STUDY OF SCRAPED-SURFACE HEAT EXCHANGER ICE-MAKING …€¦ · 1.3 ûrganization 2.1 History 2.2 Ice Making Application 3.1 Heat Transfer in Scraped-Surface Heat Exchanger 3.2 Heat

A STUDY OF SCRAPED-SURFACE HEAT EXCHANGER IN ICE-MAKING APPLICATIONS

Alexander C hi-To C hong

A thesis submined in conformQ with the requirwients for the degree of Master of Applid Science Graduate Depamnent of Chernid Engineering and Applied C h e m i .

University of Toronto

O Copyn'ght by Alexander Chi-To Chong 2001

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Master of Applied Science 700 1

Alexander Chi-To Chong

Graduate Department of Chemical Engineering and Applied Chemise University of Toronto

The performance of an industrial scnped-surtàce ice generator. supplied by

Sunwell Technologies Inc.. has been investigatcd both expenmentall> and numericaily.

The aluminium ice genentor consisted of multiple retiigerant passages sumiunding a

centrai scraped-surface channel. in which an ice-slurry was made from a brine solution.

A computer model was developed to simulate the steady-state operation. This

simulation model consisted of three sections: i) computational mesh generation. ii)

steady-aite parameters evaluation. and iii) radial temperature presentation.

Tests on the ice generator were performed within a broad range of operating

conditions. both with and without ice formation. and the data were compared with the

simulation results.

The model could predict the esperimental heat exchange rates and tefigerant

pressure drop data with standard deviations of =14?/0 and 110%. respectively. Since the

mode1 could give a reasonable prediction. the radiai temperature profiles were examineci

in detail to understand the degm of temperature variation on the scraped-sdace.

I wodd like to take this opportunity to thank for those who have helped me in the pmgress of this thesis. Especially,

Prof. Masahiro Kawaji as the thesis supervisor

Mr, Vladimir GoIdstein as the supporter of the current research

Dr. Ming-Jian Wang as the research supervisor

Mr. Roger Castonguary for assistance in the testing faciiities

Ms. Connie Zhang as my love

Abstract Acknowledgernents Table of Contents List of Figures List of Tables

** 11

iii iv vi ix

1 .1 Background 1.2 Objectives and Scope 1.3 ûrganization

2.1 History 2.2 Ice Making Application

3.1 Heat Transfer in Scraped-Surface Heat Exchanger 3.2 Heat Transfer in Forced Convective Boiling 3 -3 Pressure Drop in Two-Phase Flow 3.4 Properties of lce SIurry

4.1 Heat Exchanger Design 4.2 Experimental Apparatus 4.3 Test Procedure 4.4 Data Evaluation

5.1 Cross Section of Heat Exchanger 5.2 Process Simulation 5.3 Temperature Distribution (cross-section)

6.1 Results Experirnents Simulations

6 2 Discussions Cornparison of Experimental and Simulated Results Simulations Experimental Observations Final Remarks on the Simulation Mode1

7. CONCLUSIONS AND RECOMMENDATIONS 90

7.1 Conclusions 7 2 Recommendations

Nomenclature

References

Appendix A : Correlation and Experimentai Emrs

Appendix B : Computer codes

Appendix C : Experimental resuits

Appendix D : Simulation d t s

Figure

2.1

2.2

3.1

3.2

3.3

4.1

4.2

4.3

4- 4

4.5

4.6

4.7

4.8

4.9

5.1

5.2

5.3

5.4

5.5

5.6

5.7

6.1

6.2

6.3

6.4

Description

Illustration of a typical scraped-surface heat exchanger

Generai types of interfacial crystaI structures

Illustration of the ice generator

Parameter F in Chen's correlation

Parameter S in Chen's correlation

Scraped-surface heat exchanger (cross section and flow channels)

Photograph of the heat exchanger column tested

Photograph of the heat exchanger cross section

Photograph on the testing apparatus

Generai experimental test layout

Photograph of the rotor stiaft with scrapers

Measurements on test loop

Process of heat exchange rate detemiination

Temperature change dong heat exchanger

Structure of simulation mode1

Process of heat transfer rate evaiuation

Process of heat and mass bdance evaluation

Condition for correlating the ice M o n

Illustration of cooiing process h m brine to ice slurry

Heat exchanger with n axial elements

Process of pressure gradient evaluation

Variation of temperature ùi brine channel

Variation of ice fiadon in brine cbannel

Variation of salt concentration (liquid phase) in brhe channel

Variation of temperaîure in ceEgerant cchanels

Page

LIST OF FIGURES (cont'd)

Figure

6.5

6.6

6.7

6.8

6.9

6.10

6.1 1

6.12

6.13

6.14

6.15

6.16

6.17

6.18

6.19

6.20

6.21

6.22

6.23

6.24

6.25

6.26

6.27

6.28

6.31

Description

Variation of pressure in refigerant channels

Variation of vapour quality in refiigerant channels

Variation of heat transfer rate in each fluid channe1 (ice making mode)

Variation of heat transfer rate in each fluid channe[ (chilliig mode)

Wail temperature profiles (brine channel)

Wall temperature profiles (refiigerant channel a)

Wall temperature profiles (refrigerant channel b)

Wall temperature profiles (refiigerant channel c)

WalI temperature profiles (refiigerant channel (i)

WalI ternperature profiles (refngerant channel e)

Wall heat flux profiles (brine channel)

Wall heat flux profiles (refiigerant channel a)

Wall heat flux profiles (refiigerant channel 6)

Wail heat flux profiles (refiigerant channel c)

Wall heat flux profiles (refiigerant channel d)

Wail heat tlux profiles (refigerant channel e)

Cross sectionai temperature profile (element I )

Cross sectional temperature profile felement 4)

Cross sedona1 temperature profile (element II)

Cross sectional temperature profile (element 20)

Cornparison of brindice slurry exit ternperature

Cornparison of the predicted and measured exit ice fiaction

Comparison of refiigerant exit pressure

Comparison of refiigerant exit temperature

Cornparison of refiigerant exit vapour quaüty

Page

vii

LIST OF FIGURES (cont'd)

Figure Description

6.32 Cornparison of refigerant pressure drop

6.33 Comparison of heat exchange rate

6.34 Camparison of change in refrigerant pressure (resr run 19)

6.35 Effect of rotationai speed (vertical orientation)

6.36 Effect of rotationai speed (horizontai orientation)

Page

Table Description

5- 1 Process variables in simulation

6.1 Range of experimental parameters for simulation

6.2 Summary of experimental results for simulation

6.3 Range of overail heat transfer coefficient

6.4 Generai summary of simulation results

Page

INTRODUCTION

An ice-slurry can be used for air-conditioning of large and small buildings,

cooling h h fish aboard fishing vessels, and refngeration in supermarkas. among

others. For these applications, a large amount of ice crystds has to be not only produced

continuously, but aIso automatidiy removed. There are few industrial ice generator

desigus that are commercially avaiiable yet. One such ice generator is a scraped-surface

heat exchanger king developed by SunweIl Technologies, Inc. of Woodbridge, Ontario.

This heat exchanger is made of an extnided aluminium column, with multiple passages of

cefigerant surounding a central brine channel for ice making.

One of the major advantages of ice-slurry compared to conventional ice storage is

the ease of transportation where it wi be easily punped. In cornparhg with chilIed

water system, ice slurry has a higher themal storing unit and a steady kezing

temperature. Other advantage would be its liquidized f o n where it can cover completely

the surface of the materials king cooled, hence, providing a better cooiiig rate. tn

addition, an ice sliny machine operates more efficiently compared to the conventionai

ice machines since the scraped-surfke rnechanism provides higher rate of convection on

the heat trander surface.

The objectives of the present study were to experimentally detamine the

performance characteristics of Sunweli's scraped surface heat exchanger aizd to better

understand the heat transfer characteristics by deveioping a numerical simulation model.

capable of predicting the steady state operation under different inlet conditions of the

working fluids. Since multiple reûigerant channels surround the centrai cooling channel

(like a fînned heat exchanger), the wall temperatwr profiIe may not be uniform. A non-

uniform temperature distribution on the channel inner wall is ûelieved to be responsible

for ice crystal buildup and fieeang up of the scraper blades rotating inside the brine

channel. Thus, an understanding of the temperature distribution dong the heat uansfer

surfixes is an important objective in this study.

To this end, a set of cornputer codes were written to predict the boundary

temperature and heat flux profiles at different axial locations of the heat exchanger. In

addition to the boundary temperature and heat flux profiles prediction, the codes are able

to generate the cross-sectional temperature pmfües and the local conditions of the

working fluids (brindice sluny and refiigerant).

Testing on the scraped d a c e heat exchanger was conducted to obtain data that

could be used to veri@ the cornputer model. The heat exchanger was conaected to a

customized hot gas bypass refiigeration unit, Operathg skills were acquired to gain MI

control of the system, in order to achieve various operathg conditions for testing. The

heat exchanger was instrumenteci for data recording and cali'brations were performed on

the rneasuring instruments. A data acquisition system was set up to record the steady

state operating parameters during the tests. The binary solution used for ice m a h g was

sodium chloride solution. Tests were conducted within a broad range of operating

conditions. Seved test nins were chosen and cornparisons wete made between the

experimental and simulation results for vecïfïcation O t the cornputer mode!.

As a preview of the whole pmject. the organization of the current thesis is

presented in the following paragraphs.

h "Chapter 2: L i t e r m e Review ", the history of the scraped-surface heat

exchangers is presented. This summary is based on the pubiicaîions fiom the past to the

most recent. The descriptions of the scraped-surface heat exchangers used in the ice-

making industries are dso provided.

In "Chapcer 3: Theoretical Background". the heat -fer theories on both the

ice-slurry side and the codant side (refngerant) of the heat exchanger are summarized-

The concepts involveci in evaluating the thermal pmperties of ice slurry (a two-phase

solidAiquid mixhue), are a h ptesented,

In "Chaprer 4: Erperimenral Appuraîus and Procedure ", the consiruction of the

prototype heat exchanger and the experimental setup are described. The apparatus, the

working fluids, refiigerant loop, instnimenâation and testing procedure are summarized.

in "Cbpter 5: Simularion Mode!", the development of the simulation mode1

(cornputer code) is presented and expIained in detail.

In "Chaprer 6: Results and Discussions ", both the experimental and simdation

results are presented and c o m p d . The coUected data, experimental observations and

Page 4

the simulation results are discussed in detail. Furthemore, the aitemative research areas

are also discussed.

in "Chapter 7: Conclusions and Recommendations", the fmdings and the

suggestions based on bis research are summarized. Further research directions are

recommended for subsequent work.

Supporting materiais are coiiected in the "Appendices ", UicIudig the computer

code for the simulation model, the themai pmperties of the fluids used, and the Ml

record of both the experimental and simulated results,

2. LITERATURE REVIEW

Scraped-surface heat exchangers (SSHE) were fmt introduced to the industry for

use in applications involving viscous materiais. The scraping action dong the heat

transfer wall disturbs the formation of the boundary layer; moreover, it enhances the fluid

exchange rate between the boundary and the biilk. Hence, the heat transfer coefficient is

greatiy improved due to continuous surface renewal. in Fipire 2.1, a typicai type of

scrapers is shown.

Figure 2.1: Illustration of a typical scraped-surface heat exchanger

in later years, scraped-surface heat exchangers have been used fiquentiy in

crystallization processes due to their ability to remove crystais h m the h a t transfer

surface. Generaiiy, high super-cooiing of the fluid on the heat mns5n surface will result

in certain undesired crystai growth properties like deadritic structure, and scaiing

problems. The continuous movement of saapers can prevent the crystais h m growing

Page 6

- -

on the wall of the heat exchanger. in Figiire 2.2, the general types of crystal structures

formed during h e z h g are iiiustrated.

Figure 22: General types of interfacial crystal structures

In ment years, the scraped-surface heat exchangers have k e n used in the ice-

making indusûy, where a two-phase homogeneous solidAiquid mixture (icz slurry) is

directiy produced due to the continuous movement of scrapers.

In this section, the research history of sçraped-surface heat exchangers is

describeci.

Huggins (1931) was the first researcher who introduced the idea of scraped-

surface k a t exchangers. His study invo!ved processing of both viscous and non-viscous

materids. During his study, he found that the heat transfer coefficient was eipatly

enhanced with the use of scrapers instead of stirrers in viscous apptications.

Page 7

- -

Hodton (1944) used water as the heat traosfer medium and studied the effets of

flow rate and rotationai speed of the scraping action. Fmm bis tests, he suspected that the

heat transfer coefficient could be a hction of scraping frequency.

Kool (1958), Latinen (1958) and Hariott (1959), al1 came up with a theoretical

model for intemal heat transfer coefficient derivecl h m the penetration theory with

surface renewal. This model predicted well ttie experimental data h m Houlton (1944).

Frorn this model, neither the viscosity nor the axial velocity would influence the heat

m~s fe r coefficient. Hariott confinned that the penetration model could describe the low

viscosity application resuits adequately well.

Skelland (1958) studied the behaviour of a votator, which was a high-speed

scraped-surface heat exchanger. The scrapers he used were not spring-loaded but were

allowed to float with a stagnant liquid layer on the hait exchanger wall. He derived a

general correlation for the heat transfer coefficient using a dimensional analysis. The

exponents for the dimensionless groups were derived empirically fiom the experimental

data. However, this mode1 did not agree well with the data of Houlton (1944). Due to

the lack of agreement, Skelland (1962) later proposeci a more reiiable correlation, which

was in satisfactory

the next chapter.

Penny and

agreement with the penettation theory as d e s c r i i in more detail in

Be11 (1967) suggested that axial dispersion probabiy played an

imporiant role in the data of Skeliand because of the low axial flow rates used.

Tromrnelen (1971) studied flow patterns and heat ma4er properties in a scraped-

surface heat exchanger. He used a viscous medium in his experiments, and developed an

Page 8

empincai mode[ based on the p e n d o n theory. He suggested an empirical multiplier

that accounted for the axial dispersion, incomplete tempetanire equalization and mixing.

De Goede et ai. (1991) studied the heat transfer properties of a scraped-surface

heat exchanger in the turtiuient regime. in their experiments, a sufficiently short heat

exchanger was used where the axial distance was shorter than the diameter of the

exchanger. A CFD mode1 was developed using a commercial code, PHOENICS. for

predicting the development of a vortex between the scraper blades. They aiso developed

a theoretical model based on the approach pmposed by Trommelen (1971).

Bel et al. (1996) studied a scraped-surface heat exchanger for an ice-slurry

application. They accounted for the thermodynamic pmperties of ice slurry in the

development of their model, which was formulateci in terms of both the axial and

rotatiod ReynoIds numbers. The experimental resuIts showed that the influence of

rotational speed on the heat transfer coefficient was not direccly proportional. Moreover,

the influence of axial flow rates in 1amina.r region was ody on the ice production, where

the heat transfer coefficient was not significantIy affécted.

3. TrnORETICAL BACKGROUND

In the following sections, the heat transfer theones required in consmicting the

present simulation mode1 are summarized. The contents inckuded are, a) the evaluation

of convective heat transfer on the brine side for a scraped-surface heat exchger, b) the

evaluation of convective boiiing heat transfer on the freon side, c) the evaiuation of

pressure drop in a two-phase fieon flow (liquid/gas), and d) the evaiuation of the ice

slurry thermal pmperties.

Figure 3.1 : Illwmation of the ice gcnerator

3.1 ~ E A T TRANsm M SCRAPED-SURFACE HEAT EXCHANGER

To account for the enhanced heat transfer on the srraped-surface on the brine side,

previous researchers used two main approaches: ï) a theoretical approach or ui an

empirical approach, The theoretid or semi-theoretical approach involves the

peuetration theory with d a c e renewai. The empiricai approach involves dimensionai

analysis. These two approaches are discussed below.

Page 10

The theory of unsteady heat ûansfér to a semi-infinite solid is applied to the liquid

adjacent to the heat transfer sirrface before king scraped. The i i e interval that the

liquid stays on the surface will be the tirne between two scraping actions. The liquid

previously staying on the surface is removed totally and new liquid Erom the hulk wilI be

redeposited with each scraper pas. Combining the penetration and the sirrfàce renewal

theories, the theoretical equation for the Nusselt number, Nu, which is the dimension~ess

heat transfer coefficient, is given by,

where h, is the scraped-surface heat transfer coefficient, k is the fluid thermal

conductivity, B is the number of scrapers, hi is the rotational speed and D is the internai

diameter of the heat exchanger. The Peclet number. PeR, represents the product of the

Reynolds number and the Prandti number.

However, this theoretical equation oversimplifies the actual phenomena occurring

in the scraped-surface heat exchanger. In most cases, the liquid at the heat m s f e r

d a c e is not weil mixed with the bulk fluid, and in the case of viscous liquids, the fluid

h m the bulk can be partiy cdeposited behind the scraper blades. The hcat transfer

coefficient for viscous fiquids, therefore, is Iess than the prediction h m Eu. (3.1). in

addition, the heat transfer coefficient can also be idluenceci by other parameters such as

axiaI fluid velocity, the diameter and length of the exchanger.

Page I I

Skelland (1958) deveIoped the first empirical correlation, Eq. 13.3), using a

dirnensionaI anaiysis for a votator which is a srnail high-speed (axial fluid vekity) unit.

In his first correlation, he accounted for the liquid viscosity as well as the axial fluid

velociy and the geornetric parameters of the exchanger. However, considerition o f the

number of scrapers was neglected during his derivation of the following correlation.

where h, is the jacket heat transfer coefficient (inner shell), k is the fluid thermai

conductivity, D, is the diameter of scraper ( a h equal to inside dimeter of ihell). v is

the bulk average axiai fluid velocity, p is the ffuid dynamic viscosity, C, is the fiuid

specific heat capacity. IV is the number of scrapers and L is the axial length of the heat

exchanger.

SkeiIand et ai. (1962) then derived another empiricai equation. Eq. '3.4). again

using the dimensional analysis. In this correlation. he included the influence of the ff uid

viscosity in the empiricai constants. In addition to thc: viscosity effect, he &O included

the effect of the number of scrapers. He suggested two sets of empirical constants to

distinguish between viscous Iiquids and thin mobüe Iiquids, however, he did not specif) a

concrete relationship between the heat transfer coefficient and viscosity. ïhe correlation

is given by,

Page 12

where the Reynolds number, Re, is defined with the internai diameter, D,, of the heat

exchanger, Dr is the diameter of the rotor shafi and B is the number of scrapers.

From Skelland's suggestion, the constants a and b in Eq. (3.4) were Iound to be

0.014 and 0.96, respectively, for viscous liquida For thin mobile liquids. the constants

are O. 039 and 0.7.

Since cefigerant was used as the cooIant in the current research heat transfer in

forced convective boiling ne& to be reviewed. In the following, the theories adopted

for constructing the simulation mode1 are summarized.

Chen (1963) sumrnarized the experimental data obtained by other rexmchers for

convective boiling and compared their proposed correlations with al1 the experimental

data available at that time. However, the comlations he examineci could nct provide a

satisfactory result. Hence, he developed a new correlation that could satisfj a broad

range of forced convective boiling heat tramfer data for water and organk liqlid syterns.

His correlation. &en by Eq. (3.51, consisted of both the saturated nucleate boiling heat

transfer coefficient (hVCd and the two-phase forced convection contribution (h3.

h, =hW +hc (3.5)

He suggested that the forced convective contribution couid be represented by a

Dim-Boelter type equation. He also mggesteci tbat the liquid properties shoitid be used,

instead of two-phase mixture pmperties, in the heat transfer terms. in addition he

defineci a parameter. F, to account for the acceleration effect of liquid due to vapour shear

stress.

Page 13

Here, G is the total two-phase

diarneter, and subscripts f and

G(I - X)D, Re, =

Pf G - D,

Re, = - P m

m a s flux, x is the vapour quality, Dh is the hydraulic

p refer to single-phase liquid and two-pke mixnue,

respectively. The two-phase Reynolds number, Re,, is defined using the total m a s flux,

G, the hydrauiic diameter, Di, and the two-phase mixture viscosi~, pm.

Since the parameter, F. given by Eq. (3.9), was defined as a flow parameter. he

expected that it could be expresseci as a function of the Lockhart-Martinelli parameter,

X,, given by Eq. (3.10). The parameter. F. was experimentally determined and it was

presented graphicaily as iliustrated in Figure 3.2.

Figure 3 2 Parameter F m Chen's correlation

Page N

For the nucleate boiling contribution, Eq. (3.11), Chen used the nucleate pool

boiling correlation h m Forster and Zuber's (1955) analysis. in order to account for the

flow effect on the temperature gradients near the wall, which controls nucleation, he

incorporated a nucleation suppression factor, S, into the comlation where the factor

would approach unity at low flows and zero at high flows. Since the parameter. S,

represented the flow effect, it was expected to be a function of the two-phase Reynolds

number, Rep The parameter. S, was experimentally determined and presented

graphically as illustrateci in Figure 3.3. In Chen's comlation. the suppression factor. S.

wodd approach unity at low flows and zero at high flows.

0.79 O 45 0.49

h , = 0.00 122 kr PI ATwO 24 ApwO 75 (s) (3.11) 03. O H 0.24 1 b o s ~ , ik Pr

Figure 33: Parameter S in Chen's correlation

Evaiuation of the pressure drop in the refngerant channels is aiso important in the

current construction of the simulation model. in generai, the total pressure drop, Ap, can

be separated into three components, a) fnctionai, dpf, b) accelerationai, .&,, and c)

gravitationai, dp,, Lockhar-Martinelli (1949) mode1 w-as adopted since it provides

accurate pressure drop estimates in the low mass velocity range (G < 1360 kg/m2-K).

&=Q/ +Qa+Qr (3.13)

Lockhart and Marthelli (1949) correlateci the two-phase friction pressure drop.

dp/, using a two-phase friction multiplier, @'. as follows:

7

&r = Q p ' 9 3 1 - (3.1 4)

Ln this correlation, the two-phase fiction pressure drop is related to the single-

flow aioae in the channel. Altematively, it can a h be correlateci assuming the vapour

phase to be flowing alone.

For smooth tubes, the fanning fiiction factor, f, is caiculated using a Reynolds

number based on the iiquid phase flowing done.

f = 0.079 Re, -02s

Page 16

In the Lockhart-Martinelli comlation, the two-phase fiction multiplier, &, was

defined as a function of the Lockhart-Martinelli parameter, &. The fiction multiplier

comlation is given by,

where c is a constant. Four regimes are dehed based on the behaviour of the flow

(viscous or turbulent) when the respective phases are considered to pass alone through

the channet. Different values of c were recommended for, a) both phases in turbulent

flow (C = 201, b) liquid phase turbulent and vapour phase viscous (c = IC), c) liquid

phase viscous and vapour phase turbulent (c = IZ), and d) both phases viscous (c = 5).

Turbulent condition is achieved when the single-phase Reynolds number exceeds 100.

The accelerationai pressure drop, 4, is evaluated h m a momentum balance

between two points in the channel. Lockhart and Martinelli (1949) derived a correlation

based on a separated flow model, in which liquid and vapour phases are considered to

flow independentiy with their own velocity, in contrast with the hornogeneous rnodel. in

which both phases are a r e e d to flow at the same mean velocity. Using the separated

flow model, the accelerational pressure drop is given by:

where a is the void tiaction and x is the flow quaiity.

Page 2 7

Buttenvorth (1975) suggested a generaiized form, Eq. (3.19), for the void k t ion ,

which encompasses severai frequently used models, depending on the values of A and

exponents p, q and r.

The constants were found to be, A = 0.28, p = 0.64. q = O. 36, and r = 0.07 for the

Lockhart-MartineIli model.

The gravitational pressure &op, &y. (3.20). results from the hydrost~tic head of

the fluid. The change of void fraction due to the heat transfer effect is accowited for by

the integral. The negative sign denotes the pressure recovery, caused by a d o m tlow in

an inclined or verticd pipe.

in order to construct a simulation model or to operate the scraped-surface heat

exchanger described previously, the evaluation of ice slurry pmperties is critical. Bel et

ai. (1996) suggested the following rnethods for evaluating various properties. The

subscripts, îs, î and 1 will be used to cepresent ice sluny, ice, and liquid phase,

rqxcti.ely.

Page 18

Assurning an ice-slurry is an immiscible multiphase mixture, the mixture density

can be expressed as a hct ion of weight fraction and density as foiiows,

where the weight hction and density are denoted by Xand p, respectively.

Jeffrey (1973) suggested an equation for evduating the thermal conductivity of

two-phase solid/liquid mixture. His model is based on a randorn suspension ofspheres in

a liquid phase, and is given by,

ku =k , ( l+3a , f l+3p , , ' f l ' y ) (3.77)

where k and p represent the thennai conductivity and volume Fraction, respectively. The

parameters a, p, and yare evaluated with the following equations,

Bel et al. (1996) developed an enthalpy model for evaiuating the specific heat

capacity of an ice-slurry. This mode[ works only in ice durries produceci using binary

solutions, where the nonaystallizing component wouid be concentrated ir the liquid

phase wheneva ice fraction inmases. Based on the eutectic behaviour of the binary

Page 19

--

sdution, freezing temperature wili shifi as concentration changes in the liquid phase.

Therefore, the fieezing temperature wilI change as the mixture enthalpy changes while

the fluid is undergoing a phase change. The specific heat capacity is then evaluated as

follows,

where cp denotes the apparent specific heat capacity evaluated under constant pressure,

and h and T represent the ceference enthalpy and temperature of the mixture, respectively.

Thomas (1965) developed a relation for two-phase liquidlsolid mi.uture viscosity

in tems of the liquid phase viscosity. This relation assumed Newtonian suspensions of

uniform spheiiçal particles in fluid. Assuming the ice slurry to have the same

characteristics, the viscosity of ice slurry can then be expressed as follows,

p rn =p,(l+2.5pl +10.05~~~+0.00273exp(l6.6~~)) (3.2 7)

where p and p represent the dynarnic viscosity and the volume fraction, respectively.

4. EXPERIMENTAL APPARATUS AND PROCEDURE

The sçraped-dace heat exchanger under investigation was testai over a broad

range of conditions. Descriptions of the exchanger are provided in the next section and

the experimenîal techniques are presented in the rest of this chapter. ui order to

understand the operating parameters in greater detail, a set of cornputer codes was written

to simulate the heat exchange pmess. The development of this simulation model is

discwsed in the next Chapter.

The scraped-dace heat exchanger tested was provided by Sunwell

Tecbnotogies, Inc. This exchanger was made out of an duminum alloy, 6063-T6. and

was extrudeci with a specific cross sectional contiguration. Basicaily, the heat exchanger

consistai of muitiple channels on the outside boundaq for refiigerant circulahon and one

centrai channel for ice-slurry. The scrapers were mounted onto a rotating shaR which

was located at the center of the centrai chamel. ïüe cross section of the hex exchanger

is illustrated in Figure X I .

The exchanger was constructecl with a total of 20 channels for passing a cdd

retiigemt flow. These 20 channeis were made up of 4 ideaticaI circuits, and sach circuit

incIuded 5 chaanels with diiïkrent sbape con6gurations. The cold fluid wa spIit into

four streams and each stream passed h u g h the kat excbanger five times (five channels

wtth different shapes). The cold fluid p a s d b u g h the five channels in a specific

se~uence as indicated in Figure 4.1, with the Ietters h m a to e identwg the channeis.

Page 2 1

a b c d c

Figure 4.1 : Scraped-surface heat exchanger (cross section and flow channeis)

In addition to the shape configuration, each refngemnt channel was constructed

with micro-fins on the inner wall for the purpose of heat trauder enhancement. The

central channel with a surface scraper mechankm provided the passage for the brine or

hot fluid. This was the channel where the ice making process twk place. Pho tographs of

the heat exchanger and its cross section are presented in Figures 1.2 and 4.3.

Figure 4 2 Photograph of the heat exchanger column tested

Page 22

Figure 4.3: Photograph of the heat exchanger cross section

The experimental setup included the test section and test loop as shown in

Figure 4.4. The general testing faciIity is schematically illustrateci in Figure 4.5. A

refngeration loop was used in the experimental setup for ice making application and R22

(dichlorofluromethane) was used as the codant. Brine solution (sodium chloride

solution) was the processing fluid to be cooled. Therefore, the apparatus was constmcted

with two circulating loops, brine circdation and refngeration cycle.

The scraped-sudace heat exchanger was 149.86 cm (59 inches) in ien@h covered

with a 2.54 cm (1 inch) thick insulation material. The scrapers mounted on the rotor shaft

in the heat exchanger were made of dm high molecdar weight polyethylene. The rotor

shaft with scrapers is illustrateci in Figtve 4.6.

Page 2.3

Figure 4.4: Photograph on the testing apparatus

Figure 45: General expxhentai test Iayout

Page 24

Figure 4.6: Photo on the rotor shaft with scrapers

Thennocouples and pressure transducers were installed on the test loop for

recordiig the operating conditions. Calibration was performed on tke pressure

transducers to ensure accurate measurements. A variable area flow meter was used to

measure the flow-rate of brine and a turbine flow meter \vas used for measuring the flow-

rate of refiigeiant. The measuring locations and parameters are shown in Figiirr 4.7 h

addition to the turbine flow meter, an orifice was designed to measure the flowrate of

refrigerant using a differential pressure transducer and this setup was çalibrated using

water. The instruments were comected to a data acquisition system, which was set up to

record the measurements during the tests.

The refiigeration system used in the testing apparatus was a hot gas bypass

system. Due to an unknown heat d e r rate of the heat exchanger before the tests, the

refrigeration system was oversized. The system capacity couid then be regulated with the

use of the hot gas bypass system. Manuai control was required to regdate the system

Page 25

operating conditions, which included the saturated pressure and temperature, the inlet

quality, the ffow rate of reûigerant, the evaporating pressure, etc.

In a storage tank, a batch of brine (sodium chloride solution) was prepared by

rnixing tap water with salt. A temperature controller was used to maintain the brine

temperature. if the temperature dropped below the set point, the controller would initiate

a pump comected between the batch and a plate heat exchanger. The warm water

leaving the condenser of the refngeration unit was continuously passed through the plate

heat exchanger to heat up the brine as it was circulated through the plate heat exchanger.

l r rne Circularion Refrrgeratum Cycle

r-0 @-* Temperature / Flowrate

1 1 A U - Q ~emperature 1 Scraped-

A 1 Brine ' I a Tenperacure 1 Pressure I Flowrate

surface Refrrge- , 1 Hedt ' rarlon @ Temperature 1 Pressure

! Exdianqer

Figure 4.7: Measurements on test loop

The general testing procedure is surnmarized in this section. During each test, the

iüst requirement was to adjust the brine storage temperature by using the rz6igeration

system. Afier setting the temperature of the brine, the refiigerant passing Uito the

evaporator was shut off. The fIow rate of b ~ e passing tbrougb the evaporator was

adjusteci and the cefigerant was introduced back into the evaporator. Moreover, certain

adjustment was required to configure the refiigeration system with the desireci testing

Page 26

parameters. Refrigerant was once again shut off for clearing any ice lefi inside the

evaporator. This was done to make sure there were no ice crystalr Ieft on the heat

transfer surface that could not be removed by the scrapers. Refrigerant was -heu slowly

introduced at the previous flowrate, About one hou. was passed for the systecl to reach a

steady-state condition; however, Longer periods were required in some cases. The data

acquisition system was then initiateci to collect the required parameters for 15 minutes-

Diuing this tirne, three samples of ice slurry at the evaporator exit were collected. A

cidorimetric measurement was made on each sample to determine the ice fraction and the

overail heat transfer rate of the evaporator. h r s due to instnimentation during

experhents are sumrnarized in Appendix A.

For venfication of the simulation model, the overall heat exchange rate is one of

the main parameters of interest. The performance of the heat exchanger can be

characterized by the overail heat transfer coefficient. The procedures for evaluating both

the heat exchange rate and the overail heat transfer coefficient are presented in this

section.

The most important parameter to be detemhed is the heat exchange rdte between

the working fluids (brine and refiigerant) across the channels. This section concerns the

procedure for evaluating the heat exchange rate that involves ice making. The heat

exchange rate is evaluated h m the thermal capacity of products (ice slmksj on the hot

Page 2 7

side (scraped-surface channel) of the kat exchanger. In order to determine the heat

exchange rate, samples of ice slurries were coffected at the exit of the heat exchanger

during the tests.

Figure 4.8: Process of heat exchange rate determination

The procedure for determining the heat exchange rate is illustrated in Figure 4.8.

The ice slurry sample is mixed with a certain amount of warrn water until al1 the ice has

melted. The thermal capacity is then evaiuated by performing an energy balance on this

mixing process. An overall energy balance is given by,

Q, +Q, =Q,

where the energy content of warm water is.

Qw =m. - c P w - ( ~ * -T,)

and that of the fhai mixture,

b Qm = m. - c,, - (w, ) (4.3)

Here, Q represents the energy content, m reptesents the mas, cp represents -Sie specific

heat capacity, and T represents the temperature. The subscripts, is, w, m and ref represent

the ice slurry sample, the added water, the final mixture and the refemce point,

respectively, The Superscnpts, w and b, represent water and brine, respectively.

Page 28

The heat exchange rate, qh, is given by,

4h =4 -Qu/mu

where K, represents the rnass flow rate of brine in the sera@-surface channel.

Combining Eqs. (4.1) to (4.4, an expression for the heat exchange rate resdts as

given by Ey. (4.51. The heat exchange rate, qh, is then evaluated by assigning the brine

idet temperature as the reference temperature, TM

qh = ni, bme; km - 7-4 )- mwc,'k. - Lf 111 mm

The overall heat transfer coefficient, U,, is nonnaliy used to indicate the

perfomance of a heat exchanger in engineering practice. The development of' the overall

heat ûansfer coefficient for the current heat exchanger is presented in the remainder of

this section. The evaluation is based on the ice making operation of the heat exchanger.

An illustration of the changes in temperatures (brine and refngerant) aionp !he

heat exchanger is presented in Figure 4.9. As illutrated the k t transfer pmcess can k

separateci into two regions: i) Region I. the chilling process and ii) Region II. the ice

making ptocess.

Page 29

Figure 4.9: Temperature variation along heat exchanger

In order to perform the analysis of the overall heat transfer coefficient. the

following assumptions had to be made,

0 The super-cwling effect of brine during the transition fiom chilling to ice making

process was neglected.

The boiling temperature of reîiigerant. T, was assumed constant and the average

saturation temperature between the refngerant channel iniet and outiet was used.

0 The ûwing temperature of brine in Region II was assumed constant at the average

temperature in the ice-making region.

Remon Ir Chilling

The heat transfer rate in the chilling region. q,, is given by,

- b qc = mb - c,, - ATc (4- 61

qc =U, - A , .AT, (4.71

where Eb is the mas flow rate of brine. c i is the specific heat capacity of brine, ATc is

the change in brine temperature h m inlet to initial m g , Uc is the overall heat

trarlsfer coefficient in the chilling region. A, is the hait transfer area (scraped-dace) in

Page 30

the chiIlhg region and ATLu is the log-mean temperature difference. The log-mean

temperature difference is defined as follows,

ATLu = (AT AT, ' ) /~ (A~; /AT;) (4 8)

where AT, and ATh' are the temperature ciifferences between the brine and the refkigerant

at the inlet and initial freezing, respectively.

Page 31

Region II: Ice MaRng

The heat transfer rate in the ice making region, qi, is given by,

4, ' 4 , -4, (4- 9)

4, =ut -4 *ATrn (4.1 O)

where q, is the measured total heat transfer rate, U, is the overall heat transfer coefficient

in the ice-making region and A, is the heat transfer area (scraped-dace) in the ice-

making region. The temperature diEerence in the ice making region, AT,, is defmed as

follows,

um = 0.5 - (AT; + AT;) (4. I I )

where dTfiO is the temperature difference between the working fluids at the exit of the

scraped-surface chanael.

Overall Heat Tramfer Coefficient

The overall heat transfer coefficient. Un, is defhed by averiighg the performance

in both the chilling and the ice making regions. The coefficients. Il, and II, in Eqs. (4. ))

and (4.1 O), are used to define the average overall heat transfer coefficient. II,, üs follows.

s = u ~ A & ~ -0 +I/A,)-' +a - (I+ A J ) (4.12)

where A, is the total heat transfer a r a (scraped-surface) and A, is the area ratio between

chilling and ice making regions. The a m ratio is given by Eq- (4.13).

4 = Ac14 = ( q h , )-(a /AT,\, )

SIMULATION MODEL

A set of computer codes for the simulation model was developed consisting of

three individual programs: a) to generate a mesh for the evaiuation of the temperature

profile at each cross-section of the heat exchanger, b) to simulate the steady-state

condition based on the tluid conditions at the iniet, and c) to present the caiculated two-

dimensional temperature profiles graphically. The Listings of computer codes can be

found in ..lppendi.r B.

In order to calculate the heat transfer rate between the working fluids. the

temperature distribution in the heat exchanger cross section must be evaluated. A finite

mesh was then required to perform the calculation of the temperature distrïmtion over

the heat exchanger cross section as shown in Figure. 4. I .

A computer program was written to generate the cross sectional mes11 for use in

the simulation model. This geometric mesh was formed with square mesh in the interna1

portion of the heat exchanger and triangdar mesh on the boundary. The purpose of using

triangular elements on the bounda~~ was to inmase the smoothness around the cwature

driring the transformation of the actuai figure into the h i t e ciifference mesh.

The pmgram accepteci the size of mesh as input and generated the mesh with

diffierent resolutions in some output fües, which stored the information required durhg

the e x e d o n of the simulation program. in addition, mesh with different resolution

could be easiiy chosen during the simulation.

Page 32

5.2 PROCESS SIMULATION

A finite ciifference approach was employed in the simulation model. The heat

exchanger was divided axiaily Uito a finite number of segments. Each segment consisted

of the same axial length and the same cross section.

in order to evaiuate the steady-state operating condition, an iterative approach was

used. Three tasks were perfomed at each stage of iteration: a) heat transfer rate

evaiuation, b) heat and m a s balance evaim'on, and c) pressure drop evaiuation, The

flow chart for the simulation model is ilIustrated ia Figirre 5.1, and the process vaiables

are listed in Tuble 5.1.

Table 5.1 : Process variables in simulation rine ne channel i tb Temperature of f l u i d i XS s a l t concen t r a t ion of l i q u i d 1 XI ice f r a c t i o n l ~ e f r i g e r a n t channels

tc Temperature 1 PZ Pressure x, vapour q u a l i t y

l ~ e a t t r a n s f e r r a t e s 1 gb b r i n e channel I

1 q Ref r ige ran t channels

The main pirrpose of this step was to evaluate the heat exchange rate between the

fluids (brine and refiigerant) in each axial ekment in addition to the heat transfer rate,

the temperature profiles and the boundary heat flux profiles were aiso of in~erest. The

general structure of this subroutine is üIustrated in Figure 5.2, and the variables wd are

as foiiows. The descriptions of the variables are summarized in Tuble 5. I above.

input variables: 4 X* ph xr

Output Vanables: qb q,

Page 33

+ To simulate the steady natc proces of the scraped- surface heat exchanger

End of Simulation u

Pqamtion Stage

+ Operating condition of procrss and q u i & parameters for simulation

+ Predetennination of ammon constants used during ituation

+ Prrparation of initial values for simulation

Tempcrature and heat flux diim%ution gencraticm Rocedurcs for evaluating h e intcnncdiate h a t transfcr rates at each axial elcmcnts

Evnluation of fluid condition at each node in b i h brine and refrigmt channels

Evaluation of pressure drop in i t M g m t channrls Cdculation of fluid pmperties at cach nodc in refrigerant channeis

Check if al1 pnmss variabia have canvqeû

Output staec + Output of pmcess condition 4 Output of t empaam and heat !lux ditniution

Tamination of scraped-surface heat exchanger simulation

I Figure 5.1 : Structure of simulation mode1

Page 31

Temwrature Profile

The convection heat d e r coefficients wae calculated before the evaluation of

the temperature profiie. The averaged fluid properties beâween the inlet and the exit of the

axial segment were used to evaluate the heat ûansîkr coefficients. The thermal properties

of ice s l q were evaluated using the methods descni in section 3-3.k "Propclrties of

Ice Sltmy''. The ernpirical correlation, Eq, (3.9, devetoped by Skelland et al. (1962) was

used to calculate the heat transfer coefficient h the sçraped-surface channel.

The heat transfer coefficient in the refiigerant channels was evaiiiated using

Chen's correlation, Eq. (3.9. for flow boiling heat transfer under two-ghase flow

conditions. To account for the heat transfer coefficient of superheated cefigerant vapour.

the DittusBoelter correlation, Eq. (3. 1)' was used.

Nu, = 0.023 Ftedo8 Rn (5.1)

The constant n was equaI to O. J or 0.3, for heating or cooling of fluid, respectively.

The Gauss-Seidel iterative rnethod was appiied to the evaiuation of the

temperature distribution. However, the nucleate boiling portion of the mode1 included

the terni ( A T ~ ~ ~ ~ ~ A ~ ~ ~ ' * ) ' ) . Therefore, the nucleate boiling tenn rnw be rekshed to a

new vahe at each iterative sep.

Heat Flux Profile (boundmJ and Heat T r d e r Rote

The boundary heat flux pro6ies were evaiuated using the convective heat d e r

coefficient, h, and the temperature diffefence between the heat M e r surface, T,di, and

the buik fluid temperature, Tfid, as fouows.

Evaluate the heat iransfcr rates at each axial elemcnt of the scraped-dkce kat exchanger

Evaluation of the heat transfer nues in axial element N

Hau T m f a coemc1enr l

1 Evaluanon I

6T-6 i .R

i

Evalutacion of the heat transfer cocfticient for convedon (both brine and refngerant channels)

Convergence criteria in 2-D temprranuc profile iteration

Evaluation of steady state 2-D temperarure profile Usage of Guass-Seidel i t d v c appmach Nuclme boiling suppression in r r f i g m t channels

Chcck ifdl the nodal temperatures converge

Hcat Fiux Profiles and H a t Transfm Rate Calcularion

EvaIuation ofthe boundary heat fluxes in both brine and ntkigemt channcis

The heat gain or loss by cach Buid channcl will bc tte summation OC the imundary heat fluxes

Check if the hcat transfer baween the channcls is balanccd If ihc heat transfer is not b a l a n d the convergence cntrria for temperature profile rvaluation will be rcàuced by a fmr R

- - - --

End of each axiaI element evaluation

Tammab'on of the hem transfer rate evaluation

Figure 5 2 : Process of heat transfer tate evaluation

Page 36

The heat transfer rate in Watts h m each channel surface was equal to the

summation of the heat fluxes h m aü axial nodes,

where i represents the node location dong the channel wall, and A, is the heat transfer

area associated with the node i.

Heat and mass balances for each fluid in each axial segment were perforrned

accordmg to the procedure summarized in Figwe 53 . The variables us& were as

follows:

Input variables: th pn 46. q r

Output variables: x, x, x,

Brine Side

As brine was partially crystdized, a unifonn distribution of salt concentration

was assumed in the resuiting liquid. By performing heat and mass balances on the brine

side, the ice hction distribution was calculated as a function of the heat transfer rate and

the initial salinity of brine. A general relation between the salinity and the ice hction is

given by,

where x,' is the initiai salinity, x," is the fiml saünity (liquici phase), and x,O is the ice

weight fi;iction.

Page 37

P e d m mars balances for the fuu'te segrnais of each channe1 (bine and rciiigaant), ming the htal tmnsfer rate

Mass balance for an axial clmenr N of brire

Evaluate l e exit condition for the auid elanoit h m the heat and mass b a i a m

End o f ihe axial elmient N calculation

EvaIuation for the Refngcmt Channeli

Evalumion fbr axial elment N in channel C Shc scqwntial ordcr dqxnds on the flow direction

Evduation of the a i t vapour quality h m f ie h a t and mass baiancc

End of calculation for ihc a d elemcnt N frofess in chamiel C

I i

I

EvaIuarion of the quality chanse via ttic bend wnnecting channcl C and chanacl C + I

C = 5?

-

End of the ticat and mass batance evaluation for the nFriguant chmcis

i

Figure 53: Rocess of heat and mass b c e evaIuation

Page 38

A correlation required for ice fiaction, xi", was developed using the process shown

in Figue 5-4, by dividing the coohg process into two individual portions: a) chilling of

brine and b) phase change of water, as shown in Figure 5-5. The balances are $ven by,

(5 7)

where q is the total heat ttansfm rate- q b is the heat ûansfèr rate to brine, q, is the heat

transfer rate to produce ice, mb is the total m a s flow rate, x,O is the 6nai ice weight

hction, c, is the specific heat of brine, TB0 is the final freezing point, h, is the enthalpy of

water, and Ho is the final enthalpy of ice.

/ 1 s: in liquid p h u I

Figure 5.4: Condition for correlating the ice fraction

1 !

Figure 55: UIustration of coolhg process h m brine to ice slurry

Page 39

Substitution of Eqs. (5.6) and (5.7) into (5.5) yields,

y mb =(~-x,o).c, -(T/ - o " c ) + ~ ~ ~ . ( I ~ . -HJ

So, h m Eqs. (5.4) and (5.8),

The generai form of the ice fraction correlation was derived as follows,

A=A,,-(X;P+B~-X;+K,, (3. 11)

K = A, -x,' (5.12)

w h m the coefficients Ah Bo, Ko and Ar in Eqs. (5.1 1) and (5.12) were determineci t'rom a

regression analysis of various ice hction daci.

in evaluating the exit fluid condition at an axial segment f as show in Figiire 5.6.

the summation of the term (q /mb) is required (fiom elernent 1 tofi. This terni represents

the heat transfer rate to a unit mas of fluid passing through the segment. Two

contributions are predetennined as given by Eqs. (5.13) and (5. Id) for evaiuating the exit

ice ûaction. The first one is the heat ûansfer rate required for cooling the fluid d o m to

bC. The second one is the heat transfer rate required for cooling the fluid down to the

initial fieezing point.

Figure 5.6: Heat exchanger with n axial elements

Page JO

where Tb1 is the iniet brine temperature.

rtie terrn ( y ) was used to check if any ice was produced. The sub-cooling eff'ct of mb /P

brine before fieezing was assurned to be negligible.

The term, (q/mb)i, in Eq. (5.15) was wasd in combination with the ice fraction

correlation for calcuiating the exit ice hction. For the case of no ice generation. a

simple heat balance was perfonned to cdcdate the change in brine temperature.

Refiineranr Side

A general heat balance, Eq. (j.16), was performed based on the change in thermal

properties of the refrigerant between the idet and the outlet of an axial segment.

(5.161

where hf is the liquid enthalpy, hJg is the latent heat of vapobtion, x, is the vapour

quality, q is the heat transfer rate and Er is the m a s flow rate. The superscripts i and O

represent inlet and outlet. respectiveIy. The enthalpies were calculated b d on the

saturated conditions of the fluid.

Page 41

533.2 Pressure Gradient Evduation

The pressure drop calcuiation procedure is summarhed in Figure 5.7. The

pressures and temperatures of the refrigmt at each node were evaluated. Each axial

element was M e r divided into a certain number of segments for pressure gradient

integration, using the following parameters.

input variables: qn Xr

Output variables: r, p,

The comlations and models for evduating the two-phase pressure gradients have

been described in section 3.3: "Pressure h o p in Two-Phme Flow ". in ihe case of

superheated vapour, the Fanning equation, Eq. ( > . I - ) , was used for calculating the

fictional pressure gradient.

Micro-fins were fabricated in the refîigerant channels for heat transfer

enhancement. The effects of the micro-fins on the hydradic diameter and heat transfer

coefficient were taken into account durùig the construction of the simulation model. A

correction factor was used to modify the evaiuated heat ûansfèr coefficients and the

hydrauiic diameter. The correction factor.&, used was the ratio of the heat trrinsfer areas

with and without micro-6ns. The e f k t of the micro-fins altered the hydrauiic diameter

due to the increase in the wetted perimeter, P,. However, the change in the cross

sectional area could be negiected

Page 42

End of the Iih damnt & w o n

Figure 5.7: Rocess of pressu~e gradient evduatioa

Page 33

Due to iasufficient information avaiiable on the heat transfer enhancement effect

of the micro-tins, the folIowing assumption was made in constnicting the computer

model. It was assumed that the heat transfer enhancement was maidy caused by the

in& internd surface am; hence, the heat tramfer enhancement factor IW defined

to yield the enhanced heat transfer rate,

q=h.f,.A-AT (5- 19)

where A represents the surface rn without accounting for the micro-fins and ( h f i

represents the enhanced heat transfer coefficient.

53 TEMPERATURE DISTRIBUTION (CROSS-SECTION)

A computer code was written to plot the predicted two-dimensional temperature

profile at each mss section of the heat exchanger. Some of the output files from the

geometry program and the simulation prograrn were used as the input to this graphical

output program. The resuits h m this prograrn would dispIay the temperatun: variations

througholrt the whoIe cross-sectionai plane of the heat exc hanger in more detai 1.

6. RESULTS AND DISCUSSIONS

Several experimental test runs were chosen to verify the simulation mdel. These

test ruus and the simulation resuits are presented in the next section, "Results". in

addition to the recorded experimental data, observations made during experiments are

also presented. A cornparison of the resulting steady-state conditions. fiom the

experimenîs and the SimuIations, is discussed in the Iater section, "Discussions ". This

section also discusses the difûcuities confiontecl during the experiments and the

simulation mode1 deveIopment.

The exphenta1 d t s h m test runs conducted to verio the shulütion mode1

and the correspondhg computer simuiation resuits are presented in this sectioc

The heat exchaager in the current research was operated over a broad range of

conditions. Aithough an enormous amount of test data has been o b t a only several

experimentd resuits replesenting the typical operating conditions of the heat exchanger

can be presented here.

Experiments on the scraped-dace heat exchanger were conducted and the data

fiom a total of 65 test nms were tecocded using the data acquisition system. Due to the

extensive simulation time reqilired, ten test nms out of 65 were chosen to verify the

computer simulation modef. Table 6. I presents the ranges of the operating parameters

Page 45

for the test runs. Within these ten test m, two were conducted without producing any

ice, in order to check the applicability of the simulation mode1 to both chilling as well as

ice-making modes.

Table 6.1: Range of experimental parameters

&??rig.Wt si&

aIrz t l a r r r e Lpb 0-15 0.059 met Taper*ure

. C -2.7 -7.7

ut P r e s s u r e k ~ a 450.8 379.1

ut V p m r alauty - 0.20 0.12

mlT COmIIIOIO

Briae side kit T ~ e r n t u r e c O. 7 -3.5 kit Ice fmuion - 0.21 O. 0

h*tguunt s2de

kit T q - t u r e 'C -7.1 -U. 9 hit P T e s ~ e k& 330.1 328.6 kit Vapeur PCILLxty - 0.86 0.38

The operating parameters for each of the chosen test runs are summarized in

Table 6.2. They are divided into two categories, "Inlet Condition" and "Ourlet

Condition ".

Table 6.2: Summary of experimental results for sirnulalion

l ü U T C OUIITI OH

*An. SU. Hi18 I l o r r a c e ka/. *.@IV a . 1 6 1 a . 1 9 6 @ . A 3 1 a .A96 0 . T t 3 @ . I I 3 @ . A t 9 O . l t 9 0 . 1 1 9

In l r t T a p . r a t u r e * c 0 . a * O .O O a @ 1 . 3 0 C S @ 9 . 9 0 l a . O0 0 . 8 1 a . 7 1 @ . 7 1

In l r t S a l i n i t y D . O40 1. ( 4 8 e . 048 a . *Sb a #6@ D . a b @ a .OS@ O . 028 0 . 0 3 8 O . a 3 0

-LI C O l i l T l Q

& k g sldr ExLo T r * p r r + t u r r 'E - 3 . S @ -1 . D O - 3 . a @ - 3 . 1 0 - 3 . 0 ) O , a9 e . 7 ) -;.OS e l .O7 -t . O 1

E x i t I r e t r a c t i o n - 0 . 1 1 O.#& O . @ J O . 1 4 a . A4 1 . 0 ) 0 . 0 6 O .11 O . 1 3 O .11 - E M ~ S t i q e r r s u r r 'r - 1 . 3 8 - 9 . 0 @ - Y ? a 4 a -11 SI - a . $ * - 8 , sa - 7 0 9 -1 . 1 9 - 7 . 1 1

L u i s Pre i r u r r Ma 3 1 6 . 1 > ? a . S $ 6 8 b f l e . 0 l t b . 6 3 7 4 . 3 3 7 3 . 6 3 8 0 . 4 3 e b . 1 3 9 e . l

E n i r Vapour O u a l i t y O . ) @ 0 . 4 1 0 . S l O 6 1 I ?7 1. 41 O . O 0 . 1 4 O . 6 1 e . 6 k

Page 17

The "Inlet Condition" of the workùig fluids (brine and refiigerant) and the

rotationai speed of the rotor shaft (scraping speed) were used as inputs to the simulation

program. The "Ourler Condition " indicates the exiting siatus of the working fluids and

the heat exchange rate. These parameters were used in verifjhg the accuracy of the

simulation model.

The resdts f?om the typical test mw are a h presented as the overall k a t -fer

coefficient in Tuble 6.3. In evaluathg the coefficients for the two test nins in the chitling

mode, certain modifications were made to Eq. (4.12). The area ratio, A, was assigned to

be infinity in the case of no ice making and the term. ATh', in defining the log mean

temperature difference was taken as the temperature gradient at the exit of the heat

exchanger (scraped-surface channe!). Within the range of typical operating conditions,

the o v d I heat transfer coefficient was found to be 3.4 kw/rn2-OC (*lS%). The relation

betweea the other parameters and the overail kat transfer coefficient is Firrther discussed

later in the section, "Diswsion ".

Table 63: Range of overall h a t transfer coefficient

m r a l l k a t Trarwf cr Cacf f icient

in conducting the experiments, some results could be obtained only as qualitative

observations due to the difficulties in qwntifyuig the operation, One of the observations

made was the sound generated by the scraping action on the heat transfer sirrface during

the crystallization proces. As Iow sdt concentration hine was & during the ice-

making operation, a distinctive sound is produced by the mechanism of scnpùig. The

sound started mostly during the transition h m the c h i h g process to the ice-making

process. The chiliiig process represents the operating period where there is CO ice king

forrned. During the transition period, the fluid first undergoes a super-cooling period

where the temperature falls below the k z h g point. This sound started imediately

&er the super-cwling period and just at the beginniDg of the crystallization period.

A layer of ice is thought to have fiozen onto the heat tramfer surfacc but could

not be fully removed by the scrapers. This is probably the cause of the distinctive sound

king heard. This suggestion is proven by observing the flow at the exit of the brine flow

channel after shutting d o m the refrigeration system. Layers of ice were collected at the

exit of the brine flow channel. and the shape of the ice conformed to the curvature of the

scraped-surface. The formation of an ice layer is not desirable h m an operaton point of

view since it would likely decrease the k a t transfer rate because of low thermal

conductivity of ice (2.21 W/mK) and Iead to an unstable operating condition. Besides

decreasing the heat d e r rate, ice iayer formation causes lowering of refngerant

bailhg temperature and damages to scrapers. In some cases, the distinctive sound might

not be easily detected when the frozen ice is unifonnly b d t up on the iaternal wall. The

uniformity of the ice layer cteates a generally smooth surface where the scrapers are

skating on top of the ice layer.

Simulations were perfomed in accordance with the experimental conditions of

the chosen test runs descri'bed in the previous section. The d t i n g outputs h m the

Page 49

-

computer simulation for each of the test nins can be subdivided into four categories;

generai resuiîs, axiai distribution, boundary profles and cross sectional profiles. Each

category of the simulation results is M e r explainai below.

The general d i s of aü the simulations performed are f i presented. These

resuits are later compared with the experimental d t s in order to veriSf the feasibility of

the computer simulation model. Due to an enormous amount of simulation results

obtained, "test run mcmber 19" is used to illustrate the remaining categmies of the

results. This simulation represents a typical o p t i o n of the heat exchanger during ice

making. The simulations of other test nins are presented in Appendix D, "Simularion

Results ".

6.12.1 Summary of Ceneral ResuIts

The general resuits obtained fiom the computer simulation are the exit conditions

of the working fluids and the heat exchange rate as summarized in Table 6.4. In

veriwg the feasibility of the simulation mode[, these parameters were compared with

the d t s of the experimentd test ctms shown earüer in Table 6.2.

The information on the mesh useci in the simulation is fûst presented Ùi Tuble 6.4.

The "mesh site" indicates the dimension of the me& used for the evaluation of cross

sectional temperature distriiutions. The "no. of Gaial elemenrs " represents the number of

segments that the heat excbanger is divided ZuCialIy. Tite "aial element size" under

Process Inform~ltion shows the axiai length of each element axiaily.

(8 ) hafnb+a buttoo* - aar; rajruira ar#

(-1 Aov1rfib aitano - (rd) a r n r t i ~ d nit- - (a) isnarirdriba .rbtmo -

r r ~ r d p; ig

1-) rsyor.i# rrg aet?no - (-1 IhfUt l* i 381'mo - ( a ) i snn lhhaa ailmo -

Jut *a

: ( W u ) p i id* fiuoflr(â41 - ioatoqey

: (rlbq) aav:atp r r w - : (-1 hattrnb ailuf . : (rd] a~nerasd ailu! - : (1) i f m r s d w a ~ 4 p f i

W.. 6 1.) ~II:

: ( r p q ) bWlWt$ * r w - : (-1 Aafuttrr aituf - : (3) irnarridwa ailut -

aul:u ~01%-

Page 51

in Tuble 6.4, there are 8 parameters in the input condition and they are divided

into 3 sections: brine, refngerant and agitator. The brine section indudes the inlet

temperature, d i t y and mass flow rate. Most of the test nins were perfonned with the

b h e inlet temperature of (close to) O°C in the ice-making mode. Runs 20 and 2 1 were

performed with a higher temperature (-IPC) for testing in the chilling mode. The

refngerant section includes the inlet temperature, pressure, vapour quality and mass tlow

rate. The agitator section has oniy one parameter that is the rotational speed of the rotor

shaft holding the scrapers. Based on these input parameters, the simulation mode1

predicted the output conditions which will be compared with the experimentd results in

the later sections. On the brine side, the output condition includes the temperature,

salinity and ice fraction. The output temperature wiH be associated with the salinity

(freezing temperature). The salinity represents the salt concentration in the liquid phase

and increases if water freezes and turns into ice. n e salinity is then directly related to

the change in ice tiaction. On the refigerant side, the output condition ilcludes the

temperature, pressure and vapour quality. The heat iransfer rate will be the rnost

important result used in comparing the predictions with the expenmental resuirs.

6.1.2.2 Axial Distribution

The "Axial Distribution " results contain the axid vananations in the properties of

the working fluids (brine and reûigerant) dong the flow channels. in addition to the

working fluid properties, the heat transfer rate to each fluid channel at each cross section

dong the heat exchanger was also obtained. The redts h m the simdation of "test nm

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number 19" are presented below as an illustration. The parameters are plotted against a

dimensionless distance h m the brine inlet.

The conditions of the working f i d s are shown in F i p e s 6.1 through 6.k in Figure

6.1, the brine temperature faUs iiaeariy in the chilling region (nonnalized axia! distance <

02) and then remains relatively constant thereafter (nod ized axial distance :* 03) at the

freePng temperature. ïhe iœ W o n Figure 6.2) and salt concentration /Figure 6.3)

increase Linearly in the ice-mahg region (nomalized axial distance > 03). In Figure 6.4.

the legend trI-û5 represents the tempaature in the refrigerant channets, where h n n e l 1 is

equal to channel a (presented earlier in Figure 4 I ) , 2 to b and so on. The m w s indiate the

flow direction of the refiigefant. The saturation temperature (Figure 6.4) and pressure

(Fipire 6.5) MI linearly in each individual chamel. The figures show a larger reduction rate

in the beginning channel, then smaüer at the end This is relateci to the size of the channels

and it WU be discussed in later sections. ûn the orher hand, the vapow q d t y variation

(Figrue 6.61 shows a greater reduction rate at the last channe1 cornpareci to the first channel.

The heat d e r rate in each refigerant channel is plotted in Fip-2 6. ' dong

with that for the brine channel. Once the fluids enter the ice-rnaking region, the heat

transfer rates are predicted to remah neady constant. in contrast, a plot of the heat

m f e r rates for "test run d e r 20" (chilling mode) is provided in F i g m 6.8 as a

counterpart to the "test run number 19 ", shows continuously varying heat tmsfer mes

thughout the heat exchanger. in Figtrres f i 7 and 6.8. qb and qr represent the heat

transfer rates for the brine channe1 and the refiigerant channels, respectively, where the

numbers after qr represents the refngerant channel numkr.

Page 53

O. O 0 . 2 0-4 0.6 0.8 1. O

N o r e r l i t s d Axial Distance

Figure 6.1 : Variation of temperature in brine channel

Figure 62: Variation of ice fraction in brine channel

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Figure 63: Variation of salt concentration (liquid phase) in brine channel

Figure 6.4: Variation of temperature in refngerant channels

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Figure 6.5: Variation of pressure in refrigerant channeb

Figure 6.6: Variation of vapour quaiity in refngerant channels

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Figure 6.7: Variation of heat transfer rate in each fluid channel (ice making)

Figure 6.8: Variation of heat W e r rate in each fi uid channel (chilling)

6.123 Bounda y Profdes

Both the temperature and the heat flux distributions calculateci for each axial

segment of each fluid channel are plotted against the angle around the heat transfer

d a c e of the fluid channel. The boundary temperature distribution plots are presented in

Figures 6.9 through 6.11, while the boundary heat flux distribution plots are presented in

Figures 6.15 through 6.20. The nomenclature and the orientation of the channels c m be

found in the previous section (Figure 4. II. The distriiution plots for the brine channel

cover ody one quadrant of the cross section since al1 four quadrants have the same shape.

The zero-degree Sine corresponds to the line, as illustrated in Figure 4.1, besides the

rej?igeranr channel a; and the distribution is plotted against the angie in the an&

clockwise direction. For each refiigerant channel. the zero-degree angle is the point

closest to the brine channel with a plus sign corresponding the anti-clockwise direction.

Boundary profiles at different axial locations are plotted on the same p p h to

illumate the variation of profiles dong the heat exchanger. These results are taken h m

the simulation results of "test run number 19". For the wall temperature distribution

plots, for both the brine and refrigerant channels, the top contour is the profite at the

segment closest to the brine inlet. For the heat flux profiles, the sarne trend as the wall

tempemm plots exists where the top contour is the profile closest to the brine idet for

the refngerant charnels, however, the bottom contour of the heat flux plot for the brine

channel is the profile at the segment closest to the brine idet.

Page 58

Figure 6.9: Wall temperature profiles (brine channel)

r i p . r i e a n (üatraqmramt C h u n d #l lut h m u ï u M . a l : U 9 T e m t Run m i œ d i t d m a t h 20 mami d t 8 .sd rih PO1

-*xi-i

--cd-:

rxl-l nx? -4

- axL-i -,,.xi-a -a:-:

-ui-i

ml-+ 4x1-ici

ml-:

axi-1-

a ? - i . ai-:.

ml-i!

ul-Li.

-Ml-1-

- -1-11,

axl-i'' d - 2 1 i

~ - - - -

- i eo - 5 -oa -4s s 6: 90 135 IOC

Figure 6.10: Wall temperature profiles (refngenuit channe1 a)

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Figure 6.1 1 : Wall temperature profiles (refiïgerant channel 6)

Figure 6.12: Wal t temperature profiIes (refngerant channel c)

Page 60

Figure 6.13: Wall temperature profiles (refigerant channel d)

Figure 6.14: WdI temperature promes (refngcmt channel e)

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Figure 6.15: Wall heat flux profiles (brine channel)

Figure 6-16: Wall heat flux profiles (refrigerant channel a)

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Figure 6.17: Wail k a t flux profiles (refrigerant channet 6)

Figure 6-18: WaiI k a t flux pmfiles (refigerant channel c)

Page 63

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

Figure 6.19 Wall heat flux profiles (refigerant channel d)

Figure 620: WaiI heat flux profiles (tem'gerant channel e)

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In Figure 6.9, the temperature profiles show a slight temperature variation dong

the heat ûansfer sinface. Fmm this sample, the temperahm varies with a maximum of

about O.9C near the inlet of the heat exchanger and a minimum of about O.?C at the

exit. The heat flux varies h m a maximum of about 6000~/m' near the inlet to a

minimum of 2000w/m2 at the exit, as c m be seen in Figure 6.15. The figures show that

the maximum wall temperature occurs at the angle of O degree which has the minimum

heat flux. The maximum heat flux occurs at the angle between 40 and 50 d e p s . which

bas the minimum surface tempewture. The O degree angle is the closest to the fim

retngerant channel (channel a) which is the incoming channel and has the highest

refngerant boiling tempera=. The exiting refngerant channel is located ktween 10

and 50 degrees and has the lowest refrigerant boiIing temperature. The temperature and

heat flux variations are rnainly caused by the temperature difference in the retngerant

channels resulting h m the pressure drop witfün these channels.

6.1.2.4 Cross Sectional Profiles

The two-dimensional temperature disaibutions of the heat exchanger cross-

section at a different axial location obtained h m the simulation of "resr run tiumber 19 "

are presented in Figures 6.21 througb 6.24. During the simulation, the heat exchanger

was separateci into twenty equd axial segments, which were consecutiveiy numbered

fiom 1 starting with the segment closest to the inlet of the brine channel. Figure 6-22

shows the temperature profile at the fourth segment where ice crystaiiization commenced.

These profiles will be discussed in detail in the next section,

Page 65

F i p 6.2 1: Cross sectional temperature profile (elemenr 1 )

Figure 6.22: Cross sectional temperature profile (element 4)

Page 66

Figure 6.23: Cross sectional temperature profile (element I I )

Figure 6.24: Cross sectional temperature profiIe (element 20)

Page 67

in the followhg sections, the d t s h m the experimenis and the simulations

will be compared and discussed. The observations made during the heat exchanger

operation and ice slurry production will also be discussed.

62.1 COMPARISON OF EXPERIMENTAL AND SIMULATION RESULTS

Due to the difficulties in measuring the distn'bution of the operathg parameters

axially dong the heat exchanger, a comparison is made on the exit conditions of the

working fluids. The comparison is divided into three parts, the brine channel, the

reûigerant channels, and the heat exchanger as a whole. Each part will be discussed

separately in detail below.

62.1.1 Brine Channel

The brine channel represents the inside channe1 of the heat exchanger where ice-

sIurry production took place. The three parameters of interest are the temperarure, the ice

fraction and the salt concentration.

Exiî Temperature

The exit temperature of brine, predicted by the simulation model. is plotted

against the experimental data in Figwe 6-25. The plot shows that the estimated values

are quite close to the experimental resdts, except for the two data points at the

experimental temperatures close to 0.5"C. These two data points are associated with the

nins not undergoing any ice-making process. Since the m g temperature of ice sliirry

was assumed to be a specific function of the saIt concentration in the liquic phase, the

Page 68

exit brine temperatures predicted by the mode1 for test iuns undergohg ice-making

pmess are close to the experimental values. This is due to the insensitive nahue of the

h z h g temperature with respect to the ice hction in the low ice fraction range.

However, this assumption is valid only if the salt concentration is uniform througbout the

liquid phase of the ice slurry.

Figure 625: Cornparison of brindice sluny exit temperature

EXt Ice Fraction

A cornparison between the experimental and predicted exit ice fiactions is shown

in Figive 6.26. As can be seen, the prediction of the test nms is satisfactory. As shown

in the plot, a trend-line drawn through the data points ushg linear regression, shows an

underestimation of 12%, however, the exit ice fraction is oniy representing the latent heat

transfèr portion of the whole heat exchange proces. Hence, the assumption of uniforni

Page 69

salt concentration in the liquid phase, which is used for predicting the lÏeezing

temperature, may infiuence the prediction of exit ice fraction, since the distribution of the

sensible-to-latent heat transfer ratio is affected causing inaccurate prediction.

Figure 626: Cornparison of the predicted and measured exit ice fiactions

Exit SuIf Concen~ation (Li& PortÏonL

A change in the sait concentration in the liquid phase results directly h m the

change in ice fraction. The sait concentration accounted for in the simuIation mode1 can

be thought of as the mean value throughout the iiquid phase. The calculated value is

based on the initial sait concentration of the brhe and the resuiting ice fraction. Since it

is diacult to evaiuate the sait concenûation IocaiIy, the use of this mean vahi: is the best

approach at this stage. Hence, the cornparison of the exit ice fiactions shown in Figure

6.26 can be considered to r e p m t the acçuracy of saIt concentration prediction as weiI.

Page 70

62-13 Reirigerant chanael

On the refngermt side, three main variabIes are the pressure, temperature and

vapour quality. These parameters are interretated to each other, where one parmeter

might alter the outcorne of another.

f i t Pressure

In Figure 427, a cornparison of the simufated and experirnental exit ?ressures is

presented. The pressure is the main variable used to determine the reiïigerant

temperature since the experiments are performed under a saturated conktion. The

cornparison shows teasonable trend of prediction by the simulation mode1 where the

mean deviation of the prediction h m the data is only about 0,816. Since the accuracy of

the exit pressure prediction depends on the pressure drop evaiuation. this aspect will be

discussed in more detail in the later section, "Heat Exchange Rate and Pressura Drop ".

Figure 627: Compatison of refrigemt exit pressure

Page 71

&it Temuerunue

B a d on the saturated condition of the refngerant, the tefigerant temperature at

the exit can be correlated as a function of pressure. As long as the exit refngerant is not

superheated, the temperature would be given by the saturated temperature. In

Figrire 6.28, the predicted refrigerant temperatures at the exit are plotted against the

experimentai results. The predictions are seen to be in good agreement with the

measurements. Since the saturated temperature is directIy evaiuated h m the pressure,

the comparison in this figure should correspond to the comparison of the pressure.

However, the refngerant temperature is one of the main parameters used to evaiuate the

heat exchange rate. Hence, the pressure drop evaiuation dong the reûigerant channels is

important in the simulation mode1 since it will affect the temperature gradient between

the b ~ e and refigerant channeIs.

Figure 628: Comparison of refiigerant exit temperatwe

Page 72

Exit @di@

The exit vapour qualiîy predicted is compared with the experimental results in

Figure 6.29.

Figure 6.29: Cornparison of refiigerant exit vapour quality

From the cornparison, it can be seen that the simulation mode1 yielded reasonable

estimates of the exit vapour quality, showing a slight over prediction (+2.6%) on the

average. The predicted vapour quality would have been rnainly af3ected by the heat

transfer rate. Although the change in temperature (or pressure) might aiso have an

impact on the vapour quality due to change in enthalpy, the effect would be minor in the

current study; therefore, other than the effect on heat exchange rate due to temperature,

the change in vapour quaüty caused by pressure drop is wmparatively negiigiile. in

observing the changes in the predicted heat transfer rate at different cross sections (in the

Page 73

ice-making region where the temperature dEerence is more or less constant), the change

in the vapour quality has not favoured or negative!~ affected the heat exchange pmcess in

a great manner. Hence, the vapour quality of refiigerant changing h m 0.12 to 0.86, over

the range of testing, can be neglected h m the generd characterization of the heat

exchanger since the heat exchanger is more or less dominated by the temperature

difference.

6.2.13 Pressure Drop and Heat Exchange Rate

The heat transfer rate in the heat exchanger and the pressure drop in the

refiigerant channels are the two most significant parameters for evaluating the validity of

the simulation model.

Pressure D r o ~

The pressure gradient is one of the most important factors in refigemtion due to

saturated flow boiling of the refngerant in the heat exchanger. Based on the saturation

pmperties of refngerant, the Ming temperature changes with the refrigerant pressure.

Moreover, the refngerant temperame is the most important parameter in the heat transfer

rate h m brine in cornparison with the others. Hence, the accurate evaluation of the

pressure gradient is a major conceni in developing the simulation model. Thecefore, the

total pressure drop amss the inlet and the outiet of the heat exchanger refngerant

channel is used to characterize the heat exchanger in general.

in Figzue 6.30, the predided pressure drop behiveen the iniet and the outlet

refiigerant channels is compared with the data The comparison shows a standard

deviation of about *IO% for the ten typicai test runs. This suggests that the simulation

Page 74

model is giving good predictions of the general operating condition of the heat

exchanger. The assumption used in simulating the two-phase pressure gadient for

channels having micro-fins appears to be valid, however, it c m be seen in the figure that

the model shows a siight mder-prediction of the data. The pressure gradient is

dorninated by the two-phase fictional pressure gradient, which is predicted using the

Lockhart-Martinelli comlation. in the litemture, this correlation is known to usually

underestimate the pressure drop when applied to tefigerant two-phase flows. This might

be one of the reasons for the slight under-prediction; however, this correlation is widely

used and is the easiest to apply. Nonetheless, the simulation model is able to provide

adequate predictions of the experimental results.

Figure 630: Cornparison of refngerant pressure drop

Page 75

Heai Exchawe Rare

The heat transfer rate between the working fluids characterizes the performance

of the heat exchanger. Since the heat exchange rate is the h a 1 resulting parameter, this

main steady-state process variable is used to evaluate the feasibility of the simulation

model. The heat exchange rate evduated using the simulation model for each test run is

ploned qainst the experimentai results as shown in Fi.ctrre 6.31. The simulation mode1

is seen to give misonabte predictiom where most ot' the predictions f ' l tvithin 114% of

the data

Figure 6.3 1: Cornparison of heat exchange rate

However. there is one simulation resuit (about 6000W) significandy less than the

expimentai vahe of8000W. The pressure drop in the refhgerant channets predicted for

this data was less than the experimentai value: hence. the temperature merence between

the working fluids wodd also be Iess than the experimental data This is probabiy the

Page 76

probably the main reason for underestimating the heat m e r rate since the heat transfer

is directly proportionai to this temperature merence. As has been discussed in the

previous section, the under-prediction of the refiïgerant pressure &op would be the main

cause of any deviation in the heat tramfier prediction.

Based on the results of cornparisons between the predicted and m m e d heat

transfer rate and pressure drop, the simulation model has been shown to generdly provide

satisfactory predictions of the sjeady-state operation of the curent scraped-surface heat

exchanger. Therefore, the present computer model was further used to grun a better

understanding of the heat exchanger operation in detail as described in the nelr. section.

The simulation results obtained with the present model and described previously

are now discussed in detail in the following sections.

633.1 Axiai Distributions

The axial distributions generated h m the simulation model represent changes in

the fiuid conditions dong the heat exchanger. The reIations between the chg ing fluid

conditions and the heat exchange rate will be discussed in the rest of this section.

Chunnes in conditions of the brine flow

The change in ice fraction dong the heat exchanger was found to be linear fkom

the simulation. This can be explainai by examining the rate of heat removai h m the

brine channe1 dong the heat exchanger. As seea in the heat transfer rate plots, (Figures

6.7 and 6-81, the heat transfer rate dong the heat exchanger in the latent heat removal

zone (crystallizaîion region) was relatively constant in cornparison with the chihg

Page 77

region. The ciifference between the axial heat transfer rate distributions associated with

the ice-making process (Figure 6.7) and the chilhg process (Figure 6.8) suggests that

the locaI heat transfer rate is rather dominated by the radial temperature gradicnt in brine

since crystallization took place generally at a constant fieePng temperature. Although

the fieezing temperature was continuously decreasing dong the heat exchanger. it was

still quite constant,

The continuous brine temperature reduction was caused by the change in salt

concentration resuiting h m the increasing ice ûaction. Since the change in ice fraction

was linear, it led to a linear change in salt concentration. An observation can be made as

well that the effect of ice ûaction on the heat d e r rate is not a dominating factor as

long as the ice sluny is in a homogeneous state. This was assumeci in the use of the mean

properties of ice sluny for the calculation of the scraped-surface heat transfer coefficient

during the construction of the simulation model. In addition, the change in vapour

quality dong the retkigerant channels was not affecting the heat transfer rate significantly

in cornparison with the boilhg temperature.

As has been mentioned above, the total heat transfer rate at each crcss section of

the heat exchanger (ice slurry zone) was quite constant due to the constant temperature

difference between the brindice slurry and the refiigerant. From the axid distribution

plots, the temperature difference between the ice-sluq and the refiigerant can be seen to

remain approximately constant dong the latent heat removai zone (ice making region).

Aithough the temperature of ice sltrrry was decreasing towards the exit of the heat

exchanger, the mean boihg temperature of the refiigerant was aiso decreas-ing axially

due to the pressure drop. Therefore, the mean temperature of the refiigerant might be

Page 78

used as a gened parameter to account for the heat M e r performance of the heat

exchanger. Based on this, an overaii heat transfer coefficient (for evaiuating the overall

performance of the heat exchanger) codd be defined using the mean temperdture of the

refiigerant.

Along the sensible heat removal region (brine chilling zone). the change in heat

ûmsfer rate was found to steadily decrease. Due to the absence of latent heat, the brine

temperature was r e d d s ign i f idy . However, this did not affect Ihe boiling

temperature of the refiigerant greatly since the pressure drop in the refiigerant channels

was a more dominant factor. The heat transfer rate was once again contrclled by the

temperature dierence between the brine and the refigerant. Hence, the rate of change

in the brine temperature decreased dong the heat exchanger; this can be clervly seen in

the simulations of the "test runs 20 and 21 " both of which were associateci with only

chilling of brine rable 6.4). Due to the decreasing bfine tempetanue, a suggested

approach would be to take the log mean temperature difference as the temperature

difference.

The change in pressure was alrnost Iinear in each of the refigerant channels. An

example was given previously h m the simulation of "tesr run nwnber 19 " in Fipire

6.5. The retiigerant c b e I s were large in cross section for the flow-rates testeci, so that

the acceleration pressure drop due to the change in vapour quality did not conmiute

significantly to the overdi pressure drop. in addition, the change in vapour @ty was

quite linear since the heat mms6er rate was fairiy constant throughout each refiigerant

channel.

Page 79

Pressure chan~e in the re151'gera.t c h e h

Comparisons were made between the simdated and measured cefigerant pressure

variations in each channel. in Figure 6.32, the reûigerant pressure measured dong the

heat exchanger is plotted against the distauce h m the reûigerant inlet for "test run

mmber 19". Pressures were rneasured at the bends where two consecutive channels

were connected. The simulation d t s foilowed the experimental results quite well

except for the second last channel where iess pressure drop was predicted.

Figure 6.32: Cornparison of change in rem'gerant pressure ltest nm 19)

6333 Boundary Profiles

As iiiustrated in the "Rrsulrs " secrion 6.2, the boundary profiles of temperature

and heat flux were outputted h m the simulation modeL ïhe changes in the botmdary

profles were used to ver@ the variations of the other parameters dong the heat

exchanger. Additional boundary profile &ta can be found in the -4ppedx Dr "The

Simulation Remlts'',

if the refiigerant channels are be rearranged in an ascending order of the minor-

to-major axis ratio, the resuiting sequence would be channels 6, a. c. d and e. According

to this sequence, the temperature ciifference between the innermost and the outermost

positions of the refiigerant channels (with the center of heat exchanger as the reference

point) was increasing as well as the heat flux differentiai. This can be wel! explained by

the conduction cesistances h m the shape of the channels and the separation between the

channels.

Differences can be found between refngerant channels' waii temperame fFigwes

6.10 to 6.14) and heat flux fFigures 6.16 to 6.30) plots where the temperatm profiles

spread out wider than the heat flux profiles. Hence, the heat tramfer mostly took place

on the inner wail closest to the brine channel. in addition, the heat exchanger can be

thought of as a fin-srpe heat exchanger since the outennost surface of tbe nfiigerant

channels did not contribute signifïcantly h m a heat tmde r point of view. 3 s type of

heat exchanger would create certain non-uniformity in temperature and heat f l u

distributions dong the scraped-dace. In the industry. this non-uniformity of

temperature and heat flux distributions is believed to cause instability in ice making with

the scraped-dce heat exchanger. However, during the experiments, signiticant

evidences could not be found to support this hypothesis.

The predicted waii temperature and kat flux profles of the brine channel

(Figures 6.9 and 6.15) show a discontinuous distribution. However, this dkcontinuity

was possiiIy caused by the definition of the finite mesh. Looking at the variations in the

Page 8 1

temperature and heat flux profïies, the scraped-surface condition was not affectai

significantly (IOS°C on the scraped-surface) by different boiling temperatures in the

refngerant channels. Throughout the plots, the heat transfer coefficient in the ice-making

region remained constant. The changing pmperries of the two-phase flows (on both

sides) did not affect the heat transfer rate since both the cold and hot fluids were at

saturation. Hence, the temperature difference was the main driving force for heat

trader.

6.2.23 Cross Sectional Profiles

One of the purposes of examining the temperature distributions in the aluminium

channel wall cross sections (at different axial distances) is to confinn the simulation

model predictions by observing the common change in temperature profiles. From the

detailed temperame profiles, the simulation model can be judged to k performing well

since the predictions show clearly decreasing temperature gradients h m the brine

channel to the refiigerant cbannels.

The resuits show that the channel shape is an important factor in affecthg the heat

transfer efficiency, and in theory, the fm-like structure of the k a t exchanger will enhance

the heat transfer rate, The temperature dimiutions show that most of the heat was

transferred into the reçigerant cbannels before penetrating to the outside of the heat

exchanger. Hence. the outside thiclcness is not as signifiant except during t;ie transient

state (due to heat capacity of the material). Although the fin-like structure of the heat

exchanger wouid have certain enhancement, it did not enhance the efficiency of heat

Page 82

transfer significantly where a comparatively smaii amount of heat was transferred

through the solid between the refrigerant charnels.

in addition to the simulation mode[ development, extensive observations have

been made on the collected data, Le. the effect of rotational speed of the scrapers. Aside

h m the quantitative results, observations based on the operating experiences are also as

important. These topics wiil be discussed m e r in the following sections.

6 3 . 1 Effet of RotationaI Speed

Experiments on the heat exchanger were performed under three different

rotational speeds of scraping. Since the main focus of the experiments was to evaluate

the simulation model, the effect of the rotational speed was not investigated in M e r

detail. However. a brief summary of the effect can d l be made h m the experimental

data. Although most of the experiments were conducteci with the heat exchanger in the

verticai position, severai tests were done in the horizontal orientation as well. The

corresponding test results on the effect of the rotationai speed are presented in Figures

6.33 und 6.34, for vertical and horizonta1 orientations, respectively. in the figures. the

overail heat transfer coefficient, which indicates the eficiency of the heat exchanger was

plotted against the rotationai speed.

Page 83

- - -

+Keac 'kafisfer ~ o e f f i c r e n c O E u t Ice Fraction - --

Figure 633: Effect of rotational speed (vertical orientation) (Test conditions: brinc flow rate, O.I3kg/s, R22 flow rate, 0.059kg/s, inlet salt concentration 3wt9bNaCI)

-- - . -- +-KG~ Earufer Coefhuent m Exlt [ce Fract~an --- -

Figure 634: Effect of rotational speed (horizontal orientation) (Test conditions: brine ffow rate, 0.13kg/s, R22 flow rate, 0.067kgls, inlet saIl concentration, 3wto4NaCI)

Page 84

in Figure. 6.33. a trend of decreasing overall heat d e r coefficient with the

rotational speed is shown for the vertical orientation of the heat exchanger. This senes

was conducted under the same operating conditions; i.e. the same refngerant flow-rate

and the same brine flow-rate. In contrast a peak was found in the horizontal orientation

test series as shown in Figure 6.34. Thus. an increase in rotationai speed does not always

benefit the overall performance of a scraped-surface heat exchanger in ice-making

applications. as previously reported by Bel et al. ( 1996).

However. the expenments performed in the present work are insuficient ro

iIlustrate the general effect of the rotational speed. since the experimentai conditions

covered were quite limited. In addition. one of the dificulties in showing the effect of

the rotational speed is the operating condition of the working fluids. Any changes in the

heat transfer rate will affect the fluid condition such that the effect of the rotationai speed

cannot be obsewed clearly. Hence. further investigations shodd be performed

specifically to smdy the effect of rotationai speed on the scraped-surface heat transfer for

ice crystallization.

633.2 Scraped-Surface Heat Exchanger in Ice-Making

Two major dificulties have been found so far with the binary solution (sodium

chloride solution) during the operation of the ice generator. Both were found to be

cIosely related to the heat transfer and the crystallization processes at the scraped-surface

channel wall. One of the problems was described previously in Secrion 5-2.1.2, where a

hardened ice-layer. which couid not be scraped off. built up on the heat mander surface.

Page 85

-

Another issue was an inmashg drag force that was exerted to the scrapers during the

ice-making pmcess. These points are discussed in more detail below.

Hmdened ice-laver created on the scraped-strrfuce

During the experiments under certain operathg conditions, a hardentd ice layer

was found on the scraped surface of the heat exchanger. The ice Iayer adhering to the

surface could not be removed by the scrapers. It could be removed only when the

refiigeration system was shut down to reduce the adhesive force between the ice and

the surface. The presence of the fiozen ice layer on the heat transfer surface was

determined by observing the exiting flow; the flakes of ice were found to have the

sarne curvature as the d a c e . During the experiments, this ice freezing effect dso

created a distinctive sound. distinctly different h m the mechanicd vibra-ion. Other

than shuning off the refiigeration systern, there was no apparent method to resolve

this problem afler the h z e n ice layer had forrned. A possible future solution wodd

be to invent a sensor that could indicate formation of a fiozen ice layer hy detecting

the sound.

During the experiments (including the operation of other scraped-surface heat

exchanger units), it was found that the fiozen ice layer f o m more aisiiy as the

concentration of the brine is reduced. in addition, each specific design of scmped-

surface heat exchanger would be susceptible to ice-layer build-up rit different

minimum brine concentrations. The formation of the h z e n ice layer was found

mostly during the period h m chiiling to ice crystallizhg modcs at which

the distinctive sound could be heard h m the transition region inside the heaî

exchanger.

Page 86

Some techniques could, however, be used to prevent the h e z h g effect. During

operation, the chilling to ice-making transition region could be spread h u g h the

heat transfer area by increasing the flow rate and the severity of the effect couid be

reduced. A batch type operation in which the heat exchanger was comected to a

closed loop with a fixeci amount of brhe was used in the present experiment; hence,

the brine could be concentrated in a fast manner exceediig the minimum required

operating concentration. In this case, the hzen ice layer would slowly be scraped off

h m the heat transfer surface. There were more operathg methods to deal with the

fiozen ice effect; however, they were al1 based on the idea of increasing the bine

concentration. Prevention or resotution of this situation wouid require another

investigation.

Increused dran force exerred ont0 the scrawrs

Another operating problem with the scraped-surface heat exchangers in ice-

making applications is the increasing drag on the scrapers. This increased h g force

necessitated the usage of a higher power drive motor.

During the experiments, it was found that the pwer required to drive the scraping

rotor increased whenever the heat exchanger was in ice-making operation. Ln

addition, the power was increasing as the operating ice fiaction increased. However,

the power dropped to the same IeveI as in the c W g operation whenever the

refrigeration systern was shut down, This suggested the ice fraction was the main

factor causing the i n d operating power.

As the ice fraction increased, ice crystais codd no longer be easily repositioned

h m the near-Wall region into the core region of the brine flow channel, and the

entire near-Wall region wouid be occupied by an icehrine mixture. As a result,

surface renewai would become inefficient and the rate of heat removal wouid be

reduced tremendously due to a reduced temperature gradient (because of saturated

state during m g ) . Hence, most of the hait was probably removed from the outer

region where a higher density icehrine mixture and greater adhesion (or viscosity)

would exist. One of ttie evidences supporthg this scenario was findine a highly

dense icelbrine hollow cylinder inside the heat exchanger, after opening up the heat

exchanger following a long period of ice-making operation,

From above, the heat capacity distribution is an important factor for the

performance of a scraped-surface heat exchanger. However. a limited degree or

mixing would be a characteristic of ice slurry. At certain ice Fractions. ice slurry cm

become rigid enough (capable of holding liquid) that mixing would be difficult to

achieve.

The comrnon factor involved in both phenomena is the adhesion of ice onto the heat

transfer surface caused by continuous cooling. Formation of a hardened ice layer on the

surface occurs dmost instantaneously during the transition from a chilling to an ice

making mode. in contrast, the increasing drag force occurs in a slower manner where it

is caused by continuous build up of an ice crystai population dong the hcat transfer

surface. Hence, bey can be considered as situations where insufficient thermal

distribution occm between the bulk and boundary regions in accordance with specific

heat transfer rate. However, the crystaiiiition characteristics of binary solutions at

different concentrations are aiso a significant factor in the phenomena d e s c n i . Further

indepth research should be made to confhm this.

Page 88

6.233 Overail Heat Transfer Coefficients

A major benefit of a method capable of evaluating the overall heat trader

coefficient is for characterization of the generaI performance of the heat exchmgers with

a minimal number of parameters. A more general comparison can then be made after

data reduction (i.e. temperature difference). A complete set of overall heat transfer

coefficients obtained in the present work cm be found in Appendir C-2. These heat

transfer coefficient data will be highly beneficial for later cornparisons witii other ice-

making heat exchangers.

62.4 FINAL REMARKS ON THE SIMULATION MODEL

During the development of the simutation model, there existed certain dificutties.

One of the difficuities was to find suitable heat transfer correlations that couId be

incorporated into computer codes. Since the simulation model was constmcted using a

finite ciifference technique for calcuIating the heat conduction rate across the channel

walls, a well defined boiling heat transfer coefficient correlation (that could account for

the effect of wall temperature) for the coolant was needed. On the refiigerant side,

Chen's correlation that accounted for both nucleate boiling and forced convection

vaporization was suitable for the present purpose. The contributions of later mearchers

to the development of correlations for nucleate boiiiig suppression and forced convection

vaporization made it easy to incorporate Chen's correlation into a computer algorithm.

Another difficulty was in estimahg the two-phase pressure dmp for the

refiigerant flow. The Martinelii-Nelson correlation was chosen due to its applicability

Page 89

- - -

over a ~ i d e range of vapour quaiity, while other correlations were specifically targeted

for a certain vapour quality range and flow regimes.

On the brine side, Skelland's (1962) correlation for scraped-dace heat transfer

coefficient was used due to a lack of alternative correlations for ice making applications.

The Skelland's correlation had been used in practice and suggested by the industry for the

scraped-surface heat exchanger. Since the correlation was developed mainly for a single-

phase process, ice slurry was assumed to be homogeneous.

Due to insufficient data to account for the effect of micro fins on heat transfer in

refrigerant channels, the enhancement effect was estimated using the increase in heat

transfer area in the pressure drop calculation, the increased pressure drop was predicted

based on a slightiy increased wetted perimeter.

Finally, the effect of pressure variation in the brine channel was neglected due to

the sIow flow in the experimental apparatus. The pressure effect would not lx enough to

have a signifiant effect on the saturation characteristics, such as the freezing point.

7. CONCLUSIONS AND RECOMMEMDATIONS

In the present work, the heat transfer characteristics of a scraped surface heat

exchanger for ice-sluny generation was investigated both experimentally and

numericaiiy. The main focus was on the operation of the scraped surface heat exchanger

in ice making with a binary solution af sadiun chtoride (salt) and m e r .

A set of cornputer pmgrams was written in Fortran 90 to simulate the operating

conditions of the heat exchanger. Past literature in the fields of scraped surface heat

excbanger and two-phase flow were reviewed for constructing the simulation model,

During the model development, certain assumptiotts were made in order to simplifj the

simulation model. The rnodel consisted of k e main parts; generatioa of the geometric

mesh, evaiuation of the steady state conditions, and graphical display of the cross

sectional temperature profiles.

The evaluation of the steady state conditions in the ice-slurry and refiigerant

channels was the main part of the m d d since it involveci many computationi steps and

generated most of the results. An iterative approach was used in the simulation model,

where the iteration consisted of three procedures; the evduation of heat exchange rate,

the heat and m a s balances, and the evaluation of two-phase ptessure drop in refigerant

channels.

Physicai testing of the scraped d a c e heat exchanger was conducted over a broad

range of operating conditions. The expmhental d t s were then used to ver@ the

Page 9 1

predictions of the simulation modei. Tests were conducted in both ice-making and

c h i h g modes to enlarge the verifjing range of data, The most important parameters for

cornparisons between the expeximentai and simulation teSul& were the heat exchange rate

and the pressure drop in refiigerant channeb. The heat exchauge rate was found to be

more sensitive to the temperature differeace between the refngerant and ice-slurry,

compared to the heat hausfer coefficient. This is due to the relatively constant nature of

the heat transfer coefficient compared to the fluid tempera-. Pressure drap in the

refiigerant channels was also important for the heat exchange rate evaiuation because the

saturation temperature depends on the IocaI pressure.

Cornparisons of the measured and predicted heat exchange rate and pressure drop

showed mean deviations of 4 . 3 % and -8.2%, and standard deviatious of *14% and

&IO%, respectively. Therefore, the simulation mode1 developed was found to be capable

of reasonably predicting the performance of the scraped surface heat exchanger tested in

the current work. in addition, the consideration of micro-fius in the model was

acceptable, even though it was a simple appmach. Fdermore, the simulation model

can be used in the future for improving the existing ice generator design and the

development of new ice generator designs.

Mixing of ice-sluny and brine between the buik region and the wail d a c e in a

scraped d a c e kat exchanger is still poorly tmderstwd. The thermal

penetration may be low due to the saturateci properties (kezing point) of ice and

tiquid in a binary solution. Further resewch shouid be performed on this topic,

Page 92

- -

since the eahanced mixing is one of the most importaut factors in increased heat

d e r rates for ice makiug appiicatiom.

The process of ice formation on the scraped surface is not well understood and

needs to be fully investigated. if the ice crystah forming on the surface cannot be

scraped off easily, then the efficiency of heat exchanger would be significantly

reduced. Other pmblems associated with high ice fraction processing include

increased drag force on the scraper blades. Understanding how the ice crystals

fom on the wall in b h u y solutions with a Iow concentration of the secondary

component can be an interesthg field of research.

Sym bol Description Unit

Constant for comlation

Constant for Butterwonh (1975) void W o n

H a t -fer area in chilhg region

H a t uansfa arca in icc rnaillng region

Ha! uansfa arca of the heat achanga

Heat transfa arca ratio bctwcen chilling and ice making regions

Nodal heat transfct ana

Coefficient of icc fiaaion correlation Coefficient in the function for the coefficient of the ice fraction comlation Coefficient in the function for the constant of the icc fraction comlation

Constant for comlation

Coefficient in the hc t ion for the coefficient of the icc fraction comlation

FIuid spccific hcat capacity

Liquid spccific heat capacity

Constant for two-phase fiction multiplia correlation

Ice slwry specific heat capacity

S@6c heat capacity of w a t a

Specinc hcat capacity of mixture

Spccific heat capacity of brine

Specific hait capacïty of brine

Diameter of interual wall

Page 94

Symbol Description

Hydraulic diaoieta

Parameta F in Chen's comlation

Fanning fiction &or

Codon tactor for micro-fuis calculation

Totd two-phase mas flux

Constant of proportionaiity

Gravimionai constant

Intemai wall convective htat tfBi1Sfer coefficient

lackct heat transfa coeficiait (inna shell)

Two-phase forced convcctivc boiling hcat Wnsfa coefficient

S ~ e d nuclcatc boiling hcat transfcr coefficient

Twephase fomd convmivt h m transfa cocfficimt

R c f m a enthalpy of mixture

Convcctivc heu transfer coefficient

Enthalpy of water

Final aithalpy of icc

tnla liquid cnthaipy of rcûigerant

Inla latent hcat of vaporization

Outlet liquid cnthalpy of rc&igaant

Ouikt latent heat of vaporîzation

Latent hcat capacity (liquidlvapour)

muid themial conductivity

Liquid thamai conductivity

l a slurry thermal conductivity

Liquid thamai conductivity

I a aicrmai conductivity

Constant of i a naction correlation

Constant in the function for the d c i a i t of the icc W o n comlation

-

Axial laigh of hcat exchanger ni

Dinancc travclai m

Mass of wami watcr t Mas of mixturr t

Symbol Description

Brine mass flow rate

Inncr wal1 N w l t numba

Rotor mtational spccd

Nusscit number for cimlar channel

Randtl number

Saturatcd p m r c diffacna baxd on saturatcd and wall mpaanires

Total pressure drop

Friction pressure drop

Amlemtion pressure drop

Gravitation pressure drop

Singîc phase prssurr drop

Constant for Buncrworth (1975) void M o n

Wdted pcrirneter

Pccla number of internai wall

Constant for Buiiawonh (1975) void fraction

hcrgy conlent of ice slurry

Enagy contcnt of warm watcr

Encrgy conimt of mixture

H m exchange mc of heat achangcr

Hcat archange ratc in chiliiig region

Hcat atchangc rate in i a making

Hcat exchange rate for the hcat exchanger

Heat flux on the scrapcd wîkœ

Heat transfér rate

Noda1 hcat flux on the saapcd suriàcc

Hcst transfa rate to brine

Hcat tran-Ffér rate to produa iœ

Hcettransaratcatscgmntf

Symbol Description

Reynolds nurnbcr (ddincd with D,)

Two-phase Reynolds numba

Liquid Reynolds numba

Constant for Bunerworth (1WZ) void fracrion

Reynolds numbcr for circular channcl

Nucleatc boilig suppression factor

Diffcrence bctwccn sammcd and wall tanpaanuc

Temp«anirr of mixture

Change in brine tempaaturc h m inla to initial frrtPng

Log mcan tcmpcnuurr diffaencc

T cmpemhuc diffemce bawmi i c t slurry and rrfrigcrant at inla of hcat exchanger Tmpcraturc diffacnu bctwcen ice s l ~ z and rcîiigerant at mitial h z h g

Tempemue diffctcncc krwm i u slurry and rckigcraat at exit of hcat orchanger

Temperanirc of the fluid

Fmzing tempetanue at the exit of hcat exchanger

Tcmpaanire diaaicc bctwmi wall nrrfacc and ilow

Ovaall heat transfcr cocfficicnt in chilling rcgîon

Ovcmll hcat Dansfa coefîicicnt

Avuagc o v d l heat mander coefficient

Bulk avaagc axial fluid velocity

Vclocity in the Channel

Salinity in the Iiquid phase at thc ait of heat ucdiangcr

1ce haion at the ait of htat atchanga

Page 96

Unit Eq-

Page 97

Symbol Description Unit Eq-

x," Outlet vapour quality oinfngsant

4 Initial salinity of he brinc bcing cooled

xr' Inlet vapour quaiity of nûigaant

Panunetci for k, evaluation

Void CiPaion

PmkmCtn for k,, evaluation

Parameter for k,, evaluation

Icc volume fiaction

Two-phase Ection multiplia

Icc s l m y dynamic viscoaty

Liquid dynmic viscosity

Vapour dynamic viscosity

Huid dyiiamic viscosity

Liquid dynamic viscosity

Constant 3.14159 ... Angle h m vatical position

Icc slurry dcnsity

Icc dcnsity

Liquid dcnsity

Liquid dcnsity

Vapour dcnsity

Fiuid daisity

Vapour dcns-ty

Liquid dmsity

Liquid surface tension

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Institute, 1996, pp. 12.1-12.5,

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Air-conditioning Engineen, [m., 1997

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Johnston, Jr, RC, et al, "Heat W e r and pressure drop of refirgerants evaporating m

horizontal tubes," RTHRAE T r m , 70, 1964, pp. 163472

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McCabe, W-L., et al., Unit Operazions of Chemical Engineermg, 5* ed USA: McGraw-Hill Inc.,

1 993.

Murata, K., et al., "Forced Convective Boiling of Nonazeotropic Refrigerant M h r e s inside

Tubes,'' J . Hea! Trmtsfer, 115, August 1993, pp. 680-689.

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Hamion, P., "Heat Transfer in Scraped Surface Heat Exchangers," Chem. Eng. Progr. Symp. Ser.,

29, 1959, pp. 137-139.

Skelland, A.H.P., "Correlation of Scraped-Film Heat Transfer in the Votator," Chem Eng. Sci., 7,

1958, pp. 166-175.

SkelIand, A.H.P., et al., "Heat Transfer in a Water-Cooled Scraped Surface Heat Exchanger,"

Bric. Chem. Eng., 7,1962, pp, 346-353.

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CORRELATION AND EXPElUMENTAL ERROR

The properties of the working materials, bine solution (sodium chloridc solution),

refingerant (fieon 22) and ice are fomulated into functiom. These hctions were used in

the construction of the simulation model. The correlation of the properties is cummarized

in Appendix A- 1.

The hction of exit ice hction was used in the heat and mas balance of brine durhg

the iterative process of the simdation model. This function simplifies the computationd

process of the simulation; heuce, reducing the executing t h e of the program. This

hction is presented in Appendir A-2.

ï he experimental errors due to instrumentation are surnmarized in Appendix A-3.

A-1.1: Propenies of Sodium Chloride Solution

A- 1.2: Properties of F m n Z!

A- 1 3: Properties of Ice

A-2.1 : Function of Exit [ce Fraction

A-3.1 : Experimental E m r

AI. I

AI.2

Al.4

A L I

A3.1

The fimction is vaiid for the range of sait concentration fmm O to lûwt??. Fonns of

fiuiction are as foiiows:

y=l-(l+m)-b (A. 1)

where the pmperty, y, is a function of the salt concentration, x. The salt concentration is

dehed as the weight fraction of salt in soIution. The form of function in eq. A. I is used

for evaluating the fireePng point and the remaining pmperties will follow the fom of

function in eq. A.2. The coefficients to be used with the functions are listed in table A. 1

and the legend for prophes is provideci in table A.2.

Table A. 1: List of coefficients for function of promes (hne solution)

Prouedes Unit Cocficicnts

Table A 2 Legend for pmpecties (bine solution)

T,j, - the mitial ûeezing point

c, - the specific heat apacity

p -thedensity

p -thedynamicviscosity

k - the thermal conductivity

Saturated Pro~erties

The hct ion is valid for the range of saturated temperature fiom -M°C to 18OC. Forms

of fiuiction are as follows:

y =y , +m+bx2 +cx3 64.3)

where the property, y, is a function of the satrrrated temperature, x, except the function of

the saturated temperature. The saturateci temperature is correlated as a fiuiction of the

saturated pressure. Ail properties are correlated in the fùnction form as eq. A.3, except

the fiuiction for (Wdp)) , in the form as eq. A.#. And (Wdp)), is a fiuiction of the

saturated pressure. The coefficients for the function are listed in table A.3 und A.I. A

Iegend of symbols is provided in table A.5.

Table A.3: List of coefficients for function of properties (saturated iU2)

Propcrtia Unit Coeficienfs

Table A.4: Coefficients for hction of (dv Ji+),

Table AS: Legend for pmperties

- the santrated temperature

- the saniniteci pressure

- the liquid density

- the gas specific volume

-the liquid enthalpy

- the gas enthaipy

- the liquid specific heat capacity

- the gas çpecific heat capacity

- the liquid viscosity

- the gas viscosity

- the liquid thermal coaductivity

- the gas thermal conductivity

- the latent heat of capcity

- the surtace tension

SaturatedPtoDerties

The fùnction is vaiid in the range of 239 to 515kpa and -18 to I8OC. Form of function is

as follows:

y = y , + ~ P + ~ T + c P T + ~ P ' +eT2 (A. 5)

where P is the pressure in Pa and T is the temperature in OC. A list of coeEcients for

each of the superheated properties is provided in table A.6.

Table A.6: List of coefficients for function of properties (supeheated R22)

m~ertY ûensity Enthalpy Specific Heat

Unit k g h 3 J/kg J/kg/K

The properties of ice stated in the following are transformeci as fiuictions of the

temperature, T. The density of ice can be evaiuated h m eq. A.6. The thermal

conductivity of ice can be evaiuated h m eq. A.7, which it has been zupplied by

AIexiades and Solomon. The enthaipy of ice is correlateci in eq. A.8 using the data h m

the ASHRAE handbook. The tùnction of the enthaipy is valid within the range of -10 to

0°C.

p=917-(l+173-10' - T ) (A- 6)

k = 224 + 5.975 - lo5 (273 + T ) ' . ' ~ (A- 7)

h = 2.0769. T - 333.43 (A-@

In the previous section ''4.3.2.2 Hcat and Mus Balance Calculation, Brine Side ", the

consideration of exit ice fiaction evaluation was discussed. A function of exit ice f'ractioa

was formed to ease the computational work during the simdating process. The relation

between the initial salt concentration, the exit ice hction and the heat d e r rate is

shown in table A. 7.

Table A.7: Relation for the formation of exit ice fractian function

Based on the data calculated in table A.7, the exit ice fraction cm be correlated as a

function of initial salt concentration and heat d e r rate. The bct ion is stated as

follows:

where X: is the exit ice M o n , q is the heat transfer rate and m is the mass flowrate of

brine. The hc t ion was used to predict the exit ice fiaction with an inlet bine

temperature of O°C. The constants Kr and Kz in eq. A.9 are defined as func5ons of the

initial salt concentration of brine, and they are correlated as follows:

K, = A -(x,' )l + B -(x,')+C (A. IO)

K2 = D -(x,') (A. 11)

where x,' is the initial salt concentration of brine. The constants A, B, C and D were

evduated to be -8.6836, -l.6721,2.9999 and 4.6366, respectively.

Description Emr

Temperature 0.5 deg.C

Pressure 1 psi

THE COMPUTER CODES

The computer codes, written in Fortran 90 language, are provided in this appendix.

These cornputers were used to sunulate the steady-state operation of the scraped-muface

heat exchanger in study. Three sets of computer codes are presented:

a Geomeüic model, the codes for constnrcting the fmite elemental mesh of the heat

exc hanger cross-section.

O Simulation rnodel, the codes for evduating the steady-state operatine conditions

of the heat exchanger.

O Distribution model, the codes for displayhg the cross-sectional temperature

profiles in graphical forms.

The detailed usage and construction of the computer codes have been provided in

Chapter 5: '"Simzdation Model". A table of contents of this appendix is supplieci in the

following:

5 1: Geomewic Model

B-2: Simulation Model

B-3: Dimiution Model

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a i - : . I :r . : 4. f "l m. : E. 7 i l - Y+ - W. Ult

3R .- F9U :. 'A- ."?

-. -- - . .- -- XI:.. -

. . . -:= *,. ..: .. -.. ........ C I . ........ 9.. .-E

: -0 ......

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W iL.-U?d .. l m ; . U . 6 b........

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Ut- -

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a.. u:iYli. 2. UIW. <i

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:r .. -7 , ,- mi. .. C ........ - 2 "'" *. = . . - -,. r* C: mi. r. C . . - m.. C x. m .r

m '"

LI.. C 1 . 1. C<r.., CI - l W... Ci .. m,., C. - 1.

YT .m cm =

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THE EXPERIMENTAL RESULTS

The entire experimental data bank consists of a total of 65 test nuis. The testing

conditions of these 65 test runs were recorded. In addition to these test rum, extensive

testing was also perfomied in order to gain the knowIedge in system control; however, no

data was recorded other than observation was made. The entire data bank was stored in a

soft copy version, where it was not published.

Amongst these 65 recorded test runs, I O srpical experimental results were chosen to

represent the performance and the characteristics of the scraped-surface heat exchanger.

These 10 typical test runs were used to verify the feasibility of the developed simulation

mode1 in the curent research. Two test ntas out of these were conducted in the chilhg

mode of operation, where ice was not produced during the testing.

The overall heat transfer coefficient for =ch of the chosen typical tes nins was

evaiuated. The coefficient was cdcdated using the method stated in the section,

"Overail Heat Transfer Coeficienr ", which is in the chapter, "Results and Discussion".

This coefficient cm be used to denote the general performance of the smped-surface

heat exchanger.

The contents in this appendix are listed below:

C- 1.1 : Experimental Data

C-12: Overal! Heat Transfer Coefficient

The 10 typical experimentai results for îhe verification of the simulation mode! of the

scraped-surface heat exchanger are Listeci in table C. I . The operating conditions and the

resulting conditions are provided. The system parameters as of interest include:

The mass flowrate of bnne and h n 22

The inlet and o d e t temperature in both brine and k o n channels

The pressure via the lieon channels

The inlet salt concentration and the exit ice fiaction in the brine charnel

The rotational speed of the rotor

The heat exchange rate of the heat exchanger

Table C. 1 : Typicai test runs

The range of the typicd test runs is listed in the table C.2. This list shows the range of

the operating parameters, where this broad range tepfesents the most typ id operating

conditions. The typical range of conditions was reasonably enough to indicate the

feasl'b'ity of the cornputer simulation model.

Table C.2: Range of typical operating conditions

TmT CQiDmaa iarrrar - Mass flowrate

Temperature

Salinity

Hass flomate

Temperature

Pressure

Quality

Rotational speed

Temperature

Ice fraction

Tempe rature

Pressure

Quality

k u w E c u x Keat transfer rate

The evaluation of the overd heat transfer coefficient for the scraped-dace heat

exchanger was discussed in the section, "Overail Heat Trausfèr Cwfficient"~ previously

in the chapter, "Resdts and Discussion". The overall heat trader coefficients of the

typical experimental test nms were caicuiated. The required parameters and the heat

tms6er coefficients are preswted in table C.3.

The range of the evaluated overail heat rransfer coefficients for the sample (IO chosen

experimental mm) is summarized m table C I . The average value of the overaii heat

d e r coefficients aod the standard deviation were taken as a representation of the

sample size.

Table C3: Summaq of overall heat transfer coefficients Run Il 12 13 15 19 20 2 1 53 54 5 5

kq/a - - C

W C . crac .

uc. f r a c . HZ

t q / s

C - kPJ

kPa - . f r a c . Ut. frac.

kW

'C

C

kJ/kq/ ' : kW

kY

- i

C

C

i

k W = ' /' C

Table C.4: Range of sample size ~~ - - -

Overall Heat Trans f er coefficient

Maximum 4.117 Minimum 2.722 Average 3.397 Standard deviation 0.503

Range 3.397 ? 15%

The description of the symbols used in fable C.3 is Iisted in table C.5. Since ice was not

produceci in test nms 20 and 21, certain modifications were made. The area ratio. A, for

the case without ice production was taken as innnity in the overidi heat transfer

coefficient model, eq. 5.12. The ta- gradient, a&', was taken as the temperature

gradient beîween the brine exit and k n 22,

TabIe C.5: Legend for sumrnary of ovedl heat trBi1Sfér coefficients

- racia o f chilling to i c e makiag hcat t r a m f e r atea

- specif ic hear capacity af brine

- difference betueen brine i n l e t ccmpcrature and i n i t i a l freezinq palnt

- temperature gradient betneen brwe m i t i a l freezinq and frean 22

- temperacnre qradient between brine e x i t temperature and frean 22

- temperature gradient between brine l a l e t temperature and freon 22

- temperatuse gradrent (sensible. chilLing) - log mean temperature d ~ f f e r e n c e

- temperature gradient Lhtcnt , i ce makinq) - average temperature difference

- mass flowrate of brine

- mass flovrace of Creon 22

- rotational speed of rotor shafr

- in le t pressure of freou 1 2

- exit pressure of frean 22

- heat rtmoval ra te (sensible, chilling)

- haat r m v a l ra te (latent, :ce maklng)

- hear remaval ra te ( t o t a l )

- in le t tcmparacvrc of brinc

- exit temperature of brrne

- in i t ia l freezing point of brine

- amraqa r q e r a t u r e of frean 72

- i n l e t tanpersturc of frcon 22

- e r i c tcmperature a f frcou 22

- average o r e r a l l heat cransfer coeff icient

- exit i ce fract ion of f r e m 22

- inlec vapour quality of I rmn 22

- e x i t vapour qual i ty of frean 22

- i n l e t s a l t concentration af brlne

TEE SIMüLATION RESULTS

Simulation of the cbosen test nuis was perfonned using the cornputer codes in the

uppendix B, "The Cornputer Codes ". The simulation msults for each of the test runs are

presented in the current appendix. Each set of simulation resdts consists of the

followings:

a) Fluids (both brine and fteon 22) conditions at different cross section of the heat

e x c h g e r

b) The heat hansfer rate of each h i d channe1 at different cross section of the heat

exc hanger

C) Wail (kat -fer surface) temperature and heat flux profdes for each working

fluid channet at each axial element

d) Cross-sectionai temperature profiles at each axial eIement

The £hid conditions were plotted against the distauce measured h m the inlet of the

channels. The distance displayed on the plots is the fiaction of the totai lengtb of the heat

exchanger. A totai of 6 graphs are pmvided for the parameters, brindice slurry

temperame, brine sdt concentration, ice M o n , fieon temperature, fieon pressure and

h n vapour qdity.

One plot is provideci for the presentation of the heat d e r rate of each c h e l .

Positive kat tranSfer rate indicates an input of heat iato the channeI, while negative is the

heat loss. For each individual shape of the h n channels, thete is a totd of four

identical channeis. The heat transfer rate of each individuai h n channels represents the

rate through one channet out of four identical channels.

For each individuai fluid channel, the waii temperature and heat flux profiles at each

axial element wete plotted on the sarne graph to show the variation of the profiles. In the

plots, the contour with the highest magnitude of temperature and heat flux is the closest

etement to the inlet of the heat exchanger. The magnitude is decreasing, as it is getting

closer to the exit of the heat exchanger. Since the cross-section of the heat exchanger

was formed with four identical quadrants, only one quarter of the profile was pIotted for

the brine chamel. The zerodegree line is lying between the freon channel a and c, and

the positive angIe is in the anti-clockwise direction. An illustration of the cross-section is

Iocated injgure 3.1. For the profiles of the freon charnels, the zerodegree h e starts at

the point cIosest to the brine chamel. The positive angle represents the antictockwise

direction where negative for clockwise.

in the cross-sectional temperature profiles, twelve temperature zones were used to

illustrate the distribution of temperature throughout the heat exchanger. S a m as

mentioued in the above paragraph, only one quarter of the cross-section plane is shown

since it is made up of four identical quadrants. The name, "ternp-Ol.dct ", represents the

profile of the first element closest to the inlet of the heat exchanger, where "temp-20.m "

is the last element at the exit.

The organhtion of the simulation results are listed in the following:

Test Run 1 I Simulation

Test Run 12 Simulation


Test Run 1 S Simulation






Test Run 55 Smulation

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Documents

STUDY OF SCRAPED-SURFACE HEAT EXCHANGER ICE-MAKING …€¦ · 1.3 ûrganization 2.1 History 2.2 Ice Making Application 3.1 Heat Transfer in Scraped-Surface Heat Exchanger 3.2 Heat