating costs for that reason.

But also the increasing fouling and heating of the rivers keep on deteriorating the water quality. It has to be added that untreated fresh water contains aggressive and hardening constituents, which lead to corrosion and scale deposit and make frequent cleaning necessary.

In contrast to this air cooling has many advantages : Air is available everywhere without any charge. Scale and corrosion on the coolant

side are no more possible. In air cooled plants no cooling water can seep into the refri­geration cycle. Cleaning of the cooling surfaces at the cooling air side can easily be car­ried out during operation of the plant by blowing out the surfaces with compressed air or steam. This guarantees an undisturbed operation and avoids a possible shutting down of the plant. The energy costs for the fan drives of air cooled plants are lower than the costs for supply and treating of cooling water for water cooled heat exchangers.

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The air cooled heat exchangers operate also the more economically the higher the temperature difference between condensing temperature of the refrigerant and the ambient air temperature. If an alteration in the condensing pressure is allowed, a con­siderably lower condensing pressure can be obtained during winter by air cooling in contrast to water cooling. The water temperature can fall at most to approx. +5°C. The considerably lower cooling air temperature may bring during winter a higher effi­ciency for the refrigeration plant.

It is necessary for the erection of a condenser, that the distance between the compressor and the condenser is as short as possible, so that the expenses for the piping are not too high. For industrial plants this is no problem, as the refrigerator can always be arranged in such a way that the air cooled condenser can either be installed outside or on top of the building. Till now the refrigerators for air-conditioning plants were mostly installed in the basement. In view of a possible saving of basement area, the refrigeration plants are today often installed on the roof. These are the erection possibilities of an air cooled condenser. It can be erected on top of the roof; cooling water inlet and outlet pipes are not necessary. They would be necessary for a water cooled condenser. Furthermore there is no danger of freezing of the water pipes. [ifli There are three different constructions for the arrangement of air cooled refrigeration condensing elements. The elements may be arranged vertically, horizontally, or in roof type. A horizontal arrangement (according to Fig . . 1) is .. the simplest and most econo-

-== ·--_,_ .-:;-I

/.

\

Fig. r. Air cooled refrigeration condenser in horizontal construction. Fan drive by multi polar, slow-running fans, fan blades adjustable when fans are in operation.

mica! one. The axial fan for moving the cooling air is arranged below the condensing element and forces the cooling air through the fin tube surfaces. The ratio of the heat transfer coefficient of the outside and the air side is for Frigen 1 : 10 and for ammonia approx. 1 : 50. To compensate these differences in the heat transfer coefficient, the application of fin tubes is necessary for economical reasons. The motor for the fan drive is arranged below the axial fan and is accessible for maintenance during operation. This diapositive shows a special drive, which will be discussed later on.

The following Fig. 2 shows vertically arranged heat exchange elements. This means, that the required ground space is much smaller. The axial fan can be of the forced or induced type. For this element arrangement, however, there is the danger of wind in­fluence, as by strong air velocities the velocity, produced by the axial fan, may compen­sate the air velocity. The figure shows an ammonia-condenser with a heat duty of 400.000 kcal/hr; the condensing temperature is 34°C, design air temperature 20° C, power requirement approx. 2 x 6 kW max.

Fig. 3 shows a condenser of the roof type, which is very often applied. The re­quired ground space is only 60 % approx. of the space that is necessary for the horizontal arrangement. To provide the elements with steam, resp. vapors, one steam pipe on top of the roof is sufficient. The heat from the compressor is transferred by operating the

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Fig. 2. Ammonia Condenser in vertical construction. Q = 4000,000 kcal/h, Condensing tern· perature 34 °C, design air temperature 20°C.

Fig. 3. Air cooled condenser in roof type construction. Q = 2.ro• kcal/h, Condensing temperature 40°C, design air temperature 25°C.

first elements, which are connected in series, with high steam velocity, to obtain high heat transfer coefficients. The condensing elements are connected in parallel. The two axial fans on this figure are driven over reduction gears, the fan drive lies inside the roof structure. The heat duty for this plant is 2.000.000 kcal/hr for F 12, at a condensing temperature of 40° C and a design air temperature of 25° C.

Important for the design and economy of an air cooled condensing plant is the chosen air temperature. Unfortunately, however, for refrigeration plants, the maximum output of the refrigerator falls together with a maximum ambient air temperature. For especially during the summer months the refrigeration plant is in full operation. That means, that the design air temperature has to meet the max. annual temperature. It is, however, uneconomical to choose for the design air temperature the absolute maximum of the occurring temperatures, for this maximum occurs mostly only for a few hours during the year.

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Therefore one chooses as a rule a temperature, which will be exceeded for approx. Yi% of the annual operating hours, corresponding to 40 hours per year. For these few hours a small output loss is acceptable. The temperature difference between the condens­ing temperature of the refrigerant and the design air temperature should not be less than 10°C because of economy. The precise limit of economy depends upon costs and quality of the available cooling water and energy costs. The economical limit for the temperature difference differs from case to case. For industrial plants, which need continuous main­tenance, one cannot choose for economical reasons as design air temperature the max. dry bulb temperature, but the max. wet bulb temperature. One then designs the plant for an air temperature, which lies few degrees above the max. wet bulb temperature. If this wet bulb temperature is exceeded at the dry bulb thermometer, water is sprayed into the air flow in front of the axial fan. By a fine distribution of water its latent heat cools the air to its limiting cooling temperature.

Of major importance for the application of air cooled refrigeration condensers is the question of noise intensity. There are several sources of noise for the axial fans incl. drives. The axial fan produces, by turbulence and separation of flow at the blade tips a sound level which is approx. proportional to the cube of the tip speed. There are in addition noises in the fan by separation of flow at the fan casings or at the supporting structure for the gears. These fan noises can be influenced by proper choice of the tip speed and be mostly kept within the required tolerance limits.

In order to reduce the motor speed to the fan speed, application of gears is very often necessary. Such gears have noise levels between 80-95 db, depending on speed and construction. To avoid the operation noises gear drives can either be replaced by belt drives, which is possible up to certain fan diameters, or by slow-running motors, which operate with the fan speed. The motor noise level mounts up to 60-85 db and is con­siderably increased by the built- in cooling fan of the motor.

Summarizing one can say that the usual noise requirements can be met by correspond­ing choice of fans incl. drives and proper design. Important for refrigeration plants is the precise control of the condensing pressure. Independend from the charge of the refri­geration plant and temperature conditions of the coolant in the condenser, the condens­ing pressure has to be kept constant. The sprayer valves for the evaporators operate only precisely at a constant pressure ratio. Besides this, the control of condensing pres­sure is necessary for saving energy costs. There are many possibilities to control the condenser.

1. Stepwise control is possible by application of a great number of small fans or by application of polechanging motors for the fan drives. This control is power-saving, but the condensing pressure mostly cannot be kept precisely constant, as by switching over the motors or single fans, deviations occur in the heat duty of the condenser and in the condensing pressure. This control is economical and cheap but only suitable for large plants.

2. An infinite control can be obtained by switching off several elements or by partly flooding the condenser. This possibility is exact enough, but complicated in construc­tion; energy costs cannot be saved, as all the heat exchange surfaces have to be continu­ously provided with cooling air.

3. An infinite control is also possible by controlling the cooling air quantity by louvres or control of cooling air temperature by partial recirculation of air. The louvres are put on the heat exchange elements or arranged at the inlet side of the axial fan. By changing the flow resistance in the louvres the cooling air quantity can be varied according to the characteristic of the axial fan. Louvre setting is controlled by the condensing pressure by servomotor, which is pneumatically or electrically driven. The control is very precise, but power-saving very low. During control of partially recirculating air, the cooling air, induced by the axial fan, is continuously mixed with exhaust air from the heat exchange elements, so that a mixture temperature arises, which meets the required condensing pressure. With this relatively precise control power-saving cannot be obtained.

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4. Infinite control by means of infinitely variable motors or gears. By alteration of axial fan speed the air quantity is varied proportionally to the speed. This control is very economical and precise. But the necessary motors or gears are relatively complicated, therefore the stability and the reliability of operation is not always guaranteed.

5. The best control are adjustable fans. The cooling air quantity, moved by the axial fan, can be continuously altered by adjusting the blade angle. This results in a minimum power consumption, as the fan power meets the necessary air quantity.

By adjustment to smaller blade angles the fan produces a lower air quantity under a lower static pressure. However, the fan does not always operate at its best efficiency. But since the power requirement of the fan increases with the cube of the air quantity, this results in considerable power-saving. The blade angle can either be pneumatically or electro-mechanically adjusted. Control of operators i. e. motors is directly done by condensing pressure.

In the already shown Fig. 1 the principle of such an adjustable fan is given. It shows an electro-mechanical adjustment by an electric motor, which adjusts the blade angle over a spindle and a setting arm.

In addition the blade angle can be altered by manual adjustment. In consideration of a low noise level of the fan a very low tip speed had been chosen and the fan is directly driven by a multipolar motor. With it the relatively high noise level of the gear dis­appears. These multipolar motors operate with the fan speed and have therefore at lower frequency a lower noise level.

Fig. 4 shows the annual air temperature curve, valid for the area Karlsruhe. The curve shows the number of annual hours for a certain temperature. The other curve shows how many hours per year a certain temperature is not exceeded. The dotted curve

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Fwl12' 3tJRPlt. � ...... uw 41J"t: Alnlwd Al-- � 2/J"C ---­r-----

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•II •• ·A ·• ·• 0 1 1 1 ·• • • • • 1· • • • ..,. _ _ (Ur)

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Fig. 4. Annual air temperature curve, valid for the area Karlsruhe. Maximum and minimum power requirement for the fan drive at automatically adjustable fan blades.

shows the power requirement for the fan drive in relation to the air temperature of 40° C and an air temperature of 20° C at an output of Q = 800.000 kcal/hr. At a maximum power requirement at fan shaft of 27,5 kW the mean annual power requirement at application of an adjustable fan is only 9,1 kW, that is 33 % of the max. power requirement. This shows that the application of a fan with adjustable blades during operation in conside­ration of the power requirement is economical, especially when for operational reasons the condensing pressure has to be kept constant.

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The Influence of Partial Pressure Difference and Supersaturation on the Frost Formation during the Cooling of Gas -Vapor-Mixtures in Counterflow Heat Exchangers

Influence de la difference de pression partielle et de la sursaturation sur la for­mation de givre pendant le refroidissement de melange gaz-vapeur dans les echangeurs de chaleur a contre-courant

DR.-ING. D. DIBBERN Lehrstuhl und Institut fiir Thermodynamik und Verfahrenstechnik der Tech­nischen Hochschule, Hannover, Welfengarten I A, Germany

SOMMA/RE. Dans les echangeurs de chaleur utilises pour le refroidissement des gaz contenant des vapeurs, il se produit souvent une formation indesirable de givre sur !es murs. Pour etudier la formation de givre, des experiences ont ete effectuees precedemment dans ce domaine au College Technique de Hanovre.

Ces nouvelles experiences montrent que la quantite et la qualite du givre dependent consi­derablement de la difference entre la pression partielle de la vapeur dans le melange et de la pression de saturation a la surf ace de refroidissement. Plus cette difference etait grande, plus !'on observait de givre. Le givre devenait a la fois tri!s grossier et poreux. La repartition du givre dans l' echangeur de chaleur etait regie par la difference de pression partielle.

Avec des differences de pressions partielles egales, un melange sursature f orme davantage de givre qu'un melange non sature. A la surf ace de refroidissement apparaissent des aiguilles et des dendrites, se developpant jusqu' a une certaine grosseur, mais ensuite emportees par le courant du melange. Lorsque la sursaturation atteint un degre critique, on peut voir de la neige et du brouillard au centre du courant.

On a ca/cute les nombres de Nusselt et de Sherwood a partier des experiences et on les a compares aux resultats obtenus avec des equations sans dimensions. Les ecarts entre les expe­riences et la theorie dans les melanges sursatures peuvent s' expliquer par la formation d' aiguil­les et de dendrites.

In heat exchangers used for cooling of gases containing vapors there often occurs an undesired formation of frost on the walls. In addition to this there is sometimes the dan­ger of fog or snow formation in the centre of the tube. As a consequence of frost depo­sition pressure drop increases and heat transfer decreases. Therefore the heat exchan­gers have to be thawed off in regular periods.

In order to study the frost formation in the tubes of counterflow heat exchangers earlier experimental works of Hilz, Linde, and Rische have been continued at the Tech­nical University in Hannover. Some details of this new research have been recorded already by Hausen at the sessions of the International Institute of Refrigeration in Eind­hoven and Cambridge. Meanwhile the work has been completed and some of the latest results and conclusions may be discussed here.

In the experiments benzene was used as vapor and nitrogen or hydrogen as inert gas. The apparatus is schematically shown in Fig. 1. The main part of the apparatus consists of a vertical double tube heat exchanger. In the inner tube, 6 m long and of38 mm inner diameter, the mixture flows downwards. This tube is cooled from outside by cold air passing in counterflow in upward direction through the annular space. Before entering the heat exchanger the mixture passes a hydrodynamic starting length of 1,4 m. The gas vapor mixture is circulated through the test tube by a blower. Benzene is supplied on a bubble tray.

Temperature, flow rate, and the partial pressure of benzene in the entering mixture were held constant during each experiment, as well as the flow rate and the temperature of the entering cooling air. Temperatures of mixture and cooling air at the ends of the

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test tube and the temperatures of the wall were measured with thermocouples. The partial pressure was controlled by interferometer readings. At the end of each experi­ment the overall quantity of frost was melted and weighed. A special equipment of the apparatus is a removable part of the tube of 180 mm length, which may be seen in Fig. 1

H�t exchon er L-1

_.rest tube

-Removoble po.-t / ol the tulie

� Flash �1 tamp -Cooling air

J:•mero

Flow of mixture

BenzeM suff'ly

Fig. r . Schematic diagram of apparatus

near the lower end. After each experiment this part of the tube was removed, so that the frost deposited on the inner surface could be photographed, measured, and weighed.

To be able to photograph fog or snow in the centre of the tube from below by a tele­photographic lense, the removable part of the tube can be lighted from outside by a flash lamp (Fig. 1). In this case the removable part of the tube was made of glass and' a part of the outer wall usually consisting of iron was replaced by a half-cylindrical pie�e ofplexiglass. However in normal runs the formation of fog and snow could be seen clearly by a mirror from the bottom of the tube, if the interior of the test tube was lighted from below.

All experiments were performed with Re-numbers between 1 100 and 1 10000. These Re-numbers were related to the condition of the mixture entering the frost free tube.

CONDENSATION IN THE CENTRE OF THE STREAM

At laminar flow formation of fog and snow in the centre of the stream was observed during many experiments. Fig. 2 represents a photograph of snow in the centre. The bright circle is the lighted glass wall. The annular space between the snow and the tube

Fig. 2. Snow in the centre

wall contains no condensed benzene. As already discussed in Cambridge [1] the conden­sation in the centre of the stream was reproduceable to such a degree, that the critical

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supersaturation necessary for the formation of little nucleus in the centre could be de­termined quite exactly.

From theoretical considerations involving the profiles of velocity, temperature, and partial pressure follows, that for condensation in the centre the mixture must be super­saturated to such a degree, that the vapor is instable with regard to the vapor pressure of undercooled liquid. A comparison with the much higher critical supersaturation based on Volmer's [2] theory of spontaneous nucleation in vapors shows that in a technical apparatus condensation begins at dust particles of microscopic size.

Also in some experiments belonging to the turbulent region condensation in the centre of the flowing mixture was observed. The amount of the critical supersaturation was the same as in laminar runs.

FROST GROWTH ON THE WALLS

It has been found as a final result of the experiments that quantity and density of the frost on the wall depends considerably on the difference between the partial pressure of benzene in the mixture and the saturation pressure at the wall or at the surface of the frost. The greater this difference the more frost was observed; whilst the frost became more rugged and porous.

In a supersaturated mixture even at high Re-numbers there grew needles and dend­rithes on the wall and on the surface of the frost layer. After growing to a certain size these needles were detached from the wall by the pitot pressure. Some of these dend­rithes on the lighted wall and falling frost particles are to be seen in Fig. 3.

The difference of partial pressure which is variable along the test tube was systemati­cally varied in experiments with equal Re-numbers. In the removable part of the tube, where specially the frost layer was investigated, partial pressure differences could not be

Fig. 3. Dendrithes and falling particles of frost

measured directly. Therefore they were calculated on the basis of theoretical conside­rations. The well known equations of heat and mass transfer were used to calculate the coefficients of heat and mass transfer, which are assumed to be constant. With regard to the facts, that the measured temperature of the walls governs heat and mass transfer only at the beginning of the runs and the deposition of frost is neglegible at this time, the calculations were carried out only for the beginning. Moreover it was assumed for these calculations, that the surface of the test tube was still free from frost.

In spite of these restrictions the calculations revealed some very interesting results. One of these is shown in Fig. 4, where the quantity of frost frozen out in the removable part of the tube during 45 minutes is plotted against the calculated initial difference of partial pressure. In this case the Re-number was 5500 and benzene was frozen out from a mixture with nitrogen. The results of runs without supersaturation could easily be represented in Fig. 4 by a slightly bended curve. Runs with supersaturated mixtures delivered more frost at the same partial pressure difference. This is a remarkable state­ment. One should expect, that in the runs with supersaturated mixtures less frost re­maines on the wall, because in these cases many particles were torn off the wall and car­ried away by the flow. An investigation of the amount of supersaturation has shown that

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� 1 " � g � � a1----1----1----1--��l""it'--'"-1 � z Re-5500

� :S NiC5f1> � � 61----1----1----1--���-=i � � � f ,f---t-----�G--�f----f----1 � � ] :S 2•1----+-"''=----1----1----1-------1 6 !';:

10 40 60 80 .!fli __ Initial partial-pressure-difference

1nme remo11abfe part or the tubf

Fig. 4. Quantity of frost deposited in the re· movable part of the tube

there is the more frost the higher the measured supersaturation. A very significant result consists in that in unsaturated flowing mixtures the rate of mass frozen out depends only on the partial pressure difference. In supersaturated mixtures however there is more frost to be expected on the wall in spite of the fact that a considerable amount of dendrithes and needles is torn away. The frost left on the wall is of relative high density.

HEAT AND MASS TRANSFER

Of great importance is the question, whether dimensionless equations, for instance the following ones of Hausen [3] can be used to calculate the heat and mass transfer under frosting condition

wherein

Nu

Sh

Re

Pr

Sc

L d

d Nu = 0,037 (1 + L

2/3) (Re0,75 - 180) Pro,42

d Sh = 0,037 (1 + L

2/3) (Reo,75 -180) sco,42

a d

;. NuBelt number, a heat transfer coefficient,

fl d ;. thermal conductivity,

D Sherwood number, w velocity of Nu mixture,

w d fl Reynolds number, mass transfer coefficient,

v v

Prandtl number, D diffusion coefficient, a v

D- = Schmidt number, v kinematic viscosity,

length of the tube, a thermal diffusivity.

diameter of the tube.

Calculation of the quantity frozen out in the removable part of the tube with help of the second equation as function of initial partial pressure difference yields the straight line in Fig. 3. At small partial pressure differences this line is in good agreement with the experimental curve for the unsaturated runs. The deviation at higher values can be explained as follows. The more frost is formed during a run, the more the temperature of the frost surface is increased above the initial temperature of the wall. This causes a steady decline of the rate of mass transfer. Fig. 4 serves only as example for all other experiments with varied Re-numbers. The runs made with the hydrogenbenzene mix­ture gave similar results.

The validity of dimensionless equations under frosting conditions can better be checked if one compares the Nu- and Sh-numbers calculated on the one side from Hau­sens equations and those taken from the experiments. Also for this purpose the Nu-

366

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and Sh-numbers have been calculated for the beginning of each run and assumed to be constant. In Fig. 5 the results are plotted against the overall quantity of frost formed during 45 minutes. The dotted straight lines represent the Nu- and Sh-numbers from Hausens equation. At small frost quantities the agreement is satisfactory. With rising frost quantities however in spite of scattering of experimental values there is a remarkable deviation to be seen. The Sh-numbers are greater, the Nu-numbers seem to be smaller than the theoretical ones.

,.r===t===t===-i=-=-=-=-=-=,_i=-=-=-=-=ll-11

150 JOO 6009 750 Fig. 5 . Nu- and Sh-numbers o�,,ral/-quantity of frost frozen out wifhin 45minutes

This statement cannot be explained by the different thickness and roughness of the frost layer. For rough surfaces are expected to influence heat and mass transfer in the same way. It is a characteristic of the experimental results that the rising of partial pressure differences, and therefore the rising of frost quantities is connected with an increase of supersaturation. As already mentioned supersaturation causes the develop­ment of dendrithes and needles on the walls. The heat of sublimation originates on the surface of these particles and has to be transported by heat conduction to the cooled wall of the test tube. In supersaturated mixtures the surface temperature of long thin needles may rise above the temperature of the surrounding mixture. Also in this case mass transfer to the needle takes place. The normal heat transfer direction from the mix­ture to the frost surface however is conversed and the needle itself becomes a heat source.

Sh-numbers calculated from experiments with supersaturated mixtures under the assumption, that the surface is still free from frost, are necessarily greater than the the­oretical ones. However the Nu-numbers may be smaller than the theoretical ones. The influence of enlargement and roughness of the surface caused by the growing of needles and dendrithes is surpassed by heating the mixture with that part of the heat of subli­mation, which is not transported to the wall.

Concluding this paper two other results of the investigation may be pointed out. In this research work two kinds of inert gases were used, nitrogen and hydrogen. Besides other physical properties the diffusion coefficient proved to be of great importance. The observed increase of frost density with time is strongly influenced by the magnitude of the diffusion coefficient. The diffusion coefficient of hydrogen-benzene is nearly four times greater than that one of nitrogen-benzene. Therefore the density of frost in expe­riments with nitrogen was always lower than in the runs with hydrogen. In some runs with hydrogen the frost consisted of clear ice free from gas, whereas the highest density of frost in runs with nitrogen-benzene was about 0,8 times the density of clear ice.

An investigation of the changes of the partial pressure differences along the tubes of a counterflow heat exchanger with locally variable temperatures has shown, that under certain circumstances a maximum of this difference is to be expected within the ex­changer. Therefore in heat exchangers of the investigated kind there is the possibility of blocking by frost in the middle part of the tube while the ends are nearly free from frost.

REFERENCES

I. Commission 2, 3, 6b and 8, Cambridge 1961, Annexe 196!-3 Supplement au Bulletin de l'In­stitut International du Froid-Extrait.

2. A. Volmer, Kinetik der Phasenbildung (1939) Steinkopff, Dresden. 3. Allgemeine Warmetechnik, Bd. 9 (1958) S. 75.

367

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DISCUSSION

H. Hausen, Germany : As supersaturation is greatest on the axis of the tube, conden­sation in the central parts of the tube is very probable.

D. Dibbern, Germany : This, in fact, is to be seen in Fig. 2.

G. Haselden, U.K. : 1 . Is it possible for the suspended crystals existing in the gas stream at the bottom of the

cooled tube to pass through the blower, and round the remaining part of the circuit, without melting and evaporating ?

2. Have you any additional information to show that the suspended crystals are pro­duced within the gas phase by supersaturation and do not in fact form on the wall in the upper part of the tube and then become detached into the gas stream ?

D. Dibbern, Germany : 1. No, all crystals were melted before entering the blower.

2. It was possible in the experiments to have the critical supersaturation only in the centre of the outlet cross section. A slight change in experimental conditions gave rise to the appearance or disappearance of fog and snow. One can clearly observe that snow and fog as shown in Fig. 2 originate in the centre and are not formed on the wall in the upper part of the tube.

H. Jungnickel, Germany : I should like to add some observations made by Ullmann who is working for me on a similar subject. He made studies of the subcooling of a nitrogen­carbon dioxide mixture and found that :

The sub-cooling varies according to :

1 . the relations mentioned in Volmer's book: "Kinetik der Phasenbildung",

2. the stream velocity. With increasing velocity the sub-cooling ratio decreases quickly. As surface tension enters into the above-mentioned equations it is possible to predict

solid or liquid condensation. A paper will be presented at the Dresden Centenary of Mollier's birthday.

D. Dibbern, Germany : In our work, frost formation on the wall itself began without su­persaturation. Perhaps the supersaturation measured by Ullmann would have been observed at much lower velocities. (This fact was stated by Prof. Jungnickel after the discussion. He said that the velocities in Ullmann's work were very low. The condensation in the centre of the stream was not considered in Ullmann's investigation.)

L. M. Hamaker, Netherlands : It was understood that the measurements of the amount of frost formed were made at the removable part of the tube. Did the supersaturation, when it was present, only exist at the entrance of the tube or also at the removable part ?

D. Dibbern, Germany : When entering the test tube the gas-vapour-mixture always was unsaturated. Whether saturation or supersaturation occurred within the tube was governed by experimental conditions.

L. M. Hamaker, Netherlands : Is there always supersaturation then?

D. Dibbern, Germany : No, but if supersaturation originated at some length of the cooled tube it remained up to the outlet.

L. M. Hamaker, Netherlands : Do the heat and mass transfer coefficients OG and {J refer to the temperature and the pertaining pressure at the tube-wall or at the surface of the ice needless ?

D. Dibbern, Germany : The calculated heat- and mass transfer coefficients refer to the temperature and the pressure of the tube wall at the beginning of each run. The thickness of the frost layer on the wall was negligible then.

368

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Infiuence of Oil on Heat Transfer of Boiling Freon 12 (Refrigerant 12) and Freon 22 (Refrigerant 22)

Influence de l'huile sur la transmission de chaleur de Freon 12 (refrigerant 12) et de Freon 22 (refrigerant 22) bouillants

Dr.-Ing. K. STEPHAN Kaltetechnisches Institut, Technische Hochschule, Karlsruhe, Germany

SOMMAIRE. Pour determiner /'influence de l'huile sur !'evaporation des frigorigenes, des experiences ont ete f aites sur le processus de transmission de chaleur dans !es melanges d'huile et de Jrigorigenes.

Pour le Freon 12 et le Freon 22 pur et un liquide contenant moins de 3% d'huile, les coeffi­cients de transmission de chaleur sont a peu pres les memes. Par contre, ces coefficients sont diminues considerablement, lorsque les liquides contiennent de plus grandes quantites d'huile.

L' elevation de la temperature d' ebullition est negligeable, tant que le liquide contient moins de 10% d'huile.

Des observations ont montre que, par suite de !'addition d'huile, une ecume intense se pro­duit dans le Liquide. Cette ecume f ait augmenter le coefficient de transmission de chaleur considerablement tant que les surfaces chauffees sont couvertes de liquide. Tout au contraire, ce coefficient est diminue lorsque l'evaporateur entier est rempli d'ecume.

L'effet provoque par l'ecume s'accroit a mesure que la chaleur emise par l'evaporateur est augmentee. En plus, il depend de la quantite d'huile dans le systeme binaire ainsi que de la temperature d'ebullition. Par consequent, /'influence de la chaleur et de la temperature d' ebullition sur le coefficient de transmission de chaleur est beaucoup plus compliquee pour des melanges d'huile et de frigorigenes que pour d' autre5 systemes binaires.

In refrigeration plants working with oil-lubricated compressors there are always cer­tain amounts of oil circulating in the refrigerant system. In general this oil produces a decrease in the efficiency of the plant. The oil in the heat exchangers often causes a considerable decrease in the heat transfer coefficients.

The purpose of the present investigation is the experimental determination of the effect of different oil-concentrations on heat transfer coefficients if evaporation takes place out­side a horizontal tube or on a horizontal fiat plate. Some previous work in this field has been carried out by Tschernobylski and Ratiani [l], Witzig, Penney and Cyphers [2] and by Worsoe-Schmidt [3] . Tschernobylski and Ratiani performed experiments on oil-Freon 12 mixtures evaporating outside a horizontal tube of 14 mm outside diameter. As they found out, the heat transfer coefficient decreased about 15 % when they used mix­tures with 10 % oil in the Freon 12. For small amounts of oil the heat transfer was not influenced by the heat flux density whereas Witzig, Penney and Cyphers when investiga­ting 1 % mixtures observed a considerable decrease of the heat transfer coefficient for great heat flux densities and a negligible effect for small heat flux densities. Contrary to this Worsoe-Schmidt concluded as a result of his experiments that heat transfer coefficients increase if oil is added to Freon 12.

The results cited above demonstrate, that there are contradictions as to the effect of different amounts of oil on heat transfer in evaporation. Furthermore the influence of the boiling temperature or vapor-pressure on heat transfer coefficients of oil-Freon mixtures has not been explored. For this reason new measurements have been conducted.

In the experiments evaporation took place on a horizontal fiat plate. In a previous pa­per [ 4] it has been demonstrated that heat transfer coefficients outside horizontal tubes and on horizontal fiat plates do not differ considerably from each other, so that the results from the fiat plate measurements may be applied for evaluating heat transfer coefficients when evaporation takes place outside a horizontal tube.

369

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APPARATUS

The apparatus for the experiments is shown in a schematic diagram in Fig. 1. The oil-refrigerant circuit consisted of the test evaporator and a brine-cooled condenser. Evaporator, condenser and the tubes in which the mixture circulated were placed in an insulated room. The temperature of this room was controlled by a contact thermome­ter, which interrupted or closed the electrical circuit of a relay. The relay itself stopped or started the compressor of the refrigeration plant for the room. By this method the tem­perature difference between the room and the test apparatus could be kept very small and heat losses could be reduced to a minimum. Moreover the test-evaporator, the con­denser and the pipe lines were insulated so that the temperature of the oil-refrigerant circuit could not be influenced by the small fluctuations of room temperature.

The test evaporator was a horizontal flat copper-plate of 130 mm diameter and of 20 mm thickness. The test-evaporator was provided with a glass-cylinder for observing and photographing the evaporation process. Heat was supplied by means of a resistance wire, which was insulated electrically from the plate. The electrical voltage was kept constant by a stabilizer. The heat flux could be regulated by an electrical transformer and was measured by a wattmeter. The surface roughness of the test evaporator, which as is well known, is of great importance in boiling heat transfer problems, was determined by means of a profilometer. The mean value of the surface roughness was about 1 µm ; this value may be obtained on very smooth surfaces.

Thermocouples for measuring wall temperatures were embedded directly below the surface of the plate. The temperature of vapor was measured by means of thermocouples. The vapor pressure was determined simultaneously.

Before charging with refrigerant, the entire apparatus was evacuated and thoroughly purged because it is known that even a small amount of air in a boiling liquid would greatly affect the heat transfer coefficient by forming nuclei in the small grooves of the heated surface and thus increasing the number of bubbles.

MEASUREMENTS

Measurements were made with Freon 12 and with Freon 22 mixed with oil. The oil mainly used in the experiments has the trade mark Shell Clavus oil 129. According to the information provided by the German Shell Corp. Hamburg it has the following properties :

density at 20° C

kinematic viscosity at 20°C

kinematic viscosity at 50° C

pour point

0,8813 g/cm3

134,4 cSt

26,1 cSt

-44° C

The first measurements were carried out with oil-free Freon 12. In Fig. 2 the heat transfer coefficients at different boiling temperatures are plotted against the heat flux density. The points plotted in this and in the following figures are average values from five different measurements. The heat transfer coefficient h is given by

h = _J__ Lit

q is the heat flux density, Lit is the difference between wall temperature and boiling temperature. The boiling temperature has been measured by means of thermocouples. The heat transfer coefficients corresponding to a boiling temperature of -25°C were obtained by extrapolation of the measurements. These values are found to be in good agreement with those calculated from an equation by Kruschilin [6] and with the values from experiments by Danilowa and Mastikewitsch [7]. Further experiments on heat transfer in boiling Freon 12 outside tubes have been carried out by Lawrowa [8] in the range of heat flux densities between 200 and 12000 J /m2s and of boiling temperatures between -25°C and -5° C. The values given by Lawrowa are extrapolated to heat

370

V) -i ,_.

� Ul

t �-n S?. Jf. I g, � f

1 · - · -· - · - · -· -· 1 I . n

J 1 :::-, -B r i-n E>

�I j E vaporator

I

I I Heat Exch�nger I I !

I -- .

l ! i I i i . .

L.-r· - · - · - · _J Brine Cooler

I

Ll l l ' Needle Valve

�a� l l

L-...1 L..-.....l C ool ing trap

- · - · -· - · - · - · -,

Condenser

Oil

Thermocouple

Mano ­meter

I I I

H igh Pressure Safety Switch

Thermocoupl e I

f

II-6

fluxes above 10000 J/m2s and plotted in Fig. 2. They are in good agreement with the heat transfer coefficients corresponding to low boiling temperatures. As Lawrowa pointed out there is no influence of the boiling temperature on the heat transfer coefficient. This result is inconsistent with the conception of the heat transfer mechanism in boiling [ 4].

10 +----�---r-��---.--�-�-r-��

I 9 -i------+--____,l---+--+----+---+-____,--;.,...+-+--J B +------+---+----

7-t------+--+---5 -t-----+----+---r----r------ti�-hl"-+��-+-l "U1 o 5 +-----+---+--+---+-n<-4"'--+--.....+-.,...+--+--4 o This paper

-. .,, x Kruschilin [6],-25°C � " and Danilowa -� MasUkewitsch [7]

3 +----+.1��'--hl'"'-f'>�"'+�.F--+---+---+-+-l � Lawrowa [8] -25°C � t s � -5°C

and this paper

- 25°C

.c

c ClJ :Q � 2 +--�--l-----bo'"--------+--'

� c � o w3-1-....,....�-+-�-+�+--+�-+�-+--+--+--+....,+--1 -1: 7 . 70 4 2 3 4

5 6 7 · 104 Heat Flux Density

Fig. 2 Heat Transfers Coefficient h of Freon 12 as a Function ol Heat Flux Density for Different Boiling Temperatures t,

After the experiments with oil-free Freon 12 a definite amount of oil was filled in from a measuring glass into the liquid. The oil concentration is defined as the ratio : weight of oil in the evaporator divided by weight of pure refrigerant. It is given in per cent. It was increased gradually to the amount of 50%. The boiling temperature was gradually increased from -18° C to + 15° C. In Fig. 3 and Fig. 4 the heat transfer coefficients are plotted against the heat flux density.

It can be seen from Figs. 3 and 4 that less than 3% oil in the refrigerant does not considerably affect the heat transfer coefficients, especially not at low boiling temperatures.

If the oil concentration was less than 10%, boiling temperature and vapor pressure were in good agreement with the values for oil-free Freon 12. In this range of oil concen­tration the increase of boiling temperature is insignificant if compared with the boiling temperature of oil-free Freon 12. The differences were below 0,3° C.

As may be derived from Figs. 3 and 4, the differences in the heat transfer coefficients of oil-free Freon 12 and of mixtures with less than 3 % oil are independent of the heat flux density. This result is in agreement with the experiments of Tschernobylski and Ratiani.

For higher oil-concentration the slope of the curves h (q) decreases. This effect is connected with the bubble motion, which itself is the decisive factor for the value of the heat transfer coefficient. As the bubbles near the wall grow work must be done to displace the liquid. If the oil-concentration is increased there is a very oil-rich layer near the wall with higher values of surface tension and adhesive tension, whereas the bubble itself continues to consist of an oil-poor vapor. Because of this the growing bubble has to do more work. The bubble therefore grows more slowly. This effect increases with in­creasing wall temperature and increasing heat flux respectively. Hence the heat transfer coefficient and the slope of the curves decrease if they are compared with the results of oil­poor mixtures.

372

KJ-' 9 8 7 6 5 4

3

2

3

2

Il-6

1 1 I 1 1 'o%-I �

� [m2·� . 0c J i,.,0% E(in/s· 0c]

� v 1%= � v 3% � ... � .. 3%

I � !A v 1 % / ]JI � I I� � II' !,..-' 6 % I/ � v � .. �01� � ...

� � l/ / Id" � D"" [/"' /_/ / ... / b ,,.� i.o"" / 17 /. � v

/ � � j.t( 9°10 '�/ � !/ L/ I / .A i.. 9%

� v / I 7 /D y' I I

� ../' 20% SY � 20% / / ;T .,,.,.. ""' I I v ..,... y / l...---' / y _,,,,.,.. � 30% _;;J 9. - 30%

,,,, �- - -- ..- u -� v !...-- � - !.-

� ....... --- .-4" n- --�- �-

� ..... 50 % .1.....-' 50 % � .,_ """' -

� 1..--" - -!....-�- Lo---"-

t5 = 4 °C fs = 15 °C

2 3 4 s 5 1. 1o4 no-' 2

Heat Flux Density -- q [m�. 5 J k Oi l Oi l Con cent ration = _..;.;.;g;;z__..;;;..:.;;.__ ___ _

kg l iquid Freon 12 Fig. 3 Heat Transfer Coefficient o f Oil-Freon 12 Mixtures as a Function o f Heat Flux Density

for Different Oil Concentrations. ta=Boiling Temperature.

373

I

-c: .9! u

Il-6

10 " i 9 1 B I 7 7{ J 11

6 ,__ m2· s "'C 5 I

l " 3

� � Po"'"

A / 2

/ "" v

_ 3% 'l' 0% ,L, 1%

�% � A � 9 % L..

A � _./ � , v � v

• I I h [*-c] 0 01.

� 30 0_ I ./: 1 'lo o/o_ I � � � 6°/o

�

� [/ ,

/. v 9 % ,,& ii"'� I/ /

� � .. v /

I� v I/ L.P' v � 20%

3 / / v 20% � J.,.oO' 30?o i......-- i.-- 30% -� � 10

0 u 9 8 7 � 6 � ....

" � 3

2

10 2 1· 1o4

_ ... W""' v --= I n � ._..,.,.. �� � -

- � ---� "- 50% .-fr' 0- 50 O/o

-� i.o-' .... � o -�

2

I 1�-i

fs = - 18 °C --

3 " 5 6 7"10 1'10 -' 2 Heat Flux Density q [ J J mrs

kg Oi l

i ts = - 7 ° C

3 " 5 6 1104

Oil Concentrat ion = kg l iq u id Freon 12 Fig. 4 Heat Transfer Coefficient o f Oil-Freon 12 Mixtures a s a Function o f Heat Flux Density

for Different Oil Concentrations. t1=Boiling Temperature.

374

II-6

INFLUENCE OF FOAMING ON HEAT TRANSFER

It can be seen from Figs. 3 and 4 that the heat transfer coefficient of oil-free Freon 12 at a boiling temperature of -18°C is smaller than that of Freon 12 containing 3% oil. At boiling temperatures of -7° C, the heat transfer coefficients of a 1 % mixture are smaller than those of a 3 % mixture. Hence in some cases heat transfer coefficients in evaporation increase if oil is added to the refrigerant. This result may be explained by observation of the boiling mixture. When oil is added to the refrigerant an intensive foa­ming is produced. In previous experiments Kirschbaum [9] has demonstrated a consi-

5 6

7

Fig. 5 Boiling Freon 1 2

q = 4,25 . 10< _l_ ; ts=15° C m2 · s

Fig. 6 Boiling Freon 12 with 3% Oil

q = 4,25 · r o • _l_ ; ts=15° C m2 . s

Fig. 7 Boiling Freon 12 with 50% Oil

q = 4,25 . 104 _l_ ; ts=15° C m2 . s

375

11-6

derable increase in heat transfer coefficients in foaming liquids. He observed a threefold increase when adding only 1 % Erkantol to boiling water.

Some systematic experiments on the effect of foam gave the following results. Foaming increases with heat flux density but it decreases when the boiling temperature is increased. It increases too when the oil-concentration is increased, attains a maximum value at oil­concentrations of about 6 to 10 % and decreases at higher oil concentrations. The latter effect is demonstrated by Figs. 5, 6 and 7. The observations are in agreement with the results of Figs. 3 and 4, where in some cases heat transfer coefficients are increased with increasing oil-concentration.

The above cited observations offer an explanation of some other surprising results. By comparison of the curves for different boiling temperatures but for the same oil-concen­tration which are plotted in Fig. 8 it is seen that the heat transfer coefficient does not in all cases increase with increasing boiling temperature. This result is quite different from results obtained by the boiling of pure liquids. Fig. 8 shows that the heat transfer coefficients of mixtures with 9 % oil are smaller than those of a 3 % mixture. But as a result of the strong foaming effect in mixtures of 6 to 10 %, especially at low temperatures, the heat transfer coefficient of a 9 % mixture has its highest value at the lowest boiling temperature. If the amount of oil exceeds values of about 10 % the foaming effect decreases and the heat transfer coefficient increases then with increasing pressure ; but the distance between the curves h (q) becomes smaller, so that there is only a little influence of vapor pressure on the heat transfer coefficients. This effect is caused by the mechanism of bubble formation. As has been stated above the layer near the wall is an oilrich one. If

I 10'��������������������

9 +-�-�-t-�--+---��--+---�--+--�+---+---+-+-1---t-��1 6 -t----��-t-�--+---�-�--+---�--+--�+---+---+-+---1---+--�----" 7 +-��-t-�--+---��--+---�--+--�+---+---+-+---1---+--�--j 6 -t----��-t-�--+---��--+---�--+--�+-�"--++

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] ro--t----""-7"""+------------i���P""'"'f--------::±__.��"--+''+-�--1 9 8 +-'�=--'1':--:�.,...,':::..-1'!1"'P'---C.f---l--f--�-+-l----� 7 ����+=---l------l---l�-l-----1-1---1-----t----f-----l 6�""'--------1--5

2 Heat

3 ' 5 Flux Density

6 7 8 9 0 2

q [;;;is]-Fig. 8 Heat Transfer Coefficient o f Freen 1 2 Oil. Mixtures as a Function o f Heat Flux Density·

t1 = Boiling Temperature.

376

v4

9 8 7 6

5

4

3

2

3

2

10 ,(.

11-6

the bubble grows, the refrigerant evaporating through the oil layer into the interior of the bubble has to overcome a great resistance of diffusion. The velocity of diffusion itself is almost independent of the vapor pressure. The velocity of bubble growth and hence the heat transfer coefficients are practically independent of the boiling temperature.

By comparison of the results with those of Tschernobylski and Ratiani it is seen from Fig. 9 that the experiments described in this paper have been conducted with higher

0 % -! � lo" 3 % -�h[in/s · 0c] / A � 1 %

-· · � 6 % # ,.. ,/ ..

I i � � / I I i / I / fi / , o� � / L..- 9% --I I / I � " I I

I � r � '/ I / ! � I "/' � I / / / 1,....... ,,,,,,. 2 0 % r"O%

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,/ r/' . "" .. ,.. 28 %

5 0 % -· ' --..... ...... . ..... 37 % .--..........

:,.. · .. . ........ L--' I . ' - . / � 45 % .....----- l I / � i,.....- · - 52 °/o i � · -- · I -� · · ...... .-- : � • - 58 °/o I _... '

-::::. �-.... .

103 . 3 4 5 6 7 8 9 104 2 3 4 5 6 7 8 9 1

Heat F lux Density

Wors�e - Schmid t ts = 0 °C

Tschernobylski ts =-13 °C 7-+16 °C

this paper t s = + 4° C

Fig 9 Comparison with Results of Tschernobylski and Ratiani [r] and of Worsoe-Schmidt [3]

377

I

Il-6

heat flux-densities. The values join continuously those of Tschemobylski and Ratiani. Nevertheless it is difficult to compare the values because Tschemobylski and Ratiani did not indicate the boiling temperature corresponding to their curves and from their work it is not evident if the heat transfer coefficients were calculated with the boiling temperatu­re of oil-free Freon 12 or with the boiling temperature of the oil-Freon-mixture. The cur­ves in Fig. 9 are also in good agreement with that published by Worsoe-Schmidt, which has been determined in boiling heat transfer experiments inside tubes.

EXPERIMENTS WITH ANOTHER OIL AND WITH FREON 22

Some experiments were performed with another oil and Freon 12 and with the new oil and Freon 22. The new oil has the trade mark Fuchs-KM (producer Fuchs-Mann­heim). It has the following properties [5] :

density at 15°C

kinematic viscosity at 20° C

kinematic viscosity at 50° C

pour point

0,910 g/cm3

95 cSt

19 cSt

-48°C

By experiments it has been found out that the heat transfer coefficients agree very well with those of Figs. 3 and 4, if there is only a small amount of oil (less than 6%) in the Freon 12.

For higher values of oil concentration, a greater difference in heat transfer coefficients between the two different oil-Freon 12 mixtures has been observed. This fact is probably due to the different physical properties of the two oils, which evidently are of some impor­tance at high oil concentrations.

All relations described above between heat transfer coefficient, boiling temperature and oil-concentration could be confirmed by the new series of experiments.

Even the experiments with soluble oil-Freon 22 mixtures gave results qualitatively similar to those in the experiments with Freon 12. The only difference was that heat transfer coefficients of oil-free Freon 22 and of Freon 22 with less than 6 % oil are about 20% higher than the corresponding values in the Freon 12-experiments. If the amount of oil in Freon 22 is increased the differences between the Freon 12 and the Freon 22 values become smaller. This is a result of the increasing influence of the physical pro­perties of oil on heat transfer.

It should be emphasized that some of the results discussed in this paper can be extended to heat transfer rates of liquids in forced convection boiling inside tubes. However, such an extension can be applied only to certain flow patterns of the two-phase mixtures. It has been shown in recent publications [10, 1 1] that a considerable error can be made in evaluating heat transfer coefficients in boiling inside tubes if the effect of the flow pattern is neglected.

ACKNOWLEDGMENT

The work reported in this paper was supported by the Deutsche Forschungsgemein­schaft, Bad Godesberg. The author is indebted to Th. Dickmann, H. P. Klein and E. Hollack for their assistance in the experiments.

378

11-6

REFERENCES

1. I. Tschernobylski and G. Ratiani, Cholod. Techn. 32 (1955) 3-48/51. 2 . W. F. Witzig, G. W. Penney and ]. A. Cyphers, Refrig. Engng. 56 (1948) 153/157. 3. P. Worsae-Schmidt, Proc. 10 th. Inst. Congr. Refrig. (1959) 1-46. 4. K. Stephan, Beitrag zur Thermodynamik des Warmeiiberganges beim Sieden. Habilitations-

schrift Techn. Hochschule Karlsruhe (1963). 5. Fuchs, Technische Mitteilungen der Rudolf Fuchs Mineralolwerk Mannheim. 6. F. N. Kruschilin, Izv. Akad. Nauk. 555R. Otdel. Tekk. Nauk. (1949) 5-701/712 . 7 . G. Danilowa and ]. Masukewitsch, Cholod. Techn. 31 (1954) 62/65. 8. W. Lawrowa, Cholod. Techn. 34 (1957) 3-55/61 . 9 . E . Kirschbaum, Chem. Ing. Techn. 2 4 (1952) 393/400.

10. N. Zuber and E. Fried, Joum. Amer. Rocket Soc. 32 (1962) 1332. n. P. Sachs and R. A. K. Long, Int. Journ. Heat Mass Transfer 2 (1961) 222 .

DISCUSSION

H. Heckmatt, U.K. : Would the author give some indication of the method used in his assessment of vapour pressures ? Variation in such pressures, when different amounts of oil are added to Refrigerant 22, can be very small and can present difficulties in measurement especially at lower temperatures.

With regard to the heat transfer coefficients (HTC) of liquid Refrigerant 22, it is generally accepted that when oil is added the HTC is reduced. This is, to a great extent, due to the formation of an oil film on the boiling surfaces. However, should a miscible oil (as in the case of an aromatic base oil) be used, the oil film will not form and the HTC remains high. Has the author any comments on oils of different miscibility?

K. Stephan, Germany : Vapour pressures have been measured directly by means of a manometer. As to the effect of very small variations of the vapour pressure they are not interesting in the present paper because the saturation temperature has been measured and not been evaluated from the vapour pressure measurements.

It has been pointed out that HTC may not only be reduced by the addition of oil. In some cases HTC is even enlarged. This depends on the amount of oil added to the Refrigerant. The different sorts of oil used in the experiments were completely miscible within the temperature range of the experiments. In so far the effects observed are not due to the formation of an oil film. Experiments within the gap of miscibility have not been performed.

J. Ross-Jensen, Denmark : It seems that all work of this kind has only been concerned with R 12 and R 22.

Have you considered whether the tendencies established for these refrigerants also apply for other Freon refrigerants such as R 13, R 13 B 1 or even R 502 ?

K. Stephan, Germany : Experiments have been done only with the Refrigerants 12 and 22. But it is to be supposed that similar tendencies will apply also to other Freon Refri­gerants.

J. J. Kowalczewski, Australia : Could Dr. Stephan give some details of the experimental technique ? It is known that the heat transfer in pool boiling is strongly affected by the heat transfer surface.

What was the surface roughness and the material of the heat transfer surface ?

K. Stephan, Germany : The mean surface roughness was about 1 micrometer. This value corresponds to a very smooth surface. Further experiments on the influence of surface roughness have proved that the HTC may change by a factor of two or more by alteration of the surface roughness. The results on these experiments will soon be published in another paper.

The heat transfer surface was a copper plate.

J. B. Chaddock, U.S.A. : I should like to make just a short comment on the subject of Dr. Stephan's paper, and ask him for his opinion concerning this comment. Dr. Stephan's investigation of the effect of oil addition on boiling heat transfer was concerned with,

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what we call, "pool boiling" or "pot boiling". As he has shown, oil addition has a signific­ant effect on heat transfer rates for such boiling action. This is certainly applicable to heat transfer in flooded evaporators as a previous question here has brought forth. In the case of evaporation inside tubes, where the refrigerant is flowing with considerable velocity, however, the process of evaporation is considerably different. In such cases the energy transfer is not by the mechanism of ebullition (bubble action), but rather by a turbulent convective process similar to turbulent flow of a pure liquid or gas. (Provided there is more than a few per cent gas present, as will be the case in most refrigerant evaporators). Now, I should think that in such forced convection evaporation processes the presence of small amounts of oil would have very little effect on the heat transfer rate, as compared to the pool boiling experiments. Would the author care to express his opinion?

K. Stephan, Germany : Some results on boiling of oil-refrigerants mixtures inside tubes have been obtained by Worsoe-Schmidt [3]. As he pointed out, in boiling processes inside tubes the influence of the flow pattern on heat transfer may not be neglected. An extension of the results on pool boiling heat transfer to boiling inside tubes can be done only for certain flow patterns of the two-phase mixtures. Up to now it is not quite well known in what cases such an extension may be done. It seems to me that further experiments must be carried out on this problem.

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Separation of Oil from Refrigerant Vapor

Separation de l'huile de la vapeur de frigorigene

H. B. SAEMUNDSSON Kaltetechnisches Institut, Technische Hochschule, Karlsruhe, Germany

SOMMA/RE. Pour comparer l'efficacite de divers types de separateurs d'huile, on a effectue des essais sur une installation expbimentale speciale. Pour obtenir des conditions reproductibles, on injectait de l'huile atomisee dans un ecoulement constant de vapeur de frigorigene (R 11) et l'on mesurait la quantite d'huile retenue dans le separateur d'essai.

L' efficacite etait mesuree pour differentes quantites d'injection d'huile, pour trois types diffe­rents de separateurs d'huile, l'un contenant un filtre plat en treillis metallique, un separateur centrifuge avec un anneau de directrices et un plus petit separateur centrifuge courant avec une entree tangentielle.

Les resultats ont montre que l'efficacite variait de 82 a 99%. Chaque separateur d'huile avait une serie de niveaux optimaux d'ejficacite, pouvant correspondre a une sbie plus ou moins etendue d'ecoulements de vapeur defrigorigene et de quantites d'entrainement d'huile, dependant du type et de la charge.

1. AIM OF THE EXPERIMENTAL WORK The oil carryover in refrigerating plants is detrimental to the operation, especially in

refrigerating plants for low temperatures. The oil can be separated by several different methods : By sudden changes in direction of flow, caused by baffle plates or filters of wire mesh, fibrous media, packing beds or a sinter cake (baffle plate separators, filter separa­tors), and by centrifugal force (cyclones).

Even with the most effective and expensive separator a fractional part of the oil carried over still remains in the refrigerant. This is mainly due to the fact that separation becomes less effective the smaller the liquid particles are [1]. Particles smaller than 0,5 µ follow the streamlines and are not influenced by an inertial force [2], and the influence on larger droplets, up to 5 µ is very small [3, 4, 5]. Separation is impossible if the oil is present in the vapor state. This may occur if the discharge temperature of the compressor is very high [6].

Another cause of a bad efficiency is the fact, that an oil film can be drawn upwards a wall against the gravity force, even at vapor velocities as low as 2-3 m/sec [7] and leave the separator through the outlet. The amount of the oil carryover, the form and properties of the oil and refrigerant, the velocities and guidance of the flow in the separator are essen­tial for the separation efficiency.

It was the aim of the experiments carried out in the Institute of Refrigeration at Karls­ruhe to determine the efficiency of some types of conventional oil separators at repro­ducible conditions.

2. RESUME OF PREVIOUS WORK

The literature gives only a few results on the efficiency of separators for separating liquids from gases and vapors. These papers concern the separation of water-droplets from air [1]. For oil separation, especially from refrigerant vapor, no data about the efficiency of separators are available.

3. DESCRIPTION OF THE TEST PLANT A test plant shown schematically in Fig. 1 operating with the low pressure refrigerant

R 11 was erected. The form, distribution and amount of the oil carryover from a com­pressor can hardly be determined and moreover these properties vary with the operating conditions and design of the compressor. Therefore the compressor was replaced by an injection pump capable of delivering a defined amount of oil into a refrigerant vapor stream.

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Fig. 1 . Schematic diagram of test apparatus for oil separators.

I Evaporator 6 Rectifier 2 Superheater 7 Condenser 3 Injection Pump 8 Subcooler 4 Injection Nozzle 5 Oil Separator

9 Flow Meter 10 Rectifying Apparatus

The refrigerant vapor was generated in an electrically heated evaporator (1) and super­heated about 10° C in a super heater (2). A certain amount of oil was injected into the refri­gerant vapor stream through an injector nozzle (4) by means of a Diesel fuel injection pump (3). This nozzle is situated in a straight pipe just ahead of the separator (5) in order to prevent the droplets from forming a film on the wall of the pipe before entering the separator. The amount of oil could be varied by adjustment of the injection pump. The oil retained was collected at the bottom of the separator and was conducted together with the refrigerant dissolved in the oil to a rectifier (6). Here the oil was separated from the dissolved refrigerant and in this way the weight of the oil retained in the separator could be determined. The refrigerant vapor containing a small amount of oil was lique­fied in a condenser (7) and purified in a rectifying apparatus (10) consisting of a rectifier and a condenser. The condensate was subcooled in a subcooler (8) and its amount measu­red with a flow meter (9).

382

Oil Separators Nominal Size P::l 20 mm

Type A

w = 51 - 116 kg/h 1) = 96 - 99 %

Type B

w = 63 - 135 kg/h

'1) = 99 %

Fig. 2. Types of oil separators tested.

Type C

w = 63 - 128 kg/h

1) = 82 - 97,5 %

4. DESCRIPTION OF THE TESTED OIL SEPARATORS

The following types of oil separators shown in Fig. 2 were tested :

Type A : Separator containing a flat filter of wire mesh.

Type B : Cyclone with a ring of guide blades.

Type C : Cyclone with a tangential inlet.

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Type A : The separator consists of a vertical cylinder divided into two halves. Having passed the inlet in the lower half, the vapor strikes against the cylindrical vessel containing the retained oil. A part of the refrigerant dissolved in the oil is supposed to evaporate due to heat exchange between the entering superheated vapor and the oil vessel. Opposite the inlet the vapor flows to the upper half and through a flat filter of wire mesh 130 x 1 10 x 30 mm, which fills the whole cross-section of the upper half, and leaves through the outlet. In the filter the oil is retained on the surface of the wires, from which it trickles to the bottom of the oil vessel.

Type B : The entering refrigerant vapor strikes at first on the central outlet pipe. By means of a ring of guide blades mounted between the outlet pipe and the casing the vapor stream is set in rotary motion. The droplets are taken to the interior wall of the casing by the centrifugal force and the separated oil trickles as a film into the oil vessel. The rapid rotation causes a low pressure in the middle of the cyclone, which may lead to a creep of an oil film down the exterior wall of the pipe. In order to prevent this the outlet pipe is not prolonged to the top and moreover it bears a drip ring, on which the oil film forms droplets which are taken to the interior separator wall by the centrifugal force occurring in the rotating vapor. A small cone is arranged above the oil vessel in order to prevent the accumulated oil from being whirled up again.

Type C : This is a common type of cyclone often used to separate dust from gases. The refrigerant vapor enters the cyclone tangentially and is hereby set in rotary motion causing the droplets to form a film on the inside wall by means of the centrifugal force. The vapor, now free of oil drops, leaves through the central outlet pipe, which has no drip ring for preventing the escape of an oil film. The oil is prevented from whirling up again by a small centered cone.

5. TEST PROCEDURE

Several series of experiments were carried out. During one series of experiments the refrigerant vapor stream was held constant by adjusting to a certain power input, and the amount of oil injected was varied throughout the range of the injection pump capacity. In the other series of tests measurements were again made at seven different amounts of injected oil, but at different constant values of vapor flow. The weight of the injected oil and the weight of the separated oil which had been purified were measured. The amount of the refrigerant circulating in the plant was found by measuring the condensate with a flow meter. In order to avoid inaccuracy caused by bubbles in the liquid refrigerant, the condensate flowed through a subcooler before passing the flow meter. The pressures and thus the temperatures of the refrigerant vapor varied from one test to another between 1,6 and 2,7 kp/cm2 which corresponds to a saturation temperature of 36° C and 54°C. The oil was injected through the spraynozzle at a pressure of200kp/cm2• The temperature of the oil was held constant at 70° C in order to avoid a condensation of refrigerant on the oil drops. The size of the droplets produced by the injection nozzle depends to a large extent on the injection pressure. The actual size of the droplets could not be measured, but according to other measurements [8] on the size of oil droplets produced by atomizer nozzles, the most frequent droplet diameter decreases from 20 µ at 100 kp/cm2 to 8 µ at 300 kp/cm2 injection pressure. Comparison tests with injection pressures of 30 and 60 kp/ cm2 were run on the oil separator type A in order to study the influence of the size of the droplets on the separation efficiency, but no influence could be observed. Probably this influence in the present tests is within the limits of accuracy.

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6. RESULTS

From the measured values the separation efficiency ri has been calculated. This is the ratio of the weight of oil retained in the separator to the weight of oil injected. The separa­tion efficiency depends on the ratio of weight of oil injected to the weight of refrigerant vapor flow, which is called concentration of oil in refrigerant �.

The test-results of the three oil separators are presented in Figs. 3, 4 and 5. In these diagrams the separation efficiency ri is plotted against the concentration of oil in refrige­rant vapor �.

The efficiency curve of the jilter separator type A is shown in Fig. 3. In the range of low concentrations the separation efficiency increases with increasing concentration. A broad maximum of 99 % efficiency is reached at 4-7 % oil concentration. In the range of higher oil concentrations the efficiency decreases again. As the actual oil carryover in refrigeration plants seldom exceeds 10% L9] the higher concentrations are of little practical interest. For all amounts of refrigerant vapor flow measured no influence of the vapor flow on the efficiency could be found.

.

-;95f-'y�· ----+-"'

·� � sor-------t-- Flow Rollo 11 6 kglh ' ,. , F/ow Ra"o 100 kg /h

� o: Flow Ratio 77 kglh &asr--------t-- � Flow Rolio 51 /cglh

··.�����____,,���-'--,,�. �����,,� Oil Conc•nlr ation J ["'•)

Fig. 3. Characteristics of oil separator type A.

The test-results of the cyclone with a ring oj guide blades type B are scattered about a straight line (Fig. 4). The separation efficiency remains constant at an average of 99 % for all concentrations of oil in refrigerant and for all refrigerant vapor flows measured. The excellent efficiency is probably due to the design of this separator, to the guide blades, the outlet pipe leaving at the side and the dripring. Further it can be assumed that the range of flow measured in the present tests corresponds to the range of nominal load of this separator, and that the efficiency will decrease for refrigerant vapor flow below and above the nominal flow.

384

�•sr-----�-+------+-------i

.[ � sor-------+-'

� !ur------+--

e : Flow Rotlo 135 kglh �

• : Ffow Rallc 1oa ltglh • • Fl o w Ratio 66 kg/h " ' Flow Rollo 8J kglh

TO 15 Oil Conuntrallon f { %]

Fig. 4. Characteristics of oil separator type B.

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The results of the tests of the cyclone type C are shown in Fig. 5. The separation efficiency depends very clearly on the rate of the refrigerant vapor flow and on the con­centration. At increasing concentration of oil in the refrigerant the separation efficiency improves until a maximum is reached at about 10 % oil concentration. The influence of the rate of refrigerant vapor flow on the efficiency shows that this small separator is overloaded except at the lowest rate of flow, where the curve is rather flat, having a flat efficiency maximum of 97,5 % at 5-10% concentration of oil in refrigerant.

00-------- --- -

c l/"/1:1Ffaw Ratio 10,JiS kglh

� llf o • Flow Ratio 66,6 kglh

�as<--�----'- IV • · Flow Ratio 6112 kglh "' y •• 0 10

Oil Conctnlral1on J ["1·} 15

Fig. 5. Characteristics of oil separator type C.

7. CONCLUSIONS

As the tests show there are certain differences in the behavior of the various types of separators. Therefore it would help the designers of refrigeration units if they could be informed by the manufacturers about the characteristics of the separators under various operating conditions.

ACKNOWLEDGEMENT

The work reported in this paper was supported by a grant from the Landesgewerbeamt Baden-Wiirttemberg.

REFERENCES

r. W. Barth: Abscheidung von Fllissigkeitsnebeln aus Gasen. Alig. Warmetechnik 9 (r960) 252-256.

2. ]. Brink: New fiber mist eliminator. Chem. Engng., 66 (r959) r83-r86.

3. H. A. Leninger: Phase separations and classification. Cyclones in industry. Elsevier Pub!. Co. (r96r) r -22.

4. W. Barth: Grenzen und Moglichkeiten der mech. Entstaubung. Staub 23 (r963) r 76-r80. 5. K. R. Schmidt: Stand und apparative Grenzen der techn. Feinstaubabscheidung. Staub 23

(r963) r8r-r95. 6. Lebaux and Bartlett: Refrigerating oil carryover at high temperatures. Refrig. Engng. 53

(r947) 203-207.

7. 0. Glas: Oljemedslapning i kylanlaggningars vertikala sugledningar. Kylteknisk Tidskrift r9 (r960) ro-r6.

8. Robert Bosch GmbH.: Personal communication. 9. R. Plank: Handbuch der Kaltetechnik IV. Springer-Verlag r956, r5r .

DISCUSSION

S. Tauber, Nether/ands : Large droplets generally tend to collect on the inside wall of the pipe leading to the separator. It is possible that a large portion of the oil reaches the separator not in droplet (mist) form but as a stream of liquid oil flowing over the bottom of the pipe. Do the efficiency (YJ) figures refer to the droplets only or to the total mass of oil injected? In other words, do the figures ( 'Y/) refer to the pipe and oil separator or to the separator only ?

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H. B. Saemundsson, Germany : It is quite right that the oil reaching the oil separator in normal refrigeration units is present both in the form of mist or droplets and in the form of an oil film moving along the imide wall of the discharge pipe. In our test plant this is not so. The injection nozzle was used in order to have the oil in only one form, namely in the form of droplets of reproducible size. The nozzle had a jet angle of only 8° and was situated in a straight pipe about 250 mm ahead of each separator spraying the oil into the centre of the vapour flow. Therefore almost the total amount of oil injected reached the oil separator in the mist or droplet form.

Normally the efficiency figures refer to the total amount of oil reaching the oil separa­tor, as a film and as droplets. The efficiency figures of the present tests refer to the system consisting of the short pipe and the oil separator, and whereas almost no oil collects as a film on the inside wall of the pipe they are practically identical with the efficiency figures of the separators alone.

L. M. Hamaker, Netherlands :

1. In Fig. 2 the nominal size for type A oil-separator is given as ""' 20 mm. Is this the diameter of the inlet and outlet tube, or what other dimension does it refer to ?

2. Can the author give data on the pressure-drop for the different oil separators ?

H. B. Saemundson, Germany :

1. These three oil separators are the customary types as sold by manufacturers. The manufacturers define the size of oil separators according to the size of the piping. Thus nominal size ""' 20 mm means that the inlet and the outlet pipe of all the three separators have a diameter of about 20 mm.

2. The pressure drop was not measured.

D. Weissbarth, Germany :

1. What was the refrigerant ?

2. What were the velocities in the tubes and in the separator ?

3. What was the pressure drop ?

H. B. Saemundsson, Germany :

1. The refrigerant was R 1 1 (CFC13), as mentioned in the paper. 2. The velocities in the tubes can be calculated from the refrigerant flow and the

system pressure. For a 20 mm tube the velocities range from 3 to 12 m/sec. 3. The pressure drop was not measured.

0. Doubek, Germany : Were there any observations of the effect of oil temperature upon the efficiency of the oil separator and which oil temperature was used for the measure­ment ?

H. B. Saemundsson, Germany : The temperature of the injected oil was kept constant at 70° C and therefore no observations on the effect of oil temperature were made.

D. W. Higham, U.K. : Has the temperature of the incoming vapour any effect on the efficiency of the separators ?

H. B. Saemundsson, Germany : With increasing temperature the vapour pressure of the oil increases and the viscosity decreases [ 6]. Therefore it can be assumed that the temperature of the incoming vapour does have an effect :on the efficiency of oil separators. In the present tests the temperature range was rather narrow and far below the boiling range of the oil. Therefore no effect of the temperature could be found.

E.J. Perry, U.K. : Would the efficiency of oil separation of the type A separator be increased by passing the oil-gas-mixture vertically upwards through a horizontal demister pad; the oil falling to the base of the vessel and the gas leaving from the top.

H. B. Saemundsson, Germany : According to the observations of Professor Barth (Karlsruhe), a gas stream of a certain velocity flowing vertically upwards through a horizontal pad can take the liquid through the pad and at the upper surface of the pad

386

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tear it again to droplets, which again mix with the gas stream. In order to avoid this, the velocity has to be so low that the drag force does not exceed the opposite by directed gravity force.

In addition to the velocity, the amount of liquid is critical. A large amount of liquid can form an obstruction of the filter pad, which results in an increasing velocity through the non-obstructed area of the filter pad, again causing an increasing entrainment.

Passing the oil-gas-mixture horizontally through a vertical demister pad, the gravity force and the drag force have an angle of 90°, and the resulting force is directing the liquid downwards and out of the stream. Professor Barth found this arrangement to be the most favourable, e. g. for separating water droplets from air.

V. Chlumsky, Czechoslovakia : I am afraid the wire mesh in the separator type A is a relatively poor filtering material. I wonder if better filtering material, such as paper layers or sintered cake of a correct porosity, were also investigated ?

H. B. Saemundsson, Germany : The separation is not an effect of a real filter process at all. In a filtering process you hold back particles which are too large to pass the pores of the filter. When separating liquid particles from a gas in a wire mesh pad, the free sectional area between the wires is much larger than the droplet size. The wires induce the oil-gas-mixture to make rapid changes in direction, which cause the particles to strike upon the wires because of their inertia. There are also other factors involved, but this is the main one. The oil forms a film on the wires and has to be rapidly conducted out of the zone of vapour flow.

It is possible that separators with a sinter cake of the highest possible porosity could be used if the amount of liquid is very small. But the pressure drop will be very high.

N. H. McCall, New Zealand :

1. Were the effects of different superheated temperatures of the refrigerant taken into account with these experiments ?

2. Were any experiments carried out by injecting oil first into the refrigerant vapour and then superheating the vapour and oil mixture ?

H. B. Saemundsson, Germany :

1. The effects of different superheating temperatures were not taken into account. The superheating temperature was always kept about 10° C above the evaporator temperature.

2. Such experiments were not carried out.

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The Diffusional Penetration of Humidity in the Insulation of Pipes

Penetration par diffusion de l'humidite dans !'isolation dex tuyaux

Prof. Dr.-Ing. H. GLASER Institut ftir Technische Thermodynamik, Technische Hochschule, Stuttgart, Germany

SOMMAIRE. La vapeur d'eau diffusant dans !'isolation des tuyaux froids se condense habituellement a la surface du tuyau. Par suite, /'isolation s'humidifie par l'interieur. On presente !es equations s'y rapportant. Celles-ci s' appliquent aussi au cas ou /'isolation consiste en divers materiaux de proprietes diff erentes. Les equations soulignent le besoin d'un ecran d' etancheite efficace, si le flux d' entree de diffusion doit rester au-dessous d'une certaine limite.

On indique une methode particulierement simple pour la solution de ces problemes par une methode graphique. Ce/a permet une etude rapide de la situation physique ainsi qu'un ca/cul quantitatif de toutes /es donnees importantes. Si !es tuyaux sont tres froids, la vapeur se condense a l'interieur de l'isolant, provoquant ainsi une augmentation du flux de diffusion total. La repartition locale de l'humidite condensee, de meme que la quantite, estfacilement evaluee par la methode graphique.

Under steady state conditions a certain heat flux is fl.owing through the insulating material. This process of heat conduction is linked very often in the field of refrigerating engineering with a diffusion flux of water vapour. The rate of diffusion depends on the total vapour gradient and on the porosity of the wall. Methods for the computation of the diffusion rate through plane walls were presented in [1, 2, 3, 4, 5]. The results of these computations show, whether one has to expect under given conditions condensation within the wall and if yes, how much vapour actually condenses.

Water vapour is leaving the wall on the cold side, unless that side is not impermeable to the vapour. Without condensation the in- and outgoing fluxes are the same under steady conditions. Naturally, if water vapour does condense, less vapour is leaving the wall as is going in. However, if the material at the cold side of the wall is impervious to water vapour, then the total amount of incoming vapour must condense before this barrier. This is always so if cold pipes are insulated. The insulation is getting wet from the inside and as a consequence it becomes less effective because of an increase in the heat conductivity. It will be shown mathematically how large the diffusion rates through the insulating materials of pipes are.

The equations for the description of heat conduction and diffusion processes are of the same type. In cylindrical coordinates the equation of heat conduction runs

dq = 2 n A r ( iJ2t__ + ! ij t) dr iJr2 r iJr

with q = heat flux per length through a cylindrical shell r = radius t = temperature A = heat conductivity

(1)

With no heat sources, q is a constant. Then an integration of equ. (1) is easily perform­ed and leads to the relation (2)

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q t = t1 + 2 n ,1. ln(r/r1)

The subscript i refers to the internal radius ri of the pipe.

(2) From equ. (2) the temperature t is easily calculated if t1 at r = r; is known. There exists a logarithmic relation between t and r.

The diffusion process is described by the following equation

o ( 02p 1 op) dg =

µRn T 2 n r ¥ + -;: Tr dr

with g rate of water vapour diffusion p partial pressure Rn gas constant of water vapour T absolute temperature o diffusion coefficient of the system water vapour/air µ ratio of the diffusion coefficients for water vapour in still air

and within the wall [6].

(3)

Normally the heat insulating material is surrounded by a vapour barrier. The vapour pressure within the insulation is then usually lower than the saturation pressure. Only on the cold side of the insulation, that is practically at the surface of the pipe, the sat­uration point is reached and only here a condensation takes place. Therefore the dif­fusion rate remains constant within the insulation, that is dg = 0. Under these conditions an integration of equ. (3) leads to the following equ., which describes the partial pres­sure as a function of the radius r.

g µ Rn T P = P1 + 2 n 0

In (r/n)

with ri internal radius of the pipe Pi saturation vapour pressure at the temperature t = ti ti practically the surface temperature of the pipe

(4)

The vapour pressure p at the cold side of the insulation must be equal to the satura­tion pressure P1 given by t1 because there we have the following physical situation : A steady stream of water vapour is moving to the cold pipe surface. This would cause the vapour pressure rising. But, sii;ice the pressure p cannot exceed the saturation pres­sure p; at the temperature ti, the vapour arriving at the cold wall either will condense or form a layer of ice. Under the assumption that the insulation material is homogeneous, the diffusion flux entering the insulation is given by

g = µ RnT In (ra/r;) (5)

vapour barrier

Fig. r. Insulation of a cold pipe.

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with Pa = partial pressure of water vapour just outside of the insulation at a radius r = ra. Usually this pressure is equal to the partial pressure of vapour in the surrounding air.

In practical cases however the heat insulation is formed by several layers as shown in Fig. I. Then the rate of diffusion may be calculated from

g = 2 n (Pa - Pi) (µ1 T1 µ2 T2 µn Tn ) RD T ln(r1/n) + � (ln r2/r1) + .. . + ----;r,;-ln (ra/rn-1)

(6)

if the diffusion resistance factors µ1 are known. The subscripts refer to the different layers. As already was pointed out, the amount of water vapour condensing or forming

ice inside the insulation will be equal tog. This has no grave consequences, if the vapour barrier at the outside is very effective and is keeping the diffusion rate relatively small.

Equ. (6) enables us to compute, how good the vapour barrier has to be, so that a . l' . f . d d A dd' . 1

µs Ts In /( ) ' h certain upper 1m1t o g IS not excee e . n a 1t10na term � · rs rs - s m t e

denominator would take account of this. rs is the external radius, µs the diffusion re­sistance factor and s the thickness of the vapour barrier. From (6) the above mentioned term can be calculated, because under certain given conditions the other quantities entering (6) may be assumed to be known. Then an appropriate design of the vapour barrier is possible.

In some cases, particularly if the water vapour is condensing before it reaches the cold pipe surface, it is of interest to know the partial pressures inside the various layers of insulating material. Condensation beginning at the radius r' takes place in those regions where the partial pressure equals the saturation pressure. This effect has to be expected always in insulations of very cold pipes. As long as the temperature profile is not noticeably affected by the heat of condensation generated by formation of water or ice the problem may be solved using a graphical method.

For a very thin cylinder shell equation (5) can be written 2 n c5 dp

g = µ RD T d ln r With the definition

µ RD T --- ln r = D 2 n c5

(7)

(8)

that may be interpreted as a diffusion resistance related to the unit of length, then equation (7) becomes

dp g = dD (9)

Without condensation g is constant throughout the insulation, therefore p is a linear function of D.

A graphical presentation p = f (D) (Fig. 2) must show a straight line. The same holds if the pipe insulation consists of several layers of various materials. Then however one has to be careful to insert in equ. (8) the appropriate values ofµ, Tand c5 for each layer.

The simplest way for the evaluation of the partial water vapour pressure is this : First compute Di at r = r1 with the data referring to the inner layer. Then take

the saturation pressure Pi at the temperature ti out of a steam table and plot this against D1 in the p-D diagramm (Point J in Fig. 3).

As the second step the sum of all diffusion resistances

µ1RD T1 µ2RD T2 :ED =

2 n 151 In(r1/r;) + 2 n 152 In (r2/r1) +

µn RD Tn + In (ra/rn-1)

2 n c5n (10)

391

11-15

2000

v / l/v 1500

v v

[7 500 i

7.5 -7,0 -45 -qO -� -5,0 �� -4, 0 0 10-50

Fig. 2. The partial pressure of water vapour as a function of D.

has to be calculated. As a rule it is acceptable to replace in (10) Tp T2 ••• Tn by their average temperatures and to insert the diffusion coefficients 61, 62 • • • 6n in relation to these temperatures. Then add D1 + l: D, find the outside water vapour pressure Pa and plot Pa againstD1 + ED. The straight line connecting both points A and J, represents the distribution of vapour pressure within the insulating material. The vapour pressure at the boundaries of the layers then is easily taken out of the diagram. The slope of the line is identical with the diffusion rate, which so is being obtained very quickly (Fig. 3).

2000

1500

Q. � 1000 l i ' p

500

0

-v 7! '

' ' I '

I ! i r·-0i I -8 D -6

/v k-1--Po i I I 1

v ! v ! i I

I ' , 1 : : ' ' 1 :

0, I JrPJ ' ' - · -2 0

10-'o Fig. 3. The vapour pressure within an insulation consisting of three layers.

It is easily shown that the temperature also varies linearly with D inside each layer. However at the boundaries of the various materials the temperature gradient is exhib­iting a discontinuity because of the different heat conductivities. Once the temperature profile is known, the saturation pressures of the water vapour are easily plotted. This is done in Fig. 4.

If t1 is very low, it can happen that the straight line A-J crosses the curve representing the saturation pressures. From the point of intersection onwards the actual vapour pressure would be larger than p., which is impossible for the same reasons as it was

392

II-15

shown for the plane wall [1]. Therefore the line A-J (Fig. 4) has to approach the sat­uration line in a tangential way. From the point onwards, where both lines touch one another the actual pressure is identical with the saturation pressure. Also the condensa­tion process is starting at this point because the vapour gradient promoting the diffusion process is getting steadily smaller.

Fig. 4. The vapour pressure in the presence of condensation.

The amount of water vapour condensing is given by

µ Rn T d2p G =

2 n b r dD2 I 1 )

and therefore depends on the rate of change of the slope of the saturation line in the p-D diagram. The diffusion flux entering the insulation is equal to dp/dD. From Fig. 4 follows, that dp/dD is increasing if condensation of the diffusing vapour occurs at some point before the pipe surface. Then more moisture enters the insulation as compared to those cases where condensation takes place only at the surface of the pipe.

The graphical procedure just presented allows a quick orientation and in addition leads to quantitative evaluation of all important data.

REFERENCES

r. H. Glaser, Wii.rmeleitung und Feuchtigkeitsdurchgang <lurch Kiihlraumisolierungen. Kalte­technik lo, 86-91, 1958.

2. H. Glaser, Temperatur- und Dampfdruckverlauf in einer homogenen Wand bei Feuchtigkeits­ausscheidung. Kaltetechnik 10, 174-179, 1958.

3 . H. Glaser, Vereinfachte Berechnung der Dampfdiffusion <lurch geschichtete Wii.nde bei Aus­scheidung von Wasser und Eis. Kaltetechnik lo, 358-364, 386-390, 1958.

4. H. Glaser, Graphisches Verfahren zur Untersuchung von Diffusionsvorgangen. Kaltetechnik l r, 345-349, 1959·

5. H. Glaser, Graphic Method for Determination of the Vapour Pressure Within a Wall of a Cold Store and the Mass Rate of Water Vapour Diffusing Through Walls Consisting of Several Layers. Proc. X. Int. Congr. Refr. Copenhagen 1959· Vol. I, 387--394.

6. 0. Krischer, Trocknungstechnik, Vol. l, Springer-Verlag, Berlin/Giittingen/Heidelberg 1956.

DISCUSSION

F. L. Levy, U.K. : The same principles apply to the reverse case of passenger aircraft flying at high altitude. Passenger comfort requires an inside temperature of +20°C with 50 % relative humidity, while the outside temperature is generally of the order of -50° C. As there is no vapour seal at the inside, all water vapour condenses within the insulation of the airplane.

H. Glaser, Germany : Dr. Levy has mentioned an interesting example. The equations referred to in my paper can be applied without difficulty to study the conditions existing in an aircraft flying at high altitude.

393

11-1

On the Correlation of the Thermal Convection Coefficients

Sur la correlation des coefficients de convection thermique

Prof. CESARE CODEGONE Politecnico di Torino, Istituto di Fisica Tecnica, Torino, Italy

SOMMA/RE. La correlation des coefficients de convection thermique pour spheres, cylin­dres et surfaces planes peut etre obtenue dans tout le champ des experiences, et meme aux basses temperatures, au moyen d'une equation avec deux constantes numeriques seulement, et preci­sement :

pour la convection libre: N Nu = (N Gr · N Pr) · exp { a + b log (N Gr · N Pr) )

pour la convection forcee : NNu · (NPr) - '/, = NRe · exp (a' + b' log NRe)

FREE CONVECTION : The common equation :

(1)

must be used with many couples of values ofCand n, which are variables according to the considered interval ; moreover these intervals are of uncertain breadth. The expression :

NNu = (NGr · NPr) · exp (a + b log (NGr · NPr) )

allows to adopt a unique couple of numerical costants in the whole experimental field * . From (2) we have :

log NNu = {a + b log (NGr · NPr) ) · log (NGr · NPr)

d (log NNu) d [log (Nar . NPr)] = n1 = a + 2 b log (NGr · NPr)

which gives the inclination of the tangent to the line in logarithmic diagram. The value of n1 increases with log (N Gr · N Pr).

(2)

(3)

(4)

The experimental value of n in (1) is ""' 0 for log (N Gr · N Pr) < -5 and it increases to ""' 0,4 for log (N Gr · N Pr) ""' 12.

From the numerical correlations reported by M. Jakob (s. Heat Transfer, 1949, I, p. 524) and represented in the Fig. 1, lines I, the following fundamental values are adop­ted :

log (NGr • NPr) = -6 0 + 12 log NNu = -0,4 0 + 3,2

from which : b = 1/90 = 0,011 1 . . . . . . . . . a = 12 b = 0,133 . . . . . . . . . . n = [12 + log (N Gr · N Pr)] : 90

and therefore : n1 = [6 + log (NGr · NPr)] /45 (5)

NN u = (NGr · NPr) exp { [12 + log (NGr · NPr)] /90 ) (6)

*A exp B = AB. 395

II-I y

-5 10

Fig. r . Comparison between different correlations

I-Free convection : X = log (NGr - NPr) ; Y = log NNu l A Nusselt B McAdams (horiz. cylinde1s) C King (vert. surfaces, horiz. cylinders, spheres, blocks) D McAdams (low vertical surfaces)

II-Forced convection : X = log NRe ; Y = log [NNu • (NPrfY:i]

The equation (5) gives :

f! H i l pert l'O" Ulsamer

ll!I King <I> Reiher

for n1 0 1/10 2/10 1 /4 3/10 1 /3 4/10

log (Nar · Nrr) = -6 -1,5 +3 5,2 7,5 9 12

The comparison between the known correlations and Eq. (6) is given in the Table and in the Fig. 2.

C.[log Nw.J r----.---.--,--,-----.---'i"--.,,.,------,�.,.---,-.�---.--.�--,-----,

+ 0, 1 0

- o, 1 0

- 6 -3 + 3 + 6 + 9 + 12

A --0-- Nusselt [ horizontal cy l ind ers) B -<>- Mc Adams [ horizon tal cylind ers) c --0- W. J. King ( vertical sur�aces , horiz. cyl inder s , blocks, sphe res) D ----0---· Mc Adams [ low vertical sur face s )

Fig. 2 . Comparison between equation (6) and other correlations for free convection .

396

Table 1. (Values of log NN u)

W. J. King's correlation

Nusselt's McAdams for vertical log · correlation correlation surfaces,

(N arN Pr) for horizontal for horizontal horizontal cylinders cylinders cylinders,

blocks and spheres

-6 -5 -0,349 -4 -0,325 -0,31 -0,31 -3 -0,283 -0,26 -0,26 -2 -0,220 -0,18 -0,20 -1 -0,134 -0,075 -0,10

0 -0,025 +0,035 + 0,03 1 + 0,100 0,18 0,19 2 0,258 0,325 0,36 3 0,455 0,50 0,56 4 0,68 0,73 0,78 5 0,94 0,97 1,00 6 1,20 1,21 1,24 7 1,46 1,46 1,52 8 1,72 1,71 1,80 9 1,97 2,10

10 2,41 1 1 2,78 12 3,06

"_" _______________ "

• 1-hlped Ji! King -e Ulsamer

¢ Reiher lSl Nottage c Boelter 121 William�

11-1

McAdams correlation for Equation

low vertical (6) surfaces

-0,400 -0,388 -0,355 -0,300 -0,222 -0,100

+0,16 0,000 0,28 +0,144 0,42 0,311 0,59 0,500 0,78 0,71 1,02 0,94 1,27 1,20 1,52 1,47 1,77 1,77 2,02 2,10 2,33 2,44 2,70 2,81

3,20 - ------- --·---- ----

Fig. 3. Comparison between experimental values of n1 and equation (9) for forced convection.

397

11-1

FORCED CONVECTION :

Also for forced convection of gases and liquids external to cylinders, planes and spheres, the commonly adopted equation

(7)

must be used with many couples of values of C' and n1, which are variables according to the considered interval (m is always ""' 1/3).

The exponents n1, following M. Jakob (op. cit. p. 560, 561, 563), are indicated in the Table 2 and represented in the Fig. 3.

Table 2.

J. Ulsamer R. Hilpert J. Ulsamer J. Ulsamer R. Hilpert J. Ulsamer

Author

G. C. Williams (spheres) H. Reiher R. Hilpert W. J. King H. B. Nottage and L. M. Boelter R. Hilpert

log (NRe) (middle values of

the intervals)

-0,3 +o,3 +0,35 + 1,15

2,6 2,8 3,2 3,2 4,1 3,5 4 5,1

n

0,305 0,330 0,385 0,41 0,466 0,50 0,60 0,56 0,618 0,585 0,52 0,80

It must be noted that Hilpert's and Ulsamer's value of C" in the equation :

NN u = c• N�. (8)

for air (NPr = 0,71) and NRe ""' 1, corresponds to (0,71)Ys ; then NN u · N-t; = 1 for NRe = 1 .

The experimental values of the exponent n1 follow the approximate equation :

n1 = (IO + 2 log NRe) /30 (9) Therefore I NN u · (NPr)-Ya = NRe · exp ( [IO + log NRe] / 30} I (IO)

which is represented in the Fig. 1, curve II, and gives the following values :

log NRe n1 log [(N N u) • (N Pr)-Ys]

-1,25 1/4 -0,36 -1 0,26 -0,30

0 1/3 0,00 + 1 4/10 + 0,36

2 0,46 0,80 2,5 1/2 1,04 3 0,53 . . . 1,30 4 6/IO 1,86 5 2/3 2,50 6 0,73 . . . 3,20

398

BIBLIOGRAPHY :

M. Jakob, Heat Transfer, New York, I (1949), 524, 528, 560, 563. W. Nusselt, Ges. Ing. 38, (1915), 477, 490 ; Z.V.D.I., 73 (1929) 1475. W. H. McAdams, Heat Transmission, New York (1942), 243, 248, 237. W. ]. King, Mech. Eng. 54 (1932) 347, 410. R. Hilpert, V.D.I. Forsch. Heft, N. 355, 1932. ]. Ulsamer, Forsch. Geb. Ing. Wes. 3 (1932) 94. H. Reiher, V.D.I. Forsch. Heft, N. 269, 1925. H. B. Nottage, L. M. Boelter, Trans. Am. Soc. Heat. Vent. Eng. 46 (1940) 4r.

DISCUSSION

11-1

H. F. Th. Meffert, Netherlands : As Krischer has given a correlation between Re and G (Trocknungstechnik, Vol. I, 1956), I should like to ask Prof. Codegone whether he has also compared the proposed equations with the values given by Krischer for both the types, natural and forced convection ? It seems to me that the recommendation of two equations is a step backwards.

C. Codegone, Italy : Les formules donnees se referent aux correlations numeriques faites par les auteurs : Nusselt, Jakob, McAdams, King (voir bibliographie).

H. Hausen, Germany : For some years I myself have tried to represent the same curve by a simple equation of the form

Nu = 0,11 (N Gr • N Pr)1/3 + (N Gr • N Pr)O,lO

but I am not quite sure whether I know the coefficients exactly by heart. I think it would be very interesting to compare your quotation and mine with the experiments and with each other.

C. Codegone, Italy : Je ne connaissais pas cette equation qui a trois constantes. Je me propose de la confronter avec la mienne.

G. Walker, U.K. : Have you any information on recent experimental work on convection in gases at very high pressure in narrow spaces ?

C. Codegone, Italy : Les experiences (voir bibliographie) se referent a des champs tres vastes de temperatures et de pressions et a un grand nombre de fluides (gaz et liquides).

0. Lyng, Sweden : Have you tried to use your new type of equation at all for convection inside vertical, closed gas spaces ? I suppose this is not possible for spaces with high width-height-ratio, but may be for others.

C. Codegone, Italy : Mes experiences sur des plaques planes exposees a l'air en con­vection libre concordent avec l'equation (6) (Cf. C. Codegone, Ric. d'Ing., 3 (1935) n. 4).

La correlation ne se refere pas aux couches d'air.

399

11-21

Mesure en regime variable du coefficient d'echange thermique en surface

Measurement of the Coefficient of Surface Heat Exchange under Transient Con­ditions

ANDRE GAC, Docteur-Ingenieur Centre de Recherches et d'Experimentation de Genie Rural, Antony, France

SUMMARY. We studied 2 methods for the measurement of the coefficient of surface heat exchange under transient thermal conditions.

The first method is a direct method based on the determination of temperature gradient in the superficial layer of a material. This method, already adopted by various authors, presents the serious drawback of inaccuracy, if the temperature gradient is not measured very carefully. Results of measurements made under these conditions are presented.

The second method, an indirect one, is based on the determination of the cooling of a def­inite geometric solid. Measurement is possible with solid plastics. We mention the results of measurements made on spheres with 20 to 30 mm diameter, of polyvinyl chloride, Nylon and Teflon. These results agree, to within 10%, with those provided by calculation, from the measurement of air speed.

We mention the disadvantages and the causes of inaccuracy of the method, as well as the advantages of the process.

Le coefficient d'echange thermique en surface depend d'une part de la nature, de la vitesse, de la temperature du fluide et d'autre part des dimensions, de la forme, de l'etat de surface, de la region de l'objet qui echange de la chaleur avec le fluide. La definition du coefficient d'echange thermique en surface est done assez delicate. Or, en pratique, il est necessaire de disposer d'une valeur moyenne h utilisable dans les calculs.

La mesure de coefficient h est difficile, en raison du nombre des facteurs qui inter­viennent. Cette mesure, effectuee ordinairement en regime thermique stable, peut etre egalement faite en regime thermique variable.

Dans cette note, nous presentons deux methodes de determination de h en regime varia­ble, la premiere directe et la seconde indirecte.

1 - SYMBOLES :

1,1 - Relatifs a l'objet : l : distance minimum du centre a la surface (m) x : distance d'un point au centre (m) Ac : coefficient de conductivite thermique (kcal/h, m, 0 C)

Ac a : diffusivite thermique (m2/h) a = -!?c

Bo : temperature initiale (0 C) f9 : temperature au point x et au temps t (° C) T : temps de demi-refroidissement e : masse volumique (kg/m3) c : chaleur massique (kcal/kg, ° C) � ) parametres

1,2 - Relatifs au fluide : D : vitesse massique (kg/h, m2) V : vitesse lineaire (m/sec)

401

II-21

h : coefficient moyen d'echange thermique en surface (kcal/h, m2, 0 C) Am : coefficient de conductivite thermique (kcal/h, m, 0 C) µ : coefficient de viscosite absolue (kg/h, m) Cm : chaleur massique (kcal/kg, 0 C) E>m : temperature (°C)

2 - METHODE DIRECTE DE MESURE EN REGIME VARIABLE FONDEE

SUR LA MESURE DU GRADIENT DE TEMPERATURE EN SURFACE :

2,1 - Principe : Une methode, deja utilisee par divers auteurs tels que Eddie et Pearson, consiste a

LI E> mesurer le gradient de temperature Tx existant dans la couche superficielle d'un obj et

solide, au cours de son refroidissement. En effet, i1 est possible d'ecrire :

# I Ac LIE> I h e - E>m LfX x = l (1)

Cette methode est simple et seduisante ; elle devrait permettre la mesure directe de h, valable pour un objet determine.

Malheureusement, les erreurs peuvent etre importantes, notamment parce qu'il faut mesurer, avec une tres grande precision, le gradient de temperature d'une couche mince. En outre, si l'objet est peu conducteur de la chaleur, I' experience ne foumit qu'une valeur locale qui peut etre assez differente de la valeur moyenne.

2,2 - Resultats : A la suite des mesures faites par J. P. Tupin sur des carcasses de bovins, nous avons

determine le coefficient hnpar cette methode ainsi que par le calcul,hv a partir de la vitesse de l'air V (en m/sec), existant a proximite immediate des carcasses et de la relation :

hv = 5 + 3,5 V (2)

Cette relation est recommandee quand l'objet est de grandes dimensions et quand la vitesse n'excede pas 4 a 5 m/sec (Gottsche et Pohlmann, Nisolle).

Les resultats obtenus (Fig. 1) presentent une tres grande dispersion, specialement quand la vitesse de !'air est superieure a 3 m/ sec .

15

10 i

. . / 0-/ 0 fa: ' ·0 / · . jtf �/ 0

·�· · :0 A ·�·

---t------------+-------- -s 1 0 1 5

Fig. 1 , Comparaison du coefficient hn, determine a !'aide de la relation 1 et du coefficient hv deter· mine a partir de la vitesse de !'air V et de la relation 2 (La temperature de !'air est de o°C).

402

11-21

3 - METHODE INDIRECTE DE MESURE EN REGIME VARIABLE, FONDEE SUR LE REFROIDISSEMENT D'UN SOLIDE GEOMETRIQUE :

3,1 - Principe : Si un objet de forme geometrique dffinie, constitue d'une substance homogene et

isotrope, est place dans un flux a temperature E>m et vitesse massique D constantes, la temperature E> en un point distant de x du centre est exprimee par les relations suivantes :

(m x) -mn2t E> = E>m + (E>o - E>m) l:An • F la e avec

pour t = o, e = E>o, quel que soit x

pour t quelconque, h (E> - Bm) = (- Ac ��) x = l

F' mnl et i/a

F

(�) (�1) h l

= - � A

(3)

J Un certain temps apres le debut du refroidissement, le terme fondamental de !'expres­

sion de E> devient predominant. Le refroidissement est alors exponentiel, caracterise par le temps de demi-refroidissement T, tel que :

L 2 T = - (4) m12

Si la fonction F qui depend de la forme de l'objet, l, Ac, a et T sont connus, les rela­tions 3 et 4 permettent de calculer h.

En particulier, pour une sphere, la relation est :

( m1 l )

h = Ac 1 _ i/a

l m1 l ig i/a

(5)

Le coefficient h ainsi determine ne peut, evidemment, erre adopte pour le calcul des echanges de chaleur entre un fluide et un objet quelconque. Par contre, ii permet de calculer la vitesse massique du fluide D ainsi que la vitesse lineaire V.

En effet, ii est connu que le coefficient h satisfait egalement la relation suivante :

Nu = cp (Pr, Re) oil

2 h l Nu = ;:;-- est le nombre de Nusselt

Pr Cm µ -;:;;;-- est le nombre de Prandtl

2 I D Re = --- est le nombre de Reynolds µ

La fonction cp depend de la forme du corps. En particulier, Mac-Adams recommande, pour une sphere refroidie en convection forcee dans l'air, la relation :

Nu = 0,37 (Re)6•6 pour 17 < Re < 70000

OU

h Am = 0,37 2/ (Re)0•6

403

Il-21

Pr n'apparait pas, car, pour l'air, ce nombre varie peu.

3,2 - Dispositif experimental : Le principe de la methode impose que : le coefficient h ne soit pas trop grand (en pratique la methode n'est valable que pour les gaz) l'objet soit petit i1 soit en matiere homogene et isotrope le temps de demi-refroidissement soit de 2 a 3 minutes la temperature initiale ne differe de celle du milieu que de 20 ° C pour reduire les echanges par rayonnement. Les metaux ne satisfont pas ces conditions. Par contre, les matieres plastiques permettent

de faire les mesures dans de bonnes conditions. Nous avons utilise des spheres en chlorure de polyvinyl, nylon et teflon.

Le dispositif (Fig. 2) comprend une sphere de 2 a 3 cm de diametre, dans laquelle est introduite, a force, une soudure de couple thermoelectrique, dont les fils, de 5/10 mm sont soigneusement gaines pour reduire les fuites thermiques. Le seconde soudre est place dans le milieu a proximite de la sphere. La force electromotrice est enregistree avec un potentiometre enregistreur precis 1).

e-nr•9islreur­d• lempil"olur•

•oudur• dan.s. to • hire

couple thermo. ilectri ue

Fig. 2. Schema de principe du dispositif de mesure de la vitesse massique de !'air, en regime thermique variable.

L'operation consiste a rechauffer la sphere a (<9m + 20) environ, a la placer dans l'enceinte et a enregistrer !'evolution de sa temperature par rapport au milieu. La courbe est transcrite sur papier semi-logarithmique et le temps de demi-refroidissement est mesure. On peut determiner h, D et V, a partir des relations 5 et 7 ou mieux d'un abaque (Fig. 3 relatif a !'air).

3,3 - Precision de la methode : Les causes d'imprecisions sont les suivantes :

3,31 - Variation des caracteristiques thermiques : Dans !'interpretation des resultats, on admet que les caracteristiques thermiques de la

substance restent constantes. Or, en particulier Ac diminue avec la temperature. Cette variation est cependant negligeable quand le refroidissement n'excede pas 20° C. En effet, les experimentaux reportes sur papier semi-logarithmique sont tres bien alignes.

3,32 - Position de la soudure dans la sphere : La valeur de h depend de la situation de la soudure thermoelectrique. L'imprecision est

cependant negligeable : Dans une sphere de 3 cm de diametre, placee dans un courant d'air de 2,6 m/sec a 20° C, le temps de demi-refroidissement moyen mesure sur trois

1) Nous avons utilise un potentiometre electronique enregistreur de 25 cm d'echelle. Sa sensibilite est de 12 µ V pour r mv de deviation totale. Le rythme de commutation est de 64 secondes.

404

Il-21

periodes est : 188, 189, 187, 186, 189 et 186 sec. selon que la prise de temperature est situee respectivement a I' avant, a l'arriere, en haut, en bas, a droite et a gauche de la sphere par rapport au courant. Les ecarts sont ± 1,5 sec. pour une valeur moyenne de 187,5 sec.

3,33 - Constance de Gm : La relation 3 suppose que Gm est constant. Cependant, la methode reste valable si Gm

varie periodiquement autour d'une valeur moyenne constante.

3,34 - Mesure de la temperature et trace graphique : La transcription des releves sur papier semi-logarithmique et le trace de la courbe

constituent la plus grande cause d'imprecision et doivent etre faits avec grand soin. L' erreur relative sur h est de ± 5 % quand la mesure est faite sur 3 T dans un courant d'air de moins de 2,5 mfsec avec l'appareillage de mesure indique ci-dessus (§ 3,2).

3,35 - Influence de la valeur de h : L'erreur est d'autant plus grande que h est plus grand. La methode ne peut pas etre

utilisee quand i1 s'agit d'un liquide. Pour l'air, la vitesse ne doit pas exceder 3 a 4 m/sec (Fig. 3 : si V passe de 1 a 2 m/sec, T diminue de 50 sec, si V passe de 4 a 5 mf sec, T diminue de 5 a 6 secondes).

" � . "e : .

E � E e u Jf }! � � 8 :"> .� •O ,o > > > Q

12 : 60

11 j 12

12 10

11 50 11

10

10

<O

7 00

20

• • 2 2

1 1 1 T•mp• T

150 200 250

Fig. 3. Abaque donnant la vitesse massique de !'air D sec et la vitesse lineaire V en fonction du temps de 1/2 refroidissement T d'une sphere en Afcodur de 28 mm de diametre. (C.P.V.).

3,4 - Resultats experimentaux : Nous avons determine le coefficient d'echange thermique en surface selon la metho<le

ci-dessus h T ainsi qu'en mesurant la vitesse massique D et en appliquant la relation 7 :

he. Les resultats obtenus (tableau 1) montrent que les ecarts relatifs�� sont compris

entre 1,1 et 0,89 quand la vitesse de l'air varie de 1,1 a 5,7 mfsec.

4 - CONCLUSION

La methode directe fondee sur la mesure du gradient de temperature en surface ne foumit que des resultats grossiers, a moins que la determination du gradient de tempera­ture soit tres precise,

405

Il-21

Tableau 1 Mesure du cofficient d'echange thermique superficiel

hT: resultats foumis par la methode proposee de mesure en regime variable

he : resultats foumis par la relation 7' relative a la mesure en regime stable

Sphere diametre en

cm

Chlorure de polyvinyl 2,8 Am : 0,126 kcal/h, m, ° C (} : r .400 kg/m3 c : 0,24 kcal/kg, °C

Teflon 1,9 Am : 0,1656 kcal/h, m, °C (} : 2.150 kg/m3 c : 0,25 kcal/kg, °C

Nylon Am : 0,18 kcal/h, m, °C (} : r .045 kg/m3 c : 0,58 kcal/kg, °C

3,0

Air vitesse tern-

perature rn/sec oc

l,25 15 2,05 20 2,7 20 3.45 20

5,7 20 l ,l 30 1,2 30

2,7 20 3.45 20 5,7 20

2,6 20 2,7 20 2,7 20 3,45 20 5,7 20

r) Ces mesures ant ete effectuees avec !'aide de C. Andre

hT1) he

kcal/h, kcal;h, m2, °C m2, °C

31,5 30,8 49,0 44,5 46 48,5 { 57 56,2 62 75 75,5 25 28 31,4 29,4

52,5 57 61 65,6 86,5 87,5

44 46 47 47 42 47 51,6 54,6 71 7 1

--- ---

hT h e

l ,03 1,10 0,95 { l,or 1,10 0,995 0,89 l,03

0,925 0,93 0,99

0,96 1 ,00 0,89 o,95 1,oo

La mesure indirecte en regime variable, fondee sur le refroidissement d'un solidc geometrique, presente les avantages suivants :

le processus est simple et rapide la mesure peut etre faite a distance hors de la presence d'un operateur le sens de circulation de l'air n'intervient pas sur le resultat, du moins de fa9on notable l'objet etant de petites dimensions, la vitesse de !'air dans une region tres limitee peut etre mesuree. Par contre, les inconvenients sont les suivants : la methode ne peut etre utilisee que pour un flux gazeux dont la vitesse lineaire n' excede pas 4 m/sec. l'emploi d'un enregistreur tres precis de temperature est indispensable la qualite de la mesure, faite en regime variable, est moins bonne que celle de la mesure en regime stable.

5 - BIBLIOGARPHIE :

G. C. Eddie - S. F. Pearson: The freezing time of fish in airblast freezer, ]our. of Refrig. - juillet - aout 1958 -

A. Gae - ]. P. Tupin: Contribution ll. l'etude du refroidissement de cuisses de bovin - Reunion de l'l.l.F. - Washington 1962 - 8 p.

G. Gottsche - W. Pohlmann: Formulaire du frigoriste - Dunod - Paris 1946 - 531 p. W. H. Mac Adams: Transmission de chaleur - Dunod - Paris 1961 - 585 p. L. Nisolle: Cours de froid industriel - Centre de Doc. Uni. - Paris 1952 - 272 p. 406

11-21

DISCUSSION

L. Mattarolo, Italy : 1. Quelle est la position du thermocouple dans la sphere ? 11 faut preciser si la position

du thermocouple est importante ou non. 2. Peut-il evaluer !'influence de la radiation surtout aux tres basses vitesses de l'air?

A. Gae, France : La situation de la ronde thermomerrique dans la sphere n'a pas d'in­fluence sur le resultat de mesure puis que la methode proposee est fondee sur la deter­mination du temps de demirefroidissement en regime thermique exponential.

D'autre part des experiences ont ete effectues pour connaitre !'influence des echanges par rayonnement. Si l'ecart initial de temperature entre la sphere et le milieu n'excede pas 30°C, et pour les substances utilises, il n'a pas ete observe d'influence sensible due au rayonnement.

407

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Sur la diffusivite thermique des materiaux non homogenes

Thermal Diffusivity of Non Homogeneous Materials

Prof. L. MATTAROLO et M. SOVRANO Universite de Padoue, lnstitut de Physique Technique et Centre d'Etudes pour les Applications du Froid du C. N. R., Padoue, Italie

SUMMARY. The value of thermal diffusivity a of materials is shown by the relationship between thermal conductivity A and the product of specific heat c and of specific weight y, this being determined by a stationary method.

The question is whether the value a intervening in Fourier's equation for processes in transient state exactly corresponds to the above-mentioned definition in the case of non homogeneous materials, such as insulating materials and powders.

In this paper, the results of numerous tests on the direct determination of thermal diffusivity of different materials with transient state methods are given, and the values obtained in this way are compared with the values of the relationship between A and c y.

La diffusivite thermique qui regit la propagation d'une variation de temperature dans les solides est definie comme le rapport entre la conductivite thermique A et la chaleur specifique de !'unite de volume c y.

On peut determiner la valeur de la diffusivite directement, avec une methode ex­perimentale en regime variable lorsque on connait analitiquement, par solution de !'equation de Fourier, comment une variation imposee sur un point (ou sur une surface) est transmise a un autre point (ou a un autre surface) ; on peut encore calculer la valeur de la diffusivite thermique ayant determine experimentalement la valeur de la conducti­vite A (par une methode en regime stationaire), de la chaleur specifique et du poids specifique.

La question qui se pose c'est d'examiner si les valeurs de la diffusivite thermique obtenues par ces deux methodes differentes, (nous dirons la valeur dynamique et la valeur statique) sont coincidentes ou non. La question a raison d'etre pour Jes materiaux non homogenes [1, 2].

Nous avons execute bien de mesures avec differents materiaux non homogenes, et avec differentes methodes, pour la determination de la diffusivite thermique en regime variable et nous avons ensuite compare la valeur ainsi obtenue avec la valeur calculee par la conaissance de A de c et de y.

METHODES D'ESSAI Une premiere serie d'essais [3] avait ete execute sur des plaques de cm 50 x 50

ayant une epaisseur 1 de 10 cm au moins. On imposait sur une des surface (x = o) une variation de temperature ts = k T fonction lineaire du temps et on mesurait les varia­tions de temperature a distance x de 1, 2, 3 . . . cm de cette surface. En utilisant !'ex­pression :

t = 4 k T i2 e r f c x 2 V a T

qui represente la solution de !'equation de Fourier dans ce cas, lorsque ii est negligeable !'influence des conditions sur la surface x = l, on arrivait a la connaissance de la diffu­sivite thermique a.

Les materiaux qui ont ete soumis a ces premiers essais sont des isolants tres legers du type expanse a cellules fermees. On indique pour example (Table 1) les valeurs ainsi obtenus (c'est une moyenne de differentes conditions d'essais) pour le polystyrol.

409

11-23

Table 1 materiel

Produit ii base de Polystyrol expanse Densite 25 kg/m3

vitesse

k = d ts (O C/h) d T

75 ° C/h 112,5 ° C/h

diffusivite thermique dynamique (m2/h)

4,78 5,09

10-3 10-3

La conclusion de cette serie d'essais c'etait que la valeur dynamique obtenue de la diffusivite thermique est toujours plus elevee que la valeur statique qu'on obtient de la connaissance directe de A, de c et de y des materiaux memes.

La difficulte de realiser une variation lineaire absolument rigoreuse de la temperature superficielle, une analyse de l'erreur relative maximum du resultat du aux conditions

experimentaux et le desir d'examiner si la diffusivite est fonction de la vitesse �ts, nous a ammene ii une seconde serie d'essais.

•

On prenait une plaque du materiel ii examiner ayant dimensions 50 X 50 cm avec une epaisseur de 5 ii 10 cm et on imposait sur les deux surfaces de 50 X 50 une variation de temperature rigoureusement sinusoidale. On obtenait i;a faisant circuler en deux plaques metalliques avec espace vide serrees contre le deux surface susdites, de l'eau provenant de deux reservoirs thermostates l'un ii une temperature entre 40 et 50° C et l'autre ii une temperature entre 10 et 15°C.

Une vanne dont l'obturateur etait doue d'un mouvement periodique servait ii me­langer l'eau des deux reservoirs et ii l'envoyer aux deux plaques metalliques.

Des thermocouples mesuraient la temperature sur les deux surfaces x = o et x = s du materiel en essai. D'autres thermocouples mesuraient la temperature ii l'interieur

s du material, sur le plan mediant x = 2 de l'epaisseur.

Les temperatures etaient enregistrees par un potentiometre enregistreur du type Speedomax.

La comparaison entre les deux sinusoi:des des temperatures superficielles et celle a s

x = Z nous permettait d'obtenir la valeur de la diffusivite thermique (dynamique).

En effet si on a, pour x = o et pour x = s: t = A cos w •

. s on obt1ent sur le plan x = 2 :

t = 2 C e - {Js/2 cos (w T - fJ s/2 + a')

OU fJ = 1 I -Jr,� ' etant T 0 le periode, c est une grandeur qui devient egal a A lorsque V To a fJ s --+ 00 et a' est un angle qui meme devient egal a zero lorsque fJ s --+ 00 •

La determination de C et de a' peut etre obtenue par une methode de vecteurs rotatifs qui a ete coni;ue il y a quelques annees par un des auteurs de cet article [ 4] .

2 C e - fJ s/2

Dans la Fig. 1 nous avons trace le rapport B/A = --A� - - -- entre !'amplitude

s de la sinusoi:de sur le plan x = Z et I' amplitude des sinusoi:des superficielles, en fonction

de fJ s.

Dans la Fig. 2 il y a le dephasage a = fJ s/2 - rx' des deux sinusoi:des meme enfonction de fJ s. Les deux figures nous permettent d'arriver de la registration des temperatures (dont les Figs. 3 et 4 portent un exemple) a la valeur de fJ, done de la diffusivite thermique a. Les periodes des variations sinusoi:dales de temperatures etaient environ de 12, 24, 48 minutes.

410

Fig. r. Rapport des amplitudes des sinusoides B/ A en fonction de p s.

«' 140

120 ---!--------!

oL___t _ ___J_ _ _i__J_____JL____L _ _J__J__L--� 0

F,g 2

Fig. 2. Dephasage ex des sinusoides en fonction de p s.

iJC 1 1

11-23

Fig. 3. Registration de la temperature superficielle et sur le plan median avec periode To = r 2 minutes

Les materiaux sur lesquels nous avons essaye sont le polystyrol, le chlorure de poly­vinyl expansees et la sable (du type siliceux avec grains de 1 - 2 mm de diametre).

41 1

11-23

Fig. 4. Registration de la temperature superficielle et sur le plan median avec periocle To 24 minutes.

RESULT ATS

Les resultats sont indiques dans la Table 2 oil !es valeurs de la diffusivite dynamique representant la moyenne de tres nombreuses essais (dont l'ecart relatif de la moyenne etait d'ailleurs bien petit).

Table 2

Materiel periode cliffusi vi te (minute) thermique

dynamique (m'/h)

Produit a 12 6 . ro-3 base de Po

lystyrol 24 6,5 . I0-3 Chlorure de 12 3,55

polyvinyl 24 3,62 48 3,65

snble 24 10,90 . IO·' 48 10,30 . ro-•

;. c Cal Cal

m h ° C kg ° C

0,032 0,32

0,029 0,32

0,33 0,22

y diffusivite kg thermique m• statique

(rn','h)

20 5 . L0-8

1620 9,3 . I0-4 Le valeurs de J., c, y qui servent a calculer la diffusivite statique ont ere soigneusement

choisies presque toujours par determination experimentale directe.

On trouve que la diffusivite thermique dynamique est plus grande que la diffusivite thermique statique et que la constance de cette difference est, du point de vue qualitatif, inattaquable.

Aucune conclusion precise ne peut etre formulee jusqu'a ce moment pour ce qui concerne la variation de la diffusivite en fonction de la periode de la sinuso!de.

Nous devons souligner que l'idee de la difference entre les valeurs statiques et dyna­miques de la diffusivite thermique avait ere indique par des precedentes essais experi­mentales [1, 2), qui toutefois avaient ete conduits avec une des methodes completement differentes.

L'origine du phenomene est complexe : d'une part peut etre que dans la transmission en regime variable on obtient une transmission de la variation de la temperature plus vite de ce qu'on dirait la capacite thermique du materiel : tout se passe comme si 11:1

412

Il-23

chaleur trouve des voies de passage preferentielles : et "a peut etre bien imagine dans le materiel a grains : la variation de temperature avance dans le sens du flux thermique lorsque la variation n'a pas completement rejoint l'interieur de ce grains. D'autre part i1 faut rappeler que la conductivite thermique de ces materiaux non homogenes est

d t fonction autre que de la temperature t, meme du gradient d x' etant qu'on doit parler

dans ces materiaux d'une conductivite equivalente qui est le resultat de la conductivite vraie, de phenomenes de convection et de phenomenes de radiation.

Nous pensons qu'il y a bien de recherches a continuer dans ce domaine soit du point de vue theorique soit du point de vue experimentale.

BIBLIOGRAPHIE

I. A. Sellerio, M. C. Spitale, "Sul metodo ad impulsi di corrente nelle ricerche di termocinetica. Studio comparativo". La Termotecnica, n. 4.• vol. XIII, 1959·

2. M. C. Spitale, "Saggi di misure termocinetiche con sorgenti istantanee cilindriche o piane" Comptes Rendus du XV eme Congres Italien de Thermotechnique - Naples 1960.

3. M. Sovrano, "Determinazione della diffusivita termica dei materiali isolanti". Comptes Rendus du XVIII eme Congres Italien de Thermotechnique - Milan 1962.

4 . L. Mattarolo, "La trasmissione del calore nelle grandi dighe". !!Energia Elettrica, n. 5, vol. XXXI, 1954·

DISCUSSION

M. Guennoc, France : Vous avez signale que la conductivite thermique des solides peut dependre de la temperature. Pouvez-vous preciser quelle etait !'amplitude du signal thermique impose sur les deux faces de l'echantillon de mesure ?

L. Mattarolo, Italy : L'amplitude de la sinusoi:de etait de 25° C OU peu pres.

413

11-16

Determination of the Time Required for Contact Freezing of Whale­meat

Determination du temps necessaire a la congelation par contact de la viande de baleine

KAZUO TANAKA, Assistant Professor Tokyo University of Fisheries, Shiba Kaigandori 6, Minato-ku, Tokyo, Japan

JUN-ICHI NISHIMOTO, Assistant Professor Faculty of Fisheries, Kagoshima University, 470, Shimoarata-cho, Kagoshima­City, Japan

SOMMA/RE. Pour determiner le temps de congelation de la viande de baleine qui irradie sa chaleur par sa surface superieure et sa surface inferieure dans les congelateurs d plaques, on a etabli des formules d'apres la methode de R. PLANK.

OU q

r

C1 C2 f Oa, Ob Or r Oe d', d"

J.', J." rx', oc"

( D d' I ) ( D l: d" I ) � 2I; + .r: I- + � z;:-; + F + �;; 2 .1.2 (D d, d,, I I )2

- + .r:- + .r:- + - + --A 2 J.1 J.,, rx, rx,,

d" 2 .1.2 2 .1.2) q + 2 .1.2 .I: , ,, + -, + -,, -0--0 A rx rx r- e

c1 ( Oa - Ob) - ( c1 - c2) y ( Or - Ob) + f y

Or 1 - ---ob

temps de congelation de la viande par le congelateur d plaques, en h quantite de chaleur pour congeler /'unite de poids de viande, en kcal/kg epaisseur, en m

= poids specijique, en kg/m3, et conductibilite thermique, en kcal/m h ° C, de la viande, d /' etat de congele chaleur specijique de la viande d 1' etat non congele et d 1' etat congele, en kcal/kg °C chaleur latente de congelation, en kcal/kg temperature initiale et finale, respectivement, en °C

= point de congelation, en °C = pourcentage de viande congelee

temperature de /'agent refrigerant OU de /a plaque froide, en °C epaisseur d'une couche de mauvaise conductibilite thermique existant entre /es surf aces de la viande et les deux faces des plaques, en m coefficient de conductibilite thermique de cette couche, en kcal/mh °C coefficient de transmission de chaleur de 1' agent refrigerant ou de la face superieure et de la face inferieure des plaques froides, en kcal/m2h°C.

INTRODUCTION

In Japan, yearly about 100000 tons of frozen whalemeat are produced in the Ant­arctic and North Pacific Ocean by Japanese whaling expeditions. A type of freezer used on board a refrigerator mother-ship working there is almost all contact freezer (so­called pressure freezer or multi-plate freezer). Therefore, the authors induced a formula determining the time required for contact freezing of whalemeat, and carried out cal­culations to obtain some proper values.

415

11-16

INDUCTION OF BASIC FORMULA

1. Freezing of Whalemeat. The whales caught in the Antarctic or North Pacific waters are cut to a rectangular

shape of meat and put into a metal tray, then inserted between cold metal plates of contact freezer and frozen.

2. Induction of Basic Formula to Determine the Time Required for Contact Freezing of Whalemeat.

R. Plank applied the theory of heat transmission for determination of the freezing­time of foods cylindrical in shape. According to the same method of R. Plank, and re­garding the whalemeat as a plate, the authors induced a formula determining the time required for contact freezing of whalemeat.

In referring to Fig. 1, we let a1, b1, and a2, b2 = total breadth and total width of the upper and the lower surfaces of the meat as a plate respectively [m] ; the areas Ai and A2 [m2] are both rectangles with different sizes, and a1 � a2, b1 � b2, a1 b1 = A1, a2 b2 = A2,A1 � A2, h = total height of the meat. In this case, the meat is assumed to be a thin tra-

Fi3. I Trapezoidal body

pezoidal body in shape. Then, the lower surface A2 is insulated and the meat radiates its heat through the upper surface A1 only having thickness di', d2', d3', • • • , d' [m] in general of some substances with poor thermal conductivity il.1', il.2', il.3', • • • , ii.' [kcal/ m h 0 CJ in general.

c1, c2: /: E>a, E>b : E>r: r: E> e :

q : Q: t :

specific weight o f the meat [kg/m3] and thermal conductivity of the meat [kcal/m h 0 CJ in frozen state respectively specific heat of the meat in unfrozen and frozen state respectively [kcal/kg° CJ latent heat of freezing of the meat [kcal/kg] initial and final temperature of the meat respectively [0 CJ freezing point of the meat [0 CJ freezing percentage of the meat temperature of the cooling medium or the cold metal plate of the contact freezer [0 CJ heat transmission coefficient between the cooling medium or the cold metal plate and the upper surface of the meat [kcal/m2h 0 CJ freezing-heat-quantity of unit weight of the meat [kcal /kg] freezing-heat-quantity of the meat [kcal] freezing-time of the meat by contact freezer [h].

If we put a0, b0 [m] for breadth and width of the upper surface of the unfrozen part of the meat with rectangle of area A0 [m2], a0 :5 b0, a0 b0 = A0, and hi, h0, h2 [m] for height from the intersection of the extended side lines to the upper, above mentioned

416

Il-16

unfrozen surface, the intermediate and the lower surfaces A1, A0 and A 2 respectively, thickness of frozen part of the meat o0 [m] is given by

00 = h0 - h1 • •• h0 = h1 + 00 Also, as these upper, intermediate and lower surfaces A1, A0 and A2 are similar,

there are following relations between them.

a0 b0 h0 ai

= b; = hi ' and here, as h = h2 - h1

a1 bi hi = h -- = h --a2-a1 b2- b1

On the one hand, the freezing-heat-quantity dQ [kcal] taken away from the thin meat layer of which thickness is do0 [m] is given by

dQ = (c1 (Ba - Bb) - (c1 - Cz) r (Br - Bb) + fr } Y2 ao bo doo On the other hand, this dQ [kcal] is expressed by heat transmission from the inner

part of the meat to the cooling medium or the cold metal plate through the upper sur­face of the meat during dt [h] as follows.

(Br - B2) dt dQ = --c------,.----------,----c�-�

h1 00 + _

1_ (E '!_ + _!_) A.2 a1 b1 h1 + 00 a1 b1 A.' rx '

If the right terms of these equations are equal and integration is carried out under consideration of 00 = h0 - h1 = h2 - h1 = h in limit case, the freezing-time of the meat as a trapezoidal plate of which the lower surface is insulated and radiates its heat through the upper surface only can be obtained as follows.

or

( d' 2 A.2 E )( +

- � { bi + 2 b2 bi2 + �1-b2 + b22 ( '!_ -

6 A.2 h bi

h + bi2 2 A.2 E A.' +

2 A.2) } q + 7 Br - B e

Then, we choose 11 and /2 [m] for representative lengths of the upper and the lower surfaces A1 and A2 [m2] of the meat respectively and /1 S 12, 112 = AI> 122 = A2, above equation is transformed into

417

II-16

or = � h { V.A;A2 + 2 A2 h + �i + YA1 A!_ + _A2 (z .ic2 :E <!'_ +

6 A2 VA1 A2 A1 a'

2 Jc2) } q + --;;: 191 - l9e

Putting h = D, the equation is generally expressed as

here q = c1 (19a - 19b) - (c1 - c2) r (19r - 19b) + fr 19 r r = l- -eb

and D = total thickness of the meat [m] . When the difference between A1 and A2 is not so large, since we can carry out some

abbreviation in this equation, it is simplified as follows.

Putting

here

1

V = µ , the equation is generally expressed as A1 + A1 A2 + A2

t = _Y_2- D (n + 2 Jc :E d'

+ �) -- -· q

-2 µ A2 2 A' a' 19r - 19c

q = c1 (19a - 19b) - (c1 - c2) r (19r - Bb) + fr . Eh r = l - -eb

q

When A1 and A2 are equal, as µ = 1, this equation is more simplified as follows.

t = - - D D + 2 Jc :E - + -Y2 ( d' 2 A2) q 2 A2 2 A' a' Br - Be

here q = c1 (19a - 19b) - (c1 - c2) r (19r - 19b) + fr . 19r r = l - -eb

In connection therewith, if we put W as weight of the meat, the freezing-heat-quan­tity Q [kcal] taken away from the meat is given in all cases by

Q = Wq = W { c1 (19a - Bb) - (c1 - c2) r (Br - Eh) + fr ) . 19r r = 1 - --eb

here

Next, a formula determining the time required for contact freezing of whalemeat as a plate of which upper and lower surfaces are both rectangles with equal size, which radiates its heat through these two opposite surfaces is induced by using above final equation. In this case, the meat is assumed to be a thin rectangular body in shape.

In referring to Fig. 2, we let A', A" and A be surface area of the upper, the lower and the final frozen plane of the meat respectively [m2], and A' = A" = A, D, D' = total thickness and thickness from the upper surface to the final frozen plane of the meat

418

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respectively [m], therefore D-D' = thickness from the lower surface to the final frozen plane of the meat [m], d/, d2', d3', .. . d' in general and d1", d2", d3", • • • d" in general =

thickness of some substances with poor thermal conductivity existing between the both surfaces of the meat and the upper and the lower cold metal plates respectively [m], J.1', J.2', J.3', ••• ).' in general and ). t", J.2". J.3", • • • )." in general = thermal conductivity of above substances respectively [kcal/m h 0 C], rx', rx" = heat transmission coefficient between the cooling medium or the upper and the lower cold metal plates and the both surfaces of the meat respectively [kcal/m2 h 0C].

IX' -k if . .,...�,,," j o'

- - - - - - - - - -- - - -t I fm•l f�ozen plane ,\, D

j D-//

Fig. 2. Contact freezing of whalement of which both surfaces have some substances with poor thermal conductivicty,

On the one hand, freezing-time of the meat from the upper surface to the final frozen plane t [h] is given by

Y2 I ( I d' 2 A2) q t = 2--, D D + 2 J.2 E ,, + -, ;::;---:c; A2 A a ryr - t:::l' e

On the other hand, freezing-time of the meat from the lower surface to the final frozen plane t [h] is given by

y ( ){( ) d" 2 ,1, } q t = z l2

D - D' D - D' + 2 A2 E ,1," + rx"2

<9r - <9e Let the right terms of these equations be equal, and the thickness from the upper

surface to the final frozen plane of the meat can be obtained as follows.

D' =

D d" 1

2 A-2 + ;r )." + 7;," D d' d" 1 1

D

A2 + ;r "f + ;r ,1," + rx' + rx" Then, if we put this D' into the above two equations, we can obtain the same formula

determining the time of contact freezing of the meat in this case through each equation as follows.

t = Y2 2 ).2

4 1 9

II-16

here q = Ci (0a - f)b) - (ci - Cz) r (0t - f)b) + fr •

E>r r = l - -

eb

Values of Some Physical Constants of Whalemeat The values of some physical constants of whalemeat (lean meat, its moisture content :

72.7%, fat content : 3.0%, solid matter content : 24.3 %, here solid matter is a sum of protein 23.3 % and ash 1.0%) are shown in Table 1.

Table 1 . The values of some physical constants of whalemeat

Kind

-- -

lean meat

Specific Weight

unfrozen frozen

Yi Y2 kg/m3 kg/m3

---- - -

1071 1012

Freezing Point

E>r ° C

-1.03

Specific Heat

Latent Heat

of Freezing

unfrozen frozen Ci Cz f

kcal/kg° C kcal/kg° C kcal/kg

0.83 0.46 58.16

Freezing Percentage

r = l -- 1.03

<ih

Thermal conductivity

unfrozen frozen

Ai Az kcal/mh°C kcal/mh°C

I 0.71 I 1 .27

Comparison of Experimental Results and Calculated Values of Freezing-Time of Whalemeat by Contact Freezer The experimental freezing-time through the freezing-curves of lean meat of fin­

whales was taken in the Antarctic Ocean on the one hand, the calculated freezing-time of whalemeat was obtained through calculation by using above formules (especially final formula) and the values of some physical constants of whalemeat on the other hand. A type of contact freezer used there is an indirect type one, which has metal plates with cold CaC12 - brine circulating within them and oil hydraulic mechanism to up and down the plates.

1 . In case an air space has existed between the upper surface of the meat put into a metal tray and the cold metal plate (Data from the Nihon Fisheries Co. Expedition).

Three raw meats having a rectangular shape of thickness x breadth x width and weight : 0.057 x 0.329 x 0.583 m, 1 1 .6 kg, 0.043 x 0.329 x 0.583 m, 8.9 kg, and 0.038 x 0.329 x 0.583 m, 7.9 kg respectively were put into a metal tray having the same rectangular shape of height x breadth x width : 0.054 x 0.329 x 0.583 m and frozen by the contact freezer. Therefore, in this case an air space of which thermal conductivity is very poor existed between the upper surface of the meat and the cold metal plate. The thicknesses of the meat and the air spaces are : D = 0.054 m and d' = 0.000 m (it is to be D = 0.057 m and d' = -0.003 m, but practically the whale­meat is pressed and its thickness becomes D = 0.054 m and d' = 0.000 m simultaneously on freezing, this is the most popular case in the Antarctic Ocean), D = 0.043 m and d' = 0.011 m, and D = 0.038 m and d' = 0.016 m in above order, and thermal con• ductivity of an air space is approximately A' = 0.02 kcal/m h ° C. The initial temperature

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and the freezing time of these three meats and average temperature of a cold Ca Cl2-brine were the same in all : Ba = 10.0°C, t = 4.5 h, and Be = -26.0°C. But the final temperature (freezing percentage) of these three meats were different in all: Bb = -25.4°C (r = 0.9594) for the thickness of the meat D = 0.054 m and the air space d' = 0.000 m, Bb = -13.2°C (r = 0.9220) for D = 0.043 m and d' = 0.011 m, and Bb = -17.8° C (r = 0.9421) for D = 0.038 m and d' = 0.016 m.

Therefore, following relations can be put to the final equation previously mentioned

d" d' d' d" - 0 • · . E )." = O, and E }( = �-, and equations for D' [m] and t [h] are simplified either as follows.

D'=

D 1 2 A.2 + rx" D d' 1 1 -,-- + ,, + -;.;;- + A2 A '""' oc"

D ( D d' 1 )( D 1 '

Y2 2J:; + I' + (;! 2J:; + (;!') t = 2 A.2 (D d' 1 1 )2

A.2 + 1' + rx' + rx"

2 A.2 2 A.2) q + 7 + 7 Br - Be

( d' D D + 2 A.2 ).' +

Calculations were carried out using above two formulas with rx' = 6 kcal/ m2 h ° C. and rx" = 40 kcal/m2 h °C. for an indirect type contact freezer, and the freezing-times under the thicknesses from the upper surface to the final frozen plane of these meats were obtained at the following values : t = 2.98 h under D' = 0.027 m, t = 4.63 h under D' = 0.002 m, and t = 4.24 h under D' = 0.001 m.

2. In case the meat was put into a wooden box (Data from the Taiyo Fishing Co. Expedition).

The raw meat having a rectangular shape of thickness x breadth x width and weight : 0.140 x 0.348 x 0.606 m, 31.6 kg was put into a wooden box having the same rectangular shape of height x breadth x width : 0.136 x 0.348 x 0.606 m and frozen by the contact freezer.

Therefore, in this case such substance with poor thermal conductivity as a wooden layer existed between the lower surface of the meat and the cold metal plate. The thicknesses of the meat and the wooden plate are D = 0.140 m and d" = 0.010 m respectively, and thermal conductivity of a wooden plate is approximately )." = 0.12 kcal/m h ° C.

The initial temperature and the freezing-time of the meat and average temperature of a cold CaC12-brine were : Ba = 24.0°C, t = 1 1 .0 h, and B e = 27.0°C. The final temperature at each part of the meat was : Bb = -15.0°C at the part in 0.02 m from the upper surface of the meat, Bb = -13.0° C at 0.05 m, Bb = -6.0° C at 0.08 m, Bb = -2.0°C !at ;0.1 1 m, and Bb = -6.0°C at 0.14 m. It is assumed that the part 0.08 m from the upper surface of the meat is the final frozen plane. Therefore, following relations can also be put to the final equation previously mentioned

d' d" d" d' = 0 . · . E }( = 0 , and E )." )." and equations for D' [m] and t [h] are also simplified either as follows. D d" 1

i ).- + �,, + ;,;; D' 2 D = D d" 1 l A.2 + )." + a! + rx"

421

11-16

+ �oc�2 + 2oc�2) @;- q e�

( d" D D + 2 A.2 A." +

Calculations were carried out using the above two formulas with oc' = 40 kcal/m2h°C. and oc" = 10 kcal/m2h °C. for an indirect type contact freezer, and the freezing-time under the thickness from the upper surface to the final frozen plane of this meat (D' = 0.08 m) was obtained at the following value : t = 13.40 h under D' = 0.105 m.

In the above two examples, where these calculated freezing times are compared with those experimental ones, it can be easily recognized that these values agree with each other.

CONCLUSION

It is clear that the formulas and values of some physical constants used here to deter­mine the time required for contact freezing of whalemeat are quite effective.

422

IV. Commission 3

Design, construction and operation of machinery

for refrigerating and air conditioning plants.

Calcul, construction et exploitation du materiel

frigorifique et de conditionnement d'air.

SESSIONS :

Piston and Turbo

Compressors

Miscellaneous Questions

Thermoelectric

Refrigeration, Absorption

Refrigerants, Automation

Compresseurs a piston,

turbo-compresseurs

Questions diverses

Refroidissement thermo­

electrique, absorption

Fluides frigorigenes, automation

P i ston and Turbo C o m p ressors Compresseu rs a p i ston, t u rbo­comp resseu rs

III-15

Direct Measuring of the Middle Indicated Pressure pmi of Compres­sors by Electronic Methods 1

Mesure electronique directe de la pression moyenne indiques Pmt des com­presseurs

PROF. DR.-ING. V. FUNER AND R. TAUCHMANN Staatliche lngenieurschule, Karlsruhe, Germany

SOMMA/RE. L'evaluation des cartes d'indicateurs par planimetrie, pour obtenir la puissance indiquee, exige - si l'on utilise la methode electronique - une grande quantite de travaux photographiques et par consequent beaucoup de temps. En particulier, lorsqu' on fait une serie d'experiences ou des experiences comparees, il est utile de connaitre immediatement /es valeurs mesurees de la puissance indiquee. Les compteurs de travail mecaniques sont connus depuis longtemps mais ils sont mal adaptes

aux petits compresseurs a grande vitesse. Les «pi-metres» donnent presque toujours les valeurs moyennes de Pm dans le temps et non la pression moyenne indiquee Pm1.Pour une machine particuliere f onctionnant a vitesse constante, il n' est pas difficile de resoudre le probleme avec une relation empirique entre Pm et Pmi. On a applique la methode a doigt e/ectronique et la methode de mesure par ana/ogie, mais celles-ci sont couteuses et exigent un materiel complique. Dans cette etude, on a mis au point a peu de frais une methode de mesure electronique appli­cable pour tous /es types de dispositifs de lecture pour compresseurs, n' exigeant que l' etalonnage du dispositif de lecture. Les valeurs obtenues de cette maniere sont comparees a celles obtenues avec les methodes

classiques. La methode de mesure elle-meme se fonde sur l' application d'un multiplicateur electronique - dans ce cas particulier des transformateurs thermiques - et d'un membre de differentiel. Avec un renversement des elements de mesure suivi d'un etalonnage, on peut comparer !'influence de la vitesse, l' amplitude du piston et les variations des amplificateurs.

2

TT

5 6

3

Fig. I. Principal way of action of a directly working- or power-meter.

r . pressure-pick-up, 2. amplifier for the pres­

sure-signal, 3. piston-stroke- or veloc­

ity-pick-up, 4. a differentiator (ampli­

fier), 5. multiplier, 6. recording instrument.

r This research work was made possible by the financial support of the Forschungsrat Kaltetechnik, Germany.

425

III-15

The troublesome determination either of the middle indicated pressure Pmi or of the power Ni out of the indicator cards by planimetering - especially when series of expe­riments must be carried out - makes it necessary to develop a direct measuring method.

The different kinds of mechanical work-counters as for instance of W. Ashton and J. H. Storey (1869), Gumbel and Bottcher and Lehmann [1] can be considered as the first forerunners in this direction. An electrical, practically carried out assembly by A. Frisch (1919) uses for obtaining multiplying operations a watt-hour meter [1] . In recent times a method of H. H. Emschermann [2] allows the direct digital measuring of the indicated work of piston machines. From H. H. Emschermann and Ch. Rohrbach [3], too, dates the method system which essentially represents the measuring of the indicated power consumption. Fig. 1 shows the principal operation of such measuring methods. The machine to be investigated has a pressure-and a piston-stroke or a velocity pick-up. The signals coming out of these pick-ups are multiplied in a multiplier and supplied to a registering instrument which according to its construction can indicate either the work- or the power-consumption of the machine in case.

When using the method for determining the work or power it is necessary to calibrate one channel for the measuring of the pressure and another one for the way or velocity. The calibration of the way or of the velocity can be avoided if the middle indicated pres­sure Pm; will be measured to which the work respectively the power is proportional. In the latter case an auxiliary measuring is to be carried out shortly before or after the chief measuring instead of the calibrating the channel for the way or velocity.

As generally known, the middle indicated pressure can be expressed by the following equation :

T

� p dv JP iid1Jl dt Pmi = 'f = 0 dt (1)

Therein mean : Vh the swept volume and T duration of the period of the process. In measuring arrangement according to the scheme in Fig. 2 the periodical pressure-

2

s ,

426

Fig. 2. Block-scheme for measuring of Pmi· r. pressure-pick-up, 2. amplifier, 3. multiplier, 4. piston-stroke-pick-up, 5. amplifier, 6. differentiator, 7. moving-coil-instrument, 8. amount-former.

III--15

variations are forwarded by the pressure-pick-up (1) to the amplifier (2). Ki represents the transmission-factor of the pick-up as well as of the amplifier. From the amplifier (2) the signal reaches across the switch Si the entrance of the amplifier (3). The signal given by the piston-stroke-pick-up ( 4) is conducted to the second entrance of the multiplier (3), through the amplifier (5), the differentiator (6) and the switch S2• The stepping out signal of the multiplier, which represents the product of the entrance-values is connect­ed by the switches Sa and S4 with the moving-coil-instrument (7). The indication of the instrument shows this course of the signal :

T

A = K ·K ·K ·K . 1- JP ld?J_I dt i i 2 a 4 T dt

0 '.2)

Ki, K2, Ka and K4 are the transmission factors of the individual building elements.

By putting the switches S into the positions 2 the pressure-signal is cut out from the multiplier and instead of it a constant auxiliary voltage E is laid to it. From the stepping out voltage of the multiplier the absolute value is formed in the amount-former (8) and conducted to the instrument (7). The indication of the instrument is then:

T

A = K ·K ·K ·E· l f ldv \ dt 2 2 a 4 T dt

0

Introducing new limits for the integral there will be :

vh A2 = 2 K2 Ka K4 E· -� J dv = 2 K2 Ka K4 E·

0

1 After eliminating of T out of the equation ( 4) and (2) there is :

T

Ai � Az · Ki a .Ip l��l dt --- - - - - - -- -

(3)

(4)

(5)

The second factor of the equation (5) represents the sought middle indicated pressure Pml· It is :

Ai 2 · E Pm1 = A2 Ki (6)

As E is constant, there must be - in comparison with the other until now known methods - only determined the factor Ki by calibration.

An essential improvement of the measuring device for the determination of Pm1 shows Fig. 3. It enables a simple calibration of the pick-up (1) and of the preamplifier (2) by a calibration-device which generates pressure-variations. The calibration-device must be built up in such a way that the positive and the negative amplitudes have an equal value and that the shape of the curve is continual. The pressure-course in the calibration­device generally corresponds to the equation p = Po + (p2 - Pi) · f (t) whereby f (t) is a periodical function of the time. The part of the constant pressure Po is eliminated by using an alternating current amplifier or by a capacitive coupling. The pressures Pi and P2 represent the extreme values of the pressure-course in the calibration-device.

The function of both the switch positions 1 and 2 was already demonstrated with the help of Fig. 2. In the switch position 3 the signal coming from the pick-up (1) which is built in the calibration-device is conducted through the preamplifier (2) and the

427

IIl-15

7

Fig. 3. Improved arrangement for measuring of Pmi· r . pressure-pick-up, 2. amplifier, 3. multiplier, 4. piston-stroke-pick-up, 5. amplifier,

6. differentiator, 7. moving-coil-instrument, 8. amount-former.

differentiator (6) to the first entrance of the multiplier (3). Besides, the pressure-signal leaving the amplifier (2) reaches the amplifier at the second entrance. After the forma­tion of the amount in the amount-formator (8) the indication of the instrument is :

T

A = K2 ·K ·K . _!__ JP ldpl dt = K2 ·K ·K . _!__ • (P2 -P1)2 a 1 a 4 T dt 1 a 4 T 2 0

(7)

In the switch position 4 the constant voltage E is led to the second entrance of the mul­tiplier (3). After formation of the amount in (8) the indication of the instrument (7) is :

T

A4 = Ki · Ka ·K4 ·E· � 11��1 dt = 2·K1 ·Ka ·K4·E � (P2 - P1)

0

(8)

From the equations (7) and (8) we get K1 as follows :

A3 4 E K1 = - . -- - -A4 CP2 -P1)

(9)

With the measured values A1 and A2 gained with the switch-positions 1 and 2 there is :

A1 A4 P2 -P1 Pm1 = - · - · (10)

A2 A8 2

As the equation (10) shows, the final result does not more include the transmission­factors.

A device according to the block-scheme in Fig. 2 was built. As multiplier served at first thermo-transformers which unfortunately did not prove because of their large and different heath-inertia. At present a self-built diode-multiplier is used which works according to the parabola method system.

When comparing the values of Pm1 received with the above described arrangement with the Pm1-values gained by planimetering of pressure-volume-indicatorcards, there were deviations between 3 and 12 %. For these heavy faults - as shown by measurements - are evidently the variations of the auxiliary quantity E responsible, which are caused by the way of the circuit-breaking. Probably they can be avoided in the future according to the experience of the authors by using a watt meter as multiplier. Corresponding tests are planned. Fig. 4 shows the general view of the test stand.

The above described measuring method has the advantage that all kinds of pick-ups can be used, a calibration is simply to be carried out and last not least that this connec­tion can be easily produced with the analog computers.

428

IIl-15

Fig. 4. General view of the test stand for measuring of Pmi· r. electronic valve amplifier, 2. cathode ray oscillograph, 3. carrier frequency amplifier, 4. Pm1-measuring-instrument, 5. piston-stroke-pick-up, 6. compressor, 7. cable of the

piezo-quarz-pressure-pick-up, 8. electro-motor, 9. stroboscope for measuring the revolutions.

REFERENCES

r. K. ]. De Juhasz and ]. Geiger, Der Indikator, seine Theorie und seine mechanischen, optischen und elektrischen Ausfiihrungsarten. Julius Springer, Berlin, I938.

2 . H. H.Emschermann,Elektronischer indikator mit unmittelbarer Zahlenwertanzeige fiir Kolben­maschinen. VDI-Zeitschrift, Bd. IOI (I5), 589-594.

3. H. H. Emschermann und Ch. Rohrbach, Direkte elektrische Integration geschlossener MeB­kurven und ihre Anwendung zum Messen der Dampfung von Metallen und zum lndizieren von Kolbenmaschinen. VDI-Zeitschrift, Bd. 103 (5), 169-1 76.

DISCUSSION

V. Filner, Germany : The author himself opened the discussion by stating that elec­tronic indication had just been investigated in Japan and a paper on this matter had been written. He had not been able, however, to secure a copy and asked the Japanese dele­gates present to see what they could do in order to obtain a copy for him.

0. Skjeggedal, Norway : I should like to know a little more about how the calibration of the measuring chain was carried out during measurements and if the connecting lines to the pick-up are of considerable length. Would these then cause unavoidable gas oscillations and hence influence the measured P mt ?

V. Funer, Germany : The calibration was executed by the piston pump device and an oscilloscope. To the second part of the question I should like to mention that the pressure pick-up was built in so tight to the cylinder bore that no vibrations due to gas oscillations were detected whilst indicating.

429

m-2

Some Aspects of Pressure Pulse Attenuation for High Speed Reci­procating Compressors

Quelques aspects de l'amortissement de la pulsation de pression dans Jes com­presseurs alternatifs a grande vitesse

ESTHER KALETZKY, B. E., M. Eng. Sc., A. M. I. E. Aust. Senior Research Officer, Commonwealth Scientific and Industrial Research Organization, Melbourne, Australia

Description de !'attenuation des oscillations de pression des compresseurs alternatifs a grande vitesse a l'aide d'interference d'ondes coherentes (pratiquement sans pertes de charge).

SOMMA/RE. Les experiences efjectuees dans des conditions de fonctionnement dijjerentes avec le compresseur Jrigorijique surde l'air, du R 12 et du R 22 a des pressions de refoulement atteignant 19 kp/cm2 ( 1850000 N/m2} ont montre que !'amplitude des pulsations de gaz pourrait etre aissee de 5 a 15 f ois.

On a utilise des transducteurs de puissance avec un materiel electronique pour enregistrer la pression. Des experiences avec des appareils d' analyse du son f aisaient egalement ressortir un amortissement de 15 a 20 db pour des sons a bassefrequence a 1,30 m de distance du re­joulement libre d'un compresseur d'air.

Avec les compresseurs jrigorijiques a grande vitesse jonctionnant au jrigorigene, la simpli­cite et les faibles dimensions des amortisseurs d'interjerences sont particulierement seluisantes.

Pulsating flow of refrigerant, which can be defined as flow with periodic pressure fluctuations, occurs most severly in discharge pipes fed by reciprocating compressors. Oscillations in systems utilizing reciprocating compressors may cause undesirable mecha­nical vibrations of piping and equipment leading to large dynamic stresses, danger of resonance in long lines and excessive noise [1 ] . In some instances these problems become so serious that a refrigeration plant has to be shut down if remedial action against pul­sation is not taken in time. Measuring errors of instruments for flow and pressure measure­ments may also increase substantially because of the harmful effect of oscillations.

On the other hand, effective attenuation of refrigerant compressor pulses, while eminently desirable, should be obtained with minimum power and pressure losses in the pipe line. In order to investigate some practical possibilities of pulse reduction, the relative importance of various components of the pulse must be determined. The character of pressure oscillations in reciprocating machines mainly depends on the speed, number of cylinders, crank angle shift and the design of discharge or intake manifolds [2] .

The work described has comprised experimental investigations of inertial gas pulsa­tions originated by a reciprocating refrigerant compressor and also studies of the effec­tiveness of a coherent wave interference method for their suppression. The approach is based on the principle of uniform redistribution of oscillation energy in two divided flows and rejoining them in a common flow so that the shift of their phase is equal to an angle corresponding to an odd number of half wave lengths. As such wave compensation is practically free from pressure losses it appears to have some prospects of application to discharge and intake lines of refrigerant compressors with linear spectra of pulse fre­quencies.

Experimental investigations of pressure oscillations were carried out on twin cylinder reciprocating compressors for Freon 12 and Freon 22 having a nominal speed of 1500 rpm, which are extensively used in commercial refrigeration installations.

Pressure-time diagrams obtained by means of capacity pressure transducers [3], electronic preamplifier and oscilloscope showed that the pressure pulse obtained from

431

III-2

a twin reciprocating compressor could be regarded as fairly symmetrical and is mainly built up of a predominant fundamental frequency and a second harmonic. Actually, such a definitely linear character of the compressors pulse spectrum provided the most im­portant clue to further studies of the possibility of using the destructive wave interference method for pulse attenuation.

8Y-PASS l B'f· PASS TI

L.S. RECEIVER

AMPLIFIER OSCILLOSCOPE AUDIO CQJ-DATOR

H . S RECEIVER:

EVAPORATOR RE.STRICTOR

Fig. r. Schematic diagram of the experimental installation

On the schematic view of the experimental setup (Fig. 1) two double flow bypass systems are connected in series, in the first of which the bypass length differs from the main by a half wave length of the fundamental frequency and in the second bypass the difference in length is equal to a half wave length of the second harmonic. For creating equal pulse amplitudes in the branches similar conditions as regard transmission and reflection coefficients of the pressure wave were provided. The total interference effect of attenuation was measured after the junction of bypass I I with the main line.

INIT IAL PULSE AT 1 47 0 R.P. M.

ATTENUATED PULSE AT IS45 R.P.M

H-f4-..t j ... H-{=t1 H*ai@M ATTENUAT ED PULSE AT 14 10 R.P.M.

l--H-#4=H I I I ZERO LINE OF I N STRUMENTAT I O N

Fig. 2. Pulse attenuation on a Freon r 2 twin reciprocating compressor for different speeds (at average pressure of 80 p. s. i. (g))

Fig. 2 presents some experimental results of attenuation for a discharge pressure of 80 p. s. i. (g) using Freon 12 and varying the speed about 10 % of the compressors normal service speed. By means of a double bypass system presented on Fig. 3 a total attenuation

432

IIl-2

of 5 to 8 was realised. It is also noticeable from this figure that the peak to peak values of the initial pressure pulse were of the order of about 10 % of the average discharge pressure, and after using the attenuator this figure could be decreased to 1 .5 - 2 % .

Fig. 3 . Double bypass attenuator for a twin % H P reciprocating compressor (1470 r . p . m.).

Further experiments were carried out under different operating conditions on Freon 12 and 22 at discharge pressures up to 270 p. s. i. (g). Results of experiments showing the attenuation effect are presented in Table I.

Table I. Pulse attenuation on a discharge line of a refrigeration compressor under diffe-rent operating conditions.

Attenuation effect Discharge Suction (ratio of initial

Refrigerant Pressure Pressure r p m pulse amplitude to p.s.i. (g) p.s.i. (g) the amplitude of the

attenuated pulse.)

Freon 12 80 29 1470 8.0 " 100 30 1470 5.0

120 32 1470 7.0 " 140 22 1470 7.5

160 25 1450 9.5 " 180 26 1450 8.0

Freon 22 160 50 1480 5.8 " 180 52 1480 7.5 " 210 55 1500 6.8 " 240 55 1500 8.0 " 270 58 1480 6.4

On Fig. 4 an improved design of the interference attenuator for a double twin com­pressor with a nonsymmetrical crankshaft, shifting the phases of each pair of cylinders, is presented. Two identical attenuators were used for every twin discharge manifold and the flow obtained after passing both attenuators was rejoined at point C. Capacity trans­ducers were operating at point A (situated as close as possible to the beginning of the discharge line of the compressor) and at point B (arbitrarily selected on the downstream gasline after the junction C).

Resulting pressure-time diagrams are presented on Fig. 5. The initial pulse at point A is presented on the oscillogramm ,,a". In this case besides the two predominant harmonics noticeable high frequency oscillations were superimposed on the main pulse. These high frequency pulses are probably caused by mechanical vibrations of the valves.

433

III-2

On curve ,,b" a very remarkable attenuation can be seen and the remainder of the pulse practically does not differ from a comparison level of the ,,zero line" presented by trace ,,c".

Fig. 4. Interference attenuator for a double twin 7 HP compressor with 3000 r. p. m. synchro· nous speed.

I N I TIAL PULSE T "A" F IG . 4 .

kl 1 1 1 1 1 1 f1 r:rnrn ·1 T :�1. ZERO LINE OF INS TRU MENTAT ION

Fig. 5. Pulse attenuation for the 7 HP Freon 12 compressor (3000 r. p. m.) (at average pressure of I 35 p. S. i. (g))

Inasmuch as the problem of pulse attenuation is closely related to that of silencing the same approach seemed to have some prospects and therefore acoustical investigations were undertaken. Experiments with sound analyzing instrumentation showed a sound reduction of 1 5-20 db for low frequency sounds at 4 ft. distance from an open discharge of a reciprocating air compressor.

Several supplementary theoretical and experimental studies such as investigations of the effect of the attenuators on the performance of the compressor and its efficiency, should be carried out in the near future as this might have a significant influence on the extension of the destructive interference method of wave compensation to further practi­cal applications in adjacent fields.

434

Ill-2

The modem design of high speed multi-cylinder reciprocating compressors for Freon 12 and 22 has called for ever increasing improvement in performance, so that design of both pulse attenuators and silencers having negligible pressure losses merits some atten­tion.

In fact, for high speed refrigeration compressors working on Freons (gases having a relatively high molecular weight) the wave lengths of the predominant frequencies of oscillation are reasonably small and therefore the simplicity and small dimensions of interference attenuators might be especially attractive.

REFERENCES

r. Stephens, F. M. The Effects of and Corrections of Gas Pulsations Problems. Amer. Gas Journal, Vol. 164, 1946 p. p. 41-42 and 76.

2 . So/nick, R. L. and Bishop, R. H. Noise Vibration and Measurements Problems, Resulting from Flow Disturbances. Trans. A. S. M. E. Vol. 79, 1957 p. p. 1043-1056 .

.1 · Process Instruments and Control Handbook. McGraw-Hill, New York, 1957 (3.50-3.60).

435

IIl-26

Modern Measuring Technique for High-Speed Refrigerant Compressors

Technique moderne de mesure des compresseurs de frigorigene a grande vitesse

KAREL BRUNA Research Institute of Refrigeration and Food Engineering, Prague, Czecho­Slovakia

LUDVIK KUHN National Research Institute of Heat Engineering, Prague, Czecho-Slovakia

SOMMA/RE. Pour une analyse detaillee de la construction des compresseurs et de la verification de leur f onctionnement et de leur rendement, il est necessaire de connaitre la courbe de temps du mouvement de la soupape et les pressions a l'interieur du cylindre. Pour un parametre mesure, on utilise dans les deux cas la modulation de frequence d'un type

simple. Pour mesurer le soulevement de la soupape, on a mis au point un enregistreur a capa­citance et a inductance, n' ayant pas d'influence sur le mouvement de la soupape a proprement parler. On a mesure la pression avec un enregistreur a capacitance. Les efforts ont ete concentres sur la possibilite d'etalonner tout l'appareil en marche a l'aide d'une methode speciale d'enre­gistreur dijferentiel avec contre-pression. Le diagramme indicateur presente ensuite directement les points ou la pression mesuree dans le cylindre du compresseur est identique a celle d'une pression comparee de valeur connue. Cette methode permet d'eliminer /'influence nefaste des enregistreurs et des irregularites de la courbe zero. L'utilisation de l' enregistreur de pression avec etalonnage en marche permet de proceder

a des mesures, a partir desquelles it est possible de lire directement les pertes de charge dans les soupapes du compresseur. Lorsqu'on confoit des enregistreurs a capacitance pour l'action de la soupape, it faut tenir

compte du changement de la constante dielectrique des gaz de frigorigene en f onction de la pression. Par exemple, dans le R 12 la Constante dielectrique s' eleve de 2 a 3 % a une variation de pression de 1 atmg a 8 atmg. Les enregistreurs de pression indiques ci-dessus etaient utilises dans des essais sur un com­

presseur a R 12 ( alesage 80 mm, course du piston 63 mm, vitesses 960 et 1450 t/mn) . On indi­que les chiffres caracteristiques obtenus a l' aide de ces mesures. En conclusion, on presente quelques experiences d' ajustement de l' enregistreur au compresseur,

ainsi que les methodes d'evaluation des donnees oscillographiques.

1. INTRODUCTION

In detailed analyses and evaluation of the design of compressors and checking their performance and efficiency, it is necessary to know both the time curve of the valve plate lifts and the curve of pressures in the cylinder.

The task of determining these two variables has been solved by using inductive and capacitive pickups on the basis of frequency modulation of the carrier wave. A change in the mechanical quantity under study causes a change in the capacity or inductivity of the respective pickup which forms part of an oscillator resonant circuit ; thus the instan­taneous frequency of the carrier wave is a measure of the magnitude of pressure in the cylinder or of the valve lift. Czechoslovak patent No. 89268 has been advantageously used .in the design of these instruments.

437

111-26

2. MEASURING THE PRESSURE IN COMPRESSOR CYLINDERS

Capacitive method has been used for measuring pressures in the compressor cylinders. The pickup was installed in the valve seat in a manner so as to avoid increasing the dead space of the compressor or affecting the flow past the valve plate. The pickup is shown in Fig. 1 . A measuring diaphragm 31 screwed in the seat of valve 1 abuts on an insulating

Fig. r . Overall layout of pressure pickup.

ground ceramic plate 33. The thickness of this plate and the length of bronce electrode 32 are selected so as to balance the different coefficients of expansion of the materials used and render constant the initial capacity of the pickup when the component parts are uni­formly heated throughout.

Although the possibility of calibrating the pickups prior to measurement was ensured by auxiliary devices and unusual attention accorded to the stability and linearity of the apparatus, the overall accuracy of the values of instantaneous pressures read from oscillograms left much to be desired. In quantitative evaluation of the pressure curves -when comparing the efficiency of different types of machines, etc. - the above draw­backs can only be eliminated by calibrating the entire measuring equipment, i. e. the pickup, converter and oscilloscope, during the measurements.

3. CALIBRATING THE PRESSURE PICKUPS DURING MEASUREMENT

The staff of the National Research Institute of Heat Engineering in Prague developed a method for calibrating capacitive pickups of pressure in high-speed compressors during measurement as covered by the Czechoslovak patent No. 92971 of 1957. The principle of this method of calibration consists in obtaining - by means of a special differential pickup with low inertial massan impulses whenever the measured pressure p {t} is iden­tical with an arbitrarily adjustable comparative pressure p. In practical application the two pressures act on opposite faces of a piston which moves in a precision-ground cylinder. Depending on which pressure is higher, the piston assumes a certain limit position identity of pressures can then be defined by initiation of motion of the piston from its previous limit position. In one end position there is a contact seat insulated from the cylinder and the piston provides for electric connection. The entire equipment for measuring and cali­brating the pressure consists of two parallel branches according to Fig. 2. The first branch is used for continuous measurement of pressures and contains a capacitive pickup 1 and

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IIl-26

converter 2. The calibrating apparatus consists of a differential pickup 3, amplifier 4, differentiating circuits 5, 6 and delayer 7. Adding circuits 9, 10 and time base generator 8 are parts of oscilloscope 11 of common design with symmetric input of the horizontal and vertical amplifiers.

8

Fig. 2. Block diagram of equipment for calibrating pickups during measurement.

The time sequence of the inidividual signals is shown in Fig. 3 in which the following designations are used :

a) the pressure curve - p (t) b) time base voltage c) voltage on the differential pickup contact d) impulse behind the differentiating circuit 6 used for recording the horizontal

mark h

d

g

Fig. 3. Time sequence of signals in equipment for calibrating pressure pickups.

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III-26

e) delayed impulse behind the univibrator 7 for recording the vertical mark v f) voltage behind counter 10 g) voltage behind counter 9 h) curve recorded on screen of oscilloscope 11 where the point of pressure iden­

tical with the comparative value is defined by two mutually perpendicular im­pulses h, v

i) impulses from univibrator 12 for modulating the beam brightness

It follows from the above description of the differential pickup for determining the identity of the measured and comparative pressures that the source of largest errors is the finite value of the inertia of the moving piston. The difference in pressures acting on the two faces of the piston (Fig. 4) accelerates the piston. Assuming that the piston in the

Po

s

Fig. 4. Differential pickup.

first position closes the electric contact, then its very small travel so in time t0 suffices for opening the contact. According to the Newton's second law the distance travelled can at any instant be expressed by means of the known value of the accelerating force P ( t) and the piston mass m.

The maximum error in the method of calibration was evaluated for the case of pressure measurement in high-speed compressors. Since the opening of the contact of the differ­ential pickup takes place at a high-speed, this instant was used for forming the cali­bration mark and the pickup was designed so that contact opening takes place at increas­ing measured pressure, namely during compression. The values of instantaneous pressure p (t} and slope of the pressure curve can be derived by a simple calculation. During time to necessary for initiating the motion of the piston, the pressure in the cylinder changes by Lip. The error of the calibration method is consequently defined as the ratio of the pres­sure change Lip to the instantaneous value of pressure p (t) . After calculation we obtain

{ [me f sin wt]2 } 1_ � 0,0983 ;: (1 + n) " [cos2 �� + n]2-� 3

This relative error referred to the instantaneous value of pressure p (t} is plotted in Fig. 5 for frequency f = 25 c.p. s., i . e. 1500 r.p.m. A principle similar to the one used in the construction of the pressure pickup may be found in the works [l] and [2]. The proposed concept combines both the measuring of pressure and gauging pressure values in one unit.

The pressure pickup is placed in the compressor without using an indication channel with separating cock thus eliminating undesirable pressure pulsation before the pickup dia­phragm. The method of forming rectangular calibration marks on the recorded curve of pressure is quite new.

440

J [·�] ·----

__ ,__ 17

� v 2,5

, 2 I J \

I \ 7

0,5 -'--- -'-- , _ ·- ' . - �

I 0 20 40 60 BO fJO 120 U0 1ll0 180 -- •t

_ { (ir .. f sin •U' }T J - 0,0983 .. [ � ·t ] ' -It Po (f t- n J co,-- T "' n

III-26

Fig. 5. Error of calibrating equipment x = Poisson's constant (the ratio of specific heat during constant pressure and volume)

f = !1_ - Frequency of revolutions 2:n:

Po = initial pressure

n = Liv - the ratio of the dead space and

the cylinder volume.

4. MEASUREMENT OF VALVE PLATES LIFT

When analyzing the work of the compressor it is imperative to watch the movements of the plates of both the suction and discharge valves. The most simple method of measuring these values appears to be by pickups working on the capacitance principle ; from the viewpoint of working reliability then particularly those requiring least adjustments of the valve mechanism proper.

During the tests the following design of the valve-plate lift pickup proved successful (Fig. 6). A flat electrode (1) is insulated by a ceramic washer (2) from valve body (3). A rigid insulated cable ( 4) is led out from the compressor body, ending by a socket for

Fig. 6. Capacitive pickup of valve plate lift installed in discharge valve.

connecting the electronic apparatus. The characteristics of this pickup are not quite linear, particularly so in the case of valve plates of high lift. This is of no consequence, however, since prior to measuring the pickups are always gauged in connection with the measuring apparatus.

During these investigations the dependence of dielectric constant of freons on pressures has been established. The values of these changes are of such an order that they must be reckoned with when projecting capacitance pickups similar to that in Fig. 6.

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III-26

Fig. 7 shows the change in per cent of the dielectric constant for R 12 as related to pressure. The change in the value of the dielectric constant for R 22 has been established analogically.

This effect can for the most part be eliminated by suitable location of the pickup elec­trode, e. g. as shown in Fig. 6 which represents a pickup of discharge valve plate lift. Change of pressure in the active part of the pickup is represented merely by pressure loss in the discharge valve, that is not by the whole difference of pressure in the compressor suction and discharge.

I

0 4"'�--.--.--.-..-..-.--T"""-ir-i� g 6

Fig. 7. Dependence of the dielectric constant Lie in % on the pressure for R 1 2 . e For cases where this arrangement cannot be used, a lift pickup working on the in­

ductance principle is available. The motion of the valve plate affects the instantaneous value of the inductivity of the

pickup coil which forms part of the resonant circuit of an oscillator thus accomplishing frequency modulation of the carrier wave. In actual construction it was necessary to develop correct technological procedure for potting and cementing the coil which operates in an aggressive medium with a temperature of up to 150° C. Since the changes in inductivity are large, it was possible to divide the inductivity of the oscillating circuit into two parts -variable and constant which substantially improved the linearity of the pickup. Fig. 8. shows assembly of inductive pickup for plate lift of discharge valve of a R 12 compressor.

Fig. 8. Inductive pickup of valve plate lift installed in discharge valve.

The design of the valve-plate lift pickup is based on the works [2] and [3]. Elimination of additional condensers on the valve-plate proper resulted in simplifying the whole pickup and increasing working reliability. Nonlinearity of the characteristics is not a serious drawback deficiency in judging the valve performance.

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III-26

5. PRACTICAL APPLICATION OF THE MEASUREMENT TECHNIQUE

The use of pressure pickups with operational gauging permits such oscillographic records of the measured pressure values to be made, from which it is possible to read losses in the compressor suction and discharge valves. This is accomplished by the design of the pressure pickup provided with two differential pistons. As counter-pressure for the first one is chosen the value of pressure in the compressor suction branch, and the value of pressure in the discharge branch for the other.

During measurements on compressors with fewer cylinders (or running at low speed), pressure fluctuations in the compressor branches are encountered. This may unfavourably affect the performance of the differential pistons (the counter-pressure is not constant). For this reason constant pressure from refrigerant receiver is used in these cases. The num­ber of gauging values of counter-pressure is chosen according to the degree of accuracy required.

These methods of pressure and valve-plate lift pickup were used for testing a high-speed freon compressor (80 mm bore, 63 mm piston stroke, 960-1450 r. p. m.). Their application made it possible to carry out an analysis of working conditions of a compressor in the uni­flow and counterflow arrangement.

Records of oscillographic measurements must be carried out by means of film camera with continuous film feed. Photographing the individual curves of the measured values straight from the screen of a cathode oscilloscope is frequently a source of errors.

For simultaneous recording of a greater number of values measured a transistor amplifier has been developed permitting records to be made on a loop oscilloscope.

In conclusion characteristic records are given of measurements on the above mentioned compressor (Fig. 9 a, b, c). They clearly show gauging marks on the pressure curves. Pressures in the compressor suction or discharge branches were used as gauging counter­pressures. Fig. 9 a also gives a record of the valve-plate lift of the discharge valve made by means of the pickup described in Fig. 6.

Fig. 9. Record measurement of R 12 compressor a) pressure and lift of plate discharge valve b, c) pressure with gauging marks and dead point marks.

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III-26

REFERENCES

r . G. Lorentzen, On the Performance of Automatic Compressor Valves, Proc. IXth Int. Congr. of Refrig. 3078-90, 1955·

2. G. Lorentzen, Measurement of the Dynamic Performance of Automatic Compressor Valves, Proc. Xth Int. Congr. of Refrig., II, 163-170, 1960.

3. F. Sochting, Huboscillogramme von Ventilen raschlaufender Verdichter, Zeitschr. VDI, H. 3, 72-76, 1955·

4. V. Funer, E. SchOberl, R. Tauchmann und K. Bach, Indicating of Hermetically Sealed Com­pressors, Proc. Xth Int. Congr. of Refrig. II, 51-57, 1960.

444

III-33

Study of Factors Influencing the Volumetric Efficiency of Reciprocat­ing Compressors

Etude des facteurs influem;:ant le rendement volumetrique des compresseurs alternatifs

FRANTISEK SMUTNY Research Institute for Refrigeration and Food Engineering, Prague, Czecho­Slovakia

SOMMA/RE. Ce rapport traite de la determination des proportions des pertes indivi­duelles par rapport aux pertes totales correspondant au rendement volumetrique des com­presseurs a piston. Les resultats de mesures et /es calcu/s sont presentes pour deux mode/es de compresseurs a grande vitesse, c' est-a-dire un compresseur a ammoniac, avec un diametre de piston de 125 mm, une course du piston de 100 mm, a contre-courant, avec une vitesse de 960 t/mn et pour un compresseur a R 12, avec un diametre de piston de 80 mm, une course du piston de 63 mm, a flux continu, avec une vitesse de 1440 t/mn. Une bonne partie du rapport traite des pertes provoquees par la surchauff e des vapeurs passant de /' entree de /'aspiration du compresseur a l'entree des cylindres.

La relation entre /es pertes individuelles est definie ci-dessous: (1 - A) = (1 - Av) + (1 - At) + (1 - Au}, oil :

A rendement volumetrique Av rendement vo/umetrique theorique determine par l'espace mort, /'influence de la

surchauffe de la vapeur dans /es cylindres du compresseur et /'influence de l'etrangle­ment de la vapeur dans /es soupapes.

At = rendement determine par la surchauffe de la vapeur de frigorigene de l'entree de /'aspiration du compresseur a l'entree des cylindres.

Au = rendement determine par /es pertes par fuite des soupapes et autour du piston. Le rendement A a ete determine par pesee du fn'gorigene envoye par /es compresseurs. Le

rendement Av a ete determine a partir des diagrammes de pression p-v, obtenus par indi­cation de la pression dans /es cylindres a /'aide d'un indicateur electrique avec des detecteurs a capacitance. Le rendement At a ete verifie par mesure des temperatures du gaz a l'interieur des compresseurs. Le rendement Au a ete calcule a partir de la relation entre /es pertes.

Les mesures ont ete effectuees avec /es conditions de fonctionnement suivantes: temperature d'evaporation: to = + 5, - 10, - 25°C temperature de condensation: tk = + 30°C

Les resu/tats sont rassembles dans des tableaux e t des diagrammes avec un rendement },t indiquant /'influence de la surchauff e du gaz sur le rendement volumetrique des compresseurs. On passe en revue la quantite de pertes individuelles dependant de la temperature d' evapora­tion to. Les resultats peuvent etre uti/es pour la conception de nouveaux compresseurs et la proposition de leur application.

The paper deals with the determining of shares of individual losses on the total losses corresponding to the volumetric efficiency of piston compressors. Results of meas­urements and calculations are presented for two models of high speed compressors, i. e. for an ammonia compressor of counterflow design with a speed of 960 r. p. m. and for freon (R 12) compressor of uniflow design with a speed of 1440 r. p. m. A substantial part of the paper deals with the losses caused by overhe::iting of vapours from the suction neck of compressor to their entering into the cylinders.

445

IIl-33

1. INTRODUCTION

The aim of the paper was to analyze total losses corresponding to the volumetric efficiency of compressors. The results listed below were obtained from measurements of two models of compressors and, therefore, may be applied for similar compressors. The loss corresponding to overheating of vapours was determined by measurements of inside temperatures of compressors by means of thermocouples. After determining this influence it was made possible to determine the share of losses caused by leakages through suction and discharge valves and around pistons. Thus, it was possible to determine the shares of individual losses on the volumetric efficiency, corresponding to the prevalent factors of similar character.

The relation of individual losses is defined as follows :

(1 = A) = (1 - Av) + (1 - At) + (1 - Au) (1.1)

A = volumetric efficiency Av = theoretical volumetric efficiency determined by the clearance space and compri­

sing also the influence of vapour overheating in compressor cylinders and in­fluence of throttling of vapours in valves

At = efficiency determined by overheating of refrigerant vapours from the suction neck of compressor to the entrance into cylinders

Au = efficiency determined by losses through leakages through valves and around pistons

We are convinced that expressing the losses by means of summarizing formula is more suitable as compared with methods defining the resulting efficiency as product of in­dividual efficiency elements. For one thing it is logical and for the other it is possible to express by individual efficiency values directly the relation of true values to ideal ones. The differences (1 - A, . . . . . ) represent logically individual losses.

The tests and analyses were performed for single-stage compressors with following operating conditions :

evaporating temperature t o = + 5, -10, -25° C condensing temperature t k = + 30°C

Temperature of vapour in the suction neck of the compressor t1 for the evaporating temperatures t o were chosen according to Czechoslovak Standard corresponding to IIR Test Code.

446

--- • UtNr ""' #4

Fig. 1 . Temperatures in the suction neck 11 and temperatures in the entrance of cylinders 11 for NF 201 and 6 S com­pressors.

t 0 ° C

t1 ° C - NH3

t1 °C - R 12

+ 5

+ 15

+ 20

-10

± 0

+ 20

I II-33

-25

-10

+ 20

The true suction temperatures t1, which are practically identical to those of the Standard are outlined in diagram - Fig. 1 . They are mean temperatures measured on compressors according to Fig. 2 by means of a mercury thermometer and of a thermocouple, both numbered " 1". The differences between the two values were within decimals of ° C, which is a sufficient accuracy for this purpose.

Fig. 2 . Schematic illustration of the measuring of temperature

inside compressors NF 201 - longitudinal view

6 S - plane view

2. DESCRIPTION OF TESTED COMPRESSORS

The compressors concerned belong to two type-lines. Both lines are identical in design principles, differing only in one being of the counterflow, the other of the uni­flow arrangement.

2.1 Compressor NF 201 2-cylinder member from a line of 2-4-6-8-cylinder compressors with vertical, V, W

and VV arrangement. Cylinder bore 125 mm, piston stroke 100 m, counterflow design, tested at 960 r. p. m. with ammonia as refrigerant. Its photo in longitudinal section is shown in Fig. 3. Fig. 4 shows this compressor mounted in testing system. Clearance space of compressor e = 4,66 per cent.

Fig. 3. Longitudinal section of NF 201 compressor

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Fig. 4. Compressor NF 2or in test circuit

2.2 Compressor GS 2-cylinder compressor from a line of 2-4--6-cylinder compressors with vertical, V

and W arrangement. Cylinder bore 80 mm, piston stroke 63 mm. Uniflow design, tested at 1440 r. p. m. with refrigerant R 12. In Fig. 5 a photo of head section of this com­pressor is shown. Dead space e = 2,47 per cent.

Fig. 5. Sectional view of 6 S compressor head

3. VOLUMETRIC EFFICIENCY l AND LOSS (1 - A)

The efficiency and loss were determined by measurement on test circuits in deter­mining the weight of refrigerant delivered by compressors.

448

The volumetric efficiency A is obtained from the relation : G . Vi

A = -­Vp where G quantity of refrigerant delivered (kg/hour)

IIl-33

(3.1)

vi specific volume of vapours in the suction neck of compressor (m3/kg) VP = circumscribed volume of pistons (m3 /hour)

Table I . - Volumetric efficiency A and loss (1 - A) at t k = + 30° C

t 0 ° C A (1 - A) Type +5 -10 -25 +5 -10 -25

NF 201 0,86 0,735 0,57 0,14 0,265 0,43 6 S 0,89 0,835 0,695 0,1 1 0,165 0,305

4. EFFICIENCY Av AND LOSS (1 - Av)

This efficiency and loss were determined from pressure p-v diagrams, obtained by indication of pressure in the cylinders by means of an electric indicator with capacitance pickups. The p-T diagrams were obtained and transferred by means of gauging marks into p-v charts. The found values are listed in the following Table 2.

Table 2. - Efficiency Av and loss (1 - Av) at t k = + 30° C

t 0 °C Av (1 - Av) Type +5 -10 -25 +5 -10 -25

NF 201 0,931 0,886 0,824 0,069 0,114 0,176 6 S 0,958 0,935 0,885 0,042 0,065 0,115

As this efficiency has been taken from true p-v diagrams, it comprises also the influences considered in relation (1.1). A better efficiency of 6 S compressor is due before all to its smaller clearance space e. Note : For comparison, calculation of efficiency Av was made according to literature [1] from

the formula

l', c, (1 - , [(::) .:. - 1]) • (':)

' 1 + - 1 - m

P o P k . . d" d - - compression ratio correspon mg to temperatures t k an t 0• P o

(4.1)

The coefficients of expansion polytrop m were selected according to literature [2] NF 201 (ammonia) : m = 1,25 6 S (R-12) : m = 1,02

In the relation (4.1) the first part, marked A1 vi : expresses the influence of clearance space e, the other represents the influence of overheating of vapours in the cylinders of the compressor.

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The results have proved, that the values Av obtained from p-v diagrams lie between A' v and A' v1 ; e. g. at t o -10°C

NF 201 : A' v

0,781 A' Vl 0,905

Av (see Table 2.) 0,886

It can be derived, that the relation (4.1) is not quite suitable for high-speed compressors, as the vapours pass through the cylinder too quickly and have not enough time to absorb heat.

5. EFFICIENCY At AND LOSS (1 - At)

This kind of loss, as may be seen from the definition of efficiency At represents the influence of overheating of vapours from the suction neck of compressor to the entrance into cylinders or to the suction valves. As already mentioned above in the introduction (1.) the loss has been determined by measuring temperatures inside compressors.

The arrangement of thermocouples for measurements on the NF 201 compressor is shown in the longitudinal scheme of section, for measurements with 6S compressor ­in schematic plane view - Fig. 2. For better illustration refer once more to Fig. 3 and Fig. 5 mentioned above. To simplify the marking, the points of measurements on suction neck are marked "I", on cylinder inlet "2", regardless of the number of thermocouples. The temperature t2 is a mean value from 4 thermocouples. The ascertained temperatures t2 together with the mentioned temperatures t1 for both compressors are entered into the diagram - Fig. 1 . The difference Litt = t2 - t1 characterizes the overheating of vapours inside compressors and is entered into the diagrams Fig. 6 and Fig. 7.

_, � -• I, "e N

Fig. 6. Losses on volumetric efficiency A and overheating of vapours Litt in NF 2or compressor .

To determine the efficiency At the following consideration according to Fig. 2 was made :

In point "2" the true weight of refrigerant delivered by the compressor is :

G2 (kg/hour) = V2 (m3/hour) · y2 (kg/m3)

In an ideal case, without any overheating of vapours between points "1" and "2", then :

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Fig. 7. Losses on volumetric efficiency A and overheating of vapours Litt in 6 S compressor.

From that the resulting loss (1 - At) is

G21 - G2 V21 • Yt - V2 • Y2 (1 - At) = = ---=-----'--=-----"-----"

G21 V21 • Y1 As a compressor of the same volumetric characteristics is concerned, then:

V2i = V2 and :

v1, v2 - specific volumes of vapours (m3/kg).

· t 0

(5.1)

The specific volumes were read from the adjacent figure for the corresponding evap­oration temperature t0 and the temperatures t1 or t2•

From them the loss (1 - At) was calculated, e. g. for NF 201 compressor at t o = -10° C :

at: t1 = - 0,3° C,

t2 = + 14,6° C

V1 = 0,44 m3/kg

v2 = 0,475 m3/kg

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then:

0,475 - 0,44 (1 - At) =

0,475 = 0,0737 ,...., 7,37 %

and : At = 1 - 0,0737 = 0,9263 ,...., 92,63 %

Table 3. - Efficiency At and loss (1 - At) at tk = + 30°C

to ° C At Type + 5 -10 -25 + 5

N F 201 0,9772 0,9263 0,881 0,0228

6 S 0,9722 0,952 0,9282 0,0278

6. EFFICIENCY A u AND LOSS (1 - Au)

(1 - At)

-10 -25

0,0737 0,1 19

0,048 0,0718

The loss (1-Au) was calculated from the defining relation (1 .1) on the basis of all the losses evaluated above :

(1 - Au) = (1 - A) - (1 - Av) - (1 - At)

The results are listed again in the following table.

Table 4. Efficiency A u and loss (1 - ). u) at t k = +30° C

t o ° C

Type

NF 201

6 S

+ 5

0,9518

0,9598

). u

-10

0,9227

0,948

-25

0,865

0,8818

+5

0,0482

0,0402

7. CONCLUSION AND EVALUATION OF RESULTS

(1 - Au)

-10

0,0773

0,052

-25

0,135

0,1 182

The values of efficiency and losses enumerated in the foregoing paragraphs are listed in charts.

In charts Fig. 6 and Fig. 7 the individual kinds of losses in the contemplated range of evaporation temperatures are illustrated in per cents for compressors NF 201 and 6 S. The resulting curve, obtained by subtracting all losses from 100 per cent consequently represents the volumetric efficiency curve A. It may be seen that all kinds of losses expressed in absolute values are increasing along with decreasing evaporating temper­ature. For ammonia compressor NF 201 the loss (1 - At) is increasing more steeply as compared with type 6 S Freon compressor, which is caused by lower vapour tem­peratures t1 of NF 201 compressor resulting in higher difference in temperatures and also by smaller specific gravity y of vapours - or smaller molecular weight.

In this connection it may be interesting to refer to the chart Fig. 8, expressing effi­ciency At as function of vapour overheating LI t t (see also section 5). It should be borne in mind that the evaporating temperature for individual overheating values is different (see Figs. 1, 6, 7).

With increasing overheating a drop in At value, especially with NF 201 compressor, may be observed.

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IIl-33

'· .

II I '

Fig. 8. Efficiency At as function of overheating of vapours Lilt in NF 2or and 6 S compressors.

Absolute values of individual losses and individual efficiencies expressed in per cents are illustrated in the chart Fig. 9. Even here it may be seen that starting from a small value for t o = + 5° C the loss (1 - ).1) attains at t o = -25° C a greater value with the NF 201 compressor. In Fig. 10 the percentual share of individual losses on the total loss is illustrated. The greatest share - approximately 40 per cent - belongs to loss

--- N

----N

•J • (I .!,) (1 J. , } • ( 1 · '11 1 • t i " • ) •

Fig. 9 Fig. 10

Fig. 9. Values of individual losses on volumetric efficiency A for NF 2 o r and 6 S compressors. Fig. ro. Share of individual kinds of losses on volumetric efficiency A for NF 2or and 6 S compres­

sors.

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defined by efficiency Av, depending before all on the clearance space e. Less than 35 per cent corresponds to mean losses caused by leakage and a mean value exceeding 25 per cent is caused by overheating of vapours. It may be of interest to note the negli­gible share of (1 - Au) about t o -10° C, where the influence of spec. gravity of vapours and of pressure differences obviously give optimum conditions.

8. CONCLUSION From the analysis of efficiencies of two different single-stage compressors, of which

the NF 201 type was tested on ammonia and the 6 S type on R-12 refrigerant, the follow­ing conclusions may be drawn : a) for application of compressors :

With decreasing evaporating temperature t o all kinds of efficiency grow worse (Figs. 6, 7, 9) and it is therefore desirable to operate at highest possible evaporating tempera­tures (considering number of regulating stages in regulating schemes, especially in auto­matic systems). For compressors operating with R-12 the losses caused by overheating of vapours are lower than for compressors operating on ammonia (Fig. 8). b) for compressor design

The loss (1 - Av) forms the greatest part of the total loss. The clearance space should, therefore, be reduced to a minimum, especially in conditions where compressors operate at lower evaporating temperatures.

From the viewpoint of minimum losses (1 - At) care should be taken to provide the shortest possible track of vapour from the suction neck to the cylinders or to choose such flow track of vapours through compressor as to limit the time for heating of vapours to the minimum. Water cooling of compressors cylinders especially for ammonia re­frigerant should also be considered. Smaller losses in the 6 S compressor running at 1440 r. p. m. have apparently also a favourable influence on efficiencies At and A u (Fig. 9) and it i s possible to recommend from the viewpoint of volumetric efficiency A to increase rotational speed of compressors. In this respect the problems of energy balance should be studied more closely to determine the optimum.

As to the losses caused by leakage (1 - Av) care should be taken before all to main­tain sufficient tightness of the working valves. c) As concerns other models of compressors of both lines, before all of the NF line in 2-4-6-8-cylinder arrangement, inner temperature of NF 601 compressor was meas­ured. The ascertained values were lower only by 1 to 3°C as compared with the NF 201 compressor. The volumetric efficiency was found slightly improving with increas­ing number of cylinders. It means that individual efficiencies values for the rest of the models within the mentioned line will be practically identical with the values evaluated in this paper for the NF 201 compressor.

REFERENCES

r. R. Fuchs, E. Hofmann u. R. Plank, Leistungsversuche am Ammoniak·Verdichter ; VDI-Zeit· schrift Bd. 84, Nr. 16, 20. April 1940 (Especially the relation r .r).

2. I. Bendixen, The Volumetric Efficiency of Refrigerating Compressors. Proceedings of the X th International Congress of Refrigeration, Copenhagen, 1959/2 (Especially the relation 3.1).

3 . V. F. Chaikovsky, A . A. Shmiglya, K. I. Savkov, The Volumetric Efficiencies of Medium Capac­ity Refrigerating Compressors. Proceedings of the X th International Congress of Refrigeration , Copenhagen, 1959/2.

4. K. Linge, Der EinfluB des Ansaugezustandes auf die volumetrische und die spezifische Kalte­leistung. Kaltetechnik, 1956! Nr. 3.

LITERATURE

r . Backstrom M.: Kaltetechnik, Karlsruhe, Braun, 1953

r. Chlumsky Vl.: Pistove kompresory, Praha, SNTL, 1958

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Problemes de construction poses par une machine a un seul arbre comportant a un bout une turbine a vapeur et a l'autre bout un com­presseur centrifuge frigorifique a deux etages

Problems in the Construction of Steam Turbine Driven Two-Stage Centrifugal Compressors

M. Abel BEDUE, Ingenieur en chef, Groupe d'Etudes Mecaniques des Chantiers de l' Atlantique Saint-Denis-S/Seine, France

SUMMARY. The question of free space, the source of power available, and the various moving parts have lead to investigations into obtaining a compact, small-sized machine, to include a compressor with two cantilever turbine wheels, a turbine with one cantilever wheel, a reducer with two shaft outlets, an oil pump, a speed regulator with adjustable control and an overspeed device. Various problems of gas-tightness, assembly and dismantling are presented and reviewed. Resultant solutions are outlined with the aid of drawings.

Le systeme de transport de gaz sous forme liquide peut presenter des avantages a partir d'un certain tonnage par rapport a la solution classique de transport sous pression.

La transformation d'un petrolier de moyen tonnage en transport de gaz liquifie pre­sente sans conteste une etape interessante dans cette voie.

Dans !'ensemble de cette transformation, nous avions a resoudre les problemes de production frigorifique. En particulier, nous devions chercher a loger toute la centrale frigorifique dans un espace alloue qui etait tres exigu par rapport a la grande puissance qui etait demandee (2 fois 750.000 f/h a -18 + 40 aux brides du compresseur en marche normale et si possible 1 .000.000 de f/h a - 18 + 42 aux conditions poussees). De plus le bilan electrique du navire imposait que toute l'energie necessaire soit fournie par la vapeur aussi bien pour l'entrainement des compresseurs que pour celui des deux pompes principales necessaires pour chacune des installations frigorifiques.

ARCHITECTURE GENERALE DE LA MACHINE

Dans !'ensemble de la centrale frigorifique, la place retenue comme pouvant convenir aux compresseurs et aux pompes permettait tout juste !'installation d'un ensemble clas­sique : compresseur, turbine et accouplement ainsi que les pompes principales necessaires avec leurs turbines d'entrainement.

Les exigences de securite en mer et diverses autres considerations imposaient que l'on puisse installer deux compresseurs et leurs auxiliaires.

En plein accord avec les representants de l'armateur, nous avons done pense qu'une machine speciale ne comportant qu'une seule turbine motrice pour entrainer a la fois le compresseur (pour 650 CV environ) et les deux pompes principales (chacune de 50 CV a 1 .500 t/rnn) permettrait une reduction sensible de l'encombrement et par fa rendrait possible !'installation de deux ensembles identiques cote a cote dans l'espace reduit qui seul etait utilisable.

Le cycle frigorifique a utiliser (- 18 + 42) et la puissance desiree (1.000.000 de f/h) imposaient un compresseur centrifuge a deux etages. Nous avons done corn;:u une machine oil ce compresseur a deux etages etait monte en porte a faux a une extremite d'un arbre, l'autre extremite comportant une turbine monodisque a vapeur dont le rotor etait monte lui aussi en porte a faux.

L'ensemble rotorique ainsi constitue reposerait sur deux paliers dans un carter central qui contiendrait aussi le reducteur avec deux sorties d'arbre de 50 CV environ chacun a

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1 .500 t/mn. L'arbre secondaire serait perpendiculaire a l'arbre principal de fa�on a pouvoir ramasser les pompes aux maximum contre le carter et reduire ainsi l'encombre­ment total.

A partir de cette conception generale d'ensemble et en tenant compte du fair que les demontages et remontages a bord sont toujours delicats du fair d'une plateforme mou­vante et que de plus dans le cas particulier ils devaient se faire dans un espace reduit, nous nous sommes imposes d'avoir des elements de machines susceptibles d'etre de­montes manreuvrables par un ou deux hommes au maximum.

COMPRES SE UR

Le stator du compresseur est sans joint horizontal et le montage des differentes pieces s'effectue, comme un empilage d'assiettes. Les roues en alliage leger sont du type ferme a aubes fraisees dans la masse du corps. Le voile est soude aux extremites peripheriques des ailettes, la fixation est renforcee par des rivets places dans l' epaisseur de chaque ailette. Les roues des deux etages sont fixees l'une contre l'autre en appui sur une large galette portant des labyrinthes afin de former un rotor monobloc. L'etancheite est assuree par une garniture double it pression d'huile. Le faible espace entre l'arrii:re de la roue et le palier qui etait necessite par le desir que nous avions d'avoir une machine nettement sous critique, meme au declenchement, nous a contraints a dessiner et it construire une garniture double speciale avec l'aide d'un specialiste des garnitures (PACIFIC). Le circuit d'huile de cette garniture est independant et comporte sa propre pompe, son filtre et son refrigerant. Le reservoir d'huile de retour permet de controler l'etancheite de la garniture a tout moment.

Fig. 1. Intersection du compresseur

CARTER CENTRAL, REDUCTEUR DE VITESSE

11 est it joint horizontal, ce qui permet, grace a une etude appropriee de ses liaisons avec le corps du compresseur une ouverture totale pour la visite sans qu'il soit besoin de toucher ni au compresseur ni a la turbine. La butee Michell est calculee largement pour absorber les poussees dues au compresseur et a la turbine ainsi que celles dues a des coups de liquide accidentels.

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© ® © ® © ® @ © ©

Fig. 2. Ensemble de la prise de mouvernent

La position transversale de la prise de mouvement des pompes a saumure et a eau de mer qui nous avait ete imposee par des considerations d'encombrement total du groupe nous a conduits a la chaine cinematique de la reduction que nous avons adoptee a la suite d'une etude commune avec notre fabricant habituel d'engrenages. En particulier, l'arbre de prise de mouvement est compose de trois elements de fac;:on a permettre l'ac­cessibilite de tous les organes du reducteur sans deplacer ni desaligner !es deux pompes entrainees et sans avoir a deposer l'arbre principal.

La prise de mouvement des pompes commande l'arbre de la pompe a huile attelee et le regulateur a huile avec son servo-moteur.

SOCLE

II est commun a !'ensemble carter central compresseur et a la turbine. II forme reser­voir a huile pour le circuit de graissage principal.

Le corps de la turbine n'est pas fixe sur le carter central mais sur un support double faisant partie du socle. Cette fixation est assuree par des clavettes qui laissent la libre dilatation du carter tout en assurant la permanence du centrage entre le corps et le rotor de la turbine.

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�!""" hr!Pnt. t'alrf!srtU- l#dl/11'11

/ � iMnttllll1n. :!" m' ¢-/

Fig. 3. Cha!ne cinematique du reducteur de vitesse

Fig. 4. Turbine

REGULATION

Un servo-moteur assure le laminage de la vapeur apres la soupape d'admission pour maintenir la vitesse de la turbine a la vitesse affichee au regulateur quelle que soit la charge de la turbine. La vitesse affichee peut etre modifiee par la simple manreuvre d'un bouton.

La variation de puissance du compresseur et son adaptation aux besoins frigorifiques sont assurees par une prerotation dont la commande est soit manuelle soit asservie a la temperature.

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SECURITE

La soupape d'admission comporte un verin a fermeture rapide. Ce verin peut etre declenche soit manuellement par un coup de poing soit par une chute de pression d'huile du circuit principal.

La chute de pression d'huile peut etre provoquee soit par un defaut de graissage soit par le declencheur de vitesse qui agit sur la pression d'huile de retenue du verin de fer­meture rapide.

L'attitude habituelle et parfaitement comprehensible des armateurs est de n'accepter pour les installations de bord que des machines deja eprouvees a terre et presentant des qualites de robustesse a toute epreuve.

11 faut des imperatifs vraiment exceptionnels pour que l'on consente a utiliser des machines prototypes a bord des navires.

En effet, elles constituent un point sensible et accaparent a elles seules une attention que le mecanicien du bord n'a pas toujours le loisir de leur donner malgre son plus vif desir. En raison de la diversite des problemes a resoudre pour la construction des deux machines prototypes a installer a bord du navire, on comprend aisement que l'armateur et le constructeur aient prevu, d'un commun accord, la construction d'une troisieme machine destinee a des essais a terre.

Cette machine devait etre prete aux essais six mois avant la mise en place des deux machines de la centrale frigorifique ; pendant tout ce temps, elle devait servir de bane d'essais et de mises au point.

Malheureusement pour nous, nous n'avons pu, a cause de circonstances imprevisibles dues pour la plupart a la reconversion de certaines de nos activites, suivre ce programme raisonnable et c'est a bord que nous avons du proceder aux essais et a la mise au point.

Tous les gens du metier savent ce que cela represente d'efforts et de tenacite tant de la part des ingenieurs que de celle des monteurs et je n'ai rien besoin d'ajouter a ce sujet sauf de vous dire combien la collaboration avec les representants de l'armateur etait necessaire : elle s'avera amicale et efficace tout a la fois.

Maintenant avec deux annees de recul et de fonctionnement, nous pensons que les lignes generales etaient bonnes et que seules quelques modifications de detail, pour rendre le montage et le demontage plus aises, seront effectuees lorsque nous aurons a fabriquer des machines semblables.

DISCUSSION

L. Vahl, Netherlands : I should like to know how the leakage of refrigerant through the shaft packing can be observed if the seal pressure is controlled by an oil reservoir.

A. Bedue, France : The oil circuit was completely independent and at a slightly higher pressure than the refrigerant pressure. Therefore, if refrigerant went into solution with the oil this could easily be observed, and if there is any shaft seal leakage it would be oil rather than refrigerant.

A. Neuenschwander, France : My intervention was purely designed to clarify the discussions in connection with the possibility of refrigerant leaks to the outside. I confirm that the seal housing is always maintained at an oil pressure above that of the pressure which exists in the rest of the compressor housing, either by means of auxiliary pump inserted in an independent circuit or a reservoir in which the pressure can be controlled. Therefore, one does not actually have a refrigerant leak towards the outside but an oil leak towards the outside, which of course, is immediately apparent.

Gerard, Belgium : What are the dimensions of the complete assembly and also for an indication of the capacity and weight as compared with an open type system.

A. Bedue, France : The unit had a weight of 1 . 7 tons at a capacity of 1 .000.000 frig­ories and an external volume of 700 millimetres by 1 metre. The speed varied between 8.000 and 12.000 r. p. m.

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V. Fiiner, Germany : I should like to ask Mr. Bedue if there was a significant heat transfer between the very hot turbine and the rather cold compressor and if this influ­enced the compressor performance.

A. Bedue, France : There was very little transmission of heat from turbine to compres­sor and there was no significant capacity reduction.

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Ammonia Centrifugal Refrigeration Plants in Operation

Fonctionnement des grandes installations frigorifiques centrifuges a ammoniac

Dipl.-Ing. K. H. RICHTER Berlin-Tegel, ZiekowstraBe 139, Germany

SOMMAIRE. Depuis 1950 a debute un developpement inespere de l'economie europenne, qui s'est combine a un developpement rapide des industries chimiques. Lesfabricants de materiel frigorifique ont rei;u des demandes d'installation de puissances frigorifiques encore inconnues, qui ont f ait substituer aux system es frigorifiques utilises jusqu' alors avec leur multitude de grands compresseurs alternatijs des systemes frigorifiques centrifuges peu encombrants.

II existe de nombreuses raisons pour lesquelles en particulier l' ammoniac a maintenu sa posi­tion comme frigorigene pour les systemes frigorifiques centrifuges dans ces grandes installations. De plus, au cours des dernieres annees, il a ete realise des progres considerables dans la con­struction des compresseurs centrifuges qui ont une portee sur la limite des puissances, des taux de compression la securite du fonctionnement et la possibilite de reglage. Les installations frigorifiques centrifuges a ammoniac modernes sont caracterisees par leur

grande economie qui est obtenue par combinaison appropriee du cycle frigorifique utilisant la detente et le refroidissement intermediaires. Pour le fonctionnement a charge partielle au­dessous de la limite de pulsation du compresseur centrifuge on utilise les refroidisseurs a injection mis au point recemment. Les stations d' evaporation frequemment place es a une distance de I km des installations de compresseur centrifuge sont alimentees par de l' ammoniac a l' aide de tuyaux d'aspiration et de liquide non isotes. Avec les differentes possibilites de combinaison de ces installations frigorifiques centrifuges, utilisant un ou plusieurs etages d' evaporation, il resulte differents prix d'installation et d'exploitation de sorte que /'on peut trouver une solu­tion optimale pour chaque application. Pour lafacilite de surveillance et un mode defonction­nement simple et sur, ii apparait ca et la la necessite d'un compromis.

The first ammonia centrifugal refrigeration plant was built in Germany already in 1924 [1]. This was followed by a period in which ammonia lost ground to other refrig­erants which had been newly introduced. But in the years after 1950, favoured by the advent of new technological processes, there was a sharp increase in the demand for · centrifugal ammonia compressors. Many remarkable large-capacity refrigeration plants have since been built, and there has been considerable progress in this field.

From a process engineering point of view, ammonia is clearly superior to other refrigerants under economy, refrigeration engineering and operational aspects despite of its toxicity, low molecular weight and the large pressure ratios that must be handled. Moreover, ammonia centrifugal refrigeration plants are now considered as safe as other industrial equipment operated on steam, gas, fuel oil or electric power.

Modern ammonia centrifugal refrigeration plants are made for evaporating tempera­tures between + 10 and -50°C ( +50 and -58° F), these temperatures defining the range of interest in industry [2, 3].

The minimum refrigerating capacity of an ammonia centrifugal is dependent upon the volume of refrigerant vapor discharged from the final compressor stage. Other factors involved are the design of the refrigeration cycle and the mode of flash chilling of the liquid refrigerant and respectively the mode of intercooling of the compressed ammonia vapors. Judging from the large number of existing installations, one may safely state that the minimum capacities for which economic unstaged centrifugal plants can be built range from 2.6 X 106 kcal/h (860 tons) at -15/+35° C (+5/+95°F) and 3.3 x 106 kacl/h (1 100 tons) at + 10/+35 ° C (+50/+95°F). This corresponds to a final

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volume of about 1500 m3/h (900 cfm). Low-stage booster centrifugals for lower evapo­rating temperatures can be used for smaller minimum capacities depending on the pressure ratios involved.

Today, eight wheels on a common shaft is considered an upper limit by centrifugal equipment designers. A typical example of such a machine is shown in Fig. I . It must also be considered that heavier shafts necessitate larger hub and outside diameters with resulting efficiency drop. Modern unstaged centrifugals can cover a temperature

Fig. I. Eight-stage ammoma centrifugal compressor with double intercooling, for 6 x 106 kcal/h (3000 tons) at --r7/ +40°C ( + r/ + ro4°F).

range defined by evaporating temperatures from -20 to -15° C (-4 to +5°F) and condensing temperatures from + 35 to + 40° C ( + 95 to + 104 ° F). Evaporating tem­peratures of less than -20° C (-4 ° F) require staged units comprising two compressors in series connection.

When planning a new ammonia centrifugal refrigeration plant, a decision must first be reached on how to provide for intercooling of the compressed vapors in order to work out the most economical solution. In Fig. 2, five different curves are shown in the specific refrigerating capacity versus evaporating temperature diagram at the left. The corresponding refrigeration cycles are shown schematically in the smaller pressure­enthalpy diagrams at the right. Curve a represents the values which can be achieved theoretically according to the Rules of Refrigeration Machinery. Curves b through e are for practically possible ammonia centrifugal compressors of 5 X 106 kcal/h (1650 tons) refrigerating capacity at evaporating temperatures ranging from + 10° C ( + 50° F) to -20°C (-4° F) and at a constant condensing temperature of +35° C (+95°F). Pres­sure losses in the lines and in the intercooler have been considered, so that the cycle diagrams represent conditions encountered in actual practice. Curve b is for a centrifugal refrigeration plant in which the compressed ammonia vapors are intercooled close to saturation in an injection type intercooler. Curve c, which is next from top to bottom, indicates that a process including flash chilling in an economizer is somewhat less favourable though requiring less first cost for apparatus. Curve d characterizes inter­cooling of the refrigerant vapor to +35°C (+95°F) in a surface cooler by means of cooling water. Curve e is valid for singlestage compression without intercooling. As compared to this single-stage compression process the average power savings with the

462

-20 -15 -10 -s 0 •S Evaporation temperafure

+10 °C

'kl]f p l 0LL:t7 � p l 0 /l/7i L_d_

i

p l 0/f/7 �

III-20

Fig. 2 . Specific refrigerating capacity of ammonia centrifugal refrigeration plants vs . evaporating temperature for various refrigeration cycles. a theoretical specific refrigerating capacity in kcal/kWh, b refrigeration cycle with intercooling in injection cooler, c refrigeration cycle with flash chilling in economizer, d refrigeration cycle with intercooling in surface cooler, e refrigeration cycle with single-stage compression.

other designs are as follows : 7 % for injection intercooler; 4% for economizer; and 2 % for surface cooler. A surface cooler is not worthwhile with evaporating temperatures exceeding 0°C ( +32° F) as the gain produced by intercooling will be eaten up by the pressure drop in the cooler and in the associated piping.

Large chemical and processing plants normally operate their own steam and electric generating systems for reasons of heat and power economy. These centralized heat/power plants supply their steam and electric power to the individual consumers. Unfortunately it is scarcely known so far that also refrigeration can be generated in a centralized plant and distributed to remote consumers over long distances at no loss of economy. The consumers simply cover their refrigeration demands and need not care about the details of low-temperature generation. They may indirectly cool their products or processee by means of evaporating ammonia. Where this is not permissible, the cold may be trans­ferred locally to a suitable coolant (e. g. water, ethylene glycol solution, calcium chloride brine, methanol, etc.) which is then used to cool the product. Moreover, consumers have the advantage that they can draw from the refrigeration supply system at any time.

Centralized refrigeration supply depends on that all consumers adapt themselves to one or two evaporating temperatures much in the same way as steam is normally supplied by centralized generating plants at a limited number of fixed pressure levels. Refrigera­tion consumers requiring lower evaporating temperatures for special applications may have their own low-temperature refrigeration plants (which must then be operated with suitable low-temperature refrigerants) connected to the general supply system in cas­cade.

Fig. 3 shows how the centralized refrigeration plant 1 is connected to the various consumers by a system of ring mains and branches. In this particular instance it oc­curred that a number of years after completion of the basic system a new large-scale production was started at a peripheral location within the same chemical plant, bringing a sharp increase in refrigeration tonnage demand. It was decided to install an additional refrigeration plant 2 near the new production facilities to ensure direct refrigeration supply to the product chillers over distances of only 50 to 350 m (160 to 1 150 ft.).

The two centralized plants 1 and 2, which are about 1600 m (5250 ft.) apart, were however tied together by non-insulated pipe lines interconnecting both the liquid

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111- 20

Fig. 3. Refrigeration supply system in a large processing plant. I centralized centrifugal refrigeration plant 1 , 2 centralized centrifugal refrigeration plant 2 , 3 injection coolers, 4 liquid sub coolers. 5 cold exchangers, 6 evaporators.

ammonia sides and the suction sides of the two evaporating stages. This method allowed to realize compound operation in such a way that sudden peaks in refrigeration demand at plant 1 can partly be covered by the capacity of plant 2 and vice versa.

Modern practice is to convey ammonia refrigerant, both liquid and vapor, over long distances through non-insulated pipes. The hundreds of meters of pipe length connec­ting plant 1 with the consumers were installed on overhead piping bridges without any insulation whatsoever. The need for costly insulating jackets, especially on the large-diameter refrigerant suction lines, was deliberately avoided by providing each remote evaporator with a cold exchanger as shown in Fig. 3. In this exchanger the ammonia vapors emerging from the evaporator will cool the incoming warm liquid while being lifted to the ambient temperature, e. g. + 20° C ( + 68° F). Thus the refrigerating capacity is not wasted, for on the other hand the liquid ammonia is subjected to intense sub-cooling before entering the injection valve.

Before entering the centralized centrifugal compressor plant 1 the ammonia vapors arriving from the various directions are collected in suction manifolds, one for each evaporating stage, and then discharged into the injection coolers where the vapors entering at ambient temperature are cooled down close to saturation temperature by the injection of liquid ammonia.

Fig. 4 is a view of the apparatus and the compressor house of the centralized refrig­eration plant. In the lefthand foreground there are two ammonia storage tanks with the vapor collectors on top. The righthand part of the elevated platform carries four injection coolers and surface type intercoolers behind. The lefthand part of the platform carries a group of horizontal shell and tube condensers. The centrifugal units are housed in the illuminated building visible in the background.

It may be interesting to learn about the principal data of these two centralized large­capacity refrigeration plants which are the largest we know of in the world. The com­pressor house of plant 1 accomodates ammonia centrifugals for the -S° C ( +23°F) system with a combined refrigerating capacity of 22 x 106 kcal/h (7300 tons). The units are driven by backpressure steam turbines of 9000 kW combined capacity. A low­pressure centrifugal for the -2S ° C (-13°F) stage has a capacity of 2 X 106 kcal/h (660 tons) and uses a 640 kW electric motor as a driver. The same machine house further accomodates a standby capacity of 6.6 X 106 kcal/h (2200 tons) at -S° C (+23°F) consisting of six pre-war reciprocating compressors.

The centralized refrigeration plant 2 again consists of centrifugals with steam turbines and electric motors as drivers. In this case the refrigerating capacities available are S.S x 101 kcal/h (1800 tons) at -2S° C (-13°F) and 6 x 101 kcal/h (2000 tons) at -S° C (+23°F).

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Fig. 4. Centralized ammonia centrifugal refrigeration plant with machine house and apparatus.

Piping dimensions are similarly impressive. As an example, the suction mains at plant 2 have nominal widths of 600 and/or 400 mm (24" and/or 16") and the liquid mains have nominal widths of 150 and/or 125 mm (6" and/or 5"). The pressure line to the condensers has a nominal width of 500 mm (20"). The non-insulated interconnec­ting lines between plants 1 and 2, mounted on a modern piping bridge, have nominal widths of 600 mm (24"), 400 mm (16") and 150 mm (6"). Either centralized plant has injection coolers for ammonia vapor back-cooling and liquid coolers operated on evaporating ammonia in order to enable bubble-free refrigerant exchange between the compounded plants. Vapor flow velocities in the large-diameter suction mains range from 20 to 30 m/sec (66 to 98 fps) depending on the nominal widths, and liquid flow velocities are in the order of 1 m/sec (3.3 fps.).

It is about time that those concerned with modernization or new planning of large manufacturing or processing plants start thinking of centralized refrigeration supply systems; for it must be borne in mind that first costs and operating expenses are lower for a centralized refrigeration plant than for a multitude of decentralized units. Also the shortage in human labor renders it increasingly difficult to operate many small­capacity refrigeration units and to prevent troubles and failure. With a centralized system, peaks in refrigeration demand can be handled with less difficulty and all con­sumers may rely on a common standby capacity. In addition, under the current trend of development, it is more economical to generate refrigeration in bulk by means of centrifugal compressors which in turn can be driven not only by electric motors but also by steam turbines, thereby enabling better energy balances. Modern ammonia centrifugal refrigeration plants offer new possibilities to the industries.

REFERENCES

r . H. Voigt, Kompressoren fiir grol3e Kiilteleistungen ; Z. VDI 71 (1927), No. 33, pp. n45/n53. z . Hiitte des Ingenieurs Taschenbuch, 28th edition, part IIB, table 4, p. 221 , Berlin 1960, Wilh.

Ernst & Sohn. 3. N. Wemhoner, R. Goldner, Turboverdichter in Kiilteanlagen for Industriekiihlung und Klima­

technik; Kiiltetechnik 14 (1962), vol. 10 u. II, pp. 326/332 and 354/36r.

DISCUSSION

G. Lorentzen, Norway : I think that although centralization may be cheaper in capital and running costs in some cases, this is not universally true. In other cases smaller units, decentralized, might be technically and economically more feasible. It all depends on the particular circumstances, and each case has to be judged separately on its own merits.

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K. H. Richter, Germany : This depends very largely on the refrigerant being used. With Freon refrigerants it would not be possible to centralize too much, but ammonia leans itself very well to long pipe lines. In this particular case an old chemical works plant was enlarged and was required to work at various evaporating temperatures, i. e. 23°C and -30° C, and by putting a cascade system in, -70° C were reached.

H. Seller, Germany : I should like to know which was the lowest ammonia evaporating temperature employed till now on such plants for an ammonia condensing temperature of about + 30° C. The other point which I should like to raise is, that refrigerants such as propane or propylene are sometimes available as by-products in petro-chemical plants and may well be used in such plants.

K. H. Richter, Germany : The temperature was -45°C using booster compressors in conjunction with absorption systems as a high stage. The relation of such low tempera­tures to normal condensing temperatures is purely a question of the number of rotors in the housing which of course was limited to 8. It is therefore not advisable to use the multi-rotor machines direct between these low temperatures and normal condensing temperatures. With regard to the availability of propane or propylene, these are consid­ered to be suitable refrigerants.

A. Neuenschwander, France : There is really no reason why the Freons should not be used from the thermodynamic and efficiency points of view.

K. H. Richter, Germany : Freon utilization was not impossible, but the size of the pipe lines became very large indeed, i. e., instead of up to 12 inches diameter they would have to be increased to 1 Yz metres diameter. In the plant under consideration Freon was rejected after it was found that the pipe sizes were so enormous.

0. Cervenka, Czechoslovakia : What was the range of the capacity regulation of the individual compressors and the degree of automation of operation of the whole centralized plant ? I am also interested to know the system of regulation of the liquid supplied to the evaporators.

K. H. Richter, Germany : The plant was fully automatic, i. e., one booster and one high stage machine starting up together. The pumping limit was 65 % of full capacity below which injection was required. The liquid supply regulation to the evaporators was by low side level controls supplied by many manufacturers.

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Influence des proprietes physiques des frigorigenes sur les conditions de fonctionnement des turbomachines

Influence of the Physical Properties of Refrigerants on the Operating Conditions of Turbo-Machines

Dipl.-Ing. ALBERT BIERMANN Brown, Boveri & Cie, A. G., Mannheim, Germany

SUMMARY. The accurate determination of the thermodynamical magnitudes for the sizing of a thermal machine which has to operate with a true gas is only possible with the aid of accurate data, especially for Ks and z. Where the latter are lacking, the aspect of transfor­mation - and in consequence the magnitude of state - can be obtained graphically in a log P-i diagram. The relationship is determined between the principal characteristics of a compressor and the

properties of a gas such as Ks, M, z. From the required similarities for thermal machines, one derives the operating characteristics for a machine which, if the dimensions have been fixed, may be utilised with different refrigerants. Furthermore, with a similarity in flow, the same capacity is almost maintained if the ma­

chine operates at a condition where the refrigerant has the same reduced temperature. Measured values confirm calculated results.

1. BASES THERMODYNAMIQUES GENERALES

Tous les gaz qui entrent en ligne de compte comme fiuides moteurs dans les turbo­machines sont des gaz reels. Du point de vue de leur comportement thermodynamique, ils peuvent etre traites dans certains domaines en tant que

gaz ideaux

ou vapeurs ideales.

D'une manii:re generale, !'equation d'etat d'un gaz ou d'une vapeur peut s'ecrire sous la forme simplifiee suivante :

(1)

Pour un gaz ideal, le facteur de compressibilite z = 1, Pour une vapeur ideale, z est une fonction de l'entropie z = f (s), Les frigorigenes usuels se comportent comme des gaz reels dans le domaine de la

vapeur surchauffee. Le facteur de compressibilite est fonction de la pression et de la temperature pour un gaz reel, z = f (P, T).

Des relations generales sont valables pour un gaz s'ecoulant dans une machine ther� mique.

Ainsi pour une transformation isentropique (compression ou detente), le travail mis en jeu est:

(2)

Si !'equation d'etat est donnee sous la forme v = f (P, T), la valeur de Ks peut etre determinee a l'aide de Cv.

Chaleur specifique

a volume constant : Cv = Cvo + T J ( �2�) v

. dv (3)

(cvo = chaleur specifique a volume pour un gaz ideal).

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Exposant isentropique :

Ks = - � (<5 p) + I_-__:;__ (<5 p)2 = KT · � (4) P <5 V T P • Cv <5 T v Cv

Pour un gaz ideal, l' exposant isotherme KT = 1 ; pour un gaz reel, KT :f 1, d' oil il s'en suit que Ks :f cp/cv.

La Fig. 1 represente I' allure typique des courbes Ks pour du R 11 .

1 1$ ks --­r---·� • o "'

� 1 10 � z

' !OS

f. 00 0

"'""'=-::----------J0�-----1 '2

so 100 °c

Fig. Variation de l'exposant isentropique Ks en fonction de la pression et de la temperature pour le frigorigene R r r .

Mais comme l'ecoulement dans une machine thermique est toujours irreversible, le travail technique extfrieur est:

K [ n - 1 ] K [ n - 1 ] a = Ks

8 1 · P1 · v1 :n: -n-- 1 = ks

81 · z1 · R · T1 n-n-- 1

Nous avons pour un gaz ideal ou une vapeur ideale: n - 1 Ks - 1 -n- = K� .. . <p

relation dans laquelle pour la compression

<p = -1) pol pour la detente <p = 7)pot

(5)

(6)

Selon Dzung [1], cette relation doit etre completee par un facteur de correction pour les gaz reels:

n - 1 Ks - 1 m - 1 -n- = � • <p + (l-<p)-m-

ou m est l'exposant pour une transformation isenthalpique.

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(7)

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Dans le cas d'un gaz a variable Ks, il faut placer dans les relations (2) et (5) la valeur

moyenne de la transformation, moyenne arithmetique pour K � 8 1 et, dans les crochet�,

Ks - 1 moyenne logarithmique pour l'exposant � .

Le rapport des travaux isentropique Ys et reels a represente le rendement isentropique y s 1} is = a

pour la compression

a 1}is = -Y s

pour la detente

(8)

Le rendement polytropique correspond au rendement polytropique d'etage, pertes supplementaires aux baches d'entree et de sortie inclues. 11 se differencie du rendement isentropique par le fait que dans le compresseur !es pertes d'echauffement surchargent encore l'etage (1Jpo1 > 1]1s) alors que dans la turbine ces pertes d'echauffement doivent etre partiellement recuperees (1}pol < l'/is).

Le rapport de temperature dans une transformation isentropique est donne par:

Ks - 1 Ks - 1

Ks OU :n Ks (9)

Si l'on remplace dans !'equation (2) K

Ks - 1

par (9), le travail technique interieur :n s

Y• n'est plus qu'une fonction de la temperature. Si l'on prend le kmol pour unite de masse dans les relations (2) et (5), R l'identifie

avec la constante universelle des gaz ffi = 8314 J/kmol ° K. On voit que dans ce cas le travail technique interieur Ys d'un gaz quelconque, pour les memes conditions de fonc­tionnement :n et T1, depend principalement de l'exposant isentropique Ks. Lorsque

Cornprrzssion

A S 12.ls

cxpan.rt'on

i bS = .6 Si · {1 - "lpo/] IJ..is = t� - l---, :> "£' ?�.s> - z-: !:>o/

Fig. 2. Trace de la transformation d'etat pour la compression et !'expansion a !'aide de !'aug­mentation d'entropie.

469

III-24

l'on s'eloigne considerablement du comportement ideal, cette derniere influence s'ajoute a celles du facteur de compressibilite z et de l'exposant isenthalpique m.

Pour beaucoup de frigorigenes, les valeurs des coefficients Ks, z1 et m ne sont pas connues. On utilisera dans ces cas le diagramme log P - i pour determiner !'allure de la transformation et les valeurs correspondantes.

Si le rendement polytropique d'etage est donne pour une machine, on peut tracer la transformation dans le diagramme par la methode de Zweifel [2], que ce soit une com­pression ou une expansion. Les valeurs des grandeurs d'etat etant ainsi connues, on peut obtenir les chutes d'enthalpie, les temperatures et le rendement isentropique de la machine. Cette methode est expliquee dans la Fig. 2.

Les relations donnees jusqu'ici forment la base thermodynamique pour la determina­tion d'une machine thermique. Les dimensions principales qui en resultent sont :

- La geometrie de l'etage (angle de l'aube, diametres a l'entree et a la sortie, diffuseur) - Le nombre d'etages

- La conicite (rapport des volumes a l'entree et a la sortie)

- La vitesse de rotation.

Par suite des volumes relativement reduits, on utilise dans l'industrie du froid des turbomachines principalement radiales pour !es turbocompresseurs, parfois aussi axiales pour les turbines de detente. Comme le compresseur radial est de loin le plus souvent utilise, !es considerations qui suivent ont ere limitees a ce type de machine. Elles peuvent cependant s'appliquer aussi aux turbines radiales et aux turbomachines axiales.

2. DONNEES DE FONCTIONNEMENT D'UNE TURBOMACHINE

Le travail necessaire a la compression d'un gaz dans une turbomachine, nomme aussi travail au rotor, est donne par:

avec

a = _J�_s ·· · z1 · R · T1 [ n n

n 1 - 1 ] = µ · Zst · u 2 (10) Ks - 1 2

R ffi M

8314

M [J/Kg° K]

µ coefficient de travail, defini par (10) Zst nombre d'etages

u2 vitesse peripherique a la sortie de la roue M poids moleculaire

Le poids moleculaire apparait ici comme grandeur supplementaire en plus de Ks, z1 et m pour la determination du travail en jeu.

Pour dormer une idee de !'influence de la compressibilite dans les compresseurs, ii est commun d'introduire !'expression Ma = u2/a8 qui est egalement une mesure du nombre de Mach a l'entree pour un etage geometriquement donne. (La vitesse du son as est referee a l'etat a l'entree).

Pour la plupart des frigorigenes (gaz a poids moleculaire eleve), la vitesse du son as est si basse que c'est elle, et non pas des considerations de resistance, qui limite la vitesse periferique maximale admissible.

Avec Us max = Maadm · a s et u2 = Ma1 • as1

ii s' en suit : Ks1 · z1 · R · T1 a = µ · Zst · Ma12 • _____ M _____ _

OU rapporte a un kmol : A = µ . Zst . Ma12 • Ks1 . Z1 . R . T1

qui peut s'ecrire :

470

A Ys ------ = µ · Zst · Ma12 • R = ------Z1 · Ks1 • T1 z1 • Ks1 • T, 1/ is

( 1 1 )

(12)

( 13)

III-24

Ys est le travail isentropique rapporte a un kmol. Comme le nombre d'etages Zst ne peut etre qu'un nombre rond, les valeurs de µ et de Ma doivent etre choisies en conse­quence pour obtenir un certain travail de compression A.

Fig. 3· Variation du travail de compression isentropique rapporte a I kmol en fonction du taux de compression pour plusieurs frigorigenes.

y Dans la Fig. 3, !'expression K

s T

est reportee en fonction du rapport de Sr • Zr • 1

pression n pour plusieurs gaz differents, et comme on peut le remarquer, les differences entre les gaz sont minimes. D'ou la regle :

Lorsque la caracteristique d'etage est fixee, pour le meme nombre d'etages et un nombre de Mach identique, on obtient toujours approximativement le meme rapport de pression en comprimant n'importe quel gaz.

La caracteristique d'un etage de compresseur est donne par la relation qui lie le coef­ficient de travail 1� et le rendement polytropique 1Jpo1 en fonction du coefficient de

4 V debit o = ----D22 • U2

La Fig. 4 represente cette relation pour un certain type d'erage et la correspondance des valeurs optimales pour le dimensionnement de la machine.

Alors que la relation (10) donne une idee du nombre d'etages et au moyen du nombre de Mach (I I) egalement de la vitesse peripherique, le debit volumique s'ecoulant dans le compresseur est donne par :

Si au contraire le debit est fixe, on obtient le diametre de la roue par:

D2 = vJ�_'xra� v Ks-� Zr • ITT · Tr

A partir du debit volumique et de la relation generale

P1 • V1 = Gr · z1 • R · T1

(I4)

(IS)

471

IIl-24

0 f} µ. , 0 8 'l.pol 0.7 06 O.!S O.lf o. �

,.., -c�,._._.,.,_-::--..,1.iduw.J.{f .. , go• > /

�+ �--� \ \ cf . "" v D/ llz 0·2 a µ. . ";7l ,u , 0.1

0 25 0 ¥1 0..35 o.l<O

Fig. 4. Caracteristique de deux etages typiques de compresseurs radiaux.

On obtient apres une transformation convenable le debit massique

G = 2_ · o · D 2 • Ma • v- Ks . M--_ · P 4 2 1 Z1 • ffi . T1 1

La puissance necessaire a la compression d'un gaz devient done :

(16)

Vffi l /Ks1;-� • T1 N = G · a = -4- · o · D22 • Ma13 • µ · Zst • P1 • v M (17)

La vitesse de rotation du compresseur se determine a partir de (ll) et (15)

60 . u n = -

D2

t/min n . 2

(18)

Le rapport des volumes avant et apres compression est donne pour la transformation isentropique par:

Ks - I n Ks

n = n K �

et pour la transformation polytropique par:

V1

V2

n n

(19)

Ces relations seules ne suffisent naturellement pas a dimensionner un compresseur. Elles permettent cependant une approximation rapide des possibilites de fonctionne­ment d'un compresseur (conditions au point de fonctionnement normal) lorsque les coefficients µ, o et 1}po1 sont connus. Pour un type d'etage donne, !'experience a montre que ceux-ci peuvent etre admis approximativement constants pour n'importe quel gaz dans des domaines importants a condition cependant d'adapter la variation de la section de passage a la nouvelle allure de la transformation. C'est a dire qu'avec les memes

472

Ill-24

apports des rayons et les memes angles d'aubes, mais avec une section de passage adaptee, un etage pour la compression d'air donne environ les memes grandeurs carac­teristiques b, 11 et 1)po1 avec du R 1 1 .

Des relations derivees jusqu'a present resulte, lors du fonctionnement avec differents gaz, la dependance des donnees principales d'un turbocompresseur - rapport de pression respectivement travail de compression, nombre d'etages, diametre, vitesse de rotation, debit volumique respectivement debit massique, puissance absorbee, rapport des volu­mes - avec les proprietes physiques du fiuide.

Comme aucune restriction de quelque sorte n'a ete faite, ces relations sont valables d'une maniere generale. 11 en resulte cependant une nouvelle machine pour chaque application, car pour chaque gaz, l' «etage normal » correspondant est different.

3. SIMILITUDE D'ECOULEMENT LORS DE L'UTILISATION DE GAZ DIFFERENTS

Dans la pratique il est actuellement important de concevoir un type de compresseur, developpe et construit pour un cas particulier, de fac;on a lui donner des usages multiples. De cette fac;on, on peut d'une part economiser des frais supplementaires de developpe­ment et d'autre part diminuer les frais de finition par une production en grande serie d'un certain type.

La question suivante se pose par consequent : dans quelle mesure une machine con­struite pour comprimer un gaz donne peut-elle etre developpee egalement pour la compression d'autres gaz, le rendement et le comportement industrial devant rester aussi semblables que possible a ceux de la machine originale ?

La reponse a cette question est fournie par la thforie de la similitude, thforie d'apres laquelle !es ecoulements dans deux machines sont semblables lorsque toutes les grandeurs qui infiuencent l'ecoulement sont maintenues dans un rapport donne. Les expressions adimentionelles de ces grandeurs, nommees coefficients caracteristiques, doivent done etre egales dans les deux cas.

Cette condition ne doit pas etre maintenue seulement pour la machine dans son ensemble mais encore dans de ces elements si l'on veut une similitude rigoureuse.

Les caracteristiques d'un ecoulement dans une machine sont fixees, dans le cas sta­tionnaire et adiabatique, par les coefficients caracteristiques suivants :

Coefficient de debit b respectivement v Cm -- , ( b � v) u a

Coefficient de travail µ u 2 2

c Di (! Re -----17 Nombre de Reynolds

c Nombre de Mach Ma= as

Ks - 1 Exposant isentropique Ks respectivement -

K 8 -

et les coefficients caracteristiques geometriques

ou (c2 � a)

ou l/r represente la similitude gfometrique dans !es grands

et e/r represente la similitude gfometrique dans Jes petits

(j eux, rugosite, etc.)

Alors que b, µ, Re et Ma dependent des conditions de fonctionnement, les trois dernieres grandeurs doivent etre par principe egales pour une similitude rigoureuse.

11 s'en suit qu'une similitude parfaite ne peut pas etre maintenue lorsqu'une machine dimensionnee pour un certain gaz (Ks1) doit fonctionner avec un autre gaz (Ks2)

473

lII-24

Si cela devait avoir lieu entre certaines valeurs limites de Ks, on devra essayer de main­tenir au moins les conditions les plus importantes et prendre garde a ce que l'ecoulement ne soit pas ou seulement tres peu dependant des coefficients caracteristiques qui ne remplissent pas les conditions de similitude.

En principe il est done conseille, lors de l'emploi de gaz differents dans une machine donnee, de choisir ces gaz de telle fac;:on que la variation de l'exposant isentropique Ks soit tres restreinte. Malgre cette restriction, on remarque qu'une grande partie des fiuides frigorigenes peuvent etre rassembles dans un petit nombre de domaines.

Deux possibilites se presentent pour !'utilisation d'un compresseur dans un tel domaine de valeurs de Ks. : a) Les dimensions de la machine restent constantes ou geometriquement semblables

pour la compression des differents gaz. b) Les proportions radiales de la machine restent constantes alors que les rapports des

sections sont adaptes aux conditions regnantes dans chaque cas. Pour chacunes de ces deux possibilites il s'en suit certaines conditions de fonction­

nement parmi lesquelles la condition posee - egalite des coefficients caracteristiques -est remplie au mieux.

4. SIMILITUDE DE L'ECOULEMENT POUR DIFFERENTS FRIGORIGENES

On a rassemble dans le tableau 1 les valeurs resultant d'une transformation isentro­pique lorsque l'on utilise sans changement une machine de deux etages construite pour du R 1 1 pour d'autres frigorigenes.

D'apres les conditions de similitude, le rapport des volumes a l'entree et a la sortie est maintenu constant dans ce cas, alors que la vitesse peripherique varie comme la racine des chutes d'enthalpie. Lors de !'utilisation avec differents frigorigenes, tous Jes coefficients caracteristiques restants, comme le nombre de Reynolds (dont !'influence est negligee), le nombre de Mach et en particulier Ks, s'ensuivent automatiquement.

Comme on le voit, en plus des conditions fixees d'avance VA = VB et µA = µB, l'egalite des nombres de Mach est approximativement realisee. Mais comme d'autre part les valeurs de Ks varient entre 1,03 et 1, 21, la repartition des chutes d'enthalpie sur chaque etage est quelque peu differente. Ced est clairement montre premierement par le CH3Cl qui a cause de sa valeur de Ks > 1, 2, n'appartient plus a ce groupe de frigori­genes et pourrait etre integre dans des conditions semblables avec S02, CH2Cl2, C2H4 et N20 dans une machine construite en consequence, et deuxiemement par le C3H8 lorsqu'il est introduit a 0° C, c'est a dire dans des conditions oil son comportement est deja tres eloigne de celui d'un gaz ideal.

Pour tous les autres frigorigenes cites, les divergences sont cependant si petites qu'ils pourraient etre utilises dans les conditions ci-dessus dans une seule et meme machine.

Jusqu'a quel point ceci est-il interessant dans la pratique ? En climatisation, on a besoin de turbocompresseur dans le domaine compris entre

0° C, temperature d'evaporation, et 35° a 40°C, temperature de condensation. II resulte du tableau 1 que darts ce domaine et pour une machine donnee, les frigorigenes R 1 14, R 1 1, R 21 et R 142 peuvent etre utilises clans des conditions semblables. Cette meme machine ne travaille clans des conditions semblables avec le frigorigene R 1 13 que pour une temperature de condensation de 29°C, avec du R 12 et du R 22 qu'a 44,2° C, re­spectivement 47,2° C. Mais on obtiendrait egalement des conditions semblables pour ces deux derniers frigorigenes si la temperature d'evaporation etait d'environ -10° C au lieu de 0°C et la temperature de condensation d'environ 35° C.

Les considerations qui ont ere faites jusqu'ici se rapportaient a la similitude d' ecoule­ment clans une machine donne utilisant differents frigorigenes comme fiuide moteur. Mais du point de vue constructif egalement, il existe naturellement certaines exigences sur la similitude de la realisation exterieure (la similitude «interieure » etant admise !) pour la construction economique d'un type de compresseur.

Ainsi il est conseille, du point de vue de la realisation technique ( epaisseur des parois du bati, systeme d'etancheite par barrage de gaz ou par joints), de choisir les frigori­genes prevus pour un certain type de machine de fac;:on a ce qu'ils travaillent clans le

474

III-24

meme domaine de pression. D'autre part, la puissance requise pour une partie tournante donnee (roue, arbre, accouplement, reducteur) ne doit pas depasser une certaine valeur maximum.

D'apres la relation (17), la puissance de compression depend d'une part de l'expres-

K� - � - � ' . V---------- - -sion --- -1\.T-- - et d'autre part de la pression d entree P1. Les autres facteurs de

cette equation sont pratiquement constants pour des compresseurs semblables. On peut deduire du diagramme de la Fig. 5 les domaines de temperature dans lesquels un compresseur, normalement utilise dans le domaine de climatisation avec du R 1 1 ou R 21 et pour lequel la puissance maximum est ainsi donnee, peut etre utilise dans des conditions analogues par exemple pour du R 12, du R 22 ou du C3H8•

- 0 o,,, H I ;:;"' "' "' ' " � � � ..-(">.� 0 � "" t'...

w.. '�-1> '"�� "" � �� �� �_, � 0 ...... ::( .> v. ""' """ �� "'r-... " l'\. '\ � () !'..... "' lg ft !'-... r-.... ""' � !\.. ""' I\ """' I� � I � !'-.. ' �f'-. I'\ I\ \ � I\ I\ (5 I\.'\ I\

� 1-o- -... '" - - �'-� - s�- -' l\_ �LI - -r.<m.., I{;:; -It-l;ii;£.,.;,. '----- -

"' Ur orql?'o,,,., l \ � � \ \ I \ "' \ \ 0 "' '

8 c.. - · - --- ' 0 :>-.,,,_.,, \ I\

I... / � -· ' ....... ., - L\ t; Q""t C"o .. -· - \ ""- t-- I -

gi · �·�e.., . ,._ c.. \ • V' ....

Fig. 5. Variation d'un facteur representant la puissance en fonction de la temperature d'cvapo­ration pour plusieurs frigorigenes.

Si l'on considere de plus pres un tel domaine de temperatures, on s'aper�oit que pour des groupes de frigorigenes de meme nature tels que R 1 1 et R 12 ou R 21 et R 22, ce domaine est toujours situe a des temperatures reduites (rapportees a la temperature critique) a peu pres semblables.

D'ou la deduction suivante : Si un compresseur donne est exploite d'une maniere analogue avec differents frigori­

genes, il est particulierement avantageux de l'utiliser dans le domaine des memes tem­peratures reduites, ce qu'a deja fait remarquer Eiseman jun. [3].

Le tableau I donnait les conditions de fonctionnement, dans le cas d'une compression isentropique, pour lesquelles une machine donnee peut etre exploitee avec differents frigorigenes. Lors d'une compression polytropique, comme elle intervient dans une machine avec pertes, ces valeurs varient quelque peu, mais sans que le resultat d'en­semble en soit modifie.

Les valeurs ainsi trouvees pour chacun des frigorigenes sont chaque fois valables pour le point nominal (point de fonctionnement nominal) du compresseur. En general, la caracteristique du compresseur est mesuree pour un frigorigene au moins et reste valable sous une forme semblable pour !es autres frigorigenes.

475

Tableau 1 .

Vi/V2 = 3,265 = konst. V1/Vx = 1,825 Vx/V2 = 1,79

Frigorigene Dim. R 113 R 114 R 11 R 142 R 2 1 R 12 R 22

chute isentropiquc Y ' s KJ/kg 7,95 7,95 9,22 9,0 11,15 11,15 15,3 14,7 16,4 16,3 11,30 11,0 16,6 16,4 chute ie. totale Ys KJ/kg 15,90 18,22 22,3 30,0 32,7 22,3 33,0 vitesse peripherique u2 - 0,845 0,903 1,0 1,16 1,214 1,0 1,218 vitesse du son as m/sec 1 1 3 116,2 116,2 115,5 135 142 152 157,5 158 169,5 137,5 137,5 162,5 169 nomhre de Mach lla1, Ma2 1,01 0,985 1,05 1,06 1,0 0,95 1,03 0,995 1,038 0,968 0,98 0,98 1,01 0,972 K s - 1,058 1,0:) 1,104 1,07 1,138 1,050 1,098 pression a l'entree pl KN/m2 15,0 88 40,2 145 70,75 308,5 500 pression a la sortie p2 KN/m2 52,5 298,5 148,5 515 272,50 1060,0 1835 rapport de pression n - 3,5 3,39 3,7 3,56 3,85 3,44 3,67 temp6rature d'6vaporation to oc 0 0 0 0 0 0 0 temp6rature de condensation 'K oc 29 35,1 35 39,8 37,4 44,2 47,2 d6bit volumique VI m3/sec 0,845 0,903 1,0 1,16 1,214 1,0 1,218 debit massique GI - 0,43 2,52 1,0 3,18 1,61 7,15 10,45 travail isentropique N1 8 - 0,307 2,06 1,0 4,28 2,36 7,15 15,50 factem· de compressibilite z1�;2

- 0,99 0,955 0,962 0,885 0,985 0,965 0,945 0,885 0,98 0,945 0,93 0,83 0,896 0,795 poids moleculaire kg/kmol 187,39 170,93 137,37 100,5 102,92 120,92 86,48 constante du gaz R KJ/kg°K 0,0445 0,0487 0,0605 0,0828 0,6807 0,0688 0,0963

Ksl' Ks2 - 1,065 1,052 1,06 1,0 1 , 1 15 1,09 l,08 1,06 1,15 1,125 1,08 1,018 1,12 1,075

Tableau 1. (Continuation)

Frigorigene Dim. Methylchlorid Prop an Prop an Athan chute isentropique Y ' s KJ/kg 33,1 36,0 34,0 30,2 29,8 28,7 40,0 40,0 chute is. totale Ys KJ/kg 69,10 64,2 59,5 80,0 vitesse periphfaique 112 - 1,76 1,70 1,64 1,895 vitesse du son as m/sec 232 253 224,5 220 218 231 254 266 nomhre de Mach :.\fa1, Ma2 - 1,025 0,94 1.023 1,04 1,015 0,96 1,01 0,962 K s - 1,21 1,050 1,123 1,145 pression a l'entree pl KN/m2 256 468 70,7 551 pression a la sortie p2 KN/m2 1075 1620 268 2140 rapport de pression ;i; - 4,19 3,46 3,78 3,88 temperature d'Cvaporation 'o o c 0 0 -50 -50 temperature de condensation 'K o c 49,4 47,5 -17,1 -4,5 dChit volumique v m3/sec 1,76 1,70 1 ,64 1 ,895 dChit massique GI 4,33 7,08 1,145 7,80 travail isentropique Nis - 13,4 20,3 3,10 28 facteur de compre:ssihilitC Zl. Z2 - 0,938 0,895 0,885 0,778 0,975 0,95 0,88 0,796

<I' poids moleculaire M kg/kmol 50,49 44,09 44,09 30,07

ci1 constante du gaz R KJ/kg° K 0,165 0,189 0,189 0,277 ...... K sl' Ks2 1,245 1,175 1,10 1,0 1,15 1,115 1,19 1,10 \0 ...... i:--...... <!'

III-24

C'est pourquoi, pour un certain cas d'utilisation, il est possible, par variation de la vitesse de rotation, de choisir un autre point nominal comme point de fonctionnement nominal.

2..5 r/wda !? ff

t,, - 0 °C 0.5 t 0 1.S m3,�c!l1

Vt m a�i� ;:>Qr lo rno<:hincr Fig. 6. Caracteristique mesuree d'un compresseur radial a deux etages avec diffuseur a aubes

orientables pour le frigorigene R 1 r .

�.J § � � �

�o

.g J.S. ::; �

.3.0

o.s f 0 f 5 J1?3l;;--Voluma .aspire' par lo '7?0cl1v?rL

Fig. 7. Caracteristique mesuree du compresseur Fig. 6 mais pour le frigorigene R 21 . 477

IIl-24

Si une certaine machine, calculee pour du R 1 1 et les temperatures de 0°C et 35° C, doit etre exploitee dans le meme domaine avec du R 1 13 par exemple, la vitesse peri­pherique, et par consequent la vitesse de rotation (car D2 est constant), doit etre augmen-

F/wcta. · R It to • 0 °C

05 t o , 1. � m.;,l•ec Volume OsPtrq par 'o n?Ocnu7a

Fig. 8. Caracteristique mesuree d'un compresseur radial a un etage avec diffuseur a aubes orientables pour le frigorigene R r r .

478

I I

1 0 f.5 m vscir Vo/umr;: aspire parl'7 m<7::n1nrz

Fig. 9. Caracteristique mesuree du cornpresseur Fig. 8 mais pour le frigorigene R 2 r .

IIl-24

tee au delit de la valeur obtenue par la similitude; pour du R 21 par contre, la vitesse de rotation doit etre ramenee un peu au-dessous de la valeur correspondante.

5. RESULTATS MESURES SUR DES MACHINES GEOMETRIQUEMENT SEMBLABLES

Les caracteristiques mesurees d'un compresseur a deux etages fonctionnant avec du R 1 1 et du R 21 sont representees dans les Figs. 6 et 7. Les valeurs calculees de la compression polytropique sont donnees dans le tableau 2; elles concordent bien avec les valeurs mesurees.

Tableau 2. Compresseur a deux etages de dimensions constantes

Caracteristiqnes de fonctionnement pour Jes frigorigenes R I r et R 21

temperature a !'aspiration T , pression a !'aspiration r , volume specifique a !'aspiration v , pression a la sortie P, volume specifique a la sortie v, rapport de pression :JT, travail au rotor a vitesse de rotation 11

volume aspire v ,

R I I R 2 I

273 273 40,2 7o,75 0,405 0,3053 148,5 278 o,I275 0,096 3,7 3,92 28,6 4z,5 7200 8770

IOO I 2 I ,3 r , ro r,46

OK kN/m2 m3/kg kN/m2 m3/kg

kJ/kg tr/min % m3/sec

Les caracteristiques mesurees d'un compresseur a un etage fonctionnant avec du R 1 1 et du R 2 1 sont representees dans les Figs. 8 et 9 . Les valeurs calculees de la com­pression polytropique sont donnees dans le tableau 3 ; elles concordent bien avec les valeurs mesurees.

Tableau 3. Compresseur a un etage de dimensions constantes

Caract:eristiqnes de fonctionnement pour Jes frigorigenes R r r et R 2 1

temperature a !'aspiration T , pression a !'aspiration P , volume specifique a !'aspiration v , pression a la sortie P, rapport de pression :JT, nombre de Mach a la sortie Ma vitesse du son as vitesse de rotation n travail au rotor a volume aspire v ,

R I I R 2 I

273 273 40,2 70,75 0,405 O,J053 I48,5 265 3,7 3,75 roe IOO 139,5 165,5 ! 0 000 I I 850 28,6 40,r r,375 1,63

OK kN/m2 m3/kg kN/m2

% m/sec tr/min KJ/kg m3/sec

Dans les deux cas, les roues des compresseurs avaient des aubes radiales ((32 = 90°). Le reglage du debit etait obtenu au moyen d'un diffuseur a aubes orientables, qui donne une caracteristique semblable a celle obtenue par prerotation avant la roue.

BIBLIOGRAPHIE

r. Dzung, Thermostatische Zustandsanderung des trockenen und des nassen Dampf es. ZAMP 1 955, Heft 6.

2. Zweifel, Die Bestimmung des Zustandsverlaufs in Turbomaschinen mit Hilfe der Entropiezu· nahme. BBC-Mitteilungen r94r .

3. Handbnch der Kaltetechnik, Band IV Springer-Verlag I956, S. 98.

479

III-35

Etude experimentale sur les turbo-compresseurs frigorifiques mono­roue

The Experimental Study of Mono-Wheel Refrigerating Turbo-Compressors

J. PROUHET Chef du Service Recherches-Developpement, Ets Brissonneau et Lotz, 1, Rue Bellier, Nantes, France

SUMMARY. Experiments with refrigerating turbo-compressors necessitate a closed circuit testing plant in which given physical conditions can be imposed to ensure the stable use of gases. Some indications are given on the special device used.

Amongst others, tests were carried out on a backward-bladed impeller running at various Mach numbers and using vaned or vaneless diffusers.

It is shown that it is possible, under favourable conditions, to obtain better than normal air­conditioning requirements in single-stage compressor installations, using the usual mono-wheel compressor.

Les compresseurs frigorifiques mono-roue utilisant des gaz lourds, a faible vitesse du son ne peuvent etre valablement experimentes avec l'air atmospherique. Leur etude expe­rimentale pose done en tout premier lieu le probleme de !'installation d'essai. Celle-ci doit necessairement comporter un circuit ferme equipe des appareils usuels pour les mesu­res de pression et de debit, mais comprenant en outre des dispositifs permettant : d'une part, d'evacuer les calories dissipees, et d'autre part, de stabiliser a des valeurs determinees les conditions physiques, pression et temperature du fluide aspire. De plus, i1 est parti­culierement important de pouvoir modifier facilement la pression de refoulement avec un retablissement rapide de l'equilibre de marche qui est necessaire pour faire des mesu­res valables.

Differentes solutions peuvent etre adoptees, et ont d'ailleurs ete utilisees ; la premiere qui vient a l'esprit est d'utiliser un systeme frigorifique complet avec aspiration a l'evapo­rateur, et refoulement au condenseur. Toutefois, cette disposition ne permet pas de fixer facilement et rapidement les conditions physiques du gaz aspire lorsqu'on change de point de fonctionnement.

Une autre disposition consiste a utiliser un circuit aerodynamique ferme dans lequel la pression engendree est dissipee a travers une vanne de reglage, le fluide restant toujours a l'etat gazeux, et les calories etant evacuees a travers un echangeur de chaleur. Le refri­gerant de gaz requiert une surface importante ; la regulation des pressions a l'aspiration necessite un dispositif automatique special ; de plus, !'absence de phase liquide dans le cycle, rend incertaine !'evacuation tota�e de l'air en cours de marche. Une installation frigorifique auxiliaire est ici necessaire si on desire fonctionner aux temperatures habi­tuelles d'aspiration.

L'installation qui a ete utilisee differe assez substantiellement des dispositions evoquees ci-dessus (voir scheme Fig. 1 et photographie Fig. 2).

Elle comprend essentiellement : 1°/ - Un circuit aerodynamique ferme incluant le compresseur, une tuyauterie de re­

foulement avec un orifice de mesure de debit, une vanne papillon permettant de regler la pression de refoulement, et une tuyauterie de retour a !'aspiration dans laquelle est intercale un separateur de liquide dont le role apparaitra plus loin.

2°/ - Un systeme de refroidissement constitue par un condenseur branche sur une derivation du circuit ferme, en aval de la vanne de reglage. La temperature de con­densation regle ainsi la pression d'aspiration. Le condenseur etant en charge, le fluide refrigerant condense s'ecoule dans la bouteille collectrice du separateur,

481

III-35

1 .. Turbo pro1o11,.. ,.,.,,._...,. 2 .. P,,fu .,.., J>ru,-.:0,, •I k,,,,�n,N,•

ourrfoul,,,.,.n!-.5-EC:::':',,.•:::{'•::;� "' 1,_,.k. ·-' 4 • .,..,1/-C'• ·� p,.,� <* pms"": Q�//,. ,..,,.,,., • ..,00,..,, fl,,-, °'" ........,... .... hi,...

5 • vg,.,,. /X'/>'llon ti .. ln;ttltOn <M hyu1n nifo?,.,.nl 7 .. s� ... ,.,�.,,.. o¥ Aft"<:.¥

8 .. Aini '* �n .:I h""J>'._,.,,,.. a IUsl';,..,.,kM 9 .. Gnckn.n<1r

'(O .. Ahm-laho... M 1cw,,.,.,,.. fl'Od/J. /""O>-rngnJ clVn !J"'u,Pe fr''l'"'/.'f.,• a· o,..�ac

ff ... RoCt'n# �!J<lht.:ur f2 .. Ar...,.o. J'i']}rch0n � ,q,, /,�KJ. 13 .. a,..?. air fW"Y• <14,. &ul • .1/c. on:urndolri« n• f f�,. &.,/.,,//. 0CClN7>Uhhv"ctt n•/i

Fig. I

Turbo prololfP-€ monorouv Sehi'"a ck /'/nslolloh"on cli:ssa1I

Fig. 2 . Vue generale de !'installation d'essais.

d'ou une pompe annulaire a 5 etages le refoule sous pression dans deux pulverisa­teurs situes dans la conduite en amont du separateur, agissant ainsi comme un desurchauffeur par melange.

Le debit aspire par la pompe a la base du separateur peut atteindre deux a trois fois le debit condense, une notable portion du debit est done en exces, ce qui permet d'obtenir une desurchauffe complete si on le desire.

La variation de l'exces du fluide pulverise permet de regler la surchauffe du fluide aspire.

3° / - Un systeme frigorifique a ammoniac qui fournit de la saumure au condenseur a la temperature correspondant a la pression d'aspiration necessaire.

4° / - Un groupe de purge qui est branche sur le condenseur et assure la vidange prealable de !'installation ainsi que !'elimination de !'air et de l'eau qui pourraient pene­trer dans !'installation pendant la marche.

482

III-35

Cette installation permet d'avoir un fonctionnement aussi souple que possible, car !'interaction des variations de reglage aerodynamiques sur les conditions d'aspiration est faible. La pression d'entree est pratiquement fixee par la temperature de condensation.

Les differents points de fonctionnement etaient obtenus par variation de position de la vanne de reglage ; la stabilisation des regimes avant les mesures etait assez rapide, ce qui est particulierement precieux etant donne les nombreux points de mesure necessites par une etude complete.

Les pressions etaient mesurees au moyen de manometres a mercure. Le manometre differentiel de !'orifice de mesure de debit etait a tube incline. Afin d'eviter les conden­sations dans les tubes de manometre, en particulier au refoulement, un dispositif de re­chauffage par radiateurs a infra-rouge etait utilise.

Le fluide refrigerant utilise etait le R 1 1, ce qui reduisait la puissance necessaire au mo­teur et a !'installation frigorifique auxiliaire, mais etant donne la proche parente des divers refrigerants, en ce qui concerne la vitesse du son, et les lois de compressibilite, les resultats sont utilisables pour la plupart des refrigerants fluores, entre autres le R 12 pour lequel le prototype essaye etait plus specialement com;:u.

Les essais realises visaient a la mise au point d'un turbo-compresseur a reducteur incorpore dont on desirait explorer les possibilites d'utilisation a des rapports de compres­sion correspondant en premier lieu au conditionnement d'air mais aussi a des valeurs superieures aussi elevees que possible. Ce turbo-compresseur est presente Fig. 2.

Deux modeles de roues a ailes radiales et un modele de roue a ailes en arriere ont ete essayes en combinaison avec differents diffusseurs lisses ou ailetes pour des calages variab­les des ailettes de pregiration et ceci a des nombres de Mach de plus en plus eleves.

Pour ne pas developper exagerement cet expose nous ne relaterons que les essais les plus significatifs. En particulier, les roues a niles radiales ayant presente des caracteristiques moins souples du point de vue de !'utilisation frigorifique, nous ne traiterons que des essais de la roue a ailes en arriere.

Les courbes caracteristiques Figs. 3, 4, 5 et 6 resument les resultats des etapes essen­tielles de !'experimentation.

Rt1eoy de6 caroeNdsliou•s ci bas$c vikr.tt o'dne roue d q,/"s en arriirr arrc qf(/useur lisse

!!.. = (,6$

Fig. 3

Elles presentent les caracteristiques exprimees en coefficients sans dimensions de Rateau1) :

H Coefficient manometrique µ = u2/g

Qo Coefficient de debit <50 = u r 2

1) voir detail des notations a la fin du texte.

483

III-35

'"

'"

..

�e·set:1u d•1 CC¥!"0Clerisli9_uE'4 ,. o" ?tancle Vth.s;q:,. dbne n:Jur o- ah'rs rn arrifire ovec �ffenW' lisS#; � = f.88

ReSecrtJ des carocleiYilir.rues , £...,y.l'CTnc/- t'ilesse d Ynr /"'otH d al/rs m grrieCe qac ¢/fgseu,. qt/rN

g = �.88 Oo

8r'stnN des curaclefrish'rp!H. a· 6asse Yik£sc,. � coue c:l al/es en drr/ei-e gt're o'1f/.ys.tul' allcJ.c·

Carod.r;.rh� .. , � .: 'f.Sfl

J��h"· '�\

µ

"

...

Fig. 4

µ

Fig. 5

,,

Fig. 6

Les nombres de Mach de fonctionnement sont caracterises par le parametre u/ao. Afin de mieux faire apparaitre !'evolution des performances en fonction du nombre de

Mach, les rendements ont ete rapportes a une valeur de reference 100%, qui est le rende­ment au point a la puissance nominale du compresseur muni du diffuseur lisse normal

u pour = 1,55, a o

484

III-35

c'est-a-dire a un regime de rotation correspondant sensiblement au conditionnement d'air en R 12.

Dans les graphiques 3, 4, 5, 6, les courbes en pointilles sont les caracteristiques µ = f ( '50) pour differentes positions des ailettes de pregiration, les courbes en trait plein sont des lignes d'equirendements (rendements rapportes au rendement de reference indique ci-dessus). La zone hachuree marque la limite de pompage. En outre, une ligne en trait plein figure a titre indicatif le lieu des points de fonctionnement a charge partielle d'une installation frigorifique-type equipee du compresseur.

La Fig. 3 presente les caracteristiques du compresseur equipe d'un diffuseur lisse, u fonctionnant a - = 1,55. a o Le rapport de compression au point nominal est de 3,3, soit un peu plus que celui du

regime 0° + 40° en R 12. La limite de pompage est tres plate et permet une grande souplesse de reglage vers les bas debits.

La Fig. 4 presente les caracteristiques du compresseur equipe du meme diffuseur u mais fonctionnant a un nombre de Mach nettement plus eleve : - = 1,88. a o Le rapport de compression au point optimum nominal est de 5,16 (environ -15/

+40°C en R 12).

La courbe de pompage est tres favorable a un fonctionnement jusqu'a une tres faible puissance. Par contre, on note une importante diminution du rendement. Des essais complementaires, avec un diffuseur lisse plus etroit ont montre qu'il n'etait guere possible d'ameliorer les performances avec ce genre de diffuseur.

Des essais ont done ete entrepris avec un diffuseur ailete. La Fig. 5 montre qu'avec

ce type de diffuseur un fonctionnement satisfaisant a grande vitesse Cua = 1,88) peut etre

obtenu. Au point de fonctionnement optimum, le rapport de compression atteint ainsi 7,7.

La courbe de pompage est un peu plus tombante qu'avec le diffuseur lisse, mais a condition de fixer le fonctionnement du point maxi nominal tres legerement au-dessous de !'optimum (rapport de compression 7,2 soit -22/+40°C ou -30/ +30 ° C en R 12).

La souplesse de fonctionnement a puissance reduite est tres convenable.

Le diffuseur ailete utilise comportait des ailettes dont le calage pouvait etre facilement modifie, ceci nous a permis de verifier facilement le fonctionnement a vitesse moderee (auo = 1,55) , pour un reglage d'ailes adapte a un debit nominal de l'ordre de la moitie

de celui realise avec diffuseur lisse.

La Fig. 6 montre que dans ces conditions le rendement est a peine inferieur a celui ob­tenu avec diffuseur lisse et que la souplesse de marche a puissance reduite est encore tres bonne.

Notre etude, dont cet expose obligatoirement resume n'a relate que les traits les plus saillants, a ete loin d'epuiser le sujet ce qui autorise a penser que des caracteristiques en­core meilleures pourraient etre obtenues.

Elle a permis de montrer que pour le conditionnement d'air le diffuseur lisse convenait tres bien, mais qu'un diffuseur ailete permettait d'adapter la meme roue a des groupes de puissance frigorifique beaucoup plus faible (jusqu'a 50% environ) et ceci avec une baisse de rendement tres minime. Ceci permet de limiter le nombre de modeles de roue dans une serie.

Mais surtout elle a mis en lumiere la possibilite d'equiper avec des compresseurs mono­roue des installations frigorifiques presentant des ecarts de temperature depassant large­ment ceux du conditionnement d'air.

Le perfectionnement des traces permettra certainement d'ameliorer encore ces resultats a l'avenir, ce qui dans de nombreuses applications amenera a utiliser, a la place de com­presseurs polyetages, des appareils mono-etages plus simples et moins encombrants.

485

III-35

NOTATIONS UTILISEES

H Hauteur isentropique u Vitesse tangentielle a la peripherie de la roue Qo Debit volume aspire r rayon exterieur de la roue (correspond a la vitesse « u ») ao vitesse du son dans le fluide a !'aspiration

Po pression d'aspiration

Pr pression de refoulement.

DISCUSSION

G. Lorentzen, Norway : Why was condensation carried out on the low pressure side of the system, i. e. after the hand regulation valve ?

J. Prouhet, France : We really wanted the full compressor power, i. e., maintain the speed of sound over the temperature. In this case the speed of sound does not require calculation for each reading.

A. Bedue, France : What was the basis of calculation for the opening of the expansion valve ?

J. Prouhet, France : We were trying to retain a certain flexibility, but the thermostatic expansion valve value was fixed at LI = 0.192, and then we were keeping the superheat at suction to 5 to 10° C.

H. W. Fischer, U. K. : I suggest that 5 ° C is rather a low temperature to assume that there are no entrained droplets.

J. Prouhet, France : We tried to keep the superheat temperature as near 10° C as possible.

Benke, Germany : It was common practice to superheat the suction vapour before entry into reciprocating compressors for Freon refrigerants. The reason was, of course, to reduce to a minimum the number of entrained droplets which will be carried into the cylinders. With turbo compressors, however, it is usual to work with low superheat. It would be sensible to assume, therefore, that the same phenomenon applies as for piston compressors. Can the speaker therefore give any basis for suggesting whether or not the centrifugal compressors act in the same way as piston compressors.

J. Prouhet, France : This question has not fully been investigated by me, but I feel that there would not be a great deal of difference in the compressor efficiency. Of course, it would be very interesting to get exact data on this matter of droplets.

A. Neuenschwander, France : Droplets will, of course, inevitably reduce the efficiency. As far as suction superheat was concerned this, of course, was not part of the investi­gation which the author set out to make.

G. Lorentzen, Norway : In reciprocating compressors we have an entirely different effect than in turbo compressors and the penalty may therefore not be as high.

J. Prouhet, France : I completely agree with those last two contributions.

486

IIl-9

Design of Mixed Flow Impellers Operating at High Mach Numbers for Industrial Centrifugal Refrigeration Equipment

Conception des impulseurs d'ecoulement mixtes fonctionnant a des nombres de Mach eleves pour les installations frigorifiques centrifuges industrielles

I ING. VACLAV DANEK CKD, Prague, Czechoslovakia

SOMMAIRE. Les avantages incontestables des machines frigorijiques centrifuges ont recemment entraine un developpement considerable de ces machines, en particuler dans le domaine du conditionnement d'air. On a effectue des recherches pour augmenter le taux de compression dans un impulseur jusqu' a e = 3,5 - 4,5, f acilitant ainsi la conception des petits compresseurs centrifuges a un etage. En augmentant le taux de compression a un etage, il n' est pas necessaire de resoudre la question de la force de l'impulseur par rapport a la f aible vitesse du son de la plupart des frigorigenes (Freons) . C'est cependant une question d'infiuence negative de la vitesse supersonique locale sur l'ecoulement dans l'impulseur et dans le diffu­seur. Les nombres de Mach les plus eleves apparaissent du ct5te de l'entree de l'impulseur, c' est-a-dire dans l'inducteur. Des recherches ont ete entreprises sur l'infiuence des differentes formes d'inducteurs, dont on presente quelques resultats, obtenus sur des grilles fixes. On a trouve que les courbures paraboliques et lenniscatiques donnaient - en comparaison des cour­bes circulaires et elliptiques - de f aibles valeurs d' etranglement mais presentaient une courbe de rendement plus favorable en fonction de l' angle d' attaque a des nombres de Mach el eves. Ce point de vue est tres important pour les compresseurs frigorifiques centrifuges et determine la serie economique de regimes stables a un etage de compression.

Conventional construction of turbo-compressors for refrigeration equipment was characterized by impellers of backward bent blades having the outlet angle {32 smaller than 90°. Excellent results achieved in the field of radial turbo-compressors for gas turbines with mixed flow impellers initiated the introduction of these impellers for sta­tionary machines in general, and particularly for refrigerant turbo-compressors. As a matter of fact the required compression can be achieved by a minimum number of im­peller wheels together with all the convenient consequences resulting from this for both construction and operation of the machine.

The admissible value of Mach number means a limit to the maximum compression ratio attainable in one stage. Thus it is possible to achieve a compression ratio as high as 5 in one stage, when using R 1 1 or R 12 refrigerants which corresponds to the cir­cumferential velocity of 250 m/sec., whereby the highest values of Mach numbers are reached at the impeller inlet, in the so-called inducer, and at the impeller outlet. The absolute velocity at the impeller outlet achieves values neighbouring to the speed of sound, whereby shock waves originate at the leading edge of the diffuser vanes. This problem can be successfully solved by reducing the absolute speed at the impeller outlet under the value of critical speed in vaneless diffuser. Vaneless radial diffuser being so the only example of transit from supersonic speed to subsonic speed without generation of shock waves. In the vaneless diffuser it is possible to reduce the velocity and moreover to achieve an equalization of the constant velocity profile which reduces the mixing losses. Vaneless diffuser placed between the impeller and vane diffuser considerably moderates the noisy running of the machine. These circumstances lead to the fact that the most convenient solution would be to design a stage consisting of prewhirl vanes placed before the impeller wheel, mixed flow impeller, vaneless and vane diffuser and of a spiral.

When designing a mixed flow impeller it is necessary to adopt the production tech­nology to the requirements of the flow, i. e. to solve such a construction as to reach fluency between the admission part, inducer and the radial part of the impeller. The

487

III-9

impeller blade is, therefore, shaped as part a of helicoidal surface limited by the wheel and the shroud. However, it is necessary to adapt the axial depth of the impeller, blade curvation on the cylinder surfaces and meridional shape of impeller channels.

When appreciating the importance of inducer for determination of axial depth we have to analyse the energy delivered to the medium in the inducer and in the radial part. The flow in the impeller is, therefore, divided in the axial direction, i. e. along the cy­lindric surface, where the inducer gives the compressed medium angular acceleration only. In the radial part, where the medium has already got a constant angular velocity an increased energy of the passing gas is achieved by its motion in the radial direction along the radius from rs to r2, i. e. by means of Coriolis acceleration.

The ratio of energy delivered by the inducer to the total of energy delivered by the impeller wheel is defined by expression

where

.A. .. r.

488

G � f dG · I'z1

Ez 0 ------Ek G �! dG · I'k1

0

G quantity passing through the impeller

I'z1 inducer circulation over the i stream line

I'z1 = z � c1 · cos OG1 ds = z2 :n: r1 (cz1 cos OG1 = cu cos !Xu)

.. .L

$

U s

Fig. I . Meridional cross-section o f an impeller

and I'k1 = impeller circulation on stream line

I'ki = z 2:n: (r21 Cuzi - cu cos OGu)

III-9

If we concentrate the whole flow through the impeller on the medium stream line marked by s in Fig.I, we obtain a simplified expression :

Ez I'z rs (Cz s COS OG z s-Cls COS OG 1 s) Ek = I'k = r2 . C u2 s - rs . C1s cos OG1s

When considering an infinite number of blades :

C z s • COS OG z s = U s

Hence it follows

Ez I'z Ek - I'k

(rs/r2) ' - (rs/r2) • (C u s/u2) 1 - (rs/r2) · (c us/u2)

The ratio of the energies delivered depends therefore on the construction parameter rs/r2 and on the magnitude of the prewhirl expressed by the ratio c us/u2•

In Fig. 2 the ratio of energy I' z/ I'k in dependence on the rs/r2 value for different magnitudes of the prewhirl c us/ c u2 generated by the blade leading before the impeller is plotted. The magnitude of the energy pertaining to the inducer may be reduced by the positive value of the prewhirl. In this way the total energy delivered by the impeller

-0,2t---.._--�---+---_,_---+---+----I----< · O,f 0,2 0,3 O.• qs ,6 0,1 48 0,9

'i/'.l Fig. 2 . Ratio of energy delivered in the inducer to the total energy transmitted by the impeller

wheel.

is reduced. In our case the inducer is designed for a no-prewhirl impeller inlet. Fig. 3 shows the results of measurements carried out to verify the inducer work at static cas­cades in a high-speed air tunnel where a channel corresponding to the cylindric cross­section of the impeller was created. This cross-section passes through the maximum inducer radius i. e. through point A (see Fig. 1). This means a considerable simplifica­tion of the problems consisting in substitution of three dimensional streaming by two dimensional streaming. However, the task of the research was to rate the optimum angle

489

III-9

1,1 HW. ---+--�---+----+---+---+---+------' H­M*

¥'--+----t-----+----¥-,,,C-----I--

-·

......... =-�

lnt_+-------1

-i .,,, +ffl

Fig. 3. Results of inducer measurements as static cascade ; circular shape marked +--- +, lemniscate shape marked o-----o.

of attack even at different curvations, the maximum of the Mach number Mma x at which the channel is "choked" and the critical Mach number. Furthermor, we mention here the maximum value of the Mach number M*, whereby compression still takes place, i. e. value Llp has a negative character

Llp =

where

Psi - Ps2 I 2 (!1 W12

static pressure before cascade

static pressure after cascade

specific mass of the gas before cascade

gas speed before cascade

f (M)

These measurements have been carried out for circular curvature (dash-and-dash lines) and for lemniscate curvature (continuous lines) of the inducer.

In Fig. 4 different curvatures of inducers are demonstrated. The most simple is the circular curvature rendering the shortest channel. The longest channel is rendered by the lemniscate curvature, however, enabling the design of an inducer of small surface enlarging and extensive change of direction in the radial part. The principal difference between the results achieved with the circular and with the lemniscate shaped inducer consists in displacing of optimum angle of attack. In circular inducers the maximum value of the critical Mach number Merit ranges between -5° and 0° angle of attack, whereas in lemniscate inducers even at + 10° the maximum value Merit has not been achieved. Similar results have been obtained when ascertaining the maximum Mach number Mma x and values M*. Whilst the optimum value of angle of attack is at + 5 ° t o +7° for circular curvatures, for lemniscate curvatures both curves have a n increasing tendency and even at + 10° no maximum value has been reached. The displacement

490

1-----� z

a b c d

111-9

Fig. 4. Scheme of a helical line forming an impeller blade and the course of cylindric sections for the individual curvatures :

a) parabolic b) circular c) cubic d) lemniscate

of the maximum Mach number Mmax of lemniscate curvatures towards the angle of attack of + 10° or even more leads in case of unchanged dimensions to lowering of the maximum choke point by about 25 % in comparison with impellers provided with circular shaped inducer. Similar results of impeller research have been achieved in NASA laboratories where circular, eliptic and parabolic curvatures were tested, whereby para-

Fig. 5. Tested impeller without shroud

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III-9

bolic curvature rendered the widest operating range of the characteristic. The same im­portance for achieving a good efficiency of the impeller has also meridional curvature of the channel. This was carried out according to a theoretical treatise by Dr.-Ing. Hasselgruber of Hannover and according to an approximative solution designed by members of NASA laboratory staff Smith and Hamrick. In the theoretical solution there is a close dependence between the blade curvation in the cylindric cross-section and the channel curvatures in the meridional plane. The mentioned solutions, however, do not consider the complexity of the effective streaming with respect to the viscosity and clearance between the impeller and the turbo-compressor body. In Figs. 5 and 6 we see photographs of two impellers constructed with respect to some notions mentioned in this report. In Fig. 5 there is an impeller channel without shroud. Comparatively small curvation of the channel and continuous transition of both curvation and impeller surface are apparent here. Fig. 6 shows a shrouded impeller at the radial part.

Fig. 6. Tested impeller with shroud

When drawing conclusions the experimental research carried out for high Mach num­bers under application of R 12 refrigerant and partly also production processes affecting in lesser degree the shape of the impeller surface have been of significant importance. The tested group of mixed flow impellers was generally formed by impellers with or without shroud having different curve of the impeller surface expressed by the ratio

where

F2 Fi

= 0.74 - 0.34

Fi impeller inlet cross-section

F 2 impeller outlet cross-section

It appears already to-day that the most convenient ratio drops as the compression increases. It has been, however, affirmed that the outlet width does not determine the maximum choke point of the impeller wheel. This is determined by the inlet cross­section, blade curvation and their thickness only.

By measuring of characteristics it has been found that the maximum choke point does not grow proportionally to the increasing circumferential velocity. This was why cha­racteristics were analysed with respect to the angle of attack causing choking seize to the admission channels determining the maximum choke point at a certain circumferential

492

IIl-9

velocity. To this effect measurements were compared in which the maximum choke point was reached, i. e. with vaneless diffuser. By measurements with air the following results were obtained :

M U2 Mw1 aot

0.85 0.6325 + 1° 20' 1.05 0.752 +5° 20' 1 . 1 1 0.790 +60 1 .165 0.823 +6° 50'

Similar results were obtained also when measuring at high Mach numbers with R 12 refrigerant again with vaneless diffuser to ensure the maximum choke point of the im­peller when the inducer choke seize takes place.

- U2 M = -aot

1.13 1.29 1.45

Mw1

0.815 +4° 35' 0.915 +6° 10' 1.013 +7° 50'

From the mentioned comparison of the results carried out under application of the static cascade with a certain maximum choke point of the impellers appears a very good coincidence showing substantiation for measuring of the static cascades.

The last task was to verify a good efficiency and stability of flow at small quantities, i. e. at large positive angles of attack in the range over 10° for this type of impellers with lemniscate or parabolic curvature. To this effect, when measuring the system of charac­teristics at different setting angles of diffuser blades, the impeller efficiency was ascer­tained and the curve of efficiency obtained is plotted in Fig. 7. The impeller efficiency suddenly drops at maximum choke point, whilst the optimum of the impeller efficiency is reached at 60% Qs ma x it makes still more than 80%. From these results one can presume that at positive angles of attack of + 10° and higher no stall in such an extent takes place, as to cause a considerable drop of the total efficiency in a stage.

When evaluating the achieved results, we may say that a deep inducer with curvatures rendering a comparatively small admission load enables to reach a wide range regulation under application of adjustable diffuser vanes. Comparatively thin and slightly curved profiles ensure good aerodynamic properties in the range of high Mach numbers.

raq l_. ?mp -·'Ir-- . --.... + � / _.,

� 0,8

1 41 f � 3 4 5 6 1 8 9 *!)

� � Fig. 7 . Curve of impeller efficiency in dependence on the intake quantity at different setting

angles of diffuser vanes.

v for a3 = 5°

t for as = 1 0°

0 for aa = 1 5°

A for as = 20° 0 for aa = 2 5°

493

III-32

Operative Properties of Mixed Flow Impellers of High Mach Numbers Destined for Industrial Turbo-Compressor Refrigeration Equipment

Proprietes en fonctionnement des impulseurs d'ecoulement mixte a nombre de Mach eleve destines aux dispositifs frigorifiques a turbo-compresseur industriels

J. MELICHAR CKD, Prague, Czecho-Slovakia

Des recherches sur les impulseurs d'ecoulement mixte avec un angle de sortie {32 = 90° ont montre que l'on pouvait eliminer l'influence negative de la vitesse peripherique croissante par une conception appropriee.

SOMMA/RE. On a obtenu une grande serie de regimes stables par un dijfuseur ou des directrices ajustables. Pour le reglage de la puissance frigorifique, les caracteristiques de l' evaporateur et du condenseur sont determinees et la puissance frigorifique est reglee par le debit du frigorigene a vitesse constante du compresseur.

Des recherches ont done ete entreprises sur !'influence de l'angle d'ajustement des aubes et des directrices du diffuseur sur le type des caracteristiques a vitesse constante de la machine. On donne une analyse des mesures indiquant la limite de la serie de regimes stables par les conditions du diffuseur et la possibilite de variation economique du rendement frigorifique par des arbres de diffusion reg/ables. Les directrices d' entree font varier la quantite de travail transmis a diff erentes intensites du tourbillon d' entree, reduisant ainsi considerablement le taux de compression a des debits minimaux.

Pour eviter de reflux a un rendement frigorifique minimal, il est necessaire de choisir un niveau de compression nettement plus eleve au moment de la conception.

L'analyse donnee indique l'avantage economique des arbres de diffuseur reg/ables pour le reglage de la puissance des machines frigorifiques et l'utilite des directrices d' entree reglables pour les circuits de circulation.

The modern trend of development of turbo-compressors for refrigeration equipment tends to increase the circumferential velocity of impellers. The result of these efforts is to achieve the maximum compression in one stage. For this reason the development in our laboratories was aimed at designing mixed flow impellers operating at high Mach num­bers. A systematical research has shown the maximum compression value of 4.5 may be reached under application ofR 12 refrigerant without lowering the efficiency. The results are apparent from Fig. 1 . When applying the results achieved to the refrigeration equip­ment, it was necessary to solve the problem of the narrow stable range of characteristic, in order to comply with the condition of economic regulation of the cooling capacity in the widest possible range. The requirement to change the cooling capacity in a wide range calls for a secured stabilization of the turbo-compressor operation within a wide range of the intake volume, whereby the value of the compression achieved has to correspond to the characteristic of the evaporator and condenser. That was why the systematic re­search was concentrated on the study of reasons causing the narrowing of stable range characteristic when increasing the Mach number. If the quantity delivered differs from the nominal one, the angle of attack changes both at the impeller inlet as well as the diffu­ser inlet. When considering both the impeller and the diffuser as a cascade we see that the condition of a perfect flow of the elements mentioned at high Mach number in such a wide range of the angle of attack, can not be fulfilled.

The effected research of preset diffuser vanes has shown that with the increasing angle of diffuser vanes a3 the passing quantity grows until the maximum choke point of the impeller has been reached. The choke point of the impeller may be regulated by setting of the prewhirl vanes at the impeller inlet. For this reason the influence of the prewhirl

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vanes with vaneless diffuser was investigated which, however, showed that the reached compression value is lower on account of lower efficiency 'f/iad· Here the influence of change in direction of the absolute speed at regulation by inlet prewhirl vanes on the deliv­ered work has become evident in accordance with the Eulerian equation, which un­favourably influences the value of the compression reached in dependence on the quantity delivered at constant speed.

In order to judge the influence on the prewhirl in vanes in conjunction with vaned diffuser, measurements at small quantities, i. e. diffuser vanes are adjusted at OG3 = 10° and at larger quantities OG3 = 20° were carried out. It became evident that the influence of prewhirl vanes, when diffuser vanes are adjusted at OG3 = 10° is substantially lower than at OG3 = 20? when the diffuser vanes are fully opened. Here a wide regulation range was achieved, however, connected with a considerable decrease of the compression value not allowing the operation in refrigeration equipments. This regulation may be suitable for circulation systems, where the necessary compression value drops with the decreasing quantity.

From the mentioned results it is apparent that the change of the cooling capacity in a wide range may be achieved by means of adjustable diffuser vanes. The working regime at different capacities is given by the crossing point of the characteristic of both evaporator and condenser with the individual characteristics of the turbo-compressor. Should a cooling equipment operate economically even at varying values of evaporating and con­densing temperature, i. e. at different compression it is advantageous to combine the regulation by both adjustable diffuser and prewhirl vanes.

496

I I l £ -·� " •-1---- � . ·- --+---+--i·....-r--r-------1 I "'/,; 0• .......... .

·'-+--f-� : k,�-+--+--+-'\-+-t-----1 l �·--+---1---1�-c-l�--1---1-1---11----1 'i'..� ' i 0'r--.� , ��--l---l---l---l��---1-1---1---1 ! �·-t-�--l·iwc=:-1---1---11----1-Hl ----j

2/l--+---+-..--..._._:c,,.__�,_--f_-+--l·f--1---1 -•-+--+--+-+--+--+--+--I-+---1 I

� ID ,,,� [M-Y,.]

._ -

1 1---+-+--+----t---+-t-� t Fig. r. Results of research on mixed flow impellers

Misce l l a n eous Q u est i o n s Quest i o n s d i verses

III-5

Conditions of Cavitation in Liquid Pumps for Refrigerant Recircu­lation

Conditions de la caviation dans !es pompes a liquide pour le recirculation du frigorigene

Prof. Dr. GUSTAV LORENTZEN Institutt for Kjoleteknikk, Norges Tekniske Hogskole, Trondheim, Norway

SOMMAIRE. Les systernes de recirculation a pornpe a basse pression se sont repandus de plus en plus ces dernieres annees, en raison des avantages qu'ils presentent dans le conception de la tuyauterie des grandes installations avec evaporateurs serni-noyes. Ce developpernent se poursuit malgre un certain nombre de pannes dues a la cavitation. Celle-ci peut entrainer !'interruption Je la circulation et, dans certains cas, deteriorer la pornpe. Dans la pratique, on recommande couramment d'utiliser une colonne de liquide assez evee du cote de !'aspiration, en combinaison avec des tuyaux prevues assez largement. On obtiendra naturellement ainsi une securite totale dans des conditions d' exploitation stables, lorsque la temperature d' evaporation est constante.

En realite, les problemes se posent dans des conditions non stationnaires. Les considerations theoriques et l' experience pratique indiquent que la cavitation est tout-a-! ait susceptible de se produire pendant !es perioaes d' abaissement rapide de la temperature ou lorsqu' on declenche les pompes. L'analyse montre qu'il existe un diametre optimal des tuyaux d'aspiration qui donne la maximum de securite pour une mise en temperature rapide. Des tuyaux d'aspiration trop larges reduisent reellement le gradient maximal de temperature pour un fonctionnement sans cavitation. Il est aussi peu avantageux d' elever le niveau de liquide statique au-dessus d'une certaine valeur. Contrairement a !'opinion courante, le risque de cavitation est plus grand que des temperatures d'evaporation elevees qu'a des temperatures basses.

Un grand nombre d' essais ont ete eff ectues pour completer la theorie et donner des indications pratiques en vue de la conception optimale des tuyaux d'aspiration.

INTRODUCTION

Low pressure pump recirculation has become increasingly popular in refrigeration plants in recent years. This is the case both with ammonia and the halocarbon refrigerants. Especially in large installations with many and dispersed evaporators, it frequently per­mits simplification of the piping system without sacrifice of evaporator efficiency [2] . It also gives a smaller charge of refrigerant and more simple defrosting than the conven­tional flooded system, based on self-circulation. Oil return from the evaporators, when oil-solving refrigerants are used, is easily arranged, without having to resort to the use of rather inefficient dry type evaporators [ l ] .

A number of different pump designs are used in practice. During the early days posi­tive displacement pumps, frequently of the gear type, were the most common solution. They were rather insensitive to cavitation, but were easily damaged by impurities. Later centrifugal type pumps became more popular, and they turned out to have a higher aver­age life. However, the shaft seal is always a weak spot for a liquid pump, and much more so than for a compressor. Adequate lubrication is more difficult to achieve, and a liquid leak is of course much more serious than a gas leak. Altogether there are some doubts as to whether it is justifiable to use open type liquid pumps at all in fully automatic installa­tions operating without continuous attendence.

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In an effort to avoid mechanical pumps and seals, injectors and pulsometer type circulators are frequently applied [3]. However, modern trends are towards hermetically sealed pumps. Designs with the motor enclosed in the refrigerant atmosphere were devel­oped for halocarbons at an early stage, but have only gained limited application. The same is the case with pumps using a magnetic clutch for power transmission into the sealed unit. It was only at the advent of the "canned motor" type that the hermetic pumps became really popular.

Fig. r. Section of typical "canned motor" pump, somewhat simplified.

A common design of "canned motor" pump is shown, somewhat schematically, in Fig. 1. The rotor as well as the stator are separated from the refrigerant by a thin walled tube of nonmagnetic material, usually austenitic stainless steel. The outside casing and the electric power lead-ins are designed to withstand pressure, so that any risk to the surroundings in case of breakage of the stator "canning tube" is eliminated. The two bearings are lubri­cated by a stream of liquid refrigerant, which also serves to remove the motor heat. This liquid is taken from the pump discharge and returned to the suction of the last stage, in case of a multistage pump.

A large number of "canned motor" pumps have been installed in refrigerant systems during recent years. Although experience is generally good, a considerable number of failures due to cavitation have been reported. Serious cavitation stops delivery from the pump, leading to interruption of the liquid stream for lubrication and motor cooling, and bearing failure is the inevitable result when the pump is not immediately stopped. The use of bearing materials with good dry-running properties has only partially solved this prob­lem. The real remedy is to avoid cavitation.

The general recommendation from the pump manufacturers for this purpose is to use large suction line diameters without restrictions, short pipe connection and a fairly large static head from the pump to the liquid level. In spite of these precautions cavitation and resulting pump damage have occurred in many cases. In an effort to explain these seemingly mysterious happenings and find means to avoid them, a program of investigation was initiated at the Technical University of Norway. A short account of this work will be given in the following.

THEORY

Although the phenomenon of cavitation was known by Euler as early as 1754 and has been the subject of a large number of theoretical and experimental studies, it is still not fully explained. It is quite clear, however, that it can only occur when the liquid static pressure

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is lower than the saturation pressure at the actual temperature. Experiments have shown that the pressure may, under certain conditions, be considerably lower without cavitation taking place, depending on the existence and nature of evaporation "nucleae". However, when cavitation occurs, it develops very rapidly.

Conditions of considerable under-pressure without cavitation could be produced in the pump suction line under certain conditions. But normally cavitation will occur under circumstances with abundant evaporation nucleae present, and the possible underpressure is then very small. This case is therefore considered in the theoretical treatment.

Evaporators

Pump

t •c Fig. 2. Typical arrangement of liquid separator and pump. The vapour pressure/temperature dia­

gram indicates the influence of pressure drop and temperature rise during stationary opera­tion. Cavitation will not occur when Ah8 > O.

Fig. 2 indicates the common arrangement of the pump suction system. A liquid height H over the suction inlet is available to overcome the various losses and supply a static overpressure relative to saturation at the most critical point, that is at the entry of the pump wheel. The losses are of two different types, viz :

a. Pressure losses b. Temperature rise.

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Let us first consider the situation during stationary conditions when the plant is operat­ing continuously with constant temperature. It is also assumed that no gas enters the suction line with the liquid. The static pressure losses are

L w2 C w2 LI hi = A d · Zg + zg = friction pressure drop

w2 LI hd = 2 g = dynamic pressure

LI hp = due to acceleration and eddy losses at pump entry, often called the "Net Positive Suction Head requirement" ("NPSH"), which is a characteristic function of the liquid delivery for the pump in question.

The temperature rise in the suction line is caused by heat leakage from the surroundings into the system, heat leakage from the motor to the pump inlet, and friction. An analysis for typical conditions will show that the effect is usually small. The combined effect of pressure drop and temperature rise is illustrated in relation to the vapour pressure curve in Fig. 2. As long as the remaining static overpressure Llhs remains positive, cavitation is impossible.

Clearly the operation under static conditions offers no problems when the static height H and pipe diameter d are sufficiently large and the system correctly designed. It appears that it is such simple considerations which form the basis of the pump manufacturer's usual recommendations. During actual usage with fluctuating temperatures the situation is very different, and it is under such dynamic conditions that cavitation trouble usually occurs in practice. Two complicating factors, both aggravating the situation, must be considered, viz : a : Temperature reduction when compressor capacity is larger than demand. b : Gas entering the suction line in mixture with the liquid.

During temperature pull-down on the evaporator system, the liquid in the surge drum will be boiling all the way down to the bottom. This may be so even during stationary temperature conditions as a result of heat leakage through the insulation. In any case we can assume that the liquid will be at its boiling point at the entry into the pump suction line proper, which means that the height LI H in Fig. 2 is lost as a contribution to satura­tion over-pressure at the pump inlet. We have to rely on the height h only, disregarding the heat leakage into the surge drum. Some vapour will also normally follow the liquid into the suction line, depending on the intensity of boiling and the arrangement of the inlet. Let us disregard this additional complication in the first place.

The time lag from the suction line inlet to the pump corresponds to a difference in temperature in addition to that caused by heat leakage. This temperature difference is proportional to the time gradient of temperature reduction and the volume - or rather heat capacity - of the suction line system, at a given pump capacity. If the suction line is very wide, the time lag is great and the system is very sensitive to temperature fluctuation. A very narrow line will, on the other hand, have such a large pressure drop that cavitation may occur from this reason. For any given conditions of temperature and capacity there must be an optimum suction line dimension, giving the greatest possible safety against cavitation, corresponding to a maximum permissible rate of temperature pull-down. It is a question of striking a correct balance of pressure drop and temperature inertia.

The temperature lag due to suction line capacity can be calculated as follows :

where Gr

Gs

500

Cs Gr + G3 • -

Cr Yr . Vp

d t dr

the weight of refrigerant contained in the suction line from the inlet to the point of possible cavitation in the first pump wheel. the weight of the suction line system taking part in the temperature change. Usually the influence of insulation mass is negligible.

Cr Cs Yr Vp d t dr

spec. heat of liquid refrigerant. spec. heat of suction pipe material. spec. weight of liquid refrigerant. volume of liquid pumped per unit time.

rate of temperature reduction in the evaporator.

III-5

Knowing the pump "NPSH" as a function of delivery, the maximum safe temperature gradient can be calculated for different refrigerants, temperature levels, pipe sizes and arrangements, and static height h, taking into consideration flow friction and heat leakage as previously indicated, by setting Lltdyn = Lit •• Results of such calculations for one particular pump with a near ideal suction line arrangement and medium capacity are indicated in Fig. 3. In this case the optimum pipe diameter is about 26 mm, and the cavi-

eoo.--��--,���--.���-r���-.���-.-��---. OCAi d t d'f Vp 40 Ymin 500����+--.��+-�"r-�-r-���-t--���-+-����

20 30 40 50 60 70 d mm Fig. 3. Theoretical maximum rate of temperature reduction dt/dr for an ammonia pump with

capacity 40 !/min, h = r .8 m and NPSH = 0.7 m, as function of suction line diameter cl. Optimum occurs at d = ,....., 26 mm.

tation resistance drops off quite rapidly by deviation to both sides. It is also seen from this example that the maximum permissible temperature gradient increases at lower tem­peratures. At the same time the compressor capacity and the obtainable rate of temperature reduction is decreased. It is thus obvious that the most critical conditions of cavitation occur at relatively high evaporation temperatures. This is in complete agreement with the experience of the author over the years.

Rational design of liquid pump systems requires knowledge of optimum suction line velocities for different refrigerants and conditions. It appears from Fig. 3, and it has been supported by calculations for other capacities and arrangements, that the optimum velo­city varies little with temperature. The same is found to be the case when the static

501

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height h is changed. Fig. 4 shows results of calculation of maximum temperature gradient as a function of velocity for four different refrigerants. The optimum velocity is remarkably uniform, in the range 0.8 to 1 .0 m/sec, and a figure of 0.9 m/sec is recommended for design. The nearest available pipe size above the calculated diameter should be used.

"Sh dt dT

I' ...... \ A

o - 20 Ysec. l!.- -- 40 l-iec. $---60 �ec. 500 t-<-+----j--+----+--1-+-+- -+---------+--+----<

CF,Cl2

CHF,C I

0L-L-'-�--'-�--'-�J.-J'-'-�--'-�--'�-'----'-�--' 0,03 0,1 0,3 3 1 0 m/sec.

Fig. 4. Theoretical maximum safe rate of temperature pull-down as a function of suction line velocity for different refrigerants.

It is also apparent from Fig. 4 that there is a marked difference of sensitivity to cavi­tation for the different refrigerants. Propane is the most difficult of the four, while the halocarbons are fairly insensitive, much due to their high specific weight in the liquid state.

In the above calculations it was assumed that no gas bubbles entered the suction line. If this happens, they will condense during the rise of pressure in passing down the line. The heat of condensation is then liberated, and the liquid temperature is increased by the following amount :

502

v · r · y g L'.f tcond = -----Cr • Y r

where v gas volume fraction. r heat of condensation per unit weight. y g spec. weight of the gas.

111-5

This temperature rise comes in addition to the others, and reduces the maximum temperature gradient accordingly. It is clearly imperative to limit the gas entry to an absolute minimum. Table I indicates the maximum possible gas volume fraction for the investigated pump under some operating conditions with no temperature fluctuation. Again it is seen that the situation is by far most critical at high evaporation temperatures. In practice the gas fraction must, of course, be much smaller to permit safe operation un­der cycling conditions. The liquid intake to the suction line should be designed with parti­cular care to reduce the gas fraction to a minumim.

Table I. Maximum gas volume fraction v % in the suction line intake without cavitation under stationary temperature conditions for an ammonia pump system.

I Pump capacity I Pipe diameter I v max in percent by volume at t o = litres/min. I mm ! 10° C I 0°C I -10° c I -20°C I -30°C ! ----- - - -

I I I 20 19 15.9 27.2 44 103 217 40 25

I 13.3 22.8 37 83 183

I 60 I 34 9.3 16.6 27 I 59 127 I EXPERIMENTAL PROCEDURE

In order to test the theory a large number of experiments have been carried out, using a typical canned motor pump of the general design shown in Fig. 1 . The test arrangement is shown schematically in Fig. 5. A series of five pump suction lines of different diameters

To compressors --- -----

Air cooler

c B A DP

rv Fig. 5. Schematic arrangement of measurements. The pump delivery is measured by means of a

nozzle N, and may be diverted to air cooler or directly to separator S according to required temperature gradient. Liquid level in separator is measured by oil filled differential mano­meter A, suction line resistance by B and evaporator pressure by mercury manometer C. All pressure taps below liquid are horizontal and heated to avoid liquid columns. W indi­cates observation windows.

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ranging from 19 to 50 mm were used, and tests were also made with the pump fitted directly to the bottom of a 267 mm liquid leg. Two compressors of different capacity were used, one of which could be continuously speed regulated for capacity control. The pump discharge could be passed through an air cooler of adequate capacity or returned directly to the liquid separator, and this gave a further possibility to vary the rate of temperature pull-down. It was possible to obtain gradients of several hundred ° C per hour, far in excess of normal values in practice. The evaporator pressure and various pressure differen­ces were measured by differential manometers with mercury or oil filling. All pressure taps under liquid were carried out horizontally and heated, to avoid any liquid columns in connecting piping. For the same reason oil was bled off regularly. The amount of liquid pumped was measured by means of a standard nozzle in the discharge line. Unsteady

B o c a 500 ������������--.���� dt a d't �

400 t-���-+���· �l--�·��-+-���--t

Fig. 6. Maximum temperature gradient as a function of evaporation temperature for pump capa· city 40 litres/minute and suction line diameter 40 mm. Curve A represents theoretical values for gas free inlet condition. Curve B and observed points o represent condition of heavy cavitation with common inlet arrangement, where much gas is sucked in with the liquid. By changing inlet to a conical form with anti-rotation vanes the gas fraction is much reduced, and conditions with no apparent cavitation were observed well above the theo­retical curve A, points C o.

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discharge pressure and delivery were used as indications of cavitation in the pump. At severe cavitation conditions the liquid pumping stops, and a differential pressostat DP was installed to cut out the pump, in order to avoid bearing damage in such cases.

The conditions at the suction pipe inlet and the liquid surface could be studied through a number of observation windows, and films were taken to study the aspiration of gas bubbles with the liquid.

Ammonia was used as a refrigerant during all the tests. Measurements were made with different pipe diameters, liquid flow, and temperature level, and the necessary tempera­ture reduction gradients to give varying degrees of cavitation were observed.

RESULTS

Altogether several hundred tests have been carried out. Using a conventional suction line inlet design it was soon discovered that cavitation did occur at temperature gradients much smaller than calculated for vapour free liquid entry, and especially at high rates of pumping, Fig. 6. It was also found that the cavitation limit was not very well defined. The reason for the important deviation from the theory was that considerable amounts of vapour were sucked into the line with the liquid. This was due partly to the formation of a vortex and a bubble train from the surface, partly to collection of gas formed by evapora­tion in the surge drum below the suction line entry.

It was found by special experiments that a vortex and bubble train may form at pipe submersion as much as 600 mm, even with the small diameters and pumping rates used in these tests, Fig. 7. It depends on the state of rotation existing in the separator liquid, but occurs at irregular intervals and seemingly without any direct relation to the conditions of operation. The frequency and amount of vapour transported increase as the submersion

Fig. 7. Photograph taken through an observation window in the liquid separator shows how gas from the surface vortex is sucked into the 40 mm diameter pipe. The pumping rate was 60 litres/min, submersion 250 mm.

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is reduced. The effect can be eliminated at reasonable submersion by installation of proper guide vanes to prevent rotation.

Fig. 8. Vigorous boiling in the separator during temperature pull-down. It is easily observed how the gas bubbles collect and are sucked into the pipe opening.

The entry of vapour from the lower parts of the separator also depends heavily on the geometry. During temperature pull-down the entire body of liquid is in a state of vigorous boiling, Fig. 8, and the bubbles from below may collect to be sucked in with the liquid. The situation may be improved by designing the liquid inlet so that as much as possible of the gas is forced away from the opening.

Various forms of suction line entries have been tried, and good results were obtained with a design as shown in Fig. 9. Vapour suction from the surface is completely eliminat­ed at moderate submersion by the guide vanes, and the gas fraction limited to that existing immediately at the entrance as a result of temperature relaxation at this point. This de­pends on the temperature gradient, the size of bubbles formed and their velocity of ascent in the liquid. Measurements are in progress to ascertain how much vapour is present in a boiling liquid at different levels and rates of temperature reduction. Such data are much needed for several purposes, such as for instance the rational design of liquid interseptors for fluctuating temperature conditions.

A number of test runs have been made with an inlet arrangement as shown in Fig. 9, under different conditions of temperature and pumping rate. The results were highly satisfactory in so far as it proved impossible in most cases to obtain noticeable cavitation under continuous operation of the compressor at temperature gradients up to and in ex­cess of the theoretically calculated values. Some experimental points are indicated in Fig. 6, and correspond to cooling rates much higher than normally encountered in general refrigeration practice.

506

t­i Min. 300 mm

..c:

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Fig. 9. Successful design of suction line entry. The guide vanes eliminate vortex formation and bubble train from the surface completely at moderate submersion. Gas from the lower levels of liquid is diverted away from the pipe entrance.

At the moment of starting the pump special conditions prevail. If there is not a rising temperature in the system, the liquid in the suction line and pump will be at its boiling point all the way. In spite of this it was found that the pump would always start without any difficulty, when it was cut in before the compressor. Trouble may occur, however, when it is tried to start the pump during rapid temperature pulldown.

It was also found that cavitation and pump fall-out may occur as a result of sudden cutting-in on the system of a large capacity compressor. When the liquid separator has been on a rising temperature for some considerable time, it appears that a deficit of evaporation nucleae has developed. During the first stage of pressure pull-down the onset of boiling is delayed, and a state of superheated liquid develops. When boiling suddenly sets in, it occurs with an intensity to completely empty the suction line of liquid, and cavitation is inevitable.

In order to eliminate this difficulty it is recommended to use an automatic gas bypass from the condensor side during compressor starting, Fig. 10. It may also be used during starting of the pump as an extra measure of safety. By this provision a dangerous tern-

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perature gradient is avoided during the critical moments of starting, until the recirculated liquid-gas mixture from the evaporators has supplied the whole system with a sufficient concentration of nucleae.

-

s

-

Fig. 1 0. Arrangement of gas bypass from the condenser <luring compressor starting.

CONSIDERATIONS IN SYSTEM DESIGN

In accordance with the results reported above the most important points to observe in the design of a liquid pump system can be summed up as follows :

1 . The available static liquid pressure h (Fig. 2) must be well above the pump "NSPH". Ordinarily 0.5 m in excess is sufficient.

2. The suction line diameter should be chosen to give a normal velocity of 0.8 to 1 .0 m/sec. It is just as bad to use a too large diameter as it is to use a too small one.

3. The suction line should be run as directly as possible, with a minimum of bends, to give a small friction loss. Any valves or filters must be chosen with the same aim in mind. Gas locks in the line must be absolutely avoided.

4. The suction line inlet should be designed to avoid intake of gas bubbles coming from the lower parts of the separator. It is further recommended to use guide vanes to eliminate rotation in the liquid.

5. The submersion of the inlet must be sufficient to avoid formation of a vortex. By use of guide vanes it may be reduced to some 0.3 m for pipes up to 50 mm diameter. With­out guide vanes more than 0.6 m is needed.

6. The suction line system should be designed with a view to reduce its heat capacity to a reasonable minimum.

7. The liquid holding capacity of the separator must be sufficient to avoid level re­duction below acceptable submersion height at all times. The level control must be of a reliable type.

8. In order to avoid difficulties during starting of pump or compressor, an automatic gas bypass from the condensor is recommended (Fig. 10).

9. It is recommended to use a differential pressostat as a safety control to protect the pump in case of cavitation.

When these recommendations are followed, it is believed that all trouble as a result of cavitation may be avoided under any conditions normally occurring in refrigeration plants.

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ACKNOWLEDGMENTS

Most of the experimental data used as a basis for this report were collected by the mechanical engineers Rolf Pettersen and Harald Haugerud. Mr. Kare Aflekt assisted in drawing the figures. The pump was presented by the firm A/L Landteknikk.

REFERENCES

r. G. Lorentzen, " Automatic Control of two-stage Refrigerating Plants". The Danfoss Journal 1957, No. >, pp. 19-29.

2. G. Lorentzen, "Evaporator Design and Liquid Feed Regulation". Bull. IIF, 1958, Annex 2, pp. 235-256.

3 . G. Lorentzen and 0. A. Baglo, "An Investigation of a Gas Pump Recirculation System' ·. Proc. X Congr. Refrig. 1959, II, pp. 215-224.

DISCUSSION

Hirschberg, Switzerland : Having carried out some tests on cavitation problems with centrifugal refrigerant pumps which led to very similar results as described in the very excellent paper of Professor Lorentzen I would like to add some remarks concerning observations which might be of common interest. The experiments were carried out with a testing arrangement which consisted of a surge drum, a vertical connecting line to the pump and a return line with a throttling valve and a device for measuring and recording liquid flow rate. The vapour developed in the surge drum was pumped off by an oil free compressor with a by-pass for the modulation of suction gas rate. By varying the gas flow through the by-pass valve we could produce pressure drop curves in the surge drum like the one shown below :

p

z

As it turned out to be very difficult to measure the exact slope at the very beginning of the pressure drop, an average value was taken over the first 30 seconds. The tests showed that there was a critical value (dp/dz) crlt optimum liquid velocity which did not alter significantly with temperature. In the temperature range between 0° C and -25° C there was only a slight reduction of (dp/dz) crit towards the lowest temperatures. In our special case the average value was dp/dz = 0.4 at/min. between 0°C and -25° C, dropping to 0.3 at/min. at -35° C. I therefore feel it should be more convenient to use the pressure slope instead of the temperature gradient.

Another interesting observation concerned the pressure gradient at which the pump circulation recovered. It always came out to be 1/lOth to I /20th of the cut off slope, so that there was a considerable time elapsing until the liquid circulation started again.

If there are difficulties with liquid pumps the manufacturers often would give the advice to reduce the liquid flow rate by throttling. This seems to be not always the right remedy, as our tests showed that there is a great increase in sensitiveness towards pressure drops if the centrifugal pump is working near the culmination point of its characteristic curve. In this range even very small pressure drops will cut off liquid circulation.

The method of by-passing the compressor during its starting will prevent cavitation conditions caused by the compressor itself, in many cases however difficulties come from the evaporators as well and in this case the by-passing method seems to be too slow in reaction to avoid pump cavitations. My question to the author of this paper is, whether he has any experience in reducing shocks coming from the cooling systems.

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III-5

G. Lorentzen, Norway : I did not consider it very important whether one plotted the results in terms of variation of pressure or temperature. Of course, it was an advantage that the pressure gradient would be more constant, but I feel that the choice of tem­perature was useful since it brought out more clearly the rather important fact, that the problem was much greater at higher than at lower evaporation temperature, which is contrary to the belief of many people in practical engineering.

The problem of initial starting up of the plant, i. e. starting the pumps and compres­sors, is very different from that of continuous operation. When the pump has been idle, the liquid is usually at its boiling-point all the way down the suction line and pump inlet. Cavitation will then inevitably occur at the moment of cut-in, but this does not usually prevent the circulation from starting, and the situation is then immediately corrected. The problem of compressor starting after a prolonged period of stand-still is more difficult, due to the frequent development of a superheated state of the liquid as a result of delayed onset of boiling. When boiling finally sets in, it is so vigorous that it will blow empty the entire suction line, and cavitation is inevitable. It is pritnarily to avoid this that starting should take place at a very slow rate of temperature reduction, until the normal boiling condition is reached.

As for the effect of pressure pulsations generated by the evaporators, this problem has not been investigated separately. I rather imagine that the effect of a pressure drop will be the same, regardless of its origin, and that it is essential that the whole system is designed and operated in a way to limit negative pressure gradients. I am glad this point has been brought out.

J. Kowalczewski, Australia : Vapour in the liquid pumps may be formed (a) due to viscous effects (friction) in the pump suction line in steady state and non-steady state, (b) vapour production due to the pressure being lower than the saturation pressure corresponding to the liquid temperature (this effect disappears in steady state conditions). The rates of temperature decrease used in this paper to limit the diameter of the suction lines are much higher than the rates occurring in practical applications. This is because there are other thermal capacities in the system in addition to the thermal capacity of the suction line which have not been considered. It is therefore suggested that in prac­tical applications the velocity in the pump suction line may be lower than the 0.8 m/sec suggested in the paper. Could the author comment on this, please ?

G. Lorentzen, Norway : All thermal capacity associated with the suction line, from the surge drum to and including the pump inlet, must be taken into account in calculat­ing the optimum velocity. No other thermal capacities were relevant to this problem. For the systems investigated the optimum velocity would be between 0,8 and 1,5 m/sec. This does not mean, however, that pumps can not be operated at lower suction line velocities, but the maximum acceptable rate of temperature reduction will be less in that case. But, as pointed out by Mr. Kowalczewski, the calculated maximum tempe­rature gradients are rather high compared to usual values in practice, and it may there­fore be, that satisfactory operation will often be reached at suction line velocities de­viating considerably from the optimum.

W. H. Emerson, U. K. : I refer to Dr. Hirschberg's remarks in the discussion that some manufacturers of liquid circulating pumps recommended throttling of the discharge as a means of preventing cavitation, though he had not found this advice effective with a centrifugal pump. He said he was reminded of his own experience some years ago with a gear pump circulating liquid refrigerant 12. Under conditions of steady state and with an apparently adequate static head on the suction side, cavitation still occurred when the differential head across the pump was low, but it could always be arrested by throttling the discharge.

I understand that the advice quoted by Dr. Hirschberg was intended to be applied to centrifugal pumps, in which the effect of throttling was to reduce appreciably the mass flow rate. This was not the case with the gear pump, and I am at a loss to know how to explain how throttling such a pump can prevent it cavitating.

510

III-5

G. Lorentzen, Norway : I think it is the characteristic of the pump entry which might possibly have something to do with it. However, all the experiments which were carried out and described in this paper were done with a centrifugal pump and I, therefore, do not have the answer to the question posed by Dr. Emerson.

A. Neuenschwander, France : I congratulate Professor Lorentzen on this most important work which had been undertaken at the University at Trondheim, especially as these troubles occur very frequently in practice in installations with liquid re-circulation at low temperatures. I am also most interested in the work which has been done in relat­ing the pump calculations with the low pressure receivers.

There is one point, however, which I want to raise and that is the question of oil in the evaporator circuit. In practice it is practically impossible, except in very exceptional circumstances, such as oil free compressors, to keep oil from going over into the evapo­rator and thus contaminating the refrigerant. I wonder if Professor Lorentzen would be able to tell me something about the effect of oil in the ammonia and possibly also in those refrigerants which are mixable with oil.

G. Lorentzen, Norway : We have made no experiments with varying oil concentrations, however, I would imagine that in an ammonia plant, with the very small amount of oil per unit volume which is actually circulated by the pump, there should be very little fear of trouble on this account.

In Freon installations conditions are of course different inasmuch as it is a question of the influence of the oil on the vapour pressure curve and the specific heat of the liquid. However, with normal concentrations encountered, I would think that the influence here too would be very small indeed.

5 1 1

III-40

Liquid Content in Evaporator Pipes

Le contenu liquide des tuyaux d'evaporateur

Prof. Dr.-Ing. K. LINGE Maschinen-Laboratorium, Technische Hochschule, Karlsruhe, Germany

SOMMAIRE. Une question interessante pour les tuyauteries d'evaporateur est la sui­vante:

Quelle est la partie du volume d'un tuyau d'evaporateur occupee par la phase liquide, la teneur en vapeur etant variable le long du tuyau. Cette proportion de liquide depend en outre du volume specifique ainsi que du rapport des vitesses de la vapeur et du liquide.

Une equation est proposee, qui permet de calculer la proportion de liquide partant des conditions d' exploitation. Les resultats sont representes dans un diagramme pour diff erents cas d'interet pratique.

Des differences particulierement importantes de la proportion de liquide s' observent dans les tuyauteries d' evaporateur de patinoire. La proportion de vapeur a la sortie du tuyau, qui est determinee par la puissance frigorifique possede la plus forte influence sur le contenu liquide. L'infiuence des autres conditions de f onctionnement est par contre tres f aible.

La variation caracteristique du contenu liquide pour diverses puissances frigorifiques est representee dans un diagramme. On peut en deduire la capacite necessaire des bouteilles accumulatrices de liquide; il est en outre possible d' avoir une idee du transport de liquide entre le tuyau et le reservoir en regime variable.

A liquid-vapor mixture enters a pipe (length 12) with a vapor concentration x1 and lea­ves with a vapor concentration x2• The ratio of the vapor and liquid velocities, as well as the heat flow, is considered constant over the length of the pipe. The liquid volume then decreases linearly and the vapor volume increases linearly.

At a point l the liquid volume per unit mass, with a specific volume v', is

and the vapor volume per unit mass, with a specific volume v", is

Vv = [x1 + (x2 -x1)f] v"

For a liquid velocity w' and a vapor velocity w" the ratio of the pipe cross-sectional area occupied by the liquid a' to the entire pipe cross-section a is :

a' a

The volume percent of liquid based on the entire volume of the pipe then is:

la

L = :21 -x_:-2--

--��:-x-:�1 -:-:--

l_l_

+_ [ __

x_1 __

l2_

+_1]_v_"

_w_' d/

x2 - x1 x2 - x1 v' w" Ii = 0

513

III-40

Upon integration :

L J v" w'

__ 1_ In 1 v'

w" (x2- XI)

-- - 1

(v" w' ) v'

w" (v" w' _ )2 f II 1

v w + XI -- - 1

(v" w' )

v' w"

(v" w' ) j v'

w"- l

With this equation one can calculate the liquid content for any arbitrary case.

Of interest is the case when sufficient heat is supplied so that pure vapor leaves the pipe (x2 = 1). This occurs in refrigerating machines with continuous evaporators, where the control valve permits just enough liquid to enter the evaporator pipe so that the liquid leaves the pipe completely evaporated. Because of variable sub-cooling of the liquid, as well as variable condensation temperature, the vapor concentration XI of the liquid throttled into the pipe can vary within narrow limits. The question is how the liquid content varies when XI varies and x2 remains constant (x2 = 1). In this case, with the substitution v"

w' v' w"

the equation is

c

L 1 { 1 c }

C -- In - (C- 1) (C - 1)2 1 - XI l + xI (C- 1)

Fig. 1 presents the plot of L for different XI and various values of (C - 1) with x2 = 1.

10 9 8 7 6 5 4 3 2 1 0

"

"" '\. '"

.... /1

"'<.:: ....

3

"" "'-,

!'... "r-.,r-.. � ......._ '�on k ; "

r--.r--. on;::....... "" � .;:. ....... � --� -........... r--_ ...... l{...j

5 102 103

Fig. r . Liquid portion in Vol. % for different ratios of specific volumes and of velocities at various vapor concentrations x1 and constant vapor concentration x2 = r .

514

III-40

v" w' For NHa at to = -10°C, / = 272. If one sets � = 1, i. e. if one considers the vapor

v w and liquid velocities equal, then, according to Fig. 1, L = 0.8 % at x1 = 0.05, and L =

v" w' = 0.5 % at x1 = 0.15. For R 12, / = 1 1 1 at to = -10°C. For � = 1, L = 1.8% v w at x1 = 0.05 and L = 1.0% at x1 = 0.15.

The liquid content L in per cent by volume is very small and varies only slightly with x1 so that a new steady-state condition is quickly reached.

Another case of interest is that which occurs in a circulation evaporator as used in ice-skating rinks with direct evaporation. Liquid is pumped with a constant mass rate of flow from a container through the evaporator pipes and the refrigeration performance is varied according to the outside temperature. Here x1 = 0 and x2 is variable. The equa­tion then reads :

L = cc� 1) 2 {c

x

1

2 In u + x2 cc - 1)] - cc - 1)}

v'' w' Fig. 2 presents the plot of L at different x2 and C = / · � • For better recognition v w

of the relations, an example is considered: t 100

,..._ 90 -�·

x,,=00001

0 - 80 � i.::::.. 70 .....J � 60

<5 50

40

30

20

10

- ---� ('Jn,.. . � ..........

..... , " '

I'.!'.. ''-..... 0-�

"" """ '

' '- "�

" ', "" �o<-

""' !'-.., � " .....

' l'-......._Q,7 !'-r--.,._ �

� r--........__ � i--t-- I I"-...... ,.. �s --

--

........... I'--...

""' "� """

'....... "r--......... � r-- �

� -

-�-

""' ....... ,

r--\. �

" """

' \.

\

........ !"-.... r-!"--'"- ,...,..

�

3 5 102 3 5 103 v" w'

fy; · w,, 1 ) -+ Fig. 2. Liquid portion in Vol. % for different ratios of specific volumes and of velocities at various

vapor concentrations x2 and constant vapor concentration x1 = o.

515

III-40

The evaporator pipes of an ice rink are supplied with a constant flow of liquid, which is so large that, at maximum refrigeration performance, twice as much refrigerant with x1 = 0 enters as can be evaporated. The maximum vapor concentration at the end is then x2 = 0.5 and decreases with decreasing refrigeration performance. With decreasing refrigeration performance Qo, the evaporation temperature to increases, e. g. from to =

v" -15° C at Qo = 100% to to = 0°C at Qo = 0 %. Correspondingly / varies from 335 v

w' to 185. If preliminarily -,, ist set equal to 1, for ammonia the curve in Fig. 3 results. w

v" For R 12, ----, varies from 133 to 79, so that, under the same conditions, the somewhat v higher curve for R 12 results in Fig. 3.

'i;I � -� a.::::.. ......

§ ·­-�

700%

50

I

\ .\ \ \ R

'\� >- '" ""-

I I I i ' I i

! I

I

12

r---... -._ � I 0o 50 700%

I' 'H3

� � Ref rigerafion -+ g Performance

Fig. 3. Liquid portion in Vol. % for NH, and R 12 at different refrigeration performances.

w' w' 1 The curve for R 12 at -,, = 1 is valid for ammonia at -,, =

2 4, i. e. if for ammonia the w w .

vapor velocity were 2.4 times the liquid velocity. It can therefore be seen that the in-w' v''

fluence of -,, , as well as that of ----, , on the results is only slight. w v The necessary size of accumulators for desired ranges of adjustment can be read from

the curves in Fig. 3, which can be determined from Fig. 2 for other operating condi­tions as well. Fig. 3 also shows that the liquid range varies within broad limits according to the desired range of adjustment and that a new steady-state condition can be reached only after a longer period of time, after considerable amounts of liquid have been sup­plied to or drawn from the evaporator pipes. Knowledge of the magnitude of these amounts of liquid is useful for the operation of the refrigeration unit and supervision of the controls.

DISCUSSION K. Gutkowski, Poland : It has been established by experimental investigation, that

in horizontal pipes there can be three main types of two phase flow. The appearance of these depends on the Reynolds number of liquid rate and the vapour concentration,

516

IIl--40

which results in a stratified flow with separation of liquid and vapour, a waved flow or a homogeneous flow at which the liquid drops are being carried along by the vapour stream.

The calculation method presented in the paper appears to me to be based upon the acceptance that the ratio value of the vapour and liquid velocities is constant. It seems to me further, that such constant ratio is possible only in homogeneous flow. Thus for the technical application of this method to be complete there still remains the need to give the criteria for the design of evaporators in which the homogeneous flow can be expected, as well as to establish the value of the velocities.

L. Vahl, Netherlands : The calculation method on this paper only applies to the limit where the velocity ratio is one and then it represents a simplified approach to the pro­blem, but it cannot really be used at other velocity ratios and further work on this would have to be done.

P. Danig, Denmark : I want to make a further remark on the vapour/liquid velocity ratio which is assumed constant. This ratio must in fact vary a great deal along one evaporator tube and depends very much upon the flow pattern in the tube particularly in a dry expansion evaporator. Therefore, I think that the theory represented in the paper can only be applied to evaporators with re-circulation and even in that case the answer will only be approximate if the difference in inlet and outlet vapour concentration is great.

K. Linge, Germany : There is a further task in deriving the equations for other velocity ratios than the ones that I had taken. It is quite obvious from the discussion that there is great interest in carrying this investigation further.

W. H. Emerson, U. K. : The results given in the paper do not appear to be of direct value to the designer since it leaves unresolved the question of the velocities of the two phases which may vary widely. However, it seems on first acquaintance that results might be used by the experimentalists to determine the velocity ratio under different regimes of flow by measuring the liquid fraction. The latter measurement is much easier to make than the direct determination of the mean velocities of the two phases.

K. Linge, Germany : We have in fact a programme which will take care of these meas­urements and they will firstly be carried out with water and air. This, of course, will be a first step since it is a much simpler case than the evaporation of liquids.

517

III-16

The Problem of Refrigerant Return Line Calculations in Pump Recirculation Systems

Le probleme des calculs de la tuyauterie de retour du frigorigene dans les systeme a recyclage par pompe

K. GUTKOWSKI S I M P, Warsaw, Poland

SOMMAIRE. On presente dans ce rapport une analyse de l'ecoulement liquide-vapeur a deux phases dans les tuyauteries de retour du frigorigene d'un systeme a recyclage par pompe.

On montre !'importance de !'influence de la perte de charge dans la tuyauterie de retour, en presence du liquide retournant de l'evaporateur a l'accumulateur.

Dans ces tuyauteries, suivant la quantite de liquide et la temperature du frigorigene, on peut s' attendre a une elevation de la perte de charge de plusieurs f ois superieure a celle des conduites de gaz. Cette augmentation de la perte de charge entraine une elevation reelle de la temperature d' evaporation : aspect particulierement important des systemes f onctionnant a basse temperature. Dans le cas de !'ammoniac, pour une vitesse de recyclage et une tempe­rature d'evaporation donnees, la perte de charge dans la tuyauterie de retour d'un systeme avec recyclage par pompe sera environ cinq fois plus elevee que celle d'une tuyauterie semb­lable sans recyclage.

Le calcul du diametre du tuyau peut done etre considere comme l'un des aspects les plus importants de la conception des installations a pompage de frigorigilne.

L' A. considere necessaire d' entreprendre cette etude, car les ingenieurs peuvent avoir ten­dance a se contenter d'une simple methode de calcul, ce qui n'exige pas beaucoup de temps et permet d'utiliser les abaques existantes pour etablir le diametre des tuyaux. A l' appui des travaux fondamentaux sur l'ecoulement a deux phases par les chercheurs americains R. W. LOCKHART et R. C. MARTINELLI, confirmes plus tard par des recherches en marge de ce probleme, on a etabli une formule numerique. Cette formule permet de calculer facilement la perte de charge de l' ecoulement vapeur-liquide du frigorigene dans les conduites, dans des conditions de fonctionnement donnees, lorsque la pression d' ecoulement du gaz correspondante est connue. La trans! ormation de la formule perm et de calculer approximativement le dia­metre du tuyau pour l' ecoulement de vapeur et de liquide lorsque sa valeur est connue pour l'ecoulement du gaz.

On donne un exemple numerique pour habituer les utilisateurs au procede de calcul.

The problem of a proper choice of diameters of refrigerant return lines in refrigerating plants with pump recirculation deserves a special treatment owing to the high flow resistance values encountered inside these lines.

In these lines as they are filled with the liquid refrigerant returning together with the vapour form the evaporator to the accumulator, the value of the flow resistance is higher than in suction lines of plants without recirculation. The rise in the flow resistance value corresponds of course to an increase in pressure and evaporation temperature causing thus a decrease of the cooling effect and the necessity of lowering the suction pressure of the compressor, accompanied by lowering its capacity and the efficiency of the cooling cycle. This phenomenon is evidently discernible at low evaporation temperatures where an improper choice of diameters of return lines can raise the evaporation temper­ature even several degrees centrigrade.

In view of this, a proper calculation of return lines can be considered as one of the most important problems of designing of refrigerating pump recirculation systems. As only a simple and convenient computation method could be accepted and widely applied by the designers, the author of this paper found it appropriate to present a method for this purpose.

519

III-16

The problem of two-phase isothermal flow has been investigated in a general way by the American scientists R. W. Lockhart and R. C. Martinelli, who established some relations between physical values measured during the flow. Results of investigations accomplished by other investigators upon a series of mediums with very different viscosities, densities and physical properties, have been evaluated by these two scientists. It should be stressed in this connection that the treatment of the problem is similar to the mechanics of a two-phase flow as described by S. Kosterin and later by G. E. Alves and others.

As the results presented by the scientists in question have been obtained from a great number of experiments and confirmed lately by other investigators they can be conside­red as most reliable and fundamental for this kind of flows.

The following preliminary assumptions have been accepted in the consideration of two-phase (liquid-vapour) flow taking place in the return lines of refrigerants which is the subject of the present paper.

a) Small heat quantities penetrating into the insulated line don't affect the flow con­ditions.

b) The flow of each phase at a given mass output under single-phase flow conditions would be a turbulent one (at Reynolds number values exceeding or equal to 2000).

At turbulent flow of vapour, which always takes place in suction lines, and a laminar flow of liquid which would be sometimes existing at very small circulation numbers n < 2, a great length of the line and higher evaporation temperatures, the existing flow resistance value would be somewhat lower than that being computed.

Lockhard and Martinelli maintain that an increase in flow resistance as measured du­ring two-phase flow as compared with flow resistance values for the flow of the gaseous phase only, at a given mass output, expressed as a ratio of both resistance values, depends solely upon the mass rates of flow, of both phases their viscosity and specific volume. This relation can be expressed as a function:

LI Pm - = f (X) LI Pg (1)

where : LI Pm - resistance of two-phase flow related to a unit length of the line -kG/m • m LI P g - resistance of the gaseous phase related as above - kG m •/m X is a number giving quantitative and physical flow conditions given by the equation.

x2 = (G1)1,8 V1 (!!:!_)0,2 Gg V g µg (2)

where: G mass rate of flow - kg/h v specific volume - m "/kg µ dynamic viscosity - kG sec/m •

The index "1 " applies to the liquid phase, and "g" to the gaseous phase.

The ratio�; in the case of refrigerant return lines can be considered as the ratio between

the quantity of refrigerant supplied to the evaporator Gr (kg/h) minus the vapour quan­tity formed in it G v (kg/h), and the vapour quantity : G r

- - 1 Gv (3)

G the ratio ____!. is usually called the recirculation number and denoted by n. Designating Gv the difference determined by (3), as n - 1 = C, the formula (2) can after a transforma­tion be expressed thus :

520

2 1 c = x T,8 {� :(;;)°'2 } -T,8 (4)

III-16

or shorter, as : C · A = X 1,11 (5)

1 where A = {� ;(;:r2 } 1.8 is the expression taking into

account the influence of the saturation temperature of the flowing refrigerant, upon which the values of V and µ are dependent, on the flow resistance.

After denoting the product C· A · 10 • as Q, the function (1) can be written thus

��: = F (Q) (6)

The function (1) has been shown by Lockhart and Martinelli as a curve in a diagram with a logarithmic scale. The values corresponding to a series of points of the diagram have also been collected within a table, the equation of the function that would allow for a mathematic transformation was not given, however. It seems that in this situation the most reasonable solution would be the converting of the X-values contained in the table into Q-values, then drawing a curve correspondig to function (6) and establishing its equation. From the table values as given by Lockhart and Martinelli for refrigerating plants only those for X = 0.01 - 1.0, given below, have an interest.

Q LI Pm

Q LI Pm x -- = Z x = Z LI Pg LI Pg

0,01 0,00625 . 10 . 1,54 0,2 0,168· 10 2 4,975

0,02 0,01292· 1,875 0,4 0,364· 8,01

0,04 0,0282· 2,375 0,7 0,673 12,48

0,07 0,05235 · 2,92 1,0 1,000 · 17,62

0,10 0,0776· 3,42

A thin line has been drawn through the points given above, as shown in Fig. 1. As there have been difficulties in finding a mathematic formula strictly coinciding with the curve, a substitutional curve running very closely to the former has been drawn as a thick curve in Fig. 1.

,, ,, " -z

�

•

,I · -,_ . • ' / '

/ / /

.. -- ,1ri.f/l!fj t2� ,

0 ' � • "

'1\11' 1\1* ,,,.

,; "\<t':J>-�

"' .,

� A -- A

-- .... - , / '/

J .... I/ /

" .. "' " " .. " ., •<R"1 Fig. r . Thin curve corresponding to values in the above table superimposed by thick curve

given by equation (7)

The later curve is given by the equation :

LI Pm LI Pg= 1 + (1 + 2,56 · 10- • Q) ln (Q + 1 ) (7)

521

III-16

where according to the previous considerations

Q = (n - 1) · A · 10 2 (8)

In order to facilitate the determination of the A-value as a function of the refrigerant temperature, a curve depicting this function for ammonia has been shown in Fig. 2.

8 A·102

6

5

'\

3 0

" r\..

-10

'\� " !'-...

""" ' -20 -30 fo oC -40

Fig. 2. Influence of the saturation temperature on flow resistance A versus refrigerant temperature

The equation (7) allows for computation of flow resistance within a line of a given diameter and checking if the resulted pressure difference between the evaporator and accumulator doesn't exceed the admissible value. An approximate formula

1 dm = d g • Z 4•8 (9)

can be used for a preliminary choice of line diameters where d g means a diameter suitable for a gaseous phase flow, being computed or taken from a diagram intended for deter­mining the suction line diameters of refrigerating plants operated without recirculation.

L1 Pm Z =

L1 p g is determined for given work conditions by equation (7). It is, however,

advisable, after taking the preliminary choice, to calculate again the resistance value L1 Pm for the entire piping in order to establish the expected pressure drop within it. For the computation of the flow resistance value L1 P g, the coefficient of friction resistance as given by Blasius

0,184 ;. = Re 0,2 (10)

for turbulent flow in smooth pipes shall be taken into consideration.

This form of the formula has been used for the determination of the equation (2) for X, for which the function (1) has been established experimentally. It seems that the appli­cation of the coefficient binding for smooth piping is fully reasonable in the case of usual piping applied for refrigerating installations. These pipes display an even roughness and can be considered according to Ideltschik as hydraulically smooth as their relative roughness L1 < 17 ,85 · Re-0,875 for Re :S: 105 (the relative roughness is the ratio of the height of the

unevenness L1 mm at the inner surface of the pipe to its diameter d mm, being thus L1 = � ) . Furthermore the boundary layer existing at the pipe surface effectively reduces its roughness.

For computation of the capacity of the liquid accumulator, the possibility of evaluation of the liquid quantity present in the return line, which will after stopping the pump flow back to the accumulator, is of utmost importance.

In order to facilitate this evaluation the curve R1 = f (Q) which allows for reading the percentage liquid rate in the piping volume, as a function of the Q value, has been plotted in Fig. 1.

522

IIl-16

This curve coincides with a number of values established empirically and given in the papers of Lockhart, Martinelli and other investigators concerned with this subject.

EXAMPLE OF COMPUTATION

The diameter of a return pipe of an ammonia refrigerating plant equipped with a recirculation pump has to be established. This line conveys the refrigerant from evapora­tors delivering 100 000 kg cal/h. The evaporation temperature is t o = -40° C, the cir­culation number n = 6, the length of the line 30 m. The line is equipped with a straight­flow valve, fully open, and has four bends.

1 for t o = -400 C ; A = { 1,4493 . 10-3 ( 29 . 10-6 )0,2 } 1,8 0,0309

1,55 0,79 . 10-6

and Q = (6 - 1) 0,0309 · 10 • = 15,45

LI Pm Z = LI pg = 1 + (1 + 2,56· 10-2 15,35) In (15,45 + 1) = 4,9

By means of the diagram the d g value has been established for the given Q and t o, as d g = 0,1 m and the dm value preliminarily accepted as

1 dm + 0,1 · 4,9 -4,8 = 0,1 · 1,405 = 0,1405 m.

The pipe with a nominal diameter d = 0,15 m has been accepted. For this pipe dia­meter the local resistance equivalent will be l z = 34 m. The amount of refrigerant being evaporated in the evaporator

Q 100000 G = - = -- = 302 kG/h r 331,3 (:rt• µg . g ·3600)0.2 ( 4 . G )2 V g LI Pm = z . 0,184 ( 1 + lz)

4 . G 3600 :rt • 2.g d 4,8 {3,14 · O, 79 · 10-6 · 9,81 · 3600}0,2 LI

Pm = 4,9 · 0,184 (30 + 34) 4 • 302 { 4 . 302 }2 1,55

. 3600 · 3,14 2 · 9,81 · 0,15 4

,8 = 688 kG/m2

The calculated flow resistance causes an increase in the evaporation pressure from O, 7318 kG/cm2 to 0,8006 kG/cm • corresponding thus to an increase in the evaporation temperature from -40°C to -38,3°C.

At the accepted work conditions, laminar flow of the liquid phase would be expected from d ;:::: d1 v where

4 • (n - 1) · Q div = Rev • :rt • µ1 • g · 3600 r

4 · 5 · 1 · 10s

2 · 103 · 3,14 · 29 · l0-6 · 9,81 · 3600 · 381,3

div = 0,93 m.

REFERENCES

r. R. C. Martinelli, ]. A. Putnam and R. W. Lockhart, Transactions of American Institute of Chemical Engineers. Vol. 42. 681-705 (r946).

2. R. W. Lockhart and R. C. Martinelli, Chemical Engineering Progress. Vol. 45. 39-48 (1949).

3. G. E. Alves, Chemical Eng. Prog. Vol. 50 nr. 9 (r954).

4. I. E. ldeltschik, Sprawotshnik po gidrawlitsheskim soprotiwlenjem (r960).

523

111-16

DISCUSSION

L. Mattarolo, Italy : Whilst there is obviously great merit in the work of Martinelli and Lockhart on the pressure drop, in refrigerant circuits, this work was carried out some time ago, and one should not forget that during recent years a lot of more work has been done on pressure drops due to flow between two phases especially in the field of applied nuclear physics. Therefore, I feel that one should take into account not only the work of Martinelli and Lockhart but the work that has been done since then, among others the flow through capillary tubes, although the flow may be very small due to the very small diameters. The pressure drop will be very much higher due to the presence of oil which again will have an effect on this issue.

K. Gutkowski, Poland : I feel that the work, which has been done by Lockhart and Martinelli, is still valid as recently as 1962, and one was after all mainly concerned with boiling refrigerants for refrigerating purposes. They had therefore concentrated on this particular fundamental work. As far as the second part of the question is concerned, capillary tubes were really not considered, since they were talking about refrigerant return line calculations in pump recirculation systems, and have not gone into this matter, which of course is an entirely different case.

L. V ahl, Nether lands : This paper should really be regarded as a first step and from it curves and data could be developed for design engineers, who in any case would not have time to go through the complicated formula from first principles in each case. Furthermore, this particular paper deals with horizontal pipes only.

K. Gutkowski, Poland : It is quite correct to say that the paper deals only with hori­zontal pipes, but Martinelli has subsequently proved that these formulae are valid for pipes running at an angle of up to 30°.

524

Conception des evaporateurs d'ammoniac des grands entrepots frigorifiques polyvalents

Design of Ammonia Evaporators for Large Multi-Purpose Cold Stores

G. SAINT-GIRONS

III-34

Compagnie des Entrepots et Gares Frigorifiques, 42, rue du Louvre, Paris (ler), France

SUMMARY. It is necessary to provide satisfactory internal wetting of evaporator pipes to obtain good output and this may be achieved via a moderate coefficient of recirculation.

A study of the effect of flash gas and heat load variations on liquid level stability and liquid trans! er in the difjerent distribution systems via natural overflow or by a recirculation pump or by direct injection, with some special precautions or measures suggested for each case.

Generally speaking, it is advantageous to limit the ammonia charge in evaporators as much as possible and to oversize liquid separators.

I - Les etudes des auteurs cites en reference bibliographique ont mis en lumiere les trois conditions fondamentales auxquelles doivent repondre les evaporateurs d'ammoniac des Grands Entrepots Frigorifiques Polyvalents : - rendement maximal - alimentation reguliere - fonctionnement stable sans transfert de liquide en direction des compresseurs

1 1 RECHERCHE DU RENDEMENT MAXIMAL

Les conclusions de ces auteurs coincident sur quatre considerations relatives au rende­ment des evaporateurs :

- les batteries d'echangeurs representent une part preponderante du poids et par conse­quent du prix de !'installation, aussi est-il d'une grande importance d'ameliorer leurs performances

- dans les batteries utilisees pour le refroidissement de l'air, le coefficient global de transmission est limite par le coefficient cote exterieur et depend par consequent peu du coefficient d'echange interieur cote ammoniac

- par contre, les performances des echangeurs sont proportionelles a la surface interieure de tubes effectivement «mouillee » par I' ammoniac liquide

- il en resulte que le taux de recirculation de !'ammoniac doit etre fixe a une valeur moderee, suffisante pour obtenir un mouillage complet, mais assez faible pour limiter !'elevation de temperature due aux pertes de charge dont !'influence est d'autant plus sensible que le regime d'evaporation est plus bas. D'apres les experiences de D. D. Wile, ce taux de recirculation do it etre de I' ordre de 4.

12 - MODE D'ALIMENTATION

II resulte egalement de ce qui precede que les evaporateurs a injection directe alimentes par detendeurs thermostatiques a controle de surchauffe ont un rendement bien inferieur a celui des appareils a recirculation artificielle OU naturelle.

La proportion de surface mouillee dans les systemes a detendeurs thermostatiques est diminuee :

- par !'absence de liquide dans la zone de surchauffe

- par la faible proportion de liquide dans la zone d'evaporation

- par les inegalites de distribution entre nappes paralleles

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Ces trois inconvenients sont supprimes dans les evaporateurs a recirculation mais le Professeur Lorentzen souligne les irregularites d'alimentation dues a !'introduction des gaz de detente dans les separateurs de liquide. I1 en resulte un regime cyclique avec variations de niveaux tres superieures a la fourchette des appareils de reglage et de ce fait les detendeurs modulants agissent en realite par tout ou rien.

13 - STABILITE DE FONCTIONNEMENT ET PRECAUTIONS CONTRE LES TRANSFERTS DE CHARGE D'AMMONIAC

Les consequences des irregularites d'alimentation signalees ci-dessus sont d'autant plus graves que la charge de liquide est plus elevee. I1 faudra done d'une part limiter la capacite des evaporateurs en utilisant des tubes a ailettes de petit diametre et d'autre part reduire la densite moyenne du fluide, ce qui revient a moderer le taux de recirculation comme deja indique au paragraphe 1 1 a propos de !'arbitrage entre coefficient de trans­mission et pertes de charge.

D'apres D. D. Wile la circulation par pompe presente a ce point de vue un avantage decisif sur le regorgement naturel. Elle permet d' employer des tubes de plus petit diametre avec un gain de rendement suffisant pour compenser le supplement de prix de revient par unite de surface et avec une double diminution de la charge d'ammoniac, par reduction de capacite et de densite.

Le Professeur G. Lorentzen mentionne les precautions a prendre dans le calcul des dimensions respectives des evaporateurs et separateurs pour eviter les transferts de char­ges et conclut que pour assurer une modulation progressive du debit, les regleurs a flot­teur doivent avoir une course superieure a la modification de niveau entrainee par la varia­tion de production frigorifique correspondante.

Quant aux auteurs russes, leur systeme a autocirculation en cascade constitue essen­tiellement un moyen de reduire au minimum la charge des evaporateurs en ammoniac de fa<;:on a permettre leur vidange totale a l'arret dans des reservoirs de grande capacite.

II-Nous nous proposons de reprendre et developper quelques- unes des considerations ci-dessus, apropos de chaque systeme d'alimentation des evaporateurs, en fonction de not­re propre experience d'installation et d'exploitation d'une vingtaine d'Entrepots Frigori­fiques Polyvalents. I1 y a lieu de noter que tous les etablissements consideres comportent des circuits frigorifiques multiples a differents regimes de temperatures mais avec circuit haute pression commun a tous les compresseurs. De ce fait, une masse de liquide importan­te est rassemblee en amont des organes de detente, et c'est a ces derniers qu'il appartient d'en assurer la repartition convenable aux differents postes d'utilisation du froid.

21 - SYSTEME A REGORGEMENT NATUREL

Ce type d'installation est utilise pour l'alimentation individuelle des evaporateurs en particulier dans le cas des appareils ou tunnels de congelation.

@ @

- 1 ' I '

1_ - - - - - - - - - - - - - _ ..) Fig. I . Alimentation mixte

CD

1 - Evaporateurs, 2 - Liquide HP, 3 - Aspiration compresseur, 4 - Arrivee gaz chauds, 5 - Re· tour vers reservoir de degivrage, 6 - Detendeur thermostatique, 7 - Niveau contact de securite

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III-34

II semble que la disposition suivant schema Fig. 1 avec injection directe d'une par­tie du liquide d'alimentation et par consequent des gaz de detente a l'evaporateur presente d'importants avantages par rapport au regorgement integral :

le phenomene de pompage du a I' introduction des gaz de detente dans la separateur est supprime et une repartition convenable de !'injection entre evaporateur et separa­teur permet d'obtenir une modulation stable de l'alimentation aussi bien par detendeur thermostatique que par regleur a flotteur. Dans le cas d'emploi d'un niveau contact par tout OU rien, le reglage de repartition de !'injection permet d'effectuer Un arbi­trage entre la hauteur des variations de niveau et la frequence des coupures

- la densite moyenne du fiuide est diminuee, ce qui reduit a la fois la charge d'ammoniac et la variation de cette derniere en fonction de la production frigorifique.

A noter par contre, un inconvenient inherent au regorgement nature! et qui subsiste avec le dispositif propose. II s'agit de !'engorgement a l'arret et aux tres faibles charges thermiques d'ou resulte un exces de liquide dangereux au moment du retour a pleine puissance. C'est le cas en particulier des appareils et tunnels de congelation recharges en produits chauds apres avoir termine le cycle precedent a faible charge et basse temperature. II est done necessaire :

de disposer de separateurs surdimensionnes de couper l'alimentation de liquide avant la fin de chaque cycle de congelation et, dans le cas des tunnels, d'asservir electriquement cette alimentation au fonctionne­ment des ventilateurs.

Nous avons implicitement considere jusqu'ici un circuit simple compose d'un seul groupe separateur-evaporateur.

Si un meme compresseur travaille a la fois sur plusieurs groupes separateurs-evapora­teurs a controle d'alimentation individuel, la stabilite de fonctionnement est amelioree, comme le fait remarquer le Professeur G. Lorentzen, quand il s'agit par exemple d'un cer­tain nombre de chambres froides dont !'ensemble presente une forte inertie thermique. Par contre, les risques de transfert de charge sont aggraves s'il s'agit de plusieurs appareils et tunnels de congelation. En effet, apres remplissage de l'un d'entre eux en produits chauds, la temperature generale d'evaporation remonte et les autres tunnels travaillant a charge reduite, emmagasinent une quantite excessive de liquide qui refluera plus ou moins brutalement dans les separateurs individuels au moment du retour au regime nor­mal d'evaporation. De plus, l'evaporateur du tunnel chaud absorbe a lui seul pendant un certain temps la presque totalite de la puissance du compresseur et travaille par conse­quent en forte surcharge, d'ou necessite :

de surdimensionner ces separateurs

- de reduire la puissance du compresseur pendant la mise en regime d'une cellule chaude

- de couper l'alimentation de liquide des autres cellules pendant le meme temps

Si, au lieu d'atteler sur un meme circuit, plusieurs groupes separateur-evaporateurs a controle individuel on branche tous les evaporateurs sur un separateur de liquide gene­ral, on retrouve, dans le cas ou il s'agit de chambres froides, la dimension classique des Entrepots a regorgement integral . . . et ses inconvenients [ 6] .

S'il s'agit d'appareils ou tunnels de congelation, cette disposition n'est a utiliser qu'avec precaution en raison des difficultes de fonctionnement en parallele de cellules a temperatu­res differentes et en particulier des difficultes d'amor<;:age de chaque cellule apres charge­ment de produits chauds.

22 - SYSTEME A CIRCULATION PAR POMPE

Ce type d'installation connait une vogue croissante en raison des avantages qu'il presente comme on l'a deja vu plus haut et de l'excellente repartition d'ammoniac qu'il permet d'obtenir meme dans les appareils les plus eloignes.

Cependant, appliquee par definition a des circuits ramifies done pouvant representer de grandes puissances frigorifiques et de grandes capacites volumetriques, la circulation par pompe est exposee aux irregularites d'alimentation decrites par le Professeur G. Lo­rentzen.

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Comme l'importance du transfert de liquide des evaporateurs vers le separateur general, provoque par l'interruption des gaz de detente, est fonction directe du volume des eva­porateurs et de la densite moyenne du fluide, qu'ils contiennent, on voit qu'on peut assister a des pulsations considerables du niveau de liquide.

11 sera done necessaire de surdimensionner les separateurs et on trouve aussi une nouvelle raison de reduire la charge d'ammoniac, c'est-a-dire diminuer le diametre de tube des evaporateurs et limiter a une valeur raisonnable le taux de recirculation.

L'injection d'une partie des gaz de detente directement dans les evaporateurs comme preconise au chapitre precedent ne peut etre envisagee comme une solution satisfaisante puisqu'elle ferait perdre a la distribution par pompe l'average de sa simplicite, aussi sem­ble-t-il que la meilleure solution consiste a sous-refroidir intensement le liquide avant in­jection au separateur de fac;:on a supprimer la formation de gaz au moment de la detente (Fig. 2).

Fig. 2. Distribution par pompe a partir d'un separateur de liquide general r - Serpentin de sous-refroidissement du liquide HP, 2 - Liquide HP, 3 - Aspiration compres­seur, 4 - Vers Jes stations de reglage Fig. 4, 5 - Niveau contact d'alimentation, 6 - Niveau contact de securite, 7 - Alarme

A propos de la circulation par pompe, une controverse subsiste au sujet de l'alimen­tation des evaporateurs par le haut OU par le bas.

Le premier systeme a ses defenseurs en raison de la reduction de charge d'ammoniac et de l'economie d'appareillage qui en resulte et aussi de la simplicite du degivrage par gaz chauds apres vidange totale des appareils par gravite (Fig. 3).

- ®

@

Fig. 3. Alimentation des evaporateurs par le haut r - Evaporateurs, 2 - Liquide HP en provenance des condenseurs, 3 - Aspiration compresseur, 4 - Arrivee gaz chauds, 5 - Reservoir de liquide BP, 6 - Reservoir de liquide HP, 7 - Niveau contact d'alimentation, 8 - Niveau contact de securite, 9 - Alarme

528

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Nous ne sommes pas, pour notre part, partisan de ce systeme. En dehors de la perte de rendement dont fait etat le Professeur G. Lorentzen, nous considerons en effet que :

- la vidange par gravite impose des sujetions d'installation inacceptables si l'on se rapelle qu'un des avantages de la distribution par pompe par rapport au regorgement naturel consiste justement a supprimer ces sujetions de gravite.

- les transferts de charge par vidange des evaporateurs a l'arret portent sur la totalite de !'ammoniac contenu dans !'installation et necessite, dans les Entrepots de quelque importance, des reservoirs de capacite prohibitive d'une part au niveau du separateur de liquide basse pression (vidange a l'arret), d'autre part cote haute pression, en aval des condenseurs (retention de liquide tant que le separateur basse pression n'est pas revenu a son niveau bas)

- la remise en circulation de toute la charge a la fois, a chaque demarrage, excede le debit de l'organe de reglage et provoque le desamorc;:age de la pompe OU le declen­chement de sa securite a niveau bas.

En ce qui concerne le systeme a alimentation par les bas, nous ne sommes pas parti­san de la formule qui consiste a munir l'arrivee de liquide d'une eletrovanne et a suppri­mer (par economie) celle de !'aspiration. En effet, dans ce cas, !'ammoniac continue a s't�vaporer dans les batteries des chambres qui sont coupees par leur thermostat, avec risque de gel des marchandises les plus proches s'il s'agit de chambres a temperature posi­tive et d'autre part la distillation de cette ammoniac entraine un transfert de charge vers les reservoirs haute et basse pression qu'il devient necessaire de surdimensionner comme dans le cas d'alimentation par le haut.

Nous utilisons quant a nous le schema suivant Fig. 4 avec electrovanne d'aspiration doublee d'une soupape de surete et clapet de retenue sur l'alimentation, ce dernier de preference a une electrovanne dont le fonctionnement ne serait pas satisfaisant puisque, dans les modeles normaux, ces electrovannes ne peuvent etre fermees que sous l'effet d'une contre-pression amont. Le schema figure egalement les electrovannes de commande a distance du degivrage par gaz chauds.

CD

Fig. 4. Station de reglage pour alimentation des evaporateurs par le bas I - Evaporateurs, 2 - De la distribution Fig. 2, 3 - Electrovanne Aspiration, 4 - soupape de surete, 5 - Clapet de retenue, 6 - Arrivee gaz chauds, 7 - Retour du liquide de degivrage

23 - SYS TE ME MIXTE D' ALIMENTATION

Nous avons expose anterieurement [6] le probleme souleve par l'equipement de grandes chambres refroidies par plusieurs conditionneurs diffuseurs repartis sur toute la surface de la chambre, quand il s'agit d'extensions d'etablissements anciens ou il existe deja des circuits generaux a injection directe OU a regorgement, Sur lesquels les nouveaux eva­porateurs doivent etre raccordes, sans qu'il puisse etre question d'une transformation generale, pour mise en circulation par pompe, comme ce serait le cas dans un Entrepot neuf.

11 est rappele egalement qu'a notre avis, la presence de robinetteries et appareillages d'ammoniac a l'interieur des chambres doit etre prohibee, ce qui interdit le controle des evaporateurs par separateurs individuels et conduit a regrouper les stations de reglage a l'exterieur des chambres a portee du personnel de conduite, mais hors d'acces du personnel de manutention.

529

IIl-34

Fig. 5. Alimentation serie

A - Station de reglage exterieure, B - Chambre froide, r - Conditionneurs-diffuseurs, 2 -Li­quide HP, 3 - Retour aspiration, 4 - Arrivee gaz chauds, 5 - Retour vers reservoir de de,6vrage, 6 - Detendeur thermostatique a chauffage electrique

Nous avons utilise a plusieurs reprises la disposition suivant schema Fig. 5 avec un separateur de liquide propre a la chambre et regorgeant sur les evaporateurs les plus proches tandis que les appareils eloignes sont alimentes «en serie )) par injection directe.

On remarque que ces derniers sont alimentes avec un taux de recirculation de deux, trois ou plus suivant le nombre total de conditionneurs de la chambre, au benefice de leur rendement.

L'injection des gaz de detente directement dans les evaporateurs stabilise le niveau de liquide au separateur comme dans le cas expose au paragraphe 21 et le fonctionnement s' est av ere tres satisfaisant.

Un autre dispositif simple de recirculation destine a ameliorer le rendement des evaporateurs a injection directe suivant schema Fig. 6 a aussi donne de bons resultats. Son principe est analogue a celui qui fait l'objet d'un brevet danois rapporte aux reunions de 1'1.1.F. a Cambridge en 1961 [7]. Dans cette application toutefois, la ligne d'aspiration des evaporateurs passe par un piege a liquide avant d'aboutir aux compresseurs.

Fig. 6. Alimentation par injection directe avec recirculation interieure A - Station de reglage extfrieure, B - Caisson du conditionneur-diffuseur, r - Evaporateurs, 2 - Liquide HP, 3 - Retour aspiration, 4 - Arrivee gaz chauds, 5 - Retour degivrage vers piege a liquide, 6 - Dctendeur thermostatique, 7 - By-pass de recirculation interne

III - CONCLUSION

On voit !'importance des precautions a prendre pour obtenir un fonctionnement satisfaisant des evaporateurs utilises comme refroidisseurs d'air dans les Grands Entrepots Polyvalents. 11 est necessaire de leur assurer un bon rendement etant donne leur incidence

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sur les prix d'installation. Dans ce but, le taux de recirculation doit etre de l'ordre de 4, suffisant pour realiser un « mouillage » convenable des tubes mais limite a cette valeur pour eviter une perte de charge excessive.

Les gaz de detente et variations de charge thermique engendrent des irregularites d'alimentation avec risques de transferts de liquide vers les separateurs. Certaines disposi­tions speciales permettent d'ameliorer la stabilite des niveaux et prevenir les consequences des changements de regimes dans les divers types d'installation a regorgement, circulation par pompe ou alimentation mixte.

Dans le cas de la distribution par pompe, ii est preferable d'alimenter les evaporateurs par le bas plut6t que par le haut.

D'une fa�on generale, ii est toujours souhaitable de donner des dimensions tres larges aux separateurs de liquide et ii y a interet a reduire par tous les moyens la charge d'ammo­niac de !'installation.

REFERENCES

1. ]. Lorentzen, I.I.F. Commission II, Copenhague 1 959.

2. G. Lorentzen, I.I.F. Commission III, Moscou 1958.

3. D. D. Wile, I.I.F. Commission III, Washington 1962 .

4. S. Kobulashvili, I.I.F. Commission III, Paris 1955·

5 . M. Filinov, I.I.F. Commission III, Moscou 1958.

6. G. Saint-Girons, I.I.F. Commission III, Copenhague 1959·

7. A. Neuenschwander, I.I.F. Commission III, Cambridge l96 r .

DISKUSSION

A. Stradelli, Italy: The question examined by M. Saint-Girons must be studied and resolved for each individual case as one cannot give general rules. Only one rule is valid namely that one always wants to reduce to the minimum the number of evaporators. This conclusion is contrary to that adopted by most refrigeration engineers.

G. Saint-Girons, France: Of course, although this is a valid argument, there is a limit to the size of cold store which can be served by one evaporator, and the air distribution will therefore suffer unless the evaporators are placed in such a position as to give com­plete and equal distribution of air and temperature in large stores. In these stores it will therefore be necessary, of course, to increase the number of evaporators.

L. V ah!, Nether lands: What is the purpose of the heat exchanger coil in Fig. 2 ? I cannot quite see the reason for having i t at all.

G. Saint-Girons, France: It is correct that this coil is utilized for the cooling of the high temperature liquid within it, a quantity of the low pressure refrigerant thereby for­ming a certain amount of vapour. However, due to the subcooling of the liquid the amount of vapour would not be excessive. Therefore, it is not a question of the quantity of vapour produced by this coil, but the variation of this quantity due to the modulation in the feed requirements which result. It is supposed to act in a cycling system, and this device is proposed to even out the variations which occur.

This system has not yet been tried out experimentally, and it will be necessary to determine a convenient way in which the heat exchanger surface and the volumetric capacity of the circuit can be related, to be able to report on the stability of this condition.

531

III-31

Utilization of Steel Stamp-Welded Panels for Heat Exchanging Ap­paratus in Refrigerating Plants

L'emploi des panneaux d'acier etampes-soudes pour les echangeurs thermiques des installations frigorifiques

A. I. SHUVALOV "Kompressor" Works, Moscow, U. S. S. R.

SOMMAIRE. Jusqu'a maintenant on fabrique les appareils d'echange thermique pour les installations de moyenne et de grande puissance frigorifique avec des tubes. Quel que soit leur type de construction (a calandre multitubulaire ou a calandre et serpentin) la fabrication et l' exploitation de ces appareils sont assez couteuses et compliquees, a cause de leur en­combrement et de leur poids.

Une des nouvelles tendances pour le reduction du poids et de l'encombrement ainsi que pour !'intensification des appareils d'echange thermique est !'utilisation des panneaux d'acier etampes-soudes comme surface d'echange de chaleur au lieu de tubes.

On a developpe a l'usine <<Kompressor» de Moscou les constructions d'evaporateurs et ensuite de condenseurs dans lesquelles les panneaux d' acier etampes-soudes ont ete utilises pour /es installations frigorifiques a ammoniac d'une puissance de 100000 kcal/h et plus.

Les mode/es experimentaux des appareils d' echange thermique ont subi les essais thermo­techniques a l'Institut de recherches scientifiques de l'industriefrigorifique d'URSS a Moscou et sont recommandes pour la production en serie.

On peut classer /es appareils a panneaux en deux types principaux suivant les carac­teristiques de construction:

1. /es appareils a panneaux du type ferme, OU le fiuide chaud circule sous pression; ils sont destines a remplacer /es appareils a calandre multitubulaire et a calandre et ser­pentin.

2. Les appareils du type ouvert avec circulation libre du fiuide froid assuree par la pompe -agitateur a helice.

Les appareils du premier type peuvent etre utilises dans /es installations f rigorijiques comme condenseurs et sous-refroidisseurs a contre-courant, et aussi comme evaporateurs. Les ap­pareils du second type peuvent etre utilises comme evaporateurs du type ouvert, accumula­teurs d'eau glacee, generateurs de la glace en plaques.

Le rapport decrit le principe de travail, la construction des appareils nouveaux et les resultats de leurs essais.

New designs of heat exchangers with a heat transfer surface area from 20 to 320 m2 have been developed at the Moscow "Kompressor" Works for ammonia refrigerating plants as a result of long term designing search and experimental work. These heat exchanging apparatus can compete successfully with refrigerating plants manufactured of pipes.

Two main types of heat exchangers have been developed :

1. Closed packaged panel condensers, evaporators and vapour coolers, in which the heat transfer medium is under pressure. They are designed for replacing shell-and-tube and shell-and-coil apparatus.

2. Open panel evaporators, hold-over apparatus and ice makers with free circulation of the heat transfer medium and built-in agitators. They are designed for replacing the existing sectional apparatus of the immersion type.

The main element of the new heat exchangers is the stamped panel, which is unified for all types and sizes of apparatus (Fig. 1).

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11-11 B c

-� . . . . . �l t..._ L_ B C 420-z

c-c

�((((({(sssssssss\sssssss?�-Fig. r. Standard panel

The panels are manufactured as follows. Steel sheets are stamped according to the given profile and welded in pairs by contact spot welding. The edges of the sheets are welded along their long side by contact seam welding. Boiling and condensation of the refrigerant is effected in canals, formed by the welded sheets. The webs between the canals serve as fins.

Packaged panel condensers and evaporators

The condenser (Fig. 2) is a package assembled of individual sections 1 . The heat transfer surface of the section is formed of stamped panels. The package is closed on both sides with flat covers 2, which are secured on the perimeter with pins 3. Each section consists of a rectangular frame 4 welded of pipes and with welded panels. The latter are arranged in such a manner that a passage 5 remains at one end of the frame

6

Fig. 2. Schematic view of panel condenser

534

r - panel, 2 - covers, 3 - pins, 4 - frame, 5 - water passage, 6 - headers, 7 - ammonia inlet, 8 - liquid ammonia outlet, 9 - water inlet, ro - water outlet.

IIl-31

for the coolant (water, brine). The internal cavities of the sections are connected in parallel by means of top and bottom headers, which consist of separate links 6 that are welded to the frame pipes extending from the package.

Flat rubber gaskets are fitted between the sections and covers along the contours of the tubular frames.

The apparatus can be made multi-pass in order to reduce the hydraulic resistance. The rigidity of the apparatus, operating at increased pressures of the coolant, is provid­

ed by fitting additional bracing pins in the middle part of the package. The packaged panel condenser described herein can be employed as a closed evapora­

tor. The liquid ammonia is supplied in this case to the bottom header, while the vapour is sucked out via the top header to the liquid separator.

Packaged panel apparatus are lighter and more compact than shell-and-tube ones. The manufacture of these apparatus can be completely mechanized. The apparatus are easily disassembled for inspection, repairs and application of anticorrosive coatings. The packaged panel apparatus are more safe than shell-and-tube ones owing to their smaller ammonia charge.

PACKAGED PANEL SUBCOOLERS

Packaged panel condensers can be used in the capacity of liquid refrigerant subcoolers. The required sequence of ammonia passes is provided by fitting plugs in the ammonia headers.

Small panel condensers can be combined with subcoolers. The first sections in the water flow will fulfil in this case the role of subcoolers.

PANEL EVAPORATOR

The open panel evaporator is a heat exchanging apparatus, consisting of a rectangular metal or concrete tank with the evaporator sections assembled inside. A propeller agi­tator is assembled in the tank for agitating the coolant.

Each section (Fig. 3) consists of two horizontal collecting pipes 1 and two vertical standpipes 2. The panels 3 are welded to the horizontal collecting pipes. The sections are fitted in parallel to the headers for supplying liquid ammonia, sucking out the vapour and draining the oil.

5 f

Fig. 3. Open panel evaporator section r - header pipes, 2 - standpipes, 3 - panel, 4 - oil drain, 5 - ammonia vapour outlet, 6 - liquid ammonia inlet.

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HOLD-OVER APPARATUS

Open panel evaporators can be used in refrigerating plants as hold-over apparatus. The storage of refrigeration is achieved by accumulating ice on the sections. About 3500-4000 kcal of refrigeration can be accumulated per 1 m2 of surface area.

The utilization of such apparatus instead of plants for direct cooling of water improves the energy balance of the enterprises, because the accumulation of ice can be performed during the hours of minimum power consumption.

DIRECT EVAPORATION PANEL ICE MAKERS

A somewhat modified design of the panel sections, providing for rapid separation o the ice after freezing on, is used for developing a direct evaporation ice maker for the production of plate ice. The evaporator section is made for this purpose without a drier and equipped with an insulating tubular frame, which is welded to the section along its perimeter.

The frame, the temperature of which equals that of the water in the tank, does not freeze up and prevents the freezing together of the ice plates, forming on both sides of the panel.

Rapid separation of the ice plates from the section surface, as well as from its peri­phery is provided on supplying hot refrigerant vapour simultaneously to the section and into the insulating frame.

The weight of the metal per ton of production is 6-7 times less in a panel ice maker than in a can ice maker. Less production floor space is required for the ice making equipment.

'n - L - 1 ! • ->-:::: . I 0 -- --+ - · .

. - ,,- 2 --� I-•n . .. --

- - _,. ..... - 0 0 0 3--10 "".'.'.'. I-"

. - ' -1 - - ----- f---� ,_ --'- -1 -

30 0 - ---- -- � - ·

- � � -- ---·

0 20 10 0

1000 201 70 JliOO 3500

Fig. 4. Dependence of overall coefficient of heat transfer of evaporators from heat load at evapo­rating temperature of -r5°C. r - clean vertical-tube evaporator at brine velocity Vb = 0.49 m/sec; 2 - panel evaporator at Vb = 0.55 m/sec; 3 - vertical-tube evaporator with oil film at Vb = 0.49 m/sec.

The overall coefficients of heat transfer for panel apparatus are given in Figs. 4 and 5. The thermal characteristics of panel, shell-and-tube and vertical-tube apparatus are almost similar.

536

'O 2oa

0

/ ... -

� � I ,_. 1:1Lr ' " 100

...... - -

as

, . ......

v / ...

i..- .... -i.. / � � -

� :;.....-� -L....--'-

i- --- -

,_-

...... _ ...

- ... ......

-

-l...--

-

I/ / -

4

" ..... .,,

2 -

...... -

1---l t-I-ii

......

5 - '- -

... - ......

� ,,.. ·

,... ..

- -

......

0,7 0,8 0,9 1,0

.,, _ ..

.,,

.... .. - ......

-

a ' 'il.. �1

_ ,. J

- - --- -

:::r-< -"-�3

1.1 1,2 Yw m/sec

IIl-31

Fig. 5. Dependence of overall coefficient of heat transfer of condensers from cooling water velocity r - panel condenser : "a" at q = 4000 kcal/m• hr; ,,b" at q = 52r2 kcal/m2 hr; "c" at q = 6435 kcai/m2 hr; 2 - multishell r4-tube condenser at q = 4500 kcai/m2 hr; 3 - multi­shell 7-tube condenser after lengthy operation at q = 4250 kcai/m2 hr; 4 - standard shell­and tube condensers with clean tubes; 5 - ditto after lengthy operation.

CONCLUSION

The small consumption of metal, compacmess and simplicity in the manufacture of panel heat exchangers provide for their widespread utilization. They can be used not only in refrigerating plants, but also in other fields of engineering for the supply or transfer of heat during the processes of boiling, condensation, vapourization when two mediums participate in the heat exchange process of which one may be under consid­erable pressure, is toxic or explosive, and the other is under a pressure of 4 to 6 atm.ga. and is neither toxic, nor explosive.

537

IIl-4

Some Practical Tests on the Efficiency ofExtended Surface Air Coolers

Quelques essais pratiques sur le rendement des refroidisseurs d'air a surface augmentee

EINAR BRENDENG Norges Tekniske Hogskole, Trondheim, Norway

SOMMA/RE. Le rendement des refroidisseurs d'air a surface augmentee, munis de de­tendeurs thermostatiques et fonctionnant au R 12, est nettement influence par la resistance a la transmission de chaleur entre le frigorigene en evaporation et la paroi du tube. Ce type de transmission de chaleur a ete etudie par plusieurs chercheurs, mais les resultats n'ont pas ete ernierement adoptes par les ingenieurs charges de la conception des refroidisseurs d' air. Au con­traire, les distributeurs de liquide sont utilises sans discernement et, meme si une division de l' ecoulement de frigorigene dans un certain nombre de sections de tube paral/eles reduit la perte de charge, il pourrait en resulter un abaissement du rendement du refroidisseur d' air, en raison de !'augmentation de la resistance a la transmission de chaleur entre le frigorigene et la paroi du tube.

Un certain nombre de refroidisseurs d' air classiques ont ete essayes par la Section du Froid de !'Ecole Polytechnique de Norvege. Le coefficient de transmission de chaleur pour lefrigorigene en evaporation a ete mesure et !es resultats concordaient assez bien avec !es resultats des recherches obtenus avec des evaporateurs d' essai de laboratoire etudies par d' autres chercheurs. On examine I' influence de la longueur du tube, du taux de transmission de chaleur et de la perte de charge sur le coefficient global de transmission de chaleur.

INTRODUCTION

Extended surface air coolers, equipped with thermostatic expansion valves and working with refrigerant R 12, are extensively used in commercial refrigeration plants. The per­formance of this type of air cooler is greatly influenced by the heat transfer resistance between the evaporating refrigerant and the tube wall, and although this type of heat trans­fer has been investigated by several research workers, the manufacturers of air coolers seem reluctant to fully adopt the test results in practice. In some cases liquid distributors have been used uncritically, and even if a division of the refrigerant flow into parallel tubes re­duces the pressure drop, a decrease in the performance of the air cooler might be the result, due to increased heat transfer resistance between refrigerant and tube wall.

In order to confirm the heat transfer coefficient for evaporating R 12, found by other workers on laboratory type evaporators, a number of air coolers have been tested at the Institute of Refrigeration at the Technical University of Norway.

AIR COOLERS TESTED

A survey of the air coolers tested is given in Table I . The air coolers were of conventional design with aluminium fins on expanded copper

tubes. Plain tubes were used in coolers A-E, while cooler F was equipped with % • diam. inner tubes with inner fins in the annulus between the inner and outer tubes.

Air coolers C, D, E are identical, except for the number of parallel runs.

TEST APPARATUS

The test apparatus is shown schematically in Fig. 1. The compressor with watercooled condenser is placed inside an insulated box, and the temperature difference over the walls is measured with thermocouples and kept at zero by means of an additional cooler.

The liquid content in the refrigerant vapour at the outlet of the evaporator influences the heat transfer in the last part of the evaporator. By means of an electrical heat exchanger,

539

III-4

Table I. Technical data of tested air coolers.

Total Inside Total Number Cooler Tube tube area area Ae/Ai Fin Tubes of

No. diam. length Ai Ae thickness in straight m m2 m2 mm parall. tubes

A 5/ S H 12 0.561 8.53 15.2 0.7 1 24 B 5/ s

II 15 0.602 12.3 20.5 0.3 1 36 c 3 '· H 26 1 .401 33.2 23.7 0.75 4 24 D 3/ 4 H 26 1.401 33.2 23.7 0.75 2 24 E 3/ 4 II 26 1 .401 33.2 23.7 0.75 1 24 F S/4 "-a/ s " 9.5 0.472 19.2 41 0.3 3 12

1 . Bends included. 2 . Bends shielded from air flow, and excluded from calculated area.

the liquid drops are evaporated, and the quality of the refrigerant at the outlet of the evap­orator is calculated. In order to avoid heat leakage, the output of the outer heating element is regulated according to the temperature difference over the insulation.

d

r - - - - - ...,..- -, I •,) I !

I ' '

a

� b e

,.., ,..,

Fig. r. Test apparatus. a : Insulated box with compressor, condenser, DC-motor, and auxiliary cooler. b : Electrical heat exchanger. c : Ordinary heat exchanger. d : Air cooler. e : Meas­uring vessels.

The capacity of the air cooler is established through the total heat balance. As a control, the flow of refrigerant is also measured directly, in two vessels equipped with sight glasses. Tests on cooler B - F were run with oil separator, whereas no separator was used in the test of cooler A.

The temperature on the tubes was measured with thermocouples, soldered to the sur­face. The temperature of the fins was measured with butt-soldered thermocouples fixed in small holes, drilled in the fins.

540

III-4

Air temperatures and temperatures in the refrigerant lines were measured with thermo­couples and evaporating pressure and pressure drop in the evaporators with mercury manometers.

TEST RES UL TS HEAT TRANSFER COEFFICIENTS

Average values for the nominal overall heat transfer coefficient R, calculated according to the pressure at the evaporator outlet, the air velocity and the refrigerant velocity at the outlet of the evaporator tubes, and the air side !Xe and refrigerant side rx1 heat transfer coefficients are surveyed in Table II, for a nominal temperature difference of S°C.

Table II. Test results. Nominal temperature difference S°C.

Cooler k Air No. kcal/ veloc.

m2h° C m/s

A 7.3 3.0 B 15.6 7.1 c 3.S 2.S D 6 2.S E 9.S 2.S F 9.4 4.5

20

18

16

14

u 12 . .I:.

� 10 u .

r:1 ... u <I -" -" 8

6

4

2

00

ae kcal/

m2h° C

4S lS lS lS

800

Fin Refr. ai effic. �Xe veloc. kcal/

� m/s m2h°C

26 4.0 175 O.S2 40 5.5 600 0.77 14 1 .5 130 0.77 14 4.5 275 0.77 14 13.5 1500

17.0 3.5 1000

1600 2400 3200 o. kcavh

Evap. temp.

o c

-2S - S -2S -2S -2S -2S

Fig. 2 . Distribution of the various heat resistances for cooler B. a : Influence of pressure drop in tubes. b : Temperature difference between evaporating refrigerant and tube wall. c : Tem­perature drop in fins. d : Temperature difference between fins and air.

541

III-4

Since coolers No. C, D, and E are identical, except for the number of parallel tubes, the importance of a correctly designed refrigerant distribution to the evaporator is evident.

The relative importance of the various heat transfer resistances and the pressure drop in the tubes is shown for cooler B, in Fig. 2. The nominal overall heat transfer coefficient increases with increasing load up to the point where the pressur.e drop becomes more im­portant and causes a reduction in the k-value.

20

18

16 !;J ..C: 14 � 0 u

� 12

10

8

6

4

2

�-

0 2

B -------·-�� --.... ---"" 7

/fl' -·-- --- - --A / F � -./ -- ---�- - · · -E VJ

fl' / / /

V1 � / [ v / v

c

/ 4 8 10 12 14

At "C 16 18

Fig. 3 . The relation between overall heat transfer coefficient k and nominal temperature difference 6 t for the tested coolers.

In Fig. 3, the variation of the nominal k-value with the nominal logarithmic temperature difference is given for the tested coolers. In Fig. 4, the experimental values of evaporating heat transfer coefficient ai, tube diameter d, mass flow G and heat flux q are correlated as proposed by Pierre [1] .

The values of ai are calculated with the average tube wall temperature and the evap­oration temperature corresponding to the mean evaporating pressure Pm � p 0 + 2/3 6p. Po = pressure at evaporator outlet. "'P = pressure drop.

The heat transfer coefficient is dependant on the mass rate, the tube diameter and the heat flux. Further, the liquid content or superheat at outlet affects the result, but this influence is not significant in all the tests. Bottom feed is known to give higher heat trans­fer coefficients than top feed, and this condition may be the cause of the high values measured in air cooler No. E.

With liquid distributors, uneven refrigerant feed to the different sections may occur. This is the case with cooler No. C, as confirmed from the temperature measurements on the tubes.

Air cooler F is equipped with inner fins, and the calculated apparent heat transfer coefficients should be high when the fin area is not taken into account. The inner fins should also be expected to have influence on the refrigerant flow pattern and thereby on the heat transfer .

.542

5 0

0

0 >--I--- -c -

2-..............

1 -

I---0, --

-·-0

-c----

:

l0 � -

,_ --

• A � _ i..---1...-

,....,.. •

I•

- -

_,_ --

- �

0,1 0, 2 0,5

. " .. . ...

--�

.� -,. .

-

2

» ,,.,... ..-1 •• • v • • /� � la • . . .

-• -

0 b c , _ A 0 . ,._ f-

. ;: -B 0 f-

----- c "' � I D I> ..

-I I E v • 1� F O • I� 0

I I 5 10 20

�=W-·� I I I I I

50 · 106

III-4

Fig. 4. The relation between the evaporating heat transfer coefficient OC1 kcal/m 'h °C, internal tube diameter d m, mass flow G kg/h, and heat flux q kcal/m2h. Refrigerant superheated at

evaporator outlet : Symbols a, equation f, oc; d = 0.0145 ( G � q) 0'4 [r] Liquid drops in

refrigerant at outlet : Symbols b, equation e,-x;d= 0.034 • (G �_9) 0'5 [r] c : Arrangement

of refrigerant flow in the coolers. Cross section of tubes in cooler F shown. At lower loads,

i. e. � q P>! 0.35 · ro•, a transition in the flow occurs for cooler C, causing an increase in the

heat transfer coefficient.

PRESSURE DROP

The pressure drop in the tubes is dependent on the design, the heat load, the evaporat­ing temperature, and the oil content in the refrigerant. For a given cooler the pressure drop must be related to the evaporating heat transfer coefficient, Fig. 5. Cooler F, with inner fins has a comparatively low pressure drop, thus the inner fins seem quite efficient.

OPTIMAL DESIGN OF REFRIGERANT DISTRIBUTION

With a certain ratio between external and internal area, air film heat transfer coefficient, fin efficiency, and nominal temperature difference, there must be an optimum value for the heat load on each parallel section. In Fig. 6 the relation between the overall heat transfer coefficient and the number of parallel sections is calculated for a given cooler, using equa­tions for heat transfer coefficient and pressure drop from Pierre [1, 2] . The influence of a faulty designed refrigerant distribution is severe when the temperature difference is assumed to be constant. In Fig. 7 the same relation is given for cooler CDE, and here is also shown the temperature difference, overall heat transfer coefficient and capacity when the cooler is connected to a single compressor. These latter curves explain why the influence of an incorrectly designed distribution is not always detected in practice.

543

III-4

5000

2000 r:I.; ��l"C

1000

500

200

1 00

50

20 0.01

t:> ..

. .

.

c .

�I": � .. y ... •

v .. /

0.05 0.1

F E ...Qp�ll v �

,, A !/.':?

D /

0

a� ,/6

I-"� A � !/5.'.i--� 0

v"'

0.5 A12. .!59...!.'.m' v" err{/ kg

B " -

·- b

0

5

Fig. 5. The relation between 6 p/v" and a1. 6 p = pressure drop in tubes, kg/cm 2. v" = specific volume at evaporator outlet, m3 /kg. a1 = evaporating heat transfer coefficient, kcal/m 2h °C. The influence of a low Re · L/d seem evident for coolers C and F.

In Fig. 8, the optimal tube length of each section of 3 / 4 • dia tube is shown for an evapo­rating temperature at outlet of-30° C and-10° C. The optimal length depends on the de­sign and may be from about 15 m to 22 m, corresponding to a heat load of 1300 - 1600 kcal/h and a refrigerant velocity at the outlet of the tubes of 9-10 m/s. For a suction temperature of -10° C the optimal length is higher, from 20 m to 30 m, and the heat load 2200-2500 kcal/h with about the same values for the outlet velocity.

CONCLUSION

The values found for the heat transfer coefficient for evaporating refrigerant for labora­tory type evaporators may also be used in order to predict the performance of conventional extended surface air coolers. The distribution of the refrigerant flow should be carefully considered, and the mass flow in each tube should be high enough to give a satisfactory heat transfer without unduly high pressure drop.

ACKNOWLEDGEMENTS

Some of the measurements and calculations were made by Messrs. H. J. Kopstad, 0. Skjeggedal, R. Pettersen and K. Afiekt. Figures were drawn by K. Afiekt.

LITERATURE:

r. B. Pierre, Kylteknisk Tidskrift. No. 6, 1953, p. 76-Sr. 2. B. Pierre, Kylteknisk Tidskrift. No. 6, 1957, p. 225-238.

544

.c -

-

4 -

2 -

0 0

/ / / (

- - --),,,,,, -..... .....

..... ..... ...... ......

d =3/4" �t nom = 8 •c to = -30°C L = 40 m

...._ ' -.....

IIl-4

3000 $

0 u

.:.! •

0 2000

"o

1 000

ae �Ae/Aj = 300 kcaym2h oc

I I 0 2 3 N 4

Fig. 6. Calculated values for the overall heat transfer coefficient k and the capacity Qo for a given cooler, dependent on the number of parallel runs, N, at a constant nominal temperature difference of 8°C. Total tube length 40 m, nominal evaporating temperature -30°C, and ae · ; Ae/ Ai = 300. ae = air side heat transfer coefficient kcal/m 'h ° C, ; = fin efficiency, Ae/ Ai ratio between total and internal area.

545

III-4

�, I ! 12,5�---� ....._ ------0-V.-�----+1--t2500 \ t. -._ -!! ..!: canst. \ I <Ps ... --==-----� \Js.,::;. ;,;.-1 01-------0-\ -�-6-----+------ � 2000

� � l��A.,�� N E e/ \'- 'o<1t E o ------1-- -' --'"'� o 1f 7,5 c ' . v .. '0 -�---+--t1500 :::E. u ... '\. I ......._ ,

_:.:; <] , - , k , Vs = const. 0 u

.'<'.

0 _:.:: o, � .:::::: - -6 "' <It., I ......... I 5 I------+-------+-- ' e .c -o � ...... __ I

- 1000 °

- o

2,5>-----+----+-----+----+--t 500

0 o.___ __ __,_ ____ 2..__ __ __.3 ____ 4...____. 0 N

Fig. 7. Measured values for the overall heat transfer coefficient k and capacity Qo for coolers CDE.

546

When the nominal temperature difference is assumed constant 6 t = 8°C, a division of the refrigerant flow into parallel sections N results in a severe decrease in the performance of the cooler. The influence of a division of the flow on the performance is less severe when the air cooler is connected to a given compressor, Vs = constant.

1,0 1--------+---::::::� -=-=-""'=,,....--=----=·er-..- r _ __ ,

-- -- �'

� 40001------..__-----l------l------4 0 u � T O 0 0 '7' IJ +' II "'" 20 .:. � � 00 3000 _

___

_ _,_ _____ _.__

10 5

5 2,5

. ""-----'-------'------'------'- 0 0 10 20 30 L/N m 40

III-4

Fig. 8. The relation between tube length per parallel section, L/N m, and obtainable capacity per section, Qo/N kcal/h, for average conditions. The nominal temperature difference is 8 ° C , and the nominal evaporating temperature t 0 = -ro°C and -30°C. C = CGe 1; Ae/Ai = 300 and 600 kcal/m 2h ° C.

The efficiency 1) = !lf!I!:_ and the velocity of the refrigerant at the outlet of the tubes (Qo/ L) max,

are also shown.

DISCUSSION

L. Vahl, Netherlands: I am particularly impressed that the work has been done in much the same way as practical engineers would require the information. However, the K value which is given is not clear, since it does not state what surface area has been used.

E. Brendeng, Norway: I have used the external total surface of the cooler. L. Vahl, Netherlands: This should be made quite clear, since the surface used is most

important to the determination of the K value, and it is not stated in the paper. S. Touber, Netherlands: What is the heading which has been omitted from the sixth

column of table 2 in the paper ? E. Brendeng, Norway: The column of figures represents the product of fin efficiency

multiplied by the external heat transfer coefficient. L. Vahl, Netherlands: The method of measurement by putting the compressor and

condenser into an insulated box in order to get a heat balance is very original, and I should like to ask what is the experience of using this method over previously applied methods ?

E. Brendeng, Norway: We found that this particular method, although perhaps a little novel, is very simple to apply, and one can get a complete heat balance on the cooler without any difficulty. The normal methods of taking each separate reading is within a few per cent in fact of the method I have used for high heat loads, but the method which is described here, is very much more accurate for low heat loads.

547

111-30

Determination of Cooling Air Optimum Velocity, Arrangement and Area of Air-Cooled Condenser Surface

Determination de la vitesse optimale de l'air de refroidissement, des dimensions et de !'arrangement de la surface des condenseurs refroidis par l'air

D. M. IOFFE Scientific Research Institute of the Refrigerating Industry of the USSR, Moscow, u. s. s. R.

SOMMAIRE. L' accroissement de la vitesse d' air circulant au travers du condenseur abaisse la temperature de condensation, augmente la puissance frigorifique et la puissance frigori­fique specifique du compresseur incorpore dans un groupe frigorifique a air. En meme temps augmente la consommation d'energie pour l'entrainement du ventilateur. L'accroissement de la vitesse de l' air est rationnelle jusqu' a ce que l' economie de puissance electrique absorbee par I' entreinement du compresseur soit plus elevee que la consommation complementaire de puissance absorbee par le ventilateur.

La puissance frigorifique specifique etant determinee par les limites des temperatures d' evaporation du circuit frigorifique, la vitesse de l' air optimale pour les ensembles a basse temperature, a temperature moyenne et a temperature elevee n'est pas la meme.

La puissance absorbee par l' entrainement du ventilateur et la temperature de condensation dependent de !'arrangement de la surface d'un condenseur, a savoir du nombre des sections suivant la direction du mouvement d'air. On a demontre, que !'augmentation du nombre des sections avec la meme surf ace de condenseur contribue a la reduction de la consommation de la puissance pour le fonctionnement du ventilateur.

L'augmentation de la surface du condenseur augmente a son tour la puissance frigorifique et la puissance frigorifique effective. En meme temps, augmentent le cout initial de l' ensemble frigorifique, son encombrement et les depenses d' amortissement. On donne dans le rapport !'analyse de !'influence des dimensions de la surface du condenseur sur la consommation d'energie. On donne aussi les valeurs de la chaleur rejetee par 1 m2 de surface de condenseur telles que l' augmentation ulterieure de sa surface n' ameliore presque pas plus les caracteristi­ques techniques et ecomomiques de la machine.

L'emploi des valeurs de la vitesse d'air optimale, de la chaleur rejetee par 1 m2 de surface de condenseur et les recommandations sur !'arrangement de la surface pour le calcul et la construction des petites machines frigorifiques assure des meilleures caracteristiques energeti­ques et economiques.

The refrigerating capacity, performance factor and other characteristics of a condens­ing unit with an air-cooled condenser depend greatly on the velocity of the air, flowing over the condenser, the value of its heat transfer surface and arrangement of the latter (number of banks in the direction of the air flow).

An increase of the air velocity results in a reduction of the condensing temperature, increases the refrigerating capacity of the compressor and unit, as well as the performance factor of the compressor. Simultaneously, the power consumption by the fan motor increases. It is expedient to increase the air velocity until the decrease of electric power, consumed by the compressor, exceeds the additional power consumption by the fan drive.

An increase of the condenser surface area results also in an increase of the refrigerat­ing capacity and performance factor. However, an improvement of the given charac­teristics is followed in this case by an increase of the weight and overall dimensions of the unit, its cost and depreciation costs.

549

III-30

The characteristics of the unit depend also on the arrangement of the condenser sur­face. A change in the number of banks is of double effect. If the values of the air velocity and condenser surface area are unaltered, an increase in the number of banks results in an increase of the hydraulic resistance to the air flow and reduces its amount. The consumption of electric power by the fan drive reduces. However, the temperature of the air leaving the condenser and the condensing temperature increase which induces a decrease of the compressor performance factor.

The selection of the values of the condenser surface area, the number of banks and air velocity must be grounded on a study of the combined influence of these factors on the power and economic characteristics of the unit.

The following equations, characterizing the performance of the compressor, con­denser, fan and the unit as a whole are employed for deriving the required interdepen­dences between the power characteristics of the unit, air flow velocity, surface area and ar­rangement of the condenser :

Qc = F . K . 6J = C1 • Q K = C2 . C3 • (W . y) m h = C4 • ( W . y) n

Q = A - B . tc Kcom = D - E . tc

W . y . F . h . rp lVrm = ---------102 . y . 'Y/f . 'Y/fm . ib

1 lVrm ---- + -Kc om Q

(1) (2)

(3) (4) (5)

(6)

(7)

The following symbols are used in the equations: Qc - amount of heat transferred by the condenser; F and K - external heat transfer surface of the condenser and its coefficient of

Ca W y m h C4 and n A ; B; D; E Nim

f . ib f{J = ---p f 'Yjr and 'Yjrm Kcom

550

heat transfer;

6J = � - tar_

tc- ta1 2. 3 lg ---­tc- ta2

condensing temperature and temperature of air entering and leaving the condenser;

- refrigerating capacity of the unit ; - coefficient, depending on the design of the compressor, evaporating

temperature to and tc ; - coefficient, depending on the area and arrangement of the condenser

fins and pipes ; - coefficient, depending on the number of condenser tube banks h ; - air mass velocity in the narrow section of condenser; - exponent, depending on the design of the condenser; - hydraulic resistance of the condenser; - coefficient and exponent, depending on the design of the condenser; - factors, depending on to and the design of the compressor; - power, consumed by the fan motor; - coefficient, equaling the ratio of the narrow section to the heat transfer

surface of one bank; - narrow section of condenser, through which the cooling air flows; - efficiency of fan and its motor; - performance factor of the compressor, equaling the quotient from the

division of the refrigerating capacity by the power, consumed by the compressor driving motor;

III-30

Kc. u. - performance factor of unit, equaling the quotient from the division of the refrigerating capacity by the power, consumed by the electric motors of the compressor and fan.

The equations ( 4) and (5) are approximate. However, the experimental data can be expressed by the dependences ( 4) and (5) precise to 1 to 3 % according to the numerous tests carried out at VNIKhI with small hermetic compressors with tc ranging from 25 to 50° C.

In order to exclude the possibility of deriving very cumbersome final formulae, it is expedient to substitute the logarithmic temperature difference by an arithmetic one, namely fJ', adopting

fJ = C5 . fJ' where 0' = tc - 0.5 (ta1 + ta2) (8)

tc - ta1 The coefficient C5 depends on - -

tc - ta2 When this ratio changes from 1 to 5, the coefficient C5 alters from 1 to 0.83. The following expressions can be derived from the equations 1 to 8, connecting the

condensing temperature and the performance factor of the unit with the condenser surface and velocity of the cooling air.

A Bta1 - A tc =

B + ii(IfC1M+-

1) (9)

M

- - ----- - -D-Etc

C4 ( W y) n + I F . <p + f� . 102 . Y 1}f . 1}fm . ( A-Btc}

1 + <p

7200 . W . y . F . . Sa ib

Sa - specific heat of air.

(10)

(11)

The equation (10) presents the influence of the air velocity, condenser surface area and its arrangement on the performance factor. The equations (4), (9) and (11) show the relations between the same three characteristics and the refrigerating capacity of the unit.

Following are the results of calculation, made according to the derived formulae, in order to illustrate the quantitative influence of the given factors on the operation of the unit.

The calculation is made for three temperature ranges, characterizing hermetic units, manufactured in the USSR according to the GOST (State Standard) 9834-61 .

Conditions

I to = -35°; ta1 = 20° II to = -15°; ta1 = 20° III to = 5°; ta1 = 30°

According to the data, resulting from the tests carried out with refrigerating freon-12 units of small refrigerating capacity, the following values of the coefficients A, B, D, E are adopted in the calculation:

Conditions

I II III

A

2600 2900 3240

B

30 30 30

D

1725 3230 6500

E 25 43 90

Calculations are effected at W y = 1.5 to 8 kg/m2 sec; F = 6, 12, 18, 24 m2 ; ib = 2,4, 6; 1]r = 0.3; 1}tm - from 0.35 (at Nrm = IOW) to 0.74 (at Nrm = 400 W); C2 = 20; C3 = 0.95 (at ib = 2); Ca = 0.99 (ib = 4) ; Ca = 1 .01 (ib = 6) ; m = 0.32; C4 = 0.144 + 0.036 X (ib - l ) ; n = 1 .8 ; <p = 0.037; Sa = 0.24 kcal/kg °C.

The results of the calculations (Figs 1, 2, 3, 4) make it possible to state the following conclusions :

551

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552

'llr---.-----..---� 3400 � 32"' ��+--�30/J, 7

t -" ., .. l/l-,--+--+-1--1� 280

'UfA.-+"'4'�-+---.j '::J 260 .,.____,______,___ --'l�f-2 ---r<::<.J 240

-11 11

220 0 .. "' ., 11 '

0 0 � 8001-t-...:::+:=� 'O II

'111---_,_..._--l--�W--I ,!' 60•m--1-�+-J:b-I 400•1--1--+-+-I-� 200._..__..__...__....._.

i8 =6

i..-::: / ,_. t...

� II" j

I/ ,_ --I / i..-, II ' I

-

.mi

'2,, 6 '24

,, I 12 6

• ..., 'T> ·� ....

8 10 0 2 4 6 8 10 WJkg/m2sec WJ,kg/m1sec

4 6 6 10 WJ,kg/m2 sec

-- ta 1 = t?O o --- ta 1 = 30 "

� � lfllf/ a -le. 18(),•m-+-=-� i;:, 160111--��+--I

1-1oor1c--+-,�+-+-1200,__.,__1---1---f--f---1000'1--!---!---l--f--f-­BOO <--1--1--1--f-f-&a,,,..__,___,__......_..,_ 0 z 4 6 8 10 wx.1rg/m1sec

ta1 =20·

Fig. 1

1400'1-+--hoo''F-+-­f200 lc--+4+-+-+--10001--fL--+-+-+-800 l-f---1--1--+--GOO i,,....,,__,...._,.......,.....,, 0 z 6 " WJ,kg/m1sec

Fig. 2

1600't--+--+-t--1 14000---+-1--l--f--f---f--f---1200 O--�+--f--f---f---< 10001-r-+--r-i.--T. 'I WX lfi/m2sec

t;-J5 t8=6

1200o--.......... �--+--1000'1-l-Ll--I--+--

800'1--t--i-+-+-600 o�z!----l,_...,6_8_'0

W!,,kg/ml sec

" 201li

1500

1000

500

F= 19m2 1000 12 --+---t----+--1

500�--1.�.....1...�-'-�"'-----'�-' 50 100 1'0 200 �O JOO 350 fJ /F /ccaf'/m2hr F=24m2 � \ 1a I I I

" 12 + +i6 =2 }\ )( )( i6 =4 \. 6 6 lt = 6 " 1'.. to=5 0

6- ta1=30° IF=Z , �Z

� � L 12 �

Fig. 3

I --

-

t0=-15 ° 6 :--. taf20o

,_ �

IF=T� 'mz t =-35• 0 ,, ..... /Z tar20 �-� 6

I 0 1 2 3 4 5 6 7 8 9 w wx,xg/m2sec

III-30

Fig. 4

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III-30

Fig. I. Performance factor of unit vs air mass velocity, evaporating temperature, dimensions and arrangement of condenser surface. Kcu - performance factor of unit, equaling the quotient from the division of the refrig­erating capacity by the power, consumed by the electric motors of the compressor and fan. ib - the number of condenser banks. F - external heat transfer surface of the condenser. 10 - evaporating temperature. la1 - the temperature of air entering the condenser Wy - air mass velocity in the narrow section of condenser.

Fig. 2. Refrigerating capacity of unit vs air mass velocity, evaporating temperature, dimensions and arrangement of condenser surface at an air temperature of 20°C. Q - refrigerating capacity of the unit 10 - evaporating temperature ib - the number of condenser banks F - external heat transfer surface of the condenser la1 - the temperature of air entering the condenser W y - air mass velocity in the narrow section of condenser

Fig. 3. Maximum performance factor of unit vs specific heat load, condenser surface, number of banks and evaporating temperature Kcu - performance factor of unit, equaling the quotient from the division of the refrig­erating capacity by the power, consumed by the electric motors of the compressor and fan. F - external heat transfer surface of the condenser la1 - the temperature of air entering the condenser ib - the number of condenser banks 10 - evaporating temperature Q · - refrigerating capacity of the unit

Fig. 4. Maximum performance factor of unit vs air mass velocity, evaporating temperature and arrangement of conder.ser surface. Kcu - performance factor of unit, equaling the quotient from the division of the refrig­erating capacity by the power, consumed by the electric motors of the compressor and fan F - external heat transfer surface of the condenser ib - the number of condenser banks 10 - evaporating temperature laI - the temperature of air entering the condenser Wy - air mass velocity in the narrow section of condenser

1 . An increase of the condenser surface area and a reduction of the specific heat load to Q/F = 150 to 180 kcal/m2 hr considerably increase the refrigerating capacity and the performance factor (Figs. 1, 2, 3).

A further increase of the surface area improves the characteristics of the unit very slightly.

The surface area of condensers should be determined for low suction pressure units adopting Q/F = 180 kcal/m2 hr, and for medium pressure units - 150 kcal/m2 hr.

2. The air flow velocity, providing a high performance factor, at the earlier given condenser loads equals 5-6 kg/m2 sec in low suction pressure units, and 4-5 kg/m2 sec in medium pressure units (Fig. 1).

3. An increase in the number of banks reduces the refrigerating capacity and increases the consumption of electric power (Figs. 2, 3, 4). It is valid, therefore, to install a con­denser with a smaller number of banks in all the cases when the overall dimensions of the units are not prevailing.

An analysis of the influence of the air flow velocity on the costs of refrigeration in­dicates that the value W y, corresponding to the minimum cost of refrigeration, is approximately by 1 kg/m2 sec higher than the value of velocity, corresponding to the minimum consumption of electric power. This is due to the reduction of expenses relating to per unit of refrigeration for maintenance and depreciation at an increase of the air velocity and refrigerating capacity of the unit.

554

Construction des condenseurs evaporatifs et des condenseurs a l'air pour les installations frigorifiques lourdes

Construction of Evaporative Condensers and Air Condensers for Heavy Refrig­erating Plants

I V LUMIR VEPREK

CKD ChoceI'i, Tchecoslovaquie

SUMMARY. A simple method for the measurement of the relative volume of non-condensable gases in vapour has been applied. The relative volume of air in the vapour discharged by the condenser of a refrigerating plant, without automatic removal of air, is nearly always higher than 10 to 20%.

With an automatic system the relative volume of air can be kept down, for instance from 3 to 5%, thus reducing condenser water or the energy consumption of the compressor. To pro­vide efficient and economical air removal and full condenser capacity, water cooled condensers should be constructed in accordance with the directives set out in this paper. The construction of air cooled condensers should also cater for efficient air removal and ensure continuous liquefied refrigerant flow through all parallel branches of the finned tubes.

Comparative measurements show that, by improved construction, the weight and the bulk of condensers can be decreased by approximately 20 to 40'1,, .

I. MESURE DU VOLUME RELATIF D'AIR

Tout d'abord une methode simple de controle du contenu d'air dans la vapeur des fluides frigorigenes sera expliquee, qui a fait ses preuves dans les laboratoires et dans le montage et exploitation des installations frigorifiques.

Pour le controle on a besoin d'un cylindre calibre de verre avec un volume de 100 ou 1000 cm3, dont l'echelle s'etend autant que possible jusqu'a bord, puis d'un vase pla­te avec un volume de 3-10 litres, dans lequel on peut facilement noyer le cylindre calibre sous la nappe de liquide absorbant (l'eau pour l'ammoniaque, le petrole pour les Refri­gerants 12 et 22) et d'un tuyau en caoutchouc assez long avec diametre interieur de 5-10 mm.

100% 1 8 %

r�I Figs. r -3. i\fode de procooe pour mesurer le contenu d'air clans la vapeur des fluides frigorigimes

On ajoute une fin du tuyau a un robinet monte sur le point d'installationfrigorifique, oil on veut executer le controle ou la desairation, frequemment a la sortie du condenseur, au sommet du reservoir de liquide ou du regleur a flotteur haute pression. On ouvre legere­ment le robinet et on met l'autre fin du tuyau jusqu'au fond du cylindre calibre noye en biais dans le liquide absorbant (Fig. 1 .). Aussitot que !'air est ecarte du tuyau par le melange sortant de vapeur et d'air, on tourne le fond du cylindre calibre en haut (Fig. 2). On fait entrer le melange si vite que le cylindre est plein dans 1-2 secondes. Des que le me­lange commence a deborder, on arrete l'amenee du melange en retirant le tuyau et en fermant le robinet.

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Absorption de la vapeur commence tout de suite a diminuer le volume du melange des gaz, la nappe du liquide absorbant s'eleve. On peut activer !'absorption par oscillation horizontale du fond du cylindre de sorte que la nappe de l'eau absorbant l'ammoniaque se stabilise pendant quelques secondes et la nappe du petrole absorbant le refrigerant 12 se stabilise approximativement dans une minute. On peut lire sur l'echelle du cylindre le volume relatif d'air (Fig. 3), c.-a.-d. la relation de volume d'air au volume du melange apres la detente a peu pres isothermique ii la pression atmospherique.

Le calcul et les experiments repetes ont montre, que les fautes de mesurage par !'ab­sorption precoce, par les aberrations de pression et de temperature et par le reste de la vapeur dans !'air apres !'absorption sont pour les mesurements courants sans importance et qu'on ne doit pas calculer les corrections. On peut repeter le mesurage bien des fois sans echange du liquide absorbant.

Les mesurements du volume relatif d'air dans les installations frigorifiques diverses ont temoigne, que les methodes courantes de la desairation appliquees sous regime de marche sont tres inefficaces et ineconomiques et comme incertains et devalues sont les mesure­ments du coefficient de transmission de la chaleur sur les condenseurs, qui sont executes sans controle du volume relatif d'air.

Egalement le resultat de mesurage du rendement du compresseur frigorifique dans un circuit sans condensation peut etre deforme, si la vapeur contenant l'air passe les dia­phragmes ii mesurages. Par un volume relatif de 10% dans l'ammoniaque le rendement volumetrique du compresseur s'agrandit en apparence de 3,5 % et par 10 % d'air dans le Refrigerant 12 il se reduit en apparence de 3,9%.

II. MESURE DU COEFFICIENT DE TRANSMISSION DE CHALEUR

Les mesurages du coefficient de transmission globale thermique executes en 1958 sur les condenseurs de construction diverse avec volume relatif d'air variable dans le reste de vapeur sortant d'appareil ont apporte de meme des notions nouvelles. Avant tout on a constate, que les pertes du rendement par les gaz incondensables sous volume relatiffixe dans la vapeur sortant d'appareil sont tres variables d'apres la possibilite diverse d'em­porter les gaz incondensables a l'interieur du condenseur par le courant de vapeur humide vers la sortie d'appareil.

Tableau I. Resultats des mesurements des condenseurs ii l'eau.

Type de condenseur : Condenseur Condenseur Condenseur a calandre a ruissellemen t a ruissellemen t avec tubes avec tubes avec serpentin lisses verticaux contrecourant

horizontal

Fluide frigorigene : Refrig. I 2 ammoniaque ammoniaque

Volume relatif d'air dans le vapeur sortant d'appareil % IS 3 IO 4 IS IO 4

Ecart moyen de temperature (temp. de cond. mesuree par manometre) o c 9 5 5,8 2,5 6,4 4,4 2, 5

OF 16,2 9 ro,5 4,5 I I,5 7,9 4,5

Coefficient global de passage de chaleur a l'eau fg/m2h°C 250 450 ISO 350 450 650 I ISO

btu/ft2hr°F 5 1 92 31 72 92 133 235

Maintenant, les resultats principaux cites dans la table 1 seront commentes. I1 est vrai, que la vitesse de la vapeur humide dans un condenseur a Refrigerant 12 (Fig. 4) ii calandre n'est pas assez grande pour emporter !'air vers la sortie d'appareil, mais l'air, plus leger que

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le Refrigerant 12 peut etre entraine du sommet du condenseur. La quantite considerable du Refrigerant 12 liquide ruissellant sur les tubes detruit et ecarte la couche d'air con­densee de la proximite de surface froide. Le volume relatif moyen d'air a proximite de la surface des tubes est evidemment plus petit qu'au sommet du condenseur. Malgre cela, les pertes du rendement sont considerables.

Fig. 4. Condenseur a Refrige­rant I 2 a caJandre

Fig. 5. Condenseur a ruisselle­ment

La construction du condenseur a ruissellement avec un grand nombre de tubes verti­caux paralleles (Fig. 5) est entierement inconvenable. La vitesse de la vapeur dans le con­denseur est insignifiante et, en consequence de distribution inegale d'eau et d'influence inegale du vent et du soleil, le refroidissement des tubes particulieres est different. Tout d'abord, le rendement des tubes les plus refroidis est plus grand, la vapeur entre dans ces tubes-ci d'en haut et, en consequence de la resistence inegale du courant, aussi d'en bas par les tubes moins refroidites. A ce moment, une desairation des tubes les plus refroidites n'est pas possible, !'air s'accumule ici et en consequence de cela, le rendement des tubes particulieres se dirige vers le petit rendement du tube le plus moins refroidi. Le vo­lume relatif moyen d'air a proximite de la surface refroidissante est plus grand que le volume relatif dans la sortie du condenseur. Une desairation de ce condenseur-ci n'est pas possible sans evacuation totale. Le coefficient global de transmission thermique sous regime de marche est un fragment du coefficient calcule sans consideration d'influence des gaz incondensables.

Egalement fausse est la fonction des condenseurs a evaporation forcee avec un grand nombre de tubes paralleles horizontaux.

Au point de vue d'une desairation efficace et d'obtenir un coefficient de transmission de chaleur assez haut, le condenseur a ruissellement avec serpentins horizontaux a contre­courant (Fig. 6) est tres convenable. On peut choisir une vitesse assez haute de la vapeur a l'entree du condenseur et de plus obtenir une vitesse considerable aussi en haut a la

c:M=#\=IK=JK=l\=lt'=:::ll '-Fig. 6. Condenseur a ruisselle­

ment avec serpentins horizontaux

sortie du condenseur, oil le profil du courant de vapeur est reduit par le fluide frigorigene condense, qui detruit par ses mouvements en meme temps completement la couche condensee d'air a la surface refroidissante. Ainsi on obtient chez cette construction-ci aupres de condenseur multitubulaire vertical sous les memes conditions de calcul et sous le meme volume relatif d'air de 4-10% a la sortie d'appareil un coefficient global de transmission de chaleur a l'eau deux - trois fois plus grand.

Seulement apres une desairation parfaite, qui ne se trouve pas sous regime de marche des installations frigorifiques, les coefficients de transmission de chaleur de ces deux con­structions Ia ne sont pas considerablement differents.

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III. CONSTRUCTION DES CONDENSEURS A L'EAU

Alors, il faut construire les condenseurs a l'eau qu'il est possible d'apres les directives suivantes :

1. Determiner le nombre de tubes ou de branches des serpentins tubulaires paralleles et leurs profils de sorte que l'ecoulement de la vapeur humide est turbulent dans toute la longeur des tuyaux et qu'une circulation de la vapeur a l'interieur d'appareil en conse­quence du refroidissement different des tubes ou d' ejection de la vapeur a l' entree du con­denseur (voir point b -Fig. 5) contre la direction determinee du courantn'est pas possible.

2. Assurer une distribution proportionnee d'eau aux branches particulieres des tuyaux. 3. Disposer un contrecourant de fluide et d'eau, qui assure une transmission de chaleur

assez bonne a la fin du condenseur, ou le volume relatif d'air est plus grand et diminue l'ecart de temperature effectif.

4. Diriger le courant d'ammoniaque d'en haut a bas et le courant de Refrigerant 12 autant qu'il est possible d'en bas a haut OU horizontalement, afin que la difference des poids specifiques de l'air et de la vapeur donne aide au courant d'air vers la sortie d'appareil ou vers le robinet de desairation.

IV. CONSTRUCTION DES CONDENSEURS A L'AIR

Dans la construction des condenseurs a l'air de grande puissance il faut respecter les directives analogues. La disposition de contrecourant (Fig. 7) est necessaire, car dans un condenseur a l'ecoulement transversal (Fig. 8), utilise souvent pour lespuissances pe­tites, l'effet des tubes a cote d'air sortant, echauffe par exemple de 5°C, est de 30-50% plus petit que l'effet des tubes a cote d'entree d'air. Comme les mesurages ont montre, la

Fig. 7 . Condenseur a contrecou­rant

-

.­Fig. 8. Condenseur a L'ecoule­

ment transversal

vapeur entre d'abord aux branches les plus refroidites d'en haut et aussi d'en bas en passant les branches moins refroidites, dont la resistance d'ecoulement est plus petite, jusque la puissance de ces branches -ci est diminuee par accumulation du fluide conden­se et de !'air approximativement au niveau de puissance des branches chaudes.

•

Fig. 9. Entrainement de fluide liquefiea !'aide des tubes verticaux

Ainsi on perd par exemple 20% de la puissance totale du condenseur. L'accumulation du fluide liquefie dans les condenseurs grands a l'air est facilitee par le diametre relative-

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ment petit et la grande longueur des serpentins particuliers des tubes horizontaux. La resistance d'ecoulement du Refrigerant 12 clans les branches de puissance pleine atteint par exemple 0,2 kg/cm2 et clans les branches avec 70 % de puissance pleine 0,1 kg/cm2•

Si la desairation parfaite et permanente du circuit etait possible, on pourrait com­penser les differences de resistance d'ecoulement et entrainer le fluide liquefie du con­denseur a l'aide des tubes verticaux a compensation par exemple 2 m longues, ajoutees a la sortie des branches particulii:res (Fig. 9), ce qu'est recommande clans la litterature professionnelle pour l'adjonction de plusieurs condenseurs paralleles. Mais ce procede-ci est entierement inefficace, quand les gaz incondensables entrent clans le condenseur et commencent a s'accumuler clans les tubes a compensation sur la nappe du fluide liquefie. C'est pourquoi le chargement egal de toutes les branches paralleles par contrecourant d'air est necessaire.

v v V. CONDENSEURS CKD CHOCEN.

Sur la base de ces notions-ci on a projete et mis en production les condenseurs a evaporation forcee (Fig. 10 et 1 1) avec un petit nombre de serpentins longs en forme de

Fig. 10. Section transversal

v Fig. 1 r. Condenseur I!. evaporation forcce, puissance 200.000 fg/h (66 tons) CKD Choceti

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vis, qui sont arroses d'un systeme de tubes tournants par effet reactif de l'eau sortante . Le poids net de ces condenseurs-ci est de 40 % plus petit que le poids des appareils bien construits de systeme habituel.

v Fig. 12 . Condenseur a l'air, puissance 50.000 fg/h (16 tons) CKD Choce�.

Le condenseur horizontal a l'air a contrecourant avec un ventilateur vertical (Fig. 12) possede non seulement l'avantage d'entrainement continue! d'air et de fluide liquefie, mais en comparaison avec un condenseur vertical (Fig. 8) en outre plusieurs avantages de projection et d'exploitation deja assez connus.

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Etudes comparatives des circuits de condensation frigorifique refroi­dis par refrigerants atmospheriques OU par condenseurs a evaporation

Comparative Studies of Condensing Refrigerant Circuits Cooled by Atmospheric Coolers or Evaporative Condensers

CH. FONTANEL, lngenieur Conseil, et M. VAUZANGES, lngenieur, Ste. Hamon Paris, France

SUMMARY :

1. Introductio n : The development of the refrigeration industry has presented a problem on the cooling of condensing water in closed circuits. This problem has to be studied from both the technical and economic aspects.

2. The s tudy of typical problems with atmospheric coolers and multi-tubular condensers :

2.1 Advantages and disadvantages 2.2 Types of coolers used : characteristics and performance 2.3 Cost of initial installation and functional expenses for these systems.

3. Study of the problem with evaporative condensers :

3.1 Summary description of evaporative condensers; operational advantages and disad­vantages.

3.2 Characteristics and performance of modern types of condensers 3.3 Cost of installation and operation. Estimation of expenses and compulsory servicing.

4. Recent developments and p rogress of the two compared sys tems -

4.1 Existing trends in the construction of atmospheric coolers and evaporative condensers. 4.2 The use of new materials in the spray surf aces of coolers in order to reduce weight and

other encumberances of this equipment. 4.3 The problem of water treatment and the protection of the circuits against corrosion and

furring. 4.4 Noise problems of fans.

5. Conclusions :

Field of application for each system :

- evaporative condensers for small or average capacities, with certain precautions. - atmospheric coolers and multi-tubular condensers for use in important installations,

due to reasons of economy and operational facilities.

1 : INTRODUCTION

Le developpement tres rapide de l'industrie frigorifique pose dans le monde entier le delicat probleme des ressources en eau et des systemes a employer pour economiser l'eau de refroidissement des condenseurs.

Nous comparerons l'emploi du systeme classique refrigerant atmospherique/conden­seur multimbulaire avec celui plus recent du «condenseur a evaporation », en tenant comp­te des progres recents de ces deux systemes, de leurs aspects economiques et techniques.

Dans notre conclusion nous essaierons de rechercher le domaine propre a chacun en fonction des avantages et inconvenients respectifs.

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2 : ETUDES DES SOLUTIONS CLASSIQUES AVEC REFRIGERANT ATMOSPHERIQUE ET CONDENSEURS MUL TITUBULAIRES

2.1 . : Avantages et Inconvenients -

Les principaux avantages sont : a) separation des circuits de refrigeration de l'eau et du fluide frigorigene facilitant

!'application de traitements appropries contre la corrosion et l'entartrage ; b) facilites d' entretien et de nettoyage ; c) possibilite d'utiliser partiellement le refrigerant atmospherique pour un autre circuit

( cas des laiteries et industries chimiques ou un liquide peut Ctre refroidi prealablement par un echangeur avec l'eau du refrigerant) ;

d) adaptation aux problemes de condensation de circuits frigorifiques de grande puissance ;

e) seule solution utilisable dans les circuits frigorifiques multiples au freon 12 ou au freon 22 ou la separation des circuits facilite le retour d'huile ; Ia centralisation des circuits d'eau sur un refrigerant atmospherique unique est alors une necessite.

Les inconvenients sont les suivants : a) encombrement global plus grand ; b) frais d'investissement parfois plus eleves ; c) frais de pompage d'eau plus importants par augmentation des pertes de charge.

2.2. : Types de refrigerants employes : caracteristiques et performances -

Pour les condenseurs frigorifiques, il faut une basse temperature d'eau refroidie, done un refrigerant a ventilation artificielle,

a) soit a tirage induit (ventilateur aspirant) pour les grandes installations (2.000.000 F/H) ou pour !'implantation (diminution d'encombrement, de bruit, de hauteur, de pompage).

b) soit a tirage force (ventilateur soufflant) pour les moyennes et faibles puissances (moins de 2.000.000 F /H).

D'autres classes de refrigerants sont definies suivant le sens relatif des courants d'air et d'eau :

- a contre-courant - a courants transversaux (courants croises)

ou le systeme de ruissellement : - gouttes (pulverisation, dispersion sur lattes) - films (ruissellement sur plaques) - combinaison des deux types.

2.2.1 . : Caracteristiques :

Les caracteristiques actuelles des refrigerants sont : - enveloppe : amiante-ciment, beton ou mai;onnerie, bois traite, plastique, tole

convenablement protegee. - systeme de ruissellement : amiante-ciment, plastique, bois traite, aluminium. - groupe moto-ventilateur : helice en tole d'acier, alliage d'aluminium, plastique (en

vue de rechercher les qualites de silence)

2.2.2. : Performances :

La figure 1 indique les courbes d::s temperatures d'eau chaude et d'eau refroidie pour deux refrigerants differents. Les allures des courbes different suivant l' «approche » consideree.

Normalement, pour une temperature ambiante moyenne diurne et nocturne de 25° C et 60% d'humidite, la plus basse temperature d'eau refroidie donnee industriellement est 22°C.

2.3. : Gout de premier etablissement et cout d'exploitation -

Le tableau I montre l'accroissement des dimensions, prix et puissances de ventilation pour differentes approches d'un m eme probleme thermique.

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...

Fig. r . Combes des temperatures d'eau chaude et d'eau refroidie pour deux refrigerants differents

Tableau I : Refrigerant atmospherique -Comparaison pour un meme probleme thermique. Debit : 50 m3/h ; Ecart : 5° ; Nombre de calories evacuees : 250.000 ; Ambiance : 25° - 60% d'humidite -----

Temperature Encombrement Puissance de Prix d'eau refroidie Approche L x l x H ventilation approximatif

(metres) (Ch.)

22° 2°5 3,50 x 2,50 x 5,00 5 p 23° 3°5 3,50 x 2,15 x 5,00 4 0,8 p 24° 4°5 3,00 x 2,25 x 4,80 3,5 0,7 p 25° 5°5 3,00 x 2,00 x 4,20 3,0 0,65 P

Tableau II : Refrigerant atmospherique Comparaisons pour un meme probleme thermique - Nombre de calories evacuees : 250.000 ; Eau refroidie : 22° ; Ambiance : 25° -60% d'humidite

Ecart de Tempera- Debit Encombrement Puissance Prix Temperature ture m3/h L x l x H Ch. approxi-entre entree moyenne (metres) matif

et sortie de l'eau d'eau

20 23° 125 6 x 2,50 x 4,50 6,5 1,2 p 30 23°5 83 5 x 2,15 x 4,50 6 1,15 p 40 24° 62,5 4 x 2,25 x 5,00 5,5 l,05 P 50 24°5 50 3,5 x 2,50 x 5 5,0 1 p 60 25° 41,5 3,5 x 2,25 x 5,25 4,5 0,90 P 70 25°5 36 3,5 x 2 x 5,50 4,0 0,85 P

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Le tableau II donne pour un meme probleme thermique et pour une meme approche, la variation de ces caracteristiques en fonction du debit, c'est-a-dire de l'ecart entre tempe­ratures d'entree et de sortie d'eau du condenseur.

La figure 2 montre, pour l'ensemble refrigerant - condenseur et pour une meme temperature d'eau refroidie, les variations de prix et de puissance en fonction des debits. On voit que I' «ecart optimum » LI tw est compris entre 5 et 7°C, mais depend aussi d'une fa�on etroite du choix de la surface d'echange du condenseur.

• i

A ... . I .____.. �� ''" J11 .1 .. lf �·- .

Fig. 2 Variation du prix d'installation et de la puissance absorbee pour un circuit frigorifique de 200000 F/H -Refrigerant eau refroidie 22°C (tableau II) et condenseur (S = 60 m2/A, S = 40 m2 /B)

3. : ETUDE DE LA SOLUTION AVEC CONDENSEUR A EVAPORATION -

3.1. : Description Sommaire des Condenseurs -

L'etude de Monsieur BERLINER (rapport 352 IIF) a fait le point recemment des progres dans la construction de ces types d'appareils. 11 n'est done pas necessaire d'insister sur un sujet traite de fa4'on tres complete.

3. 1 . 1 . : Avantages :

- encombrement reduit (construction compacte), - moindre cout de premier etablissement, - economie de force motrice theorique, - moindres frais d'installation (tuyauteries d'eau), - possibilite pour !'ammoniac de brancher Jes circuits haute pression de differents

condenseurs sur un seul appareil, - possibilite d'implantation a distance eloignee des compresseurs, - possibilite d'utilisation en condenseur a air, l'hiver (basses temperatures ambiantes),

evitant ainsi les inconvenients dus au gel de l'eau de ruissellement et du bassin.

3. 1.2. : Inconvenients -

- Difficultes d' entretien : - le nettoyage est tres difficile et le detartrage delicat : on est done conduit a adop-

ter des tubes lisses. - un traitement de l'eau d'appoint et une purge de deconcentration reguliere et

contr61ee sont indispensables. - Corrosion atmospherique :

- !'utilisation de materiaux inoxydables est plus difficile que dans le refrigerant atmospherique. Dans le «condenseur a evaporation » les actions corrosives de l'air et de l'eau s'additionnent au contact meme du serpentin de condensation du fluide frigorigene.

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3.2. : Caracteristiques et performances des types de Condenseur a Evaporation modernes -

Le tableau III donne les caracteristiques principales de plusieurs types de condenseurs a eau et a air.

A signaler que les difficultes de nettoyage font eliminer les tubes a ailettes au profit des tubes lisses convenablement proteges (generalement galvanisation a chaud).

Tableau III : Condenseurs a evaporation - tableau de comparaisons

Regime de Fonctionnement : -18°C (0° F) / + 35° C (95°F) - Bulbe Humide de l'air : +21°C ( +70° F)

Dimensions Surface Puissance Prix approxi-L x l x H d'echange to tale matif

Type F/H (metres) (tubes lisses) (Ch.) par 1 .000 F /H en m2

A 75.000 2,50 x 1,10 x 2,70 80 6 P' (25 tonnes)

B 1 15.000 3,00 x 1,25 x 2,80 145 8 0,85 P' (38 tonnes)

c 460.000 4,70 x 3,00 x 3,50 370 30 0,75 P' (152 tonnes)

Les tubes a ailettes ont l'avantage de mieux se preter au fonctionnement d'hiver sans eau (Berliner - op. cit. § 5).

Deux diagrammes indiquent les variations de puissance calorifique de ces condenseurs en fonction de !'ambiance et du thermometre sec ou humide (fonctionnement en conden­seur a air). Ces 2 diagrammes sont etablis dans le cas de condenseurs avec tubes lisses.

, ... , •. .,,. '"'•" tr ti u r, 1.1 U t.J f,4 t.J U f,f U .10 .ao .ra , 6' M

,. #0 If '·' ,,, u u (l ,, t.• .,, ,,. J• .40 _,. r.,..,.:r,1 .. ,.. - •""'• lotirtt•'"'- •c r-.,;l•ur Ntr1 'f � · a-pl• Qe 1'\lt .. uu ....... u. ,. Q P'T"l&/9 C J'\f.AHun trtprlU.,o) Q • • 1 .000 .,,.

'r till '• c., ..... l T., T .. T. )

<i, l"llU .... u •-tu.lo •• l ' •ll'll&"'"ll fell •tu •1110 H 011 .. rl•w� l I• ..,, .. _.,, o:•l.::W•• Oc

.. lH•t l• •&.,....­... tMll' F( : 111

�- ·!� ... .... --11 :.O-t l ' ..... 1l.:a af 111 ,!� • IC :r.�,•n

Fig. 3. Diagramrne de variati011s de puissance calorifique ; air humide

On notera qu'en hiver le fonctionnement en air sec (sans evaporation d'eau) n'est possi­ble pour la temperature de 0° que si la charge thermique est inferieure au tiers de la charge thermique normale admise.

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!.IR SEC

•10 -ts -to -4 O •S tfO •1S Tc'"fl. J. , .. ,. ....

, .. ,_,. '" ,, .... ,, J tt ,. .,o '° .,P

. ro .10 r•clo11r- r•• N'I�

"'

I

1 I ·"

Fig. 4. Diagramme de variations de puissance calorifique ; air sec

4. : DEVELOPPEMENT ET PROGRES RECENTS DES DEUX SYSTEMES COMPARES -

4.1. : Tendances actuelles dans la construction des refrigerants atmospheriques et des conden­seurs a evaporation

4.1.1. : Refrigerants Atmospheriques :

a) L'enveloppe, autrefois en tole pour les petites unites, en beton pour les plus grosses.

Pour ameliorer la durabilite des petites unites et alleger les unites plus importantes, les enveloppes sont construites maintenant en amiante-ciment ou plastique, plus faciles a monter.

b) Les systemes de ruissellement, tres divers et generalement brevetes, sont constitues par :

- le bois maintenant traite par trempage ou sous vide et pression avec des solutions chimiques complexes (chloronaphtalenes, sels de cuivre ou arsenic, . . . ) ; la dura­bilite ainsi fortement accrue depend neanmoins de la nature del'eau (pH non alcalin, teneur en chlorures et en carbonate de soude, etc . . . . ) ; elle est excellente en eau acide.

2 - la tole d'acier galvanisee, le zinc, !'aluminium.

3 - l'amiante-ciment en plaques planes ou ondulees ou lattes, solution resistant le mieux a la corrosion pour des eaux a acidite non permanente.

4 - les matieres plastiques, plus recentes, sur lesquelles un jugement definitif ne peut etre porte.

c) Les ventilateurs helicoides normalement utilises ont ameliore leur rendement (atteignant 0,8) avec turbines en alliage coule ou matiere plastique a profil de pales specialement etudie. Des helices a pales orientables en marche apportent un perfectionne­ment interessant aux problemes de regulation de temperature d'eau en fonction de I' ambi­ance.

4.1.2. : Condenseur a evaporation :

Le rapport de Monsieur BERLINER etant deja tres explicite a ce sujet.nous precise­rons seulement que les constructeurs sont generalement fideles a une execution entiere­ment en acier galvanise avec ventilateurs centrifuges, valables a la fois pour les fluides freon ou ammoniac et donnant les qualites de silence requises.

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4.2. : Utilisation de materiaux nouveaux comme surface de ruissellement des refrigerants atmospheriques pour reduire poids et encombrement de ces appareils -

Deux sujetions d'implantation sont le poids et l'encombrement. Les systemes «films » (amiante-ciment, metal, plastique) diminuent le volume. L'amiante-ciment, pour l'enveloppe, les matieres plastiques pour le ruissellement,

diminuent le poids.

4.3. : Probteme du traitement de l' eau et de la protection des circuits contre la corrosion ou [' entartrage -

Les surfaces de ruissellement des refrigerants et les tubes des condenseurs sont soumis a differentes attaques pour lesquelles differents types de traitement sont a etudier, a appliquer et a controler.

La durabilite des appareils en depend etroitement (corrosion, entartrage, algues, etc . . . ) . La sujetion de silence conduit frequemment aux ventilateurs aspirants a axe vertical

places sur Jes appareils, les ondes sonores dirigees vers le haut etant alors peu genantes. Les ventilateurs centrifuges permettent d'obtenir, d'ameliorer le silence mais sont plus

onereux. En resume, pour avoir de bonnes conditions de silence (50 dbs a 10 m), acceptables dans

la p!upart des cas, il faut limiter la vitesse d'air dans la section nette du ventilateur a 7 m/s et la vitesse peripherique en bout de pales a 35 m/s.

La solution condenseur/refrigerant atmospherique permet de mieux choisir !'emplace­ment le plus approprie pour le refrigerant, ce qui est plus delicat pour le condenseur a evaporation.

4.4. : Probleme du bruit des ventilateurs -

Ce probleme interessant les refrigerants atmospheriques et les condenseurs a evapora­tion est primordial particulierement pour les installations de conditionnement d'air en pleine agglomeration (immeubles collectifs, magasins, hopitaux, etc . . . . ). De plus les tendances actuelles, parfaitement louables, a !'amelioration des conditions humaines de travail deplacent le champ d'application de ces questions vers l'industrie.

II est normalement utilise des ventilateurs helicoi:des mieux adaptes aux forts debits d'air mais faibles pressions demandes. Le niveau sonore peut atteindre 90 a 100 dbs (equivalence : atelier bruyant, motocyclette, voiture de metro), inacceptable en ville la nuit.

II y a deux sources majeures de bruit : - le choc de !'air sur Jes pales - (que caracterisent la vitesse d'air dans la section de

passage et le nombre de pales) - le frottement de !'air sur la virole - (que caracterise la vitesse peripherique en bout

de pales). II faut done, pour un meme probleme debit-pression, augmenter le diametre de

l'helice et diminuer sa vitesse de rotation. L'on abaisse ainsi le niveau sonore a 45/55 dbs valeurs genera!ement acceptables pour

Jes implantations considerees (equivalence : interieur d'un appartement calme). Les helices en plastique dont l'inertie sonore est sensible donnent d'excellents resultats.

En outre les ventilateurs sont montes sur plots anti-vibratiles, et Jes roulements sont prevus silencieux.

La sujetion de silence conduit frequemment aux ventilateurs aspirants :1 axe vertical places sur Jes appareils, les ondes sonores dirigees vers le haut etant alors peu genante�.

Les ventilateurs centrifuges permettent d'obtenir d'ameliorer le silence mais sont plus onereux.

En resume, pour avoir de bonnes conditions de silence (50 dbs a 10 m), acceptables dans la plupart des cas, ii faut limiter la vitesse d'air dans la section nette du ventilateur a 7 m/s et la vitesse peripherique en bout de pales a 35 m/s.

CONCLUSION

Le tableau IV resume !es variations de prix specifiques (c'est-a-dire ramenees au 1 .000 F /H) les deux systemes.

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Tableau IV : Tableau comparaison de variation de prix - des deux solutions en fonction de la puissance frigorifique

F/H Prix P par 1.000 F /H Prix P' par 1.000 F /H pour condenseur a pour refrigerant evaporation atmospherique et

condenseur multitu-bulaire

80.000 P x l P' x 1

230.000 p x 0,95 P' x 0,80

460.000 p x 0,90 P' x 0,60

P et P' sont des prix specifiques, c'est-a-dire ramenes sur la base du cout par 1.000 F/H.

II en resulte que le prix de la solution Condenseur Multitubulaire + Refrigerant dimi­nue beaucoup plus vite que celui de la solution avec condenseurs a evaporation.

En conclusion, et a la suite de J'enquete effectuee aupres des constructeurs et exploi­tants de ces deux types d'appareils et que nous remercions pour Jeur collaboration, ii apparait :

a) Dans les installations frigorifiques «commerciales » - Jusqu'a 30.000 F/H le condenseur a evaporation a tendance a etre le plus employe,

sauf dans le cas de circuits multiples au freon ou il est plus economique d'installer un seul refrigerant atmospherique desservant Jes condenseurs multitubulaires de chaque circuit.

Le condenseur a air presente de nombreux avantages dans ce domaine, sauf pour Jes climats chauds ou une refrigeration complementaire a eau est necessaire en ete.

b) Dans les installations «industrielles » - De 30.000 a 300.000 F/H, Jes deux systemes sont competitifs avec avantage de prix

et encombrement pour le condenseur a evaporation, et avantage de poids, de robustesse et durabilite pour le refrigerant atmospherique. Done, dans chaque cas, le choix dependra de differents facteurs et notamment de !'implantation possible.

c) Dans les installations industrielles de grande puissance centralisees. - Au-dessus de 300.000 F/H, la solution classique, refrigerant + condenseur multi­

tubulaire est sans conteste la plus valable pour des raisons de simplicite d'installation et d'entretien et de cout annuel d'exploitation.

Une collaboration etroite, dans ces differents domaines, est done souhaitable entre les constructeurs de refrigerant atmospherique, les constructeurs et installateurs de materiel frigorifique et l'exploitant de J'installation frigorifique.

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Reciprocating and Turbo-Expanders for Low Temperature Refrigera­tion

Detendeurs alternatifs et turbo-detendeurs pour le refroidissement a basse temperature

J. DE W AELE, Manager of Engineering Worthington Corporation, European Operations, 61, rue des Belles-Feuilles, Paris (16°), France

SOMMAIRE. Limites a l'origine a la liquefaction et a la separation de l'air, !es procedes a basse temperature se sont repandus rapidement dans d' autres industries impliquant la pre­paration, l' entreposage, le transfert, le transport et /'utilisation de gaz liquefies.

Les cycles frigorifiques a compression de vapeur, avec detente directe d'un liquide condense par un robinet d'etranglement sont tres ejficacesjusqu'a -20° C, mais le rendement s'abaisse rapidement en Jonction de l'abaissement de la temperature. L'utilisation de deux ou plus de ces cycles en cascade permet d'ameliorer le rendement, mais le systeme qui en resulte est sou­vent encombrant, couteux et difficile a reg/er.

Les cycles bases sur /'utilisation de l'effet Joule-Thomson ont le merite de la simplicite et permettent d'obtenir de tres basses temperatures avec quelques gaz, mais pas taus, suivant la forme des courbes isothermes sur le diagramme de Mollier. Le cycle gaspille de l'energie puisque toute celle utilisee pour la compression est perdue.

Le cycle de Claude, dans lequel le gaz est detendu dans une machine confue pour recuperer une grande proportion de l' energie disponible fournit un moyen simple et direct de reduire sa temperature, facile a reg/er et exigeant peu d'accessoires. Il est remarquablement ejficace lorsqu'il est utilise pour obtenir de forts abaissements de temperature, a condition que le com­presseur et le detendeur soient des machines bien confues, bien adaptees aux conditions d'utili­sation. Le rendement total du cycle dependant de la difference entre l' energie fournie au com­presseur et l'energie recuperee au detendeur, de tres petites variations du rendement d'un ele­ment ant une influence accrue sur le rendement du cycle.

Les detendeurs alternatifs et les turbo-detendeurs sont maintenant capables de repondre aux differents besoins de divers procedes. La detente variable a l'entree des machines alter­natives, l'ajustement des turbo-detendeurs avec un reg/age de l'ajutage et /'attention soi­gneuse aux pertes de chaleur dans /es deux categories de machines couptees avec de nombreux autres perf ectionnement moindres ant permis une augmentation remarquable du rendement et une securite egale a celle des au tr es machines industrielles.

INTRODUCTION

The laboratory experiments of Cailletet and Pictet in 1877 laid the foundations for the production of liquid air on an industrial scale, by Linde and Hampson in 1895. The cycle adopted was based on the use of the Joule Thomson effect, with heat exchange between the expanded air and the high pressure air. The air being recycled continuously, the cool­ing effect is cumulative until equilibrium is reached at a temperature at which some of the air is liquefied and has to be compensated by external make-up.

In 1906, Claude replaced the free expansion by direct expansion in a reciprocating machine converting part of the available enthalpy into external work, thereby improving the cycle efficiency considerably. Since then the air liquefaction and separation industry has not only increased tremendously in total capacity, but has sought for, and obtained lower operating temperatures in order to recover the valuable inert gases, neon and helium, of which the liquefaction temperatures at atmospheric pressure are 27,2°K and 4,2° K, respectively.

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At the same time, the demand for low temperature refrigeration has extended to a wide variety of other industrial fields concerned with the production, storage, transfer and long distance transport of liquid ethylene, methane, fluorine and hydrogen in addition to the oxygen, nitrogen and inert gases obtained from air.

This demand has stimulated the need for expanders, first of reciprocating design, later of turbine type.

Comparison of the Claude cycle with others

The time allotted to this paper unfortunately does not allow a formal demonstration of the following comparative figures, which although of necessity approximations, since many assumptions must be made as to individual component efficiencies, can nevertheless be substantiated broadly. - A modern ammonia compression unit, condensing at 30° C and evaporating at -Z0° C will yield approximately Z500 frigories/KWhr (1 .65 HP per ton).

The equivalent Claude cycle, operating in closed circuit with air, the cooled air being used in a heat exchanger to cool brine, will only give ZOO frigories/KWhr. On the other hand, if used in open cycle to cool atmospheric air from Z0° C to -Z0° C, the yield in­creases to 750 frigories/KWhr.

The first conclusion is therefore that the use of an expander is uneconomical at high temperatures, and at the same time it can be supposed that whatever the refrigeration level, the Claude cycle should be used for direct cooling of the working fluid rather than with a heat exchanger for cooling another fluid. - At -100°C, a typical Rankine cycle system will give about 470 frigories/KWhr. whilst the Claude cycle gives 635 frigories/KWhr. - At -Z00° C, these figures become respectively 72 and 395 frigories/KWhr.

It is therefore evident that below the cryogenic level, usually considered to be -100° C, the Claude cycle is already more efficient than vapour compression, the advantage in­creasing with further reduction in temperature. Furthermore, two or more vapour compression cycles must be used in cascade at such temperatures, and the resulting system is cumbersome, expensive and rather difficult to control when the load varies.

In contrast with this, a single expander of the reciprocating type can now be used to cool a gas through a temperature range of Z00° C.

We have seen that the Claude cycle is wasteful at high temperature. The use of a vapour compression cycle to pre-cool the gas down to around -30° C before expansion will improve the overall cycle efficiency.

Furthermore, since the efficiency of the Claude cycle is largely determined by the difference between the power required for compression and that recovered during ex­pansion, it is evident that minor changes in efficiency of each component will be reflected by magnified changes in the cycle efficiency. For optimum results, therefore : - The power required for compression should be reduced by correct staging, efficient intercooling and, if possible, by sub-cooling the suction gas to each stage below ambient conditions, thus keeping as near as practicable to isothermal compression. - The power recovered from the expander should be kept as near as possible to the adiabatic power by correct expansion in the machine and minimum heat pick-up from the surroundings.

It is not necessary to dwell for long on the comparison between free expansion and expansion with recovery of energy. The power required for compressing the gas is identical in the two cases, and a glance at the Mollier diagram will show the great differ­ence between the temperature drops obtained on the one hand from the Joule Thomson effect and, on the other hand, from an expansion in which from 75 to 85 % of the en­thalpy is taken out in the form of work, modern machines achieving such results quite easily. Furthermore, the Joule Thomson effect is negligible or even negative in certain pressure ranges.

After this very brief comparison which shows the conditions under which the Claude cycle offers very appreciable advantages over others, we will examine the salient character­istics of expanders which have already been used extensively and which will certainly be called upon to play an increasing share in cryogenic industries.

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Field of application of different types of expanders

Excluding the specialized ultra-high speed, low weight turbo-expanders which have been developed for aircraft and missile refrigeration duty, the field of application of the various types of expander available commercially can be subdivided broadly as follows : - Reciprocating machines for small flow, high inlet pressure, high pressure ratio. - Radial flow turbo-expanders for large flow, low pressure ratio. - Axial-flow turbo-expanders for large flow, low or medium pressure ratio, and in partic-ular when partial liquefaction occurs in the machine.

A certain amount of overlapping occurs in practice between the ranges covered by each class, particularly in the case of the two types of turbo-expander.

RECIPROCATING EXPANDERS

Basic design features The reciprocating expander, although incorporated in a cycle which also includes a

compressor, and thus apparently a machine of similar principle, used in reverse, is in fact more akin to a steam engine in some respects. In a compressor, the cylinders must be adequately cooled to keep the work cycle as nearly as possible isothermal, the valves are operated automatically by differential pressure and the clearance volumes must be sufficient to avoid mechanical damage in the event of condensation or liquid carry-over. Compression ratio is limited by discharge temperature, and despite good cooling the shaft power exceeds the adiabatic power. In an expander, the valves must be actuated mechanically, preferably by valve-gear allowing variation of cut-off, to maintain good efficiency under variable load ; clearance volume must be as small as possible to minimize losses from mixing the fresh charge with cold gas before expansion, and the gas end must be well insulated from its surroundings and from the frame and running gear, to reduce heat losses.

The pressure ratio is high, and the power recovered is less than the adiabatic power. The ratio of frame load to power is therefore much higher than in a compressor (75 -180 kg/kW as against 30 - 40 kg/kW), which accounts for the more massive proportions of the expander frame, crankshaft and running gear.

The high pressures impose the use of small poppet valves, with lengtl;iy packing not only to hold the pressure but to provide the necessary temperature gradient. This feature in turn demands unusually strong valve springs, supplemented on some machines by pneumatic loading.

Warming jackets are sometimes needed to prevent valve seizure due to freezing of atmospheric moisture.

The valve operating mechanism is the heart of a good expander. Since the valves operate with small lift, the linkage must be designed to maintain its clearances inde­pendantly of the operating temperature. It must be very robust because of the large forces it has to transmit, and yet of low inertia.

The cut-off control must be precise and robust. Machines with lubricated cylinders are fitted with cylinder warming jackets, of partic­

ular benefit when re-starting after shut-down during which the whole of the gas end may become subjected to temperatures close to exhaust temperature.

Non-lubricated cylinders are used for very low temperatures, such as those required for helium service, at which all known lubricants become solid ; machines are now available for temperatures as low as 5° K.

Low temperatures present other problems amongst which can be cited the loss of resilience of many materials, and differential contraction of parts tending to modify clearances.

In addition to external lagging, heat insulating barriers are sometimes inserted between the cylinder and the main frame.

Reciprocating expanders are made with one or two cylinders, fitted with plungers or pistons according to the operating pressure, arranged either horizontally or vertically, the choice being largely dictated by considerations of space and piping layout. The non-

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lubricated machines are always vertical. Fig. 1 shows a single cylinder horizontal expander suitable for 200 bars (2900 psi), and Fig. 2 part of a single cylinder vertical expander designed for 40 bars (590 psi).

-··-

0

Fig. r. High pressure horizontal reciprocating expander

Fig. 2. Medium pressure vertical reciprocating expander

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Power absorption

Reciprocating expanders are low speed machines. They can be coupled, either directly or through belted transmission to any suitable power absorbing machine such as a friction brake, a hydraulic brake, a hydraulic pump or a generator.

Being single acting machines, sometimes single cylinder and with relatively high friction losses owing to the large number of rings required, they are seldom self-starting. Consequently, the use of a reversible machine, such as a motor/generator or a pump/ turbine is far more convenient than a brake as it permits immediate starting.

Another solution which is frequently resorted to is the direct coupling of the expander to the extended shaft of the motor-compressor unit. This greatly simplifies power control, any variation in available power being automatically compensated by a corresponding adjustment change in the motor load. Fig. 3 illustrates a unit of this type.

Fig. 3. Twin cylinder expander directly coupled to compressor

Capacity control

In certain circumstances, capacity control by speed variation may be preferred, but it is usually more convenient to vary the cut-off, and this method allows high efficiency to be maintained over a wide range, although it is bound to fall off slightly with decreasing load, owing to the fixed losses in the machine. Provision can be made for manual variation, usually through a screw mechanism, which is necessarily slow. Most modern machines are fitted with automatic control, either pneumatic or oleo-pneumatic. The control system is interconnected with the various safety devices which reduce the cut-off to zero and so stop the machine, in the event of abnormal conditions.

Safety devices In addition to the pressure relief valve or bursting disc fitted on the exhaust manifold,

reciprocating expanders are fitted with overspeed, underspeed and low oil pressure trips.

Foundations The massive reciprocating parts which are a characteristic of expanders create inertia

forces which cannot be completely balanced. Even more care must be taken in designing the foundations and in grouting the machine than when dealing with a reciprocating compressor of similar size.

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Surge drums

The single acting cylinder and low cut-off are responsible for wide and sudden variations in the flow of air in the inlet piping, and adequate surge drums must be used to eliminate piping vibrations and pressure surges. Conditions are much better on the exhaust side, particularly of a twin cylinder machine, from which the flow is almost constant, hut this problem must not be overlooked since correct operating of instruments controlled by exhaust pressure may be upset, even if the piping is apparently undisturbed.

Operating range

The range of capacity, pressure ratio and maximum operating pressure has to a certain extent been determined by demand, and varies from one manufacturer to another. Without representing absolute limits, the following approximate figures may be regarded as typical :

Inlet pressure Pressure ratio Shaft speed Piston speed Shaft power Capacity at exhaust conditions

20 - 210 bars (290 - 3000 psi) 5 - 75

100 - 600 rpm 1 - 3.5 m/sec. (200 - 700 ft/min.)

15 - 400 kW 0.2 - 6 m3/min. (7 - 210 cfm).

The maximum values are individual limits, not necessarily attainable simultaneously.

Performance

A typical pressure-volume diagram obtained from a machine taking in air at 195 bars, -23°C (250° K) and exhausting at 3.5 bars and -193°C (80° K) is shown in Fig. 4. The efficiency of this machine, expressed as the percentage of enthalpy drop actually

Fig. 4- Typical reciprocating expander indicator card

recovered is 83.4%. Good design of the valve gear, as evidenced by the continuity of slope before and after complete closure of the inlet valve, despite the rapid acceleration of the plunger in this region, is one of the main contributing factors to attaining this high efficiency. In general, efficiencies of SO to 85 % can be expected.

Operating precautions The very close running clearances and small clearance volumes which are essential fea­

tures of high efficiency expanders make them sensitive to any solid particles entrained by the gas, which therefore must be very clean.

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Furthermore, although the machines are designed so that they can operate into the saturated vapour region and handle quite a high proportion of condensate, there must be no constituent of the gas capable of freezing in the machine, since this will reduce the clearance volume and soon result in mechanical trouble due to hammering, unless by some fortunate hazard it first freezes the exhaust valve in the open position.

TURBO EXPANDERS

Design considerations

In theory, the variety of different types of turbo-expander which can result from the choice between axial flow and radial flow inwards or outwards, together with the distri­bution of heat drop between the nozzles and the blades, is almost unlimited.

In practice, two main types have been adopted : - The radial, inward-flow expander, in which the heat drop is about equally divided between the nozzles and the blades, and which can therefore be classed as a 50 % reaction turbine.

Fig. 5 shows a wheel of this type.

Fig. 5. Radial inward-flow turbo-expander wheel

- The axial flow expander, which is fundamentally an impulse turbine, but in which a small amount of reaction (about 10%) is often allowed, to limit entry into the liquid region during expansion.

Fig. 6 shows the rotor of a two-stage axial flow machine.

Fig. 6. Two-stage axial flow turbo-expander rotor

Although very different in conception, and to a certain extent in operating characte­ristics, these two classes of machine have much in common.

The ideal velocity Co of the gas, corresponding with the isentropic heat drop L1 1 is given by

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K being a constant equal to 91.5 when velocity is measured in m/sec. and heat drop in cal/kg, or equal to 223.7 for ft/sec. and BTU/Lb units.

As for steam turbines, optimum efficiency can be expected when this ideal velocity bears a certain relation to the tip speed Ut (radial expander) or pitch-line speed UP (axial ex­pander).

In the radial machine, Ut/Co should be about 0.65, whereas in the axial machine it should be between 0.35 and 0.45, according to specific speed and degree of admission.

At the same time, the parameter Ut or UP is limited by mechanical considerations to 150 - 250 m/sec. (500 - 800 ft/sec.) according to design and material of the rotor. From the foregoing it follows that the maximum theoretical heat drop per stage which can be allowed without loss of efficiency is about 17.5 cal/kg (31 BTU/lb/) for the radial machine, and 60 cal/kg (108 BTU/lb) for the impulse turbine.

In practice, these figures must be modified for the following reasons : - In the radial expander, the optimum tip speed can be exceeded considerably without appreciable loss of efficiency. - The specific speed of the machine has a considerable influence on efficiency, partic­ularly below a certain limit. Since the heat drop features in the denominator of the spe­cific speed formula :

N Qo.5 N Q o.5 Ns = H O.i5 = (Lil) o-:75

increasing the heat drop excessively can, according to the capacity, shift the specific speed into a region of inefficient operation since the rotational speed cannot be varied beyond certain limits depending on other design considerations.

In general, therefore, extreme values of heat drop (high or low) can only be handled efficiently when accompanied by corresponding extremes of capacity. - The true heat drop is related to the ideal heat drop by the efficiency. In machines well adapted to operating conditions, an efficiency of 83 to 85 % can be expected from a radial expander, and 65 to 75 % from an impulse expander.

Since the designer's hands are not as free as he would wish in this respect, good adaptation of an expander to the service conditions may at times require taking liberties with the latter. If the process permits such a change, the improved performance may well justify it. As pointed out before, a minor improvement in component performance will result in considerable increase in overall cycle efficiency.

Operating range

From the foregoing brief summary of basic design considerations, the difficulty of indicating the range ofrational application of turbo-expanders will readily be apprecia­ted. Moreover, although the following figures represent present conditions, they are in no sense final and the field of use of these machines is still being extended.

Isentropic heat drop, per stage 1 to 30 cal/kg (2 to 55 BTU/lb) Shaft speed * 3000 to 25,000 rpm Shaft power 5 to 500 kW Capacity at exhaust conditions 5 to 5000 m3/min. (3 to 300 ft3/sec.)

* Very small special purpose units operate above 25 000 rpm. It is worth noting that in respect of one of the main parameters of interest to a process,

the exhaust capacity, turbo-expanders are complementary to reciprocating machines, extending the capacity range of the latter upwards. On the other hand, the second important parameter is the heat drop, and in this respect the behaviour of the two classes of machine is very different. The temperature drop in the reciprocating machine is deter­mined by the pressure ratio and the ratio of the specific heats. Consequently, under simi­lar pressure conditions, the temperature drop will be the same when two such different diatomic gases as nitrogen and hydrogen are handled. The heat drop per kilogram will however be about 14 times as much in the second.

In the case of the turbo-expander, however, a machine designed for one of these gases will be quite unsuitable for the other owing to the difference in molecular weights and

576

III-11

specific heats ; the heat drop per unit of mass being the determining characteristic of the machine. Radial flow versus axial flow

There is considerable overlapping in the above range between radial flow and impulse machines. Choice will be determined by :

- Efficiency The radial flow expander achieves optimum efficiency when the specific speed,

calculated from its constituent parameters expressed in RPM, m3/sec. and metres, exceeds 25, and the impulse expander at specific speeds of 12 and over. Note that the corresponding specific speeds, calculated from RPM - ft3/sec. - feet head units are approximately 60 and 30 respectively.

- Percentage condensation The radial inward flow machine is unsuitable for operation in the saturation region,

whereas the axial flow machine will handle up to 15 % liquid. - Heat drop The axial flow machine will maintain good efficiency for higher heat drop per stage

than the radial flow, and furthermore lends itself more readily to multi-staging. - Minimum dimensions If flow conditions are such that the entry width of a radial expander should be less than

8 mm, it is preferable to use an axial flow machine, with partial admission if necessary.

Construction details, Radial flow machines

The radial inward flow expander has a certain resemblance to a centrifugal compressor with reversed flow. This direction is adopted in preference to outward flow since it enables the necessary larger exhaust area to be obtained quite easily, particularly when the wheel is overhung.

High operating speed calls for exceptional balance, close bearing clearances and high rigidity of the component parts. These qualities are obtained by the use of special patterns giving good surface finish, even in the inaccessible passages, dynamic balancing at high speed, and a shaft dimensioned to keep the first critical speed well above the maximum operating speed.

The radial machine lends itself very well to direct drive of a centrifugal compressor, the rotor of which can be overhung and mounted on the opposite end of the common shaft (Fig. 7).

Fig. 7 . Monobloc radial expander and centrifugal compressor unit

Axial thrust is slight, almost negligible in fact if partially balanced by the compressor thrust, so that a simple shouldered bearing is sufficient. Lubrication does not offer the

577

111-11

same difficulties as in a reciprocating machine, one bearing being at quite normal tem­perature and the other being sufficiently far from the exhaust gas to avoid running too cold. The oil reservoir can be heated if necessary. Sealing at each end is by labyrinth rinp;s, which can be pressurized to prevent oil leaking into the gas.

Load regulation is achieved by inlet guide vanes which can be operated manually or automatically.

Construction details, Axial flow machines

The basic requirements are very similar to those of the radial expanders. The pattern of flow and shape of casing, however, favour mounting the wheel or wheels between bearings rather than overhung, and it is also more convenient to utilize a coupled drive to the power absorber.

Axial thrust, though a little greater than in the radial machine, is still low enough to be taken by a simple bearing since the machine is basically an impulse turbine.

The expander is centre-line mounted to eliminate coupling mis-alignment due to thermal expansion.

Load regulation is effected by varying admission, certain nozzles being controlled by valves which may be operated manually or automatically.

Safety devices

All turbo-expanders are fitted with low and high speed and low-oil pressure trips which usually shut down the machine by closing the main inlet valve. Temperature tappings are provided in the bearings and can be used either for informative purposes or to shut­down the machine in case of trouble. Most machines are fully protected for outdoor use and can be operated continuously without attention.

Power absorption Reference has already been made to power absorption by a compressor. This is the

simplest possible method, since the speeds of the two machines can often be adjusted by suitable design to permit direct drive. Generator drive through reduction gearing can also be used.

Foundation and piping requirements Turbo-expanders and their power absorbers are highly compact, often supplied as

packaged units, sometimes on sprung frames, ready to be connected up to piping (Fig. 8).

Fig. 8. Packaged turbo-expander and centrifugal compressor units

Normal foundations, as used for any other rotating machinery devoid of shaking forces, are quite adequate, but expansion joints must be used in the piping to isolate the relatively light casings from expansion and contraction stresses, particularly at start up.

CONCLUSION Reciprocating and turbo-expanders have a definite role to play in industrial low tempe­

rature refrigeration. They are compact, reliable machines capable of prolonged continuous

578

111-11

duty with little or no attention, and when incorporated in a Claude cycle for direct re­frigeration of the working fluid allow use of a much simpler system than with conventional vapour compression cycles used in cascade, thus reducing capital investment and at the same time simplifying load regulation and giving better overall efficiency.

However, every refrigeration system has its advantages and its drawbacks, and it is hoped that this brief summary of the characteristics of expanders will help them to be seen in their true perspective.

579

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Production du froid artificiel par la detente du gaz naturel

Production of artificial cold through the expansion of natural gas

Dipl.-Ing. V. MUNTEANU Service d'Installations Frigorifiques de l'Institut Proiect-Bucuresti, Bucarest, Roumanie

SUMMARY. The A discusses the possibility of obtaining mechanical and electrical power, as well as artificial cold, by expanding natural gas in a turbine prior to its further use. The process is intended to replace the usual throttling-pressure-reduction method.

The paper contains: theoretical development, possible applications and figures for power and cold production, as well as an analysis of the factors on which these processes depend, i. e. initial gas temperature, ratio between initial and final pressures, and the nature of the expansion cycle.

The possibility of obtaining very low temperatures by means of a multi-stage expansion plant is also mentioned.

INTRODUCTION Quelques pays d'Europe possedent de nombreux et riches gisements de gaz nature!

dont !'exploitation enregistre un developpement croissant (Fig. 1).

IMDD 16000 14000 12000 fOOOO 8()00

1952 1954

R0Uf1RNll.

r<JN£CW0&0VlltlLJIC f"tU1N<Y£ ()/JT/;IC!W£ &LGIQu'

Fig. 1 . Variation des quant1tes de gaz exploitees en quelques pays de !'Europe.

Dans ces gisements le gaz se trouve a des pressions parfois tres elevees, atteignant en Roumanie des valeurs de l'ordre de 100 at. qui representent d'enormes quantites d'ener­gie, dont !'utilisation complete fait l'objet d'etudes approfondies. Cette energie est a l'heure actuelle utilisee pour la circulation du gaz dans les reseaux de transport a distance et parfois pour entrainer le gaz des gisements a pression plus basse a l'aide de compres­seurs a jet de gaz montes dans les installations d'elevation de pression.

Le transport du gaz s'effectue a des pressions suffisamment elevees pour ne pas recla­mer de trop grandes sections de tubes. A la suite de reductions de pression et de pertes par frottement dans les tubes, le gaz arrive habituellement chez les grands consomma-

581

IIl-44

teurs sous des pressions comprises entre 8 et 15 at. La distribution interieure chez les consommateurs se fait d'habitude a max. 2 at; a cette fin on pratique a l'entree dans le reseau un abaissement de pression par laminage a l'aide de reducteurs de pression a soupape.

LE NOUVEAU PROCEDE

Si, au lieu de la reduction de pression par laminage on fait subir au gaz dans une machine une detente avec production de travail exterieur, on peut obtenir le gaz a la meme pression finale que par laminage, mais avec une baisse sensible de temperature par suite de la diminution de l'enthalpie.

La production d'energie sortant du cadre de notre preoccupation ne sera mentionnee qu'en passant. D'apres les informations detenues jusqu'a present, la recuperation d'ener­gie par la detente du gaz est utilisee dans les centrales thermo-electriques de Tavazzano (ltalie) et Lacq-Artix (France). Dans cette derniere, afin d'augmenter la quantite d'ener­gie obtenue ,le gaz est detendu apres un rechauffage prealable de sorte que les tempera­tures finales sont suffisamment elevees.

La production et !'utilisation du froid par la detente est un cote de la question qui n'a pas encore ere etudie.

Soient p1 et T1 la pression et la temperature du gaz au point d'entree dans !'installation d'abaissement de pression par detente et p2 la pression finale, dictee par le reseau de distribution; ii s'ensuivra une temperature finale T2• Si la paire de valeurs p1 et T1 sont suffisamment eloignees du domaine de saturation, le gaz peut etre considere comme etant un gaz parfait et suivra la Joi :

n-1

�: = �:) -� (1)

La temperature finale T2 en resulte :

(2)

Le gaz obtenu a cette temperature T2 constitue une source de froid. En le faisant passer par un echangeur de chaleur, le gaz se rechauffera par pre!evement sous pression constante p2 de la chaleur d'un agent intermediaire; la temperature finale sera T3•

La plus simple installation de ce genre est celle de la Fig. 2 qui comprend une turbine de detente M couplee a un consommateur ou transformateur d'energie G (generateur

582

H G

e

- nerhane i t

- · - agent inlermed1a1f-e Ftg. 2

Fig. 2. Schema de principe d'une installation a turbine de detente de gaz.

Fig. 3-a Representation de !'evolution en diagramme T-S pour le gaz parfait.

Fig. 3-b Idem pour le gaz reel.

111-44

e lectrique, compresseur etc.) et un echangeur de chaleur E. La Fig. 3 represente !'evolu­tion en T - S du gaz considere soit comme gaz parfait (3 - a), soit comme gaz reel (3 - b).

En analysant !'expression (2) de la valeur de la temperature T2 apres detente, on trouve les conditions necessaires pour obtenir de basses temperatures finales :

- La temperature T2 etant proportionnelle a la temperature initiale T1, il s'ensuit que de basses temperatures T 2 peuvent etre obtenues en partant de temperatures initiales T1 convenablement basses.

- Avantage d'une detente pousee (une grande valeur du rapport f3 = p \ Pour une P 2

pression d'entree p1 obligatoire, la valeur de f3 croit lorsque la pression d'utilisation p2 diminue. Par contre pour une pression d'utilisation p2 donnee, f3 croit pour des pressions de transport p1 croissantes.

- L' exposant polytropique « n » do it etre proche de I' exposant adiabatique « " » pour que la detente se rapproche le plus possible d'une detente adiabatique.

Une detente adiabatique n'est pas realisable dans une machine reelle : on ne peut obtenir qu'une detente polytropique avec n < "· La temperature finale T2 reelle sera ainsi plus elevee et l'ecart des temperatures moindre que pour une detente adiabatique. II faut done definir le rendement adiabatique de la detente :

'f/ ad 1 - (�) �� 1 - (�) "�l

La temperature finale reelle T 2 sera :

T2 = Ti [ I - 'Y/ad (I - (: :r�l ) ] ou plus simplement :

OU f3 = P 1

P 2

T2 = Ti [I - 'Y/ad (I - p-m)l n-1

et m = n

(3)

(4)

(5)

La variation des temperatures finales T2 pour differentes temperatures d'entree T1 en dependance des valeurs de f3 apparait en Fig. 4 (On a estime un rendement 'f/ad = 0,8)

583

III-44

'"" •w

l\ 2 'r" � i'\. � �[\ ' ' "'" II..

�[\ ..... ' r--... I'... � '2() II.. ""'" ....._ -\[\ ' ["".. � ": � ' I'... -"'" "- � -

\ ..... ' ["".. -.....: � � r---.. I'--. ""'" --

' ["".. -i-- � � r--.... r---.. I'--. -- -..... i'... � - �

i,O ..... � �· - -- �

120 --�--

f 2 j 4 5 5 7 8 9 ID II 12 13 14 f5 Fig. 4. Variation des temperatures finales T. pour differentes temperatures d'entree T, du gaz

en dependance des valeurs du rapport des pressions.

APPLICATION PRATIQUE

Le gaz nature! de Roumanie contient du methane (CH4) presque pur, pour lequel:

n = 1.32

n - I m = -- = 0,2424

n

YN = 0,716 kg/Nma

Cpm = 0,516 kcal/kg° C a 0°C

Dans une installation realisee selon le schema Fig. 2 la quantite de chaleur absorbee dans l'echangeur sera :

Qr = G . CPm (Ta - T2) [kcal]

ou G est la quantite de gaz en Nm3• Pour le methane, cette relation se simplifie comme suit :

Qr = 370 (T3 - T 2) [kcal/1000 Nm 3]

(6)

(7)

d'ou il apparait clairement que la production de froid Q sera d'autant plus elevee que l'ecart (T3 - T2) sera grand. La temperature de sortie Ta depend de l'echangeur de chaleur et s'approche de la temperature initiale T1•

Les productions de froid Qr en dependance de valeurs de €._i et des temperatures P 2

initiales T1 apparaissent en Fig. 5 (pour 1J a d = 0.8 et T3 = + 2° C).

584

Q; Kcalfh IIl-44

60 L--�����������������

0

Fig. 5. Productions de froid Qr pour differentes temperatures d'entree T, du gaz et differentes valeurs du rapport des pressions.

La production d'energie sera :

YN . R . T1 N =

367 m (l - (J·m) 1'/i . 1'/m [kw] (8)

Pour R = 52,9 kg . m/kg, °C et m = 0,2424 l'expression devient

N = 0,426 T1 (1 - (3·8•2424) 1'/I . 1'/m (9)

Les productions d'energie apparaissent en Fig. 6,d'ou l'on deduit que les productions d'energie N augmentent avec la temperature initiale T1 tandis qu'au contraire les pro­ductions de froid Qr (Fig. 5) diminuent.

Quelques chiffres pratiques vont illustrer l'avantage economique du procede de de­tente :

Une grande usine de produits chimiques utilise regulierement 50.000 Nm3/h gaz tant comme combustible que comme matiere premiere pour la fabrication. Le gaz arrive avec

Pression d'entree p1 = 10 kg/cm 2 Temperature

Pression du reseau de distribution

Rapport des pressions

II s'ensuit :

Temperature de sortie de turbine

Production d'energie

Production de froid

T1 = +7°C ou 280° K

p2 = 2 kg/cm2

f3 =P 1 = 5 P 2

T2 = -66°C ou207°K

N = 1470 kW

Qr = 1 .290.000 kcal/h

585

III-44

H 60�-�����������������

I 2 3 4 5 6 7 8 9 fO II 12 13 14 15 _!L Pa

Fig. 6. Production cl'energie.

Si dans la meme usine le rapport p 1 augmente a fJ = 10 (par diminution de la pression P 2

d'utilisation a 1 kg/cm2, OU augmentation de la pression d'arrivee a 20 kg/cm 2) les valeurs deviennent:

Temperature de sortie

Production d'energie

Production de froid

PROBLEMES CONSTRUCTIFS

T2 = -89°C ou 184°K

N = 1 .950 kW

Qr = 1 .740.000 kcal/1

Theoriquement, la detente peut etre realisee indifferemment dans une machine a piston ou dans une turbine. En pratique les machines a piston sont exclues a cause des grands debits de gaz a detendre, qui entrainent des dimensions excessives pour les machi­nes dont les problemes de lubrification et de construction sont tres difficiles a resoudre. La detente du gaz reste done un domaine reserve aux turbines.

Les problemes de construction d'une turbine pour la detente du gaz sont les memes que pour une turbine de detente d'air qui sont a l'heure actuelle resolus avec grand succes.

L'emploi lirnite des turbines a detente d'air n'est pas du a des complications d'ordre technique mais exclusivement au faible rendement des installations avec cycles a com­pression d'air.

II resulte des diagrammes du methane et des essais effectues que, pour la gamme des pressions usuelles dans les reseaux de transport de gaz, il est exclu d'atteindre le domaine des vapeurs humides de methane. De meme, la formation de derives du methane est peu probable a cause des grandes vitesses d'ecoulement du gaz.

586

III-44

Seule reste possible la condensation des traces de vapeur d'eau qui pourraient etre entrainees par le gaz. Pour eviter ces condensations qui peuvent devenir genantes on a etudie et mis au point des procedes commodes et efficaces de sechage du gaz avant la detente.

OBTENTION DE BASSES TEMPERATURES

Les basses temperatures necessaires aux procedes industriels peuvent etre obtenues par la detente du gaz a l'aide d'installations simples similaires a celle decrite precedem­ment, OU d'installations a plusieurs etages de detente et de refroidissement.

Une premiere possibilite d'obtenir de tres basses temperatures est de pousser la de­tente jusqu'a la pression finale correspondante a la temperature souhaitee, quitte a obte­nir une pression finale plus faible que celle du reseau de distribution. On est oblige dans ce cas de proceder a une elevation de la pression du gaz par une compression, operation a eviter parce que chere.

La seconde possibilite est de recourir a une installation a plusieurs etages, comme par exemple I' installation a 2 etages de la Fig. 7. Une fraction du debit de gaz se detend dans une premiere turbine M1 et le gaz ainsi detendu et refroidi sert a un prerefroidissement de l'autre fraction de gaz non encore detendu dans un echangeur de chaleur E1• La seconde fraction de gaz refroidi a la temperature T3 sous haute pression p1 se detend dans une autre turbine M 2 et atteint en fin de detente une temperature plus basse que celle de la premiere branche, mais sous la meme pression finale p2• On peut obtenir une large gamme de temperatures finales en variant le rapport des debits de gaz dans les deux branches de I' installation tout en respectant les limites de pression p1 et p2 imposees. Le gaz de la premiere branche peut etre utilise en totalite pour refroidir la gaz de la seconde branche(Fig. 7-a)ou seulement en partie, le reste continuant son circuit par un echangeur de chaleur E2 (Fig. 7-b).

II '•

' llf 6 i'?..r2 "'·'' "' '

T L.

0

Pt, Tr

nil hone agenf mlermed1011 e

r.9 7

f , J._ ,

Fig. 7 . Installations d e detente a 2 etages.

En groupant de cette maniere plusieurs etages de detente on peut obtenir pour le gaz et pour les agents intermediaires utilisees une multitude de temperatures differentes et basses a souhait; les pressions du gaz dans les differentes branches n'ont par contre que deux valeurs : p1 et p2•

La limite inferieure d'utilisation du procede est la temperature de liquefaction du methane qu'on doit eviter pour maintenir la simplicite de !'installation.

587

IIl-45

Design Problems of Supersonic Ejectors Operating as Booster Com­pressors in Refrigerating Systems

Probll:mes de la conception des ejecteurs supersoniques fonctionnant comme compresseurs booster dans les systl:mes frigorifiques

A. PALIWODA, B. Sc. (Eng.) Central Laboratory of Refrigeration ul. Grochowska 84 m 10, Warsaw, Poland

SOMMA/RE. Dans ce rapport on indique les principes de la conception des ejecteurs supersoniques fonctionnant comme compresseurs booster dans les systemes jrigorifiques, ainsi que les principaux problemes qu'ils entrainent. Dans /'introduction on indique les principes de la construction et du fonctionnement de l'ejecteur, on decrit brievement et l'on explique le processus thermodynamique et le processus dynamique du gaz se produisant dans l'ejecteur. L'A. developpe ensuite la methode de ca/cul des principales dimensions geometriques de l' ejecteur et en meme temps etablit une equation generale determinant le coefficient d' ejection; ii poursuit son analyse et examine les differents cas particuliers des ejecteurs. Il donne aussi une equation pour I' acceleration optimum du jet d'aspiration. En conclusion, l'A. presente un exemple de calcul d'un ejecteur fonctionnant comme compresseur booster dans une in­stallation frigorifiques a compression d' ammoniac - completant les calculs avec quelques renseignements pratiques sur la main d'reuvre, /'installation et le fonctionnement du com­presseur a ejecteur.

NOMENCLATURE

a - local velocity of sound, m/s c P - specific heat at constant pressure, kcal/kg ° C cv - specific heat at constant volume, kcal/kg ° C d - diameter, m f - cross-section area, m2

F - function g - acceleration due to gravity, m/s2 G - weight flow rate, kG/h Lli - isentropic drop or increase of enthalpy, kcal/kg, cf. Fig. 2 k - c p/cv - exponent, dimensionless l - length, m M - w/a - Mach number, dimensionless p - absolute static pressure, kG/cm2 Q - refrigerating capacity, kcal/h r - WoB/WmB - optimum ratio of driven jet velocity to average mixed jet velocity at

the entrance of the mixing chamber, dimensionsless s - entropy, kcal/kg ° C t - temperature, ° C U - Go/Gk - ejection coefficient, dimensionless v - specific volume, m3/kg w - jet velocity, m/s a - angle of convergence of jets prior to mixing, deg. a3 - angle of divergence of driving nozzle, deg. a6 - angle of divergence of diffuser, deg. 1) e - ejector efficiency 17d - diffuser efficiency

589

IIl-45

/t - ratio of the increase of enthalpy during isentropic compression in diffuser to the entire increase of enthalpy during isentropic compression in mixing chamber and diffuser, dimensionless, cf. Fig. 2

cp - coefficient of flow, dimensionless

Subscripts :

A, B, C, D, er - main sections of ejector, cf. Fig. k - driving jet L - liquid refrigerant injected into evaporator, cf. Fig. 2 m - mixed jet o - driven (suction) jet 3 - driving nozzle, cf. Fig. 4 - acceleration nozzle, cf. Fig. 5 - mixing chamber, cf. Fig. 1 6 - diffuser, cf. Fig. 1

Compound subscripts : oA, kA, oB, kB, mB etc. - the first letter refers to jet and the second one to the main

section of ejector

Abbreviations:

WAN - With Acceleration Nozzle, cf. Fig. 6a NAN - No Acceleration Nozzle, cf. Fig. 6 b

1. INTRODUCTION

The efficiency and performance of jet compressors and of the systems in which they work depends in a high degree upon the correct design of the ejector itself. The design usually requires an individual approach to each particular ejector taking into account the different operating conditions of each case.

The principle of ejector operating as booster compressor in absorption and compres­sion refrigerating plants was discused so far by a few authors, viz. : [1, 2, 3, 4, 5, 7, 8, 12).

2. THERMODYNAMIC AND GASDYNAMIC PROCESSES

An ejector is shown schematically in Fig. 1, and the thermodynamic processes taking place in the ejector are presented in ln p-i diagram in Fig. 2.

Fig. r. Schematic drawing of an ejector indicating the terms and the notation used in the tex t .

590

IIl-45

enlhalp!I i, ltr:lll/119

I ig. 2. Simplified log p-i diagram showing the thermodynamic and gasdynamic processes taking place in an ejector, and the nomenclature used in the text.

Here are the main processes : PkA (tkA)1 PoA(taA), and PmD (tmn) - condensing, evap­oration and intermediate pressure (temperature) respectively; p1 and Pm c - cf. Fig. 1 and 2; kA-kF and kA-kB - theoretical and respectively actual expansion process in driving nozzle 3 ; oA-N and oA-oB - theoretical and respectively actual expansion process in acceleration nozzle 4; kB --+ mC +- oB - mixing process taking place at in­creasing pressure, entropy and enthalpy in mixing chamber 5; mC-H and mC-mD­theoretical and respectively actual compression process in diffuser 6; J-H - entire isentropic (theoretical) compression process in ejector.

The enthalpies at the most important cross-sections kB, oB, mD and mC may be calculated, respectively, as

i k B = jkA - <p23 • LJjk

io B = io A - <p2 4 • L1 io jkA + U · ioA

imn = - --!+u- (by the ejector heat balance)

im C = imD - µ · L1im/<p26 The velocities at the above mentioned cross-sections will be respectively

W k B = 91,5 · <p3 • v LJ jk

Wo B = 91,5 · <p4 • v-.&;;-

(1)

(2)

(3)

(4)

(5)

(6)

591

IIl-45

91 5 ;--Wm c = -'- l µ . Llim

'P6 The coefficient µ used in (4) and (7) is

µ') = [ 9�:5] 2 · 104 ;L k � l • PmD • VmD

(7)

(8)

The velocities WoA, WkA, and WmD may be established or calculated according to the general principles used in refrigeration piping calculations.

3. EJECTION COEFFICIENT

Ejection coefficient may be calculated for the WAN-type ejectors by means of the equation.

where:

Kk VLJi� - Km V Liim

Km V Llim - Ko V Llio

Kk = <p3 . <p5 . cp6 � 0,834

Km = V� [l + + (1 - p:10)] � 1,0

Ko = <p4 • 'Ps . <p6 • cos a � 0,812 (for a = 0°) ar d respectively for the NAN-types by means of the equation

ua) = K k 1 I Ll�k - 1 Km V Lltm

1) - The equation was derived from the following equality

Wm C = 9;:5 v=-� Llim am C = v 2g k � l�:: · VmD

where: am c ""' F( p . . v, k) for the stagnation condition in mD

(9)

(10)

(11)

104 (A)

The equality (A) is in agreement for the supersonic ejector with the principle expressed by the following formula known from dynamics of fluid flow

(1 - M2) -� = 1 _ (1 - _1_) ��- = - df

(B) w k M2 p f As it arises from analysis of ( B) the section f m c has to be a critical one in case of super­

sonic ef ectors as initial M > 1 converts into M < 1 at section f m c.

2) - Starting-point for derivation of equation (9) is the momentum equation which as related to the cylindrical mixing chamber, characterized by the compression ratio Pm c/P1 may be written in the following form

<p5 [G.!t · Wkn + �o . Wo B . cos a] - Ggm

. Wm c

Vm c 4 = Gm - (Pm c - P1) 10 (C) Wm c

[91,5] 2 µ . Llim where: Vm c = -qi-; . g.k"": Pm c . 104

(D)

Eq. (D) was obtained in similar way as eq. (8) by means of equality (A) - footnote 1 - but here Wm c = am c = F (p, v, k) for the local conditions in mC.

By introducing (5, 6, 7) and (D) in (C) - footnote 2 - one gets on elementary rearrange­ments the equation (9) .

3) - Equation ( 11) is a particular form of the generalized equation (9) . The transforming factor of (9) into ( 11) is the approximate relation L1i0 � 0 which holds good for the NAN­types.

592

III-45

The values of ejection coefficient obtainable by means of (9) and (11) at Km < 1 determined according to (10) prove to be maximal (theoretical) ones and only obtainable when the mixing chamber is cylindrical (as assumed in momentum equation (C) - foot­note 2) and when the compression process in the mixing chamber is void of transversal compression shock. The conditions stated above cannot exist in supersonic ejectors. Therefore, the values calculated according to (9) and (11) at Km < 1 prove to be some­what too high as compared with experimental results. Experiments [13] show a good coincidence of analytical and experimental results when Km = 1 is put into (9). The above experiments were carried out with a WAN-type ejector, but as it arises from the transformation of (9) into (11) they apply also for the NAN-types. Equation (11) has been graphically illustrated in Fig. 3 (for ammonia) by means of the semi-graphic method. The diagram will be discussed at the end of the present paper.

-40 -----1-

-��o---�-�.l;------4J�---__,�o,___ ____ --o!•o evaporating temperature, foA 0c

Fig. 3 . Diagram for estimating ejection coefficient for ejectors of the NAN-type in different operating conditions. It refers to ammonia.

4. ACCELERATION OF DRIVEN JET The purpose of acceleration of driven jet in the WAN-type ejectors is to diminish

the impact loss occurring between molecules of the jets. The acceleration is brought about due to extra pressure drop by Po A - p1 which may be expressed by the enthalpy equivalent .dia. As is seen from (9) the higher .dia the higher the ejection coefficient. Increase of .dia involves, however, increase of pressure drop which the ejector is to restore again during compression. Optimum value of .dia may be calculated according to the follow­ing equation

(12)

4) - Starting-point for determination optimum .dia is the momentum equation for the mixing process, which reads WKR + U . Wo B . cos a = (1 + U) WmB By putting in (E) the eq. (6) and r = won/wm n as well as W k R = 91,5 . <p3 • one gets on rearrangement the eq. (12) .

(E) V i�-;.:.. i;;,

593

IIl-45

optimum velodflJ ratio,� Fig. 4. Diagram for <letermination of the optimum acceleration of the driven jet (r = w0n/wmB)·

The unknown value of U in (12) may be calculated with sufficient accuracy by means of (1 1) or may be determined by the diagram given in Fig. 3. The value of r which would satisfy the condition of maximum ejector efficiency may be easily determined according to the following equation :

(13)

with the corresponding coefficients ab and c given in the form of equations in the left­upper corner of Fig. 4 being the graphical illustration of the eq. (13).

5. DELIVERY CAPACITY

The design of ejector operating as booster usually requires the determination of the ejector delivery capacity, which at the same time, is the refrigerating (suction) capacity of a high-stage compressor co-operating with ejector. General relation between ejector delivery capacity Qm and its suction capacity Q0 is as follows [12] :

1 1 where : Qm = Qo (1 + --) . (1 + -)

U e

ioA - iL e = (1 + U) -. ----.- - coefficient, dimensionless

1kA - 1oA ioA, ikA, iL - according to Fig. 2

5) - Equation ( 13) was derived from the following Kamenev's formula (R. 9) determining ejector efficiency:

1 - r2 (2 - 'Y/d) 'Y/ e = U . ------- (F)

(2 - 'Y/d) (1 + U - r . U . cos a)2 - 1

By differentiating (F) with respect to r and by equating the first derivative to zero [ d 'YJ e/dr = = 0} one gets on rearrangement a quadratic equation of the form ar2 - br + c = 0. One of the solutions of this equation is (13).

594

III-45

6. GEOMETRICAL SIZES

The following set of initial data is the starting-point for the calculation of ejector, viz. : Qo, P kA, PoA, PmD, t'kA, t'oA·

The above data enable us to determine beforehand by means of the standard ln p-i chart the following further data : V kA, VoA, ikA, ioA and Go. A very useful tool in cal­culations of ejectors proves to be ln p-i chart for the refrigerant in question.

All the cross-sections are calculated according to the following well known equation :

f = G . v/w . 3600 (14)

An exception is the critical section of the driving nozzle, which is calculated by means of the equation

l/P kA fer = Gk/3600 . b . �

V kA where : b = 203 as an average value [15] .

(15)

The main task depends now on the determination for each cross-section being cal­culated of the necessary values of G, w and v in eq. (14) and (15) - having at disposal the previously given initial data. Some experimental data indespensable for ejector calculation are given below

Pa (/!5 10° 70 0,95 0,925 0,975 0,90

Longitudinal dimensions are calculated according to the following experimental formu-{ 8 . dm c . . . . . . . . . . . . for NAN-types [14] ; lae: 15 = 12 (dm c - dk B) . . . . . . . for WAN-types [9] ; 1 6 = 6,5 [dmn -dm c] ;

[14] It is to mention that the determination of the refrigerant condition in mC (Fi�. 2)

corresponding to section C*) (Fig. 1) is the main problem of supersonic ejector design, as f m c seriously a1fects the ejector efficiency and performance. The condition of re­frigerant at mC (Fig. 2) is determined by the intersection point of im c and Sm c = SH, where SH may be determined at the intersection of PmD and iH ; (iH = im c + µ . Llim). The value of Lfim is determined according to operating conditions, as difference of enthalpy at the points of intersection of an isentrope passing through, say, oA and corresponding isobars Pmn and p1.

In this way we have discussed and completed the set of equations and data necessary for the calculation of ejector.

7. PRACTICAL HINTS

The ejector may be designed as shown in Fig. 5. A peculiar feature of this design is that it can be WAN-type when the position of the driving nozzle is as in Fig. 6a, or NAN-type when the position of the nozzle is as in Fig. 6b. Determination of the distance 1 3-5 (Fig. 6b) is a very important detail as it substantially affects the perfor­mance of ejector. Although there were attempts of analytical determination of the distance, e. g. [14], nevertheless each such procedure proves to be not only complex but also unsafe. Therefore, a NAN-type design should enable experimental determina­tion (setting) of the distance by trial and error. The design shown in Fig. 5 has been adjusted for that purpose. The question is what type of ejector is better and in what circumstances. WAN- or NAN-type? Well, both types enjoy their followers. The WAN is somewhat more efficient in respect to energy, but its competitor proves to be simpler in construction, and, as experience has shown, it is more flexible in operation, and is distinguished by a higher adaptability for variable operating conditions. There­fore, the NAN-type may be recommended in such cases, where considerable heat load changes occur, e. g. for intermittent freezing, etc. It is also better fitted for the domain

*) Cross-section C (Fig. r ) calculated in the above given way should be enlarged by ro% .

595

III-45

Fig. 5. An ejector design witlt an annular nozzle o! ex = o0• It can be WAN·

or NAN-type depending

on the position of the

driving nozzle, cf. Fig. 6.

of low- and ultra low temperatures. The WAN-type is superior in cases of insignificant heat load changes, e. g. for continuous freezing, for freezer-storage, etc.

As is known, the condensing pressure (temperature) depends first of all upon the ambient. The influence of the condensing temperature on the ejection range has been shown in Fig. 3, from which is seen that the higher the condensing temperature at constant U and tmn, the greater will be the ejection range, i. e. the difference between tmn and toA· The influence of the condensing temperature is more apparent when the ejection coefficient is lower. Consequently, the application of ejectors as boosters is more justified in hot climates, during summer time, as well as in cases of evaporative condensers. In other words in all such cases when a raised condensing pressure results from justified circumstances.

In order to preserve the design conditions during the operation of ejector it is re­commended to keep the condensing pressure constant by means of the appropriate controls.

The ejector does not need permanent supervision and lubrication, its parts are robust, and operation reliable, provided it is well designed and assembled.

596

REFERENCES

III-45

Fig. 6. The positions of the

driving nozzle of the

ejector shown in Fig. 6.

a-WAN type ; b-NAN

type.

r. Badylkes, I. S., On the application of the vapour-jet de,·ice as a refrigerating booster com­pressor. Kholodilnaya Tekhnika. No. r, 1956.

2 . Badylkes, I. S., A two-stage compression ammonia plant eq1.1ipped with a vapour-jet device. Paper presented to Com. 3, 4, 5 of I. I. R. Moscow 1958.

3 . Badylkes, I. S., Practical utilization of vapour-jec device for the production of low temperature in single-stage plants. Paper presented to Com. 3. I. I. R, Budapest r96r.

4. Badylkes I. S., Danilov, R. L., Refrigerating cycle with vapour-jet device in the capacity booster. Gosenergoizdat. Moscow r960.

5 . Badylkes, I. S., Danilov, R. L., Refrigerating systems with vapour-jet devices in the capacity of boosters. Kholodilnaya Tekhnika. No. 4, 1958.

6. Danilov, R. L., Sysowv, L. P . • Experimental investigation of ammonia vapour-jet device operating as booster compressor. Paper presented to Com. 3. Marseilles 1960.

7. Girardin, P., Considerations sur !'utilisation de l'ejecteur comme etage basse pression d'un cycle frigorifique a gaz liquefiables. Paper pre5�nted to Com. 3. Marseilles 1960.

8. Heller, L., Farago, ]., Absorption refrigeration at very low temperatures. Proc. 9 th Int. Congr. Refrig. Paper no. 3. 68. Paris 1955.

9. Kamenev, P. N., Hydroelevatory i drugiye strujnyye apparaty. Mashstroiyzdat. Moscow 1950.

ro. Kroll, A. E., The design of jet-pumps. Chem. Engng. Progr. r , 21, 1947.

r r . Lozkin, A. N., Transformatory tiepla. Moscow 1948.

12. Paliwoda, A., Application of ejectors as booster compressors for freezing purposes. Paper presented to Com. 3. Cambridge 196r.

13. Paliwoda, A., Experimental investigation of the ejector operating as booster compressor in ammonia refrigerating plant. Trans. of Central Lab. Refrig. Warsaw 1962.

14. Sokolov E. ]., Zinger, N. M., Struynye apparaty. Gosencrgoizdat. Moscow 1960.

r5 . Sumelisskii, M. G., Ezektomye cholodilnye mashiny. Gosenergoizdat. Moscow 1960.

597

III-45

DISCUSSION

L. Vahl, Netherlands: In my opinion the theory put forward is a little simplified since there is in fact no sonic region at all. The velocity increases from subsonic velocity to supersonic velocity by a decompression shock and therefore will not pass through a sonic stage, and I should like to know how this theory can therefore be applied in practice.

A. W. Paliwoda, Poland: If you have subsonic velocity in one region and supersonic velocity in the other, then the velocity must go through the sonic velocity at some point. I quite agree that I have not investigated the problem as to where the sonic velocity occurs, and I equally agree that the change in velocity is due to a compression shock. Nevertheless, I feel that this is quite sufficient for practical design purposes. On the contrary I feel that, if I had gone into the calculations of shock, then no practical engineer would have been interested in it, since it would be an exceedingly complicated paper.

598

IIl-42

New Definitions Needed in Refrigeration

Nouvelles definitions necessaires dans le domaine du froid

Prof. E. B. PENROD Mechanical Engineering, University of Illinois, Urbana, Illinois, U.S.A.

SOMMA/RE. On propose d'adopter les expressions: rapport d'energie calorifique HER, rapport d'energie frigorifique GER, force electromotrice de Seebeck E ab, et coefficient de Seebeck Sab = dEab/d t·

For many years the following terms have been used to designate the performance of a refrigeration cycle or system, namely, coefficient of performance COP, coefficient of amplification, and Leistungsziffer e. None of these terms indicate that the concept refer­red to is an energy ratio.

Since 1948, the author has been using the terms heating energy ratio HER and cooling energy ratio CER for designating the performance of a heat pump or refrigeration cycle. It happens frequently in many areas that a heat pump supplies heating in the forenoon, and cooling in the afternoon in a building during the same day [1] . In some areas, how­ever, a heat pump is used for heating only; and in other locations, it is used only for cooling. Thus, it appears that two concepts are really needed in engineering design and analysis of experimental data [2] .

In an excellent paper, Davies and Watts used the term performance energy ratio. They also stated [3],

"In the United States, the term Coefficient of Performance is used both for refrige­ration and for heat pumps, clearly an undesirable custom. Dr. 0. Faber employs the term 'Advantage' (Proc. I. Mech. E., 1946, Vol. 154, p. 144). Mr. J. H. Sumner pro­posed 'Reciprocal Thermal Efficiency' (Proc. I. Mech. E., 1948, Vol. 158, p. 22). The French use 'Coefficient of Amplification'. None of these expressions, however, indicates that a ratio is in question; Performance Energy Ratio is preferred, since what is measured in actual cases is the ratio of those quantities of energy which determine the performance of a heat pump." Since the term suggested by Davies and Watts does not distinguish between the per­

formance of a heat pump when used for heating from that nsed for cooling, it appears that two different ratios are involved, one for heating anc' the other for cooling. It is proposed, therefore, that the following definitions be adopted :

For a refrigeration cycle (or heat pump cycle), the heating energy ratio HER is de­fined by the expression :

heat delivered to the condenser HER = - ·

heat equivalent of the work of compression ' and the cooling energy ratio CER is defined by the expression :

C heat absorbed in the evaporator

ER = heat equivalent of the work of compression ·

The data listed below were obtained from a test of 12-hours duration on an earth 1'eat pump designed and built at the University of Kentucky [ 4]

Carnot cycle Refrigeration cycle Heat pump Heat pump system

HER 5.26 4.24 2.94 2.40

*) Professor of Mechanical Engineering, University of Illinois, Urbana, Ill., U. S. A.

599

IIl-42

These results show clearly that from the customer's viewpoint, more information is needed in addition to the heating energy ratio of the refrigeration cycle.

In analyzing data obtained in the above test, the following definition of terms were used : 1. A compression refrigeration plant is a system consisting of an evaporator, condenser,

precooler, expansion valve (or capillary tube), and a compressor. 2. A heat pump is a refrigeration plant and the electric motor used to drive the compres­

sor. 3. A heat pump system is defined as a heat pump plus the necessary auxiliary equipment,

which consumes electric energy, needed for air conditioning. 4. The HER of a Carnot refrigeration cycle is the ratio of the heat absorbed by the hot

body to the heat equivalent of the net work of the cycle. 5. The HER of a refrigeration cycle is the ratio of the heat absorbed by the condenser

to the heat equivalent of the work of compression. 6. The HER of a refrigeration cycle is the ratio of the heat absorbed by the transport

fluid which removes heat from the condenser to the heat equivalent of the electric energy that is supplied to the compressor-motor.

7. The HER of a heat pump system is the ratio of the heat absorbed by the transport fluid which removes heat from the condenser to the heat equivalent of the electric energy that is supplied to the entire system.

The following equations are given in the International Critical Tables [5] : MER = at + ibt2 (10-2) + 1/3 ct3 (10-5) + d with one junction of the thermo­

couple at 0°C, or MER = d/dt ( MER) = a + bt (10-2) + ct2 (10-5),

where M and R are two metals connected in series to form a closed circuit, and if one junction is kept at temperature, 0°C, and other at the temperature t (t :f 0), there will he a thermo emf around the circuit. By convention, MER is regarded as positive if the current so produced flows from M to R at the junction which is at 0°C.

REM = - MER; MER + aBs = ME s . MQR = o/ot ( MER) is called the thermo­electric power of M with respect to R. Obviously, since power is a time derivative of energy , it is clear that thermoelectric power is an inappropriate term.

Currently, it appears that the Peltier or thermoelectric refrigerator, in the low capac­ity bracket, will have wide acceptance due to the fact that it provides capacity modu­lation, and also requires less space than other types of electric refrigerators. Due to the importance of thermoelectricity in refrigeration, it appears that the following recom­mendations are in order: 1. Adopt the term "Seebeck electromotive force" for the thermoelectromotive force

that is developed when there is a temperature difference between the two junctions of a thermocouple. That is, call Ea b in the equation below, the Seebeck emf. E a b = c1 (t -tr) + ic2 (t2 - tr2) + 1/3 c3 (t3 - tr3), where tr is the reference

junction temperature. 2. Adopt the term "Seebeck coefficient" for the temperature derivative of E a b in the

above equation, where the Seebeck coefficient is defined by S a b = dE a b/dt = = c1 + c2t + c3t2, microvolts per degree C.

REFERENCES

r. Penrod, E. B., "Earth Heat Pump Research'', Proceedings of the Midwest Power Conference, 12 , 3751-394, April 1950.

2. Penrod, E. B., "Theoretical Analysis and Performance Characteristics of a Peltier Refrigerator", Institut International Du Froid Bulletin, XL, (2), 517-539, April 1960.

3. Davies, S. ]. and Watts, F. G., "Fuel and Power Economy with Special Reference to Heat Pumps", Engineering, 166, 285-287, 3051-312, 352-353, July-December 1948.

4. Penrod, E. B., Knight, R. B., and Baker, M., "Earth Heat Pump Research. Part II", Univer­sity of Kentucky Engineering Experiment Station Bulletin, 5, (18), l-ro4, December 1950.

5. International Critical Tables of Numerical Data, page. 214, Vol. 6, Mc Graw - Hill Book Company (1929).

600

Thermoe l ectric Refrig erat i o n , Absorption

Refro i d i s s e m e n t t h ermo­e l ect riq u e , a bsorption

III-17

Performance of a Thermoelectric Refrigerator as a Function of Characteristic Parameters

Rendement d'un refrigerateur thermoelectrique en fonction de paramerres carac­teristiques

Prof. E. B. PENROD Department of Mechanical Engineering, University of Illinois, Urbana, Illinois, U.S.A.

SOMMA/RE. On considere un refrigerateur thermoelectrique tres ideal, consistant en un seul thermocouple, et dans lequel on neglige les pertes de chaleur et l'effet Thomson. On introduit un parametre qui montre l' eff et de coup/age des deux termes de la puissance frigori­fique irreversibles de Fourier et Joule dans les elements thermoelectriques. Ce parametre de couplage est utilise pour montrer que la puissance calorifique Joule ne se divise pas generale­ment en une moitie a la soudure chaude et l' autre moitie a la soudure froide comme il l' a ete publie.

Des effets interessants se produisent sur le rapport d' energie frigorifique, le refroidissement net, la temperature maximale et sa distance de la soudure chaude par variation du parametre de couplage. On considere egalement les effets sur le rendement du systeme en maintenant constante la surface de section transversale d'une branche tandis que /'on fait varier celle de l' azure pour diverses differences de temperature entre les soudures chaude et froide.

In 1909, Altenkirch inferred that about one-half of the internal Joule heat goes to the warm junction of a thermoelectric refrigerator [1].

Three years later he stated that "the influence of this Joule heat extends, as may be very accurately assumed if slight temperature difference exists, to the extent of one-half to the cold junction and the other half to the warm junction" [2] . It should be recognized, how­ever, that when the temperature difference between the junctions is very small the corres­ponding refrigeration will be negligibly small.

When an electric current is sent through a thermocouple, when the junctions are at the same temperature, the system is in fact a trivial refrigerator that merely transfers heat from a reservoir at one temperature to another at the same temperature. Subsequently it will be shown that when the temperature difference between the junctions is zero ( L1 t = O) one half the Joule heat power is transferred to each junction, and on the other hand it will be shown that when the temperature difference is not zero ( L1 t :j: 0) one-half the Joule heat power does not appear at each junction.

d't /'(! The Heat Equation : dx'

- + kA' = 0

The following analysis applies to a thermoelectric refrigerator consisting of a single couple in which the prescribed conditions are : the junction resistance is negligibly small ; steady state conditions prevail ; material properties are non-temperature dependent ; no heat power is transferred to the surroundings through the lateral surface of the thermo­couple arms ; the Thomson coefficient is negligibly small ; and the thermocouple arms are of equal length and constant cross sectional area.

601

IIl-17

fx + t.X I I I I

Fig. r. Section of an arm of the thermoelectric refrigerator shown in .Fig. 2.

The heat power balance for the section shown in Fig. 1 is

where

and

. . . Q1- Q2 + dQJ = 0,

• dt Qi = -- kA dx' . . 8Q1 Q2 = Q1 + ax Llx

l"e Llx A

Combination of the above equations yields

d't I'e ax• + kA-. = 0

·

(1)

(2)

(3)

(4)

(5)

To determine the temperature at any point distant x from either Junction, a particular solution of (5) is sought subject to the boundary conditions

t = th for x = 0,

t = tc for x = L, where th = tc and Lit = th - tc = 0. The particular solution is

t = t + �'_f!L� (_:_ - �) c 2 kA' L L2 The temperature gradient is

dt l'eL ( x ) dx = 2 kA• l - 2 £ .

dt Let x = Xm, t = tm for dx = O, then the maximum temperature is given by the

equation

J•ev tm = tc + S kA"

(6)

(7)

(8)

L and occurs at a distance Xm = -2 from either junction. Multiplication of (7) by -kA

gives dt -kA dx

602

_ l' eL (1 _ 2 _:_) 2A L .

Hence,

where

• 1 ( x ) • Qx = 2 2 y; - 1 QJ,

dt . l'eL Qx = -kA dx

and QJ = A- .

IIl-17

(9)

Therefore, for x = 0 and Ll t = O,

Q x � o (10)

and for x = L and Ll t = O,

(1 1)

Equations 10 and 11 clearly show that when no refrigeration is produced one-half the Joule heat power is transferred by conduction to each junction [3, 4, 5] .

THERMOELECTRIC REFRIGERATION: Li t :j: 0 For this case a particular solution of the heat equation (5) is sought subject to the

boundary conditions (Fig. 2 and Fig. 3),

HOT JU'CTION

' •c

10"-----I----

O I 1 .0 21 x,cm fx 1 lit .A le

�:......._ ___ ..__;_�-�-

' ' X + A X

I p

Fig. 2. Diagram of a thermoelectric refrigerator consisting of a single couple, including graphs showing the variation of tempera­ture along the two arms for optimum area ratio Ab/2Aa2=ka(!b/kb(!a· The upper and lower graphs are for Lit = o and

Ll I= 20 °C, respectively. The Peltier em f's

are equal for Lit = o, i.e., :n:abh = :7tabc for lh = le ; and the Peltier heat power terms at the junctions are also equal,

i. e., Qabn = Q�bc· (See Fig. 3)

Fig. 3. A thermoelectric refrigerator consisting of a single couple in which the thermal insulation material is not shown. For Ll t :j: o, Qabh and Qabc represent the Peltier heating and cooling at the junctions, respectively ; and Qabh = I :n:abh and Q0abc = I :n: abc· Q0 is the net refrig­eration produced and Qh is heat power transferred from the hot junction to a transport fluid.

603

III-17

t = th, for x = 0, t = tc, for x = L,

where Llt = th - tc [6, 7] . Introducing the dimensionless ratios C x/L, the heat equation transforms to

d'C d-+ (J = O, where

17' f' eL

A L1 t

kA T

irreversible Joule heat power irreversible Fourier heat power

The general solution of (12) is 1

C = C2 + C1 17 - j·- (J r1', or

t x 1 x• L1 t = C2 + C1 L - -2 fJ I•

Applying the boundary conditions to (14) gives

t = th + L1t [(L- 1) -"--L �] 2 L 2 L' '

( x ) J• eL' ( x x' ) or t = th - Llt L + 2 kA' i - j,2 .

t/ Llt and 17

(12)

(13)

(14)

(15)

(16)

Equations 15 and 16 express the temperature at any point in a thermocouple arm as a function of x, material properties, and parameters that describe the geometry of the sys­tem.

Using a combination of thermoelements of nickel and copper, Fig. 4, Whitlow verified Eq. 15 by sending a current of 20 amperes through the system [8] . When steady state conditions prevailed he recorded temperatures of28.5 and 25.5 C at the warm and cold ends

Fig. 4. Experimental arrangement of a nickel-copper thermoelement used by Whitlow to verify

equation t = th + Llt [ ( ?- r) � -? �: l (Courtesy of John B. Whitlow and the Uni­

versity of Kentucky.)

of the nickel rod. Maximum temperature occurred at a point in the rod 16.33 cm. from the warm end. The warm and cold temperatures were interchanged on reversing the direction of the electric current, and maximum temperature again was observed to be at a point 16.33 cm from the warm end of the nickel rod when steady state conditions ob­tained. In both trials the maximum temperature occurred at the position predicted by Eq. 15, however the value was slightly lower than the calculated value of 28.56 C. The difference between the observed and calculated maximum temperature was attributed to heat leakage through the insulation material.

604

III-17

It is of interest to know that Eq. 15 applies to the thermoelectric generator as well as the thermoelectric refrigerator [10]. The equations below were derived from Eq. 15 and are applicable to a refrigerator consisting of a single thermocouple [ 6, 9] .

at =

Ll t (L_l x - 1) (17) dx L 2 L

(18)

(19)

(20)

(21)

(22)

(23)

By varying the magnitude of the parameter {J, the above equations are useful in deter­mining the performance of the thermoelectric refrigerator under consideration. When x =

L/2, Eq. 20 shows that the conduction heat power at a point half way between the junc­tions is always equal to the Fourier heat power for all values of fJ as shown in Table 1 .

.. .g]X•D *l•L Q • Q -i o •"ci Qll• O o .... Q,.,

( 6f" O ) t + 00 6, 00 L JlPL1 +..ff.b. _ 1¥L � 6r•O � ( t.t = O ) T 111+ 8 kAi 2kA2 2 kA2. Q, •o 6,•o

L tll•* + !.!. _ ,M! - t o., i 6,, •tO,, '6 2L 2L -ta, ci, •t6,

'• -&!.! 0 .!.Q,, ci, L 0 Q, + 2 Q,.

1.5 L •• ., 74! •tci, tQ, •-i6,, -. t11 • 48 -4[ - TL +.J.Q,. ci, +J.Q,

1.0 L 111 +� ., .. , •f Q,, ci, •ta .. -. - rr -rr •t ci, Q, •tO,

0.5 - �L t� + ��t ••• _gi •1 6,, 2 6 .. �a. - TL 4L � <i. <i, 'l<i, .. ., Q,,•o o, � - T - T <-----7-- Q, (A t P O ) Q,

(tr. t • O ) Table I . Variation of maximum temperature, temperature gradient, Joule heat power transferred

ot the junctions of a thermoelectric refrigerator versus the coupling parameter {J.

If the Joule heat power is twice the Fourier heat power, fJ = 2, then the conduction heat power is equal to the Fourier heat power, or one-half the Joule heat power when x = L/2. Thus, when the system operates under these conditions the maximum tem­perature in the arm occurs at the hot junction and is equal to th, and no Joule heat power is conducted to that junction.

When the Joule heat power is three times the Fourier heat power, fJ = 3, Xm = L/6, then one-sixth of the Joule heat power appears at the hot junction and five-sixths goes to the cold junction. Maximum temperature occurs at a distance of L/6 from the hot junction and is equal to th + At/24. For 0 < fJ < 2 there are no real maxima of temperature in the

605

III-17

thermocouple arm ; also, Joule heat power is conducted away from the hot junction a� il­lustrated by data listed in Table 1 .

EFFECT OF VARIATION OF PARAMETERS ON PERFORMANCE Calculations were made to study the effect on refrigeration capacity, maximum cooling

energy ratio, and operating current by changing the geometry of the system. In Fig. 5, I, maximum CER, and Qc were plotted against A b/A a. It is apparent that the refrigeration

0 N

. 0

� -

I

� �

\

CUR RE.NT AND MA)( I MUM CER .

� 1\ \

\

� � �

1� I

. m

I\

\ \ .-' I\\ ,-,,. \

- ·--

f\ I\

-

I\

o e p 9 � N iii � REFRI GERATION 0c

;;

\ Fig. 5. Variation of refrigeration capacity,

maximum cooling energy ratio, and operating current versus Ab/ Aa for a thermoelectric refrigerator consisting of a single couple for a figure of merit of I .93 x ro·• c-•, tc = 0 °C and Llt = 5 °C.

capacity and current requirement increase linearly with increase of A b/A a, while maxima CER increases until the ratio of the cross sectional area is 1 .5 after which it gradually de­creases. Increase in refrigeration capacity may be obtained therefore without sacrifice in efficiency by increasing over a broad range in the variation of ratio of the cross sectional areas.

For A a = 1 cm", contour curves, Fig. 6, were obtained for several values of A b by

plotting Qc versus I. As shown in Fig. 5 also, maximum refrigeration capacity increases with increase of current.

2.a�-�--�-�--�-�-�--�

� � � ; 2 of----+---f---*"0,...""'1F==:...._,-J--"-M-->.,� .J

� � � 1.6 j � L2 f---+--1-b----+---<f--'r--+--T-<-..,_.,

� � o el----h'--l----+--"-'1----+'<-----1-\-___, � � <r

606

140

Fig. 6. Contour curves showing the varia· tion of refrigeration capacity versus current for eight different couples, like the one shown in Fig. 3, in which the cross section of arm "a" is 1 cm2

and that of arm "b" was changed from 0.5 to 4cm2. The figure of merit was I .93 x 10·• c-1, tc and Lit are o °C and 5 °C, respecticely.

III-17

Contour curves are shown in Fig. 7 for different values of Ab in which CER is plotted versus I. For each value of Ab the cooling energy ratio increases with increase in current until it reaches its maximum value, after which it decreases to zero with increase in current ;

� 31--�-U---J-_J_l--_J_-"J-��'--�-1--�--'�---" u

oo�����--+ 1 0 1 2 J, AMPERES

Fig. 7. Contour curves showing the variation of the cooling energy ratio versus current. Here Aa = I cm2, and Ab = 0.5, r .5, 2.5, and 3.5 cm2 and the figure of merit, cold junction temperature, temperature difference between the junctions are r .93 x ro-• c-1, o °C, and 5 °C, respectively.

further increase in current produces a negative cooling energy ratio, i. e., heat power is transferred to the space to be cooled. Maxima CER increase with increase of Ab until Ab/A a = [k aQb/kbQa] t = 1.4993, after which there is a decrease in their values. For a system having the geometry L a = Lb = 2 cm and Ab/A a = 1.4993 and material proper­ties yielding a figure of merit of 1. 93 x 10-3 c-1, the maximum CER and the corresponding Carnot CER are 5.33 and 54.6 respectively when the supply current is 4.74 amperes ; under these conditions the cooling capacity and coupling parameter fJ ab are 0.132 watts per couple and 0.171, respectively.

In Fig. 8 the temperatures in the two arms of the thermoelectric refrigerator are plotted against x for Ab/A a = 4 and I = 120 amperes. The hot and cold junction temperatures

;> 50 w "' ::l r-� 40

" w r-

04 0 8 1 2 X,Cm

16 2 . 0

Fig. 8. Variation of temperature versus distance from the hot junction (LI t :j: o) in the two arms of a thermoelectric refrigerator in which Aa = I cm2, Ab = 4 cm2 for Zm = r .93 x ro-• C-1, fla = 109.14, {lb = 1 5.33, Li t = S °C, le = o °C, and Qc = 2.313 watts.

607

IIl-17

are 5 and 0 °C, respectively. Operating under these conditions, 2.313 watts of cooling are produced at a cooling energy ratio of 0.284.

Two sets of contour curves are shown in Fig. 9. One set (dashed lines) is for Lit = O and th = tc = O, 20, 40, and 70 C. For this case the optimum current increased from 61.69 to 77.50 amperes and the corresponding heat pumping capacity increased from

X , C M

Fig. 9. Variation of temperature versus distance from the hot junction for

Ab/A a = [kaQb/kbQa] l = I .4993, that is for optimum current I 0' and f3ab= f3a = f3b· (Refer to references 6 and 9). Dashed curves are for Lit = o and th = le = o, 20, 40, and 70 °C. Solid curves are for le = o °C, th = 20, 40 and 70 °C, and LI t 20, 40 and 70°C.

1 .69 to 2.66 watts. In the other case tc = 0 °C and th = 20, 40, and 70 °C. For this case the optimum current is 61 .69 amperes for all values of th and the corresponding refrigeration decreased from 1 .218 to 0.05 watts.

The analysis given above illustrates how the theory presented here and elsewhere can be used for investigating the performance of thermoelectric refrigerators under different conditions. A procedure has been introduced for determination of the effect on perform­ance characteristics by changing the geometry or material properties of the system.

REFERENCES

r. E. Altenkirch, Uber den Nutzeflekt der Thermosiiule. Physikalische Zeitschrift, ro, 560-568, 1909.

2. E. Altenkirch, Elektrothermische Kalteerzeugung und reversible elektrische Heizung. Zeit­schrift filr die gesamte Kalte-Industrie, 19, 1-9, 1912 .

3. A. F. lofje, Semiconductor Thermoelements and Thermoelectric Cooling. Infosearch Limited, London, 1957·

4. F. E. Jaumot, Jr., Thermoelectric Effects. Proc. Institute of Radio Engineers, 46, 538-554, March 1958.

5. I. B. CadofJ and E. Miller, Thermoelectric Materials and Devices. Reinhold Publishing Cor­poration, New York, 1960.

6. E. B. Penrod, Theoretical Analysis and Performance Characteristics of a Peltier Refrigerator. Bulletin L'Institut International Du Froid, XL, No. 2, 517-539, April 1960. See also Papers 3-23 and 3-24 that were read at the Tenth International Congress of Refrigeration at Copenhagen, August 20, 1959·

7. E. B. Penrod and Cho -Yen Ho, Theorie Mathematique D'Un Refrigerateur A Eflet Peltier Et D'Un Generateur Thermo-Electrique. Le Journal De Physique Et Le Radium Physique Appliquee Supplement Au No. 7, Tome 21, 97 A, Juillet 1960.

8. J. B. Whitlow, Jr., Thermoelectric Studies. University of Kentucky Master of Science Thesis in Mechanical Engineering (1959).

9. E. B. Penrod, A Theoretical Analysis of a Peltier Refrigerator. A.S.M.E. Paper No. 59-A-266, Annual Meeting of the American Society of Mechanical Engineers, Atlantic City, N. J., No­vember 29 - December 4, 1959.

ro. E. B. Penrod and Cho-Yen Ho, Mathematical Theory of a Peltier Refrigerator and a Thermo­electric Generator. University of Kentucky Engineering Experiment Station Bulletin, Vol. 15, No. 58, 1-35, December 1960.

608

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Thermoelectric Refrigeration - Possibilities and Problems

Refroidissement thermoelectrique - Possibilites et problemes

THORE M. ELFVING, M. S. President of Isofiex Corporation, Burlingame, California, U.S.A.

SOMMA/RE. Le rendement total du refroidissement thermoelectrique peut, dans certains cas, etre ameliore par des moyens dont on ne dispose pas pour les procedes classiques. Par exemp­le : application de deux temperatures. Les caracteristiques sont l'independance de la puissance, la sensibilite a LI t et la commodite du systeme a cascade. On etudie la conception et les condi­tions defonctionnement necessaires pour obtenir un rendement maximal.

L'ecart du fonctionnement reel avec !es valeurs theoriques est surtout du a l'etevation de Li t en raison des resistances thermiques de la chaine de transmission de chaleur. L'abaissement de la temperature entre les surfaces en liaison avec !'isolation electn·que des plaques de soudure dans les groupes autonomes peut etre fortement reduit par !'utilisation de plaques de couverture en ceramique. On etudie les problemes d' application et le coilt du maten'au semi-conducteur dans les groupes autonomes normalises a venir.

At present thermocouple materials with a combined Z-factor of 3 x 10-3 ° C-1 are commercially available. Manufacturers believe that the prevailing bismuth - telluride system can be improved. The goal is a Z-factor of at least 4, the ceiling being established at Z = 5 - 6 [1] .

In comparison with conventional processes, the widely accepted opinion is that with a Z-factor of 3, thermoelectric refrigeration in regard to efficiency is equivalent to ab­sorption units of the small hermetic type and that it would take a Z-factor of 10 to match smaller hermetically sealed compressor units and a Z-factor of not less than 20 to match larger compressor units with a capacity over 1000 watts [2] .

However, in many important cases it is possible to improve the overall coefficient of performance by means which are not available at conventional processes. A thermo­electric two-temperature application can with no extra cost be operated at the theoretical maximum efficiency by using separate thermocouple assemblies for each temperature. Conventional processes operating at two different temperatures would for maximum efficiency require a costly duplication of the equipment as, with a single unit, all the refrig­eration capacity would be delivered with an efficiency or coefficient of performance corresponding to the lowest temperature. Thermoelectric refrigeration is in this respect more flexible and better than comparisons considering only the maximal LI t would indicate

DESIGN AND OPERATION REQUIREMENT FOR MAXIMUM EFFICIENCY

Previoulsy published studies have shown that thermoelectric heat pumps should use a mode of operation leading to maximum coefficient of performance at maximum operat­ing conditions, which also means minimum temperature differences and reduced costs for power supplies and heat dissipating surfaces. It has been demonstrated that inter­mittent operation is advantageous and possible without extra losses. In order to utilize the expensive semiconducting material fully at the relatively low input current corres­ponding to maximum coefficient of operation the thermocouples should be built with short legs (large A/1 ratio) which leads to minimum cost for the semiconducting material in relation to capacity [3].

The heat transfer chain of thermoelectric heat pumps must be designed for minimum temperature drops. The specific problem of interface temperature drops in sealed thermocouple packages shall later be dealt with.

609

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CHARACTERISTIC FEATURES

1 . The independency of the efficiency on the capacity of a thermoelectric heat pump is perhaps the most striking virtue of thermoelectric refrigeration. Small ice-freezers and other cooling devices can be operated air cooled with a total DC-input effect of 15-25 watts range. What the future market for refrigeration appliances in this capacity range will be can not even be guessed.

2. Another characteristic quality of thermoelectric heat pumps is their sensitiveness for temperature differences or operating Lit. This is immediately evident from the formula for the maximum coefficient of performance :

ii 1 + z Tc Emax = ii 1 + z

It is also obvious that the strong influence of Li t upon the maximum coefficient of performance will gradually diminish with growing Z-factors. For the low Z-factors re­presented by presently available thermocouple materials, however, the influence of Li t i s so dominating that the results of practical applications are largely depending upon the success with which the Lit has been kept at a minimum.

3. The convenient way in which thermoelectric heat pumps can be arranged in stages or in so called cascades is another characteristic feature of thermoelectric refrigeration different from conventional processes. Cascades ordinarily only slightly improve the efficiency over a single stage, since the increase in coefficient of performance even for perfectly balanced cascades using the most favorable temperature split is seldom more than 10 %. However, cascades where the middle temperature can be utilized for useful refrigeration purposes while only a fraction of the refrigeration capacity is used at the lowest temperature level offer simple solutions to many design problems.

1 5

1 4

1 3

1 2

I. I LO 0.9

0.8 "

0.7 '" 306

0.5

0.4

0.3

0.2

0.1

( !hermoel�t"c heol pump )

�l= lcond.- tevop. €.it:is. or TH - TC

Tc= -IO°c = +263 1<

Cc= :� = Carnot eff.

ID 20

\ I \ \ \ \ \ \

30

\ \ \ \ \ \

\

\ \ ',�,

' "· ',:o<t-.

\ \ \ \ ',�/o '�''o':I ',�:,

'.f:',' ,, ,

40 ll. I 50 60

'', ,,' ,, ,

70 80 t

Fig. r . Theoretical efficiency versus LI t for thermoelectric heat pumps (Z = 3 x 10-a 0 c-1) in comparison with absorption and hermetic compressor units of smallest type.

Fig. 1 illustrates the efficiency at various LI t's of the smallest absorption (Babs) and vapor compression ( Bcomp) units in relation to thermoelectric heat pumps with a combined Z-factor of 3 x 10-3 ° C-1 (Bmax.).

The thermoelectric heat pump is more dependent on the LI t than the absorption unit. While Bmax· and Babs are about equal at Lit = 60°C, the thermoelectric heat

610

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pump is decidedly more efficient at lower Llt's. The small hermetic compressor is in a different class, except at rather low Ll t's, where the thermoelectric process is catching up,_. with emax,Mapproaching_.50% of the compressor efficiency.

A two-temperature refrigerator-freezer can serve as an example : Let us assume that the food compartment operates at Tc = + 5 ° C and the freezer compartment at Tc =

-15 ° C and that the relative refrigeration capacity of the two compartments is 2 : 1 . A thermoelectric refrigerator using a fractional cascade can for a total LI t = 60 operate in two istage(with (Lit) 1 = 40°C. (TH1 = +45°C, and Tc1 = + 5 ° C) and (Llt)2 =

20°C (Ta2 = + s, Tc2 = - 15). Using the theoretical formula for emax, and a com­bined z = 3 x 10-3 0c-1 we find for the first stage a theoretical efficiency emax 20 = 1 .22 and for the second stage (main stage) emax40 = 0.675. The combined emax, for the cascade is then emax 60 = 0.28 which represents the coefficient of performance of the freezer compartment refrigeration at Tc = -15°C. With 1/3 of the refrigeration capacity delivered with this coefficient of performance and 2/3 at emax, = 0.675 we find that the overall theoretical coefficient of performance for a two-temperature refrigerator of this type is not less than 0.54 or twice a� good as the absorption unit for the same total LI t = 60° C and more than half as good as the small hermetical compressor.

The same calculations using a Z-factor of 4 x 10-• ° C-1 leads to an overall coefficient of performance = 0. 79 or three times better than an absorption unit and in class with a small hermetic compressor.

DESIGN PROBLEMS OF MODULES AND PACKAGE UNITS

A minimum of both heat losses and temperature drops all the way from the core material in the thermocouples to the final heat dissipating surfaces is desirable.

1 . The electric contact resistances between the semiconducting material and the junc­tion plates are important, especially the resistance on the cold junction side as it not only increases the energy input but also lowers the cooling power by the same amount. One bad contact on the cold junction side can ruin a whole thermocouple assembly. Normally, the contact resistances are no problem except for short-legged thermocouples operated with a current close to maximum output. Here the current density can amount to 200-300 amps/cm2 and the contact resistance could lead to a deterioration in perform­ance. In commercial installations thermoelectric heat pumps will be operated with current for max. coefficient of performance rather than for maximum output. The current density will, therefore, in practice never reach more than half of the above values. According to manufacturers, the technique of soldering the junctions is now so well developed that the contact resistance under the above conditions would not have any appreciable effect even for thermocouples as short as 2 mm in length.

2. Electric insulation of junction plates. The thermocouple junction plates (copper straps) must be electrically insulated from the metallic heat absorbing or heat dissipating cover plates or radiator bodies necessary to transfer heat to or from the limited surfaces of the thermocouple assembly. Thin laqcuer films, serving as bonding agents, cause considerable thermal resistance and are vulnerable for mechanical pressure unless applied to extremely plane and smooth surfaces, which thermocouple assemblies seldom offer. Aging and moisture are other problems and a solution along these lines has not been found. Epoxies as bonding agents are mechanically strong and pressure safe but have relatively high thermal resistance even with filler materials for increased heat conductivity.

Other insulation methods include using thin films of mica, mylar, epoxy etc. usually applied together with greases or oils (silicones) in small amount for better heat contact. Insulation agents of this type require bolting between the hot and cold junction side causing added heat losses.

Anodizing or hand coating is a simple and relatively inexpensive process for providing an adequate electric insulation. The problem here is again the vulnerability for pressure unless applied between very carefully machined and smooth surfaces. Excessive or

611

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uneven clamping pressure can puncture the anodizing film and short circuit the thermo­couple assembly. Therefore, when used in thermoelectric packages the anodized cover plates on top of the junction plates on both sides are bolted together under careful control so that no shorts are caused during the assembly of the package. The bolts or screws serve to secure the cover plates in a fixed position while keeping the thermo­electric arrays under compression and the finished package can later be clamped or pressed against other bodies without risk for shorts. As indicated above the main ob­jection to this construction is the interface temperature drops between the anodized surfaces and the junction plates. (An average figure for the temperature drop in bolted packages is 5°C total interface drop when the heat load on the hot junction side amounts to 1 watt/cm2 junction area.) The price of such package units is also high in relation to the cost of the incorporated semiconducting material itself.

3. Ceramic package units. A radical solution of the short circuiting is obtained by enclosing the thermocouple array between ceramic materials instead of metal cover plates. Suitable materials with both electric insulating properties and relatively high heat conductivity are, for instance, Aluminum Oxide, Al203 and Berrylium Oxide BeO, which both can be manufactured in thin wafers of sufficiently large dimensions to cover a thermocouple assembly of usual size. Al203 of Alumina wafers have a dielectric strength of 230 volts per mil, are impervious for water vapor, vacuum tight and have a heat conductivity of 16 kcal/hr, m2, °C/m, a conductivity of the same magnitude as that of stainless steel. BeO has equally good dielectric properties and a heat conductivity even exceeding that of aluminium. Al203, being less costly, is sufficiently good for our purpose.

I (.

Fig. 2 . Expanded view of ceramic package units built from short-legged high capacity thermo­couples.

l l. Junction plates on hot junction side 12. Junction plates on cold junction side 13. Semiconducting core material 14. and l 5. Electrodes for DC supply 16. Ceramic Al203 wafer on hot side l 7. Ditto on cold side 18. Non-conductive frame member r9. Slot for electrode

612

22. Border line for outside metalized area of ceramic wafer on hot side

23. Ditto on cold side 27. Inside metalized areas matching I I 28. Inside metalized areas matching l 2

All parts except frame 18 soldered toge­ther.

III-22

A heat dissipation of 1 watt per cm2 of junction plate would cause a temperature gradient in an Alumina wafer with a thickness of 0.020" of approximately 0.3°C. This heat load represents approximately the maximum conditions on the hot junction side of modern thermocouple arrays operated at max. coefficient of performance. On the cold junction side the temperature drop would be considerably less. Interface tempera­ture drops because of thermal resistances between the ceramic wafers and the junction plates can be almost completely eliminated by first metalizing the ceramic wafers in a pattern corresponding to the junction plates and then soft soldering the partly metalized wafers to the junction plates, the solder eliminating variations in the surface of the assembly and establishing a metal connection with negligible temperature drops. The ceramic cover plates are metalized also on the exposed side and sealed around the edges to form a simple vapor tight package which can be soft soldered to heat sinks or other metallic heat transfer surfaces. (Fig. 2).

APPLICATION PROBLEMS

The technique of handling metalized ceramic package units differs somewhat from bolted packages in that they require mechanical support after being assembled between heat transferring bodies so that no stress is transferred to the semiconducting crystalline material between the junction plates. The heat pumping unit must also be applied to the practical refrigeration device in such a way that undue heat losses and temperature drops are avoided.

These problems are by no means easy to solve. Using the same specific heat load as in a previous example, 1 watt per cm2 junction plate area, we find that air cooling with a heat transmission coefficient of 10 kcal/m2, hr, ° C (natural convection) would require almost a hundred times larger surface to keep the temperature difference down to l0° C. Enormous fins and forced air circulation is one solution to this problem but a more elegant solution is, no doubt, the use of boiling and condensing media. Secondary hermetic heat transfer systems using a medium like Freon do not always provide the expected high heat transfer rates. Inside temperature drops of several degrees centigrade are often encountered in the boiler and condenser portion of the system due to "vapor binding", condensate blockage and other surface phenomena. A series of experiments have shown that double-walled panels of the "rollbond" type provide ideal hermetic heat transfer systems. Special rollbond patterns provide a multitude of passages for gas and liquid with a minimum of pressure drop while boiling at a high heat transfer rate can take place without vapor binding. The total temperature drop in a secondary heat transfer system of this type can in practical applications be limited to 1-2° C. Only hermetic heat transferring panels of this type can provide the effective and large surfaces necessary for air cooling of thermoelectric heat pumps without the use of forced air circulation [4] .

COST OF SEMICONDUCTING MATERIAL

It is evident that short-legged high capacity thermocouple arrays can be standardized in package units of say 8-10 couples to suit every heat pumping need. Capacity require­ments in various devices can be met by choosing the number of units in one or more stages and additional flexibility is obtained by choosing a suitable fl-value for the opera­ting current. For thermoelectric heat pumps suitable for household refrigerators and similar devices the rod diameter of the semiconducting material is likely to be standard­ized around a magnitude of 7 mm and it can also be predicted that the corresponding leg length at present will be kept around 1 /8". With these dimensions and with a price for the alloy rods of $ 40.00/lb. the net worth of semiconducting material per 100 watts of refrigeration in air cooled ( Ll t = 40°C) devices would amount to approx. 25 U.S. dollars. With still shorter leg length this figure could be considerably reduced. With 5 mm rod diameter and 1/ 16" leg length the price would theoretically be only 6-7 dollars per 100 watts. Many problems must be solved, but there should be no unsur­mountable difficulties for standardizing on such thermoelectric assemblies which may become the real low cost heat pumps of the future.

613

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REFERENCES

1. Dr. G. Haacke, Improvement of the effectivity of material for Peltier elements (in German) Kaltetechnik, No. 3, March 1960.

2. M. Backstrom, The Peltier effect contra the conventional refrigerating processes. Proceedings of the Xth Int. Congress of Refrigeration, Vol. 2, page 96 (1960).

3. Thore M. Elfving, Desing and operation of an air cooled thermoelectric refrigerator (in German) Kaltetechnik, No. 3, March 1962.

4. Thore M. Elfving, A study of design problems and mode of operation for thermoelectric refrig· erators, ASHRAE Journal, Vol. 5, No. 2, Feb. 1963.

614

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Thermoelectric Refrigeration and Prospects for its Wide Scale Tech­nical Application

Le refroidissement thermo-electrique et les perspectives de son application sur une grande echelle

V. S. MARTINOVSKY, V. A. NAER and S. A. ROZHENTSEVA Odessa Technological Institute of the Food and Refrigerating Industry, Odessa, U. S. S. R.

SOMMA/RE. Les progres realises dans le domaine de la preparation des composes semi­conducteurs pour les dispositifs thermo-electriques permettent actuellement la fabrication des modeles industriels.

Le rapport presente les domaines les plus rationels d' application du refroidissement et du chauffage thermo-electriques: le refroidissement des noeuds isoles de l'appareillage et des instruments electroniques, la production du froid aux temperatures variables, les installations de pompe a chaleur a semi-conducteurs, le conditionnement d'air dans les cas particuliers, la production de la glace dans de petites quantites. De plus, /es dispositifs semi-conducteurs pour /es variateurs de flux de chaleur sont etudies.

On decrit aussi des constructions de dispositifs thermo-electriques experimentaux et les donnees d' essai.

De plus, on presente les resultats de l'effet de pulsations du courant electrique sur Jes indices de temperature et d' energie des dispositifs thermo-electriques. Les calculs et /es essais effectues ont montre que /'influence des pulsations du courant dans certains cas etait assez grande.

La mise en application du refroidissement et du chauff age thermo-electriques depend dans une grande mesure de la solution trouvee aux principales questions suivantes: !'extraction de nouveaux composes semi-conducteurs eff ectifs; comment /es produire en serie a prix re­duit; l'automatisation de la production des thermobatteries, et le developpement des construc­tions des installations avec consommation minimale de composes semi-conducteurs. Enfin la solution du probleme d' alimentation des installations en courant continu de basse tension et la mise au point du reg/age automatique de la temperature et du rendement.

On peut penser que si taus ces problemes etaient resolus prochainement, /es systemes thermo­electriques seraient largement employes dans l'industrie.

The development of semi-conductor materials with high thermoelectric properties and advance in the designing of thermoelectric coolers make it possible to start industrial production of semi-conductor cooling devices.

Semi-conductor microrefrigerators are widely used in the USSR in the fields of radio electronics, vacuum engineering, biology, medicine, measuring engineering, etc. The cooling of radio tubes, transistors, photoconductors, vacuum traps, tables for freezing biological tissues, ultrathermostats, null devices, hygrometers are by far not a full list of microrefrigerators manufactured at present. The hourly consumption of cold in all the enumerated devices are measured by units of kilocalories per hour and semi-con­ductor refrigeration systems are the most appropriate. Their power characteristics are of no great importance.

Larger thermoelectric refrigeration systems are developed simultaneously with micro­refrigerators, in particular, coolers for liquids, gases and solid bodies, for which a source of cold is required with an alternating temperature.

Thermoelectric refrigeration systems make it possible in the latter case to develop simple by design sources of cold with an alternating temperature, operating according to a cycle which is close to the theoretical one.

615

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Multi-stage compression, which is not always possible, is required for such a cycle in compression plants with pure refrigerants.

A semi-conductor thermal battery, which is assembled of dozens and hundreds of thermocouples, imitates a compressor with a very large number of compression stages. As a result, the power efficiency of semi-conductor coolers, manufactured of modern efficient semi-conductor material and operating with alternating temperatures, ap­proaches the efficiency of a compression plant.

Semi-conductor heat flow variators represent a separate group of devices, which can be used for intensifying the heat transfer process at small temperarure differences be­tween the mediums participating in the heat exchange process; also in thermostats, in which the temperarure ought to change according to a given programme; in some setups of air conditioning, etc. Such devices include also semiconductor vaporizers or distillers, in which the temperature of the condensing vapor does not practically differ from the temperarure of the boiling liquid; thermostats for radio electronic appararus ; air coolers, in which the temperature of the water, transferring the heat from the hot joints of the thermal elements, is lower than the temperature of the air entering the cooler; liquid and air coolers with an evaporative system of heat transfer from the hot joints, etc.

Semi-conductor microrefrigerators are described in literature in detail, therefore, only experimental models of liquid coolers, which have been investigated in the semi­conductor laboratory of the Odessa Technological Institute of the Food and Refrigerat­ing Industry will be discussed herein. Some of these coolers can be used also for other purposes, for instance, for the production of ice. Some of these models operate partially as heat flow variators.

The design peculiarities of semi-conductor liquid coolers are determined by their purpose (cooling of liquid in a limited volume or in a flow) and the method of heat transfer from the hot joints. The transfer of this heat can be effected by cooling with running water, or with an evaporative system, or with air, etc. Running water and an evaporative system were used in our experimental models.

The refrigerating capacity of the tested apparatus ranged from 30 to 500 kcal/hr. Fig. 1 presents a general view of a cooler for liquid in a limited volume. The semi­

conductor thermal battery, consisting of 16 thermocouples, is built-in the cover of a 2 litre vacuum flask. The thermal elements are 2 mm high with a cross section area of 1 cm2• The weight of the cooler without the rectifying device is 350 g. The weight

Fig. r . General view of liquid coolers in a limited volume.

of the semi-conductor materials is 45 g. The hot joints are cooled with running water The direction of the water flow coincides with the direction of the electric current The thermal battery operated with pulsating current from a single-phase full-wave rectifier.

616

�

c J

0

� 20 -lo:: � 15 10 " 0

ta.oc 15 10 5

' -

� �

0

r--.. ,

I-.-. ..,._, � ;z: �

� /// � ;,....-,._ "S�

N

30 Fig. 2. Results of testing liquid cooler.

, '4t =5IOC -..... .........

r,;-... -::.......---�

i--...: -

'J,5 �"'" ,,..-

(/ /{' /' _..,., / -.--

..__ � t'--� .........

�

/

7 -...... � � -- -

-

<bl/ i.....-.,,....,.

..,..... .--

-

10 13 15

45 �

•u�

1=20 r--r-� 30 ..._ I"-,__ �-!

r-- � 60 90 120 T min

III-28

10 - cooled water temperature, Lit- alteration of cooled water temperature, Q0 - refriger­ating capacity, s - coefficient of performance, T - time, J - current strength.

The results of testing an apparatus when cooling water are given in Fig. 2. The apparatus can cool 5 kg of water by 10 to 15°C in 25 to 45 min. The current, passing through the thermal elements, is at 75 amp, while the consumed power is 50 W. With a reduction of the current strength, the apparatus coefficient of performance increases with a simultaneous extension in the cooling time.

If direct current is passed through the thermal battery instead of pulsating current, the coefficient of performance increases by 20 to 25 %.

The photograph of another cooler is given in Fig. 3. The thermal battery is of the same design as in the previous device but consists of 32 thermocouples. The height of the thermal elements is 3 mm with a cross section area of 1 cm2• The thermal battery is built-in the bottom of a flask with a capacity of 0.5 !.

Fig. 3. General view of ice makers.

617

III-28

The device is designed for cooling and freezing wine for laboratory investigations and for cooling water. It can cool 0.5 1 of water by 15 to 20° C in 5 to 7 min. This type of cooler can be used to success in the capacity of an ice maker.

The results of testing the apparatus when freezing water are stated in Fig. 4. The dotted lines refer to the tests with pulsating current when the thermal battery has been supplied through a single-stage full-wave rectifier without an electric filter. The solid lines characterize the operation of the ice maker supplied with direct current. The capacity of the apparatus reduced by 25 to 30 % when shifting from direct current to pulsating.

e

1,5 I I

�25 I I \

1,0 I \ .\ c \

0,75 '( \ 'O \ 0 0 0,50 � 0 I '0- 025 / , 'O I ,./

0 h / 0 /' �./'

0 41

� IL:'/ -

/er I v l,Af' -, I/ .... I x" /;v '·

;: " ·, k'" 17 .,,, -� k -�1 ... / ,._.

_,,, '.lr-:' '

80 t<O fW

O.u

G

Ja

�. � � "!:> 800 600 400 200

Fig. 4. Results of testing ice maker. Q0 - refrigerating capacity, e - coefficient of performance, G - amount of ice, J - current strength.

It is clear from Fig. 4 that when operating with non-pulsating direct current (60-100 amp), the apparatus produces from 400 to 700 g of ice per hour and its coefficient of performance is 0.62 to 0.37. When operating with maximum capacity, the ice maker produces 500 g of ice in 40 min., the coefficient of performance being 0.22.

An evaporative system of heat transfer from the hot joints is applied in a liquid flow cooler. The results of testing such a cooler are given in Fig. 5. The cooler consists of two thermal batteries connected in series with 120 thermocouples in each. The sizes of the thermal elements are the same as in the previous case. The thermal batteries are turned to each other with the hot joints to which the fins are soldered. The fins are sprayed from the top with water and form vertical ducts. The fan feeds the air against the water flow. The cooled liquid flows along the canals subsequently over all the cold joints.

The apparatus was tested in the capacity of a water cooler. The tests were carried out for cooling 10 to 12; 21.5 and 30 kg of water per hour. The initial temperature of the water was to = 25°C. Pulsating current was supplied through the thermal battery. The power, consumed by the fan drive (10 W), was not taken into consideration when calculating the coefficient of performance e.

An interesting peculiarity of this type of coolers is the possibility of cooling water, running along the hot joints, utilizing the evaporation of water at the hot joints without consuming electric power.

618

<..

e s 8 7 6 5 4 3 2 D

JOO 200

� 100 <:I " "<

r;;:§' 4 8 12 16

.a t °C

III-28

\\ \\ 3

\ ·"'""' / 2 I' ,...._ ,.:::::: :,.::'._ 1 -- -r-. -::::: --,.-3

L--' ..--i.,.-o- " I ' ./'. _v-- +, IX / I

� -- I

�---W-f-----1-._,___ -�al 0 I 170 :JO i 140 !Ja I I

�� 1--.,_ I>--.. �;-.._ 3 --r- ,_ - 2

lttf 2.5 � 1

Fig. 5. Results of testing liquid cooler with output ro to rz kg/hr, 2r .5 kg/hr, 30 kg/hr (desig­nations see in Fig. 2) .

When cooling the water by 10 to 12° C, a considerable part of the thermal battery operates as an intensifier of the heat transfer process for the temperature of the water at the hot joints is lower than at the cold ones. The consumption of electric power in this case is quite negligible. For instance, the power consumed by the thermal battery is only 20 W and the refrigerating capacity equals 100 to 120 kcal/hr when cooling 10 to 12 kg of water per hour by 10°C. Changing the current, supplied through the thermal elements, it is possible to obtain the required temperature of the liquid being cooled within a wide range of heat loads.

An increase of the temperature of the water, fed to the cold and hot joints of the thermal battery, improves the power efficiency of the apparatus mainly owing to the evaporative system.

Fig. 6a. General view of liquid cooler thermal batteries.

619

IIl-28

Fig. 6a presents a general view of a liquid flow cooler, in which the heat is transferred from the hot joints by running water. The cooler consists of two thermal batteries with a refrigerating capacity up to 500 kcal/hr. Each battery consists of 48 thermocouples. A reduction in the number of thermocouples is possible owing to a reduction of the relative height of the thermal elements. The sizes of the thermal elements are as follows : 20 x 20 x 3 mm.

Fins are welded to the connecting copper plates on the hot and culd sides of the ther­mal battery.

The cooled liquid and the cooling water run in special canals respectively along all the cold and hot joints.

The weight of one thermal battery is approximately 5 kg. The results of testing one thermal battery of the cooler are given in Fig. 6 b. The

temperature of the water at the thermal battery inlet was 23.5° C, and the amount of cooling water was 90 kg/hr. Electric current was supplied to the thermal battery from a three-phase fullwave rectifier.

t 3

2

1

1!lo 25 5ii 15 10

I/.

\ � L_.I ,__

- ' T 1""11: �

\ ,\ \' I � \

� Q \ \ ' I � J � 'ef

I/ � 'J /� v

� _./ �

� � I

I J To

-' I ---

w J Qo \ r I'

I � ".'.: µ.. � � � t

't:::: ..... �

O $0 100 1�0 :Ja •-G a 18 kg/hr; o-(; = 1.5 kg/hr

Q,flo w kcaejhr 500

400

300

200

100

Fig. 6 b. Results of testing liquid cooler. T0 and T - temperatures of cooled and cooling water, Q - heat, liberated at hot 1omts, Q0 - refrigerating capacity of thermal battery, W - consumed power, e - coefficient of performance, J - current, supplied through thermal elements, G - amount of cooled water.

Investigations have been completed in the laboratory of experimental models of air conditioners with a refrigerating capacity of up to 1,000 kcal/hr. The hot joints of the air conditioner batteries were cooled with running water. When cooling air by 8 to 10°C the coefficient of performance of the apparatus was e = 2.

All the semi-conductor coolers described herein can be applied at present in the in­dustry. However, the development of mass production of these coolers requires a further improvement of semi-conductor materials, automatization of the production of thermal batteries, elaboration of special supply systems with low pulsation, as well as automatic devices for temperature and capacity control. Some of these problems are solved with success. For instance, a production line is developed for the automatic manufacture of semi-conductor thermal batteries for liquid coolers. It may be expected that all the other problems will be solved in the nearest future and thermoelectric devices will be widely used in the industry.

620

III-28

SUMMARY OF THE DISCUSSION (Papers 111-17, 111-22, 111-28)

H. Guennoc, France: I should like to ask Professor Martinovsky which order of magni­tude of the variation of thermoconductance makes the device work in this particular way.

A. Gelbtuch, U. K. : I should like to ask Professor Martinovsky if the heat transfer would be improved by inserting closed circuit thermoelements into a copper wall across which there is a thermo gradient.

V. S. Martinovsky, USSR : In reply to these two questions I should like to give an example showing the order of improvement of heat transfer to a wall of constant thickness caused by replacing copper with semiconductors as follows :

1. Q = to 120 kcal per hour for copper 2. Q = to 1 10 kcal per hour for semiconductors electric circuit open 3. Q = to kcal per hour semiconductors electric circuit closed 4. Q = to 500 kcal per hour semiconductors electric circuit closed and energized.

J. Kowalczewski, Australia : I refer to Professor Martinovsky's reply to the previous question giving an example showing the order of improvement of heat transfer through a wall of constant thickness caused by replacing copper with semiconductors as wall materials. In cases two to four above, the elements were provided with additional heat exchanger area, i. e., fins as compared with case one. I wonder if the increase of heat transfer rate in this example, when copper was replaced by semiconductor materials, was caused only by improved thermoconduction or by the increased heat transfer area, i. e., between case one and case three.

V. S. Martinovsky, USSR: This, of course, was so.

J. F. Downie-Smith, U. S. A . : I am very glad to hear the results of tests performed in Russia regarding the most efficient length of thermoelectric materials in practical devices. The lengths Professor Martinovsky quoted are comparable with the results obtained by my own company and are considerably smaller than those listed as minimum in the presentation by Professor E. B. Penrod. Clearly, there have been different geometrics or properties involved in the theoretical discussion and the practical devices.

I can also support Professor Martinovsky's recommendation of the application of thermoelectric panels to increase the heat transfer from surfaces for special cases, where first cost was not the prime factor. This has been applied successfully in the United States in a limited way.

A. Gelbtuch, U. K. : I should like to make the point to Mr. T. M. Elfving, that the limit imposed on branch height is the ability to remove heat from reduced surface areas, i. e., with increased heat fluxes.

A. B. Newton, U. S. A. : Mr. Elfving has mentioned some interesting ways of meeting the problems we face in making good use of thermoelectric equipment. However, based on well over 2 % years of using aluminum oxide - both with and without metallized images of the thermoelectric junction plates - on hundreds of subassemblies I believe we must consider other factors under the heading "Application Factors" on page 5 of the paper. 1 . For valtoge greater than about 10 volts edge sealing of the Al203 wafers is very im­

portant. Otherwise leakage paths develop over a period of time, and one cannot take full advantage of the good electrical insulating properties of the material.

2. Linear expansion of the Al203 relative to the module and cooled or sink surfaces are a problem. The Al203 chips will fracture if soldered to the adjacent surfaces, or for that matter to only one adjacent surface, if more than about 2 %" in any linear dimen­sion. This shows after as few as 500 to 1000 cycles of temperature change then some 75° F.

3. Solder bonds must cover all the area of the junction of electrical bus to metallized surface. Otherwise poor heat transfer may result and hot side junctions will become overheated.

621

lll-28

In practice the process is difficult to control since low temperature soldering is usually employed, and good noncorrosive fluxes like rosin to not activate at low tem­peratures. We found it necessary to develop low temperature fluxless solder techniques in order to achieve useful life.

4. If pretinned, the metal surfaces must be cleaned and held in inert or reducing atmo­sphere during fluxless assembly.

5. The problems are amplified, overall, by increased heat flux as more heat is transferred by each couple, and for compact designs one would like to keep the bus bars as small as possible. We have used couples of less than 0.125" length (3 mm) for over two years and feel that all assembly and reliability problems are controlled down to lengths of 0.070".

Below 0.070" the losses between hot and cold bus bars are usually out of pro­portion to gains, even at low L1 t or in a vacuum. We do control these losses quite well in the range of couple lengths of 0.070" to 0.120", and for all types of cooling, -liquid to liquid, air to liquid, air to air, etc.

622

III-10

Diagrams of Dimensionless Equations Determining Two Basic Work­ing Regimes of Peltier Heat Pump

Diagrammes des equations sans dimension pour determiner les regimes de fonctionnement des pompes a chaleur a effet Peltier

I Ing. KAREL SYROVY Research Institute of Refrigeration and Food Engineering, Prague, Czechoslovakia

SOMMA/RE. On essaie de simplifier le probleme de la proposition et du calcul des regi­mes de fonctionnement optimal et maximal d'un element frigorifique thermoelectrique Peltier. Le caractere sans dimension des equations exprimees graphiquement permet un calcul facile par rapport aux parametres physiques arbitraires du materiau semi-conducteur. Les equations sans dimensions acceptees resolvant les valeurs necessaires, et dont la validite a ete confirmee par derivation retrograde, sont presentees sous forme de diagrammes individuels. Ceux-ci etaient ca/cutes numeriquement pour une grande serie de temperatures Th et Tc aux soudures et des coefficients d' efficacite Zer du materiau.

Pour la lecture pratique d'une proposition concrete de refroidisseur Peltier, tous les gra­phiques partiels sont concentres en un seul diagramme combine. Ce diagramme permet de lire f acilement les conditions physiques choisies de l' element Peltier pour des valeurs Zer ( coeffi­cient d'efficacite) et Th (temperature de la soudure chaude) et Tc (temperature de la soudure Jroide), exprimees par une valeur sans dimension Z = i Zer (Th + Tc) et le rapport de tem-

peratures �:. Les conditions physiques comprennent la puissance frigorifique optimale de

!'element Qem ax, l'apport d'energie optimal de /'element Pem a x, le courant optimal I em ax, la puissance frigorifique maximale de l' element Qm a x et l' apport d' energie maximal de l' ele­ment Pm ax.

Les diagrammes obtenus ( ou le diagramme combine) f acilitent nettement le calcul en vue de la proposition du materiel frigorifique thermoelectrique et donnent un aperfu visuel des conditions physiques variables de l' element en f onction de la variation des parametres. Les erreurs de lecture dans un diagramme de 60 x 50 cm sont inferieures a 1%.

1 . INTRODUCTION

The present calculation technique in frequently repeated projects of thermoelectric devices is comparatively laborious and does not give rapid information on concrete changes of states under simultaneous changes of physical parameters of semi-conductor materials. The problem has been solved by the use of dimensionless equations (1) valid for Peltier heat pump system, that were numerically computed for a wide range of tem­peratures Th (hot junction temperature) and Tc (cold junction temperature) and for the "effective figure of merit" of material Z e r. The calculations have been concentrated into individual diagrams and/or into one conjugate diagram and solve two basic states of Peltier cooler - the optimum and the maximum.

Thus, it is possible to read from one conjugate diagram the optimum refrigerating capacity Q e max, optimum current Ie ma.x, optimum input of element Pe max, maximum refrigerating capacity Q max and maximum input of element Pmax, all expressed in the dimensionless form.

These physical states have been plotted in dependence on the dimensionlessly expressed "figure of merit" Z = 0.25 Zer (Th + Tc), the parameter being the ratio of junction

Th temperatures

Tc. Reading error in the diagram in the order of approx. 50 x 60 cm is less

than l %.

623

IIl-10

2. ELEMENTARY DEFINITIONS

The specific thermal conductivity of the whole Peltier element is given by the ex-pression

u = u p + UN (W. grad-1• cm-1)

Specific resistance of element (excluding contact resistance) is

e = e P + (J N (Q · cm)

Thermoelectric power of both semi-conductor parts is defined

rx = rx r + O: N (V · grad-1)

Let us designate the resistance of the contact "cold junction-semi-conductor leg" by the symbol Re (Q).

The relation of the leg length to its section will be

l (! = -- (cm-1) s

Let us also define without derivation the so-called "effective figure of merit" of the element material, which includes the effect of two contact resistance "cold junction-semi­conductor legs".

z Z e r = - - ------ (grad-1) 2 Rc 1 + --

(! (! The following calculation concerns Peltier element consisting of two semi-conductor

legs of equal length l and equal section s. This means that in practical selection of material approximately equal specific conductivities of both legs must be chosen, in order to ob­tain uniform current distribution in both parts.

3. CALCULATION OF PARAMETERS OF COOLING ELEMENT

The basic equation for refrigeration capacity of an element according to Ioffe's theory is

1 u Q = rxTcl - 2 ( g (! + 2Rc)l' - (! (T11 - Tc) (1)

and element input

P = rx (T11 - Tc)I + (ef! + 2Rc)I' (2) On the grounds of the same theory the equation for the optimum current of the element

(. h" h . Q ) . m w 1c rat10 -p = max 1s

rx(T11 - Tc) le max = (M-1) (efJ + 2Rc)

and for maximum current of element (in which Q = max)

a Tc Imax = efJ + 2 Rc By substituting (3) into (1) and by introducing dimensionless expressions

624

- 1 Z = 4 Ze1 (T11 + Tc) - fJ Q = �------ . Q

u (T11 - Tc)

(3)

(4)

(5)

(6)

IIl-10

we shall obtain (after adjustment) the dimensionless expression for the optimum refrig­eration capacity of the Peltier element

where

Q s max =

-- [ 1 ( Th)] 4 Z M - 2 1 + -r; - l (M - l)" ( 1 + �:) (7)

1 t M = [l + � Zer (Th + Tc)] 2 (1 + 2 Z) (8)

- Th By numerical evaluation of equation (7) for the wide range of values Z and Tc

, graph in

Fig. 1 was obtained, from which it is possible to read Q s max for given values Z and Th Tc .

Fig. r. Curves of dimensionless values of optimum refrigerating capacity Qs max in dependence on dimensionless value 7,

The actual optimum refrigeration capacity for the selected element dimensions results from equation (6)

u (Th - Tc) ···· Qs max = · {J • Qs max (6a)

By a similar adjustment it is possible to obtain the expression for maximum refrigera­tion capacity

Tc• Q max 2 Z T a T 2 - 1 (9) h - c which is graphically expressed in Fig. 2, also for valuesZ and�:. The effective maximum

refrigerating capacity is again given by equation (6).

The equation for dimensionless optimum current of the element is

Th - Tc lc max = 2 Th + Tc • (M + 1) (10)

625

III-10

Fig. 2 . Curves of dimensionless values of maximum refrigerating capacity Q in dependence on dimensionless value Z

where rx{J

le max = - • le max u

and is expressed by graph in Fig. 3, again for given values Z and ratio�:.

( ll)

Fig. 3. Curves of dimensionless values of optimum current I emax in dependence on dimensionless value Z

By eliminating current I from equations (1) and (2) and by adjustment we shall obtain a more complex expression for refrigerating capacity of the element in dependence on the element input, from which it is possible to calculate the input of element for the calculated refrigeration capacity Q of element.

626

1 -Q + 1 = (i + !:_

• Th + Tc) 2 _ _I._ � _ l

z z Th- Tc 2 z (12)

111-10

This dependence is plotted in Fig. 4, for temperature ratios �: = 1 .01 to 1 . 13 from

left to right and for temperature ratios �: = 1.09 to 2.00 at another scale from right to

left.

Fig. 4. Curves of values Q _: 1 (refrigerating capacity) in dependence on values � (input of Pel-

tier element) z z

The effective input of the element is again given by the equation

x (Th - Tc) p = fl

· P (13)

Fig. 5· Conjugate diagram of values Qemax, Qmax, lemax, Pemax and Pmax of Peltier refriger­ating element in dependence on value Z. The parameter is the ratio of junction temperatures Th Tc

.

627

III-10

For practical projects and calculations of the existing thermoelectric devices the ranges of the operating values have been limited and diagrams 1 to 4 have been concentrated into one main diagram in Fig. 5. The calculation of the element by means of this diagram, as compared with the conventional calculation system, is considerably simplified and offers a visual survey of the changes of the values sought with the changing parameters given.

4. CONCLUSION

Graphical expression of dimensionless equations solving a Peltier refrigerating ele­ment permits a rapid calculation of thermoelectric devices generally with physical para­meters of any description. It does away with a laborious numerical calculation for the changing temperature conditions, or for various material constants of semi-conductors.

LITERATURE

r. ]. R. Andersen, Thermoelectric Air Conditioner for Submarines, RCA Review, XXII, June 1961, No 2.

APPENDIX

A calculation example according to conjugate diagram in Fig. 5

For the sake of an example let us solve a thermo-electric refrigerating element with the following material parameters available :

a = 3.7 · 10-4 V · grad-1 e

= 2 · 10-a [} • cm

u = 2.6 · 10-• W • grad·1 • cm·1

Section s P = SN = 0.25 cm•, selected length l = 0.5 cm wherefrom

f3 = 2 s

Assuming that

2Rc ef3 = o.o5

then Zet = 2.5 · 10-a grad-1

According to preliminary calorific calculation the required temperatures on the junc­tions are selected at

Th = 307. 1 ° K ; Tc = 274.4° K Let us work out

Th Th + Tc = 581 .5°K and = 1 . 12

Tc as well as the term

u .d T

f3 2.6 • 10-• . 32. 7

2 = 0.425

According to equation (5) the dimensionless number of merit will be

� 1 Z = 4 Zer (Th + Tc) = 0.25 • 2.5 • 10-3 • 581.5 = 0.364

For the parameter Th

= 1 . 12 on the individual curves all the necessary dimensionless Tc

values are then read on the ordinate Z = 0.364 of diagram in Fig. 5.

628

Ill-10

From these dimensionless values will then be obtained the actual working regime of the selected Peltier element by using simple equations (6, 1 1, 13).

We shall read :

The optimum refrigerating capacity Q e max (the full curve running from left to right for Th r--; = u2)

u LI T -Q e max = 0.768 wherefrom Q e max =

-{3- • Qe max = 0.326 W (6a)

The maximum refrigerating capacity Q max (the dash line running from left to right for Th T; = 1 . 12)

u LI T Q max = 1 .86 wherefrom Q max = f3 • Q max = 0.79 W (6)

Q e max + 1 . Q + 1 . For ---=-- = 4.855 will be read on curve -- __ -_-- (the full curve with parameter

z z

!_T__11 = 1 . 12 running from right to left) the ordinate � = 3 and therefrom Pe max = c z

1 .091.

The optimum input of the element will then be

u LI T -Pe max =

-{3- Pe max = 0.464 W (13)

Q max -1- 1 p For ----=--- = 7.855 we shall read on the same curve the ordinate -=- = 17.5 and

z z

therefrom Pmax = 6.37.

The maximum input of the element will then be

u LI T -Pmax =

--{3- • Pmax = 2.71 W

Coefficients of performance will be

Q e max 0.326 e max =

Pe max =

0.464 = 0.704

P max 0.79 B Q max =

Q max=

2_71 = 0.292

Finally, let us read for ordinate Z = 0.364 the optimum feed current I e max (the dash-

d d . h Th .

an - ot curve wit parameter Tc = 1 . 12 runmng from left to right)

- u I e max = 0.262 wherefrom I e max = ct.{3 • I e max = 9.2 A (11)

629

IIl-10

The table below gives a comparison of results of the classical calculation (Ioffe) and of those read and calculated from diagram in Fig. 5.

Classical Calculation from calculation conjugate diagram Error in %

in Fig. 5

Qe max (W) 0.329 0.326 -0.91 Pe max (W) 0.463 0.464 +0.218 Bmax 0.709 0.704 -0.706 fe max (A) 9. 174 9.2 +0.284 Q max (W) 0.8 0.79 -1.25 Pmax (W) 2.753 2.71 -1.56 8 Q max 0.291 0.292 + 0.378

The error in % was calculated by taking values obtained by classical calculation as 100%.

Dimensionless values were read from the conjugate diagram sized 50 x 60 cm.

630

III-1

Transient Temperatures in a Thermoelectric Refrigerator Following a Step Change in Current

Temperatures passageres dans un refrigerateur thermoelectrique suivant un changement progressif du courant

W. F. STOECKER, Associate Professor of Mechanical Engineering, University of Illinois, Urbana, U.S.A.

J. B. CHADDOCK, Professor of Mechanical Engineering and Associate Director of the Ray W. Herrick Laboratories, Purdue University, Lafayette, Indiana, U.S.A.

SOMMA IRE. On met au point une expression analytique pour prevoir la repartition de temperatures passageres dans !es elements d'un refrigerateur thermoelectrique dans lequel se produit un changement progressif du courant. Le modele mathematique represente un couple : ;, , rmoelectrique dans lequel les temperatures des soudures chaudes etaient maintenues con­� tantes, mais aux soudures froides la transmission de chaleur par convection et la puissance thermique dominient. Les mesures experimentales d'un couple thermoelectrique, qui utilisaient un radiateur electrique pour simuler la transmission de chaleur par convection et la puissance thermique, concordaient generalement bien avec les resultats ca/cutes. L'ecart existant etait attribue aux caracteristiques thermiques de la soudure chaude et de !'assemblage du radiateur a sm•,1ir 1ue la temperature n'etait pas uniforme a tout instant dans !'assemblage.

introduction. Interest in the transient behavior of a thermoelectric refrigerator stems primarily from the need of predicting the start-up operation of a refrigerator and of analyzing its operation under changes in load. A contribution to the solution of the ge­neral problems of unsteady-state operation of a thermoelectric refrigerator is the solu­tion of the temperature distributions within the thermoelectric elements following a step change in current.

This paper presents an analytical solution for the transient temperatures in an idealiz­ed thermoelectric refrigerator following step changes in current. The test results obtained from an experimental model which approximates the conditions of the mathe­matical model are also reported and compared with the analytical results.

PART I. ANALYTICAL SOLUTION

Mathematical Model. As the first step in establishing the controlling differential equation and boundary conditions, a number of restrictions were imposed upon the thermoelectric refrigerator, including : (1) the flow of heat and electric current in the thermoelectric elements were one-di­

mensional in the x-direction (2) there is no heat transferred from the exposed surfaces of the thermoelectric elements (3) the thermoelectric elements had equal lengths, cross-sectional areas, thermal con­

ductivities, electrical resistivities and densities (4) there is no electric resistance at the junctions (5) the thermal conductivities, electrical resistivities and Seebeck coefficients are

independent of temperature (6) all points throughout the cold-junction assembly are at a uniform temperature

which, however, changes with time (7) heat conduction from the hot to the cold junction occurs only through the thermo­

electric elements

631

111-1

(8) the hot-junction temperature Th and the temperature of the fluid being cooled Tr are constant with time

(9) the electric current remains constant after its jump-change at zero time.

Differential Equation and Boundary Conditions. The controlling differential equation is the familiar one for the unsteady, one-dimensional conduction of heat in a solid which experiences Joulean heat generation,

where T x

a

t j p k

absolute temperature, ° K

1 o T j2p

a ai - -k- (1)

distance along thermoelectric element measured from the hot junction, cm thermal diffusivity = thermal conductivity divided by the product of density and specific heat, cm2/sec time, sec current density, amps/cm2 electrical resistivity, ohm-cm thermal conductivity, watts/(cm) (°K)

The boundary conditions to be imposed upon Eq. 1 are the conditions at the hot junction, the cold junction, and the temperature distribution within the elements at the instant of the step change in current. The first boundary condition is that the tem­perature of the hot junction is constant at a value Th, thus,

T (O,t) = Th (2)

The next boundary condition specifies a conservation of energy at the cold junction among the following quantities : conduction from the thermoelectric elements, con­vection from a constant-temperature fluid being cooled by the refrigerator, the rate of change of energy stored in the cold-junction materials, and the Peltier cooling. Com­bining these terms into an equation representing the quantities for the couple consist­ing of two thermoelectric elements,

where A h A c Tr !Xp,!Xn I M

-2 kA [ o T (L, t)/ o x] + 2 hA c [Tr - T (L, t)] = (rx p - rx n) I T (L, t) + 2M [ o T (L, t)/ o t]

cross-sectional area of element, cm2 coefficient of heat transfer by convection, watts / (cm2) (°K) area for convection heat transfer per element, cm2 temperature of fluid being cooled, ° K Seebeck coefficients of the p- and n-type materials, V/° K current, amps

(3)

thermal capacity of cold junction = summation of products of mass and specific heat, joule/° K

The remaining boundary condition expresses the temperature distribution within the elements at the instant of the change in current when t = 0

T (x, 0) = Th + Y (x/L)2 + Z (x/L) (4)

where L = length of the elements, cm Y and z = constants determined by solving Eq. 1 with the time-dependent

term, oT/ ot, eliminated.

When applying Eq. 1 to the two thermoelectric elements, the current density j will be the same for both elements if the cross-sectional areas of the two elements are equal. The properties, such as electrical resistivity and thermal conductivity, are in general different for the two materials. An average value of the properties of the two materials

632

III-1

will be used in solving Eq. 1. This constitutes an approximation that would become objectionable when the properties of the two elements are widely different.

Solution. The solution of Eq. 1, subject to the boundary conditions of Eqs. 2, 3 and 4, may be made by the use of the method of Laplace Transformations. For convenience let

and express Eqs. 1 through 4 in terms of U instead of T. Next transform t to s in Eq. 1 such that

1 j2 p U xx (x, s) = ·-· [su (x, s) - U (x, O)] - -

k a s (5)

where u (x, s) = L [U (x, t)], the Laplace transform of U (x, t), and the subscript x designates partial differentiation with respect to x.

Substitution of the boundary condition of Eq. 4 into Eq. 5 gives s Y (x )2 Z ( x ) j2p

U x x (x, s) = -; u (x, s) - -; L - -; L - ks

A Laplace transformation is now made of x to y in Eq. 6 such that

u (y, s) = L [u (x, s)]

(6)

In performing this transformation the expression u (0, s) arises which can be evaluated by transforming the boundary condition in Eq. 2. Thus

u (0, s) = L [ U (0, t)] = L [O] = 0

Making this substitution into the expression for u (y, s) gives

c, 2 y z 1 j2 p 1 u (y, s) =

y2 _ .� - aL2

ya (y2 -�) - aL y2 (y2

-� �)-h y (y2 - ·�) (7)

The remainder of the solution* consists of inverting Eq. 7 back to the x and t vari­ables. :fhe constant of integration C, can be determined by transforming t to s in the boundary condition Eq. 3 and combining with the inverse transform of Eq. 7 after inverting y back to x.

The general solution for the temperature T as a function of the time t and the posi­tion x, measured from the hot junction, following a step change in current is

L { j2pL2 (1 - cos f3n) rxjL + 2 ------.-- + - Th + HL (Th - Tr)

k f3n sm f3n k n = l [ 2 rxjL ] + Y /3n sin /3n (cos /3n - 1) + 2 + k + HL

( rxiL ) ) + z 1 + �k- + HL

* The details of the final mathematical steps are given in reference (4).

(8)

633

III-1

where fJ n is the nth root of the equation

(9)

and where H = hA conv/kA. Of particular interest is the temperature at the cold junction T c, which is obtained

by substituting the value of x = L into the general solution, Eq. 8. Comparison of Work of Other Investigators. Some comparisons may be drawn between

the analytical solution, Eq. 8, and several other studies of transient operation of thermo­electric units that are reported in the literature. Stil'bans and Fedorovich [1] considered the start-up operation where a thermoelectric refrigerator begins operation from a state of uniform temperature throughout the elements. The electric current was maintained constant, thermal capacity of the metal at the cold junction was considered, but the cold junction was insulated so no heat transfer by convection was permitted. Equation 8 becomes the solution of Stil'bans' problem when x = L, H is set equal to zero to de­note the absence of convection heat transfer, and Y = Z = 0 to express the initially uniform temperature distribution.

Gray [2] studied the step change from one operating current to another, which is similar to the problem reported in this paper, although his approach was to represent the per­formance equations in terms of incremental changes from steady-state operating con­ditions. For example, this approach gives the incremental change in the cold-junction temperature when the current is suddenly changed from one constant value to a slightly different constant value. The results of this approach are especially convenient for applications to control systems, and the approach also permits an approximate solution when the properties vary linearly with temperature. The accuracy of the solution using small-signal changes diminishes as the change becomes large.

Alfonso and Milnes [3] investigated the start-up operation where a portion of the lengths of the thermoelectric elements are surrounded by the insulation of the enclosed volume that is to be cooled. A special case of the Alfonso and Milnes problem where the thermal capacity of the insulation is ignored coincides with the special case of the problem studied in this paper where in Eq. 8 the values of H, Y and Z are equated to zero.

PART II. EXPERIMENTAL INVESTIGATION Object. The purpose of constructing and operating the experimental model was to

determine the magnitude of the deviations from the mathematical model. Deviations would be expected because of idealizations in the mathematical solution such as neglect­ing variations of properties with temperature and between materials, electrical re­sistance at the junctions, and deviations from one-dimensional flow of electricity and heat.

Apparatus. The function of the experimental equipment was to measure the cold­junction temperature of a thermoelectric couple following a jump change in current while the temperatures of the hot junctions were maintained constant. The facilities. were designed to control the rate of heat added at the cold junction with an electric heater to duplicate a variable refrigeration load and therm::.i capacities other than that inherent to the cold junction and heater assembly. The experimental work [4] was con­ducted in the Ray W. Herrick Laboratories at Purdue University where a thermo­electric couple of p-type and n-type bismuth telluride materials 2,89 cm long and 1 .26 cm in diameter was mounted as shown in Fig. 1 . The two thermoelectric elements were mounted on the same axis with the cold-junction and heater assembly soldered in po­sition at the center. The hot junctions on either end were cooled by water flowing through thin copper tubes soldered to the thermoelectric elements (see Fig. 1). The temperatures of the water streams were independently controlled to maintain constant hot-junction temperatures. To minimize heat transfer by convection, an evacuated chamber surrounded the couple. The vacuum chamber was then immersed in a water bath of a constant temperature equal to that of the hot junctions.

634

III-1

Measurement of Properties. Knowledge of the properties of the thermoelectric ma­terials was needed so that appropriate values could be substituted into Eq. 8. The property measurements were made as follows :

Fig. r . The Test Thermoelectric Couple Mounted in Position with the Vacuum Tank Removed.

a) Electrical resistivity. By sending a known current through the couple while supply­ing heat at the cold junction to maintain its temperature equal to that of the hot junc­tions, the thermoelectric emf vanished and the voltage drop through each element was due solely to its resistance. These measured resistivities are shown in Fig. 2.

1 . 1------------------------�

TEMPERATURE, 0c

Fig. 2 . Electrical Resistivity of Thermoelectric Elements.

b) Seebeck coefficient. The Seebeck coefficients, shown in Fig. 3, were measured during open circuit tests where the temperature of the cold junction was elevated by supplying energy to the cold-junction heater. The difference in emf between the hot and cold junctions of each element per ° C temperature difference is the Seebeck co­efficient.

c) Thermal conductivity. This property was determined in two ways. The first method was to use the known rate of heat flow and the temperature difference between the hot and cold junctions obtained in the tests of the Seebeck coefficient. The second method was to combine the values of the Seebeck coefficients and resistivities in Figs. 2 and 3 with steady-state tests where the cold junction temperature was measured at specified refrigeration loads. Substituting these data into the steady-state performance equation[5j, the equation was solved for the one unknown, the mean thermal conducti­vity. These measurements of k from steady state tests are also shown on Fig. 4.

Some deviation between the two methods of determining k was expected because of radiation between the thermoelectric elements and the surroundings and because of

635

III-1

thermal conduction through the power leads that brought current to the cold-junction heater. The effect of this radiation and conduction causes the values of k from steady­state tests to be too high and those from direct measurements to be too low. Although the true value of k probably lies somewhere between the values indicated by the steady­state tests and those by direct measurements, the values used to predict transient per­formance were those from steady-state tests since the radiation and conduction effects would be present in somewhat the same degree during the transient tests.

u 0

� 0.24 Q. VJ '::i � 0.22 0 0: <.) :1 ,__.. 0.20 z w u Li: � 0.1 8 0 <.) "' <.) w el 0. 1 6 10 m

i--- • .

-----� ---

�

15 20 25 30 TEMPERATURE, oG

35

N - ELEMENT

I P- ELEMENT

---�

40 45

Fig. 3 . Seebeck Coefficients of Thermoelectric Elements.

Transient Tests. The conditions under which the experimental transient tests were performed are summarized in Table 1 . The tests are in three groups : (I) tests 1 through 4 where no electrical energy was added at the cold junction, (2) tests 5 through 10 where heat was supplied by the electrical heater at a rate proportional to the temperature difference between an imaginary fluid and the cold junction to simulate convection, and (3) tests 1 1 and 12 where the cold-junction heater simulated additional thermal capacity in raising the value of Ma/kAL from the value of 0.197 inherent to the equip­ment to an artificial value of 1 .0.

� 0.030 ' ::;: '2 VJ I- 0.028 �

,: I- 0.026 � I-(.) ::::> 0 z 0 0.024 <.) _, <I ::;: 0: w 0.022 I >--

z <I w ::;: 0.020 0

636

I I

0 DIRECT MEASUREMENTS • STEADY STATE TESTS

� v • __,,.,.-. .

• . v • T. � . • • / � .

__,.,. r"' .......

5 10 15 20. 25 30 35 40 45 TEMPERATURE, 0G

Fig. 4. Mean Thermal Conductivity of Thermoelectric Elements.

Table 1. Transient Test Conditions

Initial Final Ma ajL j2pL2 Simulated Test Th Tr y z HL Fig. No.

Current, Current, kAL k k No. Conv. Thermal

amps amps O K O K O K O K O K Capacity

0 10 310 0 0 0 0.197 0.19 19.9 5 No No

2 10 0 310 -9.95 -31,2 0 0.197 0 0 5 No No

3 0 20 310 0 0 0 0.197 0.38 79.9 6 No No

4 20 0 310 -40,0 -16.5 0 0.197 0 0 6 No No

5 0 10 310 310 0 0 1 .08 0.197 0.183 18.7 7 Yes No

6 10 0 310 310 -9.35 -11.6 1.08 0.197 0 0 7 Yes No

7 0 20 310 310 0 0 1 .14 0.197 0.377 79.0 8 Yes No

8 20 0 310 310 -39.5 8.6 1 . 14 0.197 0 0 8 Yes No

9 10 20 310 298.2 -13.3 -9.59 1 .785 0.197 0.367 77.0 9 Yes No

10 20 10 310 298.2 -38.5 7.78 1 .785 0.197 0.189 19.2 9 Yes No

1 1 0 10 310 0 0 0 1 .00 0.19 19.9 10 No Yes

12 0 20 310 0 0 0 1 .00 0.38 79.9 10 No Yes ...... ...... °' ...... (;) I

'I -

III-1

Comparison of Experimental and Calculated Results. The results of the experimental tests described in Table 1 are shown in Figs. 5 through 10. The calculated results stem from Eq. 8 for the value of x = L using the experimentally-measured property values. The agreement between the experimental and calculated results is generally good. The greatest deviation appears in Fig. 9 for the tests with simulated convection where the current jumps between 10 amps and 20 amps. The experimentally measured cold­junction temperature in both of these tests even overshoots the steady-state temperature.

40

(O AMPS) - - -_ __ l -

\ -� �TEST 2 (O AMPS) 30 \ v � ., >( / I �� r"---

u 0 w" a: :J � 20 a: w Q. :lE w I- 1 0 z 0 ;::: u z :J

I - TEST I (10 AMPS) (IO AMPS) D ., ' Cl _J 0 u

��

1w w (.!) --- CALCULATED a: z

�,� -- EXPERIMENTAL u u I I I

-10

0 100 200 300 400 500 600 700 TIME, SECONDS

Fig. 5. Cold-Junction Temperature for Step Changes Between ro Amps and Off.

40

30

� 20 :J � a: w Q. � 1 0 ,__ z 0

,___ (0 AMPS)

1 \ I \, /} /f � �

I } \.

-- ---:::-----

� �EST 4 (O AMPS) /. ..

,,.-:;::.-v

-- - CALCULATED -- EXPERIMENTAL

ti 0 z ;::; ' Cl _J i ,y J " � I � � r---_2EST 3 (20 AMPS)

8 - 1 0

(2Q AMPS - 20 I-z w w to

-30

a: z

ll<{ :J :i:: u u -100 0 100 200 300 400 500 600 700

TIME, SECONDS

Fig. 6. Cold-Junction Temperature for Step Changes Between 20 Amps and Off.

638

-

BOO 900

-· --

BOO 900

III-1

The deviation between experimental and calculated results is at least partly due to the special characteristics of the cold-junction heater. Temperature measurements on the resistance heater wire indicated that this wire and perhaps a portion of the material surrounding the wire were at a higher temperature than the junctions of the thermo­electric material and copper. During a pull-down of the cold-junction temperature in

40

(0 AMPS} 3 5 --�----

<.:> 0

w � 30 �

/ i..-·

\ 0: w a. ::;: � 25

z 0 ;::

! \/' j /\ � � 20

;s ' D _J (10 AMPSt 8 1 5

10 -100

";jw � � 315

0 100

'""'::: � :------- - ---- CALCULATED -- EXPERIMENTAL

I I 200 300 400 500

TIME, SECONDS

I I TEST 6 (O AMPS}

TEST 5 (10 AMPS} ·- - � ·-

600 700 800 900

Fig. 7. Cold-Junction Temperature for Step Changes Between 10 Amps and Off with Con· vection.

40

35

w· � 30 � 0: w a. � 25 ... z 0 f; 20 z ::J ? D _J 8 1 5

10

(0 AMPS}

�I - "'

�I _ !zl

!I I

I (20 AMPS}

5 -100 0

/ -::::'.,,,.

/;/ I/ I \ \\ � '-I'\..',

� �- --100 200 300 400

TIME, SECONDS

I 8 I TEST .JO AMPS}

-· - -

--- CALCULATED -- EXPERIMENTAL

TEST 7 (20 AMPS} I ...,

500 600 700 800 900

Fig. 8. Cold-Junction Temperature for Step Changes Between 20 Amps and Off with Con· vection.

639

111-1

'-' 0 w· a: :::> � a: w 0.. ::;: w >-z 0 ;::: u z ;; 0 _J 0 u

16

14

12

10

8

6

4

(IQ AMPS)

I

I v ,,,,..,.,--

1 // \ 1\ \' ............ --

� O AM:I " .....

z w -�1� u u

------------- --- i-----TEST 10 ( 10 AMPS)

----i.- ·- .2.fST 9 �� AM�� -- I I I

- -- CALCULATED -- EXPERIMENTAL

-50 0 50 100 150 200 250 300 350 TIME, SECONDS

·-

400

Fig. 9. Cold· Junction Temperature for Step Changes Between IO Amps and 20 Amps with Convection.

u " w a: :::> � a: w 0.. ::;: w >-z 0 ;::: u z :::> ? 0 _J 0 u

40 (0 AMPS) � 30

20 \� \ '\ I

10

0

I I I

·10 �,_I z w ::!I� ·20

·30 ·200

a: <( :::> :r u u

I 0

\ ' ' � "' � � ::-- TEST I I ( 10 AMPS)

' r--=-- --- - --I-- - - -

� �, !'--.JEST 12 (20 AMPS)

- - · CALCULATED r--=�- -� - ·- - ·--- EXPERIMENTAL

I I I 200 400 600 800 1000 1200 1400 1600 1800

TIME, SECONDS

Fig. Io. Cold-Junction Temperature for Step Changes from Off to IO Amps and 20 Amps with Thermal Capacity.

640

III-1

a test with convection the rate of energy supplied to the heater increases rapidly. Some of the energy is needed, however, to build up the excess temperature of the heater wire and its surroundings. Because less energy than anticipated is flowing to the true junc­tions from the resistance heater, the cold-junction temperature decreases more rapidly than expected. The opposite process occurs when the current is reduced from 20 amps to 10 amps. In the latter case some of the energy in the excess temperature of the heater assembly flows out to the junctions at a greater rate than expected.

Summary. An analytical expression is developed for the transient temperature distri­bution in the elements of a thermoelectric refrigerator following a step change in current. Certain assumptions were made in the derivation of this analytical expression including the assumption of constant values of each of the properties - the Seebeck coefficient, the electrical resistivity, and the thermal conductivity. Experimental measurements on a test thermoelectric couple of the transient cold-junction temperature showed close agreement with the analytical results. The property values substituted into the analytical expression were mean values for the two thermoelectric materials and over the range of operating temperatures. The deviation between experimental and calculated results that did occur is attributed more to the failure of the cold-junction assembly to be at a uniform temperature at any instant than to variations of properties with temperature.

Acknowledgments. The Purdue Research Foundation provided the grant to support the purchase of equipment and supplies for the experimental study. The thermoelectric materials were provided by Ohio Semiconductor Products, Columbus, Ohio.

REFERENCES

r . Stil'bans, L. S. and N. A. Fedorovich, ,,On the Operation of Cooling Thermoelements under Non-Stationary Conditions", Zhur . Tekh . F iz . , v. 28, Mar. i958, pp. 489-492.

2. Gray, P. E., The D yn amic Behavior of Thermoelectric Devices , Technology Press and Wiley and Sons, Inc., New York, i960.

3 . Alfonso, N. and A. G. Milnes, ,,Transient Response and Ripple Effects in Thermoelectric Cooling Cells", E lectr ical Engineering, v. 79, no. 6, June i960, pp. 443-449.

4. Stoecker, W. F., ,,Transient Behavior of Thermoelectric Refrigerators under Step Changes in Current", Ph. D. Thesis, Purdue University, August i962.

5. Altenkirch, E., ,,Electrothermische Kalteerzeugung und Reversible Elektrische Heizung", Phys. Zeits., i2, 920, i9r r .

641

Come-Back of the Absorption Refrigerator?

Retour de refrigerateur a absorption ?

Dipl.-Ing. HANS STIERLIN SIBIR-Ktihlschriinke, Schlieren, Switzerland

IIl-6

SOMMAIRE. Il est de plus en plus demande au refrigerateur menager de contenir un veritable compartiment de congelation a -18 ° C ou mains. Des conditions recommandees corres­pondantes sont en cours d'elaboration dans de nombreux pays.

Meme parmi les specialistes, il s'est repandu l'idee que le refrigerateur a absorption ne pouvait pas remplir ces conditions, ou ne le pouvait que tres difficilement. On montre au contraire ici que le refrigerateur a absorption presente, dans ces conditions, la possibilite de Zutter parf aitement avec le refrigerateur a compression.

Les raisons qui nous ont amene a cette affirmation sont les suivantes : 1. Une comparaison des cycles de Carnot des deux systemes montre une amelioration relative

considerable de l' absorbeur quand la temperature a l' evaporateur s' abaisse. Z. Si seule une petite par tie de la marche doit se faire a basse temperature, le groupe a

compression des refrigerateurs doit effectuer toute la marche a cette temperature basse. II en va autrement avec le groupe a absorption pour ses temperatures defonctionnement. La temperature moyenne de l' evaporateur est alors beaucoup plus elevee et il en resulte une autre amelioration de ces groupes.

3. Les rendements de Carnot des absorbeurs construits en pratique sont environ deux fois superieurs a ceux des groupes a compression comparables. C' est le resu/tat de la relation parti­culiere entre la temperature du bouilleur et de la temperature de l' evaporateur dans le cycle de Carnot des absorbeurs a NH3-H20.

4. Les possibilites specifiques de reglage de la temperature et du degivrage des groupes a absorption entrainent de nouvelles ameliorations par rapport au compresseur.

5. La puissance frigorijique du compresseur a un etage s' abaisse brusquement en f onction de l'abaissement des temperatures de l'evaporateur. La puissance frigorifique de /'absorber reste presque constant. La comparaison de la puissance frigorifique a basse temperature devient done tres favorable pour l' absorbeur.

6. Taus ces facteurs sont enfaveur des veritables groupes frigorifiques a deux temperatures (-l8°C, +5°C) dans lesquels : a) la consommation de courant de l'absorbeur n'est que 1,7 fois celle du compresseur (1 : 3,4

a -5°C) b) La puissance frigorifique des deux systemes est presque le meme ( 1 : Z,8 a -5° C) .

7. En utilisant du polyurethane expanse au Freon comme isolant, on peut obtenir la meme consommation et la meme puissance frigorifique avec les refrigerateurs a absorption et les refrigerateurs a compression.

8. Meme avec un groupe un peu plus couteux et une isolation un peu plus couteuse, le cout d'exploitation de l'absorbeur est encore plus faible que celui du compresseur de ZO% environ.

Conclusions: Si en fait on peut obtenir la meme puissance frigorifique et une meme consom­mation pour les deux systemes, il est possible que l' absorbeur entraine par electricite puisse finalement, en raison de ses avantages traditionnels (absence de bruit, longue duree, faible cout de production, remplacer le compresseur pour les refrigerateurs menagers jusqu'a ZiiO l.

In the last few years the absorption refrigerator was falling far behind the compression refrigerator in most countries. The two main reasons for this may be known : high power consumption ; small reserve of power. This meant, at least for responsible firms, that electrically driven absorption-refrigerators with a volume greater than 120 litres could scarcely be put on the market. In addition the name of the absorber was unfortunately badly damaged by innumerable defective products.

643

III-6

Recently also in Europe a household refrigerator is more and more demanded which offers a real deep-freeze compartment at not more than -18°C. Corresponding standards are in preparation in many countries.

Among experts themselves the view is spreading that such conditions could hardly be fulfilled by the absorber, and that therefore it would finally have to disappear.

It is hence time that once more someone breaks a lance for the absorber. Careful in­vestigations led namely to the somewhat unexpected result that exactly under such con­ditions the absorber offers the possibility of being fully competitive with the compressor in respect of performance and running costs, even when electrically driven.

This short paper does not claim to contain highly theoretical investigations which even the expert might scarcely follow. It is much more concerned to point out certain facts and tendencies of a thermodynamic, technical and commercial nature, which could lead to changes in the field of domestic refrigeration.

COMPARISON OF THE COP OF THE CARNOT-CYCLES FOR COMPRESSORS AND ABSORBERS

Fig. 1 shows the COP of the Carnot cycle for the compressor and the one-step continuous NH3-H20 absorber.

The calculation of the COP of the Carnot-cycle for compressors is known.

, Tevap e C = -------·

Tcond - Tevap

For the absorption system the COP of the Carnot-cycle is : ( I I ) • (-J- _ __ !_) _ LI tho:_ • Tevap . Tcond Cc =

T Abs: -TBoller ' Tevap Tcond - LI tcold Tboller Tabs.

1\� -,.__ � r-·" 1, �s. � 1. <( -· (;J<("

l'--i "" .0 .;> � ..:· -

1/ .. · -

l 0 -5 -8 -zo -25

"'- ......... ' --.r

-40 tevap.°C -60

Fig. r. COP of the Carnot-cycle for a compressor (ec) and a one-step, continuous NH3-H20 ab­sorber (Cc) at different evaporator temperatures. (tcond-abs = 45°).

Tboller results from the phase-diagram and is therefore strictly linked to the other temperatures. (For proof see "Handbuch der Kiiltetechnik" vol. VII.)

Consideration of Fig. 1 at once shows that the COP of the Carnot-cycle of the com­pressor drops off very rapidly with falling temperature of evaporation, while the COP of the Carnot-cycle of the absorber only changes a little. The latter is somewhat unexpect­ed and should therefore be explained briefly.

644

..! •c t I

E I C I \ \

-so --+--+--:::'.'!11.l!IHH

0 -20 -40 -60 •c tevap.

III-6

Fig. 2. The temperatures of the Carnot-cycle of one-step, continuousNH3-H20 absorbers at differ­ent evaporator temperatures.

Fig. 2 represents the temperature of the boiler of the one-step NH3-H20-absorber as a function of the evaporator-temperature at given condenser and absorber temperatures. One sees from this that each evaporator-temperature has a fixed corresponding boiler­temperature. The COPc therefore only drops off slowly because as Li t of the cold stage increases the LI t of the hot-stage also increases. With the compressor-considered similarly as a pure heat-engine with infinite LI t of the hot stage - LI t of the hot-stage on the contra­ry naturally remains constant.

In Fig. 1, for the evaporator temperatures -5° and - 25°, the corresponding ratios of the COPs are indicated. At -5° C it is 1 : 6.3, at -25°C 1 : 4.7. From this it follows that with the transition from a normal refrigerator to one with a deep-freezing compart­ment, the relative power consumption of the compressor in comparison with the absorber increases sharply, namely about 34%. This is valid when both machines are carrying out the entire performance at the same evaporator temperature.

For a refrigerator with a deep-freezing compartment only approx. 20% of the cold produced has to be used at the temperature of the deep-freezing compartment. The rest serves to cool the food compartment at a considerably higher temperature. Since the com­pressor cooling system normally used for refrigerators can carry out the cooling process only at one and the same temperature, the whole cooling process must take place at -25° C. It is otherwise with the absorber as effectively built. Here in consequence of the range of de-gassing, the cooling takes place over a wide range of evaporator-temperatures, from which suitable parts can then be chosen for the two different purposes. It is therefore admissible, as has been done in Fig. 1, to compare the COP c of a compressor with an evap­orator temperature of -25°, with the COP c of an absorber with an average evaporator temperature of -8°. This ratio is 1 : 4.1, and the relative increase of power consumption of the compressor rises to 53 %.

645

IIl-6

COMPARISON OF THE EFFECTIVE COP FOR COMPRESSORS AND ABSORBERS

It must now naturally be asked whether such purely theoretical comparisons correspond with actual practice. For the comparison a commercially produced 1/12 HP compression­unit and a 120 Watt continuous one-step NH3-H20-H2 absorption unit were used.

The cooling-process of a compressor unit can be taken as known.

•c

0 4

-2•

one - step ga•-heat -exch.

two - step gas - heat-exch.

theoretical

a�c Watts

l'ig. 3. Temperature ranges within which the cold is produced in absorption-units. (lcond"abs = 45°).

Fig. 3 shows the temperatures at which the cold is available in an absorber. At a maxi­mum boiler-temperature of 180°C, a de-gassing range between 21 % and 9%, and an actual condensor and absorber temperature of 45°, the theoretical cooling process falls between-54° and-35° in the way shown. As built until now, it becomes in consequence of irreversibilities measurable for example between -12 and -2°. With somewhat in­creased expense, especially by changing over to a two-step gas heat-exchanger, it is possi­ble however to split this evaporator-temperature-field into two different areas. These lie between -30° and -25°, and between -16° and -5°, and their size is suited to the actual cooling requirements. Fig. 4 shows schematically the present and the improved design of such an absorber. (For interest it may be added that even temperatures of -52° can be made measurable.)

After the explanation of this fact, we can now put forward the practical results of the two systems being compared, deduce the Carnot efficiency and compare the two systems together.

In Fig. 5 the results mentioned are shown as a function of the evaporator temperature. From this it is first of all striking that the Carnot efficiencies of the absorber are essen­tially better than those of the compressor. This is not a result of more careful construction of the absorber, but is once more a result of the peculiar relation of the temperatures of the Carnot-cycle of a one-step NH3-H20 absorber. It will be profitable to return to this shortly.

646

111-6

SIBIR

Fig. 4. Conventional and improved layout of absorption units. r = boiler ; 2 = condenser ; 3 = ice making evaporator ; 3 a= deep-freezing evaporator ; 4 = cabinet evaporator ; 5 = one-step gas heat exchanger ; 5 a = supplementary stage of gas heat exchanger ; 6 = absorber ; 7 = reservoir.

COP

.. \ COMP "'ESSOR ABSO RBER \5,4 ' �- -----..

' ' ' ' .. ..

'rts .. ,

E.c "' 0 :;j- �-.. u .. f "'"

� " ... ;;- �

o" .... :;u ......_ I I ---...ie.z... �--c"" -- - -.. --1�5 -t

E.efl 36 \11 0 0 -10 -20 -30 - 1 0 - 2 0 -30 •c tevap

Fig. 5. COPs of the Carnot-cycles and the practical cycles of a 1/12 HP compressor and a 1 2 0 Watt one-step continuous NH3-H20-H, absorber at different evaporator temperatures. At -5°C and at -25° C the Carnot efficiencies are indicated. (lcond-abs = 45°).

In Fig. 6 three possible temperature patterns are shown in the Carnot diagram, whereby the evaporator temperatures lie between -20° and -5°, and the boiler-temperatures between 160 and 190°. With the absorption-refrigerator, which is driven not by residual

647

III-6

heat, but predominantly by electricity, the practical height of this boiler-temperature is limited mainly by the danger of corrosion.

-20 -20 -40 -&o •c '"""· Fig. 6. Temperatures of the Carnot-cycle of one-step NH8-H,O absorbers and compressors, the

latter being considered as a pure heat-engine with infinite Lit on the hot side. The practical cycles I, II and III show different "areas of tolerance" A, B and C.

In case I a large temperature-margin exists on the boiler side ; in case II this is shifted to the cooling side. By a suitable choice of the cycle this temperature-margin can be divid­ed as required, which is shown in case III. Hence there arise in contrast to the Carnot process definite areas of tolerance (A, B, C), which in practice make it easier to approach the Carnot process. With the compressor such tolerance areas are not available, since the upper temperature lies uniformly at infinity and the lower temperature directly represents the actual evaporator temperature.

From Fig. 5 the effective COP for the compressor and absorber at different freezer-evap­orator temperatures can now be deduced. In diagram 7 these are shown for -5° and -25°. From this comes an interesting fact : At -5° the COP of the compressor is still 3.4 times higher than that of the absorber, but at -25°C this ratio is in contrast 1 : 1 .9, therefore reduced by almost a half. The change in favour of the absorber is therefore even essentially greater than by the theoretical comparison of the Carnot processes.

COMPARISON OF THE PERFORMANCES OF THE COMPRESSORS AND ABSORBERS

The tendency of the COP, which is very favourable to the absorber at falling evaporator temperature, is supplemented by a still more favourable tendency in the comparison of the performances. Again a normal commercial 1/ 12 HP compressor is compared with a 120 Watt absorber. The relations of the performances thus arising are shown in Fig. 8, again as a function of the temperature of the freezer-evaporator. The performance of the absorber remains approximately constant, the compressor on the contrary drops very steeply because the COP worsens, and especially because the volumetric-efficiency becomes worse. The resulting performance ratios amount to 1 : 2.8 at -5°, and to 1 : 1 .35 (less than half) at -25°.

648

COP

2.0

1.5

1.0

0.5

0 0

Iff-6

� rt

"' ti

I � compressor __, � I jbsorber -1 ·10 -20 -30 •c 1Freezer

Fig. 7. Comparison of the COPs of a 1/12 HP compressor unit and a r20 Watt one-step continuon NH8-H,O-H, absorption unit at -5°C evaporator temperature and --25° freezer temper ature. (lcond-abs = 45°) .

Qo (Wat� 200 •

..i "'

- �

00���-�10:---�--�20,..-�--�30.....,.1 r-.-•• -.-'.,

Fig. 8. Cooling performance of a 1/12 HP compressor and a 120 Watt one-step continuous NH3-H20-H8 absorber at different freezer temperatures. (lcond-abs = 45°).

FURTHER CORRECTION FACTORS IN FAVOUR OF THE ABSORBER

In a true two-temperature refrigerator a temperature of-18° at the highest is required in the freezer. Variations in this temperature, which are due to the automatic temperature control, must therefore take place below -18°.

The compression refrigerator is today exclusively controlled by an ON-OFF regulator. The absorber on the other hand can be regulated continuously (with gas-heating) or by switching back the input to a low power, still sufficient for the freezer (e. g. 120 Watt-40 Watt - 120 Watt - 40 Watt . . . . ). In consequence of the decrease of the absorber tem­perature and the partial-pressure-differences at lowered input it is in this way possible to obtain the equivalent of the whole 40 Watts at temperatures below -25°. Fig. 9 shows the effect on the evaporator temperature of these two methods of regulation. From this it is clear that the average evaporation temperature with on-off regulation lies approx. 2-3 ° lower than with the absorber regulation. Hence the effective COP of the com­pressor falls from O. 7 to0.62, while that of the absorber remains at the previous value (0.36).

649

rn--6

time l I I

reguliltion (comp,..,ssor) I

120W-'°W-120W···

t Freezer

regulation -Ee-+--+,�'� (absorber)!

-10 -15 -20 -25 -30

.Fig. 9. Temperatures of the freezer and the evaporators during thermostatic tcrnperaturc control with either on-off or 120 W - 40 W - 1 20 W - 40 vV . . . . . switching.

In this connection may be mentioned another very important advantage of this possibi­lity of regulating the absorber. Without any difficulty the temperature of the storage compartment can be regulated quite independently of the freezer. This is known to be very difficult with the compressor, which can often only be controlled by periodic introduction of external heat into the refrigerator.

The periodic defrosting of the evaporator raises similar problems. By the 120 - 40 - 120 - 40 Watt switching, the storage compartment evaporator can, without affecting the freezer, be held at temperatures above 0° until it is free from frost. With the compressor a defrosting of the storage compartment evaporator without effect on the freezer is only possible by the introduction of external heat. This however produces an increase in the average power consumption of the compressor in comparison with that of the absorber, which can be estimated at approx. 5 o/o.

THE FINAL COMPARISON OF THE TWO UNITS

Fig. 10 shows the actual ratios of the energy-consumption and also the ratios of performances of the absorber and compressor units after consideration of the corrections for regulating and defrosting. Hence results for the true two-temperature unit :

1. The cooling performances of the two systems are almost equal. 2. The power consumption of the absorber is only 1 . 7 of that of the compressor, if

both refrigerators are similarly insulated. It can be proved that with a power consumption ratio of less than 1 : 2, the difference

can be eliminated by improved insulation of the refrigerator. In our case to attain this the cabinet heat loss-factor (KF) of the absorber must be about half that of the compressor. Since certain constant heat-losses cannot be corrected, this means a trebled improvement in the actual insulation, or if the same insulation material is used, an approx. trebled wall-thickness.

Compression refrigerators are so insulated today that condensation on the outer sur­face of the refrigerator is just definitely eliminated. This condition results for glasswool

650

lII-6

1 Q o WP (Watt) ., 3 - � � 150 "' .., - I

<--- absorber __ __. 2 '-- -· ------ -r-- 100

4ffrMti"'l l---t CHl1'9\ -- -h

50

compr. h 0 0

0 -10 -20 OC tFreezer 0 -10 -20 OC tFreezer

fig. 1 0. Comparison of the power-consumption ( C�

P) ofa 1/12 HP compression unit and a 120 Watt

NH3-H20-H9 absorption unit at -5°C and -25°C freezer temperatures and comparison of the performance of these units at these temperatures. (lcond·abs = 45°).

in an insulation thickness of 5-cm. According to the above corrected conditions this would require an insulation for the

·absorption refrigerator of 15 cm glasswool, a construction

which would naturally be unsaleable. However for some time an insulating material has been on the market which has a heat

conductivity about half that of glasswool, namely freon-foamed-polyurethane. With this insulation-material it is now actually possible to provide an absorber with an insulation­layer only 7 cm thick (the corresponding insulation thickness for the compressor would be 2.5 cm), making it equal to the compressor. Such an insulation layer of 7 cm is per­fectly acceptable, especially when one considers that the compression unit takes up more room than the corresponding absorption unit. The ratio of gross volume to the storage volume is therefore approximately equal for both types of refrigerators, inspite of the increased insulation of the absorber.

THE COMPARISON OF TWO REFRIGERATORS

By the addition of this last modification it is therefore possible to produce an electrically driven absorption refrigerator whose power consumption and performance are equal to that of a comparable compression refrigerator.

A refrigerator built in this way was first handed over in 1962 to the "Institut du Froid" in Paris for testing, and fulfilled the strict French conditions for power consumption and performance of compression refrigerators.

COMPARISON OF THE PRODUCTION-COSTS OF SUCH REFRIGERATORS

It was always a traditional advantage of the absorption refrigerator that its production­costs were basically less than those of a compression refrigerator. One must inquire whether this advantage may not have been lost by the production of a somewhat more complicated unit and the more expensive insulation.

Since the basic cabinet, accessories, thermostat and packaging etc. are the same for both types, the investigation of this question can be limited to the price differences of the units and insulation. This results in :

Production price for

Unit price Insulation price

compressor approx. 90.- SFr. approx. 8.- SFr.

98.- SFr.

absorber 45.- SFr. 20.- SFr.

65.- SFr.

651

IIl-6

The compressor is therefore still approx. 33.- SFr. more expensive than a similar absorber, and that is about 20% of the whole cost of the refrigerator.

CONCLUSIONS

If in fact performance and power-consumption parity can be attained for the two systems, then there is no reason why the absorber - also the electrically powered absorber - should not be fully competitive with the compressor. This is especially so because of the tradition­al advantages of the absorber- noiseless, long life, low production costs. It is even possible that the absorber will surpass the compressor in the region of cabinet volumes of less than 250 litres, as has been the case up till now with refrigerators of less than 100 litres volume. But, indeed, it would be a great pity if once more inexperienced firms brought inefficient products on to the market and so once more the absorber fell unjustifiably into disrepute.

652

IIl-27

Analysis of Actual Processes in a Lithium Bromide Absorption Ma­chine

L'analyse des processus reels de la machine frigorifique a absorption a bromurc de lithium

L. M. ROSENFELD and M. S. KARNAUKH Leningrad Technological Institute of the Refrigerating Industry, Leningrad, u. s. s. R.

SOMMAIRE. La machine jrigorifique a absorption a bromure de lithium a ete essayee par les AA. On indique les resultats d'essai de la machine fonctionnant suivant le cycle frigorifique et suivant le cycle de la pompe a chaleur. On a determine les valeurs des pertes des cycles reels dues a la saturation incomplete, a une echange de chaleur partiel et a la dif­ference finale de temperatures correspondants aux pertes de charge a l'interieur du systeme.

L'analyse des donnees obtenues avec la machine experimentale a ete obtenue a l'aide de la theorie thermodynamique de la machine a absorption.

A lithium bromide absorption machine is a highly efficient heat utilizing system for the production of cold and transformation of heat to a higher temperature level [1, 2, 3, 4]. An investigation of its working processes and their actual losses is of theoretical and practical importance [5, 6, 7, 8] .

The determination of the losses in the actual processes has been effected by means of an experimental plant, operating for cooling and heating. The results of the tests are given in Table 1 .

Table 1 . Results o f testing a refrigerating machine and heat pump

Refriger-Item Name of value ating Heat No machine pump

1 2 3 4

1 Water temperature at evaporator outlet, t0 ° C 4.9 53.4 2 Heating medium temperature, th ° C 102 59.8 3 Water temperature at absorber inlet, twa ° C 24.9 75.7 4 Water temperature at absorber outlet, t wa1 ° C 29.0 77.8 5 Water temperature at condenser inlet, t w ° C 29.0 1.4 6 Refrigerating capacity, Qo kcal/hr 27400 7 Heating capacity, Qa kcal/hr 21400 8 Maximum solution temperature in generator, t2 ° C 88.6 52.0 9 Minimum solution temperature in absorber, t4 ° C 33.8 86.5

10 Vapour pressure in condenser, P mmHg 43 8.0 1 1 Vapour pressure in generator, P h mmHg 43 8.9 12 Vapour pressure in absorber, Pa mm Hg 5.5 109.0 13 Vapour pressure in evaporator, P''o mmHg 5.8 109.0 14 Vapour pressure, corresponding to water temperature 6.5 109.0

at evaporator outlet, Po mmHg 15 Actual value of strong solution concentration, gr % 62.3 58.5 16 Actual value of weak solution concentration, ga % 58.6 54.5 17 Mixed solution concentration at absorber inlet, gm % 59.1 55.6 18 Theoretical value of strong solution concentration, gr* % 65.2 63.5 19 Theoretical value of weak solution concentration, ga * % 58.0 54.5 20 Actual heat coefficient, C 0.595 21 Actual transformation coefficient, M 0.38 22 Degree of reversibility, 17 0.244 0.1 17

653

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The power losses in the actual processes are evaluated by the degree of reversibility 11 which for the refrigerating machine is as follows :

t, 1'} = t, 0

and for the heat pump

M 1'} = Mo

(1)

(2)

where : t, and M - actual values of the heat coefficient and coefficient of transforma­tion, which are determined experimentally.

t,0 and M0 - heat coefficient and coefficient of transformation of the system of combined direct and reversible cycles

Thm - Tm Tom t,o = ---- ----

Thm Tm - Tom (3)

where : Thm; Tm; Tom - mean temperatures of heating, cooling and cooled mediums, OK

Tum - Tm Tkm Mo = --- - - ----

Tum Tkm - Tum (4) where : Tum; Tkm; Tm - mean temperatures of heat supply sources, heated body and

cooling medium, ° K When shifting from ideal reversible cycles to actual ones, energy losses are observed

in heat exchange processes in apparatus with a final temperature difference, when throttling and mixing, in the absorption process, when the solution is evaporating in the generator and water is boiling in the evaporator, and as a result of heat losses to the surrounding medium. The relative value of the energy losses are given in diagram 1 (Fig. I) according to the data of one of our tests.

654

Fig. I . Diagram of energy losses in actua 1 processes in refrigerating machine . I - losses in generator in heat transfer process II - losses in condenser in heat transfer process III - losses in absorber in heat transfer process IV - losses in evaporator in boiling process V - losses in throttling processes VI - losses resulting from incom­plete recuperation in heat exchan­ger VII - losses resulting from incom­plete vaporization of solution in generator VIII-losses resulting from incom­plete saturation of solution in absorber IX - losses in solution mixing process X- losses to surrounding medium

III-27

The processes of the theoretical cycles of the refrigerating machine in the liquid zone are characterized in the t;-i diagram by the following curves (Fig. 2) :

4* - 1 * - heating of weak solution in the heat exchanger; 1 * - 1°* - 2* - heating and vaporization of weak solution in the generator; 2* - 3* - cooling of strong solution in the heat exchanger; the solution in the state

3* is throttled, resulting in the formation of moist vapour, containing liquid in the state 3°* (curve 3* - 3°* - isotherm); 3°* - 4* - absorption process.

/ 70

Fig. 2 . Main points of theoretical and actual processes in refrigerating machine in !; - i diagram .

A study of the deviations of the actual processes from the theoretical ones makes it possible to draw some conclusions on the operation of the apparatus.

a) Generator. The theoretical process of vaporization of the solution in the generator follows the isobar P = const, and the state of the solution at the end of the process is defined by the point 2* at the intersection with the isotherm t2• In reality, under the influence of hydrostatic pressure the evaporating process 1° - 2 is effected in a flooded generator at a pressure changing from P to P' and the state of the liquid phase is defined at the end of the process by the point 2. The latter results in incomplete vaporization of the solution, which is measured by the difference of concentration t;,* - !;,. The negative influence of the hydrostatic pressure is intensified owing to the small absolute value of the pressure in the apparatus (35 to 50 mm Hg), while the specific weight of the solution is high (1.6 - 1 .8 kg/I).

The losses are less if the solution boils in several layers or vaporizes in a spraying generator. However, the latter is possible only if the possibility of crystallization on the pipes is eliminated.

b) Absorber. Theoretically the absorption process follows the isobar Pa = const and the change of the liquid phase state terminates at the point 4* at the intersection with the isotherm t4•

The actual absorption process does not follow the isobar and terminates at the point 4, which does not correspond to the equilibrium state. Incomplete saturation of the solu­tion, which is measured by the difference of concentration /;a - /;a*, is observed in the absorption process as a result of irreversible losses.

655

III-27

The degree of incomplete saturation depends on the efficiency of the nozzles. Special attention should be given to the latter when designing the machine.

Recirculation of the weak solution through the absorber is applied in the actual circuit for the purpose of intensifying the process (1). The weak solution in the state 4 mixes in this case with the strong solution in the state 3 and a mixed solution in the state 5 enters the nozzles of the absorber. The mixed solution throttles to the pressure Pa owing to which moist vapor forms, containing liquid in the state 3°. The process in the absorber follows the curve 3° -· 4. The maximum temperature in the absorber drops owing to recirculation of the solution.

c) Evaporator. The vapor pressure in the evaporator Po' is higher in reality than the vapor pressure in the absorber Pa by the value of the pressure losses in the connecting line. The influence of these losses in a lithium bromide machine is quite essential because the specific volume of the water vapour in the working zone of the boiling temperatures is high (125 to 160 m3/kg), while the absolute pressure in the apparatus is low (5 to 7 mm Hg).

The cooling temperature of the water, recirculated through the evaporator, is higher than the temperature corresponding to the vapor pressure therein : Po > P' o and to > to' owing to the irreversible losses in the evaporator during the boiling process, connected with the limited surface area of evaporation and final rate of the process.

The temperature conditions of operation of the heat pump and refrigerating machine are quite different, resulting in a difference in the character of losses in the actual pro­cesses.

The theoretical working processes of the heat pump in the liquid zone are given in the ; - i diagram as follows (Fig. 3) : 4* - 1 * - cooling of weak solution in the heat exchanger; solution in state 1* throttles, forming liquid in the state 1°* (curve 1* - 1°* -isotherm) 1°* - 2* - vaporization of solution in the generator; 2* - 3* - heating of strong solution in the heat exchanger; 3* - 3°* - 4* - process in absorber.

I \

I / / f; fm fz fr fr Concentration f Y.

70

Fig. 3. Main points of theoretical and actual processes in heat pump in � - i diagram.

656

III-27 The influence of the hydrostatic pressure reveals stronger in the heat pump generator

than in the refrigerating machine because the absolute value of the pressure vapor is lower. The actual boiling process 1° - 2 is characterized by a higher degree of in­complete vaporization /;r* - /;r. A loss of pressure, measured by the difference Ph - P, has been observed in the connecting line between the generator and the condenser.

Partial vaporization of the strong solution (process 2 - 3) was effected in the coil heat exchanger of the experimental plant owing to the heat supplied by the weak solution and resulting in a drop of the maximum absorption temperature. The utilization of a multishell heat exchanger makes it possible to eliminate this effect.

It can be considered that the state of the solution at the end of the absorption process corresponds to the equilibrium state (point 4*), and the boiling temperature in the evaporator corresponds to the vapour pressure in the absorber.

During the process of recirculation of the solution through the absorber in the heat pump, the solution in the state 4* mixes with the solution in the state 3, and the mixed solution in the state 5 is supplied to the absorber where the process follows the curve 5 - 3° - 4. * Recirculation lowers the maximum absorption temperature, thereby low­ering the maximum temperature of the hot water, produced in the absorber.

When effecting thermal calculations of the machine, plotting of the main points of the working processes (Figs. 2 and 3) should be performed with consideration of the losses in the actual processes. An assumption is made during the calculations that vap:irization of the solution is effected at maximum pressure P' for the refrigerating machine and Ph' for the heat pump, i. e. the actual process in the generator 1° - 2 is replaced by two processes : 1° - 6 - heating of the solution and 6 - 2 - vaporization at constant pressure.

An analysis of the actual processes points out to ways of improving the lithium bromide absorption machine.

REFERENCES

1 . Berestneff, A. A. New Development in Absorption Refrigeration, Refr. Eng., 1949, No 6.

2 . Plank, R. Amerikanische Kiiltetechnik, Tei! II: Absorptionskiiltemaschinen for Klimaanlagen, Kaltetechnik, 1956, No 10. -

3. Chernobylsky, I. I., Kremnev, 0. A., Chavdarov, A. S. Heat utilizing plants for air conditioning. Kiev, Academy of Sciences of Ukrainian SSR, 1959·

4. Rosenfeld, L. M., Karnaukh, M. S. Dynamic heating by means of a reverse absorption lithium bromide machine, Journal of Technical Physics (USSR), vol. XXIII, iss. 7, 1958.

5. Rosenfeld, L. M., Karnaukh, M. S. Concentration-enthalpy diagram of lithium bromide-water solutions for designing of absorption refrigerating machines, Kholodilnaya Tekhnika, No l , 1958.

6. Hartl, R. Novy druh absorpcniho chladiho zarizeni pro klimatizacni a prumyslove ucely, Potravinarske strojirenstvi a chladici technika.

7. Lower, H. Thermodynamische Eigenschaften und Wiirmediagramme des binaren Systems Lithiumbromid/Wasser, Kiiltetechnik, l96r, No 5 .

8. Zinger, N. M. Calculation of absorption lithium bromide refrigerating plants under alternating conditions, Kholodilnaya Tekhnika, No 2, 1962.

657

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L'emploi de la machine frigorifique a absorption comme « pompe a chaleur »

The Use of an Absorption Refrigerating Machine as a Heat Pump

Prof. A. RAS! Universite de Padoue, Institut de Physique et Centre d'Etudes pour les Appli­cations du Froid, Padoue, Italie

SUMMARY. Some considerations on the principle of an absorption refrigerating machine with the possibility of its use in some particular applications, giving the cycle of operations.

Le developpement toujours plus encourageant de la machine frigorifique a ab­sorption, que l'on a constate en ces derniers temps, fait prevoir un reveil merite de son emploi qui, en verite, avait eu une marche stationnaire clans un passe pas tres loin­tain. Les motifs d'un certain desinteressement pour ce moyen de production de froid ont ere, en plusieurs occasions, toujours au centre de discussions de grande importance, et la situation etait presentee tous ies aspects : de l'aspect technique a !'aspect econo­mique, en enveloppant ainsi les principes de fonctionnement qui ont toujours ete l'objet de vastes et profonds traites en ne negligeant pas la technique de la realisation avec rapport aux depenses pour !'installation et aux frais d'exploitation. Sous ce dernier aspect, un interet particulier a toujours ete provoque par la consideration en rapport a l'energie disponible oit a choisir pour son fonctionnement.

Dans le temps, peu a peu, on a pu ainsi etablir un cadre complet de !'ensemble de !'application de fa<;on a fournir tous les facteurs et les valeurs aptes a determiner un jugement sage et prudent sur le choix a faire. Par consequent, l'emploi avantageux de la machine frigorifique a absorption est connu et on peut, a ce propos, en indiquer la valeur du coefficient d'effet frigorifique ideal qui varie entre les valeurs de 0,994 et 0,616 pour un fonctionnement entre les temperatures de 0°C et -60° C avec des tem­peratures correspondantes de 31,7° et 203,6° et de 15° et 50° au generateur et a l'ab­sorbeur (les valeurs susdites se referent au groupe binaire eau-ammoniaque aujourd'hui presque exclusivement employ{: pour les proprietes de l'ammoniaque, facilement absorbee par l'eau tandis que, avec autant de facilite, on en obtient la disociation avec la chaleur.)

Pour la production de froid a basse temperature en utilisant la chaleur de recuperation au generateur, la machine a absorption presente, par rapport a la machine a compres­sion, un avantage hors de doute et si nous pensons a l'emploi constant et toujours plus grand des basses temperatures demandees pour la conservation et la congelation des produits exposes a la deterioration et pour les nombreuses applications industrielles, nous devons trouver ici le motif d'une affirmation qui s'ebauche de fai;on certaine.

Cependant l'emploi de la machine a absorption a ete peu traite comme "pompe a chaleur", tandis qu'il nous semble, a ce propos, que pour certaines applications quelques aspects interessants devraient etre mis en relief et fournir un orientement certain sur le choix de ce moyen pour la production de la chaleur. Dans ce but, il s'agit de voir, pour la machine a absorption, quelles sont les possibilites d'emploi comme "pompe a chaleur", comme on a fait pour la machine frigorifique a compression.

Ici nous nous proposons de determiner ces possibilites par rapport a la chaleur de fonctionnement et a celle qui est fournie par la machine.

Afin d'entrer rapidement clans les questions de thermodynamique qu'il faudra faire au cours de cette exposition, il ne sera pas inutile de parler rapidement des parties qui constituent la machine a absorption et a son fonctionnement : la machine comprend:

659

III-41

- le generateur, l'absorbeur et la pompe de circulation de la solution qui prennent la place du compresseur clans la machine frigorifique a compression;

- le condenseur, le robinet d'expansion et l'evaporateur qui ont la meme fonction clans la machine a compression et representent !'ensemble ou se produit le froid.

Voici la conception plus simple de la machine a absorption a un seul stade et sur l'etude que l'on en a fait et en posant des elements de comparaison, le cycle theorique est toujours le cycle ferme de Carnot comme d'autre part on fait pour les machines a compression, et on suppose que les transformations successives se produisent a une temperature constante.

Pareillement aux procedes d'etude des cycles de la machine a compression, le cycle interieur d'une machine a absorption peut etre represente clans le plan T, S clans lequel, en considerant !'ensemble comme une machine thermique accouplee a un compresseur ou le travail fourni par la machine est completement absorbe pour le fonctionnement

du compresseur, le coefficient d'effet frigorifique e = �; est egal au rapport de deux

surfaces. Qo est l'effet frigorifique et Qg est la chaleur fournie a la machine thermique.

1 Cependant a l'ordinaire pour la machine a absorption on se sert du plan lg p,

T clans

lequel on met en evidence de fa9on tres claire la relation qui passe entre la pression, la temperature et le titre de la solution frigorifique.

Le cycle parfait pour la machine frigorifique a absorption peut alors etre represente 1

clans le plan lg p, T comme montre la Fig. 1, ou l'on a suppose parfait le groupe binaire

eau-ammoniaque, que la phase de dissociation se produise avec un gradient de la con­centration infiniment petit de fa9on que le point 1 represente la dissociation et le point 4 !'absorption. Dans l'exemple de Fig. 1 les temperatures :

- 20° = to a l'evaporateur + 30° = t au condenseur OU a I' + 91,5 = tg au generateur

-20 .,, 91.5 / o c

Fig. r . Cycle ideal de la machine frigorifique a absorption pour le groupe eau - ammoniac pour temperatures de -20 et 30°C respectivement it !'evaporation et a la condensation ou absorption.

definissent le cycle et indiquent en meme temps les pressions au generateur et a l'eva­porateur. Habituellement on fixe Jes temperatures de vaporisation et de condensation et immediatement du diagramme on a celle au generateur qui est celle qui interesse le plus aux effets du calcul et des dimensions de l'appareil.

Dans le cycle susdit on a suppose t = tc = ta ou tc et ta sont respectivement Jes temperatures au condenseur et a l'absorbeur. Ceci peut etre le cas clans lequel on alimente le condenseur et l'absorbeur avec de l'eau de refroidissement a la meme temperature. On sait que :

1

Ile 1

660

III-41

oil avec Ta To Tc Tg nous avons indique les temperatures absolues a l'absorbeur, au vaporisateur, au condenseur et au generateur, mais que l'on est parvenu a lui en posant

Ta = Tc, et Qo = Qc et Qa = Qg. En partant de ces donnees et en agissant comme on le fait pour la machine frigorifique

a compression, nous pouvons trouver les relations qui fournissent les caracteristiques de la machine a absorption quand on considere qu'elle fonctionne comme « pompe a chaleur ». Dans la machine a compression on a nomme « coefficient de multiplication » le rapport entre la chaleur cedee au condenseur et l'equivalent thermique du travail de compression; c'est autant dire : les calories que nous pouvons obtenir par l calorie necessaire au fonctionnement de la machine; en conservant pour la machine a absorption la meme denomination, le coefficient de multiplication s'exprime de la fac;on suivante :

Qa + Qc Qo ecm = = 1 + - = 1 + cc

Qg Qg pour une machine qui relaise un cycle parfait.

Dans l'application de cette relation pour un champ de temperatures de -20° C a l'evaporateur et + 50°C a l'absorbeur on peut obtenir un tableau, pour le groupe binaire eau-ammoniaque, comme celui qui suit clans lequel sont exposes les valeurs de ecm et de la temperature au generateur :

+ 5

30 58,5 1,968

35 70 1,95

40 81,8 1,932

45 95 1,914

50 108 1,895

to 0 - 5 - 10

65,2 71,7 78 1,942 1,916 1,89

76,4 83 89,5 1,925 1,897 1,873

88,5 94,8 101,6 1,907 1,881 1,855

101,5 106,6 1 13,4 1,889 1,863 1,839

1 12,2 1 19 125,3 1,872 1,847 1,823

'·' +++++++-+-+-+-�-+-!-++-+++++++++-+-+-<

30 JS 'O '5 so • c

- 15 - 20

84,8 91,5 1,865 1,841

95,8 102,6 1,848 1,825

1081 1 15 1,831 1,808

120 127,3 1,816 1,792

133 140 1,799 1,775

Fig. 2. Allure du coefficient de multiplication d'une machine qui realise un cycle ideal pour le groupe binaire eau-ammoniac entre Jes temperatures de +s et de -20°C a l'evaporateur et de +30 et de +so0( a !'absorption OU a Ja condensation.

et que le diagramme de Fig. 2 traduit en y indiquant !'allure en fonction de la temperature d'evaporation et d'absorption.

661

III-41

Dans le fonctionnement de la machine frigorifique a absorption !es temperatures connues sont celles a !'evaporation en tant qu'elles determinent la temperature de production du froid, et celles a !'absorption ou a la condensation, car, pour ces dernieres nous connaissons la source a laquelle doit etre cedee la chaleur de condensation et d'absorption. Ces temperatures sont determinantes meme dans le cas de la machine quand elle fonctionne com me « pompe a chaleur » parce que, en somme, si telle est sa fonction plus importante, elle determine la valeur de la temperature de la chaleur que !'on a a disposition. Comme ii s'agit de chaleur et a temperature relativemem basse, le champ d'emploi de la machine en est limite et c'est en tenant compte de ce facteur que !'on a etabli le tabieau expose ci dessus ou la temperature i est comprise entre Jes valeurs de +30° et +S0° C, intervalle qui, comme on dira plus loin, interesse d'importantes

1 applications. En suivant alors cette idee dans le diagramme logp,

T on peut circonscrire

le domaine dans lequel la machine peut etre utilisee comme « pompe a chaleur » et entre Jes limites considerees auparavant, on a trace le diagramme de Fig. 3 ou l'espace ombrage se rffere a des_ conditions d'utilisation possibles.

Fig. 3. Limites d'application de la machine frigorifique a absorption pour le groupe binaire eau I

ammoniac qui fonctionne comme "pompe a chaleur". Diagramme log p, -T Bosnjacovic-

Wucherer).

Avec l'emploi du diagramme susdit, en partant de la temperature d'evaporation et de condensation ou d'absorption et en nous rfferant au cycle interieur ideal, on arrive a determiner immediatement la valeur de la temperature au generateur, evidemment theorique, mais qui peut fournir un important element caracteristique du fonctionne­ment de la machine. Si ce procede sett a determiner la temperature au generateur, on doit aussi considerer le cas que la chaleur a disposition ait sa temperature bien definie et alors on pourra, avec l'emploi du diagramme, determiner la temperature d'evapora­tion, si la temperature a la condensation ou a !'evaporation est connue ou fixee.

Fig. 4. Temperatures au generateur en rapport a celles a l'evaporateur et a l'absorbeur.

Jusqu'ici nous avons traite le cycle ideal qui fournit !es temperatures theoriques dans Jes differents erats physiques du melange, auquel correspondent naturellement des pressions theoriques, qui doivent se retenir de base pour une preparation orientative

662

III--41

de calcul de !'ensemble qui constitue la machine a absorption. En effet dans le cycle ideal considere on a suppose une circulation infinie de la solution et que par consequent le L1 � entre la solution riche et la solution pauvre tend a zero, et en outre on supposent adiabatiques les expansions soit du fluide frigorifique soit de la solution supposee cons­tituee par un groupe binaire parfait dans lequel le solvant n'ait pas une pression propre de la vapeur.

Dans la pratique on a une circulation finie de la solution de portee bien definie et determinee par les concentrations de la solution riche et de la solution pauvre; les expansions sont isoenthalpiques et les echanges thermiques se produisent toujours avec des pertes plus ou moins considerables ect. ect. De la derive l'utilite a considerer la comparaison entre une machine ideale et une machine reelle, comparaison qui indiquera le merite de la realisation. On pourra, a l'ordinaire, appeler rendement entre une machine reelle et une machine ideale le rapport entre les correspondants « coefficients de multi­plication >) c' est a dire :

cm 'Y/ = ccm

Dans la plupart des cas, en tenant compte de la valeur moyenne du dit rendement, on peut le fixer equivalent a 0,75.

L'etude de la machine a absorption se fair, et il est aise de !'executer, avec l'emploi du diagramme i, � (enthalpie, concentration) qui fournit les grandeurs necessaires pour le calcul des differentes parties de la machine; on peut lui inscrire le cycle de la solution liquide et de la vapeur naturellement dans une transposition du domaine ideal au do-

. ·a. .. c "'

300

250

200

150

100

50

Oom4 ine d• vapeu hum de I I

"'-200° C ; / I Q. � (1. � f--"�+--+--+---+-+---+--+r�--+----+---+---+� � �so� courb d'e u lition

- !OO 0:---,:':o--2:':0,---,,:':o,------oc� oo--�5 oo------fs o=---=17-0 -�ao-o -�9'=0-�100

Fig. 5. Cycle entre les temperatures de +Jo°C a l'absorbeur et -20°C a l'evaporateur dans le diagramme i, � (Merkel-Bosnjacovic) (enthalpie, concentration dans le melange eau­ammoniac).

663

IIl-41

maine reel, surtout en ce qui concerne la quantite de solution en circulation. Evidem­ment ceci est valable meme quand la machine fonctionne comme « pompe a chaleur » et la Fig. 5 considere le cas ou to = -20°C, ta = +30° C et tg = 91,4°C, c'est a dire les limites de temperature considerees plus haut. Avec une construction bien connue on determine les isothermes de la vapeur humide, et de Ia les quantites de chaleur correspondantes a Qa Qo Qc Qg. Dans le cycle de Fig. 5 on a suppose le cas non reel des points 5, 6 et 7 sur les isobares du generateur et de I' absorbeur. Si on determine graphiquement le « coefficient de multiplication », on trouve :

Qc + Qa 475 + 330 em = Qg 488

= 1,65

contre 1,84 du cycle ideal; par consequent avec un rendement de 0.9 environ. C'est un element de grand interet sous !'aspect du jugement definitif sur l'avantage

de l'emploi de la machine a absorption comme « pompe a chaleur » ; nous avons en programme d'executer un groupe de mesures, que nous croyons d'un certain interet, sur une machine deja experimentee a l'Institut de Physique Technique de l'Universite de Padoue, destinees a etablir des valeurs pratiques du rendement thermique.

Par cet expose ont peut deduire que le domaine d'application de la machine a ab­sorption comme « pompe a chaleur », tout en interessant un intervalle limite de tem­peratures, peut avoir des emplois importants. Cependant, sa caracteristique est de fournir de la chaleur a basse temperature, c'est d'autre part ce qu'il arrive pour Jes machines a compression dans lesquelles si on veut elever la temperature de la chaleur fournie, avec un rendement satisfaisant, il est necessaire de recourir a des compressions du fluide executees en plusieurs phases.

La basse valeur de la temperature de la chaleur, limite son utilisation en determinant en meme temps Jes caracteristiques particulieres des installations d'utilisation.

Dans les machines a compression l'avantage de leur emploi comme « pompe a chaleur » est decisif quand il y a la possibilite d'utiliser en meme temps le froid produit et on sait qu'il existe de tels ensembles qui realisent des applications de relief. II est evident que cette condition est aussi valable pour les machines a absorption, pour lesquelles il faut pourtant considerer que, meme quand elles fonctionnent comme machines pour la production de froid, elles sont decidement avantageuses quand pour le generateur on a a disposition la chaleur de recuperation, qui en certains cas, on peut l'avoir meme a des temperatures elevees.

Dans d'autres cas, des conditions particulieres peuvent cependant renare avantageuse la machine a absorption qui fonctionne comme « pompe a chaleur » et de positif on peut dire que la basse temperature a laquelle nous avons la valeur fournie par la machine, peut seulement faire penser a une utilisation :

- dans le domaine du conditionnement d'air

- dans le domaine du chauffage a panneaux - pour la production d'eau chaude pour l'emploi hygienique - sanitaire. La Fig. 6

indique le schema d'une machine simple avec !'utilisation de la chaleur fournie pour le chauffage.

Fig. 6. Schema de machine frigorifique a absorption qui fonctionne comme «Pompe a chaleur» . G generateur, A absorbeur, E evaporateur, H utilisateur de la chaleur.

Dans une autre occasion j'ai fait allusion a une realisation qui concerne !'utilisation pour la production d'eau chaude, ou on a obtenu un coefficient de multiplication equi-

664

IIl-41

valent a 1,45 ayant en meme temps utilise le froid produit; en telle application la machine a absorption a brillament accompli la double fonction.

Et nous voudrions conclure ainsi : que bien des fois on a manifeste de l'optimisme en parlant du developpement de la machine a absorption et de la « pompe a chaleur >) dans un sens general et que ici nous ne pouvons pas nous unir a ceux qui ont con­fiance dans le systeme et nous faisons par consequent un echo a leurs previsions, en esperant que toutes !es contributions portees dans ce domaine seront un encouragement au developpement de ce secteur particulier de la technique.

BIBLIOGRAPHIE

Fr. Bosnjacovic - Technische Thermodynamik

- Regles pour Machines Frigorifiques - etehlies par la Deutscher Kaltetechnischer Verezn - IT F, Paris.

A. Rasi - Passato e presente delle poll'pe di calore - Atti II° Congresso Nationale riscaldamento Padova i 962.

SUMMARY OF THE DISCUSSION (Papers 111-6 + 111-41)

V. S. Martinovsky, USSR: I should like to know why in Switzerland and in many other countries domestic refrigerators are heated by electricity and not by some other more direct heating medium which may be much cheaper, such as gas for example.

H. Stierlin, Switzerland: The price of electricity in Switzerland is in fact cheaper per unit of heat than gas, although the difference in price is very, very slight indeed. There­fore, there is no advantage in heating with gas and the second consideration is, that there is no need to have any gas laid into the house at all. Apart from that very many housewives prefer to have their refrigerators running on electricity. A further conside­ration is, that to run a refrigerator costs no more than 1 S. F. per month, thus one can imagine that the difference will be so small that there is no difference to speak of.

J. F. Downie-Smith, U. S. A. : In America, of course, gas is very often used, especially in those areas where natural gas is available and is therefore much cheaper in those areas.

665

III-13 The Feeding of an Ammonia Absorption Refrigeration System Rec­tifier with Liquid from the Evaporator

Alimentation du rectificateur d'un systeme frigorifique a absorption d'ammoniac en liquide provenant de l'evaporateur

I WALTER KRZYWON, M. Sc. Tech. Designing and Testing Centre of Refrigeration, Cracov, Poland

SOMMA/RE. Le passage de l'eau a l'evaporateur dans un cycle a absorption d'ammoniac, resultant du mauvais fonctionnement du rectificateur, est l'un des problemes caracteristiques que !'on rencontre dans la conception des systemes frigorifiques a absorption. Un drainage approprie de l'eau de l'evaporateur signifie un equilibre de masse du systeme en marche.

Toutes les methodes connues de drainage de l'evaporateur pro•voquent des pertes de chaleur et des troubles du fonctionnement. Dans les machines frigorifiques a absorption recentes en Pologne on a adopte une nouvelle methode mise au point par l'A. Cette methode s'appuie sur l'alimentation du rectificateur en ammoniac liquide provenant de l'evaporateur pendant le drainage de l' evaporateur. Les pert es de chaleur ant ete reduites considerablement et les troub­les de fonctionncment elimines. La caracteristique la plus remarquable de la nouvelle methode est obtenue lorsqu'on fait marcher l'evaporateur avec une charge reguliere, comme on a ten­dance a le faire dans !es nouvelles installations.

Le rapport donne une description detaillee de la nouvelle methode et de son analyse de la chaleur et de la masse. Les conditions de f onctionnement de cette nouvelle methode, appliquee a une f abrique de glace moderne sont etudiees en detail. L'installation dont la puissance frigo­rifique est de 200000 kcal/h (270 kW) est equipee pour la congelation rapide avec des machi­nes a evaporation directe avec des quantites de liquide typiquement faibles du cote basse pres­sion de l' evaporateur.

One of the important problems which are to be taken into account, designing ammonia refrigeration systems1>, is the penetration of water into evaporator caused by imperfec­tion of the rectification process.

That water does not vaporize in the evaporator and has to be removed from it through either continuous or periodical exchange of the evaporator filling.

In case of continuous purging ofliquid, a certain amount of solution 2) with entrained wa­ter is constantly drained from the evaporator. In second case, water is removed periodic·· ally through emptying the evaporator.

In either case, removal of the water is obtained through drainage of the aqua-ammonia solution from the evaporator. This ammonia getting into the absorber causes heat losses which decrease the specific refrigerating effect3l of the absorption system compared to a compression system. That loss of heat can be reduced, particularly by enlarging the rectifier with result that the amount of water entering the evaporator is diminished or, by providing a subcooler in which part of the ammonia drained from the evaporator vapo­rizes, but both methods are far from being satisfactory.

The author presents a fresh approach to the problem.

I only ammonia-water absorption systems are considered 2 practically in any point of the system charged with ammonia and water, we are dealing with

aqua-ammonia solution. Since in this particular case even a small amount of water contained in liquid ammonia is of consequence, the liquid charge in every point is described as a "solution" and the term "ammonia" refers to anhydrous ammonia. In case the vapor solution is considered, the term "vapor" is used.

3 refrigerating effect referred to 1 kg of the refrigerant.

667

IIl-13

The point of the matter is, that the solution continuously drained from the evaporator is fed not to the absorber, but to the rectifier instead. Resulting loss of the solution in the evaporator is at the same time compensated with corresponding amount of solution, flowing continuously from the condenser.

The amount of solution to flow from the evaporator to the rectifier is obtained from equation :

where

qR r = -- kg/kg4)

q s

qR - heat of rectification, kcal/kg4) q s - heat of condensation, kcal/kg4)

(1)

Then, the total amount of solution supplied to the evaporator is (1 + r) kg/kg4), of which 1 kg/kg is the amount of ammonia vaporized in the evaporator.

If the concentration of solution in the evaporator is � s kg/kg, the amount of water ente­ring the evaporator is :

w = (1 + r) · (1 - �.)kg/kg4) (2)

Balance of mass of the evaporation process requires that the same amount of water "r" kg/kg5) is contained in the solution leaving the evaporator. Thus, concentration of solu­tion leaving the evaporator is :

1 �o = - ( r - w ) kg/kg r (3)

The solution in the evaporator has the same concentration. For example, if r = 0.25 kg/ kg and �. = 0.995 kg/kg, then

w = (1 + 0.25) . (1 - 0.995) = 0.00625 kg/kg

1 + 0.25 �o =

0.25 ( 0.25 - 0.00625 ) = 0.975 kg/kg

Feeding the rectifier directly with the solution leaving the evaporator would mean waste of heat. To avoid it, a counterflow heat exchanger was provided, in which the solu­tion from evaporator exchanges heat with solution flowing from the condenser to the rectifier. As a result, the solution entering the evaporator has a temperature slightly higher than the evaporator temperature, and the solution entering the rectifier - slightly lower than condensation temperature. Therefore, feeding the rectifier with solution from the evaporator practically does not mean any loss of heat.

Thus, in the presented method, the removal of water from the evaporator is done with­out taking the liquid solution to the absorber, and the solution supplied to the rectifier is fully utilised. The system works as if the rectification were perfect.

In such an absorption system, the specific refrigerating effect is equal to that of a compressor system, and can be even higher if a subcooler is used (in which part of the solution flowing from the condenser to the evaporator is cooled by vapors leaving the evaporator).

To lead the solution intended for feeding the rectifier through the evaporator, its pressure has to be reduced from condensing to evaporator pressure and after­wards be raised back (on the evaporator outlet) to the condensing pressure. This way of feeding requires the use of a pump working between the vaporization and conden­sation pressure. Though this pressure difference is large, work input to the pump is relatively small ; for example, if the specific circulation of solution is 10 kg/kg6) and the amount of solution leaving the evaporator is 0.25 kg/kg, this work represents only 2.5 per cent of the work required to pump the solution from absorber to generator.

Since pumps are already used to supply the rectifier with the solution from the con­denser, the fact that an additional pump must be used is beyond reservation.

4 the amount referred to r kg of the solution vaporizing in the evaporator. 5 assumption is made, that vapors leaving the evaporator are free of water. 6 the specific circulation of solutions (marked "f") is understood as the amount of the solution

flowing from absorber to generator, referred to r kg of the refrigerant entering the evaporato1.

668

111-13

The presented method was checked partially in working conditions, and will be applied to a rearranged ice-making plant consisting of a quick-freezing ice generator (with direct evaporation of the refrigerant) and an ammonia absorption refrigeration system. The original layout of the plant is shown in Fig. 1 .

Condenser

Liquid receiver

I 1 __ _,..,,,,,--.,....., Subcoo/er

" ' <2)

Ice maker ! c ________ __ _j

Ltqwd separator

Fig. r . Layout of ice-making plant

I ts characteristics are as follows : 1. Icemaker :

Cold demand Qo = 165000 kcal/h by vaporization temperature to = -15°C. Charge GN = 1 150 kg. Periodical exchange of charges ; period between two succesive changes (cycle) T 48 hours.

2. Refrigeration plant : Cooling capacity intended for the ice production Qo 165 000 kcal/h by vaporization temperature to = -15°C.

550 550

,0 "' Z' -·- � 500 " � ' � � � � § a 450 "' -------,____ /; � � � � 400

5(JIJ

50

// JsQ-I/ I � � t J? /; I 300 � � � � � � � � �

250 � � _);I � "' _);I /, v I � "' " "' Rl "' I ijl lb I Lr)

200 I .; J i b' I � fl' . I ,fl ti J "' "'

�� // ISO

� I _,..- ""

�� 1� t� L---->--7 � � 50 ----� �:. rd ..... ,,_., �

350

3()()

250

200

150

100

\ 0

\-!a-i----r----r---\ \ -,,,

\ \ -MJO

-"'1

- 50

- 1()()

- l50 0 QIO 020 O.JO ll"" 0.50 QISO Ql!) aeo a"' UXJ � --

Fig. 2. i-� diagram for aqua-ammonia solutions

669

III-13 Condensation pressure Ps = 12 ata. Vaporization pressure at the beginning of the cycle pa = 2.4 ata and at the end of it Pa T = 1.9 ata. Final temperature of desorption td = 120°C. Final temperature of absorption ta = 40° C. Concentration of solution in the condenser .;. = 0.995 kg/kg. Gravitative feeding of the rectifier from the condenser. Specific refrigerating effect qo = 265 kcal/kg. Heat of desorption referred to 1 kg of the refrigerant q w = 560 kcal/kg. Coefficient of performance') ,; = 0.47 kcal/kcal.

The basic processes in the system are shown in the i- ,; diagram of aqua-ammonia solutions (after Bosnjakovic), Fig. 2.

At the beginning of the cycle the concentration of the solution in the icemaker is equal to that in the condenser. As a result of collection of water, the concentration of solution in the icemaker decreases with time, and after a period T hours it amounts to :

b =

Qo ,; sGN - - - T ( 1 - ,; s) qo

and particularly if the said period is 48 hours : 165 000

kg/kg

0.995 . 1 150 - - 245- . 48 (1 - 0.995) ,; 48 = ---------------- = 0.85 kg/kg

1 150

(4)

To keep the vaporization temperature constant, while simultaneously decreasing the concentration of solution, the vaporization pressure has to be reduced. After 48 hours, when the concentration reaches 0.85 kg/kg, vaporization pressure amounts to 1.9 ata.

It is understood, that by reducing evaporator pressure the efficiency decreases, there­fore assumed refrigeration effect of 165 000 kcal/h must be attained at the lowest expected vaporization pressure.

This pressure is to be taken into account when designing the plant, as is shown on diagram i- ,;, Fig. 2.

During the operation of the plant it appeared that changing the charge every 48 hours is troublesome and causes about 2 h break in the plant operation. Thus, it is obvious, that such a method of removal of water from the icemaker means that the expensive machinery cannot be wholly utilized and also excludes the possibility of full automatization.

This created the necessity of rebuilding the ice-making plant and a new method of removal of water from the evaporators has been worked out.

Fig. 3. New arrangement of the plant

the relation of the amount of heat rejected in the icemaker to the heat added in the generator .

670

III-13

The avoid irregularity in the concentration of solution and to secure equal vaporization temperature in the whole system, multiple circulation of the refrigerant with a circulating pump shall be adopted.

The new arrangement of the plant is shown in Fig. 3 and the essential changes in the processes are marked on diagram, Fig. 2.

As a result of rearrangement, following effects are achieved : 1 . The water from the icemaker will be removed continuously so that the plant will

operate without interruptions. 2. The concentration of solution in the icemaker remains constant. By equations (1),

(2) and (3).

�o = :: [i� - ( 1 + !�) · (1 - g s)] kg/kg

and after reduction

go = 1 - (:: + 1) · (1 - g s) kg/kg

from diagram Fig. 2.

assumed

then

q s = 295 kcal/kg, qR = 100 kcal/kg

g s = 0. 995 kg/kg

(295 ) go = 1 - !oo + 1 · c1 - o.995) = o.98 kg/kg

(5)

By such concentration of solution, it can be practically accepted that the icemaker is always filled with anhydrous ammonia and at the assumed vaporization temperature pressure would remain constant at 2.4 ata.

3. The specific refrigerating effect and the heat of desorption change and consequently coefficient of performance increases (i-g diagram Fig. 2).

The coefficient of performance of the existing plant, in the end of the cycle by qo = 265 kcal/kg and q w = 560 kcal/kg has a value

qo 265 � = - = --- = 0.47 kcal/kcal q w 560

and at the beginning of the cycle in the existing plant and in the rearranged plant always by qo = 285 kcal/kg, q w = 515 kcal/kg

qo 285 & = - = -- = 0.55 kcal/kcal q w 515

Assuming that by periodical exchange of the icemaker charge the rate of vaporization pressure drop is proportional o tthe working time of the plant, the mean heat utilization coefficient can be determined as the arithmetic mean of the coefficients at the beginning and the end of cycle. At the beginning of cycle the coefficient is equal to the one which shall be reached after the rearrangement of the plant, so the mean coefficient

0.47 + 0.55 & = --2 = 0.51 kcal/kcal

Absolute increase of the coefficient is then 0.04 kcal/h and relative about 8 per cent. It is easy to state, that if any of the known methods of continuous exchange of the eva­

porator (in this case an ice maker) charge is applied, the complete vaporization of the solution in the ice maker subcooler assembly would require keeping a constant concen-

671

IIl-13

tration of solution at a rate of about 0.85 kg/kg. The specific refrigerating effect would be then about 285 kcal/kg, but vaporization pressure should be kept constant at about 1. 9 ata.

In such working conditions of the plant, the coefficient of performance would amount to 0.51 kcal/kcal which is equal to the mean coefficient achieved in case of periodical exchange of ice maker filling.

672

111-12

A Method for Determining Performance Characteristics of Absorp­tion Refrigeration Systems

Une methode de determination des caracteristiques de fonctionnement des systemes frigorifiques a absorption

DR.-Ing. HARALD LOWER Maschinenlaboratorium, Technische Hochschule Karlsruhe, Germany

SOMMA/RE. On etudie graphiquement /'influence d'une modification des caractens­tiques de fonctionnement d'un OU de p/usieurs elements d'un systeme frigorijique a absorption sur le groupe tout entier avec des conditions prevues ·fixes des divers elements. La caracte­ristique principale du procede decrit est que le cycle de la solution est considere comme un compresseur thermique du groupe tout entier, comparable au compresseur d'un systeme a compression de vapeur.

On ejfectue une serie de determinations de la puissance specifique d'un groupe d'absorption utilisant de l' eau comme agent refrigerant et une solution aqueuse de bromure de lithium comme absorbant. Pour simplifier le procede, la temperature du generateur et l'ecoulement de la solution de l' absorbeur au generateur ont ete maintenus constants. Les variables considerees sont les temperatures de l'eau froide et de !'agent refrigerant. L'apport de chaleur au gene­rateur et la chaleur rejetee dans l' absorbeur et le condenseur sont determines pour dijferentes temperatures de condensation et d' evaporation. Les diagrammes finals de puissance du systeme s' appuient sur le f ait que, dans des conditions de regime stable, les puissances du com­presseur thermique et du serpentin de l'evaporateur sont les memes.

It is the usual practice to dimension the component parts of a compression as well as of an absorption refrigeration machine under the assumption of certain design condi­tions. However, deviations from these conditions occur very often in the operation due to changes in the temperature of the cooling medium or through temperature or output changes at the receiver end of the evaporator. Since the various components of a refriger­ation machine are connected together into a system, the performance of each compo­nent becomes dependent upon the conditions affecting the others.

Methods for the determination of the performance characteristics of compression refrigeration machines are already known. For absorption systems capacity calculations under different operating conditions are essentially more complicated, since the thermal compressor of an absorption machine consists of three pieces of apparatus - generator, absorber and heat exchanger - and in addition to flowing through the condenser, the cooling medium also flows through the absorber thereby effecting the compressor per­formance. The following graphs and procedures should be used as a model for deter­mining performance characteristics of absorption refrigeration systems. As an example the cycle employing the lithium bromide-water solution and using water as cooling me­dium is considered (Fig. 1). Cooling water and evaporation temperatures are introduced as variable operating conditions. To simplify the procedure, generator temperature, cooling water and chilled water flow have been held constant. Other simplifying assump­tions are :

1. The product of the overall heat transfer coefficient k and the heat transfer surface F is constant and results from the desip;n conditions.

2. Equilibrium between the strong solution and refrigerant vapor in the generator. 3. The solution flow S w from the absorber to the generator is constant due to constant

pump output.

673

III-12

7

3

' Steam l '"

Fig. r. Diagram of the Absorption Cycle

THERMAL COMPRESSOR AND EVAPORATOR

The method of determining the performance characteristics of absorption machines can be adapted from the method valid for compression systems by considering the solution circuit with generator, absorber and heat exchanger as one unit, the thermal compressor. Since in equilibrium the quantity of refrigerant R delivered by the generator equals the quantity absorbed by the solution in the absorber, the behavior of the genera­tor, heat exchanger and absorber is analyzed separately for different constant quantities of refrigerant. A diagram is drawn for the generator in which the characteristics deter­mined for absorber and heat exchanger can be entered. The intersections of these lines represent the points of operation of the thermal compressor.

Generator. On the basis of the external and internal behavior of the generator its ther­mal capacity QG can be written as follows :

QG = kG • PG (tH - t2) (1)

QG = R [h7 - h2 + f (h2 - hi)] (l a)

By comparison of these two equations and solution for the enthalpy hi, the following relation results :

kG . PG kG • PG • tH hi = (1 -R/S w) h2 -R/S w · h1 + -

S w _ __ . t2 - -----Sw-- (2)

Equation (2) allows presentation of the temperature ti of the weak solution upon entrance into the generator as a function of the temperature t2 of the strong solution leaving the generator for different generator pressures p (or condensing temperatures tc) and quantities of refrigerant R (Fig. 2). Connections between temperature, pressure, concentration and enthalpy of the solution are given by vapor-pressure and enthalpy/ concentration diagrams.

Heat exchanger. In the heat exchanger, the quantity of heat QHx is transferred be­tween the weak and strong solutions :

674

QHX = S w ' C w Cti - to)

QHx = S s • Cs (t2 - ta)

(3) (3a)

(3b)

SAT CONDENSING TEMP. =\, COOi.iNG WATER INLET TEMP. = tK1 AMOUNT OF REFRIGERANT "' R

60'-----'----'------'-----------� 70 75 80 85 90 95 100

SOLUTION TEMP. LEAVING GENERATOR t�t°Cl

Fig. 2. Solution Temperature Determination Chart

III-12

The combination of these equations leads to the calculation of one of the discharge tem­peratures t1 and r3 from the two inlet temperatures t2 and t6•

Absorber. Corresponding to the thermal efficiency of the generator the capacity of the absorber can be written as follows : (t4 + t5 tk2 + tk1) QA = kA . FA -2- - 2

QA = K . c (tk2 - tk1)

(4)

(4a)

By combination of equations (4) and (4a), the discharge temperature of the cooling water tk2 can be eliminated:

t4 + t, - 2 . tk1

2 1 kA · FA + K . c

From the heat balance for the thermal compressor:

(5)

(6)

Generator and absorber capacities Qa and QA from equations (1) and (5) are intro­duced into equation (6) :

t4 + t, - 2 tk1 ka • Fa (tH - t2) - = R (h1 - h10) 2 1 kA ° FA + K . c

(6a)

Considering the fact that t4 = t3 and ts = t6, the temperatures t4 and ts in equation (6a) can be replaced by t1 and t2 upon introduction of the relations (3) to (3b) valid for the heat exchanger. Since equation (6a) then contains only tu t2 and tk1 as variables, lines of constant cooling water inlet temperature t ki can be entered in the diagram Fig. 2. These isotherms intersect the lines of constant condensing temperature in one point, the operating point for the existing operation conditions (R, p, tk1) of the unit. The tem­perature t2 determines together with the pressure p the concentrations of the strong and weak solution. Using equations (3) to (3b) the temperatures of the solution entering and leaving the absorber, t4 and t5 respectively, can be calculated from the temperatures t1 and t2• An average solution temperature in the absorber tsm is thereby known, which,

675

III-12

together with the average concentration, determines an average vapor pressure of the solution, in this particular case considered to be about 1.5 mm Hg below the evaporation pressure p a.

'.2 100 "'-4�>1::"=.J"'-4"-4t r----r-= �-+----+---<

j 50 0 0 � Q f---+--��f--�-��-�-��--1 o' � 200C----,--.-------,,-------,---.----,�c=�r:::O'-r-c=I 13 � 150 >---+---L,o.

� 100 ����::$;::tJ 1'-±--F� 4=-----1-"<.--< !« !=" � � o r-�-�-�_,_-��f--�-��---1

12 11. 16 18 20 EVAPORATION TEMPERATURE t,. f°Cl

Fig. 3. Thermal Compressor-Evaporator Per· formance Chart

For different cooling water inlet temperatures tk1 diagrams are prepared, the thermal compressor performance (in terms of the quantity of refrigerant R or of the refrigeration capacity Q a) being plotted against the evaporation temperature t a at various saturated condensing temperatures (Fig. 3). Generator heat requirements Q G and the beat rejected in the absorber QA are shown in Fig. 4 and 5 respectively.

2001---.-�-.-------,----,---.--.---.-'="-= 4,o_• c-; --�

10 12 14 16 18 20 EVAPORATION TEMPERATURE t0C°Cl Fig. 4. Generator Performance Chart

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EVAPORATION TEMPERATURE t0[ 'Cl Fig. 5. Absorber Performance Chart

Evaporator. For constant evaporator dimensions, the evaporator capacity is dependent upon the evaporation temperature t o and the chilled water temperature t w :

Q O = w • C ( t W] - t w2) (7)

t w1 + t w2 Q o = kE • FE -- -- - t o)

2 (7a)

Elimination of the discharge temperature of the chilled water t w2 leads to the relation

Q o = kE . FE ------ (t w1 - t o)

kE • FE 1 + w . c

(7b)

With t w1 as parameter, equation (7b) results in straight lines, which are entered in the thermal compressor capacity chart (Fig. 3). The capacity of the combination thermal compressor/evaporator coil are determined by the intersection of both performance characteristics.

CONDENSER

The condenser capacity results from a calculation, carried out in the same way as for the evaporator:

Qc = K · c (tk1 - tk3)

tk4 + tks Qc = kc • Fe --

2-- - tc

kc • Fe --

k ---:-ji (tk3 - tc) l + _c __ c

K · c

(S)

(Sa)

(Sb)

The quantity of heat Qc given in equation (Sb) is the heat which can be rejected in the condenser as a result of its design. In Fig. 6 Qc is plotted as a function of the cooling water inlet temperature tks and the condensing temperature tc. For a certain tempera-

677

111-12

� wo�--��-��-��-��-�--� � � �0150 f------'d---�\11-----'t----\lc------\f<,. Cl !=: 0 if. � 100 f----+--'<-+--�-+--�--+----'c--;---'\---I "' UJ Vl z UJ 0 a so t----+----+---T---+--�--+-------u

- ABS./CONO. IN SERIES

--- ABS./CONO.IN PARALLEL

o�--+----+----�--�-----"----" 15 20 25 30 35 40 45 ENI WATER TEMP. t"t•cJ Fig. 6. rondenser Performance Chart

ture of the cooling water entering the unit tkp there is only one point on each ot Lhe�e curves in Fig. 6 for which the condenser capacity required by compressor and evaporator is the same as the capacity afforded by the condenser. These points are determined by calculating the required condenser capacity Qc from the heat balance for the entire unit :

Qa + Q o = QA + Qc (9) Q o, Qa and QA can be read from the diagrams Fig. 3 to 5 for the corresponding condens­ing and evaporation temperatures.

Normally absorber and condenser of an absorption machine are connected in series as far as the cooling water is concerned, in which case the cooling water consumption is about as great as for a compression machine of equal capacity. In this case tk2 is equal to tks· The temperaure of the cooling water leaving the absorber tk2 can be read in Fig. 5 on the righthand ordinate simultaneously with the absorber capacity. These temperatures give the abscissa values of those points in the condenser capacity chart for which the capacity required by compressor and evaporator is equal to that afforded by the condenser.

For parallel cooling water connection of absorber and condenser, the inlet temperature in the condenser tks is independent of the absorber capacity, where tk3 is equal to tk1• The condenser capacity can then be read on a vertical line through the abscissa value tk1•

SYSTEM PERFORMANCE The performance of the entire unit is presented in a diagram as a function ofthe cool­

ing water inlet temperature h1 and the temperature t w1 of the chilled water entering the evaporator (Fig. 7). For this presentation, in which series and parallel connection of the cooling water is distinguished, the heat balance equation (9) has to be fulfilled. The cycle coefficient of performance (COP) of the absorption machine for different external operating conditions (tk1 and t w1) is plotted in Fig. 8.

- - - ABS./COND. IN PARALLEL

O'----+----+-----'----'---�--� 15 20 30 40 45 35 25 Fig. 7. System Performance Chart COOLING WATER INLET TEMP. t.,t°Cl

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1,0 I I I I I I COP � ...._ ABS./COND.IN SERIES ABS./CONO. IN PARALLEL --

-� ---!!..._ '• "'\ """' ����-I---. � 9'!1tl:§n -r--- '5 I\ �

�� �J: §!>,!>.. '..l... 19.,, "'t..� --... I �· ...

0,7

0,6 20 30 20 30 COOLING WATER INLET TEMP. '.c,[°CJ

Fig. 8. Coefficient of Performance

The above procedure should be used as a model for determining the performance characteristics of a particular system. As an illustration an absorption unit has been used having the nominal capacity of 100 000 kcal/h under the following conditions : t w1 = 10°C, tk1 = 27° C, absorber and condenser in series. Since these calculations are based on a particular set of components, any generalization of the data should be done with caution.

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New Pumping Method in Absorption Refrigeration

Une nouvelle methode de pompage dans le refroidissement par absorption

Dipl.-Ing. LASZLO SZUCS Department of Energetics, Polytechnical University, Budapest, Hungary

SOMMAIRE. Methode de refoulement des fluides, d'abord destinee au refroidissement par absorption et presentant les avantages suivants : aucune piece mobile susceptible d'usure, pas de besoin d' entretien, f onctionnement possible egalement afJec de l' energie purement ther­mique.

La methode s' appuie sur le f ait que le degagement de chaleur a partir d'un corps chauff ant a une temperature donnee est plus eleve dans unfluide que dans de la vapeur.

L' ecoulement de jluide - pratiquement a sa temperature initiate ne presente que des inter­ruption terisdiques tres breves si bien qu'il est approprie a un fonctionnement continu.

Le fonctionnement du cycle est le suivant : le fluide s' ecoule a travers un clapet de retenue, se rassemble dans le reservoir, s'ecoule dans un bac d'alimentation et passe dans un serpentin chauffe exterieurement. La vapeur a l'interieur du serpentin provoque une elevation de la pression du bac d' alimentation; cette elevation de la pression fait passer ce fluide par l'inter­mediaire d'un autre clap et de retenue dans un bac a haute pression ( seule la surf ace du fluide prendra la temperature la plus elevee correspondant a la pression la plus elevee) .

Quand le fluide aura ete evacue du bac d'alimentation, la surface de chauffage elle-meme disparaitra et la formation de vapeur cessera.

La haute pression s' ecoulera finalement de bac d' alimentation dans le recipient a basse pression a travers un etranglement de dimensions appropriees, terminant ainsi le circuit.

On indique dans le rapport les methodes de calcul des dimensions de l'etranglement et de production du rendement optimal.

It is common knowledge that the pumps of absorption refrigerators, due to the in­evitable moving components, set serious limitations to the application of such equipment. Pumps not only affect safety of operation and service life but their electric drive makes the refrigerators, which would otherwise require thermal energy only, dependent on the availability of an electric network.

A safely operating high-efficiency pump which does not require manipulation and has no moving components, would go a long way towards expanding the field of ab­sorption-type refrigerating machines, primarily in heat pumps or in high-output house­hold refrigerators.

From the foregoing, the significance of a pump without moving parts, suitable for use in absorption-type equipment and operating exclusively with thermal power, will be evident.

The present paper lays down a system* for fluid transfer which a) if appropriately dimensioned, ensures good efficiency in any arbitrary size (large

or small) of the equipment; b) ensures complete sealing of the absorption equipment; c) is practically free from all kinds of moving components and consequently has practic­

ally unlimited service life; d) requires only thermal energy for operation and does not restrict the application

of absorption refrigeration to locations where electric power is available; e) does not set any restrictions whatsoever to the layout of the absorption equipment.

Any optional condenser or evaporator type etc. may be used in it, in any optional arrange­ment and any optional media pair (refrigerant - absorbent) may be applied.

* Patent applied for : L. Sziics : Method for Thermic Fluid Transfer Without Moving Components.

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The pumping process elaborated is based upon the fundamental physical phenomenon that the heat output of a radiator when evaporating fluid exceeds by 2 to 3 orders the output it yields when superheating steam.

This phenomenon can be utilized in the following way : A radiator is located in a space and heated by the same medium that heats the generator

of the absorption machine, and is flooded by the rich solution coming from the absorber. The radiator heats the rich liquid whereupon the arising vapour increases the pressure within the space and precludes more of the rich solution to enter. When the pressure has attained the pressure level of the generator, it passes the rich solution into this latter.

When the fluid has discharged from the space, high-pressure vapour through an appropriately dimensioned hole, will also blow down to some part of the absorption machine where low pressure prevails. Pressure having decreased, the fluid collected in the container of the absorber will once more flood the radiator and the process begins all over again.

With appropriate dimensions, the periodicity of the fluid transfer will not disturb the continuous and smooth operation of the equipment.

The technical execution of the above outlined theoretical system is as follows. The radiator is developed in the form of a natural-circulation single-tube boiler (see

K in Fig. I). The boiler is heated by the heating medium of the absorption equip­ment. It shall be so dimensioned that the maximum pressure attained shall be in excess of the pressure in the generator.

J

delinr.v resset

4

Fig. r . Experimental set-up

The liquid to the boiler is supplied from the delivery vessel via a throttle (1). The throttle must be so dimensioned that the heating surface should not receive fluid and no vapour should be produced until the delivery vessel has been filled up. In this manner the supply of the rich solution from the container will not stop until the vessel is fuU. The proper dimensioning of the (1) throttle will further ensure that the fluid remains in the boiler and continues to produce steam until the delivery vessel has been completely evacuated.

If the delivery vessel is correctly dimensioned, it will just fill up with the rich liquid collected in the container in any operating period. Should the vessel be of too small capacity, the pump will not be capable of passing the required quantity of fluid, while a too large vessel will cause higher than necessary losses arising from the alternating heating-up and cooling-down of the vessel.

Steam from the boiler will reach the delivery vessel through the separator (2). The (3) and (4) check valves guide the fluid on its path.

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IIl-25

Completing the fluid transfer cycle, the boiler discontinues to generate steam. Through the appropriately dimensioned throttle (5), high-pressure steam will blow down within a predetermined time and the pressure will adequately decrease. When dimensioning the (5) throttle it must be borne in mind that while steam is being generated in the boiler, the throttle checks the rate of pressure increase.

The trottle (5) will evidently cause losses and could be easily dispensed with con­structionally. But the elimination of eventual losses will hardly make up for the consi­derably more involved control.

Let us now examine the fluid transfer process from the energetics point of view. Since a throughgoing investigation would go beyond the scope of the present short description, I shall restrict myself to outline the factors which have some bearing on the efficiency of the equipment.

It is obvious that the thermal power input covers the heating of the rich liquid, the difference between entering and leaving work, and the heat needed to produce the steam that exits through the throttle.

In the most simple arrangement the steam passing through the throttle is a dead loss. but experiments hold out good promise for its utilization.

The difference between leaving and entering work is, by and large, equal to the theoretical power consumption of the fluid transfer.

In optimum cases the thermal energy input is spem to cover this theoretical power consumption (work) and to produce the leaving steam. Due to certain irreversibilities (for instance the periodical heating-up and cooling-down of the vessel, etc.), fluid is not only transferred, but heated too.

Should no rich - weak solution heat exchanger be applied, the thermal energy spent in the heating up of the fluid cannot be considered lost. In such cases namely the rich solution must be heated up by valuable heating medium in any case, and so whatever heat is spent on heating up the rich solution while it is in the pump, may be saved while the solution is in the boiler.

The situation will be completely different if a heat exchanger were applied. In such case the matter would be fairly similar to when part of the rich solution (a quantity approximately equal to the weak liquid) was heated up by the weak liquor - practically without any energy input - and only the rest by the heating medium. In such process that part of the heat which had in the pump heated the latter part of the rich solution, is usefully spent while the first-mentioned part lost.

It is clear from the above that heating up the transferred fluid at a possible minimum rate will, in the majority of cases, improve the efficiency of the equipment.

For this purpose : the delivery vessel must be so developed as to absorb a possible minimum of heat;

and the mixing of the hot fluid film on the surface with the bulk of the cold fluid, must

be prevented; etc. In consideration of the above points and on the basis of the given principle, it is pos­

sible to design for the purposes of absorption refrigeration and heat pumps an absolutely safe operating pump, operated by solely thermal energy at a comparatively good ef­ficiency, having a practically unlimited service life and not requiring any manipulation or maintenance whatsoever.

683

Refrig erants, Automation Flu ides frigorigenes, automation

III-3

Material Stabilities in Vapor Compression Refrigeration Systems

Stabilisation des materiaux dans les systemes frigorifiques a compression de vapeur

H. 0. SPAUSCHUS, G. C. DODERER, R. S. OLSEN and R. A. SELLERS Major Appliance Laboratories, General Electric Company, Louisville, Kentucky U.S.A.

SOMMAIRE. La duree des systemes frigorifiques a compression de vapeur est determinee par /es combinaisons de materiaux qui sont choisies ainsi que par la temperature de fonctionne­ment. Les elements a surveiller sont le frigorigene, l'huile, !es adjuvants de l'huile, !es isola­tions de moteurs et !es agents deshydratants. La stabilisation des combinaisons de materiaux est tres souvent inferieure a celle des elements individuels. Les reactions specifiques du R 12 avec !es huiles de petrole, du R 22 avec la cellulose et du R 22 avec le tamis moleculaire 4 A peuvent etre citees en exemple.

La vitesse des reactions chimiques dans !es systemes frigorifiques etant tr es f aible, ii est necessaire d'employer des methodes analytiques extremement sensibles pour deceler et mesurer !es quantites de produits de decomposition formes. Notre laboratoire a ete l'un des premiers a utiliser !es techniques d' analyse des gaz pour l' etude de ces reactions en laboratoire ainsi que dans /es systemes en fonctionnement. On etudie plusieurs reactions specijiques pour demontrer l'utilite de cette etude.

On presente en conclusion une comparaison des limites chimiques dans !es systemes utilisant du R 12 et du R 22. On demontre que la reactionfrigorene-huile a une importante primordiale dans !es systemes utilisant du R 12, tandis que pour les systemes a R 22 la stabilite des isolants classiques de moteurs est le f acteur limitatif.

INTRODUCTION Materials used in modern vapor compression refrigeration systems may be degraded

by two distinct mechanisms : chemical reaction

A + B catalyst C + D (g)

thermal decomposition A catalyst C + D (g)

In many of these reactions a gaseous decomposition product, D (g), is formed. Methods have been developed for separating, identifying and measuring these gaseous products as a means of monitoring rates of material degradation [1].

Since refrigeration systems are designed for many years of maintenance free operation, very stable materials and material combinations must be used. Conversely, very sensitive analytical methods are required to study rates of degradation of these materials, if reaso­nable test temperatures are to be maintained. As an illustration, it is assumed that a degradation reaction produces a liquid nitrogen non-condensable gas and that refrigera­tion capacity of the system is noticeably affected when 100 standard cubic centimeters of this gas have been produced. If the gas formation rate doubles for each 10° C temperature rise and a one week test period is desired, the requirements of an analytical method appli­cable to this problem are summarized in Table 1. Testing at the normal usage temperature (T) of the material demands a method capable of quantitatively determining 0.1 standard

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IIl-3

cc. of the gaseous degradation product. If the test temperature can be accelerated as much as l00°C, with assurance that the mechanism of degradation is not altered, over 100 standard cc. of gas will have formed. Although this is a relatively large amount of gas, it represents less than 0.05% of the total gas in the system assuming a refrigerant charge of 10 moles.

Table I. Analytical Requirements for detecting decomposition

Test Temperature

(OC)

T T + 40 T + 70 T + 100

Quantity of Gaseous Product Formed

(St. cc.)

0.1 1.6

12.8 102.4

* Assuming refrigerant charge of 10 moles.

Mole Percent* Of Refrigerant

0.0004 .0006 .005 .04

Gas analysis methods of impressive versatility and capable of satisfying the rigorous requirements outlined above are available. In our Laboratory these methods have been used extensively for analysis of the contents of sealed glass tubes [2], metal bombs with provisions for refluxing the refrigerant [l], "motorette" type devices [3], and operating refrigeration systems [l].

The effect of temperature on material degradation rates is important in predicting time to failure. Two characteristic effects are illustrated in Fig. 1 . Many material degrada­tions obey the general rule that the reaction rate is doubled for each 10° C rise in tempera­ture (Example A, Fig. 1 .). In other instances (Example B) degradation proceeds at a

z 0 ;: <( 0 <( a: (!) "' 0

E X A M P L E A

T,

r, T,

T I M E

E X A M P L E B

Fig. r. Examples of degradation rates

Ti T,

negligible rate for a long period of time - then accelerates suddenly to catastrophic dimensions. Reactions of this type are difficult to identify and the reaction kinetics are usually more complex than those illustrated in Example A.

EXAMPLES OF TYPICAL DEGRADATION REACTIONS Table 2 illustrates a variety of degradation processes which have been studied. The

gaseous products produced by each reaction are arranged in order of decreasing amounts (left to right). Temperature and duration of the experiment denote the conditions under which the gaseous products are produced. Since experimental conditions vary substan-

686

IIl-3 tially, extensive stability comparisons between various materials or material combinations are precluded. The data in Table 2 reveal that : 1) Material degradations or reactions generally produce characteristic gaseous products, 2) the nature of the products formed is determined, to a large extend, by the environment ; i. e. metals and atmosphere, 3) orga­nic insulation materials usually produce CO and H2, which may be used to monitor the rate of degradation, 4) refrigerant reactions are generally one of two types, substitution or disproportionation, and the halogenated products which are formed provide a means of measuring the reaction rate.

Table 2. Gaseous products formed by degradation of refrigeration materials

Type Refe-Reactant Materials Gaseous Products of Test Conditions rence

Test1

CC12F2 CF4, SiF42, CO G. T. ya 4 CCl2F2 (Fe, Fe203) CF4, C02 M. T. v 4 CHC1F2 CF 4, SiF 4, CO, HCl, H20 G. T. v 4 CHC1F2 (Fe, Fe203) CF4, C02 M. T. v 4

CCl2F2 + Oil (Cu, Fe) CHClF2, C02, H2, CH4 G. T. 150°c, 14 Days 2 CHC12 + Oil (Cu, Fe) CHF3,CH2F2, CO, H2, C02, G. T. 175 ° C, 100 Days 5

CH4 CHClF 2 + Oil (Al) CHF3, H2, CH4, CO, HCl G. T. 175 ° C, 100 Days 5

CC12F 2 + Molecular Sieve 4A co M. B. 70° C, 30 Days 5,6 CHC1F2 + Molecular Sieve 4A co M. B. 70° C, 30 Days 5,6

Cellulose C02, H20, CO, H2, CH4 G. T. 175° C, 8 Days 5 Polyvinyl formal C02, H20, CO, H2, CH4 G. T. 175 ° C, 8 Days 5 Aromatic polyimide C02, H2, H20 G. T. 175 ° C, 8 Days 5 Aromatic polyamide C02, H2, H20 G. T. 175 ° C, 8 Days 5 Polyacrylic C02, CO, H2, CH4, NH3, G. T. 175 ° C, 8 Days 5

H2COHCN

Cellulose + CHClF 2 HCl, SiF4, HC02H, G. T. 150°c, 4 Days 5 CH3C02H

Aromatic Polyamide + CHC1F2 CO, C02, H2 G. T. 175 ° C, 8 Days 5 Aromatic Polyimide + CHC1F2 CO, C02, H2 G. T. 175 ° C, 8 Days 5 Aromatic Polyamide + CC12F2 C02 G. T. 175°C, 8 Days 5

Polyacrylic-polyethylene CO, H2, C02, CH4, terephthalate H2COHCN c. s. 120°C, 20 Days 5 Polyvinyl formal-cellulose CO, H2, C02, CH4 C. S. 104°C, 20 Days 5

1 Type of Test : G. T. - Glass tubes, M. T. - Metal tubes, M. B. - Metal bomb with refluxing, C. S. - Complete Refrigeration system.

2 Silicon tetrafluoride, found in some glass tube experiments, would not be produced in a refrigeration system.

3 Test condition V refers to measurements over a wide temperature range.

Several examples of the application of gas analysis in the study of refrigeration material degration will be cited.

687

III-3

It has been known for some time that molecular sieve Type 4A (synthetic zeolite of sodium alumino-silicate type) reacts with CH ClF 2, [ 6, 7, 8] resulting in the formation of halogens, which are retained by the drier, and carbon monoxide. Previous workers have studied the rate of this reaction by analysis for chloride ion. At early stages in the reaction it is difficult to obtain accurate rate data by this method since it is necessary to determine very small quantities of chloride. For example, in the work reported by Mays, [8] 0.3 mg. of chloride corresponded to 0.01 % R 22 decomposition, which was the lower limit of chlori­de values reported in his work.

In our Laboratory, the CHC1F2-molecular sieve 4A reaction has been studied by measuring the rate of formation of CO. Fifteen grams of the drier were transferred into a stainless steel bomb of the type previously described [1], the bomb was evacuated and 300 g. of CHClF 2 were condensed into the bomb. The lower part of the bomb was immersed in a bath maintained at constant temperature (59 or 69° C). As the reaction proceeded, samples of gas were removed periodically from the long stem where the CO collected. These periodic analyses were conducted without disturbing the experiment since very small quantities of CHClF 2 were removed with each sampling. Typical results, obtained at two temperatures, are shown in Fig. 2. Extensive studies have shown that reproducibility for duplicate experiments is extremely high and that as little as 0.001 % CHClF 2 decomposition is readily detected.

1 2

.. .,. 8 .,; ,.

� ' 0 (.) 4 u u

1 0 T I M E ( D A Y S )

1 0 0

6 9 " C

5 9 ' C

Fig. 2 . Reaction of molecular sieve 4 A with CHC!F2

1000

The great sensitivity and accuracy of the gas analysis method, as described in this example, makes it feasible to study many variables with a minimum expenditure of time and without accelerating temperatures beyond the range encountered in actual service in refrigeration systems. The method has proven particularly useful for studying modi­fied synthetic zeolites, such as molecular sieve - 4A XH [6], which react much more slowly with CHC1F2•

Gas analysis is also an excellent method for detecting chemical reactions between insulation materials and refrigeration system environments. Considerable information concerning new insulation materials can often be obtained by use of the relatively simple ,,sealed-tube" tests [9], complemented by gas analysis. Upon completion of high tempera­ture aging under controlled environments in the tube, the insulation material is inspected for visual signs of degradation. The tube is then opened by a technique which permits ana­lysis of the gaseous products [2] . Finally, the material under study is removed from the tube and subjected to physical property measurements for comparison with properties of the material before test. Experience has shown that gas analysis often uncovers chemical reactivity or degradation before it proceeds to a degree where these effects are obvious by

688

111-3

visual examination or changes in properties. Furthermore, a knowledge of reaction products often provides a clue to the mechanism of degradation and, as a corollary, suggests formulation changes to enhance stability.

Studies of an aromatic polyamide fiber sheet material, a reaction product of isophthalic acid and m-phenylenediamine*, serves as an example. Initial screening tests were conducted with material vacuum dried ( < 0.5% water) and not dried (5% water). Samples of insulation were heated in three environments ; vacuum, CC12F and CHClF 2 and the tube contents analyzed after heating. Analytical results for CO are summarized in Table 3. No other gases were found in substantial quantities.

Table 3. CO formed in sealed Tube Tests of Aromatic Polyamide (8 Days, 175°C)

Condition of Material

Vac. Dried Not Dried

Standard cc. of CO Formed In Vacuum CC12F2 CHC1F2

< .l < .l

< .l < .l

3.9 4.3

These studies revealed that the aromatic polyamide reacts with, or decomposes in the presence of, CHC1F2, and that CO is produced as a gaseous degradation product.

A second series of experiments was conducted to determine the rate of CO formation. The results, shown in Fig. 3, indicate only a slight decrease in the rate of CO pro­duction with time. Calculations reveal that the amount of CO produced from the quan­tity of insulation required in a hermetic motor would be excessive if this material were contemplated for CHC1F2 systems operated at high temperature.

,.. <( 0 ' E .,. ' 0

1 .5

1 .0

� 0.5 u

"' ,_ <( a:

0 5 1 0 T I M E ( D A Y S )

1 5

Fig. 3. Reaction of aromatic polyamide with CHCIF,

2 0

The most likely source of CO would appear to be the carbonyl and/or terminal carboxyl groups since qualitative tests indicate no free halogens, thus eliminating the remote possibility of transforming CHC1F2 to CO. The role of CHC1F2 is visualized as that of a Lewis acid which catalyzes the decomposition of the polymer at high temperatures. Assuming this mechanism and a molecular weight of 20,000 for the polymer, simple calculations reveal that elimination of CO by chain scission at amide linkages is not

* Material produced by E. I. du Pont de Nemours and Company, Inc. and designated as HT·l . 689

III-3

likely. The quantity of CO formed would dictate substantial reduction of molecular weight accompanied by gross changes in physical properties. On the contrary, it has been observed that the aromatic polyamide, if initially well dried, exhibits little loss of physi­cal properties after heating in CHClF 2•

A second possibility is that CO stems from attack of carboxyl end groups. Calculations again negate this consideration since there is little change in the rate of CO generation even after 20 days. Further studies are required to firmly establish the source of the CO and to relate the effect to properties of the polymer.

The final example is selected from sealed tube studies of refrigerant-oil reactions. Long term aging tests at 175 ° C were conducted with the following combinations of materials :

A

CHC1F2 (0.43 g, containing 4 ppm water) paraffinic oil (0.45 g) aluminium, type 380 (0.2 g)

B

CHClF 2 (0.43 g, containing 63 ppm water) paraffinic oil (0.45 g) aluminium, type 380 (0.2 g)

The tubes were inspected at two-week intervals for visual signs of reaction. After 16 weeks of heating there was no change in appearance of the contents of the tubes but in the next two-week interval the liquid in combination A was transformed to a black sludge and the aluminium surface was covered with a black deposit. The contents of the tube with combination B remained unchanged for another month at which time the test was terminated and the gases in the tubes were analyzed. The results are given in Table 4.

Table 4. Reaction Products of Aluminium and CHC1F2 ; Gases Formed (Standard cc.)

Materials Combination A Combination B

H2 9

.004

CH4 2.1 .007

CHF3 66.2 .035

HCl 0.2

.0001

For material combination A only a small amount (1.4 cc.) of unreacted CHClF 2 remain­ed in the tube. The principle product, CHF3, was found to be present in about 90 % yield, based on the available fluorine. For combination B, which differed from A only in that a higher moisture level was present, very little chemical reaction had occurred.

These observations can be explained as follows : Prolonged heating of CHClF 2 in the presence of oil and aluminium produces trace amounts of free halogen which react with the aluminium surface. The A1Cl3 (or, less likely, AlF 3), if anhydrous, is an effective agent for promoting disproportionation reactions. Murray [10] has shown that CHC1F2 can be converted, in high yields, to CHF3 by A1Cl3• The equation proposed by Murray to account for the reaction is :

AICl3 3 CHC1F2 --� CHC13 + 2 CHF3

The reaction observed in the tube tests was under different conditions than those described by Murray and this may explain why no CHC13 was produced in the tube tests.

This reaction is clearly representative of Example B, Fig. 1, where the reaction rate attains catastrophic proportions after a long induction period. These results emphasize the hazards associated with the use of aluminium for components exposed to halogenated methanes at high temperature ; particularly under conditions of extreme dryness.

CONCLUSIONS : The examples which have been cited represent only a small part of a large number of

material combinations which have been studied for chemical stability. Since CC12F2 and CHC1F2 are by far the most widely used refrigerants for home refrigeration and air conditioning service, it is of interest to compare and contrast the chemical problems most frequently encountered with these two refrigerants.

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Chemical instability in system using CC12F2 generally results from splitting of a C-Cl bond in the refrigerant. CC12F 2 will react with petroleum oils at elevated temperatu­res (150-175°C) and among the ultimate products of this reaction are high carbon residues (tar, coke, etc.) and HCl [2] . The acid, in turn, is indirectly responsible for copper transfer [1 1], metal corrosion, etc.

CHC1F2 is a much less effective halogen donor than CC12F2• Thus reactions with petroleum oils do not commence until very high temperatures (225-250° C) are attained. Chemical problems in systems using CHCIF 2 can generally be related to the ,,positive" hydrogen, which accounts for enhanced solvent properties and susceptibility to hydroly­sis by alkaline media. In refrigeration systems these effects manifest themselves as extrac­tion problems (preferential solution of insulation material components) and insulation degradation (for example, the chemical degradation of cellulose [12]).

REFERENCES r. H. 0. Spauschus and R. S. Olsen, Refrig. Eng., 67, No. 2, 25 (I959). 2. H. 0. Spauschus and G. C. Doderer, ASHRAE J. 3, No. 2, 65 (I96I). 3 . For example, motorette test devised by Working Group For Hermetic Motor Insulation,

Insulation Subcommittee, A. I. E. E. Rotating Machine Committee. 4 F. ]. Norton, Refrig. Eng., 65, No. 9, 33 (I957). 5 . Experiments conducted by the authors at the Major Appliance Laboratories, General Electric

Company, Louisville, Kentucky. 6. R. L. Mays, ASI-IRAE J., 4, No. 8, 73 (I962). 7. P. Cannon, J. Phys. Chem., 63, I6o (I959). 8. H. Steinle, Kaltetechnik, I3, No. 4, I50 (I96I). 9. For example : H. M. Elsey, L. C. Flowers and J. B. Kelly, Refrig. Eng. 60, No. 7, 737 (I952) ;

D. E. Kvalnes and H. M. Parmelee, Refrig. Eng., 65, No. r r , 40 (I957) ; W. 0. Walker, S. Rosen and S. L. Levy, ASHRAE J., 4, No. 8, 59 (I962).

IO. W. S. Murray, U. S. Patent 2, 326, 638. I r . H. 0. Spauschus, in press. 1 2 . H. M. Elsey and L. C. Flowers, Refrig. Eng. 64, No. 4, 3I (1956).

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IIl-29

Utilization of Refrigerant Mixtures in Refrigerating Compression Machines

L'emploi des melanges de fluides frigorigenes dans les machines frigorifiques a compression

V. F. CHAIKOVSKY and A. P. KUZNETSOV Odessa Technological Institute of the Food and Refrigerating Industry, Odessa, u. s. s. R.

SOMMAIRE. Des melanges non-azeotropes des jluides frigorigenes peuvent etre employes dans certains cas comme substances de travail.

Les A. A. ant etudie deux problemes lies a l'emploi de tels melanges:

1. l' accroissement de puissance d'une installation fonctionnant suivant le schema ordinaire a un seul jluide frigorigene;

2. l'obtention de basses temperatures dans le schema particulier de !'installation au degre optimal de compression dans le compresseur a un seul etage.

Les recherches experimental es concernant le probleme n° 1 et f aites avec des melanges differents des fiuides frigorigenes ant montre qu'on peut augmenter le rendement de l'installa­tion en choisissant des melanges correspondants. Ainsi, le rendement d'une installation fonctionnant a la temperature d'evaporation -20°C et de condensation +30°C et avec le melange de F-22 (30% au poids total) / F-12, au lieu de F-12, augmente de 33% . Avec ces conditions, les rapports de pressions dans le compresseur ne dependent pas pratiquement de la composition du melange, et la puissance frigorifique specifique (kcal/kwh) augmente suivant !'augmentation de teneur en poids du F-22 dans le melange.

Les recherches concernant le probleme n° 2 ant ete effectuees sur !'installation dont le schema a ete propose par les A. A.

Comme substance de travail on emploie un melange non-azeotrope de jluides frigorigenes dont l'un avait un point d'ebullition eleve (F-12) et l'autre avait un point d'ebullition peu eleve (F-13).

Le composant a point d' ebullition peu eleve assurant l' effet frigorifique dans l' evaporateur, etait condense a !'aide de la chaleur d'evaporation du composant a point d'ebullition eleve condense dans le condenseur a eau.

Cette installation ayant des coefficients energetiques satisf aisants, etait caracterisee par la simplicite de construction par rapport aux installations a cascade et a plusieurs etages. On peut obtenir des basses temperatures d'evaporation dans cette installation a un seul etage (par exemple, -65°, -80°C, teneur en poids du F-13 dans lemelange egale aux 35%) sans vide dans l'evaporateur.

Many engineers are highly interested at present in the problems of utilizing mixtures of different refrigerants as working substances for refrigerating machines.

Azeotropic mixtures have already been applied in practice. Considerably less research has been carried out on investigating non-azeotropic mixtures and the fields of their most effective utilization.

Non-azeotropic mixtures are characterized by the following main advantages : 1) the list of working substances extends, thereby increasing the possibilities for unifying refrigerating equipment; 2) the power characteristics of the machine improve when effecting non-isothermal processes of heat transfer; 3) low temperature cycles are effected with single-stage compression and no vacuum in the evaporator; 4) control of the composition of the working substance provides for achieving optimum working conditions for a singlestage compressor in a wide range of refrigerating capacities and temperatures in the evaporator.

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Theoretical and experimental investigations of the thermodynamic properties and operating peculiarities of some non-azeotropic mixtures of refrigerants have been carried out at this Institute during the recent years. The possibilities have been studied of intensifying a plant, operating according to the ordinary circuit, and of producing low temperatures in a single-stage compressor.

An experimental stand has been designed for these investigations on the base of a condensing unit with a displacement of 12.1 m3/h at n = 560 r. p. m.

The diagram of this stand, which provides for an ordinary or low temperature cycle when switchindthe respective valves, is given in Fig. 1 .

6

Fig. r. Diagram of experimental stand for ordinary or low temperature cycles r - compressor, 2 - water condenser, 3 - first fraction receiver, 4 - evaporator-condenser, 5, i - expansion valves, 6 - second fraction receiver, 8 - electric calorimeter, 9 - rege­nerative heat exchanger, ro - stop valve, a, b, c - valves, A - ammeter, V - voltmeter, W - wattmeter, CH - voltage stabilizer.

The tests have been effected with closed valves "a" and "b" (valve "c" - open) when operating according to the ordinary cycle.

The test programme included preliminary investigations of such a plant with pure refrigerants freon-12 and freon-22 and subsequent comparative investigations with different non-azeotropic weight concentrations of freon-22 in the mixture F 12 - F 22 and of freon-13 in F 12 - F 13. The tests have been effected with evaporating tempera­tures ranging from -15 to -35°C (measured directly after the expansion valve). The temperature of the leaving condensate has been controlled by altering the water consump­tion and was maintained at +30°C.

Experimental investigations have indicated that it is possible with refrigerant mix­tures to increase essentially the refrigerating capacity of the plant as compared with its capacity when operating with freon-12 at the same temperatures, or to lower the tem­perature in the evaporator with the same refrigerating capacity (Fig. 2).

The evaluation of the experimental data has demonstrated that the volumetric and power characteristics of the compressor improve at an increase of the concentration of the low-boiling component.

Fig. 3 illustrates that the discharge and suction pressure difference increases with an increase in the concentration of the low-boiling component in the mixture. The ratio of these pressures for the mixture of freon-12 with freon-22 remains practically constant for any concentration, while for the mixture of freon-12 with freon-13 with the investi­gated concentrations (up to 35%) it increases quite negligibly.

694

Fig. 2. Results of investigating different refrigerant mixtures. Q0 - refrigerating capacity of unit, 10 - evaporating temperature

1 Pt1/Ps / .\ N'?J '-.. ��� � I � ' 101----1----+-�+---+--+-----+----!-\-+---I-----,·"�;� LI \ , L_-� "' ..;t--.._ I �...._ � 9 ·--+�-.j\-.�+-�f.---+�-P��r��A,---�

+---t..:::,.,...,.+���6 �:,, (�"' , _ B1-----+-----i--+--,��_._ (orft:ffrJc,Xt�,--cJ"�--+---1 6 ""' K ���---1 s l---l----1--+---+--+-""----f'>..---+----t-----t--P.-r\ � -

i r-�� 4 1-�l-�1-�l--4�--l�--l�-+�-+-�-+�-+�-t-��-I I II I I I ; �7'LV---6�5L_ ___ 6

�0----5

�5----5�0----45L--_4_0L--_-3�5----3i0 ____ 2�5----20'--_1�5----'10

t0, •c

111-29

Fig. 3. Changes in discharge and suction pressures with increase ill low-boiling component con­centration Pd - discharge pressure, Ps - suction pressure, t0 - evaporating temperature

Investigations for the second problem (low temperature circuit) have been carried out at the same stand (valve "c" - closed, valves "a" and "b" - open).

Let us analyse the operation of the plant according to the low temperature cycle with a mixture of freon-12 and freon-13.

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The vapour of the refrigerant mixture is compressed in the compressor (1) and dis­charged into the water condenser (2), where the high-boiling component (freon-12) is condensed at its partial pressure and drained into the first fraction receiver (3).

The vapour, rich with the low-boiling component (freon-13), is evacuated from the upper part of the condenser to the evaporator-condenser (4), where it is condensed owing to the boiling of the first fraction, entering the evaporator-condenser via the expansion valve (5).

The second fraction (which is rich with the low-boiling component) is accumulated in its receiver (6) from which it is directed via the expansion valve (7) into the evaporator of the electric calorimeter (8).

A regenerative heat exchanger (9) is provided in the circuit for the low temperature fraction to increase the refrigerating capacity and reduce the irreversible losses when the vapours mix on the compressor low side.

Design data, obtained by means of the enthalpy-concentration diagram, indicate that the weight concentration of freon-13 in the working mixture should be within the range of 30 to 45 %, in this case it is possible to obtain the minimum boiling temperature in the evaporator ranging from -65 to -75°C without any vacuum in the evaporator and at the permissible degree of compression in the compressor.

Fig. 2 presents the results of investigating at a concentration of 32% of freon-13 in the mixture with a minimum boiling temperature range in the evaporator from -40 to -60°C. The temperature was maintained at the compressor low side at +5 to +7°C, while the minimum condensing temperature of the high-boiling fraction was +30°C.

At the minimum boiling temperature of -60°C and condensing temperature of +30°C the refrigerating capacity of the same machine (with a compressor operating at 960 r. p. m.) was 565 kcal/h, the ratio and difference of discharge and suction pres­sures being respectively 1 1 .5 and 14.0 atm (Fig. 3), and with no vacuum in the evaporator

(Po = 1 .3 atm. abs.). The refrigerating effect per unit of swept volume was�:= 28 kcal

Qo m3, the performance factor Ne = 412 kcal/kwh.

Confrontation of the characteristics of two-stage and cascade refrigerating machines with those of a single-stage low temperature machine illustrates that the latter can be related to the class of simple by design low temperature refrigerating machines and be used when less capital investments and simplification of maintenance are more important than the power characteristics.

The experience of operating such a machine stresses that the tested single-stage compressor does not require any design modifications. An ordinary shell-and-tube water condenser was used in the capacity of a condenser for the high-boiling fraction though the application of a special (fractionating) condenser would make it possible to improve the characteristics of the machine.

Such a low temperature machine is characterized also by some other advantages. It has been determined, for instance, that such a generator of cold operates normally when applying oil XF-12 for lubricating the compressor. The main circuit of oil circulation in the system aligns with the circulation circuit of the high-boiling fraction. The circula­tion of oil in the low-boiling fraction circuit is insignificant. The non-freezing of the oil in the working parts of the expansion valves is provided by the presence of freon-12 in the composition of both fractions.

The discharge temperature at all the operating duties of the refrigerating machine did not exceed the range permissible for freon machines ( + 85 to + 95°).

The test has indicated also that the heat transfer from the working body improved with an increase of concentration of the low-boiling component.

The condensation of the mixture of refrigerants in apparatus with condensation inside and between the pipes has been investigated as well. It has been determined that the separation of fractions in an apparatus with condensation inside the pipe was unsatis­factory in the low-boiling temperature circuit.

The theoretical and experimental investigations carried out by us point out to the possibility of widespread utilization of mixtures of refrigerants in compression machines.

696

IIl-7

The Lubrication of Refrigerant 22 Machines

La lubrification des machines a frigorigene 22

Dr. W. P. SEEMANN Deutsche Shell, P. A. E. Lab., Hamburg, Germany

Dr. A. D. SHELLARD Shell International Petroleum Co., London S. E. 1, England

SOMMA/RE. Les huiles classiques pour refrigerateur ne se melangent pas avec le frigori­gene 22 bien au-dessous de -5°C et, a mains qu'on ne dispose de separateurs d'huile appro­pries, il peut se former dans l' evaporateur une huile visqueuse avec perte possible de rendement.

Certaines huiles synthetiques sont tres miscibles avec le frigorigene 22 mais ne sont pas tres repandues chez les fabricants de refrigerateurs, probablement pour des raisons economiques.

Une autre classe de composes, les benzenes a/kyles sont seduisants du point de vue economi­ques et, lorsqu'ils sont traites specialement, ils se melangent bien avec le frigorigene, contien­nent peu de substances insolubles dans le frigorigene et sont d'une stabilite chimique convenable. Des essais de duree du compresseur dans les group es hermetiques a 140 ° C port ant sur 3 mois ont confirme que la stabilite chimique de ces benzenes alkytes speciaux etait satisf aisante et qu'ils n' attaquaient pas les materiaux de construction usuels. La tegere usure observee sur le piston indiquait que leurs proprietes lubrifiantes sont a la limite. On peut cependant obtenir une lubrification appropriee en leur adjoignant d'autres hydrocarbures choisis, tout en maintenant les substances insolubles du frigorigene a mains de 0,02% en poids et une stabilite chimique superieurs a 120 h dans l'essai Philipp. On obtient une miscibilite complete avec lefrigorigene 22 jusqu' a -30° c au mains, ce qui permet de repondre a un grand nombre de besoins domesti­ques et industriels. Le comportement de ce lubrifiant dans les essais sur des groupes de labora­toire et dans le fonctionnement pratique est tout-a-! ait satisf aisant.

INTRODUCTION The use of conventional refrigerator oils for the lubrication of Refrigerant 22 (R 22)

machines presents a problem because of the immiscibility of oil and refrigerant at low temperatures, such as are encountered in many evaporators. At these low temperatures the oil separates from the refrigerant and, although it contains a certain amount of R 22, depending upon the temperature [1, 2] its viscosity can be sufficiently high to prevent it from returning with the refrigerant into the compressor. A build-up of oil in the evapora­tor can result in reduced efficiency (whether in a dry or flooded evaporator), restriction to refrigerant flow and even blockage and, in extreme cases, to oil starvation in the com­pressor. One solution to this problem is to provide an efficient oil separator and a means of returning the oil to the compressor. Another solution is to design a lubricant with the required miscibility properties and both Loffler [3] and Steinle [ 4] have proposed synthetic oils which are completely miscible with R 22 down to -70 or even -80°C. These synthetic oils have not, however, been widely accepted, probably for economic reasons and, in the case of one of them, because of its poor hydrolytic stability.

The use of oil separators is not always attractive to refrigerator manufacturers parti­cularly when questions of space limitation and costs are all important. There is therefore a real need for a satisfactory lubricant for use with R 22 machines and this paper summa­rises the work which led to such an oil.

MISCIBILITY OF MINERAL OILS WITH R 22 The miscibility curves for three naphthenic-based refrigerator oils A, B and C of

different viscosity are shown in Fig. 1 .

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+30

+20

+ 10

± 0

u - 1 0 0 w � -20

� ffi -30 a. � 1- -40

-so

-60

-70

-80 0

/ � J '

I / r---.... � ""' ""' // � "'

� [\. OIL A. I/ / / _ -...... � "' / � '""' "' r--.

'// '� """ "' r'\. "'

I I ' ' OIL B. I\ OIL C .

I \ OIL VISCOSITY cS \

I \ OIL D. AT 50 °C A 2 7 - 5

B 20. s

c 9 . 5

D 20. s

10 20 30 40 50 60 70 80

0/o W OIL Fig. r . Miscibility curves for mineral oils with R 22

Outside the area bounded by each of these curves the lubricant is completely miscible with R 22, whereas inside the curves the reverse is the case. The curves can be character­ised by quoting :

1. the temperature corresponding to the highest point of the curve, above which the oil and refrigerant are completely miscible in all proportions.

2. the temperature corresponding to the point on the curve above which a mixture of 10% weight of oil and 90% weight R 22 is completely miscible-this corresponds to the floe point (10% weight oil) criterion used to characterise the low temperature perform­ance of refrigerator oils in R 12 machines.

Miscibility with R 22 is greatly influenced by the viscosity of the lubricant, a change from 40cS (5.3°E) to 9.5cS (l.S0E) (measured at 50°C) reducing the maximum of the miscibility curve from 21 ° C to -S°C.

Loffler [5] had previously drawn attention to this effect of oil viscosity on miscibility with R 22 and from a study of a wide range of mineral oils had concluded that for opti­mum miscibility one needed a lubricant which

1. is dewaxed as far as possible. 2. has a high aromatic content, consistent with good chemical stability. 3. has a low viscosity, consistent with the lubrication requirements of the compressor. Loffler selected a mineral oil of high aromatic content and concentrated these aromatics

by solvent dewaxing at -70° C and in doing so reduced the maximum of the miscibility curve from -3° C to -24°C.

Using rather more practical refining methods it is possible to improve the miscibility characteristics of mineral oils with R 22 to a certain degree. For example, oil B in Fig. 1 is a conventional naphthenic-based refrigerator oil, whilst oil D is equi-viscous but has been refined to give a higher aromatic content. In doing this, the maximum of the miscibility curve has been reduced from +S°C to -14°C. Compressor life tests with oil D, however,

69S

III-7

were unsatisfactory because quantities of black deposits were formed, indicative of oil instability in the presence of the refrigerant.

MISCIBILITY OF AROMATIC OILS WITH R 22

The concentration of aromatics from mineral oils is therefore not a practical method of producing lubricants for use with R 22. By far the best way is to select from the petroleum or chemical industries, fractions which are as completely as possible aromatic in type. In choosing these it has to be borne in mind that the lubricant must comply with many other requirements apart from good miscibility with R 22. One of the most important of these is good chemical stability in the presence of the refrigerant and compressor components, over the whole range of operating temperatures. Furthermore, the selected aromatics must have the correct viscosity and volatility characteristics to meet the lubrication requirements of the compressor.

The properties of an aromatic-rich fraction, consisting mainly of alkylated benzenes, which had been refined to remove thermally unstable molecules, are given in Table 1 .

Table 1 . Physical properties of refrigerator oils based on alkylated benzenes

Oil I Test Alkylated Alkylated

Method benzene benzene fraction fraction blended

with mineral oil

Density at 15°C g/ml DIN 51 757 0.878 0.878 Flash point 0. C. ° C DIN 51 584 173 180 Viscosity at 20°E cSt/E DIN 51 562 119.9/15.8 140.0/18.4 Viscosity at 50° E cSt/E DIN 51 562 20.3/2.91 23.75/3.30 Viscosity Index ASTM D 567 -41 4 Neutralization value, mgKOH/g DIN 51 558 0 0 Colour ASTM ASTM D 1500 < 0.5 < 0.5 Pour Point ° C DIN 51 583 -42 -44 Refrigerant 12 insolubles % w DIN 5 1 590 < 0.02 < 0.02 Refrigerant 22 insolubles % w < 0.02 < 0.02 Floe point (R. 12) °C DIN 51351 E < -74 -52 Maximum of the miscibility curve with R. 22 ° C < -74 -35 Refrigerant resistance/R. 12 hr "Philipp Test" DIN 51 593 > 120 > 120 Refrigerant resistance/R. 22 hr "Philipp Test" DIN 51 593 > 120 > 120 Bosch Bomb Test*/R. 12 Satisfactory Satisfactory Bosch Bomb Test* /R. 22 Satisfactory Satisfactory

* A test to measure the sludging tendency of the oil as well as its tendency to promote copper plating [6, 7] .

PERFORMANCE OF AROMATIC OILS IN REFRIGERATOR COMPRESSORS

In every respect this lubricant has excellent properties and therefore it was subjected to compressor tests. Hermetic domestic compressors of the reciprocating type were used and the life test procedure applied [8]. The compressors were run continuously for a period of 3 months and were thermally insulated to maintain a temperature of 140°C on the motor windings. After test the lubricant and compressor components were exam­ined, paying particular attention to wear, corrosion, oil properties, state of insulation, copper plating and deposits on the discharge valve, and it was found that there were

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score marks on the pistons and wear debris in the compressor housings. Apart from this lack of lubricity of the alkylated benzenes they appeared satisfactory in respect of chemi­cal stability and had no deleterious effect on compressor lacquers and components used in this type of hermetically sealed compressor.

In order to improve lubricity, selected mineral oil components were blended with the alkylated benzenes in various proportions, as shown in Table 2. The corresponding miscibility curves with R 22 are presented in Fig. 2.

+ 10

! O

- 10

-20

u 0 w -30

a: ::::> t{ -40 a: UJ � -50 UJ t-

-60

-70

-80

-90

I '

0

( ' ""'

/ � � /,/ l'QIL H. --� I I'- OIL G. r-...

"r-... !'--.. OIL F. r -......_,. OIL E.

10 20 30 40 50 60

°loW OIL Fig. 2. Miscibility curves for alkylated benzene/mineral oil blends with R 22

Table 2. Lubricant blends of alkylated benzenes and mineral oil

Oil % w mineral oil blended with Viscosity at 50°C of the mineral designation the alkylated benzenes oil fraction

cS I OE

E I 10 9.5 1.8 F 30 20.5 2.9 G I 30 27.5 3.7 H I 38 27.5 3.7

The temperature at the maximum of the miscibility curve increases with both increas­ing mineral oil concentration and increasing mineral oil viscosity. Compressor life tests were carried out on the above oils and entirely satisfactory-results were achieved when the mineral oil concentration was maintained at 30 % w or above. The properties of the lubricant selected to give both optimum performance in the life test and miscibility with R 22 and, which for the purpose of this paper has been designated Oil I, are given in the last column of Table 1 .

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Although the miscibility of alkylated benzenes with R 22 is appreciably affected by mineral oil addition, the resultant blend does satisfy the low temperature requirements of a wide range of refrigerator compressor types. This is particularly true when one bears in mind that the concentration of lubricant circulating with the refrigerant is normally far less than 10 % weight.

SERVICE EXPERIENCE WITH OIL I

This lubricant has now been in service in a number of refrigerator compressors for periods of up to two years. One example will serve to illustrate this - a 3-stage Linde unit used for liquefying chlorine. In this unit the maximum compression temperatures af­ter the first, second and third stage are 85°C, 60° C and 70° C and the lowest evaporator temperature is -55° C. This compressor has operated satisfactorily for a period of almost two years with this lubricant.

When changing over from a mineral oil to a blended type of oil, as described above, it is always advisable to check that the various components in the compressor are compatible over the whole range of operating temperatures with the new lubricant. Compressor life tests give a good indication of any interaction of components and lubricants and these can often be supplemented by laboratory tests. It has been found, for example, that certain types of rubber gasket material can swell and lose their elasticity when heated with this lubricant, so that it was necessary to select a new type of gasket material in certain instances.

Refrigerator unit design features can of course influence the ease of return of the re­frigerant/oil mixture from the evaporator to the compressor and, in this connection, Goldner's [9] work is worthy of attention. From our experience, there is no doubt that this lubricant does offer substantial advantages in R 22 machines, but work will continue to improve still further its miscibility characteristics with the refrigerant.

REFERENCES

r. ]. L. Little, Ref. Eng., 60, r r91, r952. 2 . C. M. Bosworth, Ref. Eng., 60, 617, r952. 3. H. ]. Loffler, Kaltetechnik, 9 (5), 1 35, 1957. 4. H. Steinle, Kaltetechnik, 8, 1 2, 1956. 5. H. ]. Loffler, Kaltetechnik, 9 (9), 282, 1957· 6. H. Steinle and W. Seemann, Kaltetechnik, 3, 194, 1951 . 7. H. Steinle and W. Seemann, Kaltetechnik, 3, 90, 1953· 8. H. Steinle, Kaltetechnik, 7 (4), 101, 1955· 9. H. Goldner, Der Kalte-Praktiker, r , roo, 1962.

DISCUSSION

C. Hocking, Sweden : The lubricants described in the paper are completely of the non­polar type. I think that polar compounds would be more effective as lubricants, and therefore I want to know if such materials have been examined.

A. D. Shellard, U. K. : The paper deals entirely with polar compounds although we were studying the effect of some non-polar lubricants in the form of some additives. This work, however, is far from being complete, and it would be pointless, therefore, to introduce it at this stage.

701

III-8 Examinations on the Behaviour of Plastics in Hermetic Units

Examen du comportement des matieres plastiques dans les groupes hermeriques

Dr. H. STEINLE Robert Bosch GmbH., Giengen (Brenz), Germany

SOMMAIRE. Aujourd'hui il y a des matieres plastiques en forme de feuilles et de fibres, lesquelles sont thermiquement stables dans les mixtures des fluides frigorigenes et des huiles jusqu' a 160°C. Leurs points de fusion sont assez au-dessus de la temperature d'emploi. Dans l'etat d'equilibre avec l'humidite d'air moins d'eau sera adsorbee et aucune eau seraformee en cas de surchauffage a cette temperature. Les esters de l'acide terephtalique (Mylar, Hostaphan, Terafilm, Diolen, Trevira) et les polyamides (Nylon, Perlon, Durethan) peuvent etre seches lentement avec de l' air sec ou au vacuum a des temperatures elevees jusqu' a 100 ° C. Au-dessus de 140°C il y a de nouveau une petite separation d'eau. L'acide terephtalique et des esters oligomeres de l'acide terephtalique ainsi que des produits de decomposition des polyamides n'etaient pas a demontrer.

Dans le fluide R 12 comme avec l' huile les esters de l' acide terephtalique meme pour long­temps ne vont pas changer leurs proprietes mecaniques et electriques. En cas de formation d'acides les polyamides deviendront pailleux. A une temperature de 160°C les polyamides et quelquesuns des esters de l' acide terephtalique manquent en devenant cassants. Dans le fluide R 22 comme avec de l'huile ils se forment des separations d'oligomeres des esters de l'acide terephtalique, lesquelles cristallisent en rafraichissant. A partir de 100°C les proprietes mecaniques changent dans des limites critiques. L'eau n'a pas d'influencejusqu' a 180 mg/kg de fluide frigorigene. Elle sera aussi stipuze en usant des sechoirs avec les tamis moleculaires. L'huile elle-meme n'a aucune influence. Les matieres decrites sont a utiliser avec le fluide frigorigene R 12, mais pas avec le R 22 jusqu' a une temperature permanente de 140°C.

INTRODUCTION

The electrical insulating materials used in the motors of hermetic units preferably consist of cellulose and their derivates. Cellulose is hygroscopic and picks up water in equilibrium [1] from the surrounding atmosphere. Therefore its use is requiring careful drying processes for the hermetic units [2] . Moreover at temperatures above 120° C cellulose begins to form water [1, 3] gases and acids [4] by thermal destruction. All these contaminants lead to the failure of the refrigerating machine by reactions of the refrigerant and the oil, the corrosion of the constructing materials as by sludging and plug-ups of the metering device. Failure mostly is caused by superheating of the cel­lulosic materials [5]. Refrigerant, oil and laquers are much more stable. All reaction products make the insulating materials to become more or less conductive for electrical current. So they finally lead to breakdown and short-circuit. The efforts for using materials more stable in place of cellulose are now successful by using some plastic materials.

MATERIALS TESTED AND PROCESSING

It is not the purpose of this paper to designate materials with good conduct in hermetic units. By use of very different materials it was suggested to prove a method for testing the behaviour of plastics in refrigerants and oils and their mixtures at different temperatu­res from 100°C up to 160°C. Machine tests are expensive and they take long time. Moreover the behaviour of machines is very complex. So it is difficult to valuate the

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behaviour of the plastic material itself. Extraction-tests and bomb-tests can be found valuable for the determination of the principal behaviour of plastics. Finally selection is performed in the life-tests at elevated winding temperatures during some months.

Table 1 . Methods to test plastics for use in hermetic units.

1 . Extraction with refrigerant at normal boiling point Extraction with solvent used for cleaning

2. Water adsorption, drying process, thermal stability

3. Bomb-test at 100°C in sealed glass tube and at 140°C and 160°C in steel autoclave with refrigerant-oil-mixtures

4. Floe-test with 100°C sample

5. Elongation at break of new material and after bomb-test at 100, 140, 160°C

6 . Unit life-test 12 weeks at 140°C and at 160°C.

The following materials have been tested :

a) Esters of terephtalic acid (Mylar (R), Hostaphan (R), Terafilm (R), Diolen (R)).

b) Polycarbonate (Makrolon (R)).

c) Polyamide (Durethan BK 40 f (R)).

d) Polyformaldehydes (Delrin (R), Hostaform C (R)).

e) Phenolic resins (Bakelite (R), Resitex (R)).

Only such materials have been tested, of which softening and melting points are high enough above the commonly used temperature in the motors of hermetic units. Water content and the most commercial method for drying have been tested at the same time by flushing with dry, hot air [3]. The upper temperature limit was found by the same test.

Solvent resistance with trichloroethylene was measured by Soxhleth-extraction [6], while refrigerant resistance was made in cold-extraction apparatus [6] at normal boiling point.

Chemical stability in refrigerant-oil-mixtures at elevated temperatures and the in­fluence of these temperatures on mechanical and electrical properties have been found by bomb-test in glass tubes at 100° C and in autoclave up to 1 60°C [6]. The samples in the tubes may be used for floe test to find materials in solution precipitable by cool­ing down [7].

Finally life-test has to run in hermetic units at elevated wiring temperatures as high as 140° C and 1 60 ° C for 12 weeks.

At the same time materials had to be found which not only do not form water in the case of superheating, but also allow higher winding temperatures in service than cellu­lose does. For cellulose the upper permanent temperature limit is given at 120° C, life-test is made at 140°C. It was desirable to find materials withstanding 140°C for a long time and which allow 1 60°C for life-test. Performance of motors could be increased by using such materials. At the same time drying processes should be simplified by picking up all the water adsorbed by driers with highly effective molecular-sieves [1).

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WATER CONTENT AND DRYING

Water adsorption and drying of insulating materials have been investigated after saturation in wet air with 35% and 65 to 75 % relative humidity. The water was caused to evaporate its quantity and determined by flushing with dry nitrogen at increasing temperature. The whole water content of Mylar, Hostaphan and Terafilm is about 1 % by weight and only 10% of that of cellulosic materials [3, l]. Desorption is possible at temperatures up to 100°C. The small amounts of water desorbed at temperatures higher than 140°C can be picked up by a drier arranged in the circuit and filled with molecular sieves [l]. Only the very tight binding of water in this drying agent prevents hydrolysis of the esters of terephtalic acid [9], which leads to embrittlement.

Diolen fibres adsorb 0.4% of water at 65 % relative humidity and 0.9% at 95 % humidity, each at 20° C.

Makrolon is able to adsorb up to 0.2% water at 60 % relative humidity.

Polyamides are very different in water adsorption, characteristics depending from processing in production.

Phenolic resin products mostly contain cellulosic filling materials; therefore they behave like cellulosic materials.

TEMPERATURE LIMIT

Temperature limit in air was found by increasing temperature in drying process up to beginning deterioration. Cracking products were frozen, determined and analysed. These tests show that up to 200° C, this means near melting point at 230 to 250°C, esters of terephtalic acid are forming no terephtalic acid and no oligomers of esters of terephtalic acid.

Therefore it is not to suppose that formation of gasous products carbon dioxide, carbon monoxide and water take place. As for temperature-stability these esters may be used in the form of foils and fibres at temperatures up to 160°.

Diolen fibres may be heated up to 175° C without damage, at 230° C they begin to soften.

Polyamides with high molecular weight like Durethan BK 40 F do not crack below softening point at 220° C. But in air at temperatures more than 160 ° C embrittlement and shrinkage are to be observed. Makrolon resists up to more than 160° C, its melting temperature is at 220 to 230° C.

The upper temperature limit of some materials is reduced by the influence of refrig­erants and oil, the results were gained by bomb-tests at 100° C, 140° C and 160°C. The reasons for this are very different. Mylar and Hostaphan become brittle at 160°C. Elongation decreases nearly to zero. Extraction of softening oligomers may be the reason. Terafilm, another ester of terephtalic acid resists at 160°C. Polyamides embrittle if acid formation takes place, as acids react directly with polyamides. The use of polya­mides does not seem suitable for this reason.

SOLVENT RESISTANCE

Trichloroethylene is the most commonly used solvent for degreasing parts. Extrac­tion in Soxhleth-apparatus must not exceed 0. 1 % by weight to prevent solubility, swelling, shrinkage or embrittlement of the insulating materials. Extraction with refrig­erants at the normal boiling point was made in cold extraction apparatus; extract shall not exceed 0.1 % with R 12 or 0.2% with R 22.

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All materials mentioned above, except Makrolon, resist to the solvent action oftri­chloroethylene at 87° C, of R 12 at -29° C and R 22 at -41° C. Makrolon is completely soluble in trichloroethylene and up to some percents in R 22. Makrolon is not to be used in hermetic units of this type.

Bomb-tests with refrigerant and oil at 100° C are performed in glass tubes but at 140°C and at 160° C in V2A-steel bombs. A testing time of 14 days is sufficient to produce nearly the final state of properties in the materials. Water was added in some tests in the form of wet paper.

The results show that oil has no significant influence and water accelerates only to a small amount the reduction of mechanical properties. This is merely caused by refri­gerant and temperature.

Mylar and Hostaphan show by R 12 a very strong embrittlement so that elongation and yield point practically decrease to zero. With R 22 at 100° C the yield point of Mylar and Hostaphan is reduced practically twice as much as by R 12. Moreover R 22 pro­duces a crystallic precipitation of oligomers, which might lead to plug-up in the capillary tube or expansion valve. Mylar and Hostaphan can be used with R 12 at permanent temperatures up to 140° C, at higher temperatures embrittlement and cracking might lead to short-circuit in the motors and failure of the refrigerating machine. In machines with R 22 Mylar and Hostaphan are not to be used for electrical insulation because of the precipitations in the refrigerant oil mixture. Makrolon resists R 12 and oil. In the R 22-oil-mixtures already after seven days at 100° C a milky discoloration appears, while the surface softens and the form is destroyed. The solubility in trichloroethylene pro­hibits the use of Makrolon also in most of the refrigerating machines with R 12.

LIFE-TEST RESULTS

Life-tests of hermetic units at elevated winding temperatures [8] which normally have a duration of 12 weeks, serve for testing the chemical and physical stability of all materials. Life-tests with cellulosic insulated stators are performed at 140° C. Plastic insulations with permanent stability at 140° C should withstand life-tests at 160°C. With the same materials tests at 140°C were running over years. After several periods the materials in such compressors were tested for change in properties and impurities.

Insulation of the motors was made from Mylar, Hostaphan and Durethan BK 40 F and Diolen fibres. Seven compressors were working at 140°C from 12 weeks to nearly 3 years, some of them are still running. All 140°C test machines were running without any trouble. At 160°C winding temperature twelve compressors failed within 12 weeks by short-circuit. Electric load was the same as for 140°C test machines. The higher tem­perature was merely produced by thermal insulation on the surface of the compressor.

The results are reported in Table 1. In R 12 from the machine run at 140° C no reaction products R22 and carbon dioxide could be found. These two gases are the first indicators for reaction and chemical instability (10.). On the other hand the R 12 vapor phase with 0.67% by vol R 22 and 1.4% by vol of C02 and 0.15% of CO in 160° C test machi­nes demonstrates chemical reactions of R 12 and esters of terephtalic acid. Polyamide Durethan BK 40 F became brittle by acid reaction; Mylar and Hostaphan are much damaged and break. But in 140° C life-test all these materials are nearly without change in mechanical and electrical properties after years. Durethan BK 40 F, Mylar and Hostaphan may be used in 140 ° C temperature range over long periods, but they fail after short time at 160°C. Water content in the driers was very low in both cases, but twice as much in 160°C test machines than in 140° C ones. The difference may be estimated to be reaction water. Electrical connections being insulated with Hostaphan­foil and Diolen-fibres performed satisfactory in 160° C life test over 12 weeks.

All these tests are indicating again the value of laboratory tests for evaluating materials for use in refrigerating machines, but programm has to be on broad base.

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Table 2. Results of life-test at 140°C and at 160°C winding temperature

l40°C - 18 months l6o°C - 3 months

R 22

co. co

Yield point : before test after test

Elongation : before test after test

Non condensable gas in R 12

<0.1 Vol % R 22

0 co.

0.04 Vol % CO

Mylar-Hostaphan

91-97 Kp/cm2 80--88 Kp/cm2

140--160 % I00--120 %

Yield point : before test after test

Elongation : before test after test

Durethan BK 40 F

0.67 Vol % i .4 Vol % 0.15 Vol %

91-rg7 Kp/cm• 42 Kp/cm1

140--160 % 25- 90 %

Cable-insulation and sheet insulation yellowed or browned, but elastic. Elongation and tensile strength not measurable, because material is flowing. Slot insulation yield point changed from 16.o Kp to 25.6 Kp.

Cable insulation and slot insulation dark and brittle, breakable. Cable insulation showed ruptures and was sporadically molten. Dielec­tric strength decreased from l 2 400 volts to 10400 volts.

H,O in R 12 : 23 mg/Kg

H,O in drying agent : 0.87 %

H,O in R 12 : 40 mg/Kg H,O in drying agent : i .54 %

In Table 3 the behaviour of the above mentioned plastic materials is given as result of the tests pointed out in Table 1 . Limit values are proposed for the assessment.

Table 3. Behaviour of plastics in laboratory- and life-tests.

Materials tested

Mylar Hostaphan Terafilm

Makrolon

Durethan BK 40 F

Delrin Hostafore C

Bakelite Resitex

Limit value I

Extraction

R 12 'X I Trichloro-0 ethylene %

+ + + +

+ + ---

+ -

+ +

Thermal stability

oc

140

160

1 00

140 ? �---·--·--- - - - ----

+ + ? + 1 00

+ + + + 120

0,1 % 1,0 o/o l40/160°C

Elongation Floe-test l40°C I l6o°C

+ + ? -

+ + ? -

+ + + -- --- --- ---

- - ---- ---

+ - ---- --- ---

- + -- + -

--- ---

-�--

decrease less -40°C than 50 %

Life-Test

l40°C I l6o°C

+ -

+ -

+ + --- ---

- ---- ---

+ ? ---- -- ---

- -

- ---- ---

+ +

--- ---

3 3 months months

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REFERENCES

r. H. Steinle, Wassergleichgewichte in gekapselten Kiiltemaschinen ; Kiiltetechnik Vol. I I (I959), Nr. Io, p. 336.

2. C. C. Roberts, Proper instrumentation assures dehydration of hermetic units ; Refrig. Engng. Vol. 56 (I948), p. 23r.

3. H. Steinle, Temperaturbestandigkeit nichtmetallischer Stoffe in Kaltemaschinen ; Kiilte­technik Vol. 4 (I952), p. 28; Werkstoffe und Korrosion, Vol. 3 (I952), Vol. 4I9.

4. F. M. Clark, Pyrochemical behaviour of cellulose-insulation; Electrical Engng., Vol. 54 (I935), p. ro88.

5. R. T. Divers, Electrical insulation and the hermetic motor ; Refrig. Engng. Vol. 65 (I957), Nr. 8, p. 33.

6. H. Steinle, Priifung der Kiiltemittel-Bestiindigkeit nichtmetallischer Stoffe ; Kiiltetechnik, Vol. 3 (I95r), Nr. 5, p. r r o und Nr. 6, p. I39·

7. H. Steinle, Bestimmung des Flockpunktes von Kiiltemaschinenolen, Entwurf DIN 51351 ; Kiiltetechnik, Vol. I4 (I962), Nr. 4, p. 126.

8. H. Steinle, Versuche iiber Kupferplattierung in Kiiltemaschinen; Kiiltetechnik, Vol. 7 (1955), p. IOI.

9. ]. P. Hurtgen a. A. R. Mounce, New insulation systems for hermetic refrigeration motors ; ASHRAE-Journal, Vol . I (1959), Nr. 7, p. 60. ]. P. Harrington a. R. ]. Ward, Polyester film insulation in hermetic motors ; ASHRAE­Journal, Vol. r (I959), Nr. 4, p. 75·

10. H. 0. Spauschus, Reaction of refrigerant I2 with petroleum oils; ASHRAE-Journal, Vol. 3 (r96I), Nr. 2, S. 65.

Summary of the Discussion (Papers III-3 + III-8)

A. Neuenschwander, France : I am most impressed by these two papers and particularly by that of Dr. H. 0. Spauschus (U.S.A.), as it is most important for engineers to know, when they introduce a new material, what effect the refrigerant or oil would have on these materials.

H. W. Fischer, U. K. : I should like to ask Dr. Spauschus, if Fig. 2 on page 4 and the text do take into consideration the presence of moisture. After all the idea of fitting molecular sieve driers with Refrigerants 12 or 22 for that matter would be to dehydrate a plant usually known to have contained water or water vapour. If this water vapour breaks down, Fig. 2 may well be affected by further oxygen release.

H. 0. Spauschus, U. S. A. : The molecular sieve tested is in fact dry material. I did, however, run similar tests where the molecular sieve contained about 2 % weight of moisture. But this did not alter the results appreciably as the figures are only altered very slightly, i. e., something of the order of 5 %. This is the sort of figure that we would expect in a reasonably dry refrigeration system.

D. Weissbarth, Germany : I want to ask you, Dr. Spauschus, if you could suggest any methods which would make it not too difficult to check whether a re-distilled refrigerant from a breakdown and particularly from a burn out is still pure enough. In cases like that a vapour test might not always be possible, but a liquid test may be very much easier. You also refer to the decomposition of Refrigerant 22. Could you also give some hints as to the stability of Refrigerants 1 1 and 1 13 compared to Refrigerant 22?

H. 0. Spauschus, U. S. A. : I am quite ready to answer the second part of the question, but I prefer to leave the first part to the author of the other paper, Dr. H. Steinle. I feel that of the Refrigerants 11 and 1 13, Refrigerant 11 is most certainly the poorest as far as

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stability is concerned. Refrigerant 11 is much more chemically reactive than R 12 and also with regard to its thermal stability.

This work is well documented. R 1 13 is very similar to R 12 and R 22, as far as the oil stability is concerned it is much more superior to R 12 normally, since as the amount of chlorination increases, it is much easier to split the carbon-chlorine bond for chemical reaction. There are other problems associated with R 22, since it has the single hydrogen pattern which makes it susceptible to other kinds of breakdown.

H. Steinle, Germany : Referring to the first part of Mr. Weissbarth's question we can only suggest a test for hydrochloric acid but I do not consider this very satisfactory, as it is practically impossible to tell just how much of the decomposition is due to the com­pressor itself.

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Effect of Heat Exchange between Capillary Tube and Suction Line on the Performance of Small Hermetic Compressor Systems

Influence de l'echange de chaleur entre le restricteur et !'aspiration sur le rendement des petits compresseurs hermetiques

K. ROELSGAARD and K. V. VALBJ0RN DANFOSS A/S, Nordborg, Denmark

SOMMAIRE. L'echange de chaleur entre le restricteur et !'aspiration dans les petits compresseurs hermetiques a fait l'objet de nombreuses recherches au cours du temps et a donne des resultats assez contradictoires.

Pour essayer d'obtenir une image plus exacte de !'influence de l'echange de chaleur sur le rendement d'un systeme frigorifique hermetique, on a conr;u un systeme experimental def ar;on a permettre de determiner la puissance totale du syteme en chif.fres absolus, avec et sans echange de chaleur et sans toudier du systeme frigorifique hermetique.

De plus, on a examine dans ce systeme experimental !'influence des differents types d' echan­geurs de chaleur.

BACKGROUND FOR THE INVESTIGATION

There were two major reasons for undertaking this investigation : Firstly, the fact that the so-called single-joint evaporator during the last couple of years

has gained ground within the household refrigeration industry. This type of evaporator, which is characteristic in having only one outer tubing connec­

tion with the capillary tube running to the canal system of the evaporator inside the suc­tion line, gives a neat appearance to the cabinet in its simple layout of piping. On the other hand the possibilities of obtaining an effective heat exchange between capillary tube and suction line are not too good.

And secondly, that the tendency in the design of the compressors, as in many other fields, is directed towards better utilization of the material, and thereby in case of hermetic compressors greater power converted within still smaller surfaces which leads to a heavier thermal load on the motors. It is, therefore, interesting to learn how the heat exchange influences the capacity of the system and the motor temperature.

The investigation was made on a common combination of compressor, capillary tube and suction line for a household refrigerator, and the results thus refer primarily to this combination. It cannot be excluded that investigating other combinations of circulated amounts of refrigerants, flow rates and surfaces may show minor deviations from the results reported here.

DESCRIPTION OF THE TESTING ARRANGEMENT

By the design of the testing arrangement (Fig. 1) it was aimed at reaching a system which resembled a normal system for a refrigerator very closely ; at the same time it had to be possible to measure the capacity of the system with an accuracy of 1-2 %·

The condenser was a tube-on-sheet type mounted between two vertical walls provided with small adjustable heating elements underneath for maintaining a constant temperature of the air to the condenser.

The compressor was a 4.3 cm3/rev. (0.26 cu. in/rev.) 2 pole model with a capacity of 136 kcal/h at -15° Cf +50°C, 32° C ambient temperature (540 BTU/h at + 5° F/ + 122°F, 90° F ambient temp.) 50 c/s. It was also built-in a "lifelike" way.

The capillary tube had an inside diameter of 0.8 0 x 3500 mm (0.031 " x 1 1 .5 ft) length, a common size in conjunction with the compressor mentioned.

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IIl-14

Healing_ elements

Condens_er:

Fig. I Test Apparatus Arrangement

The evaporator, Y:!" copper tube x 1000 mm (39.5") length, was mounted in a tank with the connections to the evaporator, i. e. the capillary tube and the suction line, led through the wall of the tank in a way that made heat conduction from tubing to tank vir­tually impossible. In the tank was a secondary charge of R 21 and a system of heating elements. The input to these elements was controlled by the ambient temperature of the tank, the input being balanced in a way that kept the temperature of the tank's shell and the ambient temperature identical during the experiment, i. e. the tank acted as a calori­meter as the input measured on a watt-hour-meter corresponded to the refrigerating capacity.

By changing the ambient temperature the temperature and pressure of the secondary system in the tank were changed automatically, i. e. the load on the evaporator was altered which again led to a change of the evaporating temperature (in the primary system). By means of this it was made possible to make a number of measurements at different evap­orating temperatures. In some instances it was, however, necessary to adjust the charge of the primary system to reach stable conditions and to obtain that the temperature of the suction gas at the discharge from the calorimeter tank was on the same level as the tem­perature of the calorimeter tank. This procedure is in itself not a source of error as a varia­tion of the charge only means a more or less active surface of the evaporator.

THE INFLUENCE OF THE HEAT EXCHANGE ON THE REFRIGERATING CAPACITY

In the first series of tests comparative measurements were made on the system without heat exchanger and the system with heat exchanger (Fig. 2). The heat exchanger was, as can be seen from the sketch, built-up in the traditional way by soldering the capillary tube to the suction line. The capillary tube was placed on the suction line according to the well-known "rule of thumb" that the soldering ought to have a length of at least 1 m (40") (usually it is difficult to obtain much more in practice) and that approx. Y:i of the detached capillary tube ought to be at the condenser and the remaining % at the evaporator end.

The measurements show clearly an increase of the capacity in kcal/h when using a heat exchanger differing from approx. 3 % increase at -25°C (-13°F) evaporating temper­ature to approx. 17% increase at -5°C ( +23°F) evaporating temperature. The higher

712

III-14

I Heat exchanger design

Jia �aldering _Q J-'. 200 1000 200 .l:

Compressor� � Evaporator � lt �� ;: Suction tub e ' 11{ • 1400 CU 10"'<:;

Capillary, .. 2 ..... o,8 . 3500

190 20%

15 --180 1n r-,__ Increase in capacity in

5 r----- percent using heat exchange 170 l o .___

160

150 � 140 \' 130 " \ � I

" \ "' ___ With heat exchange 120 r-..

'\ � "" _ _ _ Without heat exchange

110 " " "\ I� ... 100 �

" , r-.. '\I\. """ ' 90 ' '\ ":.Consumption

80 l"r-.. twatts1

I " � 70 � � ' ' �""" 60 - """"'-<=--- I ' Capacity r---� - � f"--"'= "'---= " - r-- - Condensing temp.

50 r-- I- ..._ -r---...:::::.- I liquid temp. at inlet

40 ' capillary tube

·5 ·7,5 · 10 ·12,5 · 1 5 ·17.5 ·20 ·22,5 ·25 ·27,5 · JO Evaporating temperature in °C

Fig. 2 Influence of heat exchange on capacity

condensing temperature and the higher liquid temperature at the entrance of the capillary tube (sub-cooling temperature), which is measured on the system without heat exchanger can be explained by the fact that the volumetric efficiency of the compressor and thus the amount of circulated refrigerant in kg/h is increased on account of the lower suction-gas temperature at the entry in the compressor. The difference in the two systems' watt consumption is very small being within approx. 2 %· The application of a heat exchanger, therefore, gives an increase in the specific refrigerating capacity (kcal/Wh) which more or less corresponds to that of the increase of refrigerating capacity (kcal/h).

In the following series of experiments the refrigerating capacity of the system in kcal/h was measured at various designs of the heat exchange (Fig. 3), some measurements with heat exchangers soldered as before but without following the "rule of thumb" and some with a heat exchanger of the type which is used for the single -joint evaporator i. e. with the capillary tube inside the suction line. However, 1 m (40") of the capillary tube was applied for the heat exchanger in all designs. The results of the measurements confirm that it is really quite a good rule to divide the detached capillary tube on the soldered heat exchanger with approx. % at the condenser and the remainder at the evaporator end, as the system with this design had the highest refrigerating capacity of the four measured. Irrespective of towards which end of the suction line the capillary tube is moved the refrigerating capacity of the system is diminished. The differences are moderate and is the best (the "normal" design) compared with the design with internal capillary tube it can be seen that the latter gives the system a refrigerating capacity of 94-97 % of the capacity of the "normal" design.

713

III-14

� .x 160

ISO

140

130

120

110

too 90

80

70

sn

so

4n

' \:\ ,K�

The length of heat exchanger is Suction tube 114 '• 1400 capillary

200 1000

Compressor IJJ end �

--- A:800 8 : 1700 - - A :SO 8:24SO - - - A:2'SO B : SO

held at 1000mm tube • 0,8x 3SOO

Evaporator end

- - - - Capillary inside suction tube over a length of IOOOmm starting IOOO mm from evaporator

'�� '��

-� � i -� ' 0..

'�:--� , , �

'� � ( ' apacity

- s ·7.5 ·10 -12.S -15 -17,5 -20 -22,5 -25 -27,5 -30 Evaporating temperature in ·c

Fig. 3 Influence on capacity of different heat exchanger designs

THE INFLUENCE OF THE HEAT EXCHANGE ON THE MOTOR TEMPERA­TURE

Most of the standards dealing with specifications for hermetic compressors stipulate max. permissible winding temperatures which limit the compressors' range of application and, therefore, you will often be in a situation where the question of reducing the wind­ing temperature for inst. 5°C (9°F) can be decisive as to whether a system (refrig­erator or freezer) can be approved by the respective testing institutions. From Fig. 4 you will note that the increase of the suction gas temperature you have by applying heat ex­change results in a substantial increase of the winding temperature from approx. 4°C (7° F) increase at -25°C (-13°F) evaporating temperature to approx. 10° C (18°F) increase at -5 ° C ( + 23 ° F) evaporating temperature. The measurements were taken during the first series of tests with the system built-up with and without heat exchanger as stated in Fig. 2. The lower curve in Fig. 4 shows the increase in kcal/h of the amount of heat which is led into the compressor by the warmer suction gas, when heat exchange is applied, calculated on basis of the enthalpy which corresponds to the two curves for the suction gas temperature in Fig. 4 and the circulated amount of refrigerant. However, the whole amount of heat does not work as an increase in the heat which must be emitted through the housing of the compressor, as part of the heat is led out of the compressor by the discharge gas.

CONCLUSION Application of the most effective heat exchange between capillary tube and suction

line results in an increase of the refrigerating capacity (kcal/h) and the specific refriger-

714

c

' 3 2

0 0 0 0

10 .,.. - -n 14

I t? In A • l '

5 4 4 3 3 2

O n 5 n 5 0 ' 0 5 0 5 0

---

...........

- �

� --

�� ----

-

with heat exchange

-suction gas temp. at inlet compressor

'-without heat exchange

Increase in winding temperatl.l"e employing heat exchange

Theoretical increase in heat load on compressor housing employing heat exchange

-5 -7,5 -tO -12.S -15 -17,5 -20 -22,5 -25 -27,5 -30

Evaporating temperature in •c Fig. 4 Influence of heat exchange on the thermal:balance of the compressor

III-14

ating capacity (kcal/Wh) of approx. 3% at -25°C (-13°F) evaporating temperature to 17 % at -5° C ( +23°F) evaporating temperature, but in return the winding tem­perature of the compressor is increased by approx. 4 to 10°C (7 to 18°F). Less effective heat exchangers as for inst. the single-joint evaporators suction line with internal capillary tube gives 4-6% smaller increase of the refrigerating capacity, but also a smaller increase of the winding temperature.

For household refrigerators and freezers, normally working with evaporating tem­peratures between -20° C and -30° C (-4°F and -22° F) the influence of the heat exchange on refrigerating capacity and winding temperature is therefore rather small. In some cases there may be a reason for considering whether you need the lower winding temperature you obtain by omitting the heat exchange more than the slightly higher refrigerating capacity which the heat exchange entails. What remains is perhaps the most important mission of the heat exchanger in the household refrigeration industry : to prevent that the part of the suction line which is outside the refrigerated room "sweats" moisture into the insulation of the cabinet or out on the surroundings especially during max. load:as for inst. pull-down or freezing.

DISCUSSION

V. S. Martinovsky, U.S. S.R. : I would like to ask Mr. K. Roelsgaard, what he consi­ders to be the best heat exchange surface between a capillary tube and the suction line on small units which use capillary tube as an expansion device.

K. Roelsgaard, Denmark : The design which shows up best is one where the capillary is soldered to the outer surface of the suction line which gives a 2 mm. diameter for the capillary tube and 61/2 mm. for the suction line diameter. This is of course a very small surface indeed.

M. Pertzelan, Israel : In small refrigerators such as domestic refrigerators, where the capacity itself is extremely small, every small saving is of great importance. Although we have not done actual basic tests on the various types of heat exchangers, we also came to the conclusion that the method of soldering the capillary tube to the suction line is also giving the best performance in very hot countries like Israel.

715

IIl-18

A Special Method for the Measurement of Capacities and Character­istics of Thermostatic Expansion Valves

Une methode speciale de mesure de la capacite et des caracteristiques des deten­deurs thermostatiques

Prof. JUNKICHI NAGAOKA Tokyo University of Fisheries, Tokyo, Japan

SOMMAIRE. Cette methode est destinee a mesurer les capacites d' ecoulement defrigorigene liquide et les caracteristiques des detendeurs thermostatiques.

Le frigorigene provenant du detendeur penetre dans un separateur oit le liquide est separe de la vapeur produite. La vapeur entre ensuite dans /'aspiration du compresseur tandis que le Jrigorigene liquide restant est refoule par pompe a l' entree du condenseur. Le liquide refroidi ce melange a la vapeur de frigorigene provenant du compresseur, la refroidissant et la condensant en partie. Ainsi seule la vapeur de frigorigene produite par le detendeur doit etre comprimee jusqu' a une pression elevee. On peut avec ce moyen mesurer la capacite du detendeur qui est de 3 a 4 f ois celle du compresseur. La quantite de Jrigorigene liquide s' ecoulant par le detendeur est mesuree par un thermometre tandis que le thermo-bu/be est maintenu a une temperature donnee par immersion dans un bain a temperature Constante.

It is a well-known fact that the best results can be obtained when the capacity of the thermo-expansion valve is matched to that of the evaporator. When the capacity of the thermo-expansion valve is too large compared to that of the evaporator, the valve is liable to function unstably, i. e. the suction pressure changes periodically and some­times liquid refrigerant flows back into the compressor. On the other hand, if the valve capacity is too small, the evaporator will be starved.

For these reasons it is absolutely necessary to measure the exact capacity of the thermo-expansion valve, and therefore this special method has been devised.

One method of rating and testing refrigerant expansion valves is described in the A. S. R. E. Standard No 17-R.

According to this method, the expansion valve is placed in an ordinary refrigerating machine and the refrigerant which flows through the expansion valve being tested, is completely evaporated in the evaporator. The capacity of the valve is determined from the refrigerating capacity measured by the calorimeter or from the measured quantity of liquid refrigerant which flows through the valve. By this method, a valve capacity which is greater than that of the refrigerating machine cannot be measured and more­over a large heat source is needed to evaporate the liquid refrigerant which flows through the expansion valve.

When the liquid refrigerant flows through the valve, part of it evaporates and cools the remaining part of the liquid refrigerant, but most of it remains in the liquid state. If the remaining liquid refrigerant is fed back to the high pressure side by the liquid pump, only the gas which evaporates when the liquid refrigerant flows through the valve is compressed by the compressor and liquified in the condenser. We can therefore measure the capacity of a large expansion valve with a rather small refrigerating machine. This is the principle of the special method described in this paper. In this method the mixture of gas and liquid refrigerant which has flowed through the expansion valve, flows into a liquid separator where the refrigerant gas and liquid is separated.

The liquid refrigerant is fed by the liquid pump back to the condenser. The refrig­erant gas is sucked and compressed by the compressor and then flows into the condenser where it mixes with the liquid refrigerant and is condensed by cooling water. The percentage of refrigerant x which is evaporated when the liquid refrigerant flows through the expansion valve is shown in Table 1 .

717

III-18

Table 1. (Refrigerant R . 12. Condensing temperature 38°C.)

Evaporating Capacity of temperature compressor

o c R. T.

5 3.2 0 2.6

-10 1.6 -20 0.9

Refrigerant gas generated when liquid

refrigerant flows through expansion

valve x %

22 24 29 34

Calculated capacity of equipment

R. T.

15 11 5.5 2.6

For each kg of liquid refrigerant which flows through the expansion valve, we need to compress only x kg of refrigerant gas in the compressor. In this way, we can test an expansion valve whose capacity is 1/x times the compressor capacity and no heat source is needed to evaporate the liquid refrigerant. Of course this is an ideal case. Actually we must let some liquid refrigerant flow through the bypass and some liquid refrigerant is evaporated by heat leakage. The actual capacity is therefore somewhat below the calculated value shown in Table 1. For a condensing temperature of 38°C and an evaporating temperature of 5°C, we could measure the capacity of a R 12 expansion valve of about 13 R. T. (Fig. 5) using a 3 R. T. R 12 compressor.

We can therefore measure the capacity of an expansion valve which is more than 3 times that of the R 12 compressor.

/E lectric Motor

L ��r�;; ��l�� Check Valve th:�b��i�� Const. ?ressure Expansion Valve Hand. Expansion Valve Gear Pump Drain Valve Suction Pressure R'1ulator 6i,9ttse��"J.!lor Water Re3ulatif>3 Valve

Fig. r . Diagrammatic General Arrangement of Thermo-expansion Valve Testing Equipment .

In Figs. I and 2 is shown the schematic general arrangement and piping of this testing equipment. The refrigerant gas, which is separated from the liquid-gas mixture in the liquid separator, is sucked through the suction pressure regulator by the compressor and compressed. This high pressure R 12 gas flows through two oil separators which

718

Packless Valve Strainer Flow-meter

+ Expdnsion Valve E under test

---@- Solenoid Valve. under test -1L Thennomeler -.e- Si91it Glass

Fig. 2. Diagrammatic General Arrangement of Expansion Valve Tester

111-18

are connected in series, is freed from oil and then liquified in the condenser. The con­denser is cooled by cooling water regulated by a water regulating valve and the con­densing temperature of the refrigerant is held constant. On the other side the liquid refrigerant, separated from the refrigerant gas in the liquid separator, is sucked and compressed by the liquid pump and fed back into the condenser. To prevent too much liquid being pumped out of the separator and to keep its liquid level constant, the li­quid refrigerant from the liquid pump is partially by-passed and this excess liquid refrigerant flows into the liquid separator through a float valve which keeps the liquid level in the separator constant. The rest of the liquid refrigerant flows into the condenser.

The liquified refrigerant in the condenser and the liquid refrigerant fed back from the separator by the liquid pump after being stored for some time in the receiver, flow to the expansion valve tester. Some of this liquid refrigerant is by-passed to the liquid separator through a constant pressure expansion valve which keeps the pressure of the separator at the determined value. The rest of the liquid refrigerant flows through a flow-meter and then through the thermo-expansion valve being tested. To keep the separator pressure constant in this equipment, a constant pressure expansion valve in the by-pass pipe and a suction pressure regulator in the suction pipe are used but it may be better to use an expansion valve in the by-pass pipe and an evaporating pressure regulator in the suction pipe.

The condensing pressure is kept constant by the water regulating valve. When the capacity of the expansion valve being tested is large, much refrigerant gas will be sepa­rated in the liquid separator. In this case the suction pressure of the compressor must be increased so that more refrigerant gas can be sucked by the compressor. When the capacity of the expansion valve is small, the suction pressure of the compressor must be lowered. The liquid refrigerant which flows through the expansion valve must be somewhat subcooled so that there are no bubbles in the liquid refrigerant. Whether there are bubbles in the liquid refrigerant or not, can be seen with the sight glass. When there are bubbles in the liquid refrigerant, it may be cooled by the cool liquid from the liquid pump which is by-passed to the double tube heat exchanger. The degree of subcooling of the liquid refrigerant which flows to the expansion valve tester may be adjusted by adjusting the amount of cool liquid from the liquid pump by-passed to the heat exchanger. It may also be adjusted by adjusting the amount of cooling water which flows through the cooling coil in the receiver.

When a thermo-expansion valve is being tested, its thermo-bulb must be kept in a constant temperature bath. Figs. 3 and 4 are photographs of this equipment. The small size equipment uses a 3 H. P. R 12 compressor and can measure the capacity of valves of about 10 R. T. when the condensing temperature is 38° C and the evaporating tem­perature is 5°C. The large size equipment uses a 20 H. P. R 12 compressor and can measure about a 60 ton valve capacity at the above mentioned temperature conditions. In Figs. 5 and 6 are shown the characteristic curves of thermo-expansion valves taken with this equipment. The ordinate shows the flow rate in refrigerating tons and the abscissa shows the temperature of the thermo-bulb. In these valves the liquid refrigerant

719

III-18

Fig. 3. Small Size Expansion Valve Testing Equipment

720

I compressor 3 receiver 2 condenser 4 liquid separator

5 expansion valve tester 6 liquid pump

Fig. 4. Large Size Expansion-Valve Testing Equipment r compressor 3 expansion valve tester 2 liquid separator 4 liquid pump

III-18

begins to flow when the thermo-bulb temperature is about 2.5° C above the evaporating temperature and the flow rate of refrigerant reaches the rated value when the superheat changes about 2.5°C.

14 13 12 � II

.r: .10 ?;- q "G {} � 7 � 6 � 5 ;:! 4 t s � 2

I 0

-

-

-

-

-

-

-

'7

I I� // v I I I

C»tcA rana.citv ( /0 ?. T l

I '/ //; Con dens in� Temperature J8 'C _

II (/ Eva.poratin� Temperature 5 •c

11 II �

8 q 10 1 1 12 /J 14 IS 16 /7 Temperature of Thenno·bulb 'C

Fig. 5. Characteristic Curve of ro R. T. R 12 Thermo-Expansion Valve.

-

- v:::::: ----R"teJ Caoa.citv i SS RT l �

- / I I I - ;//' Condensin� Temperature 38 'C

'/ [vaporn.tin9 Terniierature s ·c - I I I I I ..... 7 ff Cf 10 11 12 13 14 !J 16

Temperature of Thermo-bulb ·c Fig. 6. Characteristic Curve of 5.5 R. T. R 1 2 Thermo-Expansion Valve.

721

III-19

Some Experiments on the Discharge Coefficients and Characteristics of Ammonia Thermostatic Expansion Valves

Quelques experiences sur les coefficients de refoulement et les caracteristiques des detendeurs thermostatiques a ammoniac

Prof. JUNKICHI NAGAOKA Tokyo University of Fisheries, Tokyo, Japan

SOMMAIRE. Dans ces experiences, on a mesure le debit d'ammoniac liquide a travers des detendeurs thermostatiques de capacites et types diff erents, avec ou sans tubes de refoulement, a /'aide d'un f/uxmetre, en p/afant /e thermobu/be dans un bain a temperature Constante et en f aisant varier la pression d' aspiration pour permettre la surchauffe.

On a recherche /'influence des tubes de refoulement sur le debit de /'ammoniac liquide. Pour determiner le coefficient de refoulement de ['orifice du detendeur, on a mesure le debit de I' ammoniac liquide avec une ouverture constante du detendeur, en maintenant le thermobulbe a l' air et en plafant une cale au lieu du ressort qui neutralise la pression de gaz du bulbe et action de soufflet. Le coefficient de refoulement qui est de 0,65 environ est peu influence par /es dimensions de l' orifice de detendeur, mais il est un peu par la forme du pointeau du detendeur. Le retard de la reaction du thermobulbe aux variations de surchaujf e a ete mesure egalement.

In these experiments, the rate of flow of liquid through thermoexpansion valves of various capacities, with and without discharge tubes, was measured by a flow-meter installed in the liquid line of an ammonia refrigerating machine. The suction pressure was varied while the feeling bulb was kept in a constant temperature bath or in open air.

From this data, the discharge coefficient of the valve orifice was calculated. The effects of discharge tubes on the flow of liquid ammonia, were also investigated. To determine the coefficient of discharge of the expansion valve, the amount of liquid ammonia which

Fig. r. Thermo-expansion valve with

distance piece instead of

adjusting spring.

723

III-19

flowed through the expansion valve being tested at constant valve opening was measured by the flow-meter. The feeling bulb was kept in open air whose temperature was be­tween 25° C and 29° C and a distance piece was placed in the valve insteadoftheadjusting spring which counteracted the gas pressure of the feeling bulb acting on the membrane or bellow of the power element (Fig. 1).

In these tests the inlet pressure of the expansion valve was kept between 10.0 and 11 .6 kg/cm2 g by changing the volume of cooling air and that of cooling water of the evaporative condenser. The outlet pressure was varied by throttling the suction valve of the compressor, by changing the air volume through the brass coil in the cold storage room and by changing the temperature of the cold storage room by opening or shutting the door. The subcooling of the liquid ammonia which flows through the expansion valve was kept between 4 and 6° C by a double tube heat exchanger installed in the liquid line. The liquid ammonia was cooled by the cooling water in this heat exchanger. The valve needles used in these experiments were the cone type and the flat type shown in Fig. 2. The diameter of the valve orifice of the cone type valve needle was 1 .6 mm and

Fla.t Type Cone Type Fig. 2. Shape of valve needle.

3.1 mm and that of the flat type valve needle was 3 mm. The opening of the valves was changed by turning the adjusting stem by 45 degrees, 60 degrees and 90 degrees. One turn of the adjusting stem opened the valve by 1.5 mm in the case of the cone type and 1 mm in the case of the flat type.

The coefficients of discharge were calculated from the following simple hydraulic equation used in the paper by Mr. D. D. Wile published in Refrigerating Engineering, Aug. 1935 :

M = 720 Cn A y 2 g o (P1 - P2) The above equation, converted into the metric system, can be written in the following

form :

Where

M = 3880 Cn A V P1 -P2

M = Cn = A P1 = P2 =

refrigerant flow, kg/h coefficient of discharge orifice area, cm2 valve inlet pressure, kg/cm2 g valve outlet pressure, kg/cm2 g

The orifice area is calculated by the following equations. In the case of the cone type valve needle:

A :n d h sin 2

In the case of the flat type: A :n d h

where d diameter of the valve orifice, cm h amount of valve needle opening, cm CG angle of valve needle shown in Fig. 2

The measured values are plotted in the curves shown in Figs. 3 and 4. The curve in Fig. 3 shows the coefficient of discharge of the cone type valve needle. This curve shows that the coefficient of discharge is little affected by the orifice area and the valve

724

� "' 70 "'"'" � 60 ] .ro Cl 4- 40 0

� .JO � 20 � !O

u

Sho..pe of Va,lve Needle Cone }ire (Nee le Anqle 45')

Inlet Pressure 10 - I /- / 1Worn'j·

-i j-f! Subclolin� f-1--

x 1 'f I I

- Orifice An�le of Amount of -- Diameter Turn Open i'!J. _

0 J· I 11111t '10' O ·J?. "'"' - " J I 45 0-187 -

,_ x 1 6 qo 0·375 -+ l ·G 45 0-187

0 2 J 4 Outlet Pressure Ka/crn' 3

Fig. 3. Coefficient of discharge of ammonia thermo-expansion valve (Cone type).

III- 19

orifice diameter. The coefficient of discharge seems to decrease slightly as the suction pressure increases. This tendency agrees with the results of Mr. Wile's experiment on Freon and chlormethyl. In Fig. 4 the curve for the fiat type valve is shown as a full line and that for the coefficient of discharge of the cone type valve needle is shown as a dotted line. The coefficient of the fiat type seems to be somewhat smaller than that of the cone type valve needle. In this case the coefficient also tends to decrease as the suction pressure increases.

� 70 "' 60 '-" .'E. .fO � 40 "i> JO � 20 <J � 10

C> u

Shll.pe of V cilve Need.le Fla.t Type Orifice Diameter 3 mm , An9le of Turn 60° Amount of Openin� J.16 mm , Inlet Pressure 10·6-l \ ·6 K'/{m•� .Subcool ini S- 6 'f_� l-n- =..:e - - � - - - - -

� o"::..::: t--

� �--�-Shape of Valve Needle

- Ful l Line Fla.t Type � --- Dotted Line Cone Type (Needle An�le 45°) -

0 2 J 4 s 7 Out let Pressure KM:m2�

Fig. 4. Coefficient of discharge of ammonia thermo-expansion valve (Flat type).

As there was no oil left in the piping when the expansion valve tested was taken away, there seemed to be no oil in the liquid ammonia which flowed through the ex­pansion valve. From these experiments the coefficient of discharge of the liquid ammonia is about 0.6 to 0.65. This is a little smaller than that of Freon 12 which is about 0.75 and nearly equal to that of methyl chloride which is about 0.65.

The formula for the coefficient of discharge of Freon and methyl chloride given in the paper by Mr. Wile gives a somewhat smaller coefficient for ammonia than the measured one. A discharge tube is used to restrict the flow of refrigerant through the small area of the valve orifice.

Sometimes the capacity of a thermo-expansion valve is changed by merely changing the discharge tube. To determine the effect of the discharge tube, the rate of flow of liquid ammonia refrigerant through the expansion valve was measured by the flow-meter keeping the feeling bulb in the constant temperature bath and varying the outlet pressure of the valve.

725

III-19

Cone Type Valve Or if ice Dia.. J I mm

(A) without d.ischar� tube (BJ with l6mm<I> discho.r�e tube (C J curve ( 8) correted in

proportion to curve (A )

<A Jv . . / . /<BJ

Superheat Chan�e ·c 5 6

� ""' � 160 () €

�120 :s:! ::i .,,.. � 80 .... () Q) "'d rr. 4lJ '"' 0 CL

0 0

Fla.t Tyrie Valve Orifice Dia. 3 mm

!Al without dischar�e tube <Bl with 2-Smm <I> d.ischo.r9e tube (C) with 2 mm<I> dischar�e tube (D) curve <C J corrected in

v

proportion to curve rAJ (8) ,/ w>:,.y

/ ,/"': c c ) //, �­g"--17 rDJ I'

2 4 6 Superheat Cha.n�e 'C

8

Fig. 5. Change of capacity of thermo-expansion valve by discharge tube. Fig. 6. Change of capacity of thermo-expansion valve by the size of discharge tube.

In Figs. 5 and 6 are shown the characteristic curves of some expansion valves with and without discharge tubes. The discharge tube effectively restricts the maximum flow of refrigerant through the valve but cannot sufficiently reduce the flow rate when the superheat change is small. It seems better to use stiffer adjusting springs when the capacity of a valve is decreased with a discharge tube.

To measure the time-lag of the power element, two constant temperature baths whose temperatures differed only slightly, were prepared. The feeling bulb which had been kept in one of these constant temperature baths is suddenly placed in the other one. The time required for the flow-rate of the refrigerant to attain the new constant value was measured. In Figs. 7 and 8 are shown the curves of the flow-rate of the thermo­expansion valves when the temperature of the feeling bulb was suddenly changed. There is little difference in the time-lag of the feeling bulb whether charged with a liquid or with some refrigerant adsorbent.

0

,,.,.-· . /

/Cl ,..£.-6--f'""o-

- � --� -o-

/ 0/ 'rAl 1/ - �- - ><- -- -I

I I I/

0

' / v \B1 /

i I JO 60 90 120

Time Ela.psed in Sec

� 401--��l--�--l��--1��-+��-+��-j & � 201--�--1��-+��--1��-+���l-�-j

0

0

'--�---'Jo��--'6-

0

���ro

��-1�20--,---

�-!�SO-:--�--:-:MO

Time El a.psed in Sec 7 8

Fig. 7. Time-Jag of thermo-expansion valve's feeling bulb. Fig. 8. Time-lag of thermo-expansion valve's feeling bulb.

726

(a) liquid charge (made in Japan) (b) liquid charge (made in U.S.A.) (c) adsorption charge (made in Denmark).

III-39

Liquid Control for High Evaporator Efficiency

Reglage de liquide pour accroitre l'efficacite des evaporateurs

J. LORENTZEN, Chief Engineer Development Department, A/S ATLAS, Copenhagen, Denmark

SO MMAIRE. Les calculs conventionnels de transmission de chaleur du fluide frigorigene aux evaporateurs ont ete discutes.

Ces calculs ne prennent pas en consideration les eff ets de la fonction dynamique des evap­orateurs.

Des f acteurs import ants dynamiques sont la perte de pression du ftuide frigorigene pendant son evaporation et la variation du debit de liquide au cours du temps. L'importance de ces f acteurs est discutee pour /es evaporateurs reg/es par detendeurs thermostatiques.

En ce qui concerne /es evaporateurs noyes ayant une circulation par gravite, on decrit un nouveau regleur de liquide ameliorant la fonction dynamique de ce type d'evaporateur.

Plusieurs types d' evaporateur avec circulation forcee de ftuide frigorigene ont ete discutes, et /es possiblites du systeme de grand debit de circulation de liquide ont ere examinees plus soigneusement. Le choix du meil/eur type d' evaporateur pour un service specifique depend d'un certain nombre de f acteurs dont quelques-uns des plus importants en relation avec le systeme de reglage du fluide frigorigene ont ere traites dans cette etude.

INTRODUCTION

Normally the heat transfer (Qo) of evaporators with plane cooling surface (F) is studied by means of the equation (1)

Qo = k . F . Lltm (1) with Lltm as logarithmic mean temperature difference, and k the overall heat transfer factor is given from the equation (2), reading

1 1 1 0 - = - + - + , + f k OG1 OG2 11. (2)

1 1 overall heat transfer resistance -

k is the sum of surface heat transfer resistances-and

OG1 1 0 - the heat conductance resistance of the wall -,- and the fouling factor f.

OG2 II. For cylindrical surfaces, the overall heat transfer factor is usually referred to one

particular of the wall surfaces (F1), and the partial resistances of equation (2) are cor­rected correspondingly:

1 1 F1 1 F1 - = - -- + - -- -- + --k1 OG1 F2 OG2 Fm

(3)

Very often the two last items are insignificant and when omitted the equation (3) is reduced to

__!__ = __!__ + F1 . L k1 OGi F2 OG2

(4)

For extended surface-coolers, the heat conductance resistance from the extended parts of the surface wall to the prime surface often plays an important part, and this resistance usually is accounted for by introduction of the fin efficiency factor, �' into the calculations [1 , 2].

727

III-39

The equation ( 4) may then be transformed as follows :

1 k1 =

1 F1 0:1 c - (1 - ;) �

1x ) + F2

1 (5)

In most of the extended surface coolers the surface extensions (Fex) form the greater

Fex part of the nominal surface (F1) and when F --+ 1, equation [5] may be simplified

1 further:

(6)

In Table 1 are put together some characteristic examples of heat transfer resistance figures for normal type evaporators, showing the relative importance of the two main resistances according to equation (6).

F 2 is in every case the surface of the refrigerant side of the evaporator and the overall heat transfer factor k, is referred to the air- or waterside surface F1•

The two last columns of table 1 show the increase of the overall heat transfer factor when the heat transfer co-efficient of the refrigerant side o:2 is increased with 10% (giving k1,1 ) and with 100 % (k2,0).

This simple static evaluation of evaporators shows, that improvements of heat transfer on the refrigerant side is very important for shell- and tube coolers, it is also important for extended surface air coolers, but it does not seem important for bare pipe air coolers.

Static heat transfer co-efficients for evaporators have been rather well investigated for various conditions as well for the refrigerant side as for the other side of the wall for instance to air or to water [3]. This normally forms the basis of evaporator designs.

In a dynamic analysis of the heat transfer of evaporators, we have to account for the pressure drops in the refrigerant system. These pressure drops may be classified into two main groups :

1. Pressure drops connected with the mass flow of refrigerant through the evaporator. 2. Static pressure of liquid.

Table 1. Relative importance of heat transfer resistances and total effect of refrigerant side improvement of 10 % and 100 %

F1 1 k1'1 k2•0 Evaporator Type

;<X1 F2 <X2 k1 k1 k1

NH3 - bare pipe coil natural convection 0,1000 0,0017 0,1017 1,002 1,01

NH3 - bare pipe coil forced convection 0,0154 0,0017 0,0171 1,01 1,05

NH3 - extended surface coil forced convection 0,0305 0,0125 0,0430 1,03 1 ,17

R 12 - extended surface coil forced convection 0,0305 0,0250 0,0555 1,04 1,29

NH3 - Shell & Tube plain tube, -watercooler 0,0004 0,0010 0,0014 1,07 1,55

R 12, Shell & Tube plain tube-watercooler 0,0004 0,0020 0,0024 1,08 1,71

728

IIl-39

The pressure drops raise the liquid boiling point as compared to the saturation temperature at the suction pressure and thus reduce the available temperature difference of the evaporators.

Another subject of dynamic analysis of evaporators forms the time variations of liquid flow through the paths of the evaporator and the corresponding variations in the characteristics of the heat transfer.

In fact the main characteristics k and LI tm are variable with time and with the spot on the evaporator surface, and complete dynamic analysis of all important factors will hardly be possible.

Although a thorough analysis may not be possible, the fact remains that these questions have great importance.

A technical approach is to study each dynamic factor separately, and choose the des­ign so that each factor only may give little adverse effect on the overall, mean perfor­mance.

When considering the dynamic behaviour of refrigerant evaporators, we come to the conclusion that the liquid refrigerant control and the heat transfer on the refrigerant side is often of much larger importance than the previous calculations show (Table 1).

In the following some characteristic" dynamic incidents will be discussed in connection with various types of liquid control systems used in industrial refrigeration.

COOLERS WITH THERMOSTATIC EXPANSION VALVES.

The dynamic problems of thermostatic expansion valve operation have been extensi­vely treated by other authors, showing the importance of the following characteristics of the regulating valve [ 4, 5].

Valve size in relation to heat load conditions, liquid pressure and sub-cooling and evaporator pressure.

Valve opening- and closing characteristic as a function of bulb temperature at con­stant evaporator pressure, including effect of internal friction of valve closing member.

Valve superheat as a function of evaporator pressure. The time constant of the bulb system as installed in the plant, including the effect

of the surroundings and the thermal coupling of the bulb to the suction line wall. When using thermostatic expansion valves, the influences of the above-mentioned

characteristics lead to cycling expansion valve operation which may often have a detri­mental effect on evaporator heat transfer. The conditions are most difficult when se­veral coolers are connected to the same suction line, and when the coolers have a low time-constant which means that the normal liquid refrigerant charge of the evaporator is relatively low as compared to the amount of liquid evaporated per unit of time.

Fig. 1 shows an example to illustrate this effect for an air cooler working with Refrigerant 22. Curve A is the measured mean air temperature in the cooler, curve B is the evaporator temperature according to the suction pressure and curve C is the mean air temperature to be expected according to equation (6) with the evaporator tempera­ture and the refrigeration load in question. In this case the mean overall heat transfer coefficient of the cooler is reduced by 70-75 per cent, because of the dynamic charac­teristics of the liquid control system. A good deal better result was obtained by exchang­ing the regulating valve with another type, having more suitable dynamic characteris­tics, but this regulation system is so sensitive to irrelevant factors that it should be avoided on industrial evaporators of importance [6].

FLOODED EVAPORATORS WITH SELF CIRCULATION OF THE LIQUID

For evaporators with self circulation of liquid refrigerant in a tube system it is known that the mean heat transfer coefficient between refrigerant and tube wall has a maximum value when liquid recirculation is at a positive, limited value [7]. The object of the liquid regulation device is thus to maintain such liquid charge in the evaporator to obtain the recirculation required.

729

III-39

10 '° lime (min_)

Fig. I . Dynamic behaviour of air coole1 with thermostatic expansion valve (A) as compared to calculated static conditions (C). Suction temperature curve B.

The normal regulation systems for such evaporators have been the constant level controls, but these are in principle unable to maintain the correct amount of liquid recirculation under variable operating conditions, and are sometimes liable to heavy cycling of operation [4] . Now a new regulating device for a flooded evaporator has been developed, Fig. 2.

The evaporator has a downcomer (11) for the liquid recirculation, and the cooler itself (8) forms the riser side of the circulation path. The recirculated liquid from the riser side is trapped in the separator (9). It flows through the orifice (13) in an interior chamber in the separator. The amount of liquid recirculation is measured in that in­terior chamber by the liquid level above the orifice. The amount of recirculation chosen for the evaporator is the flow of liquid through the orifice when the liquid level reaches the phial (10). This phial is the electrically heated bulb of a thermostatic expansion valve (2), also known as a level master control. This circulator control system, which is simple and reliable has given very good operating results showing favourable dynamic characteristics.

730

® © (!)

Fig. 2. Circulator regulator system for flooded evaporator.

III-39

EVAPORATORS WITH FORCED CIRCULATION OF LIQUID REFRIGERANT

Recently forced circulation of liquid refrigerant in evaporators has become rather common, and several methods may be used [4] . The dynamic characteristics are fa­vourable because the circulation of liquid refrigerant in the evaporator is maintained continuously by means of a mechanical pump or a gas pump [8]. The extra heat load introduced by the pump energy consumption may usually be kept so low that it will be of no importance.

For a single evaporator the installation of a pump is a rather heavy complication, but for a large number of evaporators on one system, this complication will be offset by the simplifications that can be made for each evaporator as compared to other methods. With larger distances between the evaporators and liquid separators, however, diffi­culties with mass flow pressure drops may be significant.

The ordinary bottom feed system may be easily applied with almost any form of evapo­rator and liquid distribution is simple. It gives approximately similar heat transfer conditions as the flooded evaporator, but fluctuations in regulation system do not affect the heat transfer.

The ordinary top feed system has been much advocated because the effect of static liquid head in the evaporator is reduced and because the evaporator has a smaller refri­gerant charge. The system is, however, only applicable in special cases, with certain designs of evaporators and it imposes serveral restrictions on the arrangement as well. Practical results of the top feed system, however, is usually inferior than those of the bottom feed system because of the difficulty in attaining the correct liquid distribution.

A special form of the top feed system is the film flow evaporator. This form is mainly suited for vertical evaporator surfaces and has so far been of little use in the refrigeration field. It presents some design complications with the liquid distribution system, but may give significant improvement in the heat transfer factor as compared to other systems.

Another interesting way of improving the heat transfer factor is by using a very high rate of liquid recirculation [9]. With liquid flow at high Reynolds number the heat trans­fer coefficient between wall and liquid may be much higher than between wall and boil­ing liquid of conventional evaporators. By this high rate circulation system the heat load is carried by the liquid as sensible heat to a flash chamber where the evaporation takes place. The cooler is a liquid cooled heat exchanger, and not an evaporator in the proper sense of the word. The amount of liquid recirculation must be chosen so that the liquid temperature rise does not occupy too much of the available temperature difference for the heat transfer. In this case the pump energy consumption will be of importance, and must be kept within acceptable limits. With this system, however, pressure drops along the refrigerant paths in the cooler lose much of their importance, and it has, therefore, also been applied for low-temperature installations under condi­tions when pressure drops in conventional evaporators are critical [10].

Different refrigerants have different possibilities for improved performance with this system, and the possible improvement is best with high heat loads. For a brinecooler the performance of the system has been calculated for different conditions with Refri­gerant 12 and with NH3 - and compared to the performance of conventional evaporators. The result is given in Fig. 3, where the line 100% represents the capacity (Q) and power consumption (N) of the combined compressor/evaporator-plant with conventional evaporator, and the curves represent the corresponding performance values for the same compressor and the same size cooler with this system. The curves denoted 5 is for systems with conventional evaporator dimensioned for 5 ° C mean temperature difference, and the curves denoted 10 for conventional evaporator size corresponding to 10°C mean temperature difference. To get a favourable performance, much higher rates of liquid circulation is necessary with Refrigerant 12 than with NH3 and the success of the system is then also more dependent on circulation pump co-efficient of performance. The fa­vourable range of operation is more narrow with Refrigerant 12, but on the other hand the possible gains are larger.

731

III-39

Table 2. Some important characteristics of evaporators with several refrigerant control systems.

.... i::: ·O Q) ij:J 0. ....

0 ::l tJ g .... .,

� "O ., Q) Q) u 0 Q) .... ..c:: .... Q:1 0. u � � u 'fil � ., ., ·� Pt ., ., OS Q) IS 8 � .... ci5 ti 0. CZl "O

Single Thermostatic cooler 1 3 expansion valve many

coolers 3

Flooded, single 2 2 2 level control many 2 2 2

Flooded, single 2 2 2 circulation control many 2 2 2

Pump, single 2 2 2 bottom feed many 2 2 2

Pump, single 2 2 3 top feed many 2 2 3

Film single 3 3 3 flow

many 3 3 3

High single 3 3 circulation cooler many 3 3

Legend: 3 = good, 2 = fair, 1 = bad

CONCLUSION

» u i::: Q) "O i::: Q) ....

gJl :.!:l u 6

1

1

2

2

3

3

3

3

3

3

3

3

3

3

s Q) .... ., » .,

-a 6h § ·;;; • Q) CZl "O

3

3

2

2

2

2

3

2

gJl .... g :g .... ., 0 Q) Q) u b.O � ;.s ]i � Q) .... Q) .... ..c:: Q) � e.c � S' � � 0 OS 0 OS ... :di :> 2:l :> ,g

2 2

2 2

2

1 2

1 2

1 2

2

2

2 2 2

2 2 2

3 3 3

3 3 3

3 3

3 3

Several aspects of the dynamic behaviour of industrial evaporators are of great im-portance to evaporator performance. In Table 2 a simplified evaluation of such dynamic characteristics is given together with other important factors.

732

III-39

150! I � �� � -t/' v�_y

./ / x1' .e:.,... .--.... --::::= :::::: ..... ).,

......_ � ·If/I� �_-1u2}

so:

i 0,1 as s

Fig. 3. Performance curves of a refrigerating plant with high rate liquid recirculation compared to a plant with conventional evaporator of the same size (roe %).

REFERENCES

r. M. Backstrom: Kylteknikem p. 290, Stockholm 1947. 2 . H. D. Bahr: Handbuch der Kaltetechnik III p. 1 18, Berlin 1959· 3. E. Hofmann: Handbuch der Kaltetechnik III, Berlin 1959· 4. G. Lorentzen: Evaporator Design and Liquid Feed Regulation, IIR Bull, Annex 1952-2 p. 235. 5. P. Danig: The Regulation of Thermostatic Expansion Valves, Kulde, Oct. 1962 p. 50. 6. ]. Lorentzen: Some Aspects of Flooded Type Evaporators in Marine Refrigeration. IX th Int.

Congress Refrig., Paris, Sept. 1955· 7 . ]. Lorentzen: Heat Transfer and Liquid Circulation in a Flooded Evaporator. X th Int. Congress

Refrig., Copenhagen, Aug. 1959· 8. G. Lorentzen & 0. A. Baglo: An Investigation of a Gas Pump Liquid Recirculation System.

X th Int. Congress Refrig., Copenhagen, Aug. 1959· 9. F. D. Bickel: Pumped Liquid Refrigerant in Scraped Surface Heat Exchangers. IIR Meeting,

Washington D. C., Aug. 1962. 10. A. Neuenschwander: Le Nouveau Caisson d'Altitude du Centre d'Essais en Vol de Bretigny­

sur-Orge. IX th Int. Congress Refrig., Paris, Sept. 1955·

733

III-21

Report on Liquid Level Control of Flooded Evaporator of Refrigerating System by Automatic Level Controller

Commande automatique du niveau de liquide dans les evaporateurs noyes des systemes frigorifiques

T. NAGANO and K. TAKAHASHI Research Laboratory of Japan Cold Storage Co., Ltd., Tokyo, Japan

SOMMAIRE. On eprouve parjois des dijjicultes dans les evaporateurs noyes a cause des variations frequentes des besoins de froid. Ces variations provoquent un fonctionnement inefficace et de dangereux coups de liquide dans le compresseur.

Cette recherche avait pour objectif de determiner lefonctionnement efjicace et sur des evapo­rateurs noyes dans les installations de fabrication de glace ou les installations d' entrepots frigorifiques. Pour regler le niveau de liquide dans les evaporateurs, on utilise couramment des flotteurs ou un dispositif semblable, donnant seulement un niveau de liquide constant, mais inca­pables de regler la hauteur de liquide d' apres les variations de la charge. Les AA. ont obtenu des resu/tats satisj aisants en utilisant des regleurs de niveau a dijjerentie/ de temperature COnfUS par eux. Le rapport explique brievement ces dispositijs de reg/age et donne des renseignements sur quelques experiences eff ectuees avec l'une des installations def abrication de glace.

1 . INTRODUCTION

The liquid level in the flooded evaporator of a refrigerating system is often influenced by the variation of the load or by the instability of the evaporating temperature. To obtain efficient operation of evaporators, especially flood type evaporators, we must keep its liquid level constant and also keep the superheat degree of the suction gas as low as possible. We have many ways to keep the liquid level constant; for instance, a float valve control is one of the most effective methods. However it is only able to keep the liquid level constant at a pre-determined height and is not able to adjust the liquid height for obtaining the most efficient operation of the evaporator according to the variation of the load.

If the refrigerating load suddenly diminishes, the height of liquid increases and the suction gas carries a larger quantity of liquid particles into the compressor which results in inefficient operation and often causes dangerous liquid hammer. To eliminate these difficulties, authors have improved the method of controlling the liquid supply to the evaporator by using the temperature differential thermostat which acts on the superheat temperature of the suction gas.

At first we tried to convert temperature difference between the temperature of the liquid in the accumulator and that of the gas in the outlet piping from the accumulator into the pressure difference between these two positions by using two thermo-bulbs containing the same kind of refrigerant (ammonia) and tried to have the magnetic valve in the liquid supply pipe of the accumulator open and close by this pressure difference.

From these experiments, however, we could not obtain satisfactory results because of ( 1) rather poor controlling accuracy ( ± 3 ° C) and (2) rather slow action, due to the heat capacity of thermo-bulbs, and (3) inconvenience of the installation caused by the capillary tube that connects the thermo-bulbs to the controller.

Since then, we have changed our testing plan into applying thermo-electrical processes and obtained satisfactory results.

2. THERMO-ELECTRICAL CONTROLLING

Noticing that very little difference of temperature between two temperature measur­ing resistance (sensitive elements) brings about unbalanced voltage in a Wheatstone

735

IIl-21

bridge circuit very sensitively, we tried to detect the direction of temperature difference by using a phase detector and amplifier.

We could control the liquid level in the accumulator by opening and closing the magnet valve in the liquid supply pipe using the electric current through the above Wheatstone bridge circuit into the control relay.

Fig. 1 shows our controller and two sensitive elements. Fig. 2 shows our controlling panel and Fig. 3 shows a schematic drawing of the application of the controlling device for an accumulator of an ice making plant.

E

S T

Temp. measuring resistance

Standard temp. measuring resistance

Fig. r . Controller and sensitive elements.

3. BRIEF EXPLANATION OF OUR EXPERIMENTAL PLANT

Our experiments are carried out at one of the refrigerating plants in Tokyo, having an ice making capacity of 17 tons (kilogram) per day and a few cold storage rooms.

Brief specification of the plant is as follows : 1. Ammonia compressors 2 sets

736

No. 1 compressor 95 mm (cyl. dia) x 76 mm (stroke) W type 8 cylinders Belt driven 1200 r. p. m. 37.2 Japanese tons of refrigeration (ET = -15°C, CT = +30° C)

No. 2 compressor 95 mm X 76 mm W type, 8 cylinders Belt driven, 1000 r. p. m. 31.0 Japanese tons of refrigerations (Japanese ton of refrigeration = 3320 kcal/h).

111-21

Fig. 2. Controlling panel.

2. Condenser Evaporative condenser, having a capacity of 75 Japanese tons of refrig-eration.

3. Ice making tank Number of ice cans Length of cooling coils Ice making capacity

300 lbs. 288 pcs. 11'4" X 1700 m

16900 kg/day of 24 hrs. 4. Cold storages

One cold storage room for foods, required refrigerating capacity 20.9 Japanese tons of refrigeration. One ice storage room, required refrigerating capacity 3.7 Japanese tons of refrigeration.

5. Accumulator of ice making tank Dia. 508 mm Length 2800 mm Temperature measuring resistances (sensitive elements) of liquid level controller are installed at two positions; one is within the suction pipe of the accumulator and the other at the bottom place of the accumulator itself. (Fig. 3)

4. RECORDS OF TEMPERATURE MEASURED AT THE EXPERIMENTS

We measured the temperature of the following spots with a recording thermometer. a) Temperature of the liquid in the accumulator in which one of the sensitive elements

is installed. b) Temperature of gas in the suction pipe from the accumulator where one of the

sensitive elements is installed. c) Brine temperature in the ice making tank. d) Temperature of liquid between the expansion valve and the liquid supply magnetic

valve. e) Suction gas temperature of No. 1 compressor. f) Suction gas temperature of No. 2 compressor. Results of measurement are shown in Fig. 4, Fig. 5, Fig. 6.

Test No.

1 2 3 4

Date

May 17th May 15th May 19th June 3 rd

Setting temp. difference of controller

No controller +o.2° c

0°C + o.2°c

Operating angle of expansion valve

operated manually 90 degree 90 "

120 "

737

III-21

/""°"'=-�--�>"-----�'··· temp measuring resistance - I I ll

accumlafor:.

pu/np out

""ice tank upper surface

ice tank

; to compressor ;.-----

liquide header

Fig. 3. Schematic drawing of the con trolling device for an accu­mulator of ice ma­king plant

--+�+--+--+-+--11-��-f--f--f--�-f--+--+--+--+--+--+--+--+-t----1 Liquid temp before E V

Time hr

1----t-- ·· -1-- .. 1--�t:::1--. r-:-==-._:_J____ _ _ _ ..______ __ ..____ .. . _i.__ __ .____ ··· --i '--+-1 -'----+--+---�of 8 !::::

0 0 � Fig. 5 -- Time hr

Figs. 4 and 5. Results of measurements

738

1:

:

� LU � QI ... � - QI .Q

._____ ci.. E QI ... ._____ '1:J ·:; .!? �

-.J \'-.-

: ��

( I

,__ I

I ( cP � : � ) ' +

' I I : -I 1

I 1-.� i ., ,.. ,

i t I ·: I I I ; I I II I : I I i � I I

1 1: I : i'. I

I� I I I i I

I j ! : l I i: I

. , . .r I i i : \ : I i � l l ; : ; I \ ' i � \ I\ .�f,:� . i'. I l:·, ; 1 ; l� t I ,,. . I

. ; II

r ' . ' > I I 0 : I

: Ii ' · . .. � i : ! l I j. I ( I : I

o ' I

t=' . I ; 1

!

:

cP · � -

0 I...

o/. ..(::

0

006

01

00 II

00 ll

0 ti

III-21

"' .., i:: ., s � ::i "' "' ., s " 23 "3 "' ., � '° bi>

ii;

739

III-21

S. CONCLUSION

Through experiments our automatic liquid level controller has shown satisfying performance on both sides of safety and efficient operation.

1 . We can easily regulate the liquid level in the accumulator by adjusting the indicator which denotes the temperature difference between two sensitive elements and by opening the expansion valve properly. This cannot be done with the float valve.

2. The controller is connected to the sensitive elements by electric wires, so it can be installed at any place around the machine and without complicated refrigerant piping.

3. Operation of the controller is very sensitive, e. g. ± 0.2° C. 4. Owing to the small heat capacity of the elements, we can achieve quick action of

controlling mechanism. S. The range of temperature difference between two sensitive elements is between

-S° C and +S°C. Therefore, we can expect reliable operation of a refrigerating plant at lower superheat of the suction gas even if the liquid heat in the evaporator is high.

SUMMARY OF THE DISCUSSION (PAPERS IIl-39 + III-21)

A. Neuenschwander, France : The values given in the two last columns of the Table 1 seem to have been calculated without taking into account the possible fouling of the heat transfer surfaces, which is however included in the equation (3) in the term

F1

Fr · f

It is evident that the figures indicated in the two columns would be quite significantly modified if one took into consideration the fouling factor, as one invariably has to do in practical applications.

J. Lorentzen, Denmark : Mr. Neuenschwander is quite correct in assuming that the figures given are a simplification and have not taken into account the fouling factor. The reason for this is, that Table 1 as such was only written out in an introductory manner and really has little to do with the actual graphs and calculations which were later discussed more fully.

H. W. Fischer, U. K. : I would like to ask Mr. Lorentzen about a reference which is made on page 4 to Fig. 2, but which I think really means Fig. 4. In reference it is stated that the recirculated liquid from the riser side is trapped in the separator and then flows through the orifice. Could Mr. Lorentzen perhaps give an idea as to how the size of this orifice is calculated. It appears that on maximum capacity it will have to be sufficiently large to overcome the static and velocity heads. Then at low capacity the orifice size may be too large to keep any constant liquid level in the separator.

J. Lorentzen, Denmark : The diameter of the orifice is not in fact dependent on the charge of the system but is purely calculated on the amount of internal surface that we want to wet due to recirculation. This is mainly dependent on the configuration of the evapora­tor itself. If, for instance, the evaporator is made from 3 / 4" nominal bore tube, having 10 parallel circuits then that is the deciding factor of dimensioning the orifice. We have made various tests and have established the optimum orifice sizes for various lengths of evaporator surface and for the number of circuits. This is taken as the basis of dimension­ing this orifice.

G. Saint-Girons, France : On page S of the paper it is indicated that the ordinary top feed is less favourable than the ordinary bottom feed whereas in another part of the paper it is stated that the opposite is true in fact.

740

III-21

J. Lorentzen, Denmark : I think that is quite correct, but I have also given the explana­tion for this statement on page 5, namely that in practice one gets inferior results from the bottom feed system because the design of the liquid distribution system is not usually good enough to make it act in an optimum fashion.

J. Kowalczewski,Australia : The operation of the very high rate of liquid circulation in this system is not quite clear from the text. Could Mr. Lorentzen show this cycle on perhaps a pressure enthalpy diagram ?

J. Lorentzen, Denmark : The high circulation ratio system has nothing to do with the circulation system of the plant. Then if we have a cooling system consisting of a separator unit, a liquid pump and some cooling surface then the point of the high liquid circulation system is that the heat transfer coefficient between wall and liquid may be much higher than between wall and boiling liquid of the conventional evaporators, and by this high circulation rate system the heat load is carried by the liquid as sensible heat and is then allowed to enter the flash chamber where evaporation takes place. This has been described in some other references which will be found at the end of the paper.

M. Pertzelan, Israel : In practice we have observed that the influence on the supply of liquid through the expansion valve to the evaporator is very much affected by the fluctuation of load, especially with automatic or semi-automatic capacity regulators. In this connection it should also be mentioned that the condensing pressure has a consider­able effect on the regulation of the expansion valves and therefore on the cooling capa­city at any one load setting.

J. Lorentzen, Denmark : I agree with Mr. Pertzelan that the operation of the expansion valve depends upon so many different factors which have nothing to do really with the evaporator, and therefore, that is a great weakness in this particular type of evaporator construction. Therefore, I would say that perhaps this system should not be used in any evaporator of great importance.

741

Introduction commune pour les deux rapports suivantes (111-36 et -37)

REGULATION ET AUTOMATISMES DANS LES INSTALLATIONS FRI­

GORIFIQUES

Par J. RICHARD et E. BESSE

Dans un tres grand nombre d'installations frigorifiques et, notamment, dans les Entrepots frigorifiques, surtout dans ceux qui comportent de nombreuses chambres, !es besoins du froid sont variables dans de larges proportions.

Pour adapter Ia production du froid aux besoins, tout une serie de moyens sont employes :

- moyens permettant d'obtenir une variation discontinue de la puissance, tels qu'usage de compresseurs multiples, elimination de cylindres, commande par moteurs a 2 vites-ses . . . .

- moyens permettant d'obtenir une variation continue de la puissance; retour partiel des gaz refoules a !'aspiration, variation progressive d'espace mort, laminage des vapeurs aspirees, etc., enfin, variation de vitesse des compresseurs.

Les procedes de regulation continus permettent, de maniere generale, plus de pre­cision que les procedes comportant une variation par paliers, car on ne peut evidemment trop multiplier le nombre de ceux-ci.

La variation de vitesse des compresseurs obtenue, soit en interposant un variateur entre moteur a vitesse constante et compresseur, soit en entrainant directement le compresseur par un moteur a vitesse variable est le plus rationnel des procedes. S'il est assez onereux au point de vue des frais de premier etablissement, c'est le plus eco­nomique au point de vue exploitation.

Combine eventuellement avec !'usage des compresseurs multiples et !'elimination successive des cylindres, il permet de realiser avec un seul variateur de puissance reduite, une puissance continilrnent variable dans des proportions considerables.

Le variateur doit alors permettre de faire varier la puissance des compresseurs qu'il entraine sur une plage au moins egale a l'ecart le plus grand entre les paliers successifs de puissance qui permettent le reglage discontinu. L'automatisme sequentiel intervient alors pour provoquer le passage au palier de puissance suivant lorsque la variation continue arrive en bout de plage.

On voit done que dans une regulation de ce genre, les deux techniques fondamentales de l'automatisme

- la regulation, d'une part, - les automatismes a sequences, d'autre part,

ont a intervenir simultanement. Les deux communications qui suivent, presentees sous les titres respectifs : - les automatismes a sequences dans la production et !'utilisation du froid; - regulation de la production du froid dans les installations frigorifiques,

auront done, pour objet, d'illustrer par un exemple, comment interviennent et sont realises les automatismes sequentiels et la regulation dans une installation frigorifique donnee.

II eut ete souhaitable, avant de decrire l'appareillage utilise dans cette installation et d'exposer son mode de fonctionnement, de dormer quelques indications sur les methodes d'etude qu'on Utilise pour etablir Uil projet d'automatisme OU de regulation.

Ces techniques sont assez originales et utilisent des methodes de calcul qui sont plutot affaire de specialistes. En donner, ne fut-ce qu'une idee sommaire, dans le cadre qui nous est imparti, est absolument impossible.

743

111-36 et -37

Nos communications ont done ete redigees comme si un certain nombre de notions etaient acquises qui ne le sont peut-etre pas encore par tout le monde. II en resultera, sans doute, que certains points paraitront obscurs et que certains raisonnements parail­ront arbitraires.

Nous tenons, cependant, a la disposition des lecteurs qui souhaiteraient mieux comp­rendre, une note oil est expose, de maniere tres condensee et tres simplifiee, ce qu'il est suffisant de savoir pour comprendre comment a ete etudie l'automatisme de !'instal­lation, tres simple, choisie comme exemple et comment a ete analyse le fonctionnement de sa regulation.

744

III-36

Regulation de la production du froid dans les installations frigori­fiques

Refrigeration Control in Refrigerating Plants

J. RICHARD S. E. D. R. A., 10, rue de Chartres, Neuilly-sur-Seine (France)

SUMMARY. The problem of the continuous and automatic operation of refrigerating plants is, at the present, generally achieved by intermittent methods (unloading of compress­ors, multi-speed motors) through direct acting devices.

These devices are not always uniform, resulting in some complications particularly with multiple compressors and evaporators. Furthermore they do not provide economical operation.

The refrigeration industry has, however, benefitted from control equipment already in use in other fields; signal detectors, proportional, integral and possibly differential controllers and, for performance, a compressor speed controller.

The input signal to the controller depends on the type of plant. Plants with several evap­orators require suction pressure control. Variation of refrigerating capacity by the variation of compressor speeds also raises some problems, particularly with multi-compressor systems and when it is desirable to equip each with a speed controller, the control device proper should then be supplemented by sequence automation.

Some applications of continuous refrigerating capacity control in terms of predetermined requirements have made it possible to test the performance of circuits and to arrive at certain deductions.

I - PRINCIPE DE LA REGULATION DE PRODUCTION DU FROID D'UNE

INSTALLATION FRIGORIFIQUE (entrep6t)

La Regulation a pour objet de maintenir, de fa<;:on aussi precise que possible, une grandeur physique a une valeur donnee, <lite valeur de consigne, celle-ci etant, en general, fixe OU quelquefois variable clans le temps (regulation a programme).

Dans le cas d'une chambre frigorifique oil serait entreposee une denree de nature homogene, la grandeur a maintenir constante est la temperature optimale d'entreposage de cette denree.

En fait, etant donne les inconvenients qu'il y aurait a mesurer la temperature de la denree elle-meme, on se borne a maintenir constante la temperature de l'air clans la chambre, ce qui n'est d'ailleurs parfaitement admissible que lorsque les denrees sont deja, a peu pres, a temperature d'entreposage et, de plus, ne sont pas le siege d'impor­tants degagements de chaleur.

Quand il s'agit d'un entrep6t a chambres multiples comportant des chambres a temperatures diverses, on pourrait evidemment envisager, pour caracteriser les besoins de froid, de mesurer les ecarts de ces temperatures d'atmosphere des chambres par rapport a leurs valeurs de consigne propre et de faire la somme de ces ecarts en les af­fectant de coefficients qui soient proportionnnels a la puissance frigorifique necessaire pour obtenir, clans le meme temps, un meme abaissement de temperature clans chaque chambre, clans les conditions de !'instant considere. Cette methode de ponderation des besoins de froid individuels des chambres serait evidemment tres compliquee a realiser car le chargement des chambres varie au cours de !'exploitation et les conditions d'en­vironnement dependent largement des temperatures ambiantes, des chambres adjacentes, etc . . . .

On a cherche a simplifier radicalement le probleme en affectant chaque chambre d'un coefficient fixe une fois pour toutes et qui est proportionnel a sa charge thermique

745

111-36

evaluee dans des conditions moyennes d'exploitation. C'est assez arbitraire et peu logique.

Une solution qui reste simple, mais plus satisfaisante, consiste a regler a une valeur fixe la temperature, ou, ce qui revient au meme, la pression d'evaporation dans les evaporateurs de toutes les chambres fonctionnant a une meme temperature (ou a des temperatures voisines) et qui font partie d'un meme circuit frigorifique possedant sa regulation propre de production de froid. La transmission du froid des evaporateurs, oil la temperature d'evaporation est ainsi reglee a valeur constante, aux denrees, sera reglee par thermostat d'ambiance commandant par tout ou rien !'injection de fluide frigorigene aux evaporateurs et la marche des ventilateurs.

A ce stade, on admet done que la production du froid, en vue de maintenir constante une temperature d'evaporation dans les chambres est du domaine de la regulation progressive, mais que son utilisation dans les chambres est du domaine de la regulation par tout ou rien.

L'avantage du systeme, par rapport a la regulation directe en fonction des temperatures de chambres affectees de coefficients de ponderation pre-determines est, que la mise en temperature des denrees est obtenue le plus rapidement possible sans qu'il soit necessaire de passer, pour cela, en commande manuelle ou de changer les coefficients de ponderation pendant la mise en regime et, cependant, la protection contre un refroi­dissement excessif est assuree.

Pour etre complet, il nous faudrait done etudier, d'une part, la regulation de type continu de la production du froid, d'autre part, la regulation discontinue de son utilisation dans les chambres. 11 n'est pas possible de traiter, meme succinctement, ces deux sujets en un seul expose. Aussi, la presente communication ne traitera-t-elle vraiment que de la regulation continue de la production du froid.

Toutefois, afin de ne pas completement passer sous silence la question de l'automatis­me de !'utilisation du froid, rappelons tres sommairement que la variation de la tempe­rature e des denrees ainsi que la temperature de !'atmosphere interieure dans une cham­bre froide, refroidie par frigorifere a temperature constante, est assez bien representee par la solution d'equations differentielles du 2 eme ordre a coefficients constants et sont de forme:

Ki e r1t + K2 e r2t + 61

ri et r2 etant tous deux negatifs, Ki et K2 etant des constantes dependant des condi­tions initiales et @r etant les temperatures de denrees OU d'atmosphere interieure qui seraient atteintes au bout d'un temps infini. La loi du rechauffement, quand le frigori­fere ne fournit plus de froid, est donnee par la meme equation differentielle que celle du refroidissement, mais dans laquelle le terme representant le froid fourni par le fri­gorifere est supprime et dont la solution est egalement une somme de deux exponen­tielles.

La constante de temps des thermostats etant, en general, negligeable aux vitesses de variation des temperatures de refroidissement et de rechauffement des chambres, on deduit aisement, des formules etablies, la courbe d'evolution de la temperature d'air a l'interieur de la chambre qui revet la forme d'arcs d'exponentielles en dents de scie qui s'inscrivent entre les deux temperatures d'enclenchement et de declenchement du thermostat. La courbe de variation des denrees a la meme allure mais elle est d'amplitude moindre et dephasee par rapport a la premiere.

On trouvera, dans les communications que nous avons faites anterieurement, notam­ment au IX° Congres, a propos de la pre-refrigeration, l'essentiel des methodes de calcul a utiliser pour etablir ces courbes de variation de temperature reglees par tout ou rien.

II - SCHEMA FONCTIONNEL DE LA BOUCLE DE REGULATION

Au contraire de la regulation de !'utilisation du froid, la regulation de la puissance frigorifique sera modulante et sera commandee, de fa<;:on continue, par les ecarts que la pression d'aspiration subira relativement a sa valeur de consigne.

746

111-36

L'ecart en question, convenablement mesure, commandera, par l'intermediaire d'un appareil regulateur, le variateur de vitesse qui entraine le compresseur.

De la variation de vitesse du compresseur resulte une variation de debit du frigorigene, d'ou, dans !es evaporateurs, une variation de pression qui, evidemment, doit tendre a annuler l'ecart mesure, et la boucle est ainsi fermee.

Tra<;:ons, pour !'installation simple consideree, le schema fonctionnel de la boucle de regulation.

e c o. r t �

z: tert$ton i 111oqe d.e. la � r i u 1 o n

6 cvop•rt.tiori

rwu�i-itr£ t;*U"i°" '----� & rwil""""t

Reg u l o teur

d&bit Evopon:ott1.1•t--.,...--- tompms \tOllJme

Donnons, successivement, !es caracteristiques des elements de la boucle qui permet­tent d'etablir la transmittance de chacun:

a) - Variateur: C'est un variateur electronique que nous ne decrirons pas en detail.

II suffit de savoir que c'est un appareil qui permet, a partir de l'energie foumie, a puis­sance Constante, au Stator d'un moteur triphase a bagues et transmise a de foibles pertes pres au rotor, d'en reverser au reseau d'alimentation une certaine proportion variable, cependant que le reste est utilise dans le rotor sous forme d'energie mecanique. Le prelevement d'energie electrique dans le circuit rotorique pour retour au reseau cree un glissement variable.

Les variations de glissement, done de vitesse de l'arbre du rotor, sont rendues propor­tionelles par asservissement electronique a la tension de commande de l'amplificateur d'entree.

De fa<;:on a proteger l'appareil contre les surintensites dans le circuit rotorique, en cas de brusque augmentation de la tension de commande, si la regulation venait a exiger une tres grande acceleration du compresseur, cette tension de commande est reglee par la rotation d'un potentiometre commande par un petit moteur relativement lent puis­qu'il permet de passer de l'arret a la vitesse maximale (1320 t/min.) en 50 secondes.

C'est Ia une limitation a la linearite du systeme car si le regulateur se trouve place dans des conditions telles qu'il vienne a commander une variation de vitesse de plus de 1320 50• soit 26,4 t/seconde, le signal de sortie du regulateur est sature. II tombe cependant

sous le sens qu'avec l'inertie bien connue des evaporateurs frigorifiques on ne doit pas risquer de voir le regulateur commander un taux de variation de cet ordre meme avec un regulateur tres nerveux. C'est ce qu'on pourra, d'ailleurs verifier dans la suite.

Finalement, a cette restriction eventuelle pres, la transmittance du regulateur peut etre, en negligeant la Constante de temps d'asservissement electronique de la Vitesse du variateur, consideree comme constante et egale a :

1320

300- = 4,4 t/min. degre d'angle de rotation du potentiometre. Ce dernier ayant

une rotation de 300°.

b) - Compresseur : Le compresseur a puissance continilment variable, a, a son regime moyen d'utilisation,

une puissance frigorifique de 140.000 frigories/heure. Son debit volume est de 180 m3/h. a 1320 t/min.

747

IIl-36

L . d d . 1 180

a transmittance u compresseur est one ega ement une constante egale a 1320

= 0,136 m3 t/min.

c) - Evaporateurs: Leur grandeur d'entree est le debit volume V des compresseurs ; leur grandeur de

sortie, la pression d'evaporation Po. Les variations de la grandeur d'entree s'identifient done avec les seules variations de debit volume du compresseur a vitesse variable ; on a, entre V et B0 la relation :

d B o V A. q o t h = C dt + .; o (Ba - B o)

A. etant le rendement volumetrique, q o th la production frigorifique theorique volumetrique C la masse calorifique des evaporateurs .; 0 le flux thermique unitaire des evaporateurs B a la temperature de l'air dans les chambres B 0 la temperature d'evaporation.

On sait que q ot h depend de Bo et aussi de la temperature de sous-refroidissement et que A. depend du taux de compression. I1 en resulte que A. depend, de B o et de la tem­perature de condensation; mais, on peut, sans grande erreur, considerer que, dans les installations normales ou le condenseur est assez largement calcule, la temperature de condensation ainsi que celle de sous-refroidissement, varient peu avec la charge ther­mique et sont pratiquement constantes.

D'autre part, faisant, a priori, confiance a la regulation, on peut admettre que, puis­qu'on regle Po done Bo cette derniere grandeur est effectivement sensiblement con­stante. Done, en premiere approximation, A. q ot h est constant et on peut ecrire :

d B o C at - .; o B o - .; o B a = k V avec k = A. q ot h

En ne considerant que !es variations concommittantes de V et B o et en utilisant les notations complexes, il vient :

k V = C dd�o

+ .; o B o

V V jwt avec = o e

et 0 o = G Vo e j (wt + 1P)

d'ou la transmittance G ej<p

= --�---­C jw + .; o

k que nous ecrirons :

C .; 0 I + ;Jw

c etant la Constante de temps des evaporateurs .

.; 0 C et .; o dependent du nombre de chambres en service, mais, remarquons toutefois

c que le rapport .;

0 en est sensiblement independant.

En effet, les evaporateurs sont tous a peu pres du meme type et un evaporateur moyen de 200 m2 qui pese 1 .300 kg et contient 200 1. d'ammoniac, a une masse calorifique de 360 calf° C.

Sa constante de temps est done, en admettant un coefficient d'echange de 10 cal/m2 h. ° C :

3600

2000 0,18 heure, soit a peu pres : 1 1 minutes.

D'ou la transmittance de l'evaporateur avec k = 780 frigories / m3

748

III-36

780

� o 1 + 1 1 jw � o pouvant varier dans de tres larges proportions selon le nombre de chambres en

service.

d) - Pressiometre C'est un appareil a tube Bourdon qui commande un potentiometre a microfriction.

On peut considerer que le curseur du potentiomerre suit, sans retard, les variations de pression et que ce pressiometre n'a pas de constante de temps.

Au point de vue de sa transmittance, ii nous est plus simple de l'englober dans le regulateur lui-meme, car nous avons directement mesure la variation de la grandeur de sortie du regulateur en fonction de la variation de la mesure du pressiometre en deg­res de temperature d'evaporation, sans passer par l'intermediaire de la tension d'entree au regulateur.

e) - Regulateur C'est un appareil classique a 3 actions : proportionelle, integrale et differentielle, qui

positionne le moteur du potentiometre de commande du variateur de vitesse en envoyant, a ce moteur, par l'intermediaire de relais, des impulsions dont la duree et la frequence sont relies qu'un fonctionnement lineaire est approche de fa9on acceptable.

Comme pour le pressiometre, on englobera, par la pensee, le moteur et le potentio­merre de commande du variateur dans le regulateur.

La grandeur de sortie etant !'angle IX de ce potentiometre et la grandeur d'entree, l'ecart e de la pression d'evaporation par rapport a sa consigne, le fonctionnement du regulateur P. I. D. est caracterise par !'equation en variables complexes :

t IX = a e + b/� d t + c � f

en posant toujours

-- jwt e = S o e

0

- j (wt + cp) IX = G e o e

ii vient la transmittance :

que nous allons ecrire :

IX b = = a + . + cj w e J w

[c (jw)2 + ajw + b] j w Les coefficients a, b et c sont, par essence, positifs et, en general, sauf cas exceptionnel

dont nous ne parlerons pas ici, le trinome; c (jw)2 + aj w + b a des racines reelles qui, obligatoirement sont negatives. Soient z1 et z2 ces racines, on a :

soit :

C Z1 Z2 [ 1 + �z�] [ 1 + �J]

1 1 posons [z1] = r, et [z2J = •2 (ces grandeurs sont homologues a des temps).

La transmittance du regulateur est alors :

(1 + r1 j w) (l + r2 j w) j w

b Remarquons d'ailleurs que le produit Z1 Z2 = - et que, par consequent, le produit c

749

tii-36 c Z1 Zz est egal a b, constante d'action integrale homogene a l'inverse d'un temps. La transmittance du regulateur peut done se noter :

(1 + T1 j w) (1 + T2 j w) b . - -.---- ----J W

A la mise en marche de l'installation en Janvier 1963, nous avons realise un reglage de regulateur tel que, quand on lui envoie a l'entree un ecart maintenu de 0,4 kg /cm2, soit tres sensiblement, dans la zone de reglage, 4° de temperature, le regulateur delivrait une premiere impulsion de 10 secondes qui traduisait l'action proportionelle, suivie indefiniment d'impulsions de 2,5 secondes, espacees de 7 secondes et demie, qui tradui­saient l'action integrale.

La rotation du potentiometre de commande du variateur etant de 360° minute. L'action proportionnelle etait pour 1 ° d'ecart de temperature :

1 360.10 4 � = 15 degres d'angle par degre d'ecart de temperature d'ou : a = 15.

60 Le temps d'action integrale etait de : 2,5

7,5 = 20 secondes par minute, ce qui

donnait une rotation de 120° /min. Soit, pour 1 ° d'ecart : 30°/min. d'ou b = 30. Quant a l'action derivee, on a releve qu'une vitesse de variation de pression correspon­

dant a 1 ° C. en 6 secondes, avait le meme effet que l'action proportionelle due a 1 ° C d'ecart, soit une variation de 15° d'angle du potentiometre.

Il en resulte que, pour une variation de 1 ° C par minute, le potentiometre varie de 1,5°, d'ou c = 1,5.

La transmittance du regulateur est done :

1 [1,5 (jw)2 + 15 jw + 30]

j w

qui peut s'ecrire : 30 . (1 + O,l38 jw) (1 + 0,362 jw) J w

Maintenant, nous pouvons done ecrire le schema fonctionnel avec les transmittances de tous les elements :

- + • ;'W .. ' r .. ' j W I zS (i ousj ,)(1 •uljo>)I ()(

Re� u t e> t a. u r

� .. (&,.) .l6..Q _1 -_ v 0,0 6 r ... 4 -E.. h-11J'c.>

v i' opo rat .,r,s ' Com "\elH' vorlotevr

III - ETUDE DE LA STABILITE ET DE LA PRECISION DE LA REGULA­TION

a) - Stabilite La transmittance en boucle ouverte est:

soit :

750

30 · 4,4 · 0,136 · 780 (1 + 0,138 jw) (1 + 0,362 jw)

g o j w (l + 1 1 jw)

14000 (1 + 0,138 jw) (1 + 0,362 jw)

g o jw (l + ll jw)

Dans !'installation decrite, i1 y avait :

- 6 grandes chambres avec - 9 chambres moyennes avec - 2 petites chambres avec

� o = 3.000 cal/h ° C � o = 1 .500 cal/h ° C � o = 750 cal/h ° C

Done, le flux thermique total peut atteindre: au maximum 33.000 cal/h °C quand toutes les chambres sont en service, et,

tii-36

au minimum 1.500 cal/h ° C quand une chambre moyenne ou les deux petites sont en service (le cas d'une seule petite chambre en service n'etant pas a prendre en consideration car extremement exceptionnel).

Le coefficient de la transmittance en boucle ouverte peut done varier entre 0,425 et 18,65.

Si n chambres se trouvent en service a un instant donne, la mise hors service d'une des chambres deja en service ou la mise en service d'une chambre supplementaire creera une perturbation dont les consequences, au point de vue stabilite de la regulation, ne seront pas les memes selon que n'est grand ou petit.

1122

2/22 _

6/22

12122

22122 _

SE-ule la courbe di!' Goin

relotivl!' (I 1 /22 de lo $Ul"­foce- "de.s ivaporoteura

Kl rrpl"e..s."ntee Pour lt'1.out�courb�

de Goin on a s1mpll':T1tont Figure. � osymptotH qui

perml"tte-nt de dif1n1r Jo

pulsation de coupurE- a­v« une prK1!.10n surF1 -

Fig. r . Diagramme logarithmique de gain et de phase de la boucle de regulation

Le diagramme logarithmique de gain et de phase (Fig. 1) represente les courbes logarithmiques de gain et de phase de la boucle de regulation pour : - une seule chambre moyenne en service ou les soit : 1 /22 de la puissance totale

2 petites chambres - une seule chambre en service ou 2 moyennes - 3 grandes chambres en service ou tout autre

combinaison de chambres - les 6 grandes chambres en service ou tout autre

combinaison de chambres - la totalite des chambres en service.

Les pulsations de coupure sont, respectivement

soit : 2/22 de la puissance totale soit : 6/22 de la puissance totale

soit : 12/22 de la puissance totale

0,92 0,65 0,375 0,270 et 0,195 rad/minute.

Il en resulte des marges de phase comprises entre 33° et 22°, ce qui est insuffisant car, pour assurer une stabilite rigoureuse, il faudrait, au moins, 45°.

De fait, on constate sur les graphiques d'enregistrement de la pression, que la stabilite n'est pas parfaite. C'est notamment visible sur la Fig. 2 ou on constate, qu'apres

751

IIl-36

Ia coupure des compresseurs correspondant aux heures de pointe E. D. F. la pression de consigne n'est retrouvee qu'apres une oscillation assez marquee.

Il faut cependant noter que le systeme reste quand meme a l'abri du pompage car, si la marge de phase est faible, les diverses pulsations de coupure correspondent a des frequences tres faibles comprises entre 12 et 1,6 cycle par heure. Aussi, en dehors des cas tels que la remise en marche des compresseurs, apres coupure E. D. F. avec une pression qui s'est beaucoup ecartee de la pression de consigne, on ne constate que peu d'oscillations de la pression reglee, !'amplitude de ces oscillations etant de l'ordre de 100 grammes/cm2 au maximum et leur periode trop grande pour qu'elles soient vraiment genantes.

Cependant, il sera evidemment souhaitable de modifier ce premier reglage et d'aug­menter !'action proportionnelle; l'ideal, au point de vue stabilite serait, theoriquement, de la multiplier par 20, sans toucher a !'action derivee qui, dans notre cas, ne sert prati­quement pas et est, d'ailleurs, negligeable.

La transmittance du regulateur deviendrait alors :

30 � (1 + lOjw) JW

ce qui ferait pratiquement disparaitre la pente-12 decibels par octave de la courbe de gain et amenerait une excellente stabilite. On aurait, en outre, l'avantage de raccourcir les temps de reponse qui sont un peu trop longs, puisque variables de 3 a 15 minutes ainsi qu'on peut le deduire des frequences de coupure et que la Fig. 2 permet de le verifier (15 minutes a forte charge a la fin de la coupure E. D. F.).

: �--·-

.

z 0

� � :;! . il ii' .. E � � 0 . ·-� � . z 0 .

5 � � L

' ' lo;9/<"' l

Fig. 2. Diagramme des pressions et des temperatures

On peut cependant dire que, meme au point de vue du temps de reponse, le reglage realise a la mise en service de !'installation, en hiver a faible charge moyenne, est quand meme acceptable, puisqu'apres les petites pointes de la courbe de pression regulee qui correspondent chacune a des mises en service de chambres, on constate que la consigne est rattrapee chaque fois en 3 ou 4 minutes (Fig. 3).

752

111-36

( MARC HE

YEC REGULATION

l \ Q't.U1rt1Q l N• 2 �

� � . c � . � � . . . 0 . :: '

;, c w � . . . .. ., ,,...

Fig. 3. Diagramme des pressions et des temperatures

b) - Precision.

Elle doit etre envisagee a un double point de vue. - par rapport a la consigne: On sait qu'il ne peut y avoir d'ecart residue! apres stabili­

sation puisqu'on utilise Uil regulateur a action integrale dont la Constante de temps d'integration a d'ailleurs ete, des les premiers essais de Janvier 1963, reglee assez court

15 (soit

30 = 1/2 minute). De fait, apres action du regulateur, la pression se stabilise a

sa valeur de consigne sans ecart appreciable. - par rapport aux perturbations dues aux variations de temperature ambiante des

chambres. L'equation de fonctionnement du frigorifere etant:

d B o V A q oth = - C dt + � o (B a- B o)

on voit qu'une variation sur B a equivaut a une variation egale et de signe contraire sur

la grandeur reglee Bo. Or, le rapport de l'ecart complexe s a la temperature reglee

complexe B o est !'inverse de la transmittance en boucle ouverte. Dans ce rapport, le terme d'integration, j w figure done au numerateur. Par consequent, lorsque, a la stabili-

sation, jw tend vers 0, ; 0 tend egalement vers 0 et, par suite de !'integration du regula-

c:; 0 teur, il n'y a pas d'ecart permanent du aux variations de temperature ambiante des chambres.

IV - CONCLUSION

Nous avons applique a un exemple simple de regulation d'installation frigorifique, les methodes utilisees par !es specialistes en Regulation. On con�oit, a partir des expli­cations que nous avons donnees, comment ces methodes peuvent s'appliquer a l'etablis­sement d'un projet de regulation, au reglage d'une installation realisee et a !'analyse de son fonctionnement.

A vrai dire, dans l'exemple choisi, point n'etait, au fond, besoin de tant en dire pour expliquer le fonctionnement d'une regulation que, mieux que tout, la comparaison des Figs. 3 et 4 permet de juger.

753

111-36

La Fig. 3 represente la variation de la pression en regulation, avec le reglage encore imparfait, ainsi qu'on l'a vu, realise a la mise en marche de !'installation en Janvier 1963. On y constate les pointes correspondant a la mise en service de chambres generalement rattrapees en 3 ou 4 minutes. Les paliers de stabilisation de pression ne presentent pas d'ecart superieur a 100 g/cm2, soit une precision du demi degre sur la temperature d'evaporation. On constate egalement qu'un changement du point de consigne a ete fidelement suivi.

La Fig. 4 represente la variation de pression d'evaporation en conduite manuelle, celle-ci etant faite par un mecanicien qualifie assurant un horaire de presence normal

dans un entrep6t frigorifique, !'amplitude des variations de pression est de 1 kg/cm2,

soit une variation de 10° sur la temperature d'evaporation. Aussi, pour ne citer que celui-Ia, on con9oit l'avantage d'une telle regulation pour

le maintien du degre hygrometrique dans des chambres qui, dans le cas particulier, sont des chambres a fruits qui beneficient ainsi d'excellentes conditions d'entreposage (on a, a ce propos, verifie que !'introduction de 20 tonnes de fruits chauds dans une chambre, contenant deja 30 tonnes de fruits refroidis, d'avait pas fait varier le degre hygrometrique, fixe auparavant a 85 %, de plus de 2 %).

Si, cependant, nous nous sommes un peu etendus sur certains aspects theoriques de la Regulation, c'est que, pour des installations moins simples, on risque de gros me­comptes a s'en tenir au vieil empirisme pour la conception des systemes de regulation, le choix de l'appareillage et le reglage des installations. Ce sera le cas des installations frigorifiques a qui sont imposes des programmes de regulation severes telles que les

\

� � M A RC HE

A S REGULATIO

c.•.A. " auc ""• 1 J I Q

t ..

; r' .. \ . . - • •c

: . .

Fig. 4. Diagramme des pressions et des temperatures

installations de simulation d'ambiance, celles de traitements thermiques en genie chimique, en metallurgie, de conservation de denrees dans des conditions particuliere­ment precises, des appareils d'essais de laboratoire, etc . . . .

Dans des cas de ce genre, i1 faut savoir qu'il est possible aux specialistes de prevoir exactement, a l'avance, la regulation qui convient pour realiser le programme et de la regler avec le minimum de tatonnements.

754

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Les automatismes a sequences dans la production et l'utilisation du froid

Sequence Automation in the Production and Use of Refrigeration

E. BESSE S. E.D.R.A., 10, rue de Chartres, Neuilly-sur-Seine (France)

SUMMARY. Industrial production and the use of refrigeration necessitates, unless the plant is a simple one, a given number of processes for cold production and distribution. This sequence of operations can be altered, interrupted or reversed when certain operation magni­tudes have been reached, or under certain problematic conditions.

This is referred to as automotive programming. Whilst automotive programming is comparatively simple, it is possible, with some ex­

perience, to directly establish an automotive scheme representing the control devices (push buttons, switches, thermostats, pressostats, etc.) motive power and relays (motors, motorized valves, etc.), thus establishing a liaison between these two categories of equipment.

If the program is more complicated - in the case of large modern cold stores, ice-! actories or in special applications ( i. e. ambient simulation) - calculation is necessary in order to draw up an automotive scheme, but these calculations should be applied with discrimination if it is desired to obtain the simplest of schemes.

Examples of computed schemes actually applied in practice have enabled a study of se­quence automation.

Les operations a realiser automatiquement pour commander les appareils dans les installations frigorifiques font partie d'un programme. Plusieurs programmes peuvent se succeder suivant un ordre bien etabli.

Ces programmes ou, sequences de fonctionnement, doivent evidemment satisfaire a differentes conditions qui, normalement, doivent pouvoir etre exprimees plus ou moins commodement dans le langage courant.

Lorsque ces conditions sont clairement exprimees, il appartient aux Ingenieurs charges d'etablir les schemas electriques de fonctionnement de rendre plus concis le texte qui exprime le fonctionnement de ces sequences.

A la limite de la concision, le texte d'une sequence se limite a une formule eta1'lie a partir des regles de l'algebre logique.

Or, ces formules ayant pour representations graphiques, les schemas de fonctionne­ment eux memes, on peut dire que l'algebre logique a permis a l'Ingenieur de transformer un texte, exprime en langage courant, en un schema de fonctionnement.

Ce procede etant reversible, i1 devient evident qu'un schema, fourni a un Ingenieur, peut etre transforme en langage courant pour retrouver le programme qui a ete a la base de son etablissement.

A l'aide de cette discipline, on peut done affirmer que, sauf erreur materielle, l'automa­tisme realise est parfaitement conforme aux instructions fournies.

Jusqu'a present, l'etablissement d'un schema etait reserve a des specialistes qui, mentalement, traduisaient les programmes qui leur eraient donnes pour executer imme­diatement les schemas : or, i1 se trouve qu'actuellement, avec les complications qui accompagnent les installations automatiques, ces specialistes, eux-memes, sont par instant depasses et realisent leur automatisme qui n'est pas forcement celui qui leur a ete demande.

Bien qu'il ne soit pas dans !'intention des auteurs de cette communication, de faire un rappel des elements de base de l'algebre logique, i1 a semble utile de rappeler les conside­rations elementaires qui ont permis l'etablissement de cette discipline.

Les appareils couramment utilises dans l'Industrie, sont toujours accompagnes d'une partie qui est destinee a etre raccordee aux bomes d'une source d'energie electrique.

755

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Si l'on fait abstraction du temps pendant lequel s'etablit le courant, a l'interieur du circuit, on s'aper�oit qu'il n'y a que deux etats stables, a savoir : l'appareil est sous-ten­sion ou hors-tension.

De meme, nous pouvons considerer les organes de commutation qui sont des contacts electriques qui ouvrent ou ferment un circuit.

Si, a l'image de ce qui a ete admis pour le raccordement des appareils au reseau, on neglige le temps pendant lequel un circuit se ferme ou s'ouvre, on s'aper�oit que ces organes de commutation ne peuvent prendre egalement que deux positions stables, a savoir : ils sont ouverts ou fermes.

Sous ces reserves, il a ete permis d'etablir une algebre qui ne tient compte que de deux etats stables.

Par convention, on admet qu'un circuit, repere par une lettre, sera affecte de la valeur 1 ou 0, selon qu'il sera sous tension ou hors tension ; parallelement, un circuit sera affecte de la valeur 1 ou O, selon qu'il sera ferme ou ouvert. L'algebre logique repose sur ces bases.

Le cas particulier qui est traite comme un exemple concerne l'equipement d'un Entre­pot frigorifique con� pour conserver a 0°C, dans une ambiance a 85 % d'humidite rela­tive, une quantite de fruits evaluee a 1000 tonnes qui peuvent etre reparties dans 17 cham­bres de contenances inegales.

Par contre, chaque chambre est equipee de fa�n a pouvoir refrigerer de + 25 a + 3 ° C, en 16 heures, 20 tonnes de fruits. La refrigeration globale de l'Entrepot est limitee a 200 tonnes par jour.

Le bilan thermique de cet Entrepot a permis d'etablir qu'il fallait disposer d'une puissance nominale de 350000 Frigories/H en tenant compte de la manipulation des 200 tonnes d'arrivages journaliers.

Par contre, en periode de stockage, la puissance minimale necessaire au fonctionnement de la chambre la plus petite a ete estimee a 15000 Frigories/H. La puissance frigorifique doit done pouvoir evoluer de 3, 7 % a la totalite.

II a ete egalement demande que la pression d'aspiration soit maintenue a une valeur de consigne choisie par la Direction de l'Entrepot, en fonction des besoins de refrigeration ou de stockage. En refrigeration, il faut disposer d'une grande puissance tandis qu'en stockage il faut pouvoir maintenir une humidite relative correcte.

Pour realiser ce programme, i1 a ete decide d'employer deux compresseurs ayant les caracteristiques suivantes et de reguler la pression d'aspiration en agissant sur la vitesse du compresseur ayant la puissance la plus faible.

756

11£1 rl'• < l"o.Jh ll Cl>HU, /ti '" u .. , ,..lllf JI{ <01'1"111'"' """' �" li!llAJ .. A

Fig. r . Diagramme des phases

IIl-37

Compresseur A : Puissance nominale a 1500 t/min. Nombre de cylindres pouvant etre mis en service Vitesse variable comprise entre

1 10000 Frigories/H 2 - 4 ou 6 600 et 1320 t/min.

Compresseur B : Puissance nominale a 1000 t/min. 275000 Frigories/H Nombre de cylindres pouvant etre mis en service 2, 3 ou 4. Vitesses possibles 500 ou 1000 t/min.

Les puissances utilisables sont done les suivantes, en appelant P la puissance reelle correspondant au regime de marche impose par !'Exploitation qui «affiche » une pression d'aspiration de consigne en fonction des besoins immediats.

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Fig. 2. Les schemas representent la traduction des formules (voir page 765)

757

IIl-37

Compresseur A

1 Vitesse minimale

Vitesse maximale 2 cylindres en service

f Vitesse minimale

) 4 cylindres en service

I Vitesse maximale

f Vitesse minimale

) 6 cylindres en service

l Vitesse maximale

Compresseu1 B

Petite vitesse et 2 cylindres en service

Petite vitesse et 3 cylindres en service

Petite vitesse et 4 cylindres en service

Grande vitesse et 3 cylindres en service

Grande vitesse et 4 cylindres en service

600 2 2

p A 1500 -6 15

PA

PA 1320 2 4,33 ·-- -·- = �- P A 1500 6 15

600 4 PA

1500 6 4

15 p A

1320 4 8,66 PA I500 6 = 15 PA

600 6 p A

1500 G 1320 6

PA 1500 6

500 2 PB

1000 -4 500 3

PB 1000 4 500 4

PB 1000 4

1000 3 PB

1000 4 1000 4

PB 1000 4

6

IS PA

13

15 PA

2 - PB 8

3 8

PB

4

8 PB

6

8 PB

8

8 PB

Pour la commodite de !'expose, on a appele, pour le Compresseur A, allure de marche, le nombre de cylindres en service ; par contre, les allures de marche du Compresseur B sont une combinaison du nombre de cylindres en service et la vitesse a laquelle il est entraine. Les allures de marche de chacun des deux Compresseurs sont done les suivantes :

Compresseur A

Allure n° 1 Allure n° 2 Allure n° 3

Compresseur B Allure n° 1 Allure n° 2 Allure n° 3 Allure n° 4 Allure n° 5

2 cylindres en service 4 cylindres en service 6 cylindres en service

2 cylindres en service a petite vitesse 3 cylindres en service a petite vitesse 4 cylindres en service a petite vitesse 3 cylindres en service a grande vitesse 4 cylindres en service a grande vitesse

Si l'on donne a p A et a PB les puissances nominales respectives des compresseurs A et B, on remarque que la puissance distribuee par le compresseur A, en allure 1, a vitesse maximale, est superieure a celle que est distribuee par ce meme compresseur, en allure 2, a vitesse minimale.

On remarque egalement que les paliers de puissance du compresseur B sont tous infe­rieurs a la puissance qui peut etre mise en service par le compresseur A entre !'allure 1, a vitesse minimale, et !'allure 3, a vitesse maximale.

758

III-37

Ces dispositions ont ete prises afin d'eviter tout pompage et de s'assurer qu'il y avait au

rnoins une combinaison d'allure entre les deux compresseurs qui permet de trouver la puissance strictement necessaire aux besoins de l'Entrepot.

Le jeu de la regulation a pour but de faire varier la puissance de fai;:on continue entre

2/15 PA et 13/15 PA + PB, soit en puissance nominale de 14700 Fg/h a 385000 Fg/h. La puissance minimale represente done seulement 3,8 % de la puissance maximale.

La regulation de la vitesse du compresseur A a pour obj et de creer !'appoint de puissance

strictement necessaire au maintien de la pression reelle d'aspiration a la valeur de pression de consigne ; cet appoint est positif lorsque la vitesse du compresseur A croit et negatif quand la vitesse decroit.

Pour exposer plus commodement le fonctionnement de !'installation, nous considerons,

en premier lieu, que !'installation est a l'arret, c'est-a-dire que toutes !es chambres ont atteint leur valeur de consigne.

Lorsque l'une quelconque des 17 chambres atteint une temperature ambiante superieure

a la valeur de consigne, le thermostat d'ambiance ferme un circuit electrique qui a pour role de faire demarrer le compresseur A en allure 1 a 500 t/min. On peut egalement imaginer que plusieurs chambres se mettent en marche simultanement.

Quel que soit le nombre de chambres qui se mettent en service simultanement, le compresseur A demarre toujours en allure 1 a 600 t/min.

Ce n'est que lorsque le compresseur A a termine son demarrage que la regulation

commence a jouer son role.

Si la puissance distribuee par le compresseur A, en allure 1 a 600 t/min. est trop foible, la pression d'aspiration est plus importante que celle qui est affichee. Le regulateur comman­

de alors l'accroissement de la vitesse jusqu'au moment ou la vitesse est maximale, soit 1320 t/rnin.

Ce n'est que lorsque la vitesse a atteint sa valeur maximale que le compresseur A passe, si la pression reelle est encore superieure a la puissance de consigne, de !'allure 1 a !'allure 2.

A cette nouvelle allure de marche, la pression reelle d'aspiration peut etre superieure, egale ou inferieure a la valeur de consigne ; si cette pression est superieure, le compresseur A passe de !'allure 2 a !'allure 3.

Si la pression est egale, le compresseur reste a !'allure 2 ; si la pression est inferieure, la vitesse decroit jusqu'au moment ou l'egalite entre la pression reelle et la pression de consigne est atteinte.

Supposons que la pression d'aspiration soit encore superieure a la valeur de consigne, le compresseur, en allure 3, passe en allure 4.

Cette allure 4 est une allure fictive puisque le compresseur A est deja a sa puissance maximale en allure 3, mais c'est un moyen qui a ete adopte pour permettre le demarrage du compresseur B en allure 1 .

Les conditions de fonctionnement qui ont ete enumerees pour le passage de !'allure 2 a !'allure 3 du compresseur A restent valables pour le passage de !'allure 1 a !'allure 2 du compresseur B, ce qui revient a dire que le compresseur B peut passer d'une allure «n » a une allure «n + 1 » a la condition que le compresseur A soit en allure 4 a vitesse maximale et que la pression reelle d'aspiration soit superieure a la pression affichee.

Lorsque les deux compresseurs sont en marche a leur allure maximale, la puissance en service est maximale et la regulation deviertt sans objet.

Supposons, maintenant que !'installation soit en service avec le compresseur A en allure 4, a vitesse maximale et le compresseur B est en allure 5.

Si, a un moment donne, une ou plusieurs chambres se mettent a l'arret, la pression d'aspiration decroit et devient inferieure a la pression de consigne.

Le role de la regulation est alors de reduire la vitesse du compresseur A, tout en laissant le compresseur B en allure 5 ; ce n'est que lorsque le compresseur A est en allure 3, a vitesse minimale, qu'il passe en allure 2.

De proche en proche, et toujours a la condition que la pression d'aspiration soit inferieure a la valeur de consigne, le compresseur A revient en allure 1 a vitesse minimale.

759

III-37

Ce n'est qu'aux deux conditions suivantes :

1°) - compresseur A en allure 1, a vitesse minimale, 2°) - Pression d'aspiration inferieure a la valeur de consigne,

que le compresseur B passe de !'allure 5 a !'allure 4. De proche en proche, et toujours aux deux memes conditions, le compresseur B passe de

!'allure 4 a !'allure 3, puis de !'allure 3 a !'allure 2 ; de !'allure 2 a !'allure 1 et, enfin, s'arrete.

Le compresseur A ne peut etre arrete qu'a la condition que routes les chambres soient satisfaites.

Par contre si, a un moment donne, routes les chambres s'arretent simultanement, les compresseurs A et B s'arretent immediatement, quelle que soit leur allure de marche.

De routes fai;:ons, le prochain demarrage aura lieu par la mise en route du compresseur A en allure 1, a vitesse minimale.

Pour realiser ce programme, nous avons a notre disposition : - Le contact A que le regulateur ferme si la pression d'aspiration est superieure a la

valeur de consigne ; - Le contact B que le regulateur ferme si la puissance d'aspiration est inferieure a la

valeur de consigne ; - Le contact X que l'organe d'asservissement de la vitesse ferme quand la vitesse a

atteint sa valeur maximale ; - Le contact Y que le meme organe d'asservissement de la vitesse ferme quand la

vitesse a atteint sa valeur minimale.

TOUS CES CONTACTS SONT INDEPENDANTS LES UNS DES AUTRES

Nous avons done, a notre disposition, deux couples de contacts qui peuvent etre ecrit sous la forme d'un circuit de coincidence. XA 1 lorsque le compresseur A est a vitesse maximale et que la pression d'aspiration

est superieure a la valeur de consigne ; YB lorsque le compresseur A sera a vitesse minimale et que la pression d'aspiration

sera inferieure a la valeur de consigne. XA et YB ne peuvent pas prendre la valeur 1 en meme temps mais l'un comme l'autre

peut etre maintenu egal a 1 pendant tout le temps necessaire au passage des allures jusqu'a ce que X ou Y s'annule.

Ceci est incompatible avec le fonctionnement ; puisque ces couples de contact ne materialisent pas le passage d'une allure a l'autre ; ii a done fallu transformer ces contacts maintenus en une succession d'impulsions permettant d'utiliser Jes cycles de deux con­tacts independants, l'un erant asservi a AX et l'autre a BY.

La periode de ces impulsions a ete limitee a 15 secondes ce qui represente le temps minimal necessaire au passage d'une allure a une autre dans le sens direct ou retrograde.

Ce temps a ere predetermine, en tenant compte de la transmittance moyenne des cir­cuits. Le programme permettant de realiser le schema se resume done a celui du dispositif capable de compter un certain nombre d'impulsions dans le sens direct si les allures de marche des compresseurs doivent passer d'un rang «n » a un rang «n + 1 » ou retrograde si ces allures doivent passer d'un rang «n » a un rang «n - 1 ».

Pour realiser ce dispositif, nous avions a choisir entre les differents codages des nume­rations binaires ; Jes machines a calculer, aussi complexes soient-elles, utilisent generale­ment le code d'Aiken qui presente l'avantage de dormer a chaque nombre compris entre 0 et 4 un code symerrique a celui des nombres compris entre 5 et 9 ; Exemple :

Le code du nombre 2 etant 0010

Le code du nombre 7 sera 1 101

II a ere imagine pour permettre aux machines comptables d'utiliser des decades. Par contre, etant donne que, dans ce systeme, le codage des nombres de 0 a 4 inclus

s'effectuent en numeration binaire pure et que, parailleurs, le nombre d'impulsions necessaires au passage de !'allure du compresseur A est limite a 4, le codage d'Aiken se

760

111-37

confond, dans ce cas particulier, avec celui du codage en numeration binaire pure. Par raison d'homogeneite le meme codage a ete retenu pour !'elaboration du dispositif destine a passer les allures du compresseur B necessitant 5 impulsions.

Le diagramme des phases (Fig. 1) reproduit le codage binaire des impulsions fournies par les contacts cycliques.

ID (sens de comptage direct) IR (sens de comptage retrograde)

La lecture du diagramme des phases permet de remarquer qu'il est indispensable de disposer de relais de memoire d'impulsion et de relais codeurs afin d'avoir suffisamment de combinaisons de contacts pour commander les relais necessaires au comptage des im­pulsions sans en sauter une ou sans compter deux fois la meme.

L'application des regles de l'algebre logique a ce diagramme permet d'ecrire les formu­les de commutations suivantes relatives a chacun des relais utilises.

La transformation graphique de ces formules permet d'etablir les schemas suivants qui sont la traduction du programme suivant.

- «Imaginer un dispositif capable d'additionner algebriquement des impulsions emanant de deux sources independantes ; la somme des informations enregistrees etant limitees a 4 pour un premier dispositif et a 5 pour le second. Chaque nombre, de 0 a 4 pour le premier et de 1 a 5 pour le second dispositif, disposera d'un circuit permettant l'allumage d'une lampe et la commande d'une vanne electromagnetique. La puissance de chacun de ces circuits est limitee a 80 VA sous 24 Volts alternatifs» .

Calcul des formules de commutation Les methodes a employer etant toujours Jes memes, nous traiterons le cas de dis­

positif de comptage du compresseur B qui doit pouvoir enregistrer 7 impulsions, alors que celui du compresseur A n'en enregistre que 3.

Relais de memoire d'impulsion (sens direct) Remarquons que la mise sous tension de A est asservie a ID et aux combinaisons des

circuits de c, d, e, et f. II est auto entretenu par Jes memes combinaisons que celles qui Jui ont permis d'etre sous tension.

Pour ecrire ces combinaisons, nous retiendrons celles d'entre elles dont la vie est la plus longue.

A = i b d f + i a d f + i b c f + i a c f + i b d e + i a d e + i b c e + + i a c e.

A = i (a + b) (d f + c f + d e + c e)

I A = i (a + b) (c + d) (e + f) I Le calcul de la formule de commutation de C sera traite evidemment de la meme fa;:on,

mais en remarquant que Jes contacts e ou f suffisent a l'auto entretien.

C = i a b d f + i c f + i a b e d + i c e

C = i (a b d + c) (e + f)

1 -c = i (a b d + c) (e + f) I Le calcul de la formule de commutation de E sera traite de la meme fa�on mais en

remarquant que l'auto entretien peut etre realise par le relais E lui-meme.

E = i a b c d f + i e I E = i ( a b e d r -+� On peut traiter le prob!eme de fa�on differente en appliquant tout simplement Jes

regles de l'algebre logique a tous Jes cas rencontres au cours d'une commutation ou de son entretien.

A = i b c d e f + i a b c d e f + i a c d e f + i a b c d e f + + i b c d e f + i a b c d e f + i a c d e f + i a b c d e f + + i b c d e f + i a b c d e f + i a c d e f + i a b c d e f + + i b c d e f + i a b c d e f + i a c d e f + i a b c d e f

A = i (b c d e f + a c d e f + b c d e f + a c d e f + b c d e f + + a c d e f + b c d e f + a c d e f).

761

III-37

A = i (b c d e + a c d e + b c d e + a c d e)

A = i (b c e + a c e)

De la meme fm;:on, on trouve :

IE = i a c (b d f + e) I I A = i c e (b + a) I

I C = i a e (b d + c) I Ces formules sont equivalentes et conduisent au meme resultat. On peut done choisir,

entre elles, pour conserver celle qui conduit au schema le plus simple ou au nombre de contacts le plus reduit. On peut egalement fixer son choix sur la formule qui assure la meilleure continuite du circuit et la securite de fonctionnement la plus grande.

Relais de memoire d'impulsion (sens retrograde)

Comme pour les relais precedents, on remarque, sur le diagramme des phases, que la mise sous tension de G, H ou J est asservie a I R et aux combinaisons de contacts des relais B D et F. II est, en effet, inutile de chercher une formule de commutation dans laquelle interviendraient les contacts des relais A, C et E puisqu'ils sont asservis a I n et qu'il est impensable que I R soit a circuit ferme en meme temps que I n. On ne peut tout de meme pas enregistrer simultanement une somme de deux informations de valeur iden­tique et de signe oppofe.

On peut done ecrire :

G = irbdfhj + dfghj + bdfhj + dfghj + bdfhj + dfghj + bdfhj + dfghj

l��hfCb+g)I De la meme fac;:on on trouve

et

Relais codeurs

Les memes methodes s'appliquent a l'etablissement des formules de commutation des relais codeurs.

Un remarque sur le diagramme des phases que B est mis sous tension par A, H ou J, qu'il est auto-entretenu par une combinaison des contacts repos des relais C, D, E et F, mais qu'il peut etre mis au repos par le contact travail de G.

Cette demiere condition pose un probleme qui peut etre resolu par l'un des trois moyens suivants :

I 0) - On calcule la formule de commutation sans tenir compte de I' action de G et au moment de !'execution du schema on represente un contact travail de G aux bomes du relais B pour le courcircuiter ;

2°) - On calcule le complement de la formule ci-dessus a laquelle on ajoute un con­tact travail de G ;

3°) - On calcule immediatement une formule de commutation en ne tenant compte que des combinaisons conduisant a une mise hors tension de B.

A cette formule est associee celle qui permet la mise sous tension.

1°) - Formule de commutation sans tenir compte de G B = a + h + j + b (df + cf + d e + c c e)

I B = a + h + j + b (c + d) (e + f) I Z0) - Formule comptementaire de B accouptee a G

B = a + h + j + b (c + d) (e + f) + g i B = a hf(b + cd + �f5+gl 3°) - Formule comptementaire etablie en tenant compte de G

Pour cela il faut remarquer sur le diagramme des phases que B est mis hors tension par une combinaison des contacts des relais C, D, E et F et par la fermeture de G. On peut done ecrire :

762

B = c e f + c d e + c e f + g = c e + c d e + g = c + d e + g

III-37

Cette formule est, evidemment, beaucoup plus elegante que celle precedemment ecrite et pourtant elle doit lui etre equivalente. Cela tient a ce que la premiere contient des ter­mes redondants provenant de la complementation de la premiere formule qui contient des termes dont le role est superflu dans la commutation recherchee.

En effet, pour que B soit mis hors tension, il faut que :

1 °) - A = 0 H = O et J = 0 dont le produit a h j est redondant puisqu'il ne peut pas en etre autrement

2°) - B = I sans cela on n'aurait pas a le mettre hors service,

done b est rendondant a l'interieur de la parenthese

3°) - Si C = I D = I au bout d'un certain temps, done C ou D, seul, suffit.

4°) - Si E = I F = I au bout d'un certain temps, done e au f, seul, suffit.

Ces reflexions demontrent que la seconde formule est valable et peut etre retenue sans crainte d'etre en presence d'une erreur de calcul.

On peut done associer a cette formule complementaire, la commutation necessaire et suffisante a la mise sous tension

I B = a + h + i + b I B = e + d e + g 1 --- · · ·------

Tous calculs faits, on trouve Jes formules de commutation suivantes pour les relais D et F

I D = c + i + d j n = e + h

1 : : �_:_

f

��'

Les formules qui ont ere etablies permettent d'executer le schema qui sera confie a l'Entreprise chargee de !'execution du montage.

Pour exemple, deux schemas peuvent etre executes :

Le schema 1 etabli en rentenant Jes formules suivantes :

A = id (b + a = (c + d) (e + f)

B = a + h + j + b (c + d) (e + f' C = id (a b d + c) (e + f)

D = c + j + d (e + f)

E = id (a b c d f + e)

et B = g

et D = h

F = e + f et F = j

G = ir h j (b + g)

H = ir g j (b d + h)

J = ir g h (b d f + j)

Le schema 2 etabli en retenant les formules suivantes :

A = ir c e (b + a)

B = a + h + j + b et B = c + de + g

C = ir a e (b d + c)

D = c + i + d et D = ef + h

E = id a e (b d f + e)

F = e + f et F = j

G = ir h j (b + g)

H = ir g j (b d + h)

J = ir g h (b d f + i)

763

111-37

Le schema joint reproduit cette solution pour l'un comme pour l'autre compresseur.

Decodage

Pour ecrire les forrnules de commutation des circuits de commande comprenant les vannes de reduction de puissance, les lampes de signalisation et les vitesses de rotation pour le compresseur B, on dispose des programmes suivants :

Compresseur A

2 vannes de reduction de puissance 3 lampes de signalisation

ST - sous tension HT - hors tension AL - allumee ET - eteinte.

Rang de VANNES

I' impulsion N° 1 N° 2

0 ST ST 1 ST HT 2 HT HT

3 HT HT

LAMPES DE SIGNALISATION

Allure 1 Allure 2 Allure 3

AL ET ET ET AL ET ET ET AL ET ET AL

En appelant B et D les relais codeurs du dispositif de comptage des impulsions du compresseur A, on trouve, en se reportant au diagramme des phases :

V1 = b d + b d = d

V2 = b d

Li = b d

L2 = b d

L3 = b d + b d = d

Compresseur B

2 vannes de reduction de puissance 55 lampes de signalisation

2 allures de vitesse : PV et GV

HT - hors tension ST - sous tension AL - allumee ET - eteinte

Rang de VANNES VITESSE

!'impulsion N° 1 N° 2 P V G V

0 HT HT HT HT

1 ST ST ST HT

2 ST HT ST HT

3 HT HT ST HT

4 HT ST HT ST

5 HT HT HT ST

764

LAMPES DE SIGNALISATION

All. 1 All. 2 All. 3 All. 4 All. 5

ET ET ET ET ET

AL ET ET ET ET

ET AL ET ET ET

ET ET AL ET ET

ET ET ET AL ET

ET ET ET ET AL

IIl-37

En appelant B D et F les relais codeurs du dispositif de comptage des impulsions du compresseur B, en se reportant au diagramme des phases, et en supposant que le disposi­tif est bloque a 5, on trouve :

V1 = b d f + b d f = f (b d + b d)

V2 = b d f + b d f = b d L1 = b d f L2 = b d f L3 = b d f

L4 = b d f

L5 = b d f

On peut simplifier et ecrire

PV = b d f + b d f + b d f = f (b + d)

GV = b d f + b d f = d f

I v, = r Cb d + b d) I I V. = b d i I L, = bdfl I L2 = d f l I L3 = b d l I L3 = b f j lLa = d f l

Les schemas joints representent la traduction de ces formules (Fig. 2).

765