Agma Standard: Specification For Speed Gear Units [PDF]

Citation preview

-



ANSIIAGMA 6011 H98 (Revisionof ANSIIAGMA 6011 -G92)



AMERICAN NATIONAL STANDARD



Specification for High Speed Helical Gear Units



AGMA STANDARD COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



S T D - A G H A b O L L - H S B - E N G L L778



American National



Standard



Ob87575 0 0 0 5 8 3 5 7 7 7



Specification for High Speed Helical Gear Units ANSI/AGMA 6011-H98 [Revision of ANWAGMA 6011-G921 Approval of an American National Standardrequires verification byANSI that the requirements for due process, consensus, and other criteria for approval have been met by the standards developer. Consensusis established when,in the judgment of the ANSI Board of Standards Review, substantial agreement has been reachedby directly and materially affected interests. Substantial agreement means much more than a simple majority, necessarily but not unaal views and objections be considered, and that a nimity. Consensus requires that concerted effort be made toward their resolution. The useof American National Standards is completely voluntary; their existence does not in any respect preclude anyone, whether he has approved the standards or not, from manufacturing, marketing, purchasing or using products, processes or procedures not conforming to the standards. The American National Standards Institute does not develop standards and will in no circumstances give an interpretation of any American National Standard. Moreover, no person shall have the right or authority to an issue interpretation of an American National Standard in the name ofthe American National Standards I n s t i e . Requests for interpretation of thisstandardshouldbeaddressedtotheAmericanGearManufacturers Association.



CAUTION NOTICE: AGMA technical publications are subject to constant improvement, revision or withdrawal as dictated by experience. Any person who refers to any AGMA Technical Publication should be sure that the publication is the latest available from the Association on the subject matter. Fables or other self-supporting sections may be quoted or extracted. Credit lines should read: Extracted from ANSllAGMA 6011-H98, Spikation tbr High Speed Helical Gear Units, with the permission of the publisher, the American Gear Manufacturers Association, 1500 King Street, Suite201, Alexandria, Virginia 22314.1 Approved June 19, 1998



ABSTRACT This standard includes design, lubrication, bearings, testing and rating for single and double helical external tooth, parallel shaft speed reducers or increasers. Units covered include those operating with at least one stage having a pitch line velocity equal to orthan greater 35 meters per second rotational or speeds greater than 4000 rpm and other stages having pitchline velocities equal to or greater than8 meters per second. Published by American Gear Manufacturers Association 1500 King Street, Suite 201, Alexandria, Virginia 22314



Copyright O 1997 by American Gear Manufacturers Association All rights resewed.



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



Printed in the United States of America ISBN: 1-55589-693-6



II COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



ANWAGMA 6011-H98



AMERICAN NATIONAL STANDARD



Contents Page



Foreword ............................................................... 1 Scope .............................................................. 2 3 4 5



6 7



8



Symbols,terminologyanddefinitions .................................... Designconsiderations ................................................ Ratingofgears ...................................................... Lubrication ..........................................................



iv 1



1 3



6 9



.................................................11



Vibrationandsound Functionaltesting



................................................... Vendorandpurchaserdataexchange .................................



14 15



Tables Symbolsused in equations ............................................ 2 2 Maximum length-to-diameter (Ud) ratios for unmodified leads . . . . . . . . . . . . . . 3 3 Hydrodynamicbabbittbearingdesignlimits .............................. 5 4 Recommendedoils .................................................. 10 5 Casingvibrationlevels ............................................... 14 1



Figures



..................................................



13



Service factors ......................................................



19



B A simplified method for verifying scuffing resistance ...................... C Lateralrotordynamics ............................................... D Systems considerations for high speed gear drives ...................... E Illustrativeexample .................................................. F Efficiency ..........................................................



23



1



Amplificationfactor



Annexes A



25



31 41 43



iii COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



S T D - A G N A bOLL-HSB-ENGL



L998 D O b 8 7 5 7 5 0005837 5 b l



ANSllAGMA 6011-H98



STANDARD NATIONAL AMERICAN



Foreword r h e foreword,footnotesandannexes,ifany,inthisdocumentareprovidedfor informational purposes only and are not to be construed of ANWAGMA as a part Standard 6011-H98, Specificationfor High Speed Helical Gear Units.] The first high speed gear unit standard, AGMA421.O1,was adopted as a tentative standard in October, 1843. It contained formulas for computing the durability horsepower rating of a short table of application factors. AGMA gearing, allowable shaft stresses, and included 421.O1 was revised and adoptedafull as status standardin September, 1947 and issued as AGMA 421.02. The High Speed Gear Committee began work on the revision AGMA of 421 .O2 in 1951, which included: classification of applications not previously listed; changing the application factors from "K" values to equivalentService Factors; revision of the rating formula to allow for the use of heat treated gearing; and develop a uniform selection method for high speed gear units. This Uniform Selection Method Data Sheet became AGMA 421.03A. AGMA 421.O3 was approved as a revision by the AGMA membership in October,1954.



The standard was reprinted as AGMA 421 .O4 in June, 1957. It included the correction of typographical errors and the addition of a paragraph on pinion proportions and bearing span, which had been approved by the committee for addition to the standard at the October, 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 hardness AGMA materials and sound level limits. The revisions were incorporated in 421.O5 which was approved by the AGMA membership as of October22,1963. The significant changes of 421.06 from 421.05 were: minimum pitch line speed was increased to 5000 feet per minute(25 meters per second); strength and durability ratings were changed; and some service factors were added. AGMA 421.O6 was approved by the High Speed Gear Committee as of June 27, 1968, and by the AGMA membership as of November 26,1968. ANWAGMA 6011-G92 was a revision of421 .O6 approved by theAGMA membership in October, 1991. The most significantchangesweretheadaptationofratingsper ANWAGMA 2001-B88 and the addition of normal design limits for babbitted bearings. ANWAGMA 6011 -G92 used "application factor" and not 'service factor".



AN WAGMA 6011-H98 is a further refinement ANWAGMA of 6011-G92. One ofthe most significant changesis theconversion to an all metric standard. The rating methods are now per ANWAGMA 2101 -C95 which is the metric version of ANSI/AGMA 2001 4 9 5 . To provide uniform rating practices, clearly defined rating factors are inincluded this standard (ANSI/AGMA 6011-H98). While some equations may slightly change to conform to metric practices, no substantial change has been made to the rating practice for durabilrty and m/s to 35 strength rating. In addition, minimum pitch line velocity has been raised25from m/s and minimum rotationalspeed increased to4000 rpm. AGMA has revertedto the term "servicefactor" in their standards, which is reflected in this revision. The servicefactorapproach is more descriptive of enclosedgeardrive applications andcan be defined as the combined effects of overload, reliability, desired life, and other application related factors. The service factor is applied only to the gear tooth of all components. rating, rather than to the ratings In continued recognitionof the effects of scuffing in the rating of the gear sets, additional information on scuffing resistance has been added to annex B of this revision.



COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



S T D * A G f l A bOLL-HSB-ENGL NATIONAL AMERICAN



L998



Ob87575 0 0 0 5 8 3 8 4 T 8



STANDARD



ANWAGMA 6011-H98



AGMA 427.01 has been withdrawn. The information found in AGMA included in annexD of this standard.



427.01 has been



Realistic evaluation of the various rating factors of ANWAGMA 6011-H98 requires specific knowledge and judgment which come from years of accumulated experience in designing, manufacturing and operating high speed gear units. This input has been provided by the AGMA High Speed Gear Committee. ANWAGMA 6011-H98 was approved as a revision by the AGMA membership in June, 1997. It was approved as an American National Standard on June19,1998. Suggestions for improvement of this standard will be welcome. They should be sent to the AmericanGearManufacturersAssociation,1500KingStreet,Suite 201, Alexandria, Virginia 22314.



PERSONNEL of the AGMA Committee for High Speed Helical Gear Units Chairman:L.Lloyd ............................. Vice Chairman: M.W. Neesley ....................



Lufkin Industries, Inc. Philadelphia Gear Corporation



ACTIVE MEMBERS



................................. ................................... M.J. Hardiman ................................. W.P. Pinichil ................................... F.L. Vanlaningham .............................. J.B. Amendola



W.P. Crosher



MAAG Gear Company, Ltd. Flender Corporation Philadelphia Gear Corporation Philadelphia Gear Corporation Rotating & Turbo Machine



ASSOCIATE MEMBERS F. Barresi ...................................... K.O.Beckman ................................. A.S. Cohen .................................... E. Dehner ..................................... N. Hulse ....................................... C.E.Long ..................................... D.L. Mairet ..................................... D.R.McVttie ................................... W. Naegeli ..................................... J.R. Partridge .................................. J. Simpson .................................... M.P. Starr. ..................................... FA Thoma .................................... F.C. Uherek ....................................



Flender-Graffenstaden Lufkin Industries, Inc. Engranes y Maquinaria Arco BHS-Voith Getriebetechnik Cotta Transmission Company Cummins Engine The Falk Corporation Gear Engineers, Inc. MAAG Gear Company, Ltd. euro Lufkin bv Turner Uni-Drive Company The Falk Corporation F.A. Thoma, Inc. Flender Corporation



V



COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



S T D . A G M A bOLL-HSB-ENGL



Ob87575 0 0 0 5 8 3 7 334



L778



ANWAGMA 6011-H98



STANDARD NATIONAL AMERICAN



American National Standard -



Specification for High Speed Helical Gear Units 1 Scope Thishighspeedhelicalgearunitstandardis provided as a basis for improved communication regarding:



-



establishment of uniform criteria for rating;



-



guidance for design considerations; and,



-



identificationoftheuniquefeatures speed gear units.



ofhigh



1.1 Application



Operationalcharacteristicssuchaslubrication, maintenance,vibrationlimitsandtestingarediscussed. This standardis applicableto enclosed high speed, external toothed, single and double helical gearunitsoftheparallelaxistype.Unitsinthis classification are:



- singlestageunitswithpitchlinevelocities equal to or greaterthan 35 meters per second or rotational speeds greater than 4000 rpm;



1.2 Normative references



The following standards contain provisions which, through reference in this text, constitute provisions of thisAmericanNationalStandard.At thetime of publication,theeditionsindicatedwerevalid. All standardsaresubjecttorevision,andparties to agreements based on this American National Standard are encouragedto investigatethe possibilrty of applying the most recent editions of the standards indicated below. AGMA 908-B89, Information Sheet - Geometry of Spur, Factors for DeterminingtheStrength Helical, Herringbone and BevelGear Teeth



ANWAGMA 101O-E95,Appearance of Gear Teeth - Terminology of Wear and Failure ANWAGMA2000-A88, GearClassificationand Inspection Handbook - Tolerances And Measuring Methods For Unassembled Spur And Helical Gears (Including Metric Equivalents) ANWAGMA 21O1 -C95,Fundamental RatingFacfor Involute Spurand tors and Calculation Methods Helical Gear Teeth ANWAGMA 6000-896, SpecificationforMeasurement of Linear Vibration onGear Units ANWAGMA 6001-D97, Design and Selection of Components for EnclosedGear Drives ANWAGMA 6025-C90, Sound for Enclosed Helical, Herringbone, and Spiral Bevel Gear Drives



-



ANWAGMA 9005-D94, Industrial Gear multi-stage units withat least one stage havLubrication 35 ing a pitch line velocity equal or greater to than meters per second and other stages having pitch line velocities equalto or greater than 8 meters per second. 2 Symbols, terminology and definitions The gearing and unit components are to be manufacturedwithintheconventionalprocesses and 2.1 Symbols accuracy limits normally used by manufacturers of high speed gear units. The symbols used in this standard are shownin table Limitsspecifiedaregenerallyaccepteddesign limits. When specific experience exists in excess of these limits, this experiencemay be applied.



1. NOTE: The symbols and terms contained in docuthis ment may vary fromthose used in otherAGMA standards. Users of this standard should assure



Marine propulsion, epicyclic, aircraft and automotive themselves that they are using these symbols and terms in the manner indicated herein. gearing are not covered by this standard.



Previous page is b h ~ ~ t ~ . COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



1



STD.AGMA bOLL-HS8-ENGL ANSIIAGMA W11 -H98



1778



O b 8 7 5 7 5 00058LtO 0 5 b



STANDARD NATIONAL AMERICAN



ANSI Y10.3-1968,Letter Symbols for Quantities Used in Mechanics of Solids The terms used, wherever applicable, conform to the following standards: ANSI/AGMA 1O12-F90, Gear Nomenclature, AGMA 904-C96, Metric Usage Definitions of Terms with Symbols 2.2 Nomenclature



-



Table 1 Symbols used in equations



Term Allowable double amplitudeof unfiltered vibration Amplitude atNCt Amplification factor Service factor for pitting resistance Critical response envelope Diametral clearance Nominal bearing bore diameter Pinion operating pitch diameter incremental dynamic load Transmitted tangential load Rim thickness factor Load distribution factor Mesh alignment correction factor Mesh alignment factor Lead correction factor Pinion proportion modifier Service factorfor bending strength Size factor Dynamic factor Net face width plus gap Number of stress cycles Rotor first critical, center frequency Maximum rotorspeed Initial (lesser) speed at0.707 x peak amplitude (critical) Final (greater) speed at 0.707 x peak amplitude (critical) Allowable transmitted power for the gearset Allowable transmitted power for bending strength at unity service factor Allowable transmitted power for pitting resistance at unlty service factor Power loss Service powerof enclosed drive Lubricant flow Separation margin Amount of residual rotor unbalance Journal static loading Half weightof coupling and spacer Total weight of rotor Stress cycle factor for bending strength Temperature factor



Reference paragraph



6.5



6.3.3.3 6.3.3.3 4.2.2



6.3.3.3 3.5.1 3.5.1 3.2 4.2.1 4.2.1



4.3 4.2 4.2 4.2 4.2 4.2 4.2.2 4.2 4.2.1



3.2 4.2



6.3.3.3 6.3.3.3 6.3.3.3 6.3.3.3 3.9 3.9 3.9



7.3.5 3.10



7.3.5 6.3.3.3 6.4 6.4 6.3.3.2



6.3.3.2 4.3 4.2 (conb



2 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



STD*AGMA bO11-HSB-ENGL 1 9 9 8 W Ob87575 0005841 T 7 2



STANDARD MA NATIONAL AMERICAN



6011-H98



Table 1 (concluded)



I Symbol



ZN ZR



ZW



AT aFP aHP



o



Term Stress cycle factor for pitting resistance Surface condition factor for pitting resistance Hardness ratio factor for pitting resistance Change in lubricant temperature Allowable bending stress number Allowable contact stress number Speed of rotor



3 Design considerations



Table 2



Any practical combinationof tooth height, pressure angle and helix angle may be used. However, it is I recommended that helical gears have a minimum working depth of 1.800 times the normal module, a maximum normal pressure angle25 ofdegrees, and 8 and45degrees.The ahelixanglebetween relationship of center distance to face width may vary of providedthatthereisnoundueconcentration stress arising from deflection under load.



Table 2 presents the maximum length-to-diameter (Lid) ratios for several combinations of materialsin current use. The Lldvalues shown in table 2 apply to helical gears with unmodified leads when designed Lld to transmit the rated power. Generally higher ratiosarepermittedwithleadsthataremodified when analytical load distribution methods are employed that yield load distribution values, I&, that are lower or equal to the default values in ANSI/ AGMA 2101-C95. A detailedanalyticalmethod shouldinclude,butnotbelimitedto,bending, torsional and heat distortion. When a higherLld ratio than tabulated in table 2 is proposed, the gear vendor shall submit justification intheproposalforusingthehigher Lld ratio. Lld ratios Purchasersshouldbenotifiedwhen exceed thosein table 2. When operating conditions other than the gear rated power are specified by the purchaser, such as the normal transmitted power, the gear vendor shall consider in the analysis the at which the gear unit length of time and load range will operateat each conditionso that the correct lead modificationcanbedetermined.Whenmodified leads are to be furnished, purchaser and vendor



"



"



"



"C N/mm2 N/mm2 rPm



Reference paragraph 4.2 4.2 4.2 7.3.5 4.4 4.4 6.3.3.2



shall agree on the tooth contact patterns obtained in the checking stand, housing or test stand.



3.1 Tooth proportions and geometry



3.2 Pinion proportions



Units



- Maximum length-to-diameter (Lld) ratios for unmodified leads



Minimum hardness



I Maximum Lldratio I Double helical



Single helical 302 H6 341HB 1.5 321 HB 363 HB 1.45 341 HB 1.45 363 H6 341 HB 50 HRC 2.0 1.45 50 HRC 50 HRC 1.4 55 HRC 1.7 55 HRC 1.35 58 HRC 58 HRC 1.3 NOTE: t = net face width plus gap,mm; d = pinion operGear



Pinion



ating pitchdiameter, mm; HB =Brinell hardnessnumber; HRC = Rockwell hardness (C scale)



3.3 Rotor construction



Several configurations may be applied in the construction of rotors. The most commonly used are listed below: a) Integralshaftandgearelement.Thisconfiguration is commonly used for pinions, smaller gears,orrotatingelementsoperatingabovea pitch line velocrtyof 150 meters per second. The its shaft, is machined pinion or gear, integral with from a single blank; b) Solid blank shrunk on a shaft. The shrink fit may be used either with or without a mechanical torque transmitting device (such asor key spline). When no torque transmitting device is used, the shrink fit must provide ample capacity to transmit torque when considering centrifugal and thermal effects.Whenatorquetransmittingdevice is used, the shrink fit must provide ample location support when considering centrifugal and thermal effects;



3 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



ANSI/AGMA 6011-H98



AMERICAN NATIONAL STANDARD



- environmentalelementsthatwillattackthe c) Fabricatedgear. A forgedrim is welded directly to the fabricated substructure producing a unit housing, rotating components, bearings or lubricant; one-piece welded gear. The shaft may be a part of the weldment. Fabricated gears should be ana- - inadequate support for the housing; lyzedto consider centrifugal and thermal stresses - high pitch line velocities which may affect oil and fatigue life. Maximum pitch line velocrty for distribution, create excessive heat rise, or cause 130 metersper weldedgearconstructionis other adverse conditions. second; 3.49 Shaft seals d) Forged rim shrunk onto a substructure. The Gear housings shall be equipped with replaceable substructure may be forged, cast, or fabricated. labyrinth-typeendsealsanddeflectors where a part of thesubstructure. Theshaftmaybe shaftspassthroughthehousing.Thesealsand Shrunk rims shall consider stresses and torque transmitting capacity due to fit, centrifugal, and deflectors shall be madeof nonsparking materials. thermal effects (refer to item b). The normal de- The design of the seals and deflectors shall effecsign limit for this type of construction is60 meters tively retain oil in the housing and prevent entry of per second. foreignmaterialintothehousing.Lip-typeseals have a very finite life and can generate enough heat Combinations of theaboveareoftenusedon at higher speeds to be a fire hazard. They may be multi-stage units. used only with the purchaser’s approval and surface Stresses and deflections at high speeds often dictate velocrty should be kept within the seal manufacturers limtts for a specific type of construction. A careful conservative recommendation. analysis of actual operating stresses and deflections 3.5 Bearings should be made to ensure reliable operation. Radial bearings are normally of the hydrodynamic 3.4 Gear housing sleeve or pad type. Thrust bearings are usually flat The gear housing should be designed to provide a land,taperedland,orthrustpadtype.Rolling sufficiently rigid enclosed structure for the rotating elementbearingsareoccasionallyusedwhen elements that enables them to transmit the loads speeds are at the very low end of the high speed imposed bythe system and protects them from the range.Parametersforbearingdesignconsider environment in which they will operate. The normal service power. Proper design of bearingsis manufacturer’s designof the housing must provide critical to the operation of a high speed enclosed for proper alignmentof the gearing when operating drive gear unit. under the user’s specified thermal conditions, and 3.5.1 Hydrodynamic radial bearings its to the torsional, radial and thrust loadings applied Hydrodynamicradialbearingsshallbelinedwith shaft extensions. In addition,it should be designed suitable bearingmaterial.Tinandleadbased to facilitate proper lubricant drainage for the gear babbitts (white metal) are among the most widely mesh and bearings. used bearing materials. Tin alloy is usually preferred The user’s designof the supporting structure must overleadalloysbecauseof its highercorrosion of the gearing. maintain proper and stable alignment resistance, easier bonding, and better high temperaThe alignment must consider a l l specified torsional, ture characteristics. Hydrodynamic radial bearings radial and thrust loadings, and thermal conditions shall have a rigid steel or other suitable metallic present during operation. backing and be properly installed and secured the in housing against axial and rotational movement. 3.4.1 Special houslng considerations ease of Bearings are generally supplied split for Certain applications may be subjected to operating assembly. Selection of the particular design sleeve conditions requiring special consideration. Some of or pad type bearing shall be based on evaluation of these operating conditions are: journal veloctty, surface loading, hydrodynamic film oil thickness, calculated bearing temperature, - temperaturevariations in thevicinityofthe viscosity, flow rate, bearing and stabil@. gear unit; - relativethermalgrowthbetweenmating system components; 4 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



Hydrodynamicradialbearingsshall bedesigned such that damaging self generated instabilities (e.g.,



S T D - A L M A bO11-HSB-ENGL



1 7 7 8 D Ob87575 00058L13 8 b 5 H



AMERICAN NAflONAL STANDARD



ANSIIAGMA 6011-H98



half frequency whirl) do not occur at any anticipated provided on the low speed shaft for all double helical operational loador speed. Hydrodynamic instabillty gears and on single helical gears fitted with a thrust type collars (see 3.5.5). Thrust bearings shall be occurswhenajournaldoesnotreturntoits established equilibrium position after being momen- provided on each shaft for all single helical gears not tarily displaced. Displacement introduces an instafitted with thrust type collars. bility in which the journal whirls around the bearing axis at less than one half journal speed. Known as When gear units are supplied without thrust bearbe "half frequency whirl" this instabillty occurs in lightly ings,some type of endfloatlimitationshall provided at shaft couplings to maintain positive axial loaded high speed bearings. positioning of the gear rotors and connected rotors. Heat is generated at running speeds as a result of oil Provisions to prevent contact of the rotating eleshear. Temperatureis regulated by controlling the oil ments with the gear casing shall be provided unless of the flow through the bearing and external cooling specifically agreed to by the purchaser. oil. The anticipatedpeakbabbitttemperatureas relatedto bearing oil discharge temperatures shouldThe design of a hydrodynamic bearing to sustain thrust is as complicated as the design of a radial be kept within a range that is compatible with the hydrodynamic bearing. Complete analysis requires 3 bearing material and oil characteristics. See table consideration of heat generation, oil flow, bearing for design limits. material, load capaclty, speed and stiffness. Thrust 3.5.2 Rolling element radial bearings bearing load capacity should consider the possibility of torque lock-up loads from couplings. When other Selection of rolling element radial bearings should external thrust forces are anticipated, the gear drive be based upon the application requirements and the manufacturer must be notifiedof their magnitudes. bearing manufacturer's recommendations and ratSee table3 for design limits. ing methods. For normal applications an L10 life of 50 O00 hours minimum is required.



3.5.4 Rolling element thrust bearing



3.5.3 Thrust bearings



Selection of rolling element thrust bearings shall be Thrust bearings shall be furnished with all gear units based upon the application requirements and the bearing manufacturer's recommendations and unless otherwise specified. Unless otherwise approved by the purchaser, thrust bearings shall be rating methods.



-



Table 3 Hydrodynamic babbitt bearing designlimits1)



N/mm2



Minimum oil film thickness mm



Bearing shell temperature21 3) "C



3.8



0.020



1004)



4.2



0.020



115 115



1.7 0.5 3.5



0.020



115 115



125 50



Projected unit Type of bearing



Radial bearing - Fixedgeometry - Tiltingpad Thrust bearings - Taperedland - Flat face - Tilt pad NOTES:



load



NIA L



0.015



L



115



Maximum velocity m/s



125



L



125



Limits are forbabbitt on steel backing. When other materials are used, established limits for these materials are permissible. Bearing clearances shouldbe chosen toyield proper temperature, high stiffness andstability, as well as to ensure adequate clearanceto cope with thermal gradients, whether steady, static or transient. The average ratio of diametral clearance (C), to the nominal bore size (o), UD,for radial bearings is approximately 0.002 mm/mm. 2, Bearing temperature measurements are taken in the backing materialwithin 3 mm of the backing materialbabbitt interface atthe hottest operational zoneof the bearing circumference. 3, Higher values are acceptableif supported with testing and field experience. 4) For fixed geometry radial bearings operating above 100 Ws, up to 125 m/s, special engineering is required. NOTE: Above limits will generally not occur all together. One parameter alone may dictate the design.



5 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



S T D * A G M A bOLL-HSB-ENGL ANSI/AGMA 6011-H98



must be calculatedusingthemethodshownin clause 4 (see also ANWAGMA 2101-C95, clause Thrust collars may be used to counteract the axial 1O). The lowest value obtained shall be used as the gear thrust developed by single helical gear sets. allowable transmitted power of the gear set. A l l other a power components shall be capable of transmitting Thrustcollarsarrangedateachendofasingle equal toor greater than the gearset power rating. helical pinion and having bearing surface contact diameters greater than that of the pinion outside P,, set, The allowable transmitted power for the gear diameter may be used to carry the gear mesh thrust is determined: forces. The general arrangement of thrust collars Payu have a conical shape where they contact a similarly Pa = the lesser of -and 'SF KSF shaped surface on the mating gear rim located below where: the root diameter of the gear. Other designs also exist. Single helical gear sets using thrust collars P- is the allowable transmitted power for pitting may be positioned in the housing in a similar fashion resistance at untty service factor (CSF= 1.O); to that of double helical gear elements. Puyuistheallowabletransmittedpowerfor bendingstrengthatunrtyservicefactor 3.6Threaded fasteners (&F = 1-0); Refer to ANWAGMA 6001-D97,Design and SelecCSF is the service factor for pitting resistance; tion of Components for Enclosed Gear Drives, recommended values are shown in annex clause 8. A; 3.5.5 Thrust type collars



3.7 Shafting



&F



is the service factor for bending strength; recommended values are shown in annex A.



Thepinionandgearshaftsmaynormallybe designed for the maximum bending and maximum 3.10 Senrice power, PS torsional shear stresses at unit rating (see 3.10)by is defined as the the appropriate methods and allowable values from The service power of an application maximum installed continuous power capacrty of the ANSI/AGMA 6001-D97, clause 4. In some inprime mover, unless specifically agreed to by the stances this may result in an oversized or undersized customer and gear manufacturer. For example, for shaft, and therefore an in-depth study using other electric motors,the maximum continuous power will available analysis methods may be required. be the motor nameplate power rating multiplied by 3.8 Torsional andlateral vibrations the motor service factor. When an elastic system is subjected to externally appliedcyclicorharmonicforces,theperiodic is called forced motion that results from such forces vibration. Such systems are evaluated in two ways: torsionally,foranalyzingtheeffectsoftorsional vibrations;andlaterally,fordeterminingthe influenceoflateralvibrations. Incertaincasesaxial vibrations must be considered. of geardriven Because of thewidevariation systems, clause 6 of this standard outlines areas where proper assessment of the system may be necessary. ln addition,appropriateresponsibility between the gear manufacturer and customer must be clearly delineated.



For gear units between two items of driven equipment,theservicepowerofsuchgearsshould normally be not less thanitem (a)or(b)below, whichever is greater. a. 110 percent of the maximum power required by the equipment driven by the gear; b. The maximum power of the driver prorated between the driven equipment, based on normal power demands. If maximum torque occurs at a speed other than the maximumcontinuousspeed,thistorqueand its corresponding speed shall be specified by the purchaser.



The service power shall be less than, or equal to, the allowable transmitted gearset power rating, or PS S p a ...(2) where: The pitting resistance power rating and the bending strength power rating for each gear mesh in the unit PS istheservicepower, kW.



3.9 Allowable transmittedpower for the gear set, P,



6 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



ANWAGMA 6011-H98



ANlERICAN NATIONAL STANDARD



Gm



= Usevalues from curve 3,precision enclosed gear units, see figure 7 and table 2 of ANWAGMA 21O1 -C95;



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



= 0.8; Gpm= 1.o; &qmc



4 Rating of gears



&qe



4.1 Ratingcuteria



= 0.8.



The calculated value of & shall not be less 1than .l.



Thepittingresistancepowerratingandbending NOTE: The above empirical rating method assumes strength power rating for each mesh in the unit mustproperly matched leads whether unmodified ormodibe calculated and the lowest value obtained shall be fied, teeth centralto the bearing span and tooth contact checked at assemblywith contact adjustments as reused as the power rating of the gearset. quired. If these conditions arenot met, or for wide face 4.2 Pitting resistance power rating



The pitting resistance of gear teeth is consideredto be a Hertzian contact fatigue phenomenon. Initial pittinganddestructivepittingareillustratedand discussed inANWAGMA 1010-E95.



gears, it maybe desirableto use an analytical approach to determine load distribution factor. 4.2.1 Dynamic factor, Iyy



Dynamic factors account for internally generated gear tooth dynamic loads which are caused by gear toothmeshingactionatanon-uniformrelative The purpose of the pitting resistance formula is to angular velocity. determine a load rating at which destructive pitting of Thedynamicfactoristheratiooftransmitted the teeth does not occur during their design life. The tangential tooth load to the total tooth load which ratingsforpittingresistancearebasedonthe includes the dynamic effects. formulas developed by Her& for contact pressure between two curved surfaces, modified for the effect ..(3) of load sharing between adjacent teeth. where: ZN is the stress cycle factor as calculated by the Fd is the incremental dynamic tooth load due to lower curve of figure 17 of ANWAGMA 2101-C95 the dynamic response of the gear pair to the and should be based on 40 O00 hours of service at N; transmission error excitation, 40 O00 hours is rated operating speed. If other than Fr is the transmittedtangentialload, N. used for rating,it must be with the specific approval of the customer and must be so stated along with the Dynamic forces on the gear teeth result from the rating. ZN should be greater thanor equal to 0.68. gear transmissionerrorwhichisdefinedasthe departure from uniform relative angular motion of a The pitting resistance power rating shall be per the pair of meshing gears. The transmission error is ratingproceduresandformulasof ANWAGMA caused by: 2101 -C95, clause 10, when using service factors, with the following values: is the hardness ratio factor,Zw = 1.O; is the temperature factor,Ye = 1.O; is the stress cycle factor, ZN=2.466N-0*056, where is the number of stress cycles;



& = 1.O; is the size factor, is the surface condition factor, ZR= 1.O; is thedynamic factor (see 4.2.1); is theload distribution factor. Values are to be perANWAGMA 21O1 495. The following values shall be used with the empirical method:



- inherentvariationsingearaccuracyas manufactured; and, - geartoothdeflectionswhicharedependent on the variable mesh stiffness and the transmitted load. The dynamic response to transmission error excitation is influenced by:



-



themasses ofthegearsandconnected rotors; - shaft and coupling stiffnesses; and, - the damping characteristics of the rotor and bearing system. Forgearsetsoperatingat or nearratedload, dynamic factors between 1.O9 and 1.15 may be



7 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



AMERICAN NATIONAL STANDARD



ANSIIAGMA 6011-H98



used; however,it is suggestedthat& = 1.15, unless experience dictates otherwise. TheAGMAQualityNumbersperANWAGMA 2000-A88, specifically tooth element tolerances for pitch and profile, and the pitch line veloctty may be used as parameters to guide the selection of the dynamic facfors. Within the1.O9 to 1.15dynamic factor range, the trend is for K,,to vary in nearly a direct relationshipwith AGMA Quallty Numbers from to Q15.



ell



Thedynamicfactor, &, doesnotaccountfor dynamictoothloadswhichmayoccurdue to torsionalorlateralnaturalfrequencies.System designs should avoid having such natural frequenwith cies close to an excitation frequency associated an operating speed since the resulting gear tooth dynamic loads may be very high. Refer toANWAGMA2101-C95 foradditional considerations influencing dynamic factors.



Occasionally, manufacturing tool marks, wear, surfacefatigue,orplasticflowmaylimitbending strength due to stress concentration around large, sharpcorneredpitsorwearstepsonthetooth



surface. The bending strength power rating for gearing within the scope of this standard shall be determined by the of ANSI/AGMA ratingmethodsandprocedures 2101-C95, clause 10, when using service factors, with the following values: Iyy is the dynamic factor (see 4.2.1); & is the load distribution factor (see 4.2); Ye is the temperature factor, Ye = 1.0; & is the size factor, & = 1.0; & is the rim thickness factor, & = 1.0; YN is the stress cycle factor, YN = 1.6831 N-0.03u, where N is the number of stresscycles. 4.4 Allowable stress numbers,0q.p and u ~ p



Allowablestressnumberswhicharedependent upon material and processing are given in ANSI/ The service factor includesthe combined effects of AGMA21 O1-C95 clause16.Thatclausealso overload, reliabiltty, desired life, and other applicaspecifiesthetreatmentofmomentaryoverload tion related influences. The AGMA service factor conditions. used in thisstandarddepends onexperience Three grades of material have been established. acquired in each specific application. Grade 1 is the normal commercial qualtty steel and In determining theservicefactor,consideration shall not be used for gears rated by this standard. should be given to thefact that systems develop a Grade 2 is a high qualrty steel meeting SAE/AMS peak torque, whether fromthe prime mover, driven 2301 cleanliness requirements. Grade 3 isa machinery, or transitional system vibrations, that is S W A M S 2300. Both premium qualtty steel meeting greater than the nominal torque. Grade 2 and Grade 3are heattreated under carefully of material, controlled conditions. The choice When an acceptable service factor is not known from l e f t to the gear designer; hardness and grade is experience the values shown in annex A shall be however, values of CTHPand shall be for grade 2 used as minimum allowable values. materials. 4.3 Bending strength power rating Dueconsiderationshould be giventoultrasonic The bending strength of gear teethis a measureof testing and/or magnetic particle inspection of high the resistance to fatigue cracking at the tooth root speed gear rotors which are subject to high fatigue fillet. cycles during operation. The intentof the AGMA strength rating formula is to For details on tooth failure, refer to ANWAGMA 1O10-E9!5. can be transmitted for the determine the load which design l i f e of the gear drive without causing root fillet 4.5 Reverse loading cracking or failure. For idler gears and other gears where the teeth are YN is the stress cycle factor as calculated by the completely reverse loaded on every cycle, use 70 lower curve of figure 18 of ANSVAGMA 2101-C95 percent of the allowable bending stress number, UFP, and should be based on40 O00 hours of service at in ANSI/AGMA21O1 -C95. rated operating speed.If other than40 O00 hours is resistance 4.6 Scuffing used for rating,it must be with the specific approval of the customer and must be so stated along with the Scuffing failure (sometimes referred to as scoring) rating. YN should be greaterthan or equalto 0.80. concern for has been known for many years aand is 4.2.2 Service factor, CSF and &F



8 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



S T D B A G V A bOZL-HSB-ENGL



L998 D Ob87575 O O O 5 8 L i 7 400 W 6011-H98



GMA STANDARD NATIONAL AMERICAN



high speed gear units. Mathematical methods have beendevised to assesstherelativerisk(see ANWAGMA 2101-C95,annex A); however, consensus on exact calculation methods and application of those methods have not been reached. With noconsensusonmethodsandapplication,this standard cannot include scuffing calculations. Further information is provided in annexB, but the gear designer is cautioned to use this information carefully andis reminded that the contents of annex B areforinformationalpurposesandarenot considered a part of this standard.



turetodropbelow 20°C atstart-up,thegear manufacturershouldbeadvised.Specialprocedures may be requiredto ensure adequate lubrication. 5.1.2 Environment



If agear unit is to be operated in an extremely humid, salt water, chemical, or dust laden atmosphere, the gear manufacturer must be advised. Special care must be taken to prevent oil contamination. 5.1.3 Temperature control



Theoiltemperaturecontrolsystemmustbedesigned to maintain an oil inlet temperature within design limits at any expected ambient temperature 5 Lubrication oroperatingcondition.Designinlettemperature may vary, but 50°C is a generally accepted value. Oil 5.1 Design and lubrication considerations temperature rise through the gear unit should be limited to 30°C. Special operating conditions, such Highspeedgearunitsshallbedesignedwitha as to provide lubrication and cooling to high pitch line velocity, high inlet oil temperature, pressure system A normal oil inlet pressure ofand high ambient temperature may result in higher the gears and bearings. operating temperatures. 1 to 2 bar is an industry accepted value. Special applicationsmayrequireotheroilpressures.Ifa 5.1.4 Gear element coolingand lubrication oil level in the gear gear element extends below the The size and location of the spray nozzles installed casing, it is said to be dipping in the oil. Dipping at by the manufacturer is ciiical to the cooling and high speed can result in high power losses, rapid proper lubricationof the gear mesh. overheating,possiblefirehazard,andshouldbe avoided. Spray nozzles may be positioned to supply oil at The following minimum parameters must be consid-either the in-mesh, the out-mesh,or both sides of ered to ensure that proper lubrication is provided the for gear mesh (or at other points) at the discretion of the gear manufacturer. the gear unit:



-



-



typeofoil;



5.1.5 Oil sump



oil viscosity; inletoilpressure;



The oil reservoir may be in the bottom of the gear 'case (wet sump) or in a separate tank (dry sump). In either case, the reservoir and/or gear case should be sized, vented, and baffled to adequately deaerate the oil and control foaming. In dry sump applications the external drainage system must be adequately sized, sloped and vented to avoid residual oil buildup in the gear case. Drain velocities may vary, but0.3 meters per second is a generally accepted maximum value.



inlet oil temperature; filtration;



-



oilflowrate;



-



coolingrequirements.



drainage; retention or settlingtime;



5.1.1 Ambient temperature



5.1.6 Filtration



The ambient temperature is defined as the tempera- A good filtering system for theoil is very important. ture of the air in the immediate vicinlty of the gear The design filtration level may vary, but filtration to a unit. The normal ambient temperature range for high 25micronnominalparticlesize is agenerally speed gear unit operation is from -10to 55°C. The accepted value. Finer filtration than 25 microns is gearmanufacturershouldbeinformedwhatthe recommendedwhenlightturbineoilsareused ambient temperature will be,if aor large radiant heat particularly for higher operating temperatures. It is source is located near the gearbox. Furthermore,if good practiceto locate the filteras near as possible low ambient temperature causes the sump temperato the oil inlet. Further, it is recommendedto provide



COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



STD-AGMA boll-HSB-ENGL ANSI/AGMA 6011-H98



L998



m



O b 8 7 5 7 5 0005848 347



m



AMERICAN NATIONAL STANDARD



.



speed increasers where the lubrication system is oil is integrated with the prime mover, a light turbine usually provided for the gear drive. This oil viscosity 5.1.7 Drain lines is lighter than any of the recommended AGMA oils. Thisexample of system”compromise”requires The location of the drain line must be in accordanceclose attention to the qualrty of theA oil. 90 VI oil or wlth the gear drive manufacturer’s recommendabetter should be employed. Special considerations so they are no tions. Drain lines should be sized may require the use of oils not listed in table 4. The more than hair f u l l . The lines should slope down at agearmanufacturershouldalways beconsulted minimum of 20 mm per meter and have a minimum when selecting or changingviscosity grades. number of bends and elbows. 5.2.2 Synthetic oils 5.2 Choice of oil Synthetic oils may be advantageous in some applications, especially where extremes of tempera5.2.1 Oil viscosity ture are involved. There are many types of synthetic Wahrespecttothechoiceofthe oilviscosity oils,andsomehavedistinctdisadvantages.The selected, the lube oil load carrying capacity of the gear oil manufacturer should be consulted before using film increases with the viscosrty ofoil.the Therefore, any synthetic oil. a high viscosity oil is preferred at the gear mesh. 5.3 Oil maintenance Development of an adequate elastohydrodynamic or EHL oilfilmthicknessandreduction in tooth The oil mustbefilteredandtestedorchanged roughness areof primary importance tothe life of the periodically to assure that adequate oil properties gearset. However, in high speed gearboxes, particularly those with high bearing loads and high journal are maintained. velocities,theheatcreatedinthebearingsis Prior to initial start-up of the gear unit, the lubrication considerable. Here, the lube oil viscosity must be It system shouldbe thoroughly cleaned and flushed. lowenough topermitadequatecooling ofthe isrecommendedthattheinitialchargeofoil be bearings. changed or tested after 500 hours of operation. It is obvious that the selectionof an appropriate oil 5.3.1 Change interval viscostty is a compromise of two main factors. In Unless the manufacturerrecommendsdifferent addition,lubricationsystemsareoftentimesinteintervals, under normal operating conditions subsegrated with other train equipment whose oil viscosity test intervalsshouldbe 2500 quentchangeor requirements are different from the gear. This further operating hours or 6 months, whichever occurs first. compromises the selection of the oil. Extendedchangeperiodsmaybeestablished throughperiodictestingofoils. Oilviscosrtyrecommendationsarespecifiedas With periodicoil AGMA oil numbers which are listed in ANWAGMA testingandconditioning, it isnotuncommon to 9005-DM. Recommendations for high speed operate lubrication systems without oil changes for applications are listed in table4. For turbine driven the lifeof the gear drive. a duplex filterto facilitate cleaningof the filter. Any kind of bypass of the filter is prohibited.



-



Table 4 Recommended oils



2



IS0 viscosity gmde (VG) 32 46 68



3



100



AGMA oil number O



1



Wscosity range mm2/$ ( C M )



at 40°C



28.8 to 35.2 41.4 to 50.6 61.2 to 74.8 90.0to 100.0



Reference viscosity approximate SSU at 100°F 150 215 315 465



Minimum viscosity index 90



90 90



90



NOTE: When operating at low ambient temperatures, the oil selected must have a pour point 3 to 6’C below the lowest



expected ambient temperature.



10 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



AMERICAN NATIONAL STANDARD



ANSI/AGMA 6011-HW



oped into a model in order to analyze the system as a whole and solve its torsional mechanical vibrations. Where operating conditions result in water collecting in the lubrication system,oilthe should be processed It is important to note that this result is only as good In fact,theprocessoflumping or changed as required to keep water content below asitsmodel. parameters could be the largest source of errors. the oil manufacturer’s recommendation. Failure to The result of the torsional system analysis is not control moisture may result in damage to the gear of the gear manufacturer, since the within the control unit.Some -oils arehygroscopicandmayneed gearbox itself is only one of several elements in a specialconsiderationtoeliminate orcontrolthe coupled train. water content andtotal acid number. The gearbox is only one of the influences on the modeling data and the conclusions derived from that data. For this reason, torsional vibration problems 6 Vibration and sound cannot bethe responsibiltty of the gear manufacturer. The gear unit manufacturer is seldom the system designer andin normal cases the gear unit manufac6.1 Vibration analysis turer is responsible only for providing mass elastic data to the system designer. Vibration of any component of the gear unit can result in additional dynamic loads being superim6.3Lateralvibrationanalysis . posed on the normal operating loads. Vibration of The rating equations used in this standard assume sufficient amplitude may resultin impact loading of smooth operation of the rotors. To insure smooth the gear teeth,interference in thegearmeshor operation, these rotors should be analyzed for lateral damage to close clearance parts of the gear unit. critical speeds. Also, it is imperative that slow roll, Whereunusualtorquevariationsdeveloppeak startup,andshutdownofrotatingequipmentnot loads which exceed the application power by a ratio cause any damage as critical speeds are passed. greater than the factor CSFor &F specified for the See annex C. application, the magnitude and frequency of such torque variations should be evaluated with regard 6.3.1 to Lateral critical speed map the endurance and yield properties of the materials Anundampedlateralcriticalspeedanalysis is used. sufficient in some casesto determine rotor suitabiltty. If this method is chosen as the sole criterion for The types of vibration which are generally of concern determining the suitability of a rotor, it should be forgearunitsare the torsional, lateral andaxial based upon significant experience in designing high modesof the rotating elements, since these can speed gear drives utilizing this method. It includes a have a direct influence on the tooth load. Of these, lateral critical speed map, showing the undamped the two thatarenormallyreviewedanalytically critical speeds versus support stiffness or percentduring design are the lateral critical speeds of the age of torque load. The graphic display shows all gear unit rotating shafts and the torsional critical applicableloadingconditionsandno-loadtest frequencies ofall connected rotating elements. conditions (approximately 10 percent of the rated torque) at the maximum continuous speed. 6.2 Torsional vibration analysis 5.3.2 Water contamination



Any torsional vibration analysis must consider the complete system including prime mover, gear unit, drivenequipmentandcouplings.Dynamicloads imposed ona gear unit from torsional vibrations are theresultof the dynamicbehavior of theentire system andnot thegear unit alone. Thus a coupled system has to be analyzed in its entirety by first separating its properties into a series of discrete spring, connectedmasses.Whenappliedtoa multi-mass system, this method is knownas using lumped parameters. These parameters are devel-



The critical speed map for gear rotors is used to determine potential locations of the critical speeds by locating the intersection of the principal bearing stiffness values with the undamped critical speeds. If no intersections are indicated, with experience this can be used to determine rotor suitability. Note that these undamped speeds can be significantly different from the critical speeds determined from a rotor response to unbalance analysis. The differences are due to the cross coupled stiffness and damping effectsfrom the bearings.



11 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



S T D - A G H A bOLl-HS8-ENGL



1998



ANSVAGMA 6011-H98



6.3.2 Analytical methods



= Ob87575AMERICAN NATlONAL TT5 9 0005850



STANDARD



than 20 percent; or(b) the vibration levels are within the specified limit and the amplification factor is less than 2.5 (see 6.3.3.3).



Coupling moments and shear force transfer effects between rotors with properly designed and installed In some cases a simple undamped lateral critical couplings will beminimal. As a result, each coupled speed analysis may be sufficient to properly analyze element can generally be analyzed independently. the rotor. The mathematics of this analysis are complex and beyondthescopeofthisstandard.Commercial 6.3.3.1 Forcing phenomena computersÖftwareisavailableandtheanalysts A forcing phenomenonor exciting frequency may be the method they use should assure themselves that less than, equalto, or greater than the synchronous gives accurate results for the type of rotors being frequency of the rotor. Potential forcing frequencies analyzed, Most high speed rotors are supportedin may include but are not limited to the following: hydrodynamic journal bearings; therefore, of equal importance is themethodused to analyzethe - unbalancein the rotorsystem; support (bearing) stiffness and damping. - oilfilmfrequencies; The analyses should include the following effects on the critical speeds:



- The bearing-oil film stiffnessand dampingfor the rangeof bearing dimensions and tolerances, load and speed; - The bearing structure and gear casing support structurestmess;



-



internalrubfrequencies; - gear-meshingandside-bandfrequencies, as well as other frequencies produced by inaccuof the gear tooth; racies in the generation



-



couplingmisalignmentfrequencies; looserotor-systemcomponentfrequencies;



-



asynchronouswhirlfrequencies.



- The coupling weightto be supported by each 6.3.3.2 Rotor response analysis gearbox shaft (the weight the of coupling hub plus The rotor responseto unbalance analysisis used to 1/2 the weight ofthe coupling spacers).The couof the rotor pling weight shall be applied at the proper centerpredict the damped vibration responses to potential unbalance combinations (¡.e., criitical of gravity relative to the shaft end. The weight and of a gear rotor speeds). The critical speeds the center of gravlty will be specified by the purdetermined from the rotor response analysis should chaser of the coupling; field test data. - The potential unbalanceof the gear rotor and be verified by shop and coupling. 6.3.3 Lateral critical speeds



The rotor response analysis should consider the following parametric variations in order to assure that the vibrations will be acceptable for all expected conditions:



Lateralcriflicalspeedscorrespondtoresonant frequencies of the rotor-bearing support system. 1. Unbalance,g-mm The basic identification of critical speeds is made from the natural frequencies ofthe system and of the midspan unbalance o 6400 W , forcing phenomena. If the frequency of any harmonic component of a periodic forcing phenomenon is - couplingunbalance 64 WcpI equal to or approximates the natural frequency of of resonance any modeof rotor vibration, a condition - out-of-phase unbalance 64 m Wcpl at may exist. If resonance exists at a finite rotational 3200 peak speed, the speed at which theresponse occurs coupling and at the furthermost mass is called a critical speed. The speed or frequency at station on the gear tooth portion of the gear. which these occur varies with the degree of transwhere mitted load, primarily as a result of the change in stiffness of the bearing oil film. o is the speed of rotor, rpm; Criticalspeedsarenormallydeterminedusing a rotorresponseanalysisandaredeemedtobe acceptable if: (a) the separation margin is greater 12 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



W, is the total weight of the rotor, kg; Wcpi



is the half weightofthe coupling and spacer, kg-



S T D - A G H A bOLL-HSB-ENGL



L q q A D Ob87575 OUCI5853 93L ANWAGMA 6011-H98



AMERICAN NATlONALSTANDARD



NOTE: The shape of the curve is for illustration only and does not necessarily represent anyrotor actual response plot. In most cases the amplitude does not decrease to N. (0.707 of peak); therefore calculate N. from the "flip"of NI,or use another method suchthe as amplification factor in the 'Handbook of Rotordynam¡Cs"by F.F. Ehrich, page 4.28.



2. Gearloading



-



-



unloaded and/orminimumload;



50 percentload; 75 percentload;



100 percentload.



6.3.4 Stability analysis



3. Bearingclearances



Damped eigenvalues (damped critical speeds) may occur below 120% maximum rotor speed due to a variationinload,bearingproperties,etc.These maximumclearanceandminimumpreload. damped eigenvalues are the frequencies at which 4. Speedrangefromzero to 130 percent of the rotor will vibrate if there is sufficient energy or maximum rotor speed. insufficientdampinginthesystem.Therefore,a damped stabilrty analysis is performed to insure that 6.3.3.3 Amplification factor thesedampedeigenvalueshavealargeenough Theamplificationfactorisdefined as the critical logarithmic decrement (log dec) to insure stabilrty. speed dividedby the band width of the response freThe stability analysis calculates the damped eigenquencies at the half power point. The response of a values and their associated logarithmic decrement. criitical speed is considered to be critically damped if The rotor should have minimum log dec of +0.1 at the amplification factor is less 2.5 than (see figure1). any of the damped eigenvalues to be considered stable. Operating 6.3.5 Mode shape speed



-



minimumclearanceandmaximumpreload;



irn Shaft speed, rpm Nc,



is therotorfirstcriitical,centerfrequency, rpm;



Nmc



isthemaximumrotorspeed,rpm;



NI



is the initial (lesser) speed at0.707 x peak amplitude (critical), rpm;



N2



isthefinal(greater)speedat 0.707 x peak amplitude (critical), rpm;



N2-N1 is the peak widthat the half power point;



AF



is the amplification factor= Nct



SM



is theseparationmargin;



CRE



is the critical response envelope;



kt



is theamplitude at&. Figure 1 - Amplificationfactor



.



Each finite resonant frequency has an associated mode shape. Knowing the mode shape that the rotor will assume when responding to a critical speed is important in understanding the consequencesof bearing placement and residual unbalance. In most high speed gear unit rotors the mode shape of the first criitical speed is mostly conical with a node point betweenthebearings,vibrationatthebearings approximately 180" out of phase, and the point of highest vibration at the drive (coupling) end of the shaft. Aslight bending shape of the rotor is common. The amplitude at the bearing locations is usually high enough to allow the damping inherent in hydrodynamic journal bearings to limit maximum vibration amplitudes. However, the location of highest amplitude at the coupling makes most gear units sensitive to unbalance at this location and extra care in coupling balance is recommended. 6.4 Balance



All gearrotatingelementsshallbemultiplane dynamicallybalancedafterfinalassemblyofthe rotor. Rotors with single keys for couplings shall be balancedwith their keyway fitted with a fully crowned its entire for half-key so that the shaft keyway is filled length.Thebalancingmachineshall besuitably calibrated,withdocumentation of thecalibration



13 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



S T D - A G M A bull-HS8-ENGL



L998



m



Ob87575 0 0 0 5 8 5 2 8 7 8



ANSI/AGMA 6011-H98



m



AMERICAN NATIONALSTANDARD



available. The rotating elements should be balanced at any other speed within the specified range of operating speeds, casing vibration as measured on to the level of the following equation: thebearinghousingshallnotexceedthevalues ...(4) shown in table5. where



-



U,,,, is the amount of residual rotor unbalance, W W



Table 5 Casing vibration levels



g-mm; is the journal static loading, kg; 2.5 kHz



is themaximurncontinuousspeed,rpm.



-



2.5 kHz 10 kHz



6.5 Shan vibration



During the shop test of the assembled gear unit or at any operating atits maximum continuous speed other speed within the specified rangeof operating speeds, the double amplitude of vibration for each shaft in any plane measured on the shaft adjacent and relative to each radial bearing shall not exceed the following value or50 pm whichever is less: A =



m J q



...(5)



where A



W



is theallowabledoubleamplitude of unfiltered vibration, micrometers(Pm) true peak to peak; is themaximumcontinuousspeed,rpm.



6.5.1 Electrical and mechanical runout



When provisions for shaft non-contact eddy current vibrationprobesaresuppliedonthegearunit, electricalandmechanicalrunoutshall be deterat the journal mined by rolling the rotor V-blocks in bearing centerline or on centers true to the bearing journals while measuring runout with a noncontactingvibrationprobeandadialindicator.This measurement will be taken at the centerlineof the probe location and one probe tip diameter to either side and the results included with the test report. 6.5.2 Electrical/mechanicaI runout compensation



If the vendor can demonstrate that electrical/mechanical runoutis present, a maximum 25 of percent of the test level calculated from equation5 or 6.4 micrometers,whichever is the lesser,maybe vectoriallysubtractedfromthevibrationsignal measured during the factory test.



6.6 Casing vibration During shop no-load test of the assembled gear its maximum continuousspeed or drive operating at 14 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



NOTE: The levels shown in table 5 are for horizontal offset gear unitsonly. The allowable vibration levelsfor vertical offset gear units are twice those shown in the



table. 6.7 Sound measurement



The sound level measurement and limits shall be in accordancewithANSI/AGMA 6025-C90 unless otherwiseagreeduponbythepurchaserand manufacturer.



7 Functional testing 7.1 General



Eachunitconforming to thisAmericanNational Standard should be functionally tested as a minimum atfull speed. Functional testing provides a means of evaluating operational characteristics of the unit. The procedures may the be manufacturer’s standard or one agreed upon by the manufacturer and the customer. 7.2 Purpose



Functional testing presents an opportunity to evaluatetheoperationalintegrrty of thedesignand manufacture of gear drives. Functional test procedures provide a means of evaluatingthe entire gear system for noise, vibration, lubrication, gear tooth contact,bearingoperatingtemperatures,bearing stabilrty,lubricantsealing,mechanicalefficiency, instrument calibration and other unit features and providedatathatparallelstheexpectedon-line operational characteristics. 7.3 Procedures



Functionaltestingmayalsoincludeprocedures ranging from partial speed and no load spin testing to full speed and full power testing. Following testing,



ANSIIAGMA 6011-H98



AMERICAN NATIONAL STANDARD



the unit may be disassembledfor bearing and gear tooth contact inspection. 7.3.1 Spin testing-no load



-



Onegearunit'srotatingelementsmaybe loaded on the normally unloaded flanks of the teeth; - If therotatingelementshavemodifiedhelices, one unit may be torqued in the wrong direction.



The unit under test is normally driven in the same rotational direction and at the same shaft as in the design application. The output shaft will have no 7.3.4 Special testing load applied it. töTest speeds may range from partial In the caseof very high rotational speeds or multiple speed to over speed. Thetest duration should be no less than one hour after temperature stabilization. input/outputshafts,conventionaltestingmaybecome impractical. In such cases special test 7.3.2 Partial load testing proceduresspecific to theapplicationshouldbe developed between the manufacturer and the The unit under test is normally driven in the same customer. rotational direction and at the same shaft as in the 7.3.5 Power loss testing design application. The output shaft will be connected to aloadingdevicewhichappliesa When testing for power loss in a high speed gear resisting torque less than the design f u l l load torque. unit, the normal procedure is to measure the heat Test time should be no less than one hour after removed by the oil flowing through the gear unit. The temperature stabilization. oil inlet temperature, oil outlet temperature and oil loss isthen flowaremeasuredandthepower calculated ignoring the heat dissipated from all other Full speed and full power testing can be carried out areas in except the oil. Due to aeration of the drain oil, the same manner as described in 7.3.2 for units calculate powerloss using the following equation: rated in the lower rangeof the power scale. PL = 0.027(Qum) (W ...(6) Full power testingof units with higher power ratings where may require back-to-back locked torque testing. In PL is thepowerlosses, kW; this proceduretwo identical mirror image units are Q L ~ BisEthe oil flow, Ihnin; shaft coupled together, inputto input and output to AT isthechange in lubricanttemperature torque is. applied by output. Full operational "C. from inlet to outlet, disengaging one ofthe shaft couplings, rotating the shafts relativeto one anotheruntil the proper torque The constant0.027 in equation6 is forIS0 VG 32 oil. is achieved, then re-engaging the shaft coupling. When using other VG grades, other values may be The unit shafts are then rotated at full speed. Full used, based on experience. power testing duration is usually not less than four hours after temperature stabilization.



7.3.3 Full speed andfull power testing



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



-



Bearingswith f u l l loadappliedatthestatic condition willstart with f u l l load and no hydrodynamicoil filmuntil 'some" rotationalspeedis reached;



- Bearings of one unit willbe loaded in a direction opposite normal operation; - Gear and pinion teeth withfull load applied at the static condition will start with full load and no oil film to separate the teeth until "some" rotational speed is reached.Scuffingmayoccur;special procedures suchas plating of the gear teeth with an EP lubricant may be required;



8 Vendor and purchaser data exchange 8.1 Rationale for data requirements



In order to promote consistency and reduce errors, recommended information to be furnished by the vendor is specified in this section. A detail of the schedule for transmission of drawings, curves and data should be agreed to at theoftime the proposal or order. The purchaser should promptly review the vendor's data when he receives them; however, this review does not constitute permissionto deviate from any requirementsin the order unless specifically agreed upon in writing. After the data has final approval, the



15 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



AMERICAN NATIONAL STANDARD



ANSI/AGMA 6011-H98



vendor should furnish certified copies inq uthe anti specified.



-



couplingselection;



thrust and radial bearing sizing and estimated loading;



A complete listof all vendor data should be included rotor dynamics; withthefirstissue ofmajordrawings.This list contains titles, drawing numbers and a schedule for any special paint or environmental protection transmission of all data the vendor will furnish. requirements;



-



Inquiry documents should be revised to reflect any subsequent changes. These changes will result in the purchaser’s issue of completed, corrected data sheets as partof the order specifications. 8.2



Document identification



Transmittal (cover) letter title blocks or title pages when shouldcontainthefollowinginformation, available:



-



-



thepurchaserluser’scorporate name; thejoblproject; theequipmentitemnumber; the inquiry or purchase order number;



any other identification specified in the inquiry or purchase order; - thevendor’sidentifyingproposal number, shop order number, serial number, or other reference required to completely identify return correspondence.



8.3 Hems needing resolution The items listed below normally should be resolved after purchase commitment. This may be donea at coordinationmeeting,preferablyatthevendor’s plant or by other suitable means of communication.



-



the purchase order, scope of supply, vendor’s internal order details and subvendor items;



-



anyrequireddatasheets;



applicablespecifications,standards,clarifications and previously agreed upon exceptions;



-



that the system and all its components are in accordance with specified standards;



-



schedules for transmittal of data, production and testing;



- qualityassuranceprogram,proceduresand acceptance criteria; -



inspection,expeditingandtesting;



schematicsandbillsofmaterial auxiliary systems;



(B/Ms)of



-



thephysicalorientationofequipment,shaft . rotation,pipingandauxiliarysystems;



16 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



-



whether SI or Imperial units areto be shown.



8.4 Proposal data



The following is a guide to proposal data that should be furnished.



- A general arrangement or outline drawing for each gear unit showing overall dimensions; - The purchaser’s data sheets, with completed vendor’s information entered thereon and literature to fully describe details of the offering;



- A schedule for shipmentof the equipment, in weeks after receipt of the order, and all approved drawings; - A l i s t ofrecommended start-upspares,including any items that the vendor’s experience indicates are likely to be required; - A completetabulationof the utilityrequirements, including the q u a n t i of lube oil required and the supply pressure,the heat loadto be removed by the oil and the nameplate power rating. (Approximate data shallbe defined and identified as such.);



- A description of the tests and inspection procedures, as required; - Anystart-up,shut-down,oroperatingrestrictions required to protect the integrity of the equipment; -



When requested, the vendor should furnisha list of the procedures for any special or optional tests that have been specified the purchaser by or proposed by the vendor;



- The conditions and periodof the manufacturer’s warranty. 8.5 Contract data



The following lists the contract data normally supplied by the vendor. a. Certifieddimensionaloutlinedrawingand parts list, including the following: - the size, rating and location of all of the purchaser’s connections; - approximate overall and handling weights;



-



overalldimensions;



ANSIIAGMA 6011-H98



AMERICAN NATIONAL STANDARD



-



dimensionedshaftend@)forcoupling mounting(s);



-



the height of the shaft centerline;



-



the dimensions of baseplates or soleplates with thediameter, numberandlocationofbottholesandthe thickness of themetalthroughwhichbolts must pass; (if furnished),complete



- shaftpositiondiagram,includingrecommended limits during operation, with all changes in shaftendpositionandsupport growths from an ambient reference or 15°C noted; -



journal bearing clearances and tolerances; axial rotor float or thrust bearing clearance,



as applicable;



-



the number of teeth on eachgear.



b.When lube-oilsystem is supplied,aschematic, certified dimensional outline drawing, and parts list including the following:



- control, alarm and trip settings (pressures and recommended temperatures); - utilityrequirements,includingelectrical, water and air; -



pipe and valve sizes;



- instrumentation,safetydevicesandcontrol schemes; -



size, rating and location of all the purchaser's connections;



-



instruction and operation manuals;



maximum, minimum and normal liquid levels in the reservoir.



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



-



vibration warningand shutdown limits;



-



bearingtemperaturewarningandshutdown limits;



- lube-oiltemperaturewarningandshutdown limits. d. Lateral criiical speed analysis, whichmay include any or all of the following:



-



themethodused;



a graphic display of bearing and support stiffness and their effects on critical speeds (undamped lateralcritical speedmap); -.



- a graphic display of the rotor response to unbalance, including damping (rotor response analysis); - journalbearingstiffnessanddamping coefficients; - damped stability analysis, including identified eigenvalues and associated logarithmic decrement. e. Torsional data sufficient for a third party to do a system torsional analysis. f.



Whenmechanicalrunningtestissupplied,



test reports, including the following (see clause 7):



- vibration; - lube oil flow and inlet and outlet temperatures; - bearingtemperatures. g. Nameplatesandrotationarrowsshallbeof Series 300 stainlesssteelor of-nickel-copper alloy (Monel orits equivalent) attached by pins of similar material and located for easy visibility. As a minimum, the following data shouldbe clearly stamped on the nameplate: - the vendor's name; - the size and typeof the gear unit; - the gear ratio;



-



theserialnumber; the service power, P,;



theratedinputspeed,inrevolutionsper minute;



- the rated output speed, in revolutions per minute; - the gear service factor, as deiined in standard; -



-



this



the purchaser's item number; the number of gear teeth; the number of pinion teeth;



- date of manufacture: month and year unit was successfully tested. h. A statement of anyspecialprotectionrequired for start-up, operation and periods of idleness under the site conditions specified on the data sheets. Thelist shallshow the protection to be furnished by the purchaser, as well as that included in the vendor's scope of supply. 8.6 Installation manual



Whenspecifiedbythepurchaser,aninstallation manual shall be supplied. Any special information



17 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



STDmAGMA bOLL-HSB-ENGL ANSIIAGMA 6011-H98



L778 W O b 8 7 5 7 5 00058Sb q13 AMERICAN NATIONAL STANDARD



titles, and a complete required for proper installation design that is not on index sheet containing section list of referenced and enclosed drawings by title and the drawings shall be compiled in this manual. This drawing number. The manual shallbe prepared for manual shall be forwarded at a time that is mutually is not agreed upon in the order. The manual shall contain thespecifiedinstallation;atypicalmanual at a information such as special alignment and grouting acceptable.Thismanualshallbeforwarded procedures, utilrty specifications (including quantitime that is mutually agreed uponin the order. This manual shall contain a section that provides special ties) and all other necessary installation design data, instructions for operation at specified extreme enviincluding drawings and data specified in 8.5. The ronmental conditions, suchas temperatures. manual shall also include sketches that show the location of the center of gravity and rigging provi8.8 Recommended spares sions to permit the removal of the top half of the When the vendor submits a complete list of spare casing, rotors and subassemblies that weigh more all parts,thelistshouldincludesparepartsfor than 140 kilograms. equipment and accessories supplied. The vendor should forward the list to the purchaser promptly 8.7 Operation, maintenance and technical after receiptofthe reviewed drawings andin time to manuals permit order and delivery of the parts before field When specified, the vendor shall provide sufficient start-up. list of all written instructions and a cross-referenced 8.9 Special tools drawings to enablethepurchaser tocorrectly A list of special tools required for maintenance shall operateandmaintainalltheequipmentordered. be compiled This information,when required, should be furnished. The vendor shall identify any items in a manual or manuals with acover sheet containing included in the offering as to whether arethey inch or 8.2, an allreference-identifyingdataspecifiedin metric.



18 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



S T D . A G f l A bOLL-HSB-ENGL



1798



Ob87575 0 0 0 5 8 5 7 35T 6011 -H98



GMA STANDARD NATIONAL AMERICAN



Annex A (informative)



Service factors rheforeword,footnotes and annexes, if any, are providedfor informational purposes onlyand should not be construed asa part of ANWAGMA 6011-H98,Specirïcation for High Speed Helical Gear Units.]



A1 Purpose



A.2.2 Driven equipment characteristics



Thisannexprovidesdetailedinstructionsforthe determinationanduseofservicefactorsforenas described in closed high speed helical gear units



Drivenequipmentcangenerallybedivided into rotary and reciprocating types of machines. Rotary machines generally have smoother power requirements than reciprocating machines, but type eachis unique and the equipment characteristics of each must be known to be properly evaluated.



AGMA 6011-H98.



A2 Determination of senrice factors



The determination of service factor is based on the equipment characteristic overload of the gear as unitA.2.3 System conditions a resultof operation, the desired reliability of the gear The gear unit is apart of a system and this system unit during its design life, and theof time lengththatis canhavedynamic(vibratory)responseto time consideredthedesignlife.Itreliesheavilyon varying(dynamic)powertransmissionthatmay experience acquiredin each specific application. A overload the gear unit. This is most commonly found broad explanation of the factors involved are: as torsional vibration in the rotating shafts, but can



-



beanyvibratoryresponse to dynamicexciting The causes of service overloads are broken into three broad categories: those produced by forces.Generally,overloadsareassumed to be the prime mover, those produced by the driven transmitted with no amplification through the gear, equipment, and those resulting from system con-but whenthere is a resonant responseto a dynamic siderations uniqueto the equipment train; power overload, a much higher load may occur the at gear unit. - The reliabilityof a geared system depends on many factors both internalto the gear unit itself Thus,thedynamicoverloadsthatarecausedby and externalto the unit. Increases in service factor to influence reliability normally take into con- prime movers and driven machines may be amplified in such a way as to greatly increase their magnitude siderationexternalsources offailuresuchas at the gear unit, and primarily at the gear tooth mesh. abuse and unexpected operating conditions; Thenormalratingofgearunitsandthe normal - The desired lifeof most high speed enclosed service factors used assume that these responses drives is usually longer than other types of en(resonances) do notappreciably affect the gear unit closed drives. At high operating speeds this can load. Therefore, careful system analysis is recomtranslate into a very large number of stress cycles mended to ensure that no unexpected overloads on the components. due to resonances are present. A.2.1 Prime mover characteristics A.2.4 Reliability and life requirements Some different typesof prime movers are electric or in the power rating hydraulic motors, steam or gas turbines, and single There isareliabilityfactor equations, butit deals onlywith the statistical nature or multiplecylinderinternalcombustionengines. of materialtestingandprobabilityoffailurefor Each of these prime movers is designed to produce materials at a given stress level. In a gear unit there somenominalpower,buteach will producethis power with some variation over time. The variation are many separate components that may fail, many modes of failure, and many factors that can contrib of power output with time may be lower or higher of failure.Forthisreason, depending on the prime mover and also the way Ute thetothosemodes quantifying factors associated with reliability and life prime mover is applied in a particular' machinery to account for these external issues can be extremetrain, but any variation over nominal power is an ly difficutt. overload and must be considered.



19 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



S T D - A L M A bOLL-HSB-.ENGL 6011



ANSI/AGYA



1998



-H98



O b 8 7 5 7 5 0 0 0 5 8 5 0 27b STANDARD NATIONAL AMERICAN



m



-



A.3 Service factor table Service factors have served the industry well when theyhavebeenidentifiedbyknowledgeableand experiencedgeardesignengineers.Theservice factorsshown in table A l havebeenusedwith as success in the past. These values may be used general guidelines, but they do not eliminate the re responsibility of defininganyunusualsystem quirements that would alter the listed values.



Electric motors that have electric power interrupted and then re-applied before the shafts have stopped rotating produce very high torques;



- Synchronouselectricmotorscanproduce very hightorsional forcing functions during startup. This can cause very high transient torsional torques onthe gear unit;



- Generators have extremely high loads when they areout of phase with the main system, and across-the-lineelectricalshorts can produce very high torque loads. For this reason torque lirnA3.1 General selection guidelines iting devicesor higher service factors are advisThere is no wayto list all the possible considerations able; that may affect selectionof service factors,but the - Brakes or other decelerating devices can profollowing are some guidelines. duce loads on the gear unit larger than the transmitted power. - Inductionelectricmotorscanproducehigh but the intent here torques on start-up. Therefore, on an application The list could be much longer, with many starts, higher service factors may be is to give a general idea of items to consider when selecting service factors. warranted;



-



Table A1 Service factors,CSFand &F



T Appllcation



3lowers Centrifugai Lobe Sompressors Centrifugal process gas, except air conditioning air conditioning service air or pipe line service Rotary axial flow - all types liquid piston (Nash) lobe - radial flow Reciprocating 3 or more cylinders 2 cylinders lynarnometer - test stand -ans Centrifugal Forced drafl Induced draft Industrial and mine (large with frequer It stal ìenerators and exciters Base load or continuous Peak duty cycle



1 Service factor.with Drime mover Synchronous motors lnductlon Sas or steam or internal cornbustlon engine (multi-cyllnder) motors turbine’) 1.4 1.7



1.6 1.7



1.7 2.0



1.4 1.2 1.4



1.6 1.4 1.6



1.6 1.6 1.7



1.7 1.7 1.7



1.7 1.7 1.7



1.7 2.0 2.0



2.0



2.0



2.0



2.0 1.1



2.0 1.1



2.3



1.4 1.7 1.7



1.6 1.6 2.0 2.0



1.7 1.7 2.2 2.2



1.3 1.4



1.3 1.4



1.4



1.4



1.3



1.7 (continued)



20 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



~~



~



S T D - A G H A bOLL-HSB-ENGL



~



L778



Ob87575 0 0 0 5 8 5 9 1 2 2 E



AGMA STANDARD NATIONAL AMERICAN



-H98



Table Al (concloded)



I Application



'aper industry Jordan or refiner Paper machine- line shaft



Service factor, with prime mover



I Synchronousmotors T Induction Gas or steam orinternal combustion motors turbine') engine (multi-cylinder) "



1.5



"



"



1.3



"



%rips



Centrifugal (all service except as listed below) 1.3 1.5 Centrifugal boiler feed 1.7 2.0 descaling (with surge tank) 2.0 2.0 hot oil 1.7 2.0 pipe line 1.5 1.7 water works 1.5 1.7 Reciprocating 3 or more cylinders 2.0 1.7 2 cylinders 2.0 2.0 Rotary axial flow- all types 1.5 1.5 1.5 1.5 gear type liquid piston 1.7 1.7 lobe 1.7 1.7 sliding vane 1.5 1.5 iugar industry Cane knives 1.5 Crushers 1.7 Mills 1.7 SOTES: Gas turbines seldom operate at full design power while steam turbines often operate at or Qpropriate design considerations should be made to assure adequate torque capacity. "



"



"



1.7 "



"



"



2.0 2.0 2.0



.2.0 1.8 1.8 2.0 2.0 1.8



1.8 2.0 2.3



above rated power.



21 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



S T D - A L M A bOLL-HSB-ENGL



1998 W Ob87575 00058b0 9'411



STANDARD MA NATIONAL AMERICAN



6011-H98



Annex B (informative) A simplified method for verifying scuffing resistance pheforeword, footnotes and annexes, if any, are providedfor informational purposes onlyand should not be construedas a -H98, Specifcation for High Speed Helical Gear Units.] pari of ANWAGMA 6011



B.I Purposé



C,



= 1.1 O (conservative value);



Thisannexprovidesinformationconcerningthe scuffing (scoring) of high speed gear units.



C,



= 1.15 (nominalvalue);



C,



= 1.20(maximumvalue).



B.2 Scuffing considerations



NOTE: C, values are suggested values. Manufacturer's own experience may change these values with



AGMA 6011-H98 isconcernedwith two failure modes in gear teeth. They are surface pitting and root bending fatigue failure of the tooth material for a givennumber of stresscycles.There is another known failure type: scuffing (sometimes referred to as scoring). is avery The calculationof the scuffing load capacity complex problem. While this type failure has been known for many years and mathematical methods havebeendevised to assessrelativerisk(see ANSI/AGMA 2101-C95, annex A), asimplified scuffing criterion is suggested that is suitable for general high speed design work. From the valuesof tooth loading, pitch line velocity andviscosity of thelubricant,acondensedload function, F (load), is formed,which,toassure scuffing resistance, must be less than (or equal to) the geometric function, F (geometric). The geometas ric functionis based on gear characteristics such number of teeth of thepinionandgear,center distance and gearsetratio. As long as the value of the load function, F (load), does not exceed that of F (geometric),thereis thegeometricfunction, adequate safety against scuffing.



...(8.1)



Load function,F (load):



F (load) = [w'/Cw][ v ' ] ~ [. 4~ 6 / ~ 4 ] ~ - ~...(~B.2) where W'



-



Table 6.1 Lubricant viscosities AGMA lubricant number O 1 2



IS0 viscosity



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



Nominal viscosity at 40°C, rnm*/sec (cst) 22



32 46 68



NOTE: For high speed gearset applications, lubrjcant viscosity means lightturbine oilwith littleor noadditives based on a viscosity range of: 32 5 v a 5 68. The standard FZG oil test, DIN 51354, gives approximations for the lubricant with respectto scuffing tendency.



Geometric function,F (geometric):



F(geometric) =



(50 + L ,



+ z ~ ) ( u ) ' -[C,] ~ -**(B-3)



A



where:



Therefore:



F (load) 5 F (geometric)



supporting data.



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



V'



is thepitchlinevelocity, m/s;



V@



is the viscosity of lubricant at 40"C,mm2/s (CS);



LI



is the number of teeth of the pinion;



z2



is the number of teeth of the gear;



U



isthecenterdistance,mm;



U



isthe gearratio = Z&



A



= 300 for pressure angle ap = 20 deg;



For1 s u c 3 c, = IO$ + 333 [3 - U] For3 IU



...(B.4)



I10



C, = 130 - 10.0 [lo9- (13



- u ) ~ ] ' . ~ ...(B.5)



Previous page is blank COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



23



S T D - A G M A bOLL-HSB-ENGL



L778 W Ob87575 0 0 0 5 8 b L 880 D



ANSIIAGMA 6011-H98



AMERICAN STANDARD NATIONAL



8.3 Field of application



2. Pressure angle, g p = 17.5 deg



The above scuffing criterion is applicable to:



A=300



a. High speed gears with a modified addendum resulting in reasonably balanced sliding and rolling conditions between the tooth flanks at the tip of the pinion and mating gear;



For1



Iu < 3



Cu = 90 + 30 [3



-



...(B.8)



U]



b. Gear tooth accuracy grade, per IS0 1328-1,



shall be equal to or better than: Q5 (AGMA 12) for single pitch deviation, rpl Q5 (AGMA 12) for total cumulative pitch deviation, Fp



3. Pressure angle, ap = 22.5 deg A=250



Q4 (AGMA 13) for total profile deviation, Fa



Q3 (AGMA 14) for total helix deviation,Fß c. Surface roughness of tooth flanks after grinding & s 0.5 pm (20 rms); d. Basic rack profile with: pressure angleap = 20 deg addendum hap = 1 module.



For1 S u e 3 Cu = 95 + 28.6 [3 - U]



10 Cu = 130 - 10 [112.5 - (13 - u ) ~ J O . ~



For3



5 U 5



4. Pressure angle,



The working flanks of the pinion or gear shall be providedwithprofilemodificationstoobtaina trapezoidal tooth load distribution along the path of contact. The working flanks of the pinion or gear shall be provided with longitudinal modification to compensateforbendingandtorsionaldeflectionsand thermal deformations of the gear rotors in order to obtainauniformtoothloaddistributionoverthe entire rated face width.



...(B.10)



...(B.ll)



ap = 25 deg



A=250



For1 5 u e 3 Cu = 105 + 31.4 [3 - U] For3 I



...p. 12)



U I10



Cu = 140 - 10 [133.5



- (14 - u)2]0.5



....(8.13)



BA Scuffing design criteria



1. Pressure angle, 4p = 15 deg



As stated, there are no firm criieria.for designing to prevent scuffing at this time. However, it is hoped that the use of methods such as those in this annex and those in ANSI/AGMA 2101-C95 can lead to a set of design criteria. There are other methods for predicting scuffing and there is no intentto deny the validity of any method at this time.



A = 350



8.5 Conclusion



Lubricants applied shall conform with an FZG L 6 load stage(DIN51354, IS0 14635). For applications employing other pressure angles where manufacturers have had successful experience, the following formulas are suggested:



For1 5 u < 3 Cu = 95 + 28.6 [3 - U]



For 3 s



10 C, = 130 - 10 [112.5- (13 - u)~]O-S



...(B.6)



U I



,..(B.7)



24 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



Predicting scuffing is very important in high speed gearing. It is hoped that industry consensus can be reachedonscuffingprediction. To achievethis consensus, industry must utilize available methods and gain experience.



Annex C (informative)



Lateral rotor dynamics rheforeword,footnotesand annexes,if any, are providedforinformationalpurposes onlyand should be notconstrued as a part of ANSIIAGMA 6011-H98,Specification for High Speed Helical Gear Units.]



c.1 Purpose



mainly a result of the rotor weight and is therefore constant.



Inthe dynamic analysis of a high speed gear box, it is necessary to veri¡ that the drive is inherently stable, andthatanyactualharmfulcriticalspeedsare sufficiently removed from any operating speed or load range of the equipment. This annex provides information on rotor dynamics for high speed gear drives. C.2 Modes



High speed gear drives are frequently coupled to turbomachinery. Although the gear drive operates turbomachineryspeeds, its dynamicbehavioris significantly different from compressors or turbines. Gear shafting is generally of the rigid rotor design. This meansthatthroughouttheoperatingspeed range of the machine, most vibration that occurs is causedbyshaftdisplacementsinthebearing system oil films rather than deflectionsof the rotor l ) . (see figureC



Figure C.l



- Typical modes of rigid rotor lateral vibration



-



Figure C.2 Typical modes of flexural lateral vibration .



at High speed gear drives use fluid film or sleeve type bearings.Theyfrequentlyaremanufacturedwith non-cylindrical bores. Gear drive bearings generally have a large length to diameter ratioto gain the bearing area requiredto support the torque load as wellasrotorweightloadingand stillbeable to maintainhighefficiencies.Thistypeofbearing design lends itself to asymmetrical oil film stiffness rates in thexand Ydirections. High stiffness values occur in the direction of the applied load. Relatively large cross coupled stiffness and damping coefficients are common. Bearing cross coupling spring anddamping, in simple terms, meansthat, in addition to a resulting resisting force being generated in thedirectionofdisplacementorvelocity, 90 degreesfromthe another forceiscreated direction of motion. This phenomenon has a more pronouncedeffectingeardrives than in turbo equipment, which frequently uses tilting pad type bearings. For an accurate analysisa gear of drive, a complete eight element matrix of spring and dampC.3). ing rates should be obtained (see figure



Typical turbomachinery equipment can pass through what is called flexural type critical speeds Stiffness terms: within their operating speed range. Here the rotor will actually deflect to create mode shapes similar to K, is force in X resulting from a displacement in the X direction in Newtons per millimeter; those shownin figure C.2,in addition to any vibration in its bearings. resulting from shaft displacement KT is force inXresultingfrom a displacement in the Y direction in Newtons per millimeter; C.3 Bearings iG, is force in Y resulting from a displacement in the Y direction in Newtons per millimeter; In gearrotordynamics,bearingoilfilmstiffness varies 'Wh speed as wellas torque load applied to X;. is forcein Y resulting from a displacement in the drive. In most turbomachinery, bearing load is theX direction in Newtons per millimeter.



25 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



Bearin shell



9



vibrations. Lightly loaded fluid film bearingscan get into sub-synchronous vibration problems, particularly in thequalificationtestingprocess,whichis generally a no load test. Oil whirl and oil whip are the is names for this type of problem. This vibration usually at a frequency of around 0.4 times rotational speed. If not properly detected in the analysis of the drive,undesirable orevendestructivevibrations may be exhibited in testing or lightly loaded field running.



X



Figure C.3



- Cross coupled bearing schematic representation



Damping terms:



Dnr is forceinX resulting from a velocitytheX in



D, D, L),.x



direction in Newtons per millimeter; is force inX resulting from a velocity in the Y direction in Newtons per millimeter; is forcein Y resulting from a velocity in Y the direction in Newtons per millimeter; is forcein Y resulting from a velocity in Y the direction in Newtons per millimeter.



I



e, :F ,I



'



;;I=-



source



1



:y:



'



'Hz €5



Figure C.4



- Heat balance model



C.5 Critical speed



A critical speed is defined as the speed at which the Obtaining these coefficients is the first step to an peak response amplitude actually will occur when accurate gear drive rotor dynamics analysis. the rotorbearingsystem is in resonancewitha Sophisticated bearing analysis techniques are periodic forcing frequency. There are many possible available to determine these coefficients.A typical forcing frequencies in a gear drive system but the method will solve the Reynolds and energy equais the harmonic tions over a grid network of the bearing area for theone most likely to excite the system forcegeneratedatrotorrotationalspeeddueto particular geometry in question by finite difference mass imbalance. Gears generally are designed to techniques.Theresultsfromeachgridpointare have their actual criiical speeds above 120 percent numerically combinedto produce the performance of their maximum operating speed. Undamped and characteristics of the complete bearing.A detailed dampednaturalfrequenciesmaybecalculated its heatbalance of thebearingsystemunder belowrunningspeed.Dampingmaycompletely operating conditions must be performed to ensure suppress the response of these modes or signifithat the actual oil film viscosities are being utilized. will cantly shiftthe frequency at which these modes This is normally accomplished in an iterative type actually experience peak response or critical speed technique, where an assumed temperature is choby the above definition. Damping tends to lower sen for performance calculation and then is comcalculated natural frequencies. For simple systems pared with the final calculated temperatures they are related by: resulting from the heat balance. If the two do not agree, a new assumed temperature is chosen and ...(C.1) the process continuesin the program until convergence occurs (see figureC.4). where C.4 Stability



5



isthe dampingratio; wd is the damped natural frequency; Astability analysis is required to ensure that the drive willnotexhibitselfsustainingnon-synchronous W, is the undamped natural frequency. 26 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



STD.AGMA bOLL-HSB-ENGL



L778



m



Ob87575 0 0 0 5 8 b l l 5 9 T



ANSIIAGMA 6011-H98



AMERICAN NATlONAL STANDARD



,



..(C.2)



where D istheactualdamping; Dc is thecriticaldamping. Damping, however, tends to raise the frequency at whichtheactualresponseamplitude or critical speed due to imbalance occurs. For simple systems they are related by:



Wr = wo



-4



1



m



...(C.3)



where



C.6.1 Undamped critical speed analysis



is an excellent The undamped critical speed analysis of arotor simpletoolforpreliminaryevaluation bearingsystem.Itallowstheanalysttoidentify of oilfilmstiffness approximatelythemagnitude required to obtain the desired regime of operation of thesystem (Le., rigid orflexiblerotordesign). Approximate mode shapes are obtained. Effectiveness of bearing damping can be seen. If motion of the rotor occurs at the bearing, damping will be very effective.



W,



is the actual response frequency. The damped, undamped, and response frequencies is small. will agreeonlywhenthedampingratio Large discrepancieswill be seen at damping ratios larger than0.3. Another wayof expressing damping ratio is by a logariithmic decrement which defines how quickly a vibration will decay with time.



If themotionoccursotherthanatthebearing, dampingwillbeineffective.Whiletheundamped criticalspeedmapisausefultoolinestimating performance, it is lacking in several major areas. First, it does not consider the cross coÚpled effects in and second, it does not consider the direct the oil film I or crosscoupleddampingterms.Ingeardrives L o g decrement S = ...( C.4) which generally have large damping values as well as large cross coupled terms, the result can to tend C.6 Analysis types yield critical speed predictions less than what an There are threemaintoolsused in (naturalfreactual machine may exhibit. Lastly, no indication of quency and criitical speed) analysis, each having its stability characteristics is obtained. The map should ownstrengthsandweaknesses.Theyare the display the effect of load variations. Stiffness values undampedcriiicalspeedanalysis, the damped of applied load are generally plotted on for the range critical speed stability type analysis and the damped unbalance response analysis. Mode 3 Mode 2



1 Mode 1



Figure C S - Undamped critical speed map



27 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



AMERICAN NATIONAL STANDARD



ANWAGMA 6011-H98



smalldampinginthesystemforthismode.See figure C.7.



C.6.2 Damped critical speed analysis



The damping approach is similar to the undamped map except that it is evaluated using full bearing spring and damping characteristics, including cross coupling terms. Damping in gear bearings is significant and the first two mode shapes generally show significant movement in the bearings,therebyutilizingthe available damping (see figure C.6). This tends to give a closer the to real world result when evaluated, consideringthatfrequencieswithdampingratios greater than 0.2-0.3 will not be responsive where Figure C.6 Bearing damping indicated. It gives results which agree very closely The degree of damping or likelihood of responseis with the damped response analysis for the flexural shown via logarithmic decrement or damping ratio mode of vibration whichis generally the realcridical values. This stability type analysis can also identify will occur. This is because of speed where response sub-synchronousvibrationpotentialsuchashalf little movement at the bearings and corresponding



-



E = 0.011



E= 0.01 I



I



I



I



I



I



I



I



I



I



I



I



I



I



5 = 0.011 I



I



I



I



I



I



I



I



I



5 = 0.012 I



1



I



I



I



1



I



I



I



NF3 Bend



NF1 Rock NF1



Bounce



2000 Figure C.7



3000



4000



5000 6000 7000 Shaft rotating speed (rpm)



8000



9000



- Damped critical speed mapnatural frequency versus rotational speedload



28 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



S T D - A G M A bOLL-HSB-ENGL



L778



Ob87575 00058bb 3b2 ANWAGMA W11 -H98



AMERICAN NATIONALSTANDARD



frequency of thepeakinvibrationresponseto frequencywhirl,whichcanoccurwithunloaded excitation. gears. Here a growth factor is calculated for each mode. Ifthe factor has a negative value, the system isinherentlystable.Ifthevalueispositive,the Thedampedresponseanalysisincludesallthe system may be unstable. This analysis should also effects from both damping and cross coupling. It will if applicable. The be performed over the load range not indicate stability problems. isIt generally best to damped naturalfrequencyanalysisyieldsmore specify unbalance forces several times larger than information but can be difficult to interpret if one is not the actual rotor balance specification allows in the familiar with evaluating the effect of the damping must be analysis. Unbalanced force stations ratio. selected to excite the particular mode of vibration in question.Theunbalanceshouldbeappliedat C.6.3 Damped response analysis several places along the rotor in successive runsto The damped response analysis is generally consid- ensurethateachmode will beexcited.Coupling of the tools for evaluating end, midspan, and blind end locations should be run ered to be the most useful will usually rotor synchronous vibration. It gives excellent as a minimum. Coupling end unbalance correlationwithactualmachines. By definition, a excite the most common mode seen (see figures C.8, C.9 and C.10). critical speedis the speed which corresponds to the



1



Max AMP 99 prn at 20 800 rpm



90 h



E 5 al



70



U



.-3



c ,



E. 50



E



Q



a



fi 30 9000 18 O00 Speed (rpm) Figure C.8



27 O00



- Unbalance modeled at coupling



110



90 h



E



5



Y



al U 3 CI



70 50



E



(II



Q



5 30



I"



9000 18 O00 Speed (rprn)



O



Figure C.9



27 O00



- Midspan 29



COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



S T D - A G M A bOLL-HSB-ENGL



L778



ANWAGMA 6011-H98



Ub87575 U U U 5 8 b 7 2 T 7 W STANDARD NATIONAL AMERICAN



Speed (rpm) Figure C.10



- Blind end



drive.Thisisbecause the criticalsareusually In high speed gear drives with large LID bearings, it designed to be at operatingspeedshigher than the is generally accepted thatNFI(bounce mode) and N F (rock ~ mode) are heavily damped and unrespon- rest of the drive may be ableto withstand. Bearing temperatureorcentrifugalstressconsiderations sive. When heavily damped (damping ratio greater usuallylimitthemaximumoperatingspeed.The than 0.3),these bearing modes may fall within the only thing that can usually be verified is that the 20% band width around the rotating speed-natural actual criiical is above design speeds, but not the frequencyline.Theacceptabilitymaybeproven either by response analysis or by the damping ratio actual critical speed frequency. This is determined by not measuring any peak in response over the of actual damping/criical damping. speed range of the machine. A term called the amplification factor determines when a response peak is to be treated a Evaluating the undamped and damped naturd as well as the response analycriticalspeedor if thefrequencytendstobecriticallyfrequencies dampedm factors less than 2*5 are sis is the most complete m y to determine if gear a considered to be critically damped. drive rotorwill have dynamics problems.If only one It is not the normal case to be able to evaluate the tool can be available, the mast reliable overall results will be obtained with the damped response analysis. accuracy of a critical speed calculation for a gear



-3 .-



5ï



---



ó !



n >



----------



Ac1



0,7277_-FK



I



I



------



-*-



4



I



-



Ncl N1 N2 N2- N I AF



= Rotor first critical center frequency, cycles per minute = Initial (lesser) speed at 0.707 x peak amplitude (critical) = Final (greater) speed at 0.707 x peak amplitude(critical) = Peak width at the 'half power point = Amplification factor



-



" "



*I



-



- " I 1



1.



1



I



.



1



N1 M N 2



Figure C.ll



30 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



RRE AC1



Nd



N 2 - N1 = Resonance response envelope = Amplitude at



- Amplification factor



Annex D (informative)



Systems considerations for high speed gear drives Fheforeword, footnotes and annexes, if any, are provided for informational purposes only and should not be construed a as part of ANSIIAGMA 6011-H98,Specification for High Speed Helical Gear Units.] D.l Purpose -



gear, coupling, equipment, driven other any component.



or



The need for high mechanical reliability in geared Theincreasingdemandsforsystem"mechanical drives can best be satisfied by a "systems approach" reliability" can best be satisfied by a coordinated to the entire train of machinery including foundatechnical exchangebetweendesigner,equipment tions, lubrication, vibration, the forces moments and associated with piping, couplings, etc. The purpose supplier, erecting engineers, and user. The various system analyses,in at least preliminaryforrn, should of this annex is to point out common problems that precededetailedequipmentpurchasespecificamayoccur in gearedsystems, an explanationof to be tions. This sequence will permit the design these problems, and the possible effects. based on more nearly correct load and operating It is not the intentof this annex to present detailed conditions. methods of analyzing or solving the problem, nor will Thiscoordinatedeffortcanbeproperlycalled there be any attempt to set design criteria or limits. "system engineering" andis normally Performedby D.2 Responsibility the design agent or his technical representative. Gear mtlufacturers may not have the expertise nor A gear unit is susceptible to a variety of problems the detailedinformationtoadequatelyanalyze when fi becomesapartofarotatingmachinery system overload- This function must be Performed system, the severity of which generally increases by specialists under the responsibility of the systems Ath speed.Eventhoughtheseproblemsare engineer. generally beyond the gear manufacturer's control, they adversely affect system reliability and/or perforThere is no set format for communicating this data. manceandmaycausedamagetothegearunit.TherequiredinformationisthemagnitudeofoverThe party having contractual responsibilEtyforsys-loadand a description of the operational conditions under whichit occurs, suchas when, how long, and temperformanceshouldinvestigateandresolve nature. theseproblems in thedesignstageandthereby avoid the conflicts that may develop between the Gear units and couplings can be adversely affected component manufacturers users. and one by or more system generated problems. Failuresthatresultfromthesesysteminduced It is that the having pa* categorized becan under three main head:---. responsibilityforthesystemanalysisofa ciiical II lys. service gear drive beclearlyidentifiedinthe specifications,contractorpurchaseorder.Because - Thoseresultingfromoverstressingcompoof the substantial cost involved ina system analysis, nent parts, which are grouped under "Overload"; and in some cases the system performance, it - Thoseresultingprimarilyfrom a lubrication should be emphasized that all parties supplying related failure; components to thesystemhave a responsibilltyto - Alignmentrelated,suchasdistortedfoundafurnish correct and accurate dataso that the analysis tions or poor alignment with connected machinwill be meaningful. eV* D.3 Introduction It is not uncommon to find daily process system



operating costs many times theofcost the gear unit. Thisdowntimecostmakes it desirabletoavoid failure of any partin the system-- be it prime mover,



0.4 Overloads



For the purpose of this discussion overloadw i l l be defined as:



"That load whichis in excess of the nominal design point load."



31 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



ANWAGMA 6011-H98



AMERICAN NATIONAL STANDARD



In the case of air handling centrifugal compressors, Overload can be of momentary duration, periodic, the design power is usually based on the normal quasi-steady state,or vibratory in nature. Dependmaximum ambient temperature. Consideration ing on its magnitude and the number of stress cycles should be given to cold weather operation since the accumulated at overload, it can be afatigue or ayield densrty of air varies with the absolute temperature. stress consideration. Compressorshandlingothergasesareusually Overload on a gear drive can result from internal or encounteredinprocesssystemsundergreater external causes. Internal cause of overload -- such controlwheretemperaturevariationsareless. as taum manufacturing (faults of manufacture) are However, other variables may become serious. In usually found by routine inspections before the gear refinery practice, for example, the composition of the driveisput into service. Externalsources of gas can vary widely, andin other process workthe overload result from the operational characteristics inlet pressure may not be a fixed value. of the system into which the gearisdrive placed, and are more complex and difficult to identify. Carry out is an expression used by the pump and



The gear manufacturer has little if any control over the external influences that produce overload. The system engineer who has overall responsibility for performance should include, along with output, unit cost, efficiency, etc., the investigation of overloads as they relate to potential failure, downtime, and system reliability. Thefollowingmaterialisintended to assistthe systemanalystbyhighlightingsubjectsforhis consideration, and to establish better communication between system people and the gear manufacturers. D.4.1 Estimated maximum continuous power



The operational overload characteristics of various driven equipment varies with the type of machine and should be considered on an individual basis. Pump or compressor designers, for example, can predict the power requirements at the design point with fairly good accuracy. However, the maximum continuous power (service power)is a combination of:



compressor industries to indicate performance on a headcurvebeyondthe so-called designpoint. Figure D.l illustrates a typical compressor percentage performance curve. It will be noted at100% speed as the head drops off and the flowis increased, the power increases to a level as high as 115% load. Carryout is an everyday as reality.Itcomesaboutthroughsuchthings improper estimation of system performance during designstages,alteredsystemrequirementsof existingprocesses,gradualdeteriorationofprocesses,systemsemployingmultipleunitswhere shutdown or failure of one increases the requirements on the remaining units, or through leaks or failures. Figure D.2 illustrates a similar percentage performance curve for centrifugal pumps.



Overspeed is just what the name implies, and is obviously limitedto applications with variable speed prime movers. Becausethe power absorption of the driven machine varies approximately with the third - changes in specific gravity or density of the power ofthe speed, overspeed is a large contributor media being pumped; to overload.Referringagain to figure D.l, the - carry out; 110% performance curve indicates that at speed and 100% flow, the power is increased to 125%. Carry - overspeed; out at this speed can increase the power still further, - variations in pressure ratio across a compres- to levels approaching 140%ofservice power. sor dueto abnormal operating conditions.



Changes in specificgravityof the fluid medium handled by a pump, or change in the density of the a compressor,affectthepower gashandledby transmitted in directproportion.Onboilerfeed pumps, for example, this occurrencebecan encountered during startup, upon malfunction of pre-heating equipment, or during boiler cool-down following the failure.



32 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



Normal practice for a turbine driven centrifugal pump is to set the overspeed trips at 115% design speed. Governorsettingsaregenerallyestablished to 105% and permitcontinuousoperationbetween 110% design speed. It should be borne in mind that operators can anddoresetgovernorstoavail of thesystem, themselvesofmaximumoutput regardless of the originalsettings.



S T D . A G M Ab O L L - H 7 8 - E N G L



1778



O b 8 7 5 7 5 0 0 0 5 8 7 0 873



m



GMA STANDARD NATIONAL AMERICAN



6011-H98



140



130



60



70



Figure D.l



80



90 1O0 % Flow



110



120 140 130



- Typical centrifugal compressor performancecurve



140



130



I



8 100



90



60



70 Figure D.2



80



140 90 1301O0120 % Flow



110



- Typical centrifugal pump performance curve 33



COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



ANWAGMA 6011-H98



AMERICAN NATIONAL STANDARD



condition,therecommendedsolution is toshift natural frequencies by changing stiffness or mass An essential phase in the design of a critical service to limit instead of relyingonsystemdamping system of rotating machinery is the analysis of the vibratoryamplitudes.Normally,alinearvibration dynamic (vibratory) response of a system to excitaanalysis is adequate. However, under certain tion forces. conditions nonlinear responses can occur and the possibility of their existence should be recognized. a systemresults in Thedynamicresponseof additionalloadsimposeduponthesystemand Itisalsoadvantageous to performapreliminary relative motion between adjacent elements in the vibration analysis early enough in the design procesystem. The vibratory loads are superimposed uponduretoallowforanychangeswhichmight be the mean running load in the system and, depending required for detuning purposes. upon the dynamic behaviorthe ofsystem, could lead D.4.2.2 Torsional vibration to failure of the system components. In a gear unit thesefailurescouldoccur as toothbreakage or The vibratory load caused by a steady state torsional pitting of thegearelements,shaftbreakageor vibration of a system is due to the interaction of a bearing failure. periodic excitation, and a natural frequency of the system. The magnitudeof the dynamic load caused Dueto the backlash between the geared elements of bythistype of vibrationisdependent on three a gear unit, tooth separation will occur when the factors: the magnitude of the excitation, the amount vibratory torques in the shafts exceed the average of damping in the system, and the proximity of the torque, resultingin tooth separation and subsequent excitation frequency to resonance. Typical sources impacts. Gear tooth loads due to these impacts can for steady state excitation are: be several times the vibratory torque in the gear shafts. - internalcombustionengines; A vibratorytorquewhich is synchronized to the - reciprocatingpumpsandcompressors; rotation of a gear element can form a cyclic wear - pump or compressorimpellers. pattern on the gear. This wear, which varies around A torsional vibrationin a system can also be caused the circumference on the gear element, results in by a transient excitation which is often called a shock tooth spacing errors of the gear causing noise or or impact loading. Transient conditions occur to due even can become a self-generating excitation which sudden changes in load or speed, or the accelerating reinforces the original excitation. or decelerating through system natural frequencies, Vibratory motion of gear unit components can take including the AC. component of synchronous moupclearancescausinginterferenceproblemsbetors during startup. tween gearing elements, or between shafting and This type of disturbance will produce oscillations at bearings or seals. all the natural frequencies of the system. These D.42.1 Vibration analysis oscillations will decay and eventually disappear due to damping. The peak dynamic loads occur during or Any vibration analysis must consider the complete directly after the disturbance and their magnitudes systemincludingprimemover,gearunit,driven are not substantially reduced by the damping in the equipment, couplingsand foundations. The dynamsystem. The effects of thetransient class of vibration ic loads imposed upon a gear unit are the result of the to dynamic behavior of the total system and not that ofcan be most severein the case of gear teeth due theirabilitytoseparate,thusproducingimpact the gear unit alone. The individual components of loadings onthe teeth. the system are usually supplied by different D.4.2 Vibratory overloads



manufacturers.Therefore, the responsibilityfor D.4.2.3 Lateral vibration performing the vibration analysis must rest with the Dynamic loads at a gear mesh can be caused by a designer of the total system orhis designated agent. lateral vibration of a gear element in response to an The vibration analysis must determine all significant excitationsource.Thelateralvibrationofarotor system natural frequencies and evaluate the system system should consider all flexibilies and restraints response to a l l potential excitation sources. If the which will influence the vibratory response of the analysisindicatesaresonantornearresonant rotor. In the case of a rotor system comprised of a 34 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



S T D - A L M A bOLL-HSB-ENGL



1998



AMERICAN NATIONAL STANDARD



m



Ob87575 0 0 0 5 8 7 2 bbb ANSI/AGMA 6011-H98



- establishes confidence that the rotating system will perform satisfactorily or indicate areas where corrective actions are required prior to a system failure; - provide a basis for evaluation of systems that The most common sources of lateral excitation in a may be designed or manufactured in the future; rotorsystemareunbalanceandmisalignment. - pinpoint system excitations or non-linear reTherefore, care should be given to minimize these sponseswhichwerenotconsideredinany factors in the design, manufacture and installation of theoretical evaluation. arotatingsystem.Thelateralresponseofthe In the design stages it is advantageous to provide system should be evaluated based on the design design features in the system which would facilitate tolerances for system unbalance and misalignment. testing, such as ground surfaces and proper access Considerationmustbegiventooperationinthe points for pickups or strain gages. Also in the system proximityoflateralnaturalfrequenciesbecause design, if it is feasible, consideration should be given large vibratory loads may result with relativelylow excitation. Fluid film bearings are generally used to tofieldmodificationsthatcouldbemadewitha if damaging minimumofoperationaldowntime supportrotors in criticalservicesystems.These vibratory loads were encountered. An example of bearings possess stiffness and damping properties this would be providing both access to couplings and which vary with speed and load. These non-linear additional space for coupling changes for detuning properties should be considered when calculating purposes. the lateral natural frequencies of the system. Under certain conditions of operation, these bearings can D.4.3 Alignment causeinstabilities in therotormotionwhichwill D.4.3.1 Drive train alignment impart dynamic loadson thegear mesh. A gear unit by the nature of its operation is always D.4.2.4 Axial vibration connectedto at leasttwo other pieces of equipment. The successful operation of the gear unit is largely Dynamicloadsonagearmesh are sometimes dependent on the alignment of these components. caused by what appears to be an axial vibration. There are three distinct types of misalignment which This axial motion is most often the response of the must be considered between connecting component gearelementto the unbalancedthrustforces. shafting. Common sources forthese forces are matfunction- Paralleloffsetmisalignment -- when two ing ormisalignedcouplings,electricarmatures shafts are not coaxial, but their axes are parallel; mounted off their magnetic center, face runout of thrust collarsor compressor wheels, and assembly - Angular misalignment -- when two shafts are errors. not coaxial, and their axes are not parallel; Axial misalignment -- when the ends of the D.4.2.5 Vibrationmeasurementsanddesign two shafts are not positioned to provide the reconsiderations quired shaft separation under operating conditions. The resultsof any theoretical vibration analysis are only as accurate as the mathematical model which isMisalignmentduringoperationnotonlycauses developedto perform the calculations. The correct- vibration, but superimposes bending stress on the ness of the model of the system is dependent on the shearstressduetotransmittedtorque.These accuracyto which the inertia, stiffness, damping and stressescannotbereadilycalculatedbutthey excitation can be ascertained. Since there is always warrant discussionso the designer can take precauthepossibility of the actualsystemresponding tions to minimize their effect. Perfect alignment is differently than the theoretical evaluation, consider- almostimpossibletoobtain;therefore,flexible ation should be given to physically measuring the couplings are used to minimize the effects of the vibratory loads in the system at the time of initial inherent misalignment. startup. However, "flexible" couplings, whether of the gear Obtaining test data related to operational loading on tooth, spring elements, flexing disc, or elastomeric a system has the following advantages: type, produce forces and moments on their supportgearelementandshaft, thisshouldincludethe influence of bearings, foundations, couplings, connecting adjacent rotors and the mating gear element.



35 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



S T D - A G M A bOLL-HSB-ENGL



L998



Ob87575 0 0 0 5 8 7 3 5 T 2 D AMERICAN NATIONAL STANDARD



ANWAGMA 6011-H98



avoid unequal settling or



ing shafts when operating misaligned. The analyti-



is the best insurance to



cal determination of the magnitude of these forces not fully understood. It canbe andmomentsis generalized that:



twisting from other causes.



- The sense and direction are such that they try to bring the supporting shafts in line;



-



Significantbendingmomentsmaybe posed on supporting shafts;



-



im-



Thatthemagnitude of the forcesandmoments increases with larger angularii across the coupling;



Fabricated steel bedplates make convenient shippingandhandlingframes,butaregenerallydesigned for strength, not rigidity. Theyare frequently designedwithoutconsideration for the various piping and/or oil sump thermal expansion. Out-ofdoor installations on steel bedplates are particularly subject to cyclic bowing caused by the daily "rise and fall" of the sun. When steel bedplates are used, the designer should endeavor to achieve two things:



- Notwithstandingcatalogclaimsforangular - Arrange oil sumps, piping,andweather capacity, flexible couplings should not be looked protection to minimize unsymmetrical thermal exupon as universal joints; they should be given thepansion; best possible alignment. - Thoroughly investigate elastic deformation of The designer,in order to obtain a greater mechanical the bedplate due to piping forces and moments; reliability of a coupled shafting system must make a then design the bedplate to eliminate twisting at the gear supports. comprehensive assessment of the operating alignment. This is a system study and must include all D.4.3.3 System piping elements of the system including bedplates and/or Theforcesandmomentsimposedonpumps, foundations. An accurateevaluationofthermal compressorsandturbines by theirinletanddisgrowth for all components from a valid and common chargepipingaremajorfactors in deflecting this referencelineisrequired.Journaldisplacement equipment and causing operating misalignment.A l l within bearings, though generally smallerin magniefforts should be made to minimize piping effects. tude, should be considered, particularly as it affects Inlet and drain piping to gear units should receive cold or static alignment checks. After determining similar consideration. the probable magnitude of alignment change from D.4.3.4 Installation instructions static and cold to dynamic and hot (including any periodic cyclic changes that may occur), select a The system designer should assemble and integrate coupling arrangement that provides enough length completeandcomprehensiveinstallationinstrucor span between flexible elements to keep angularity tions covering, as a minimum,such thingsas: low, in the regionof 5 minutes or lower. - soleplate,bedplate,machinerypositionand leveling details; A hot alignment checkis recommended at the time the unit is put in service. This should be performed - foundation bolting and grouting details; whenalltemperatureshavestabilized,andthe - coldalignmentdata -- includingmethod of system is transmitting rated power at rated speed. measuring,relativeposition,andsequenceof alignment; D.4.3.2 Foundations - keying,pinningandtorquingdetails as reAnother kind of alignment problem commonly enquired; countered in geared systemsis the misalignment of - pipe support and flange makeup details; pinion and gear axes due to foundation or bedplate - all other relevant details that would otherwise that twistings or deflections.It should be recognized the job sitemechanic. be left to the judgment of gear units require foundations with sufficient rigidity D.5 Additional lubrication considerations to maintain alignment under operating loads. Reinforcedconcretefoundationswithgrouted-in The continued successful operation and long life of a soleplate are generally preferable to fabricated steel gear unit is dependent on the constant supplyof a bedplates in terms of foundation stiffness, mass and lubricating oil of proper quantity, quality, and condidamping characteristics. A concrete foundation of tion. The lubrication system has five functions to adequate section, on good soil or onsufficient piling, perform: 36 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



~~~~~~



S T D - A L M A bULL-HSB-ENGL



L778



m



Ob87575 0 0 0 5 8 7 4 437



reducefriction;



-



transfer heat;



-



minimizewear;



-



transferwearparticles;



-



reducerusting andcorrosion.



The failure of lhe lubrication system toadequately perform any one or more of these functions may result in premature failure of the gear drive. 0.5.1 Type of lubricant



Two basic types of oils are used to lubricate gear drives:



-



petroleurn base;



-



synthetic.



Therecan be awidevariation in thelubricating qualities of oils within each of these general types. Oils are compounded to meet specific requirements for various applications such as gear oils, bearing oils, internal combustion oils, worm gear oils, etc. Therefore, it isimportantthatanoilbeselected meetingtherecommendationssuppliedwiththe gear unit.



~



m



ANSI/AGMA 6011-H98



AMERICAN NAilONAL STANDARD



-



~



frequently changed to avoid accumulative separation of the additives during operation. When exposed to high operating temperatures in excess of 90" C, rapid degradationwill occur.



D.5.2.2 Viscosity and viscosity index Oils refined into lubricants are generally derived from



two types of crude oil, either paraffin baseor naptha base. Paraffin based oils preferred are because they have better natural extreme pressure characteristics and better resistance to "thinning down" at higher operating temperatures. Naptha based oils, on the otherhand,requirespecialadditives in order to possess this benefit. The oil's resistance to "thinning" is measured by the ViscosityIndex.Thehighertheindexvaluethe bettertheresistance to "tinning".Oilswithout additives of the paraffin base type usually have VI values of ninety(90) or above, whereas naptha base oilswillexhibitlowervalues,Oftentimesbetween twenty (20) and thirty(30). D.5.3 Oil film



Gear elements and the supporting bearing system require a continuous supplyproperly of selected and Synthetic oils shouldneverbesubstitutedfor conditioned oil for survival. An oil film of adequate petroleum base oils without the gear manufacturer's thickness must be established between the rolling approval, since these oils not only have different and sliding component surfaces to avoid damaging lubricating qualities,but also may not be compatible wear and scuffing and to provide component cooling. with materials usedin the gear unit. Hydrodynamic and elastohydrodynamic lubrication theories are commonly usedtoday in analyzing film D.5.2 Lubricant selection thickness in bearingsandgearteeth.Theoil The correcttype and viscosity of oil must be supplied viscosity has the greatest effect on the film thickin accordance with the manufacturer's recommenness. Consequently, failure to use a oil of the proper dations. The friction, wear, film strength and viscosity or viscosity index can result in failure to corrosion protection characteristics of differenttypes produceanadequatefilm thickness for the gear of oils can vary widely. Deviation from the recomteeth and bearings. mended oil for the gear drive can result in premature Improperoilfilmthicknessmaycauseseveral wear and/or failure. operational problems. Lack ofoil film or inadequate D.5.2.1 Lubricantquality oil film thickness may cause metallurgical drawing Lubricating oils for high speed gearboxes should bedue to frictional heatof hardened surfaces, destruchigh qualrty, refined, paraffin base petroleum oils. tive wear, scuffing or pitting of the gear teeth and They must not be corrosive and must be free from frictionalmelting,plastic flow or failure of the grit or abrasives. As they are oftentimes subjectto babbitted bearing surfaces. Oil viscosity increases large flow rates and high operating temperatures, frictional power losses and therefore increases the they must have good antifoaming properties. temperature riseandmayproduceheatenergy beyond the control of the cooling system. Oils of a straight mineral type should be used. High quality ,rust and oxidation resistance is desirable. Thelubricationsystemdesignmustsuccessfully Oils with additives which enhance these characterisachieve a balance of the viscosity and the oil film ticsshouldbecarefullyselectedor,ifselected, thickness considerations.



37 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



ANWAGMA 6011-HQ8



AMERICAN NATIONALSTANDARD



box and improperoil distribution to the spray jets and bearings. The oil supply must meet the requirements set forth Whenthe oilsupplytemperatureishigherthan in the manufacturer’s recommendations. specified, the oil will be subject to rapid oxidation D.5.4.1Quantity reducing the life of the oil, and reducing the operating viscosity resulting in an inadequateoil film. This The proper quantity of oil must be supplied to the condition can result in overheating, excessive wear gear drive to ensure adequateoil film formation on and even failure. the rotor elements, and in cases where babbitted bearings are employed, in the bearing journals, to D.5.4.4 Pressurized system components preventmetaltometalcontactoftherespective Thesystemcomponents must beselectedand flow must be elements. In addition, sufficient installedto avoid problems. The following are some maintained to assure adequate cooling. Too small a quantity may cause inadequate distribution resultingsuggestions to avoid problems: - Aeration. Care must be takento avoid excesin potential overheating, whereas too large a quantisive aeration of the oil. Aeration may result in ty may result in excessive churning of the oil which and decrease the volume of oil to pump cavitation may also resultin overheating. come in contact with the elements of the gear D.5.4.2 Pressurized lubrication systems drive; When lubrication systems are self contained, the - Oilreservoir.Thereservoirmustbelarge system should be designed with a flow capacity of aenough to allow time for the air to separate from the oil. Return linesto the oil reservoir should reto minimum of10% greater than that initially required allow for pump wear, slight bearing wear with normalturn below the oil level. This also includes relief valve bypass lines and any other return lines. service, or change in oil viscosity due to temperature These lines should be located as far away from variations and change of viscosity with use. the pumpsuction lineas possible. Baffles properWhere pressurized oil is furnished from a central ly located in the reservoir will ensure the aerated supply,operating,alarmandshutdownpressures return oil does not find its way to the suction line must be in accordance with the gearbox manufactur-until air has had time to escape from the oil; er’s specifications. Pressures lower than that - Drain lines. The locationof the drain from the recommendedmayresult in reduced flow and gear drive is critical, and the manufacturers recoverheating. Pressures too high may cause excesommendationsshouldbefollowed.Drainlines sivechurningandpossiblegearboxflooding,inshould be sued so they run no more than half full creasing power loss and also resulting in of of oil. Theline should slope down at a minimum overheating. (20 mm/m, 2%) and have a minimum number of bends and elbows. It is desirableto have a vent Oil pressure to the gear drive should be measured at located in the drain line near the exit from the gear a point as near to the entry of the unit as possible, drive to insure proper drainage; thus avoiding the inclusion of pressure losses in the - Vents. Vents must be carefully located and of piping between the point of measurement and the ample size to avoid pressure buildup and allow actual gear supply. ready escape of air from the system without the D.5.4.3 Lubricant temperature loss of oil. Vents must be high enough to avoid entry of contaminants fromthe environment into The gear supplierwill normally specify the minimum the oil. Oftentimesit is desirableto place the vent allowable oil temperature for startup. If temperain the drain line near the exitthefrom gear driveto tures lowerthan this are expected, provisions must ensure proper drainage.The oil is filtered prior to be made to heat the oil prior to startup. The gear returningto the gear drive as well.In this manner drive must not be operated for extended periods at direct contamination of the gear drive the fromatthis minimum startup temperature. mosphere outside is avoided; accordance withthe Oil inlet temperature mustinbe - Suctionlines.Theselinesshouldbegenermanufacturer’s specifications. A low supply temper- ously sized to minimize the pressure loss. The ature may result in a change in viscosity causing suction pressure (net positive suction head) must higher than expected temperature rise in the gearnot beless than that recommended by the pump D.5.4 Lubricant supply



38 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



STD.AGMA bOLL-HSB-ENGL



1998



AMERICAN NATIONAL STANDARD



Ob87575 0 0 0 5 8 7 b 2 0 1 ANSIIAGMA6011-H98



particlesintheoilsupplyshouldbeeliminated. Theseforeignmattersactasanabrasiveinthe bearingsandgearteeth,causingabrasivewear. Splash lubricated drive systems may require frequent changes to remove this debris. In pressurized a filter as systems,the oil issuppliedthrough specified by the gearbox manufacturer. These filter systemsshouldbeservicedregularlytoavoid - Flushing. Before the oil is circulated through the gear drive, a bridge section containing arecirculation of contaminants with theoil and to avoid movable screenis fittedbetween the supply point excessive pressure drops through the filters which and the drain. The system must beflushed until may reduce the quantityof oil supplied to the gear there isno significant accumulation ofdirt on the drive. screen. During flushing the piping should be hamThe oil must be maintained in its correct chemical mer rapped to dislodge foreign particles. After flushing is completed, the supply and drain lines condition to properly perform. Foreign matter, dirt of and moisture can change the chemical properties are connectedto the gear drive. the oil. Additives usedin many oils are depleted with D.5.4.5 Lubricant condition use and require replacement. Since many factors Having provided the propertype and gradeof oil, it is influence the useful life ofthe oil, its condition should beanalyzedon aregularbasis 'to ensure its also important the oilbe supplied and maintainedin thepropercondition.Dust,dirt,gritandother properties are within specification. manufacturer. The total suction loss must include the loss in the piping, valves and fittings,in addition to the distance of thelift. If a check valve is used in the suction line of positive displacement pumps,apressurelimitingdeviceshouldbe of reverse installed to protect against the effects rotation of the pump;



39 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



5



Annex E (informative) Illustrative example r h eforeword, footnotes and annexes, if any, are provided for informational purposes only and should be construed not asa part of ANSIIAGMA 6011-H98,Specification for High Speed Helical Gear Units.]



E.l Purpose-



Thisannexprovidesexamplesbasedon the assumptionthatthegearsetpowerrating is the minimum component rating. In practice all component ratings must be calculated to determine the lowest rated component.



calculated per ANWAGMA 2101 -C95 equations. of one(1.0) Withthefactorsthathaveavalue deleted, the equations are:



(see ANWAGMA 21 O1-695,Eq. 275



E.2 Example #i



E.2.1 Operationalparameters The gearset to be rated transmits power from an induction motor rated at 2500 kilowatts and 1480 5000 RPM to a centrifugal compressor operating at RPM. AnnexAindicates that aservice factor of 1.4is appropriate for this service.



(see ANWAGMA 2101LC95 Eq.28) where: b



= 0.23 (see AGMA = 182.76 mm = 0.680 (pinion) = 0.720 (gear)



zl 4 1



E.2.2 Gearset parameters



ZN



The through hardened double helical gearsetto be rated has the following parameters: Number teeth, pinion 53 of 179 gear teeth, Number of normal Module, Pressure angle, normal angleHelix distance Center Outside diameter, pinion 188.75 mm Outside diameter, 623.24 gear mm Normal circular tooth thickness reference diameters, pinion (182.76 mm) and gear (617.24 mm) width Face Overall face (gap included) 350 pinion Hardness gear Hardness speed Pinion grade Material levelqualityGear mm 7.0 depth whole Cutter radiustipCutter



= 5000 rpm = 255 mm



W1



ZE



fi pinion fi gear 4



20" 29" 32' 30" 400 mm



YN



KJ Kif



Cp=&



at



OHP



4.63 mm



300 mm



OFP



H6 300 H6 5000 rpm 2



OFP



Qvl2 1.28 mm



E.2.3 Rating parameters



Thepittingresistance power ratingandbending strengthpowerratingatunityservicefactorare



908-889)



= 190 = 0.57 (see AGMA 908-889)



= 0.59 (seeAGMA908-B89)



= 3.448 (3 + COS 29" 32'30") = 0.800 (pinion) = 0.827 (gear) = 1.15 = 1.26 (seeANWAGMA 21O1 495) = 1.4 (see annex A) = 1080 N/mm2(pinion @ 350HB) = 960 N/mm2 (gear @ 300 HB) (see ANWAGMA21O1 -CS, figure 8 Grade 2) = 359 N/mm2 (pinion) (see ANWAGMA 21O1 -CS, figure 9 Grade 2) = 324N/mm2(gear @ 300 HB) (see ANWAGMA2101-C95 figure 9, Grade 2)



[



5000(255) 0.23 182.76(1080)(0.680) 1.91 x 1071.15(1.26) 190 = 5290 kW (pinion)



P,



=



P ,



=



5000(255)



0.23



1.91 x 1071.15(1.26)



[



182.76(960)(0.720)



190



= 4680 k W (gear)



Previous page is blank 41 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



I'



L978 W Ob87575 0 0 0 5 8 7 8 O B q W



S T D - A G M A bOLL-HSB-ENGL



AMERICAN NATIONAL STANDARD



ANSllAGMA 6011-H98



5000(182.76) 255(3.448)(0.57) 359(0.800) 1.91 x1.15(1.26) 107 1 = 4750 kW (pinion) 5000(182.76) 255(3.448)(0.59) 32q0.827) payu = 1.91 X 107 l.U(l.26) 1 = 4590 kW (gear)



Puyu =



5290 4680 4750 4590 or 3BokW Pa is the lesserof 1.4 ' 1.4 ' 1.4 ' 1.4



(see ANWAGMA 2101 -c95, Eq. 277 Pay.



=



Wldwl



= 8215 rpm = 260 mm



W1



E.2.4 Rating conclusions



b



Pa is equal to the lesser ofP , or Payudivided by the service factor, orPo = 4590 1.4 = 3280 kW. This is greater than the service power of 2500 kW.



ZI 6 1



= 0.19 (see AGMA = 255.9 mm



E.3 Example #2



ZN



= 0.680



+



908-889)



= 190



E.3.1 Operationalparameters



rJ pinion = 0.55 (seeAGMA 908-889)



The gearset to be rated transmits power from a gas turbine rated at 15 MW and 8215 RPM to an electric generator operating at 3600 RPM on a base load cycle. The service factoris 1.3.



YJgear



YN



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



= 1.15



Iyy



= 1.27 (see ANWAGMA2101 -C%) C.F=I& = 1.3 (see annex A)



KY



39 89 6



= 1550 N/mm2



(?HP



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



20" 23"45' 420 mm 268.8 mm 595.1 mm o. 1 0.0 260 mm 80 mm 58 HRC 2 2.4 mm 14 mm 0.25



E.3.3 Rating parameters



The pittingresistancepowerratingandbending strength power rating at unityservicefactorare calculated per ANWAGMA 2101-C95 equations. With thefactorsthathaveavalue of one(1.0) deleted, the equationsare:



42 COPYRIGHT American Gear Manufacturers Association, Inc. Licensed by Information Handling Services



=OS73 (seeAGMA908-889)



= 6.555 (6 + cos 23" 45') = 0.800 (pinion) = 0.803 (gear)



4



E.3.2 Gearset parameters



Number of teeth, pinion Number of teeth, gear Module, normal Pressure angle, normal Helix angle Center distance Outside diameter, pinion Outside diameter, gear (q), pinion Profile shift coefficient Profile shift coefficient (xd,gear Face width Gap Hardness pinion and gear Material grade Cutter tip radius Cutter depth Cutter protuberance



OFPYA'



"



1.91 X 107 K C X ~ 1 (see ANWAGMA 21 O1-C95 Eq. 28) where:



= 450 N/mm2(pinionandgear) (see ANWAGMA 2101-C%, table 4



GFP



Grade 2)



P ,



=



[



8215(260) 0.19 , 255.9(1550)(0.680) 190 1.91 x 107 1.15(137)



= 29315 kW



8215(255.9) 260(6.555)(055) 450(0.800) 1.15(1.27) 1 = 25 430 kW (pinion)



Payu = 1.91 x 107 P ,



8215(255.9) 260(6.555)(0.57) 450(0.803) 1.91 x 107 1.15(1.27) 1 = 26 450 kW (gear)



=



29315 25 430 26 450 or Pa is the lesserof 1.3 ' 1.3 ' 1.3 19 560kW.



E.3.4 Rating conclusions The allowable transmitted power, Pa = 19 560 kW, is greater thanthe service power of 15 MW.



Annex F (informative) Efficiency



rheforeword,footnotesand annexes,if any, areprovidedforinformationalpurposes onlyand shouldnot be construedas a part of ANSIIAGMA 6011-H98, Specification for High Speed Helical Gear Units.] F.l Gear unit efficiency



F.2 Calculation methods



F.2.1 Mesh losses .Mostcontractsforhighspeedhelicalgearunits requiresomeguaranteeofminimumoperational Mesh power loss(PIM), for 17'/2" or 20" NPAof basic efficiency. When high power is transmitted, a very rack, can be estimated as below: can represent substansmall increment of efficiency tial economic gain or loss over the life of the gear PM = (22 - 0.8 a,) 0.01 P [I'4+z:z] ...(F.1) a unit.To realize optimumgearunitefficiency, where detailed study of the several sources of power loss is required. a,, is normal pressure angle of basic rack; z1 is number of teeth in the pinion; Sources of power loss for high speed helical gear unitsinclude:mesh,internalwindage,radialand 22 isnumber of teethingear; . thrust bearing friction and shaft driven accessory P istransmitted power, kw. power requirements. F.2.2 Windage losses F.l.l Meshlosses Windage and churning loss can be evaluatedthe by Mesh losses result from oil shearing and frictional losses which are dependent on the specific sliding velocity and friction coefficient. Most gear meshes under this standardwill operate in the EHD lubrication regime. F.1.2 Internal windage losses



following equation:



P, =



d'2n2bcos3 ß'm, 1.42



X



A



...(F.2)



where



PW is windage power loss per gear, k W d' isoperatingpitchdiameterofgear,mm;



n is gearspeed,rpm; Because of the sensitivity to gear to unit specific b istotalfacewidth,mm; relationships - housing-to-rotor clearances, pitch line velocity, gear blank proportions and design, oil ß' is operatinghelixangle; viscosity, method of mesh lubrication and cooling, m,, isnormalmodule; horizontal or vertical offset and internal baffling -this A is arrangement constant (use 1O00 to 4000, component of gear box losses is very difficult to based on arrangement). accurately estimate without experimental data from F.2.3 Bearing losses a specific gearbox. Hydrodynamic sleeve bearing loss in kW, P&, can F.1.3 Bearing losses be estimated bythe following equation: Hydrodynamic journal bearing losses are generated d 3 ~1.723 j X 10-17 ...(F.3) Bearing losses may be through oil shearing. pBh= p nt C calculatedbyamodifiedPetroffequationorby The thrust bearing power loss in kW, P B ~is: , complex computer modeling methods. (