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ANSI/AGMA6011-I03

ANSI/AGMA 6011- I03(Revision of ANSI/AGMA 6011--H98)

AMERICAN NATIONAL STANDARD

Specification for High Speed Helical GearUnits

Copyright American Gear Manufacturers Association Provided by IHS under license with AGMA Licensee=Praxair Inc/5903738101

Not for Resale, 09/14/2005 02:40:06 MDTNo reproduction or networking permitted without license from IHS

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ii

Specification for High Speed Helical Gear UnitsANSI/AGMA 6011--I03[Revision of ANSI/AGMA 6011--H98]

Approval of an American National Standard requires verification by ANSI that the require-ments for due process, consensus, and other criteria for approval have been met by thestandards developer.

Consensus is established when, in the judgment of the ANSI Board of Standards Review,substantial agreement has been reached by directly and materially affected interests.Substantial agreement meansmuchmore than a simplemajority, but not necessarily una-nimity. Consensus requires that all views and objections be considered, and that aconcerted effort be made toward their resolution.

The use of American National Standards is completely voluntary; their existence does notin any respect preclude anyone, whether he has approved the standards or not, frommanufacturing, marketing, purchasing, or using products, processes, or procedures notconforming to the standards.

The American National Standards Institute does not develop standards and will in nocircumstances give an interpretation of any American National Standard. Moreover, noperson shall have the right or authority to issue an interpretation of an American NationalStandard in the name of the AmericanNational Standards Institute. Requests for interpre-tation of this standard should be addressed to the American Gear ManufacturersAssociation.

CAUTION NOTICE: AGMA technical publications are subject to constant improvement,revision, or withdrawal as dictated by experience. Any person who refers to any AGMAtechnical publication should be sure that the publication is the latest available from the As-sociation on the subject matter.

[Tables or other self--supporting sections may be referenced. Citations should read: SeeAGMAAGMA6011--I03,Specification for High SpeedHelical Gear Units, published by theAmerican Gear Manufacturers Association, 500 Montgomery Street, Suite 350,Alexandria, Virginia 22314, http://www.agma.org.]

Approved February 12, 2004

ABSTRACT

This standard includes design, lubrication, bearings, testing and rating for single and double helical externaltooth, parallel shaft speed reducers or increasers. Units covered include those operatingwith at least one stagehaving a pitch line velocity equal to or greater than 35meters per second or rotational speeds greater than 4500rpm and other stages having pitch line velocities equal to or greater than 8 meters per second.

Published by

American Gear Manufacturers Association500 Montgomery Street, Suite 350, Alexandria, Virginia 22314

Copyright 2003 by American Gear Manufacturers AssociationAll rights reserved.

No part of this publication may be reproduced in any form, in an electronicretrieval system or otherwise, without prior written permission of the publisher.

Printed in the United States of America

ISBN: 1--55589--819--X

AmericanNationalStandard



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ANSI/AGMA 6011--I03AMERICAN NATIONAL STANDARD

iii AGMA 2003 ---- All rights reserved

ContentsPage

Foreword iv. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .1 Scope 1. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .2 Normative references 1. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .3 Symbols, terminology and definitions 1. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .4 Design considerations 3. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .5 Rating of gears 7. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .6 Lubrication 9. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .7 Vibration and sound 12. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .8 Functional testing 15. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .9 Vendor and purchaser data exchange 17. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

Bibliography 51. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

Annexes

A Service factors 21. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .B A simplified method for verifying scuffing resistance 24. . . . . . . . . . . . . . . . . . . . .C Lateral rotor dynamics 26. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .D Systems considerations for high speed gear drives 32. . . . . . . . . . . . . . . . . . . . .E Illustrative example 41. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .F Efficiency 44. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .G Assembly designations 47. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .H Purchaser’s data sheet 48. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

Figures

1 Amplification factor 14. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

Tables

1 Symbols used in equations 2. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .2 Recommended accuracy grades 3. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .3 Recommended maximum length--to--diameter (L/d) ratios 4. . . . . . . . . . . . . . . . .4 Hydrodynamic babbitt bearing design limits 6. . . . . . . . . . . . . . . . . . . . . . . . . . . . .5 Dynamic factor as a function of accuracy grade 8. . . . . . . . . . . . . . . . . . . . . . . . .6 Recommended lubricants 10. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .7 Casing vibration levels 15. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .



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ANSI/AGMA 6011--I03 AMERICAN NATIONAL STANDARD

iv AGMA 2003 ---- All rights reserved

Foreword

[The foreword, footnotes and annexes, if any, in this document are provided forinformational purposes only and are not to be construed as a part of ANSI/AGMAStandard6011--I03, Specification for High Speed Helical Gear Units.]

The first high speed gear unit standard, AGMA421.01, was adopted as a tentative standardin October, 1943. It contained formulas for computing the durability horsepower rating ofgearing, allowable shaft stresses, and included a short table of application factors. AGMA421.01was revised and adopted as a full status standard in September, 1947 and issued asAGMA 421.02.

The High Speed Gear Committee began work on the revision of AGMA 421.02 in 1951,which included: classification of applications not previously listed; changing the applicationfactors from “K” values to equivalent Service Factors; revision of the rating formula to allowfor the use of heat treated gearing; and develop a uniform selection method for high speedgear units. This Uniform Selection Method Data Sheet became AGMA 421.03A.

AGMA 421.03 was approved as a revision by the AGMA membership in October, 1954.

The standard was reprinted as AGMA 421.04 in June, 1957. It included the correction oftypographical errors and the addition of a paragraph on pinion proportions and bearingspan, which had been approved by the committee for addition to the standard at theOctober, 1955 meeting.

In October, 1959 the Committee undertook revisions to cover developments in the design,manufacture, and operation of high speed units with specific references to high hardnessmaterials and sound level limits. The revisions were incorporated in AGMA 421.05 whichwas approved by the AGMA membership as of October 22, 1963.

The significant changes of 421.06 from 421.05 were: minimum pitch line speed wasincreased to 5000 feet per minute (25 meters per second); strength and durability ratingswere changed; and some service factors were added. AGMA 421.06 was approved by theHigh Speed Gear Committee as of June 27, 1968, and by the AGMA membership as ofNovember 26, 1968.

ANSI/AGMA 6011--G92 was a revision of 421.06 approved by the AGMA membership inOctober, 1991. The most significant changes were the adaptation of ratings perANSI/AGMA 2001--B88 and the addition of normal design limits for babbitted bearings.ANSI/AGMA 6011--G92 used “application factor” and not “service factor”.

ANSI/AGMA 6011--H98 was a further refinement of ANSI/AGMA 6011--G92. One of themost significant changes was the conversion to an all metric standard. The rating methodswere changed to be per ANSI/AGMA 2101--C95 which is the metric version of ANSI/AGMA2001--C95. To provide uniform rating practices, clearly defined rating factors were includedin the standard (ANSI/AGMA 6011--H98). While some equations may slightly change toconform to metric practices, no substantial changes were made to the rating practice fordurability and strength rating. In addition, minimum pitch line velocity was raised from 25m/s to 35 m/s and minimum rotational speed increased to 4000 rpm.

AGMA has reverted to the term “service factor” in their standards, which was reflected inANSI/AGMA 6011--H98. The service factor approach is more descriptive of enclosed geardrive applications and can be defined as the combined effects of overload, reliability,desired life, and other application related factors. The service factor is applied only to thegear tooth rating, rather than to the ratings of all components. Components are designedbased on the service power and the guidelines given in this standard.

In continued recognition of the effects of scuffing in the rating of the gear sets, additionalinformation on scuffing resistance was added to annex B of ANSI/AGMA 6011--H98.



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v AGMA 2003 ---- All rights reserved

AGMA427.01 has beenwithdrawn. The information found inAGMA427.01was included inannex D of ANSI/AGMA 6011--H98.

ANSI/AGMA 6011--I03 is a further refinement to ANSI/AGMA 6011--H98. Symbols havebeen changed where possible to conform with ANSI/AGMA 2101--C95 and ISO standards.The minimum rotational speed has been increased to 4500 rpm. Helix angle limits havechanged, and a minimum axial contact ratio limit has been added. The L/D limits havechanged, and use of modified leads is now encouraged with the use of predicted rotordeflection and distortion. Bearing load design limits have also changed. For gear toothaccuracy, reference is now made to ANSI/AGMA 2015--1--A01 rather than to ANSI/AGMA2000--A88. The Zn and Yn life factors now have a maximum rather than a minimum limitwhen the number of load cycles exceeds 1010. A table of dynamic factor as a function ofaccuracy grade has been added. References toAGMAoil grades have been removed; nowonly ISO viscosity grades are listed. To facilitate communications between purchaser andvendor, an annex with data sheets has been added.

Realistic evaluation of the various rating factors of ANSI/AGMA 6011--I03 requires specificknowledge and judgment which come from years of accumulated experience in designing,manufacturing and operating high speed gear units. This input has been provided by theAGMA High Speed Gear Committee.

The first draft of AGMA 6011--I03 was made in May, 2001. It was approved by the AGMAmembership in October, 2003. It was approved as an American National Standard onFebruary 12, 2004.

Suggestions for improvement of this standard will be welcome. They should be sent to theAmericanGearManufacturers Association, 500Montgomery Street, Suite 350, Alexandria,Virginia 22314.



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vi AGMA 2003 ---- All rights reserved

PERSONNEL of the AGMA Helical Enclosed Drives High Speed Unit Committee

Chairman: John B. Amendola MAAG Gear AG. . . . . . . . . . . . . . . . . . . . . . . . .

ACTIVE MEMBERS

E. Martin Lufkin Industries, Inc.. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .J.M. Rinaldo Atlas Copco Compressors, Inc.. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .W. Toner Siemens Demag Delaval Turbomachinery, Inc.. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

ASSOCIATE MEMBERS

A. Adams Textron Power Transmission. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .K.O. Beckman Lufkin Industries, Inc.. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .A.S. Cohen Engranes y Maquinaria Arco, S.A.. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .W. Crosher Flender Corporation. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .G.A. DeLange Hansen Transmissions. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .H. Ernst HSB. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .R. Gregory Turner Uni--Drive Company. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .M. Hamilton Flender Graffenstaden. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .L. Hennauer BHS Getriebe GmbH. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .O.A. LaBath Gear Consulting Services of Cincinnati, LLC. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .L. Lloyd Lufkin Industries, Inc.. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .M.P. Starr Falk Corporation. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .F.A. Thoma F.A. Thoma, Inc.. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .F.C. Uherek Flender Corporation. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .U. Weller MAAG Gear AG. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .D.G. Woodley Shell Oil Products U.S.. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .



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1 AGMA 2003 ---- All rights reserved


American National Standard --

Specification for HighSpeed Helical Gear Units

1 Scope

This high speed helical gear unit standard isprovided as a basis for improved communicationregarding:

-- establishment of uniform criteria for rating;

-- guidance for design considerations; and,

-- identification of the unique features of highspeed gear units.

1.1 Application

Operational characteristics such as lubrication,maintenance, vibration limits and testing are dis-cussed. This standard is applicable to enclosed highspeed, external toothed, single and double helicalgear units of the parallel axis type. Units in thisclassification are:

-- single stage units with pitch line velocitiesequal to or greater than 35 meters per second orrotational speeds greater than 4500 rpm;

-- multi--stage units with at least one stage hav-ing a pitch line velocity equal to or greater than 35meters per second and other stages having pitchline velocities equal to or greater than 8 metersper second.

Limits specified are generally accepted designlimits. When specific experience exists for gear unitsof similar requirements above or below these limits,this experience may be applied.

Marine propulsion, aircraft, automotive, andepicyclic gearing are not covered by this standard.

2 Normative references

The following standards contain provisions which,through reference in this text, constitute provisions ofthis American National Standard. At the time ofpublication, the editions indicated were valid. Allstandards are subject to revision, and parties toagreements based on this American National Stan-dard are encouraged to investigate the possibility ofapplying the most recent editions of the standardsindicated below.

ANSI/AGMA 1010--E95, Appearance of Gear Teeth-- Terminology of Wear and Failure

ANSI/AGMA 2015--1--A01, Accuracy ClassificationSystem -- Tangential Measurements for CylindricalGears

ANSI/AGMA 2101--C95, Fundamental Rating Fac-tors and Calculation Methods for Involute Spur andHelical Gear Teeth

ANSI/AGMA 6000--B96, Specification forMeasurement of Linear Vibration on Gear Units

ANSI/AGMA 6001--D97, Design and Selection ofComponents for Enclosed Gear Drives

ANSI/AGMA 6025--D98, Sound for EnclosedHelical, Herringbone, and Spiral Bevel Gear Drives

ISO 14635--1, Gears – FZG test procedures – Part1: FZG test method A/8,3/90 for relative scuffingload carrying capacity of oils

3 Symbols, terminology and definitions

3.1 Symbols

The symbols used in this standard are shown in table1.

NOTE: The symbols and terms contained in this docu-ment may vary from those used in other AGMA stan-dards. Users of this standard should assurethemselves that they are using these symbols andterms in the manner indicated herein.



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Table 1 -- Symbols used in equations

Symbol Term UnitsReferenceparagraph

A Allowable double amplitude of unfiltered vibration mm 7.5Act Amplitude at Nct mm 7.3.3.3AF Amplification factor -- -- 7.3.3.3CSF Service factor for pitting resistance -- -- 5.2CRE Critical response envelope rpm 7.3.3.3cp Specific heat of lubricant kJ/(kg°C) 8.2.5DJ Nominal bearing bore diameter mm Table 4d Pinion operating pitch diameter mm 4.6Fd Incremental dynamic load N 5.3.3Ft Transmitted tangential load N 5.3.3KB Rim thickness factor -- -- 5.4KH Load distribution factor -- -- 5.3.2KHe Mesh alignment correction factor -- -- 5.3.2KHma Mesh alignment factor -- -- 5.3.2KHmc Lead correction factor -- -- 5.3.2KHpm Pinion proportion modifier -- -- 5.3.2Ks Size factor -- -- 5.3KSF Service factor for bending strength -- -- 5.2Kv Dynamic factor -- -- 5.3.3L Face width including gap mm 4.6Ncm Initial (lesser) speed at 0.707× peak amplitude (critical) rpm 7.3.3.3Ncp Final (greater) speed at 0.707× peak amplitude (critical) rpm 7.3.3.3Nct Rotor first critical, center frequency rpm 7.3.3.3Nmc Maximum continuous rotor speed rpm 4.1nL Number of stress cycles -- -- 5.3.1Pa Allowable transmitted power for the gear set kW 5.1Payu Allowable transmitted power for bending strength at unity

service factorkW 5.1

Pazu Allowable transmitted power for pitting resistance at unityservice factor

kW 5.1

PL Power loss kW 8.2.5PS Service power of enclosed drive kW 4.1QLUBE Lubricant flow kg/sec 8.2.5SJ Diametral clearance mm Table 4SM Separation margin rpm 7.3.3.3Umax Amount of residual rotor unbalance g--mm 7.4W Journal static loading kg 7.4Wcpl Half weight of coupling and spacer kg 7.3.3.2Wr Total weight of rotor kg 7.3.3.2YN Stress cycle factor for bending strength -- -- 5.4.1Yθ Temperature factor -- -- 5.3ZN Stress cycle factor for pitting resistance -- -- 5.3.1ZR Surface condition factor for pitting resistance -- -- 5.3ZW Hardness ratio factor for pitting resistance -- -- 5.3∆T Change in lubricant temperature _C 8.2.5σFP Allowable bending stress number N/mm2 5.5σHP Allowable contact stress number N/mm2 5.5



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3.2 Nomenclature

The terms used, wherever applicable, conform to thefollowing standards:

AGMA 904--C96, Metric Usage

ANSI/AGMA 1012--F90, Gear Nomenclature, Defi-nitions of Terms with Symbols

ISO 701, International gear notation – Symbols forgeometrical data

4 Design considerations

This standard should be used in conjunction withappropriate current AGMA standards. Externalloads must be considered as acting in directions androtations producing the most unfavorable stressesunless more specific information is available.Allowances must be made for peak loads.

4.1 Service power, PS

Service power of an application is defined as themaximum installed continuous power capacity of theprime mover, unless specifically agreed to by thepurchaser and vendor. For example, for electricmotors, maximum continuous power will be themotor nameplate power rating multiplied by themotor service factor.

For gear units between two items of driven equip-ment, service power of such gears should normallynot be less than item (a) or (b) below, whichever isgreater.

a. 110 percent of the maximum power requiredby the equipment driven by the gear;

b. maximum power of the driver prorated be-tween the driven equipment, based on normalpower demands.

If maximum torque occurs at a speed other thanmaximum continuous speed, this torque and itscorresponding speed shall be specified by thepurchaser. Maximum continuous speed, Nmc, isnormally the speed at least equal to 105% of thespecified (or nominal) pinion speed for variablespeed units and is the rated pinion speed forconstant speed units.

All components shall be capable of transmitting theservice power.

4.2 High transient torque levels

Where unusual torque variations develop peakloads which exceed the application power by a ratiogreater than the service factor, CSF or KSF, specifiedfor the application, the magnitude and frequency ofsuch torque variations should be evaluated withregard to the endurance and yield properties of thematerials used. See annex D and also ANSI/AGMA2101--C95, subclause 16.3.

4.3 Torsional and lateral vibrations

When an elastic system is subjected to externallyapplied, cyclic or harmonic forces, the periodicmotion that results is called forced vibration. For thesystems in which high speed gears are typicallyused, two modes of vibration are normally consid-ered.

a) Lateral or radial vibration, which considersshaft dynamic motion that is in a direction perpen-dicular to the shaft centerline; and

b) Torsional vibration, which considers the am-plitude modulation of torque measured peak topeak referenced to the axis of rotation.

In certain cases, axial or longitudinal vibration mightalso be considered.

Because of the wide variation of gear drivensystems, clause 7 of this standard outlines areaswhere proper assessment of the system may benecessary. In addition, appropriate responsibilitybetween the vendor and purchaser must be clearlydelineated.

4.4 Tooth proportions and geometry

Any practical combination of tooth height, pressureangle and helix angle may be used. However, it isrecommended that the gears have a minimumworking depth of 1.80 times the normal module, amaximum normal pressure angle of 25 degrees, ahelix angle of 5 to 45 degrees, and a minimum axialcontact ratio of 1.1 per helix.

4.5 Recommended accuracy grade

Table 2 presents recommended ANSI/AGMA2015--1--A01 accuracy grades as a function of pitchline velocity. Based on experience and application,other accuracy grades may be appropriate.



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Table 2 -- Recommended accuracy grades

Pitch linevelocity, m/s

ANSI/AGMA 2015--1--A01accuracy grade

35 –100 A5100 – 160 A4Over 160 A3

4.6 Pinion proportions

Table 3 presents maximum length--to--diameter (L/d)ratios for material hardeningmethods in current use.The L/d values shown in table 3 apply to helical gearswhen designed to transmit the service power.Generally, higher L/d ratios are permitted whenanalytical load distribution methods are employedthat yield load distribution values, KH, that are lessthan the value calculated per 5.3.2 at the maximumL/d ratio per table 3. A detailed analytical methodshould include, but not be limited to, bending andtorsional deflection and thermal distortion.

Table 3 – Recommended maximum length--to--diameter (L/d) ratios

Maximum L/d ratioHardeningmethod

Doublehelical

Singlehelical

Through hardened 2.2 1.6Case hardened 2.0 1.6NOTE:L = face width including gap, mm;d = pinion operating pitch diameter, mm

No matter what the L/d ratio is, if the combination oftooth and rotor deflection and distortion exceeds 25mm for through hardened gears, or 18 mm for casehardened gears, then an analytically determinedlead modification should be applied in order toreduce the total mismatch to a magnitude belowthese values. Determination of the combined toothand rotor deflection shall be based on the servicepower. The modification is intended to provide auniform load distribution across the entire facewidth.Working flanks of the pinion or gear wheel should bemodified when necessary to compensate for torsion-al and bending deflections and thermal distortion.Gears with pitch line velocities in excess of 100 m/sare particularly susceptible to thermal distortion.Consideration should be given to the relationship oflead modifications to gear tooth accuracy.

When a higher L/d ratio than tabulated in table 3 isproposed, the gear vendor shall submit justification

in the proposal for using the higher L/d ratio.Purchasers should be notified when L/d ratiosexceed those in table 3. When operating conditionsother than gear rated power are specified by thepurchaser, such as the normal transmitted power,the gear vendor shall consider in the analysis thelength of time and load range at which the gear unitwill operate at each condition so that the correct leadmodification can be determined. When modifiedleads are to be furnished, purchaser and vendorshall agree on the tooth contact patterns obtained inthe checking stand, housing or test stand.

4.7 Rotor construction

Several configurations may be applied in theconstruction of rotors. Themost commonly used arelisted below:

a) Integral shaft and gear element. This con-figuration is commonly used for pinions, smallergears, or rotating elements operating above apitch line velocity of 150 meters per second. Thepinion or gear, integral with its shaft, is machinedfrom a single blank;

b) Solid blank shrunk on ashaft. The shrink fitmay be used either with or without a mechanicaltorque transmitting device (such as key or spline).When no torque transmitting device is used, theshrink fit must provide ample capacity to transmittorque when considering centrifugal and thermaleffects. When a torque transmitting device isused, the shrink fit must provide ample locationsupport when considering centrifugal and thermaleffects;

c) Fabricated gear. A forged rim is welded di-rectly to the fabricated substructure producing aone--piece welded gear. The shaft may be a partof theweldment. Fabricated gears should be ana-lyzed to consider centrifugal and thermal stressesand fatigue life. Maximum pitch line velocity forwelded gear construction is 130 meters per sec-ond;

d) Forged rim shrunk onto a substructure.The substructure may be forged, cast, or fabrica-ted. The shaft may be a part of the substructure.Shrunk rims shall consider stresses and torquetransmitting capacity due to fit, centrifugal, andthermal effects (refer to item b). The normal de-sign limit for this type of construction is 60 metersper second.

Combinations of the above are often used onmultistage units.



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Stresses and deflections at high speeds often dictatelimits for a specific type of construction. Highpitchline velocity, especially when combined withhigh loads, may require special material specifica-tions and/or testing. Construction features such asholes in the gear body should be analyzed for theirinfluence on the stress. The influence of real orvirtual inclusions and/or cracks may need to beconsidered using the methods of fracture mechan-ics, with testing of the material to ensure that thereare no inclusions greater than the assumed maxi-mum. Overall, a careful analysis of actual operatingstresses and deflection should be made to ensurereliable operation.

4.8 Gear housing

The gear housing should be designed to provide asufficiently rigid enclosed structure for the rotatingelements that enables them to transmit the loadsimposed by the system and protects them from theenvironment in which they will operate. Thevendor’s design of the housing must provide forproper alignment of the gearing when operatingunder the user’s specified thermal conditions, andthe torsional, radial and thrust loadings applied to itsshaft extensions. In addition, it should be designedto facilitate proper lubricant drainage from the gearmesh and bearings.

The user’s design of the supporting structure mustmaintain proper and stable alignment of the gearing.Alignment must consider all specified torsional,radial and thrust loadings, and thermal conditionspresent during operation.

4.8.1 Special housing considerations

Certain applications may be subjected to operatingconditions requiring special consideration. Some ofthese operating conditions are:

-- temperature variations in the vicinity of thegear unit;

-- relative thermal growth between mating sys-tem components;

-- environmental elements that will attack theunit housing, rotating components, bearings or lu-bricant;

-- inadequate support for the housing;

-- high pitch line velocities which may affect lu-bricant distribution, create excessive temperaturerise, or cause other adverse conditions.

4.8.2 Shaft seals

Where shafts pass through the housing, the hous-ings shall be equipped with seals and deflectors thatshall effectively retain lubricant in the housing andprevent entry of foreign material into the housing.Easily replaceable labyrinth--type end seals anddeflectors are preferred. The seals and deflectorsshall be made of nonsparking materials. Lip--typeseals have a very finite life and can generate enoughheat at higher speeds to be a fire hazard. Surfacevelocity should be kept within the seal manufactur-er’s conservative recommendation.

4.9 Bearings

Proper design of bearings is critical to the operationof high speed enclosed drive units. The bearingdesign shall consider normal service power.

Radial bearings are normally of the hydrodynamicsleeve or pad type. Thrust bearings are usually flatland, tapered land, or thrust pad type. Rollingelement bearings are occasionally used whenspeeds are at the very low end of the high speedrange. Bearing design shall consider start--up andunloaded conditions, as well as normal servicepower.

4.9.1 Hydrodynamic bearings

Hydrodynamic bearings shall be lined with suitablebearing material. Tin and lead based babbitts (whitemetal) are among the most widely used bearingmaterials. Tin alloy is usually preferred over leadalloys because of its higher corrosion resistance,easier bonding, and better high temperature charac-teristics. Hydrodynamic bearings shall have a rigidsteel or other suitable metallic backing, and beproperly installed and secured in the housing againstaxial and rotational movement. Bearings aregenerally supplied split for ease of assembly.Selection of the particular design (sleeve, pad typeor land bearing) shall be based on evaluation ofsurface velocity, surface loading, hydrodynamic filmthickness, calculated bearing temperature, lubricantviscosity, lubricant flow rate, and bearing stability.

Heat is generated at running speeds as a result oflubricant shear. Temperature is regulated by control-ling the lubricant flow through the bearing andexternal cooling of the lubricant. The anticipatedpeak babbitt temperature as related to bearinglubricant discharge temperatures should be keptwithin a range that is compatible with the bearingmaterial and lubricant characteristics. See table 4for design limits.



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Table 4 -- Hydrodynamic babbitt bearing design limits1)

Type of bearing

Projected unitload,3)N/mm2

Minimum lubricantfilm thickness,

mm

Bearing metaltemperature,2) 3)

°°°°C

Maximumvelocity,3)

m/sRadial bearing-- Fixed geometry 3.8 0.020 115 100-- Tilting pad 4.2 0.020 115 125Thrust bearings-- Tapered land 2.5 0.020 115 125-- Flat face 0.5 N/A 115 50-- Tilt pad 3.5 0.015 115 125NOTE: Table limits will generally not occur all together; one parameter alone may dictate the design.1) Limits are for babbitt on steel backing. When other materials are used, established limits for thesematerials are per-missible. Bearing clearances should be chosen to yield proper temperature, high stiffness and stability, as well as to en-sure adequate clearance to copewith thermal gradients, whether steady, static, or transient. The average ratio of diame-tral clearance (SJ) to the nominal bore size (DJ), SJ/DJ, for radial bearings is approximately 0.002 mm/mm.2) Bearing temperature measurements are taken in the backing material within 3 mm of the backing material/babbittinterface at the hottest operational zone of the bearing circumference.3) Higher values are acceptable if supported either with special engineering or testing and field experience.

4.9.2 Radial bearing stability

Hydrodynamic radial bearings shall be designedsuch that damaging self generated instabilities (e.g.,half frequency whirl) do not occur at any anticipatedoperational load or speed. Hydrodynamic instabilityoccurs when a journal does not return to itsestablished equilibrium position after being momen-tarily displaced. Displacement introduces an insta-bility in which the journal whirls around the bearingaxis at less than one--half journal speed. Known as“half frequency whirl”, this instability may occur inlightly loaded high speed bearings.

4.9.3 Thrust bearings

Thrust bearings shall be furnished with all gear unitsunless otherwise specified. Thrust bearings aregenerally provided on the low speed shaft for alldouble helical gears and on single helical gears fittedwith thrust collars (see 4.9.4). Thrust bearings aregenerally provided on each shaft for all single helicalgears not fitted with thrust type collars. If the axialposition of any of the shafts depends on itemsoutside the gear unit, the purchaser and vendor shallagree to the arrangement relative to the thrustbearings.

When gear units are supplied without thrust bear-ings, some type of end float limitation shall beprovided at shaft couplings to maintain positive axialpositioning of the gear rotors and connected rotors.Provisions to prevent contact of the rotating ele-

ments with the gear casing shall be provided unlessotherwise specifically agreed to by the purchaser.

The design of a hydrodynamic bearing to sustainthrust is as complicated as the design of a radialhydrodynamic bearing. Complete analysis requiresconsideration of heat generation, lubricant flow,bearing material, load capacity, speed and stiffness.Thrust bearing load capacity should consider thepossibility of torque lock--up loads from couplings.When other external thrust forces are anticipated,the vendor must be notified of their magnitudes.

4.9.4 Thrust collars

Thrust collars (also known as rider rings) may beused to counteract the axial gear thrust developed bysingle helical gear sets.

Thrust collars arranged near each end of the teeth ona single helical pinion and having bearing surfacecontact diameters greater than that of the pinionoutside diameter may be used to carry the gearmesh thrust forces. Typically the thrust collars havea conical shape where they contact a similarlyshaped surface on themating gear rim located belowthe root diameter of the gear. Other designs alsoexist and may be used. Single helical gear setsusing thrust collars may be positioned in the housingin a similar fashion to that of double helical gearelements.

4.9.5 Rolling element bearings

Selection of rolling element bearings shall be basedupon the application requirements and the bearing



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manufacturer’s recommendations and ratingmethods. For normal applications, an L10 life of50 000 hours minimum is required.

4.10 Threaded fasteners

Refer to ANSI/AGMA 6001--D97, Design and Selec-tion of Components for Enclosed Gear Drives,clause 8.

4.11 Shafting

The pinion and gear shafts may normally bedesigned for the maximum bending and maximumtorsional shear stresses at service power (see 4.1)by the appropriate methods and allowable valuesfrom ANSI/AGMA 6001--D97, clause 4, or otherequivalent standards. In some instances, this mayresult in an oversized or undersized shaft.Therefore, an in--depth study using other availableanalysis methods may be required.

5 Rating of gears

5.1 Rating criteria

The pitting resistance power rating and bendingstrength power rating for each gear mesh in the unitmust be calculated. The lowest value obtained shallbe used as the allowable transmitted power of thegear set.

The allowable transmitted power for the gear set, Pa,is determined:

Pa= the lesser ofPazuCSF

andPayuKSF

(1)

where

Pazu is allowable transmitted power for pitting re-sistance at unity service factor (CSF = 1.0);

Payu is allowable transmitted power for bendingstrength at unity service factor (KSF = 1.0);

CSF is service factor for pitting resistance; rec-ommended values are shown in annex A;

KSF is service factor for bending strength; rec-ommended values are shown in annex A.

The service power shall be less than, or equal to, theallowable transmitted gearset power rating:

PS≤ Pa (2)

where:

PS is service power, kW.

It is recognized that all prime movers have overloadcapacity, which should be specified.

5.2 Service factor, CSF and KSF

The service factor includes the combined effects ofoverload, reliability, life, and other application relatedinfluences. The AGMA service factor used in thisstandard depends on experience acquired in eachspecific application.

In determining the service factor, considerationshould be given to the fact that systems develop apeak torque, whether from the prime mover, drivenmachinery, or transitional system vibrations, that isgreater than the nominal torque.

When anacceptable service factor is not known fromexperience, the values shown in annex A should beused as minimum allowable values.

5.3 Pitting resistance power rating

The pitting resistance of gear teeth is considered tobe a Hertzian contact fatigue phenomenon. Initialpitting and destructive pitting are illustrated anddiscussed in ANSI/AGMA 1010--E95.

The purpose of the pitting resistance formula is todetermine a load rating at which destructive pitting ofthe teeth does not occur during their design life.Ratings for pitting resistance are based on theformulas developed by Hertz for contact pressurebetween two curved surfaces, modified for the effectof load sharing between adjacent teeth.

The pitting resistance power rating for gearing withinthe scope of this standard shall be determined by therating methods and procedures of ANSI/AGMA2101--C95, clause 10, when using service factors,with the following values:

ZW is hardness ratio factor, ZW = 1.0;

Yθ is temperature factor, Yθ = 1.0;

Ks is size factor, Ks = 1.0;

ZR is surface condition factor, ZR = 1.0;

ZN is stress cycle factor (see 5.3.1);

KH is load distribution factor (see 5.3.2);

Kv is dynamic factor (see 5.3.3).

5.3.1 Stress cycle factor, ZN

Stress cycle factor, ZN, is calculated by the lowercurve of figure 17 of ANSI/AGMA 2101--C95, and



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should be based on 40 000 hours of service at ratedoperating speed.

ZN= 2.466 n−0.056L (3)

where

nL is number of stress cycles.

When the number of stress cycles exceeds 1010

(i.e., speed above 4167 rpm for 40 000 hours), ZNshould be less than or equal to 0.68.

If less than 40 000 hours is used for rating, it must bewith the specific approval of the purchaser and mustbe so stated along with the rating.

5.3.2 Load distribution factor, KH

KH is the load distribution factor. Values are to be perANSI/AGMA 2101--C95. The following values shallbe used with the empirical method:

KHmais mesh alignment factor. Use values fromcurve 3, precision enclosed gear units, offigure 7 and table 2 of ANSI/AGMA2101--C95;

KHmcis lead correction factor,KHmc= 0.8;

KHpmis pinion proportion factor,KHpm= 1.0;

KHe is mesh alignment correction factor,KHe = 0.8.

The calculated value of KH shall not be less than 1.1.

NOTE: The above empirical rating method assumesproperly matched leads whether unmodified or modi-fied, teeth central to the bearing span and tooth contactchecked at assembly with contact adjustments as re-quired. If these conditions are not met, or for wide facegears, itmaybedesirable touseananalytical approachto determine load distribution factor. AGMA 927--A01provides one such approach.

5.3.3 Dynamic factor, Kv

Dynamic factors account for internally generatedgear tooth dynamic loads, which are caused by geartooth meshing action at a non--uniform relativeangular velocity.

The dynamic factor is the ratio of transmittedtangential tooth load to the total tooth load, whichincludes the dynamic effects.

Kv=Fd+ Ft

Ft(4)

where:

Fd is incremental dynamic tooth load due to thedynamic response of the gear pair to trans-mission error excitation, N;

Ft is transmitted tangential load, N.

Dynamic forces on gear teeth result from geartransmission error, which is defined as the departurefrom uniform relative angular motion of a pair ofmeshing gears. The transmission error is causedby:

-- inherent variations in gear accuracy asmanufactured;

-- gear tooth deflections which are dependenton the variablemesh stiffness and the trans-mitted load.

The dynamic response to transmission error excita-tion is influenced by:

-- masses of the gears and connected rotors;

-- shaft and coupling stiffnesses;

-- damping characteristics of the rotor andbearing system.

The AGMA accuracy grades per ANSI/AGMA2015--1--A01, specifically tooth element tolerancesfor pitch and profile, and the pitch line velocity maybe used as parameters to guide the selection ofdynamic factors. Within the 1.09 to 1.15 dynamicfactor range, the trend is for Kv to vary in nearly adirect relationship with AGMA accuracy grades fromA5 to A2 as shown in table 5.

Table 5 -- Dynamic factor as a function ofaccuracy grade

ANSI/AGMA 2015--1--A01accuracy grade

Dynamic factor, Kv

A5 1.15A4 1.13A3 1.11A2 1.09

The dynamic factor, Kv, does not account fordynamic tooth loads which may occur due totorsional or lateral natural frequencies. Systemdesigns should avoid having such natural frequen-cies close to an excitation frequency associated withan operating speed, since the resulting gear toothdynamic loads may be very high.

Refer to ANSI/AGMA 2101--C95 for additionalconsiderations influencing dynamic factors.



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5.4 Bending strength power rating

The bending strength of gear teeth is a measure ofthe resistance to fatigue cracking at the tooth rootfillet.

The intent of the AGMA strength rating formula is todetermine the load which can be transmitted for thedesign life of the gear drive without causing root filletcracking or failure.

The gear rim thickness must be sufficient for thecalculated rim thickness factor to be 1.0.

Occasionally, manufacturing tool marks, wear, sur-face fatigue, or plastic flow may limit bendingstrength due to stress concentration around large,sharp cornered pits or wear steps on the toothsurface.

The bending strength power rating for gearing withinthe scope of this standard shall be determined by therating methods and procedures of ANSI/AGMA2101--C95, clause 10, when using service factors,with the following values:

Yθ is temperature factor, Yθ = 1.0;

Ks is size factor, Ks = 1.0;

KB is rim thickness factor, KB = 1.0;

YN is stress cycle factor (see 5.4.1);

Kv is dynamic factor (see 5.3.3);

KH is load distribution factor (see 5.3.2).

5.4.1 Stress cycle factor, YN

Stress cycle factor, YN, is calculated by the lowercurve of figure 18 of ANSI/AGMA 2101--C95, andshould be based on 40 000 hours of service at ratedoperating speed.

YN= 1.6831 n−0.0323L (5)

where

nL is number of stress cycles.

When the number of stress cycles exceeds 1010, YNshould be less than or equal to 0.80.

If other than 40 000 hours is used for rating, it mustbe with the specific approval of the purchaser andmust be so stated along with the rating.

5.5 Allowable stress numbers, σHP and σFP

Allowable stress numbers, which are dependentupon material and processing, are given in ANSI/AGMA 2101--C95, clause 16. That clause also

specifies the treatment of momentary overloadconditions.

Three grades of material have been established.Grade 1 is normal commercial quality steel and shallnot be used for gears rated by this standard. Grade 2is high quality steel meeting SAE/AMS 2301 cleanli-ness requirements. Grade 3 is premium quality steelmeeting SAE/AMS 2300 cleanliness requirements.Both Grade 2 and Grade 3 are heat treated undercarefully controlled conditions. The choice ofmaterial, hardness and grade is left to the geardesigner; however, values of σHP and σFP shall be forgrade 2 materials.

Due consideration should be given to additionaltesting, such as ultrasonic or magnetic particleinspection of high speed gear rotors which aresubject to high fatigue cycles or high stress, or both,during operation.

For details on tooth failure, refer to ANSI/AGMA1010--E95.

5.6 Reverse loading

For idler gears and other gears where the teeth arecompletely reverse loaded on every cycle, use 70percent of the allowable bending stress number, σFP,in ANSI/AGMA 2101--C95.

5.7 Scuffing resistance

Scuffing failure (sometimes incorrectly referred to asscoring) has been known for many years and is aconcern for high speed gear units. When high speedgears are subject to highly loaded conditions andhigh sliding velocities, the lubricant film may notadequately separate the surfaces. This localizeddamage to the tooth surface is referred to as“scuffing”. Scuffingwill exhibit itself as a dull matte orrough finish usually at the extreme end regions of thecontact path or near the points of a single pair ofteeth contact resulting in severe adhesive wear.

Scuffing is not a fatigue phenomenon andmay occurinstantaneously. The risk of scuffing damage varieswith the material of the gear, lubricant being used,viscosity of the lubricant, surface roughness of thetooth flanks, sliding velocity of the mating gear setunder load, and geometry of the gear teeth.Changes in any or all of these factors can reducescuffing risk.

Further information is provided in annex B. Annex Bis not a requirement of this standard. However, it isrecommended that either annex B or some other



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method be used to check for the probability ofscuffing failure. See AGMA 925--A03 for furtherinformation.

6 Lubrication

6.1 Design parameters

High speed gear units shall be designed with apressurized lubrication supply system to providelubrication and cooling to the gears and bearings. Anormal lubricant inlet pressure of 1 to 2 bar is anindustry accepted value. Special applications mayrequire other lubricant pressures. If a gear elementextends below the lubricant level in the gear casing,it is said to be dipping in the lubricant. Dipping at highspeed can result in high power losses, rapidoverheating, possible fire hazard, and should beavoided.

The following minimum parameters should be con-sidered to ensure that proper lubrication is providedfor the gear unit:

-- type of lubricant;

-- lubricant viscosity;

-- film thickness;

-- surface roughness;

-- inlet lubricant pressure;

-- inlet lubricant temperature;

-- filtration;

-- drainage;

-- retention or settling time;

-- lubricant flow rate;

-- cooling requirements.

6.2 Choice of lubricant

Certain lubricant additives, such as those in extremepressure (EP) lubricants, may be removed by finefiltration. Changes to the level of filtration shouldonly be done in consultation with both the gear unitand lubricant manufacturers. Extreme pressurelubricants are not normally used in high speed units.

To avoid dependency on extreme pressure addi-tives, unless otherwise specified, the gear unit shallbe designed for use with a lubricant that fails ISO14635--1 load stage 6. The lubricant used shall passISO 14635--1 load stage 5. When an alternate

lubricant is requested, the vendor shall providecalculations and an experience list to support arequest for an alternate lubricant selection.

6.2.1 Lubricant viscosity

Selection of an appropriate lubricant viscosity is acompromise of factors. In addition, lubricationsystems are oftentimes integrated with other drivetrain equipment whose viscosity requirements aredifferent from the gear unit. This complicates theselection of the lubricant.

Load carrying capacity of the lubricant film increaseswith the viscosity of the lubricant. Therefore, ahigher viscosity is preferred at the gear mesh.Development of an adequate elastohydrodynamiclubricant film thickness and reduction in toothroughness are of primary importance to the life of thegearset. However, in high speed gear units,particularly those with high bearing loads and highjournal velocities, heat created in the bearings isconsiderable. Here, the viscosity must be lowenough to permit adequate cooling of the bearings.

Lubricant viscosity recommendations are specifiedas ISO viscosity grades. Recommendations for highspeed applications are listed in table 6. For turbinedriven speed increasers where the lubrication sys-tem supplies both the bearings and the gear mesh,an ISOVG32 is usually provided for the gear drive. Alubricant with a viscosity index (VI) of 90 or bettershould be employed. Special considerations mayrequire the use of lubricants not listed in table 6. Thegear vendor should always be consulted whenselecting or changing viscosity grades.

Table 6 -- Recommended lubricants

ISOviscositygrade (VG)

Viscosity rangemm2/s (cSt)

at 40°°°°C

Minimumviscosityindex (VI)

32 28.8 to 35.2 9046 41.4 to 50.6 9068 61.2 to 74.8 90100 90.0 to 100.0 90

NOTE:When operating at low ambient temperatures, the lubri-cant selected should have a pour point at least 6°Cbelow the lowest expected ambient temperature.

6.2.2 Synthetic lubricants

Synthetic lubricants may be advantageous in someapplications, especially where extremes of tempera-ture are involved. There are many types of synthetic



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lubricants, and some have distinct disadvantages.The gear vendor should be consulted before usingany synthetic lubricant.

6.3 Lubrication considerations

6.3.1 Ambient temperature

Ambient temperature is defined as the temperatureof the air in the immediate vicinity of the gear unit.The normal ambient temperature range for highspeed gear unit operation is from --10°C to 55°C.The vendor should be informed what the ambienttemperature will be, or if a large radiant heat sourceis located near the gear unit. Furthermore, if lowambient temperature causes the sump temperatureto drop below 20°C at start--up, the vendor should beadvised. Special procedures or equipment, such asheaters, may be required to ensure adequatelubrication.

6.3.2 Environment

If a gear unit is to be operated in an extremely humid,salt water, chemical, or dust laden atmosphere, thevendormust be advised. Special caremust be takento prevent lubricant contamination.

6.3.3 Temperature control

The lubricant temperature control system must bedesigned to maintain a lubricant inlet temperaturewithin design limits at any expected ambient temper-ature or operating condition. Design inlet tempera-ture may vary, but 50°C is a generally acceptedvalue. Lubricant temperature rise through the gearunit should be limited to 30°C. Special operatingconditions, such as high pitch line velocity, high inletlubricant temperature, and high ambient tempera-ture may result in higher operating temperatures.

6.3.4 Gear element cooling and lubrication

The size and location of the spray nozzles is criticalto the cooling and proper lubrication of the gearmesh.

Spray nozzles may be positioned to supply lubricantat either the in--mesh, out--mesh, or both sides of thegear mesh (or at other points) at the discretion of thevendor.

6.3.5 Lubricant sump

The lubricant reservoir may be in the bottom of thegear case (wet sump) or in a separate tank (drysump). In either case, the reservoir and/or gear case

should be sized, vented, and baffled to adequatelydeaerate the lubricant and control foaming. In drysump applications, the external drainage systemmust be adequately sized, sloped and vented toavoid residual lubricant buildup in the gear case.Drain velocities may vary, but 0.3 meters per secondin a half full opening is a generally acceptedmaximum value.

6.3.6 Filtration

A good filtering system for the lubricant is veryimportant. The design filtration level may vary, butfiltration to a 25 micron or finer nominal particle sizeis a generally accepted value. Filtration finer than 25microns is recommended when light turbine lubri-cants are used, particularly for higher operatingtemperatures. ISO 4406 may be used as a morecomplete specification of the oil cleanliness re-quired. An ISO 4406:1999 cleanliness level of17/15/12 is recommended if there is no otherrecommendation from the gear unit manufacturer.To remove the finer particles, systems may beinstalled downstream of the filters. It has been foundthat removing very fine particles can greatly extendlubricant life. It is good practice to locate the filter asnear as possible to the gear unit lubricant inlet.Further, it is recommended to provide a duplex filterto facilitate cleaning of the filter when the unit can notbe conveniently shut down for filter change. Any kindof bypass of the filter is prohibited. A mechanism toindicate the cleanliness of the filter is recommended.Systems that take a portion of the filtered lubricantand further clean it are acceptable.

6.3.7 Drain lines

Location of the drain line must be in accordance withthe vendor’s recommendations. Drain lines shouldbe sized so they are nomore than half full. The linesshould slope down at a minimum of 20 millimeterspermeter and have aminimumnumber of bends andelbows.

6.4 Lubricant maintenance

The lubricantmust be filtered and tested, or changedperiodically, to assure that adequate lubricant prop-erties are maintained.

Prior to initial start--up of the gear unit, the lubricationsystem should be thoroughly cleaned and flushed. Itis recommended that the initial charge of lubricant bechanged or tested after 500 hours of operation.



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6.4.1 Change interval

Unless the vendor recommends different intervals,under normal operating conditions subsequentchange or test intervals should be 2500 operatinghours, or 6 months, whichever occurs first. Ex-tended change periods may be established throughperiodic testing of lubricants. With periodic lubricanttesting and conditioning, it is not uncommon tooperate lubrication systems without lubricantchanges for the life of the gear drive.

6.4.2 Water contamination

Where operating conditions result in water collectingin the lubrication system, the lubricant should beprocessed, or changed as required, to keep watercontent below the lubricant manufacturer’s recom-mendation. Failure to control moisture may result indamage to the gear unit. Some lubricants arehygroscopic (absorb water) and may need specialconsideration to eliminate or control the watercontent and total acid number.

7 Vibration and sound

7.1 Vibration analysis

When the frequency of a periodic forcing phenome-non (exciting frequency) applied to a rotor--bearingsupport system coincides with a natural frequency ofthat system, the system may be in a state ofresonance. A shaft rotational speed at which therotor--bearing support system is in a state ofresonance with any exciting frequency associatedwith that speed, is called a “critical speed”.

Vibration of any component of the gear unit canresult in additional dynamic loads being superim-posed on the normal operating loads. Vibration ofsufficient amplitude may result in impact loading ofthe gear teeth, interference in the gear mesh, ordamage to close clearance parts of the gear unit.Where torque variations exceed 20 percent of therated torque at the service power, themagnitude andfrequency of such torque variations should beevaluated with regard to the endurance properties ofthe materials used.

The types of vibrationwhich are generally of concernfor gear units are the torsional, lateral and axialmodes of the rotating elements, since these canhave a direct influence on the tooth load. Of these,

the two that are normally reviewed analyticallyduring design are the lateral critical speeds of thegear unit rotating shafts and the torsional criticalfrequencies of all connected rotating elements.

7.2 Torsional vibration analysis

Any torsional vibration analysis must consider thecomplete system including prime mover, gear unit,driven equipment and couplings. Dynamic loadsimposed on a gear unit from torsional vibrations arethe result of the dynamic behavior of the entiresystem and not the gear unit alone. Thus, a coupledsystem has to be analyzed in its entirety. A commonmethod used is to separate the system into a seriesof discrete spring connected masses. When appliedto a multi--mass system, this method is known asusing lumped parameters. These parameters aredeveloped into a model in order to analyze thesystemas awhole and solve its torsionalmechanicalvibrations.

It is important to note that this result is only as goodas its model. In fact, the process of lumpingparameters could be the largest source of errors.The result of the torsional system analysis is notwithin the control of the vendor, since the gear unititself is only one of several elements in a coupledtrain.

The gear unit vendor is seldom the system designerand in normal cases the gear unit vendor isresponsible only for providing mass elastic data.The system designer, not the gear vendor, isresponsible for the torsional vibration analysis.

7.3 Lateral vibration analysis

The rating equations used in this standard assumesmooth operation of the rotors. To insure smoothoperation, these rotors should be analyzed for lateralcritical speeds. It is imperative that slow roll,start--up, and shutdown of rotating equipment notcause any damage as the rotating elements passthrough their critical speeds. See annex C.

7.3.1 Undamped lateral critical speed map

An undamped lateral critical speed analysis issufficient in somecases to determine rotor suitability.If this method is chosen as the sole criterion fordetermining the suitability of a rotor, it should bebased upon significant experience in designing highspeed gear drives utilizing this method. It includes alateral critical speed map, showing the undampedcritical speeds versus support stiffness or percent-age of torque load. This graphic display shows all



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applicable loading conditions and no--load testconditions (approximately 10 percent of the ratedtorque) at the maximum continuous speed.

The undamped lateral critical speed map for gearrotors is used to determine potential locations of thecritical speeds by locating the intersection of theprincipal bearing stiffness values with the undampedcritical speeds. If no intersections are indicated, withexperience this can be used to determine rotorsuitability.

Note that these undamped speeds can be signifi-cantly different from the critical speeds determinedfrom a rotor response to unbalance analysis. Thedifferences are due to the cross coupled stiffnessand damping effects from the bearings.

7.3.2 Analytical methods

Coupling moments and shear force transfer effectsbetween rotors with properly designed and installedcouplings will be minimal. As a result, each coupledelement can generally be analyzed independently.The mathematics of this analysis are complex andbeyond the scope of this standard (see C.6.2).Commercial computer software is available andanalysts should assure themselves that the methodthey use gives accurate results for the type of rotorsbeing analyzed. Most high speed rotors aresupported in hydrodynamic journal bearings; there-fore, of equal importance is the method used toanalyze the support (bearing) stiffness and damping.

The analyses should include the following effects onthe critical speeds:

-- bearing--lubricant film stiffness and dampingfor the range of bearing dimensions and toler-ances, load, and speed;

-- bearing structure and gear casing supportstructure stiffness;

-- coupling weight to be supported by each gearunit shaft (the weight of the coupling hub plus 1/2the weight of the coupling spacers). The couplingweight shall be applied at the proper center ofgravity relative to the shaft end. The weight andcenter of gravity will be specified by the purchaserof the coupling;

-- potential unbalance of the gear rotor and cou-pling.

7.3.3 Lateral critical speeds

Lateral critical speeds correspond to resonantfrequencies of the rotor--bearing support system.The basic identification of critical speeds is madefrom the natural frequencies of the system and of theforcing phenomena. If the frequency of any harmon-ic component of a periodic forcing phenomenon isequal to or approximates the natural frequency ofanymode of rotor vibration, a condition of resonancemay exist. If resonance exists at a finite rotationalspeed, the speed at which the peak response occursis called a critical speed. The speed or frequency atwhich these occur varies with the degree of trans-mitted load, primarily as a result of the change instiffness of the bearing lubricant film.

Critical speeds are normally determined using arotor response analysis and are deemed to beacceptable if: (a) the separation margin is greaterthan 20 percent; or (b) the vibration levels are withinthe specified limit and the amplification factor is lessthan 2.5 (see 7.3.3.3).

In some cases a simple undamped lateral criticalspeed analysis may be sufficient to properly analyzethe rotor.

7.3.3.1 Forcing phenomena

A forcing phenomenon or exciting frequency may beless than, equal to, or greater than the synchronousfrequency of the rotor. Potential forcing frequenciesmay include, but are not limited to, the following:

-- unbalance in the rotor system;

-- coupling misalignment frequencies;

-- loose rotor--system component frequencies;

-- internal rub frequencies;

-- lubricant film frequencies;

-- asynchronous whirl frequencies;

-- gear--meshing and side--band frequencies,as well as other frequencies produced by inaccu-racies in the generation of the gear tooth.

7.3.3.2 Rotor response analysis

The rotor response to unbalance analysis is used topredict the damped vibration responses of the rotorto potential unbalance combinations (i.e., criticalspeeds). The critical speeds of a gear rotordetermined from the rotor response analysis shouldbe verified by shop and field test data.

The rotor response analysis should consider thefollowing parametric variations in order to assure



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that the vibrations will be acceptable for all expectedconditions:

1. Unbalance, g--mm

-- midspan unbalance6350 WrNmc

;

-- overhung mass unbalance63 500 Wcpl

Nmc;

-- out--of--phase unbalance63 500 Wcpl

Nmcat cou-

pling, and3175 WrNmc

at the furthermost mass sta-

tion on the gear tooth portion of the gear.

where

Nmc is maximum continuous speed of rotor, rpm;

Wr is total weight of the rotor, kg;

Wcpl is half weight of the coupling and spacer, kg.

2. Gear loading

-- unloaded, or minimum load, or both;

-- 50 percent load;

-- 75 percent load;

-- 100 percent load.

3. Bearing clearances

-- minimum clearance and maximum preload;

-- maximum clearance and minimum preload.

4. Speed range from zero to 130 percent ofmaximum rotor speed.

7.3.3.3 Amplification factor

The amplification factor, AF, is defined as the criticalspeed divided by the band width of the responsefrequencies at the half power point.

AF=Nct

Ncp− Ncm(6)

where

Nct is rotor first critical, center frequency, rpm;

Ncm is initial (lesser) speed at 0.707× peak am-plitude (critical), rpm;

Ncp is final (greater) speed at 0.707× peak am-plitude (critical), rpm.

The response of a critical speed is considered to becritically damped if the amplification factor is lessthan 2.5 (see figure 1).

The shape of the curve in figure 1 is for illustrationonly and does not necessarily represent any actualrotor response plot. In most cases the amplitudedoes not decrease to Ncp (0.707 of peak); thereforecalculate Ncp from the “flip” of Ncm, or use anothermethod such as the amplification factor in the“Handbook of Rotordynamics” by F.F. Ehrich, page4.28.

Shaft speed, rpm

Vibrationam

plitude

Nmc Ncm Nct Ncp

CRE

SMOperatingspeed

Act

0.707 Peak

Key:

Nmc is maximum continuous rotor speed, rpm;

Ncp--Ncm is peak width at the half power point;

AF is amplification factor ;

SM is separation margin;

CRE is critical response envelope;

Act is amplitude at Nct.

=Nct

Ncp− Ncm

Figure 1 -- Amplification factor

7.3.4 Stability analysis

Damped eigenvalues (damped natural frequencies)may occur below 120%maximum rotor speed due toa variation in load, bearing properties, etc. Thesedamped eigenvalues are the frequencies at whichthe rotor will vibrate if there is sufficient energy orinsufficient damping in the system. Therefore, adamped stability analysis is performed to ensure thatthese damped eigenvalues have a large enoughlogarithmic decrement (log dec) to insure stability.The stability analysis calculates the damped eigen-values and their associated logarithmic decrement.



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The rotor should have minimum log dec of +0.1 atany of the damped eigenvalues to be consideredstable.

7.3.5 Mode shape

Each finite resonant frequency has an associatedmode shape. Knowing themode shape that the rotorwill assume when responding to a critical speed isimportant in understanding the consequences ofbearing placement and residual unbalance. In mosthigh speed gear unit rotors, the mode shape of thefirst critical speed is mostly conical with a node pointbetween the bearings, vibration at the bearingsapproximately 180° out of phase, and the point ofhighest vibration at the drive (coupling) end of theshaft. A slight bending shape of the rotor is common.The amplitude at the bearing locations is usually highenough to allow the damping inherent in hydrody-namic journal bearings to limit maximum vibrationamplitudes. However, the location of highestamplitude at the coupling makes most gear unitssensitive to unbalance at this location and extra carein coupling balance is recommended.

7.4 Balance

All gear rotating elements shall be multiplanedynamically balanced after assembly of the rotor.Rotors with single keys for couplings shall bebalancedwith their keyway fitted with a fully crownedhalf--key so that the shaft keyway is filled for its entirelength. The balancing machine shall be suitablycalibrated, with documentation of the calibrationavailable. The rotating elements should be balancedto the level of the following equation:

Umax=6350 WNmc

(7)

where

Umax is amount of residual rotor unbalance,g--mm;

W is journal static loading, kg;

Nmc is maximum continuous speed, rpm.

7.5 Shaft vibration

During the shop test of the assembled gear unitoperating at its maximumcontinuous speed or at anyother speed within the specified range of operatingspeeds, the double amplitude of vibration for eachshaft in any plane measured on the shaft adjacentand relative to each radial bearing shall not exceedthe following value or 50 mm, whichever is less:

A= 2800Nmc

(8)

where

A is allowable double amplitude of unfilteredvibration, micrometers (mm) true peak topeak.

7.5.1 Electrical and mechanical runout

When provisions for shaft non--contact eddy currentvibration probes are supplied on the gear unit,electrical and mechanical runout shall be deter-mined by rolling the rotor in V--blocks at the journalbearing centerline, or on centers true to the bearingjournals, while measuring runout with a non--con-tacting vibration probe and a dial indicator. Thismeasurement will be taken at the centerline of theprobe location and one probe tip diameter to eitherside and the results included with the test report.

7.5.2 Electrical/mechanical runoutcompensation

If the vendor can demonstrate that electrical/me-chanical runout is present, themeasured runoutmaybe vectorially subtracted from the vibration signalmeasured during the factory test. However, in nocase shall the amount subtracted exceed thesmallest of:

-- measured runout;

-- 25 percent of the test level determined from7.5;

-- 6.4 micrometers.

7.6 Casing vibration

During shop no--load test of the assembled geardrive operating at its maximum continuous speed orat any other speed within the specified range ofoperating speeds, casing vibration as measured onthe bearing housing shall not exceed the valuesshown in table 7.

7.7 Vibration measurement

Vibration measurements and instrumentation shallbe in accordance with ANSI/AGMA 6000--B96unless otherwise agreed upon by the purchaser andvendor.

7.8 Sound measurement

Sound level measurement and limits shall be inaccordance with ANSI/AGMA 6025--D98 unlessotherwise agreed upon by the purchaser andvendor.



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Table 7 -- Casing vibration levels

Frequencyrange

Velocity10 Hz --2.5 kHz

Acceleration2.5 kHz --10 kHz

Unfiltered (peak) 4 mm/sec 4 g’sFilteredcomponent

2.5 mm/sec

NOTES:1) The above vibration levels are for horizontal offsetgearunitsonly. Theallowablevibration levels forverticaloffset gears are twice those shown in the table.2) Filtered componentmeans any vibration peakwithinthe frequency range.

8 Functional testing

8.1 General

Each unit conforming to this standard should befunctionally tested at full speed. Additional testsmayalso be done at other speeds. Functional testingprovides a means of evaluating operational charac-teristics of the unit. The procedures may be thevendor’s standard or one agreed upon by the vendorand purchaser.

Functional testing presents an opportunity toevaluate the operational integrity of the design andmanufacture of gear drives. Functional testprocedures provide a means of evaluating the entiregear system for noise, vibration, lubrication, geartooth contact, bearing operating temperatures, bear-ing stability, lubricant sealing, mechanical efficiency,instrument calibration and other unit features, andprovide data that parallels the expected on--lineoperational characteristics.

8.2 Procedures

Functional testing may also include proceduresranging frompartial speed and no load spin testing tofull speed and full power testing. Following testing,the unit may be disassembled for bearing and geartooth contact inspection.

8.2.1 No load testing

The unit under test is normally driven in the samerotational direction and with the same input shaft asin the design application. The output shaft will haveno load applied to it. Test speeds may range frompartial speed to over speed. The test duration shouldbe no less than one hour after temperature stabiliza-tion.

8.2.2 Full speed and partial load testing

The unit under test is normally driven in the samerotational direction and with the same input shaft asin the design application. The output shaft will beconnected to a loading device which applies aresisting torque less than the design full load torque.Test duration should be no less than one hour aftertemperature stabilization.

8.2.3 Full speed and full power testing

Full speed and full power testing can be carried out inthe samemanner as described in 8.2.2 for units withlower operating powers.

Full power testing of units with higher power ratingsmay require back--to--back locked torque testing. Inthis procedure two identical ratio units are shaftcoupled together, input to input and output to output.Full operational torque is applied by disengaging oneof the shaft couplings, rotating the shafts relative toone another until the proper torque is achieved, thenre--engaging the shaft coupling. The unit shafts arethen rotated at full speed. Full power testing durationis usually not less than four hours after temperaturestabilization.

When performing back--to--back locked torque test-ing the following risks should be considered:

-- Bearings with full load applied at the staticcondition will start with full load and no hydrody-namic lubricant film until “some” rotational speedis reached;

-- Gear and pinion teeth with full load applied atthe static condition will start with full load and nolubricant film to separate the teeth until “some”rotational speed is reached. Scuffing may occur;special procedures such as coating of the gearteeth with an EP lubricant may be required (thisproblem may be avoided if the method of torqueapplication allows start up at low torque);

-- Bearings of one unit will be loaded in a direc-tion opposite normal operation;

-- Slave unit bearing loads are in the oppositedirection, stub shafts used to complete the torquepath may have to be removed, and if the gear ele-ments of the slave unit are not flipped end for end,they will be loaded on the flanks that are not nor-mally loaded. Therefore the slave unit, and oftenalso the tested unit, will have to be modified afterthe test;

-- For purposes of this test the slave unit may re-quire a lead and profile modification suitable for



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loading in the testing mode. When the leads aremodified specifically for test, then after back--to--back testing the slave gears may require finalmodification suitable for the contract application.The vendor and purchaser shall agree on the ex-tent of this work. At the conclusion of back--to--back tests, the slave unit will require a test of itsown, since the back--to--back configuration can-not be duplicated for that purpose. The vendorand purchaser shall agree on the test to be per-formed.

8.2.4 Special testing

In the case of very high rotational speeds or multipleinput/output shafts, conventional testing may be-come impractical. In such cases, special testprocedures specific to the application should bedeveloped between the vendor and purchaser.

8.2.5 Power loss testing

When testing for power loss in a high speed gearunit, one method is to measure the heat removed bythe lubricant flowing through the gear unit. Lubricantflow rate and lubricant inlet and outlet temperaturesare measured. Power loss is then calculated using:

PL= QLUBE cp ∆T (9)

where

PL is power loss, kW;

QLUBE is lubricant flow, kg/sec;

∆T is change in lubricant temperature from in-let to outlet, °C;

cp is specific heat of the lubricant, kJ/(kg°C).

Aeration of the lubricant may result in the indicatedflow rate being higher than the actual mass flow, sothe indicated flowmay need to be adjusted to a lowervalue.

Accuracy of the power loss calculation may beimproved if all other heat transfer to or from the gearunit is properly accounted for.

Other methods of measuring power loss may beused, such as the difference in the power in and outas measured with torque meters, if agreed to by thepurchaser and vendor.

9 Vendor and purchaser data exchange

9.1 Rationale for data requirements

In order to promote consistency and reduce errors,recommended information to be furnished to thevendor and data provided by the vendor is specifiedin this section. A detail of the schedule fortransmission of drawings, curves and data should beagreed to at the time of the proposal or order.

The purchaser should promptly review the vendor’sdata when he receives them; however, this reviewdoes not constitute permission to deviate from anyrequirements in the order unless specifically agreedupon in writing. After the data has final approval, thevendor should furnish certified copies in the quantityspecified.

A complete list of all vendor data should be includedwith the first issue of major drawings. This listcontains titles, drawing numbers, and a schedule fortransmission of all data the vendor will furnish.

Inquiry documents should be revised to reflect anysubsequent changes. These changes will result inthe purchaser’s issue of completed, corrected datasheets as part of the order specifications.

9.2 Document identification

Transmittal (cover) letter title blocks or title pagesshould contain the following information, whenavailable:

-- purchaser/user’s corporate name;

-- job/project;

-- equipment item number;

-- inquiry or purchase order number;

-- any other identification specified in the inquiryor purchase order;

-- vendor’s identifying proposal number, shoporder number, serial number, or other referencerequired to completely identify return correspon-dence.

9.3 Data provided by purchaser

To allow the gear unit to be properly selected ordesigned, the vendor must have adequate informa-tion from the purchaser. The following is a guide todata that should be sent along with a request forquotation:

-- a data sheet is provided in annex H. All of thedata on the left hand side of that form should beincluded in the request for proposal;



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-- scope of supply;

-- information on the couplings that will be used;

-- testing requirements;

-- measurement units to be used in drawingsand other communications (SI or U.S.customary);

-- list of applicable standards andspecifications;

-- copies of any applicable purchaser specifica-tions;

-- any other special requirements, such aspainting, shipping, storage or environmentalprotection requirements.

9.4 Proposal data

The following is a guide to proposal data that shouldbe furnished by the vendor:

-- general arrangement or outline drawing foreach gear unit showing overall dimensions;

-- purchaser’s data sheets, with completed ven-dor’s information entered thereon and literature tofully describe details of the offering (a suggesteddata sheet is provided in annex H);

-- if applicable, a list of requested exceptions tothe specifications;

-- schedule for shipment of the equipment, inweeks after receipt of the order, and all approveddrawings;

-- list of recommended start--up spares, includ-ing any items that the vendor’s experience indi-cates are likely to be required;

-- complete tabulation of the utility require-ments, including the required flow rate of lubricantand supply pressure, heat load to be removed bythe lubricant, and nameplate power rating(approximate data shall be defined and identifiedas such);

-- description of tests and inspectionprocedures, as required;

-- when requested, the vendor should furnish alist of the procedures for any special, or optionaltests, that have been specified by the purchaseror proposed by the vendor;

-- any start--up, shut--down, or operating restric-tions required to protect the integrity of the equip-ment;

-- conditions and period of the vendor’swarranty.

9.5 Items needing resolution

The following items normally should be resolvedafter purchase commitment. This may be done at acoordination meeting, preferably at the vendor’splant or by other suitable means of communication.

-- purchase order, vendor’s internal orderdetails and sub--vendor items;

-- any required data sheets;

-- applicable specifications, standards, clarifi-cations and previously agreed upon exceptions;

-- that the system and all its components are inaccordance with specified standards;

-- schedules for transmittal of data, productionand testing;

-- quality assurance program, procedures andacceptance criteria;

-- inspection, expediting and testing;

-- schematics and bills of material (B/Ms) ofauxiliary systems;

-- physical orientation of equipment, shaft rota-tion, piping and auxiliary systems;

-- final coupling selection.

9.6 Contract data

The following lists contract data normally supplied bythe vendor:

a. Certified dimensional outline drawing andparts list, including the following:

-- size, rating and location of all purchaser’sconnections;

-- approximate overall and handling weights;

-- overall dimensions;

-- dimensioned shaft end(s) for couplingmounting(s);

-- height of shaft centerline;

-- dimensions of baseplates or soleplates (iffurnished), complete with the diameter, num-ber and location of bolt holes and thickness ofthe metal through which bolts must pass;

-- shaft position diagram, including recom-mended limits during operation, with allchanges in shaft end position and supportgrowths from an ambient reference or 15°Cnoted;



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-- journal bearing clearances and tolerances;

-- axial rotor float or thrust bearing clearance,as applicable;

-- number of teeth on each gear.

b. When a lubricant system is supplied, a sche-matic, certified dimensional outline drawing, andparts list including the following:

-- control, alarm and trip settings (pressuresand recommended temperatures);

-- utility requirements, including electrical,water and air;

-- pipe and valve sizes;

-- instrumentation, safety devices andcontrol schemes;

-- size, rating and location of all purchaser’sconnections;

-- instruction and operation manuals;

-- maximum, minimum and normal liquid lev-els in the reservoir;

-- quantity of lubricant required to fill reservoirto the normal level.

c. Electrical and instrumentation schematicsand bills of materials, including the following:

-- vibration warning and shutdown limits;

-- bearing temperature warning and shut-down limits;

-- lubricant temperature warning and shut-down limits.

d. Lateral critical speed analysis, which may in-clude any or all of the following:

-- method used;

-- graphic display of bearing and support stiff-ness and their effects on critical speeds (un-damped lateral critical speed map);

-- graphic display of the rotor response to un-balance, including damping (rotor responseanalysis);

-- journal bearing stiffness and dampingcoefficients;

-- damped stability analysis, including identi-fied eigenvalues and associated logarithmicdecrement.

e. Torsional data for the gear unit and any shaftcouplings supplied by the vendor, sufficient for athird party to do a system torsional analysis.

f. When mechanical running test is supplied,test reports, including the following (see clause 8):

-- vibration;

-- lubricant flow and inlet and outlet tempera-tures;

-- bearing temperatures.

g. Nameplates and rotation arrows shall be ofSeries 300 stainless steel or of nickel--copperalloy (Monel or its equivalent) attached by pins ofsimilar material and located for easy visibility. Asa minimum, the following data should be clearlystamped on the nameplate:

-- vendor’s name;

-- size and type of gear unit;

-- gear ratio;

-- serial number;

-- service power, Ps;

-- rated input speed, in revolutions perminute;

-- rated output speed, in revolutions perminute;

-- gear service factor, as defined in thisstandard;

-- purchaser’s item number;

-- number of gear teeth;

-- number of pinion teeth;

-- date of manufacture: month and year unitwas successfully tested.

h. Statement of any special protection requiredfor start--up, operation, and periods of idlenessunder the site conditions specified on the datasheets. The list shall clearly identify the protectionto be furnished by the purchaser, as well as thatincluded in the vendor’s scope of supply.

9.7 Installation manual

When specified by the purchaser, an installationmanual shall be supplied. Any special informationrequired for proper installation design that is not onthe drawings shall be compiled in this manual. Thismanual shall be forwarded at a time that is mutuallyagreed upon in the order. The manual shall containinformation such as special alignment and groutingprocedures, utility specifications (including quanti-ties), and all other necessary installation designdata, including drawings and data specified in 9.6.The manual shall also include sketches that showthe location of the center of gravity and rigging



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provisions, to permit removal of the top half of thecasing, rotors, and subassemblies that have a mass(weight) greater than 140 kilograms.

9.8 Operation, maintenance and technicalmanuals

The vendor shall provide sufficient written instruc-tions and a cross--referenced list of all drawings toenable the purchaser to correctly operate andmaintain all the equipment ordered. This informationshould be compiled in a manual or manuals with acover sheet containing all reference--identifying dataspecified in 9.2, an index sheet containing sectiontitles, and a complete list of referenced and encloseddrawings by title and drawing number. The manualshall be prepared for the specified installation; ageneric manual is not acceptable. This manual shall

be forwarded at a time that is mutually agreed uponin the order. This manual shall contain a section thatprovides special instructions for operation at speci-fied extreme environmental conditions, such astemperatures.

9.9 Recommended spares

When the vendor submits a complete list of spareparts, the list should include spare parts for allequipment and accessories supplied. The vendorshould forward the list to the purchaser promptlyafter receipt of the reviewed drawings and in time topermit order and delivery of the parts before fieldstart--up.

9.10 Special tools

A list of special tools required for maintenance shallbe furnished.



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Annex A(informative)

Service factors

[The foreword, footnotes and annexes, if any, are provided for informational purposes only and should not beconstrued as a part of ANSI/AGMA 6011--I03, Specification for High Speed Helical Gear Units.]

A.1 Purpose

This annex provides detailed instructions for thedetermination and use of service factors for en-closed high speed helical gear units as described inANSI/AGMA 6011--I03.

A.2 Determination of service factors

The determination of service factor is based on theequipment characteristic overload of the gear unit asa result of operation, desired reliability of the gearunit during its design life, and length of time that isconsidered the design life. It relies heavily onexperience acquired in each specific application. Abroad explanation of the factors involved are:

-- The causes of service overloads are brokeninto three broad categories: those produced bythe prime mover, those produced by the drivenequipment, and those resulting from system con-siderations unique to the equipment train;

-- The reliability of a geared system depends onmany factors both internal to the gear unit itselfand external to the unit. Increases in service fac-tor to influence reliability normally take into con-sideration external sources of failure such asabuse and unexpected operating conditions;

-- The desired life of most high speed encloseddrives is usually longer than other types of en-closed drives. At high operating speeds this cantranslate into a very large number of stress cycleson the components.

A.2.1 Prime mover characteristics

Somedifferent types of primemovers are: electric orhydraulic motors, steam or gas turbines, and singleor multiple cylinder internal combustion engines.Each of these prime movers is designed to producesome nominal power, but each will produce thispower with some variation over time. The variationof power output with time may be lower or higherdepending on the prime mover and also the way theprime mover is applied in a particular machinerytrain, but any variation over nominal power is anoverload and must be considered.

A.2.2 Driven equipment characteristics

Driven equipment can generally be divided intorotary and reciprocating types of machines. Rotarymachines generally have smoother power require-ments than reciprocating machines, but each type isunique and the equipment characteristics of eachmust be known to be properly evaluated.

A.2.3 System conditions

The gear unit is a part of a system, and this systemcan have dynamic (vibratory) response to time,varying (dynamic) power transmission that mayoverload the gear unit. This ismost commonly foundas torsional vibration in the rotating shafts, but canbe any vibratory response to dynamic excitingforces. Generally, overloads are assumed to betransmitted with no amplification through the gear.However, when there is a resonant response to adynamic power overload, a much higher load mayoccur at the gear unit.

Thus, the dynamic overloads that are caused byprimemovers and drivenmachinesmay beamplifiedin such a way as to greatly increase their magnitudeat the gear unit, and primarily at the gear toothmesh.The normal rating of gear units and the normalservice factors used assume that these responses(resonances) do not appreciably affect the gear unitload. Therefore, careful system analysis is recom-mended to ensure that no unexpected overloadsdue to resonances are present.

A.2.4 Reliability and life requirements

There is a reliability factor in the power ratingequations, but it deals only with the statistical natureof material testing and probability of failure formaterials at a given stress level. In a gear unit thereare many separate components that may fail, manymodes of failure, and many factors that can contrib-ute to those modes of failure. For this reason,quantifying factors associated with reliability and lifeto account for these external issues can be extreme-ly difficult.

A.3 Service factor table

Service factors have served the industry well whenthey have been identified by knowledgeable and



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experienced gear design engineers. The servicefactors shown in table A.1 have been used withsuccess in the past. These values may be used asgeneral guidelines, but they do not eliminate theresponsibility of defining any unusual system re-quirements that would alter the listed values.

A.3.1 General selection guidelines

There is no way to list all the possible considerationsthat may affect selection of service factors, but thefollowing are some guidelines.

-- Induction electric motors can produce hightorques on start--up. Therefore, on an applicationwith many starts, higher service factors may bewarranted;

-- Electric motors that have electric power inter-rupted and then re--applied before induced mag-

netic fields have dissipated can produce very hightorques;

-- Synchronous electric motors can producevery high torsional forcing functions during start--up. This can cause very high transient torsionaltorques on the gear unit;

-- Generators have extremely high loads whenthey are out of phase with the main system, andacross--the--line electrical shorts can producevery high torque loads. For this reason torque lim-iting devices or higher service factors are advis-able;

-- Brakes or other decelerating devices can pro-duce loads on the gear unit larger than the trans-mitted power.

The list could be much longer, but the intent hereis to give a general idea of items to consider whenselecting service factors.

Table A.1 -- Service factors, CSF and KSF

Service factor, with prime mover

ApplicationSynchronous

motorsInductionmotors

Gas or steamturbine1)

Internalcombustion

engine(multi--cylinder)

BlowersCentrifugal 1.7 1.4 1.6 1.7Lobe 2.0 1.7 1.7 2.0

CompressorsCentrifugal

process gas, except air conditioning 1.6 1.4 1.6 1.6air conditioning service 1.6 1.2 1.4 1.6air or pipe line service 1.7 1.4 1.6 1.7

Rotaryaxial flow -- all types 1.7 1.7 1.7 1.7liquid piston (Nash) 2.0 1.7 1.7 2.0lobe -- radial flow 2.0 1.7 1.7 2.0

Reciprocating3 or more cylinders 2.0 2.0 2.0 2.02 cylinders 2.3 2.0 2.0 2.3

Dynamometer -- test stand 1.3 1.1 1.1 1.3Fans

Centrifugal 1.7 1.4 1.6 1.7Forced draft 1.7 1.4 1.6 1.7Induced draft 2.2 1.7 2.0 2.2Industrial and mine (large with

frequent starts)2.2 1.7 2.0 2.2

Generators and excitersBase load or continuous 1.4 1.3 1.3 1.4Peak duty cycle 1.7 1.4 1.4 1.7

(continued)



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Table A.1 (concluded)

Service factor, with prime mover

ApplicationSynchronous

motorsInductionmotors

Gas or steamturbine1)

Internalcombustion

engine(multi--cylinder)

Paper industryJordan or refiner ---- ---- 1.5 ----Paper machine -- line shaft ---- ---- 1.3 ----

PumpsCentrifugal (all service except as listed

below)1.7 1.3 1.5 1.7

Centrifugalboiler feed ---- 1.7 2.0 ----descaling (with surge tank) ---- 2.0 2.0 ----hot oil ---- 1.7 2.0 ----pipe line 2.0 1.5 1.7 2.0water works 2.0 1.5 1.7 2.0

Reciprocating3 or more cylinders 2.0 2.0 1.7 2.02 cylinders 2.0 2.0 2.0 2.0

Rotaryaxial flow -- all types 1.8 1.5 1.5 1.8gear type 1.8 1.5 1.5 1.8liquid piston 2.0 1.7 1.7 2.0lobe 2.0 1.7 1.7 2.0sliding vane 1.8 1.5 1.5 1.8

Sugar industryCane knives 1.8 ---- 1.5 1.8Crushers 2.0 ---- 1.7 2.0Mills 2.3 ---- 1.7 2.3

NOTES:1) Gas turbines seldom operate at full design power while steam turbines often operate at or above rated power.Appropriate design considerations should be made to assure adequate torque capacity.



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Annex B(informative)

A simplified method for verifying scuffing resistance


B.1 Purpose

This annex provides information concerning thescuffing (scoring) of high speed gear units.

B.2 Scuffing considerations

ANSI/AGMA 6011--I03 is concerned with two failuremodes in gear teeth. They are surface pitting androot bending fatigue failure of the tooth material for agiven number of stress cycles. There is anotherknown failure type: scuffing (sometimes referred toas scoring).

The calculation of the scuffing load capacity is a verycomplex problem. While this type failure has beenknown for many years and mathematical methodshave been devised to assess relative risk (seeAGMA 925--A03), a simplified scuffing criterion issuggested that is suitable for general high speeddesign work.

From the values of tooth loading, pitch line velocity,and viscosity of the lubricant, a condensed loadfunction, F (load), is formed, which, to assurescuffing resistance, must be less than (or equal to)the geometric function, F (geometric). The geomet-ric function is based on gear characteristics such asnumber of teeth of the pinion and gear, centerdistance and gearset ratio. As long as the value ofthe load function, F (load), does not exceed that ofthe geometric function, F (geometric), there isadequate safety against scuffing.

Therefore:

F (load)≤ F (geometric) (B.1)

Load function, F (load):

F (load)= w′Cw[v′]0.2546ν400.22

(B.2)

where

w′ is specific tooth load on the pitch circle, N/mm;

v′ is pitch line velocity, m/s;

ν40 is viscosity of lubricant at 40° C, mm2/s(cSt);

Cw = 1.10 (conservative value);

Cw = 1.15 (nominal value);

Cw = 1.20 (maximum value).

NOTE: Cwvaluesare suggested values. Vendor’s ownexperience may change these values with supportingdata. Value ofCw= 1.20 should only be used if total he-lixdeviationmeetsANSI/AGMA2015--1--A01accuracygrade A3.

Table B.1 -- Lubricant viscosities

ISO viscosity gradeVG

Nominal viscosity at40°C, mm2/sec (cSt)

VG -- 22 22VG -- 32 32VG -- 46 46VG -- 68 68

NOTE: For high speed gearset applications, lubricantviscositymeans light turbineoilwith little or noadditivesbased on a viscosity range of: 32 ≤ ν40 ≤ 68. The stan-dard FZG oil test, ISO 14635--1, gives approximationsfor the lubricant with respect to scuffing tendency.

Geometric function, F (geometric):

F(geometric)=50+ z1+ z2(a)

0.5

A[Cu]

(B.3)

where:

z1 is number of teeth of the pinion;

z2 is number of teeth of the gear;

a is center distance, mm;

A is taken from table B.2;

Cu is taken from table B.2.



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Table B.2 -- Values A and Cu for calculating F (geometric)

αp A Cu @ 1≤≤≤≤ u < 3 Cu @ 3≤≤≤≤ u≤≤≤≤ 10

15 350 95 + 28.6 (3 -- u) 130 -- 10 [112.5 -- (13 -- u)2]0.5

17.5 300 90 + 30 (3 -- u) 120 -- 10 [90 -- (12 -- u)2]0.5

20 300 100 + 33.3 (3 -- u) 130 -- 10 [109 -- (13 -- u)2]0.5

22.5 250 95 + 28.5 (3 -- u) 130 -- 10 [112.5 -- (13 -- u)2]0.5

25 250 105 + 31.4 (3 -- u) 140 -- 10 [133.5 -- (14 -- u)2]0.5

NOTEαp is pressure angle, degrees;u is gear ratio (z2/z1).

B.3 Field of application

The above scuffing criterion is applicable to:

a. High speed gears with a modified addendum(rack shift or x factor) resulting in reasonably ba-lanced sliding and rolling conditions between thetooth flanks at the tip of the pinion and matinggear;

b. Gear tooth accuracy grade, per ANSI/AGMA2015--1--A01, shall be equal to or better than:

A5 for single pitch deviation, fptA5 for total cumulative pitch deviation, FpA4 for total profile deviation, FαA4 for total helix deviation, Fβc. Surface roughness of tooth flanks after grind-ing, Ra ≤ 0.5 µm (20 rms);

d. Basic rack profile with:pressure angle, αP = 20 degaddendum, hap = 1 module.

The working flanks of the pinion or gear shall beprovided with profile modifications to obtain atrapezoidal tooth load distribution along the path ofcontact.

The working flanks of the pinion or gear shall beprovided with longitudinal modification to compen-sate for bending and torsional deflections andthermal deformations of the gear rotors in order toobtain a uniform tooth load distribution over theentire rated face width.

The lubricant used shall pass ISO 14635--1 loadstage 5.

B.4 Scuffing design criteria

As stated, there are no firm criteria for designing toprevent scuffing at this time. However, it is hopedthat the use of methods such as those in this annexand those in AGMA 925--A03 can lead to a set ofdesign criteria. There are other methods forpredicting scuffing and there is no intent to deny thevalidity of any method at this time.

B.5 Conclusion

Predicting scuffing is very important in high speedgearing. It is hoped that industry consensus can bereached on scuffing prediction. To achieve thisconsensus, industry must utilize available methodsand gain experience.



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Annex C(informative)

Lateral rotor dynamics


C.1 Purpose

In the dynamic analysis of a high speed gear box, it isnecessary to verify that the drive is inherently stable,and that any actual harmful critical speeds aresufficiently removed from any operating speed orload range of the equipment. This annex providesinformation on rotor dynamics for high speed geardrives.

C.2 Modes

High speed gear drives are frequently coupled toturbomachinery. Although the gear drive operates atturbomachinery speeds, its dynamic behavior issignificantly different from compressors or turbines.Gear shafting is generally of the rigid rotor design.This means that throughout the operating speedrange of the machine, most vibration that occurs iscaused by shaft displacements in the bearingsystem oil films rather than deflections of the rotor(see figure C.1).

Figure C.1 -- Typical modes of rigid rotor lateralvibration

Typical turbomachinery equipment can passthrough what is called flexural type critical speedswithin their operating speed range. Here the rotorwill actually deflect to create mode shapes similar tothose shown in figure C.2, in addition to any vibrationresulting from shaft displacement in its bearings.

C.3 Bearings

In gear rotor dynamics, bearing oil film stiffnessvaries with speed as well as torque load applied tothe drive. This is quite different from turbomachinerydriven through a high speed flexible coupling where

bearing load is mainly a result of the rotor weight andis therefore constant.

Figure C.2 -- Typical modes of flexural lateralvibration

High speed gear drives use fluid film or sleeve typebearings. They frequently are manufactured withnon--cylindrical bores. Gear drive bearings gener-ally have a large length to diameter ratio to gain thebearing area required to support the torque load aswell as rotor weight loading and still be able tomaintain high efficiencies. This type of bearingdesign lends itself to asymmetrical oil film stiffnessrates in the X and Y directions. High stiffness valuesoccur in the direction of the applied load. Relativelylarge cross coupled stiffness and damping coeffi-cients are common. Bearing cross coupling springand damping, in simple terms, means that, inaddition to a resulting resisting force being gener-ated in the direction of displacement or velocity,another force is created 90 degrees from thedirection of motion. This phenomenon has a morepronounced effect in gear drives than in turboequipment, which frequently uses tilting pad typebearings. For an accurate analysis of a gear drive, acomplete eight element matrix of spring and damp-ing rates should be obtained (see figure C.3).

Stiffness terms:

Kxx is force in X resulting from a displacement inthe X direction, in Newtons per millimeter;

Kxy is force in X resulting from a displacement inthe Y direction, in Newtons per millimeter;

Kyy is force in Y resulting from a displacement inthe Y direction, in Newtons per millimeter;

Kyx is force in Y resulting from a displacement inthe X direction, in Newtons per millimeter.



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Bearing shell

Bearing film Journal

KyyY

X

Kxx Dxx

Dyy

KyxDyx

DxyKxy

Figure C.3 -- Cross coupled bearing schematicrepresentation

Damping terms:

Dxx is force in X resulting from a velocity in the Xdirection, in Newtons per millimeter;

Dxy is force in X resulting from a velocity in the Ydirection, in Newtons per millimeter;

Dyy is force in Y resulting from a velocity in the Ydirection, in Newtons per millimeter;

Dyx is force in Y resulting from a velocity in the Xdirection, in Newtons per millimeter.

Obtaining these coefficients is the first step to anaccurate gear drive rotor dynamics analysis.Sophisticated bearing analysis techniques areavailable to determine these coefficients. A typicalmethod will solve the Reynolds and energy equa-tions over a grid network of the bearing area for theparticular geometry in question by finite differencetechniques. The results from each grid point arenumerically combined to produce the performancecharacteristics of the complete bearing. A detailedheat balance of the bearing system under itsoperating conditions must be performed to ensurethat the actual oil film viscosities are being utilized.This is normally accomplished in an iterative typetechnique, where an assumed temperature is cho-sen for performance calculation and then is com-pared with the final calculated temperaturesresulting from the heat balance. If the two do notagree, a new assumed temperature is chosen andthe process continues in the program until conver-gence occurs (see figure C.4).

C.4 Stability

Astability analysis is required to ensure that the drivewill not exhibit self sustaining non--synchronous

vibrations. Lightly loaded fluid film bearings can getinto sub--synchronous vibration problems, particu-larly in the qualification testing process, which isgenerally a no load test. Oil whirl and oil whip are thenames for this type of problem. This vibration isusually at a frequency of around 0.4 times rotationalspeed. If not properly detected in the analysis of thedrive, undesirable or even destructive vibrationsmay be exhibited in testing or lightly loaded fieldrunning.

QvT3

TG

T2

Drains

Bearingfilms

Externalsource

Bearinggroove

QeTe

Q2

T2

T1

QLG

Qi

TGQeTe

Q1 Q2 Q1

H2 P2

Px

Figure C.4 -- Heat balance model

C.5 Critical speed

A critical speed is defined as the speed at which thepeak response amplitude actually will occur whenthe rotor bearing system is in resonance with aperiodic forcing frequency. There are many possibleforcing frequencies in a gear drive system but theone most likely to excite the system is the harmonicforce generated at rotor rotational speed due tomass imbalance. Gears generally are designed tohave their actual critical speeds above 120 percentof their maximum operating speed. Undamped anddamped natural frequencies may be calculatedbelow running speed. Damping may completelysuppress the response of these modes or signifi-cantly shift the frequency at which these modes willactually experience peak response or critical speedby the above definition. Damping tends to lowercalculated natural frequencies. For simple systemsthey are related by:

(C.1)WdWo= 1 − ξ2

where

ξ is the damping ratio;

Wd is the damped natural frequency;



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Wo is the undamped natural frequency.

(C.2)ξ= DDc

where

D is the actual damping;Dc is the critical damping.

Damping, however, tends to raise the frequency atwhich the actual response amplitude or criticalspeed due to imbalance occurs. For simple systemsthey are related by:

(C.3)WrWo= 1

1 − 2 ξ2where

Wr is the actual response frequency.

The damped, undamped, and response frequencieswill agree only when the damping ratio is small.Large discrepancies will be seen at damping ratioslarger than 0.3. Another way of expressing dampingratio is by a logarithmic decrement which defineshow quickly a vibration will decay with time.

(C.4)Log decrement S= 4 π2 ξ2

1 − ξ2

C.6 Analysis types

There are three main tools used in natural frequencyand critical speed analysis, each having its ownstrengths andweaknesses. They are the undamped

critical speed analysis, the damped critical speedstability type analysis, and the damped unbalanceresponse analysis.

C.6.1 Undamped critical speed analysis

The undamped critical speed analysis is an excellentsimple tool for preliminary evaluation of a rotorbearing system. It allows the analyst to identifyapproximately the magnitude of oil film stiffnessrequired to obtain the desired regime of operation ofthe system (i.e., rigid or flexible rotor design).Approximate mode shapes are obtained. Effective-ness of bearing damping can be seen. If motion ofthe rotor occurs at the bearing, damping will be veryeffective.

If the motion occurs other than at the bearing,damping will be ineffective. While the undampedcritical speed map is a useful tool in estimatingperformance, it is lacking in several major areas.First, it does not consider the cross coupled effects inthe oil film; and second, it does not consider thedirect or cross coupled damping terms. In geardrives, which generally have large damping valuesas well as large cross coupled terms, the result cantend to yield critical speed predictions less than whatan actual machine may exhibit. Lastly, no indicationof stability characteristics is obtained. The mapshould display the effect of load variations. Stiffnessvalues for the range of applied load are generallyplotted on the map (see figure C.5).

103

104

105

105 106 107 108

1 x pinion

1 x bull gear 4000 cpm

8000 cpm

Bearing support stiffness, N/mm

Mode 1

Mode 2

Mode 3

Criticalspeed,cpm

KXX -- 50% LD

KXX -- 75% LD

KXX--100%LD

KYY -- 50% LD

KYY -- 75% LD

KYY--100%LD

Figure C.5 -- Undamped critical speed map



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C.6.2 Damped critical speed analysis

The damping approach is similar to the undampedmap except that it is evaluated using full bearingspring and damping characteristics, including crosscoupling terms.

Damping in gear bearings is significant and the firsttwo mode shapes generally show significantmovement in the bearings, thereby utilizing theavailable damping (see figure C.6). This tends togive a result closer to the real world when evaluated,considering that frequencies with damping ratiosgreater than 0.2--0.3 will not be responsive whereindicated. It gives results which agree very closelywith the damped response analysis for the flexuralmode of vibration which is generally the real criticalspeed where response will occur. This is because oflittle movement at the bearings and correspondingsmall damping in the system for this mode.

Figure C.6 -- Bearing damping

The degree of damping or likelihood of response isshown via logarithmic decrement or damping ratiovalues. See figure C.7. This stability type analysiscan also identify sub--synchronous vibrationpotential such as half

1

2

3

4

5

6

78

910

2000 3000 4000 5000 6000 7000 8000 9000

ξ= 0.43

ξ = 0.46

ξ = 0.47ξ = 0.50

ξ = 0.011 ξ = 0.011 ξ = 0.012ξ= 0.01

ξ= 0.4

Shaft rotating speed (rpm)

ξ = 0 .45

ξ = 0.5

ξ = 0.6

NF1

NF1Rock

Bounce

Bend

NF3

Naturalfrequency,cpmX10

(Max

contspeed)

Matingshaft

Shaftstudied

(Max

contspeed)

3

Figure C.7 -- Damped critical speed map--natural frequency versus rotational speed load



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frequency whirl, which can occur with unloadedgears. Here a growth factor is calculated for eachmode. If the factor has a negative value, the systemis inherently stable. If the value is positive, thesystem may be unstable. This analysis should alsobe performed over the load range if applicable. Thedamped natural frequency analysis yields moreinformation but can be difficult to interpret if one is notfamiliar with evaluating the effect of the dampingratio.

C.6.3 Damped response analysis

The damped response analysis is generally consid-ered to be the most useful of the tools for evaluatingrotor synchronous vibration. It gives excellentcorrelation with actual machines. By definition, acritical speed is the speed which corresponds to the

frequency of the peak in vibration response toexcitation.

The damped response analysis includes all theeffects from both damping and cross coupling. It willnot indicate stability problems. In the analysis, it isgenerally best to specify unbalance forces severaltimes larger than the actual rotor balance specifica-tion allows. Unbalanced force stations must beselected to excite the particular mode of vibration inquestion. The unbalance should be applied atseveral places along the rotor in successive runs toensure that each mode will be excited. Couplingend, midspan, and blind end locations should be runas a minimum. Coupling end unbalance will usuallyexcite the most common mode seen (see figuresC.8, C.9 and C.10).

110

90

70

50

30

100 9000 18 000 27 000

Max AMP 99 mmat 20 800 rpm

Speed (rpm)

Diaam

plitude

(m)

m

Figure C.8 -- Unbalance modeled at coupling

110

90

70

50

30

100 9000 18 000 27 000

MAX AMP 86 mmat 20 800 rpm

Speed (rpm)

Diaam

plitude

(m)

m

Figure C.9 -- Midspan



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110

90

70

50

30

100 9000 18 000 27 000

MAX AMP 98 mmat 20 800 rpm

Speed (rpm)

Diaam

plitude

(m)

m

Figure C.10 -- Blind end

In high speed gear drives with large L/D bearings, itis generally accepted that NF1 (bounce mode) andNF2 (rock mode) are heavily damped and unrespon-sive. When heavily damped (damping ratio greaterthan 0.3), these bearing modes may fall within the20% band width around the rotating speed--naturalfrequency line. The acceptability may be proveneither by response analysis or by the damping ratioof actual damping/critical damping.

A term called the amplification factor determineswhen a response peak is to be treated as a realcritical speed or if the frequency tends to be criticallydamped. Amplification factors less than 2.5 areconsidered to be critically damped.

It is not the normal case to be able to evaluate theaccuracy of a critical speed calculation for a gear

drive. This is because the criticals are usuallydesigned to be at operating speeds higher than therest of the drive may be able to withstand. Bearingtemperature or centrifugal stress considerationsusually limit the maximum operating speed. Theonly thing that can usually be verified is that theactual critical is above design speeds, but not theactual critical speed frequency. This is determinedby not measuring any peak in response over thespeed range of the machine.

Evaluating the undamped and damped naturalfrequencies as well as the damped response analy-sis is the most complete way to determine if a geardrive rotor will have dynamics problems. If only onetool can be available, themost reliable overall resultswill be obtained with the damped response analysis.

Act

0.707 PEAK

RRENct = rotor first critical center frequency, cycles per

minuteNcm = initial (lesser) speed at 0.707 x peak amplitude

(critical)Ncp = final (greater) speed at 0.707 x peak amplitude

(critical)Ncp -- Ncm = peak width at the “half power” pointAF = amplification factor

RRE = resonance response envelopeAct = amplitude at Nct

Vibrationlevel

Nmc NcpNctNcm

=Nct

Ncp− Ncm

Figure C.11 -- Amplification factor



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Annex D(informative)

Systems considerations for high speed gear drives


D.1 Purpose

The need for high mechanical reliability in geareddrives can best be satisfied by a “systems approach”to the entire train of machinery including founda-tions, lubrication, vibration, the forces and momentsassociated with piping, couplings, etc. The purposeof this annex is to point out common problems thatmay occur in geared systems, an explanation ofthese problems, and the possible effects.

It is not the intent of this annex to present detailedmethods of analyzing or solving the problem, nor willthere be any attempt to set design criteria or limits.

D.2 Responsibility

A gear unit is susceptible to a variety of problemswhen it becomes a part of a rotating machinerysystem, the severity of which generally increaseswith speed. Even though these problems aregenerally beyond the vendor’s control, they adverse-ly affect system reliability and/or performance andmay cause damage to the gear unit.

The party having contractual responsibility for sys-tem performance should investigate and resolvethese problems in the design stage and therebyavoid the conflicts that may develop between thecomponent manufacturers and users.

It is recommended that the party having contractualresponsibility for the system analysis involving acritical service gear drive be clearly identified in thespecifications, contract or purchase order. Becauseof the substantial cost involved in a system analysis,and in some cases the system performance, itshould be emphasized that all parties supplyingcomponents to the system have a responsibility tofurnish correct and accurate data so that the analysiswill be meaningful.

D.3 Introduction

It is not uncommon to find daily process systemoperating costs many times the cost of the gear unit.This downtime cost makes it desirable to avoidfailure of any part in the system ---- be it primemover,

coupling, gear, driven equipment, or any othercomponent.

The increasing demands for system “mechanicalreliability” can best be satisfied by a coordinatedtechnical exchange between designer, equipmentsupplier, erecting engineers, and user. The varioussystem analyses, in at least preliminary form, shouldprecede detailed equipment purchase specifica-tions. This sequence will permit the design to bebased on more nearly correct load and operatingconditions.

This coordinated effort can be properly called“system engineering” and is normally performed bythe design agent or his technical representative.

Gear vendors may not have the expertise nor thedetailed information to adequately analyze systemoverload. This function must be performed byspecialists under the responsibility of the systemsengineer.

There is no set format for communicating this data.The required information is the magnitude of over-load and a description of the operational conditionsunder which it occurs, such as when, how long, andnature.

Gear units and couplings can be adversely affectedby one or more system generated problems.Failures that result from these system inducedcauses can be categorized under three main head-ings:

-- those resulting from overstressingcomponent parts, which are grouped under “over-load”;

-- alignment related, such as distortedfoundations or poor alignment with connectedmachinery;

-- those resulting primarily from a lubricationrelated failure.

D.4 Overloads

For the purpose of this discussion overload will bedefined as:

“That load which is in excess of the nominal de-sign point load.”



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Overload can be of momentary duration, periodic,quasi--steady state, or vibratory in nature. Depend-ing on its magnitude and the number of stress cyclesaccumulated at overload, it can be a fatigue or a yieldstress consideration.

Overload on a gear drive can result from internal orexternal causes. Internal cause of overload ---- suchas faulty manufacturing (faults of manufacture) areusually found by routine inspections before the geardrive is put into service. External sources ofoverload result from the operational characteristicsof the system into which the gear drive is placed, andare more complex and difficult to identify.

The gear vendor has little if any control over theexternal influences that produce overload. Thesystem engineer who has overall responsibility forperformance should include, along with output, unitcost, efficiency, etc., the investigation of overloadsas they relate to potential failure, downtime, andsystem reliability.

The following material is intended to assist thesystem analyst by highlighting subjects for hisconsideration, and to establish better communica-tion between system people and the vendor.

D.4.1 Estimated maximum continuous power

Operational overload characteristics of various driv-en equipment vary with the type of machine andshould be considered on an individual basis.

Pump or compressor designers, for example, canpredict the power requirements at the design pointwith fairly good accuracy. However, continuouspower (service power) is a combination of:

-- changes in specific gravity or density of themedia being pumped;

-- carry out;

-- overspeed;

-- variations in pressure ratio across acompressor due to abnormal operating condi-tions.

Changes in specific gravity of the fluid mediumhandled by a pump, or change in density of the gashandled by a compressor, affect the power trans-mitted in direct proportion. On boiler feed pumps, forexample, this occurrence can be encounteredduring startup, upon malfunction of pre--heating

equipment, or during boiler cool--down following afailure.

In the case of air handling centrifugal compressors,design power is usually based on the normalmaximum ambient temperature. Considerationshould be given to cold weather operation since thedensity of air varies with absolute temperature.Compressors handling other gases are usuallyencountered in process systems under greatercontrol where temperature variations are less.However, other variables may become serious. Inrefinery practice, for example, the composition of thegas can vary widely, and in other process work theinlet pressure may not be a fixed value.

Carry out is an expression used by the pump andcompressor industries to indicate performance on ahead curve beyond the so--called design point.Figure D.1 illustrates a typical compressor percent-age performance curve.

It will be noted at 100% speed as the head drops offand flow is increased, power increases to a level ashigh as 115% load. Carry out is an everyday reality.It comes about through such things as improperestimation of system performance during designstages, altered system requirements of existingprocesses, gradual deterioration of processes, sys-tems employing multiple units where shutdown orfailure of one increases the requirements on theremaining units, or through leaks or failures.

Figure D.2 illustrates a similar percentage perfor-mance curve for centrifugal pumps.

Overspeed is just what the name implies, and isobviously limited to applications with variable speedprime movers. Because the power absorption of thedriven machine varies approximately with the thirdpower of speed, overspeed is a large contributor tooverload. Referring again to figure D.1, the perfor-mance curve indicates that at 110% speed and100% flow, power is increased to 125%. Carry out atthis speed can increase the power still further, tolevels approaching 140% of service power.

Normal practice for a turbine driven centrifugal pumpis to set the overspeed trips at 115% design speed.Governor settings are generally established topermit continuous operation between 105% and110% design speed. It should be borne in mind thatoperators can and do reset governors to availthemselves of maximum output of the system,regardless of the original settings.



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140

130

120

110

100

90

60 70 80 90 100 110 120 130 140

Head at110% speed

Power at110% speed

Head at100% speed

Power at100% speed

%Headand%power

% Flow

Figure D.1 -- Typical centrifugal compressor performance curve

140

130

120

110

100

90

60 70 80 90 100 110 120 130 140

Head at110% speed

Power at110% speed

Head at100% speed

Power at100% speed

%Headand%power

% Flow

Figure D.2 -- Typical centrifugal pump performance curve



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D.4.2 Vibratory overloads

An essential phase in the design of a critical servicesystem of rotating machinery is the analysis of thedynamic (vibratory) response of a system to excita-tion forces.

The dynamic response of a system results inadditional loads imposed upon the system andrelative motion between adjacent elements in thesystem. The vibratory loads are superimposed uponthemean running load in the system and, dependingupon the dynamic behavior of system, could lead tofailure of the system components. In a gear unitthese failures could occur as tooth breakage orpitting of the gear elements, shaft breakage orbearing failure.

Due to the backlash between the geared elements ofa gear unit, tooth separation will occur when thevibratory torques in the shafts exceed the averagetorque, resulting in tooth separation and subsequentimpacts. Gear tooth loads due to these impacts canbe several times the vibratory torque in the gearshafts.

A vibratory torque which is synchronized to therotation of a gear element can form a cyclic wearpattern on the gear. This wear, which varies aroundthe circumference on the gear element, results intooth spacing errors of the gear causing noise oreven can become a self--generating excitationwhichreinforces the original excitation.

Vibratory motion of gear unit components can takeup clearances causing interference problems be-tween gearing elements, or between shafting andbearings or seals.

D.4.2.1 Vibration analysis

Any vibration analysis must consider the completesystem including prime mover, gear unit, drivenequipment, couplings and foundations. Dynamicloads imposed upon a gear unit are the result of thedynamic behavior of the total system and not that ofthe gear unit alone. The individual components ofthe system are usually supplied by differentmanufacturers. Therefore, responsibility for per-forming the vibration analysis must rest with thedesigner of the total system or his designated agent.

The vibration analysis must determine all significantsystem natural frequencies and evaluate the systemresponse to all potential excitation sources. If theanalysis indicates a resonant or near resonant

condition, the recommended solution is to shiftnatural frequencies by changing stiffness or massinstead of relying on system damping to limitvibratory amplitudes. Normally, a linear vibrationanalysis is adequate. However, under certainconditions nonlinear responses can occur and thepossibility of their existence should be recognized.

It is also advantageous to perform a preliminaryvibration analysis early enough in the design proce-dure to allow for any changes which might berequired for detuning purposes.

D.4.2.2 Torsional vibration

The vibratory load caused by a steady state torsionalvibration of a system is due to the interaction of aperiodic excitation, and a natural frequency of thesystem. The magnitude of the dynamic load causedby this type of vibration is dependent on threefactors: magnitude of the excitation, amount ofdamping in the system, and proximity of the excita-tion frequency to resonance. Typical sources forsteady state excitation are:

-- internal combustion engines;

-- reciprocating pumps and compressors;

-- pump or compressor impellers.

A torsional vibration in a system can also be causedby a transient excitation which is often called a shockor impact loading. Transient conditions occur due tosudden changes in load or speed, or the acceleratingor decelerating through system natural frequencies,including the A.C. component of synchronous mo-tors during startup.

This type of disturbance will produce oscillations atall the natural frequencies of the system. Theseoscillations will decay and eventually disappear dueto damping. The peak dynamic loads occur during ordirectly after the disturbance and their magnitudesare not substantially reduced by the damping in thesystem. Effects of the transient class of vibration canbe most severe in the case of gear teeth due to theirability to separate, thus producing impact loadingson the teeth.

D.4.2.3 Lateral vibration

Dynamic loads at a gear mesh can be caused by alateral vibration of a gear element in response to anexcitation source. The lateral vibration of a rotorsystem should consider all flexibilities and restraintswhich will influence the vibratory response of therotor. In the case of a rotor system comprised of a



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gear element and shaft, this should include theinfluence of bearings, foundations, couplings,connecting adjacent rotors and the mating gearelement.

The most common sources of lateral excitation in arotor system are unbalance and misalignment.Therefore, care should be given to minimize thesefactors in the design, manufacture and installation ofa rotating system. The lateral response of thesystem should be evaluated based on the designtolerances for system unbalance and misalignment.Consideration must be given to operation in theproximity of lateral natural frequencies becauselarge vibratory loads may result with relatively lowexcitation. Fluid film bearings are generally used tosupport rotors in critical service systems. Thesebearings possess stiffness and damping propertieswhich vary with speed and load. These non--linearproperties should be considered when calculatingthe lateral natural frequencies of the system. Undercertain conditions of operation, these bearings cancause instabilities in the rotor motion which willimpart dynamic loads on the gear mesh.

D.4.2.4 Axial vibration

Dynamic loads on a gear mesh are sometimescaused by what appears to be an axial vibration.This axial motion is most often the response of thegear element to unbalanced thrust forces. Commonsources for these forces are malfunctioning ormisaligned couplings, electric armatures mountedoff their magnetic center, face runout of thrust collarsor compressor wheels, and assembly errors.

D.4.2.5 Vibration measurements and designconsiderations

The results of any theoretical vibration analysis areonly as accurate as themathematical model which isdeveloped to perform the calculations. The correct-ness of the model of the system is dependent on theaccuracy to which the inertia, stiffness, damping andexcitation can be ascertained. Since there is alwaysthe possibility of the actual system respondingdifferently than the theoretical evaluation, consider-ation should be given to physically measuring thevibratory loads in the system at the time of initialstartup.

Obtaining test data related to operational loading ona system has the following advantages:

-- establishes confidence that the rotating sys-tem will perform satisfactorily or indicate areaswhere corrective actions are required prior to asystem failure;

-- provide a basis for evaluation of systems thatmay be designed or manufactured in the future;

-- pinpoint system excitations or non--linear re-sponses which were not considered in anytheoretical evaluation.

In the design stages it is advantageous to providedesign features in the system which would facilitatetesting, such as ground surfaces and proper accesspoints for pickups or strain gages. Also in the systemdesign, if it is feasible, consideration should be givento field modifications that could be made with aminimum of operational downtime if damagingvibratory loads were encountered. An example ofthis would be providing both access to couplings andadditional space for coupling changes for detuningpurposes.

D.5 Alignment

D.5.1 Drive train alignment

A gear unit by the nature of its operation is alwaysconnected to at least two other pieces of equipment.The successful operation of the gear unit is largelydependent on the alignment of these components.There are three distinct types of misalignment whichmust be considered between connecting componentshafting.

-- Parallel offset misalignment ---- when twoshafts are not coaxial, but their axes are parallel;

-- Angular misalignment ---- when two shafts arenot coaxial, and their axes are not parallel;

-- Axial misalignment ---- when the ends of thetwo shafts are not positioned to provide the re-quired shaft separation under operating condi-tions.

Misalignment during operation not only causesvibration, but superimposes bending stress on theshear stress due to transmitted torque. Thesestresses cannot be readily calculated but theywarrant discussion so the designer can take precau-tions to minimize their effect. Perfect alignment isalmost impossible to obtain; therefore, flexiblecouplings are used to minimize the effects of theinherent misalignment.

However, “flexible” couplings, whether of the geartooth, spring elements, flexing disc, or elastomerictype, produce forces and moments on their support-



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ing shafts when operating misaligned. The analyti-cal determination of the magnitude of these forcesand moments is not fully understood. It can begeneralized that:

-- the sense and direction are such that they tryto bring the supporting shafts in line;

-- significant bending moments may be im-posed on supporting shafts;

-- the magnitude of the forces and moments in-creases with larger angularity across the cou-pling;

-- notwithstanding catalog claims for angularcapacity, flexible couplings should not be lookedupon as universal joints; they should be given thebest possible alignment.

The designer, in order to obtain a greatermechanicalreliability of a coupled shafting system must make acomprehensive assessment of the operating align-ment. This is a system study and must include allelements of the system including bedplates and/orfoundations. An accurate evaluation of thermalgrowth for all components from a valid and commonreference line is required. Journal displacementwithin bearings, though generally smaller in magni-tude, should be considered, particularly as it affectscold or static alignment checks. After determiningthe probable magnitude of alignment change fromstatic and cold to dynamic and hot (including anyperiodic cyclic changes that may occur), select acoupling arrangement that provides enough lengthor span between flexible elements to keep angularitylow, in the region of 5 minutes or lower.

A hot alignment check is recommended at the timethe unit is put in service. This should be performedwhen all temperatures have stabilized, and thesystem is transmitting rated power at rated speed.

D.5.2 Foundations

Another kind of alignment problem commonly en-countered in geared systems is the misalignment ofpinion and gear axes due to foundation or bedplatetwistings or deflections. It should be recognized thatgear units require foundations with sufficient rigidityto maintain alignment under operating loads.Reinforced concrete foundations with grouted--insoleplate are generally preferable to fabricated steelbedplates in terms of foundation stiffness, mass anddamping characteristics. A concrete foundation ofadequate section, on good soil or on sufficient piling,

is the best insurance to avoid unequal settling ortwisting from other causes.

Fabricated steel bedplates make convenient ship-ping and handling frames, but are generally de-signed for strength, not rigidity. They are frequentlydesigned without consideration for the variouspiping and/or oil sump thermal expansion. Out--of--door installations on steel bedplates are particularlysubject to cyclic bowing caused by the daily “rise andfall” of the sun.

When steel bedplates are used, the designer shouldendeavor to achieve two things:

-- arrange oil sumps, piping, and weatherprotection to minimize unsymmetrical thermal ex-pansion;

-- thoroughly investigate elastic deformation ofthe bedplate due to piping forces and moments;then design the bedplate to eliminate twisting atthe gear supports.

D.5.3 System piping

The forces and moments imposed on pumps,compressors and turbines by their inlet and dis-charge piping are major factors in deflecting thisequipment and causing operating misalignment. Allefforts should be made to minimize piping effects.Lubricant supply and drain piping for the gear unitshould be given similar consideration.

D.5.4 Installation instructions

The system designer should assemble and integratecomplete and comprehensive installation instruc-tions covering, as a minimum, such things as:

-- soleplate, bedplate, machinery position andleveling details;

-- foundation bolting and grouting details;

-- cold alignment data ---- including method ofmeasuring, relative position, and sequence ofalignment;

-- keying, pinning and torquing details as re-quired;

-- pipe support and flange makeup details;

-- all other relevant details that would otherwisebe left to the judgment of the job site mechanic.

D.6 Additional lubrication considerations

The continued successful operation and long life of agear unit is dependent on the constant supply of alubricating oil of proper quantity, quality, and condi-tion. The lubrication system has five functions toperform:



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-- reduce friction;

-- transfer heat;

-- minimize wear;

-- transfer wear particles;

-- reduce rusting and corrosion.

Failure of the lubrication system to adequatelyperform any one or more of these functions mayresult in premature failure of the gear drive.

D.6.1 Type of lubricant

Two basic types of oils are used to lubricate geardrives:

-- petroleum base;

-- synthetic.

There can be a wide variation in the lubricatingqualities of oils within each of these general types.Oils are compounded to meet specific requirementsfor various applications such as gear oils, bearingoils, internal combustion oils, worm gear oils, etc.Therefore, it is important that an oil be selectedmeeting the recommendations supplied with thegear unit.

Synthetic oils should never be substituted forpetroleum base oils without the gear vendor’sapproval, since these oils not only have differentlubricating qualities, but also may not be compatiblewith materials used in the gear unit.

D.6.2 Lubricant selection

The correct type and viscosity of oil must be suppliedin accordance with the vendor’s recommendations.The friction, wear, film strength and corrosionprotection characteristics of different types of oilscan vary widely. Deviation from the recommendedoil for the gear drive can result in premature wear,failure, or both.

D.6.2.1 Lubricant quality

Lubricating oils for high speed gear units should behigh quality, refined, paraffin base petroleum oils.They must not be corrosive and must be free fromgrit or abrasives. As they are oftentimes subject tolarge flow rates and high operating temperatures,they must have good antifoaming properties.

Oils of a straight mineral type should be used. Highquality rust and oxidation resistance is desirable.Oils with additives which enhance these characteris-

tics should be carefully selected and, if selected,frequently changed to avoid accumulative separa-tion of the additives during operation. Whenexposed to high operating temperatures in excess of90° C, rapid degradation will occur.

D.6.2.2 Viscosity and viscosity index

Oils refined into lubricants are generally derived fromtwo types of crude oil, either paraffin base or napthabase. Paraffin based oils are preferred because theyhave better natural extreme pressure characteristicsand better resistance to “thinning down” at higheroperating temperatures. Naptha based oils, on theother hand, require special additives in order topossess this benefit.

The oil’s resistance to “thinning” is measured by theviscosity index. The higher the index value the betterthe resistance to “thinning”. Oils without additives ofthe paraffin base type usually have VI values ofninety (90) or above, whereas naptha base oils willexhibit lower values, oftentimes between twenty (20)and thirty (30).

D.6.3 Oil film

Gear elements and the supporting bearing systemrequire a continuous supply of properly selected andconditioned oil for survival. An oil film of adequatethickness must be established between the rollingand sliding component surfaces to avoid damagingwear and scuffing and to provide component cooling.

Hydrodynamic and elastohydrodynamic lubricationtheories are commonly used today in analyzing filmthickness in bearings and gear teeth. The oilviscosity has the greatest effect on the film thick-ness. Consequently, failure to use an oil that hasboth the proper viscosity and viscosity index canresult in failure to produce an adequate film thick-ness for the gear teeth and bearings.

Improper oil film thickness may cause severaloperational problems. Lack of oil film or inadequateoil film thickness may cause metallurgical drawingdue to frictional heat of hardened surfaces, destruc-tive wear, scuffing or pitting of the gear teeth, andfrictional melting, plastic flow or failure of thebabbitted bearing surfaces. Increased oil viscosityincreases frictional power losses and thereforeincreases the temperature rise and may produceheat energy beyond the control of the coolingsystem.



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The lubrication system design must successfullyachieve a balance of the viscosity and the oil filmthickness considerations.

D.6.4 Lubricant supply

The oil supply must meet the requirements set forthin the gear vendor’s recommendations.

D.6.4.1 Quantity

The proper flow rate of oil must be supplied to thegear drive to ensure adequate oil film formation onthe rotor elements, and in cases where babbittedbearings are employed, in the bearing journals, toprevent metal to metal contact of the respectiveelements. In addition, sufficient flow must bemaintained to assure adequate cooling. Too small aquantity may cause inadequate distribution resultingin potential overheating, whereas too large a quanti-ty may result in excessive churning of the oil whichmay also result in overheating.

D.6.4.2 Pressurized lubrication systems

When lubrication systems are self contained, thesystem should be designed with a flow capacity of aminimum of 10% greater than that initially required toallow for pump wear, slight bearing wear with normalservice, or change in oil viscosity due to temperaturevariations and change of viscosity with use.

Where pressurized oil is furnished from a centralsupply, operating, alarm and shutdown pressuresmust be in accordance with the gear unit vendor’sspecifications. Pressures lower than that recom-mendedmay result in reduced flow and overheating.Pressures too high may cause excessive churningand possible gearbox flooding, increasing powerloss and also resulting in overheating.

Oil pressure to the gear drive should be measuredeither in the oil passages of the gear unit or at a pointas near to the entry of the unit as possible, thusavoiding the inclusion of pressure losses in thepiping between the point of measurement and theactual gear supply.

D.6.4.3 Lubricant temperature

The gear supplier will normally specify the minimumallowable oil temperature for startup. If tempera-tures lower than this are expected, provisions mustbemade to heat and, if possible, circulate the oil priorto startup. The gear drive must not be operated forextended periods at this minimum startup tempera-ture.

Oil inlet temperature must be in accordance with thevendor’s specifications. A low supply temperaturemay result in a change in viscosity causing higherthan expected temperature rise in the gear unit andimproper oil distribution to the spray jets andbearings.

When the oil supply temperature is higher thanspecified, the oil will be subject to rapid oxidationreducing the life of the oil, and reducing the operatingviscosity resulting in an inadequate oil film. Thiscondition can result in overheating, excessive wearand even failure.

D.6.4.4 Pressurized system components

The system components must be selected andinstalled to avoid problems. The following are somesuggestions to avoid problems:

-- Aeration. Caremust be taken to avoid exces-sive aeration of the oil. Aeration may result inpump cavitation and decrease the volume of oil tocome in contact with the elements of the geardrive;

-- Oil reservoir. The reservoir must be largeenough to allow time for the air to separate fromthe oil. Return lines to the oil reservoir should re-turn below the oil level. This also includes reliefvalve bypass lines and any other return lines.These lines should be located as far away fromthe pump suction line as possible. Baffles proper-ly located in the reservoir will ensure the aeratedreturn oil does not find its way to the suction lineuntil air has had time to escape from the oil;

-- Drain lines. The location of the drain from thegear drive is critical, and the vendor’s recommen-dations should be followed. Drain lines should besized so they run no more than half full of oil. Theline should slope down at a minimum of (20mm/m, 2%) and have a minimum number ofbends and elbows. It is desirable to have a ventlocated in the drain line near the exit from the geardrive to insure proper drainage;

-- Vents. Vents must be carefully located and ofample size to avoid pressure buildup and allowready escape of air from the system without theloss of oil. Vents must be high enough to avoidentry of contaminants from the environment intothe oil. Oftentimes it is desirable to place the ventin the drain line near the exit from the gear drive toensure proper drainage. The oil is filtered prior toreturning to the gear drive as well. In this mannerdirect contamination of the gear drive from the at-mosphere outside is avoided;



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-- Suction lines. These lines should be gener-ously sized to minimize pressure loss. Suctionpressure (net positive suction head) must not beless than that recommended by the pumpmanufacturer. The total suction loss must includethe loss in the piping, valves and fittings, in addi-tion to the distance of the lift. If a check valve isused in the suction line of positive displacementpumps, a pressure limiting device should beinstalled to protect against the effects of reverserotation of the pump;

-- Flushing. Before oil is circulated through thegear drive, a bridge section containing a remov-able screen is fitted between the supply point andthe drain. The systemmust be flushed until thereis no significant accumulation of dirt on thescreen. During flushing the piping should be ham-mer rapped to dislodge foreign particles. Afterflushing is completed, the supply and drain linesare connected to the gear drive.

D.6.4.5 Lubricant condition

Having provided the proper type and grade of oil, it is

also important the oil be supplied and maintained inthe proper condition. Dust, dirt, grit and otherparticles in the oil supply should be eliminated.These foreign matters act as an abrasive in thebearings and gear teeth, causing abrasive wear.The pressurized oil must be supplied through a filteras specified by the gear unit vendor. These filtersystems should be serviced regularly to avoidcirculation of contaminants with the oil and to avoidexcessive pressure drops through the filters whichmay reduce the quantity of oil supplied to the geardrive.

The oil must be maintained in its correct chemicalcondition to properly perform. Foreign matter, dirtand moisture can change the chemical properties ofthe oil. Additives used inmany oils are depleted withuse and require replacement. Since many factorsinfluence the useful life of the oil, its condition shouldbe analyzed on a regular basis to ensure itsproperties are within specification.



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Annex E(informative)

Illustrative example


E.1 Purpose

This annex provides examples based on the as-sumption that the gear set power rating is theminimum component rating. In practice all compo-nent ratings must be calculated to determine thelowest rated component.

E.2 Example #1

E.2.1 Operational parameters

The gearset to be rated transmits power from aninduction motor rated at 2500 kilowatts and 1480RPM to a centrifugal compressor operating at 5000RPM. AnnexA indicates that a service factor of 1.4 isappropriate for this service.

E.2.2 Gearset parameters

The through hardened double helical gearset to berated has the following parameters:

Number of teeth, pinion 53Number of teeth, gear 179Gear speed 1480 rpmModule, normal 3 mmPressure angle, normal 20°Helix angle 29° 32’ 30”Center distance 400 mmOutside diameter, pinion 188.75 mmOutside diameter, gear 623.24 mmNormal circular tooth thickness atreference diameters pinionreference diameters, pinion(182.76 mm) and gear (617.24 4.63 mm(182.76 mm) and gear (617.24mm)Face width 255 mmOverall face (gap included) 300 mmHardness pinion 350 HBHardness gear 300 HBPinion speed 5000 rpmMaterial grade 2Gear quality level A4Cutter whole depth 7.0 mmCutter 1/2 pitch addendum 3.8 mmCutter tip radius 1.28 mm

E.2.3 Rating parameters

The pitting resistance power rating and bendingstrength power rating at unity service factor arecalculated per ANSI/AGMA 2101--C95 equations.With the factors that have a value of one (1.0)deleted, the equations are:

Pazu=ω1 b

1.91× 107ZI

KvKHdw1σHPZN

ZE2

(see ANSI/AGMA 2101--C95, Eq. 27)

Payu=ω1 dw1

1.91× 107b mt YJKvKH

σFPYN1


where:

ω1 = 4998.5 rpm

b = 255 mm

ZI = 0.22656 (see AGMA 908--B89)

dw1 = 182.76 mm

ZN = 0.67313 (pinion)= 0.720 (gear)

ZE = 190 [N/mm2]0.5

YJ pinion = 0.56923 (see AGMA 908--B89)

YJ gear = 0.58766 (see AGMA 908--B89)

mt = 3.4483 mm (3/cos 29°32’30”)

YN = 0.79531 (pinion)= 0.82720 (gear)

Kv = 1.13

KH = 1.2648 (see ANSI/AGMA2101--C95)

CSF= KSF = 1.4 (see annex A)

σHP = 1079 N/mm2 (pinion @ 350 HB)= 958 N/mm2 (gear @ 300 HB)(see ANSI/AGMA 2101--C95, figure 8Grade 2)

σFP = 359 N/mm2 (pinion)(see ANSI/AGMA 2101--C95, figure 9Grade 2)



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σFP = 324 N/mm2 (gear @ 300 HB)(see ANSI/AGMA 2101--C95, figure 9,Grade 2)

Pazu=4998.5 (255) (0.22656)

1.91× 107 (1.13) (1.2648)

= 5163 kW (pinion)

×182.76 (1079) (0.67313)190

2

Pazu=4998.5 (255) (0.22656)

1.91× 107 (1.13) (1.2648)

= 4657 kW (gear)

×182.76 (958) (0.72)190

2

Payu=4998.5 (182.76)1.91× 107

255(3.4483)(0.56923)1.13 (1.2648)

× 359 (0.79531)1

= 4782 kW (pinion)

Payu=4998.5 (182.76)1.91× 107

255(3.4483)(0.58766)1.13 (1.2648)

× 324 (0.8272)1

= 4635 kW (gear)

Pa is the lesser of51631.4

46571.4

47821.4

46351.4

or 3311 kW

E.2.4 Rating conclusions

Pa is equal to the lesser of Pazu or Payu divided by theservice factor, or Pa = 4635 1.4 = 3311 kW. This isgreater than the service power of 2500 kW.

E.3 Example #2

E.3.1 Operational parameters

The gearset to be rated transmits power from a gasturbine rated at 15 MW and 8215 RPM to an electricgenerator operating at 3600 RPM on a base loadcycle. The service factor is 1.3.

E.3.2 Gearset parameters

The carburized and case hardened double helicalgearset to be rated has the following parameters:

Number of teeth, pinion 39Number of teeth, gear 89Gear speed 3600 rpmModule, normal 6 mmPressure angle, normal 20°Helix angle 23°45’Center distance 421.6 mmOutside diameter, pinion 268.8 mm

Outside diameter, gear 595.1 mmProfile shift coefficient (x1), pinion 0.1Profile shift coefficient (x2), gear 0.0Face width 260 mmGap 80 mmHardness pinion and gear 58 HRCMaterial grade 2Quality level A4Cutter tip radius 2.4 mmCutter depth 14 mmCutter 1/2 pitch addendum 8 mmCutter protuberance 0.25 mm

E.3.3 Rating parameters

The pitting resistance power rating and bendingstrength power rating at unity service factor arecalculated per ANSI/AGMA 2101--C95 equations.With the factors that have a value of one (1.0)deleted, the equations are:

Pazu=ω1 b

1.91× 107ZI

KvKHdw1σHPZN

ZE2


Payu=ω1 dw1

1.91× 107b mt YJKvKH

σFPYN1


where:

ω1 = 8215.4 rpm

b = 260 mm

ZI = 0.1730 (see AGMA 908--B89)

dw1 = 256.91 mm

ZN = 0.6547 (pinion)= 0.6856 (gear)

ZE = 190 [N/mm2]0.5

YJ pinion = 0.4722 (see AGMA 908--B89)

YJ gear = 0.4861 (see AGMA 908--B89)

mt = 6.5551 mm (6/cos 23.75)

YN = 0.7826 (pinion)= 0.8038 (gear)

Kv = 1.13

KH = 1.2369 (see ANSI/AGMA2101--C95)

CSF= KSF = 1.3 (see annex A)

σHP = 1550 N/mm2

(see ANSI/AGMA 2101--C95, table 3Grade 2)



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σFP = 450 N/mm2 (pinion and gear)(see ANSI/AGMA 2101--C95, table 4Grade 2)

Pazu=8215.4 (260)1.91× 107

0.17301.13 (1.2369)

= 26 060 kW (pinion)

×256.91 (1550) (0.6547)190

2

Pazu=8215.4 (260)1.91× 107

0.17301.13 (1.2369)

= 28 580 kW (gear)

×256.91 (1550) (0.6856)190

2

Payu=8215.4 (256.91)1.91× 107

260 (6.5551) (0.4722)1.13 (1.2369)

× 450 (0.7826)1

= 22 410 kW (pinion)

Payu=8215.4 (256.91)1.91× 107

260 (6.5551) (04861)1.13 (1.2369)

× 450 (0.8038)1

= 23 690 kW (gear)

Pa is the lesser of 26 0601.3

28 5801.3

22 4101.3

23 6901.3

or 17 240 kW.

E.3.4 Rating conclusions

The allowable transmitted power, Pa = 17 240 kW, isgreater than the service power of 15 MW.



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Annex F(informative)Efficiency


F.1 Gear unit efficiency

Most contracts for high speed helical gear unitsrequire some guarantee of minimum operationalefficiency. When high power is transmitted, a verysmall increment of efficiency can represent substan-tial economic gain or loss over the life of the gearunit. To realize optimum gear unit efficiency, adetailed study of the several sources of power loss isrequired.

Sources of power loss for high speed helical gearunits include: mesh, internal windage, radial andthrust bearing friction, and shaft driven accessorypower requirements.

F.1.1 Mesh losses

Mesh losses result from oil shearing and frictionallosses which are dependent on the specific slidingvelocity and friction coefficient. Most gear meshesunder this standard will operate in the EHD lubrica-tion regime.

F.1.2 Internal windage losses

Because of the sensitivity of gear and unit specificrelationships (such as housing--to--rotor clearances,pitch line velocity, gear blank proportions anddesign, oil viscosity, method of mesh lubrication andcooling, horizontal or vertical offset, and internalbaffling), this component of gear box losses is verydifficult to accurately estimate without experimentaldata from a specific gear unit.

F.1.3 Bearing losses

Hydrodynamic journal bearing losses are generatedthrough oil shearing. Bearing losses may becalculated by a modified Petroff equation or bycomplex computer modeling methods.

F.1.4 Accessory losses

The power consumed by shaft driven accessoriescan be computed by classic pressure -- displace-ment methods in the case of fuel or lube oil pumps.Accessories other than pumps require appropriateevaluation.

F.2 Calculation methods

F.2.1 Mesh losses

Mesh power loss (PM), for 171/2° or 20°NPA of basicrack, can be estimated as below:

PM= (22− 0.8 αn) 0.01 P z1+ z2z1 z2 (F.1)

where

αn is normal pressure angle of basic rack;

z1 is number of teeth in the pinion;

z2 is number of teeth in gear;

P is transmitted power, kw.

F.2.2 Windage losses

Windage and churning loss can be evaluated by thefollowing equation:

PW=d′2n2b cos3β′mn 1.42 × 10−11

A(F.2)

where

PW is windage power loss per gear, kW;

d′ is operating pitch diameter of gear, mm;

n is gear speed, rpm;

b is total face width, mm;

β′ is operating helix angle;

mn is normal module, mm;

A is arrangement constant (use 1000 to 4000,based on arrangement).

F.2.3 Bearing losses

Hydrodynamic sleeve bearing loss in kW, PBh, canbe estimated by the following equation:

PBh= m n2bd3bL j 1.723× 10−17

c(F.3)

The thrust bearing power loss in kW, PBt, is:

PBt= m n2b

r4o− r4i1.723× 10−17

t(F.4)

where

m is oil viscosity, mPa·s;

nb is bearing speed, rpm;

db is bearing bore, mm;



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t is oil film thickness, mm;

ri is inside radius of thrust bearing, mm;

ro is outside radius of thrust bearing, mm;

L is bearing length, mm;

c is diametral clearance, mm;

j is bearing power loss coefficient (see figureF.3).

The Sommerfield Number used in figure F.3 iscalculated by the following:

S=d2bm nb 10

−6

c2 w 60(F.5)

where

w is load per unit area, kPa.

3

2

1

40 50 60 70 80 90 1000

10

60

70

80

90

20

30

40

50

100

110

Temperature, °C

Absoluteviscosity,m

Pa

s

S

1 ISO Grade 462 ISO Grade 683 ISO Grade 100

Figure F.1 -- Viscosity of petroleum oil

F.2.5 Accessory losses

Oil pump losses may be evaluated based on oil flowfor lubrication and operating pressure:

Pp=Q p

60 000 e(F.6)

where

Q is pump displacement (l/min);

p is pump operating pressure (kPa);

e is pump efficiency (85% estimated).

Temperature, °C

4

5

3

1

2

0

10

20

30

40

50

40 50 60 70 80 90 100

60

Absoluteviscosity,m

Pa

sS

1 Dow Corning XF--258 (Silicone)2 GE Versalube F--30 (Silicone)3 MIL--L0286B (Cellutherm 2505A)4 Mil--7808D5 Mil--L--25336 (Sinclair L--743)

Figure F.2 -- Viscosity of synthetic oil



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0.2

0.25

0.375

0.5

0.75

1.01.251.5

10 5 2 1 0.5 0.1 0.05 0.02 0.01

Plot of j for elliptical bearingsSommerfield Number, S

(A)

(B)

(C)

0.5

0.75

1.0

1.251.5

0.5

0.75

1.0

1.251.5

10 5 2 1 0.5 0.1 0.05 0.02 0.01

10 5 2 1 0.5 0.2 0.1 0.05 0.02 0.01

Sommerfield Number, S

Sommerfield Number, S

Plot of j for cylindrical bearings

Plot of j for four--groove bearings

j = Power Loss Coefficient



0.2

L/db =

L/db =

L/db =

0.25

L/db =

Figure F.3 -- Bearing power loss coefficient, j



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Annex G(informative)

Assembly designations


NOTES:1. Code: L = Left; R = Right2. Arrows indicate line of sight to determine direction of shaft extensions.3. Letters preceding the hyphen refer to number and direction of high speed shaft extensions.4. Letters following the hyphen refer to number and direction of low speed shaft extensions.

Plan views

L--R L--L R--R R--L

LR--LR

LR--R R--LRLR--LL--LR

Plan views

Figure G.1 -- Parallel shaft spur, helical and herringbone gear drives, single or multiple stage



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Annex H(informative)

Purchaser’s data sheet


H.1 Purpose

Data sheets in SI and U.S. customary units areprovided to facilitate communication between pur-

chaser and vendor. The purchaser should fill in theleft side of the data sheet.



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°C

°C

cSt @ 100° CcSt @ 40° C

∆T LUBE

WR2 REFERRED TO LS SHAFT

DATASHEET: ANSI/AGMA6011--I03 JOB NO. ITEMNO.

HIGHSPEEDGEARUNITS END USER

SI UNITS SITE

PURCHASER: PROJECTNAME

REVISION NO. BY DATE

INFORMATION TO BE COMPLETED BY PURCHASER INFORMATIONTOBECOMPLETEDBYVENDOR

1 APPLICABLE TO: PROPOSAL PURCHASE MANUFACTURER

2 REQUISITION NO. MODEL NO.

3 SERVICE QUOTE NO.

4 DRIVERTYPE BASICGEARUNITDATA

5 DRIVEN EQUIPMENT FULL LOAD POWER LOSS kW

6 NO. REQUIRED MEASUREDBY: OTHER

7 SPECIFIED RATING REQUIREMENTS MECHANICAL EFFICIENCY %

8 GEAR SERVICE POWER kW PITCH LINE VELOCITY m/sec

9 GEAR SERVICE FACTOR (MIN) ANTICIPATED SPL dBA@ m

10 RATED SPEED, RPM: N.m.s2

11 INPUT SPECIFIED NOMINAL BREAKAWAY TORQUE N--m@LSShaft

12 OUTPUT SPECIFIED NOMINAL NET MASS (WT) OF GEAR UNIT kg

13 MAX CONTINUOUS SPEED RPM MAX. MAINTENANCE MASS (WT) (IDENTIFY) kg

14 TRIP SPEED RPM TOTAL SHIPPING MASS (WT) kg

15 EXTERNAL LOADS TOTAL SHIPPING DIMENSIONS X X m

16 OTHEROPERATINGCONDITIONS LUBRICATION REQUIREMENTS

17 OIL VISCOSITY:

18 CONFIGURATION REQUIREMENTS UNIT OIL PRESSURE kPa

19 SHAFT ASSEMBLY DESIGNATION (LR or RL) UNIT OIL FLOW (TOTAL) m3/hr

20 HSSHAFTROTFAC’G CPL’G CW CCW MIN. STARTUPOILTEMPERATURE

21 LS SHAFTROTFAC’G CPL’G CW CCW CONSTRUCTION FEATURES

22 HS SHAFT END: CYLN. TAPER 1-- KEY 2--KEYS TYPEOFGEAR REDUCER INCREASER

23 HYDR’LCTAPER INTEGRAL FLANGE SINGLE STAGE DOUBLE STAGE

24 LS SHAFT END: CYLN. TAPER 1-- KEY 2--KEYS SINGLE HELICAL DOUBLE HELICAL

25 HYDR’LCTAPER INTEGRAL FLANGE GEAR TOOTH GEOMETRY

26 INSTALLATIONDATA NUMBER OF TEETH PINION GEAR

27 AMBIENTTEMPERATURE (MIN / MAX): / GEAR RATIO CENTERDIST mm

28 ELEVATION m BAROMETER kPa abs AGMAGEOMETRY FACTOR ”J”:

29 ELECTRICAL AREA CLASS GRP DIV PINION GEAR

30 MAX ALLOW SPL dBA@ m HELIX ANGLE DEGREES FINISH Ra

31 UNUSUALCONDITIONS DUST FUMES NORMAL PRESSURE ANGLE DEGREES

32 INSTRUMENTATION NETFACEWIDTH mm PINION L/d

33 RADIALVIBRATION PROBES NORMAL DIAMETRAL PITCH BACKLASH mm

34 NO. AT EACH RADIAL BEARING TOTALNO. ADD MOD COEF: GEAR PINION

35 AXIALPOSITION PROBES MIN HARDNESS: GEAR PINION

36 LOCATION NO. REQUIRED MANUFACTURING METHODS PINION / GEAR

37 KEYPHASORS TEETHGENERATING /

38 LOCATION NO. REQUIRED TEETH FINISHING /

39 ACCELEROMETER TEETHHARDENING /

40 LOCATION NO. REQUIRED GEAR HUB TO SHAFT INTEGRAL SHRUNK--ON

41 BEARING METALTEMPURATURE SENSORS RIMATTACHMENT

42 TYPE TOTALNO. BEARINGS

43 NO. AT EACH RADIAL BEARING PINION GEAR

44 NO.ATEACHTHRUSTBEARING RADIAL TYPE

45 ADDITIONALREQUIREMENTS UNIT LOADING, kPa

46 JOURNAL VELOCITY, m/s

47 THRUST TYPE

48 UNIT LOADING, kPa

49 MEAN DIA VELOCITY, m/s

50



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ft lb @ LS Shaft

lb ft2

US CUSTOMARY UNITS

°F

°F

cSt @ 210° FSSU @ 100° F

∆T LUBE

WR2 REFERRED TO LS SHAFT

DATASHEET: ANSI/AGMA6011--I03 JOB NO. ITEMNO.

HIGHSPEEDGEARUNITS END USER

SITE

PURCHASER: PROJECTNAME

REVISION NO. BY DATE

INFORMATION TO BE COMPLETED BY PURCHASER INFORMATIONTOBECOMPLETEDBYVENDOR

1 APPLICABLE TO: PROPOSAL PURCHASE MANUFACTURER

2 REQUISITION NO. MODEL NO.

3 SERVICE QUOTE NO.

4 DRIVERTYPE BASICGEARUNITDATA

5 DRIVEN EQUIPMENT FULL LOAD POWER LOSS HP

6 NO. REQUIRED MEASUREDBY: OTHER

7 SPECIFIED RATING REQUIREMENTS MECHANICAL EFFICIENCY %

8 GEAR SERVICE POWER HP PITCH LINE VELOCITY ft/min

9 GEAR SERVICE FACTOR (MIN) ANTICIPATED SPL dBA@ ft

10 RATED SPEED, RPM:

11 INPUT SPECIFIED NOMINAL BREAKAWAY TORQUE

12 OUTPUT SPECIFIED NOMINAL NET MASS (WT) OF GEAR UNIT lb

13 MAX CONTINUOUS SPEED RPM MAX. MAINTENANCE MASS (WT) (IDENTIFY) lb

14 TRIP SPEED RPM TOTAL SHIPPING MASS (WT) lb

15 EXTERNAL LOADS TOTAL SHIPPING DIMENSIONS X X ft

16 OTHEROPERATINGCONDITIONS LUBRICATION REQUIREMENTS

17 OIL VISCOSITY:

18 CONFIGURATION REQUIREMENTS UNIT OIL PRESSURE psi

19 SHAFT ASSEMBLY DESIGNATION (LR or RL) UNIT OIL FLOW (TOTAL) GPM

20 HSSHAFTROTFAC’G CPL’G CW CCW MIN. STARTUPOILTEMPERATURE

21 LS SHAFTROTFAC’G CPL’G CW CCW CONSTRUCTION FEATURES

22 HS SHAFT END: CYLN. TAPER 1-- KEY 2--KEYS TYPEOFGEAR REDUCER INCREASER

23 HYDR’LCTAPER INTEGRAL FLANGE SINGLE STAGE DOUBLE STAGE

24 LS SHAFT END: CYLN. TAPER 1-- KEY 2--KEYS SINGLE HELICAL DOUBLE HELICAL

25 HYDR’LCTAPER INTEGRAL FLANGE GEAR TOOTH GEOMETRY

26 INSTALLATIONDATA NUMBER OF TEETH PINION GEAR

27 AMBIENTTEMPERATURE (MIN / MAX): / GEAR RATIO CENTERDIST in

28 ELEVATION ft BAROMETER ” Hg AGMAGEOMETRY FACTOR ”J”:

29 ELECTRICAL AREA CLASS GRP DIV PINION GEAR

30 MAX ALLOW SPL dBA@ ft HELIX ANGLE DEGREES FINISH

31 UNUSUALCONDITIONS DUST FUMES NORMAL PRESSURE ANGLE DEGREES

32 INSTRUMENTATION NETFACEWIDTH in PINION L/d

33 RADIALVIBRATION PROBES NORMAL DIAMETRAL PITCH BACKLASH mil

34 NO. AT EACH RADIAL BEARING TOTALNO. ADD MOD COEF: GEAR PINION

35 AXIALPOSITION PROBES MIN HARDNESS: GEAR PINION

36 LOCATION NO. REQUIRED MANUFACTURING METHODS PINION / GEAR

37 KEYPHASORS TEETHGENERATING /

38 LOCATION NO. REQUIRED TEETH FINISHING /

39 ACCELEROMETER TEETHHARDENING /

40 LOCATION NO. REQUIRED GEAR HUB TO SHAFT INTEGRAL SHRUNK--ON

41 BEARING METALTEMPURATURE SENSORS RIMATTACHMENT

42 TYPE TOTALNO. BEARINGS

43 NO. AT EACH RADIAL BEARING PINION GEAR

44 NO.ATEACHTHRUSTBEARING RADIAL TYPE

45 ADDITIONALREQUIREMENTS UNIT LOADING, psi

46 JOURNAL VELOCITY, ft/sec

47 THRUST TYPE

48 UNIT LOADING, psi

49 MEAN DIA VELOCITY, ft/sec

50

Ra



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Bibliography

The following documents are either referenced in the text of ANSI/AGMA 6011--I03, Specification for HighSpeed Helical Gear Units, or indicated for additional information.

AGMA 908--B89, Geometry Factors for Determin-ing the Pitting Resistance and Bending Strength ofSpur, Helical and Herringbone Gear Teeth

AGMA 925--A03, Effect of Lubrication on GearSurface Distress

AGMA 927--A01, Load Distribution Factors -- Ana-lytical Methods for Cylindrical Gears

Ehrich, Fredric F., Handbook of Rotordynamics,McGraw--Hill, Inc., 1992

ISO 4406:1999 (SAE J1165), Hydraulic fluid power-- Fluids -- Method for coding the level ofcontamination by solid particles

SAE/AMS 2300, Steel Cleanliness, Premium Air-craft--Quality Magnetic Particle Inspection Proce-dure

SAE/AMS 2301, Steel Cleanliness, Aircraft QualityMagnetic Particle Inspection Procedure



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PUBLISHED BYAMERICAN GEAR MANUFACTURERS ASSOCIATION1500 KING STREET, ALEXANDRIA, VIRGINIA 22314



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Documents

AMERICAN NATIONAL STANDARD - DrGearbox.comdownload.drgearbox.com/standards/AGMA6011-I03_Specification for... · ANSI/AGMA 2101--C95, Fundamental Rating Fac-tors and Calculation Methods