AGMA 12FTM19-2012 A Field Case Study of Whining Gear Noise in Diesel Engines.pdf

1、12FTM19AGMA Technical PaperA Field Case Study of“Whining” Gear Noisein Diesel EnginesBy Y. Kotlyar, G.A. Acosta, andS. Mleczko, Navistar, Inc. andM. GuerraA Field Case Study of “Whining” Gear Noise in DieselEnginesYefim Kotlyar, German Alonso Acosta, and Sam Mleczko, Navistar, Inc. andMarshall Guerr

2、aThe statements and opinions contained herein are those of the author and should not be construed as anofficial action or opinion of the American Gear Manufacturers Association.AbstractThispaperdescribestheworkperformedtoreduce1stordergearmeshvibrationwhichwastherootcauseofan excessive gear “whine”.

3、 It includes a discussion for introduction of an unmodified involute profile (exceptfor a short tip relief to smooth the tooth entry and exit into the mesh), inspection charts, significance of astatistical sample size for validation, vibration measurements, and correlation of tooth profiles to vibra

4、tionresults. While many articles on gear noise reduction focus on introducing profile modifications, this paperdescribes one helical gear application where an improvement was achieved by the elimination of profilemodification.Copyright 2012American Gear Manufacturers Association1001 N. Fairfax Stree

5、t, Suite 500Alexandria, Virginia 22314October 2012ISBN: 978-1-61481-050-63 12FTM19A Field Case Study of “Whining” Gear Noise in Diesel EnginesYefim Kotlyar, German Alonso Acosta, and Sam Mleczko, Navistar, Inc.and Marshall GuerraIntroductionNoise Vibration Harshness, NVH, is a broad term that encomp

6、asses the study and development of noise,vibration, and sound quality. Within engines and vehicles, the gear train is one of many sources of excitationthataffectNVH. Whilesignificanteffortisfrequentlyfocusedoninsulating(orprotecting)anobserverfromthenoise,engineersare alsoworking toreduce thelevel o

7、fstructure-borne vibrationat thesource whichcausesthe noise.As major sources of noise within diesel engines and vehicles have successfully been addressed, engineersare paying more attention to the gear train that is increasingly becoming a more “visible” source of noise.Thiscasestudyonlyfocusesonthe

8、reductionofvibratoryenergy,notasoundqualityimprovementwhichcanbe very subjective.Specifically,thispaperdiscussestoothprofilegeometryeffectson firstorder gearmesh vibration,eliminationof profile modifications to reduce vibration energy, and validation of a new design.While many articles on gear noise

9、 reduction concentrate on introducing profile modifications, this paperdescribesonegearapplicationwhereanimprovementwasachievedbythe eliminationof profilemodificationexcept for a short tip relief.The problem statementThe vehicle OEM utilizing a particular engine model within our product line was inv

10、estigating 5-10% of theirproductsthatexhibiteda“gearwhine”noisewhichwassubjectivelydeemedunacceptable. Basedonvehiclenoise data acquisition, the source of the gear noise clearly originated as the first gear mesh order within theengines geartrain consisting of three (14.23 Normal Diametral Pitch, 20

11、Normal Pressure Angle, and 20.5HelixAngle)gears:Crankshaftgear(A),Camshaftgear(B),andFuelpumpgear(C),schematicisdepictedinFigure 1. However,thecauseofnoisewasquiteelusive. Frequentqualityinspectionswereconducted;but,aclear-cut quality problem could not be attributed as the source of the gear whining

12、 noise.TheinvestigationaswellasthedevelopmentofthedesignofexperimentbecameateameffortinvolvingNVHengineers, design engineers, the gear supplier, gear cutting tool suppliers, as well as engine manufacturingengineers and technicians.Gear A Gear B Gear CFigure 1. Gear train schematic4 12FTM19Review of

13、WOW and BOB enginesThe vehicle OEM identified engines that had noise conditions described as worst of the worst, WOW, andbest of the best, BOB under light to moderate throttle conditions. The decision was made to investigate acorrelationbetweentheseenginesvibrationsignaturesandgearinspectiondata. Th

14、eengineswerereturnedto our engine assembly plant for analysis.Utilizing a mobile data acquisition system, engine vibration data focusing on the first gear mesh order (48thengineorderascrankshaftgearhad48teeth)wasacquiredononeWOWandoneBOBengines. OneWOWengineexhibiteda40%higherlevelofgearmeshvibratio

15、n(Figure 2)ascomparedtotheBOB. Thisdegreeofdifferentiation and a high level of repeatability in the vibration data provided confidence that there was acorrelation to tooth geometric parameters.The gears from each gear train were then inspected on an analytical gear inspection machine for all theNVHc

16、ritical characteristics - profile, lead, and index errors.Gear A had profile tip and root reliefs that were close to the tolerance limit (Figure 3) and gear B had tip androot reliefs that were slightly outside the tolerance limit (Figure 4). All inspected gears had lead and pitchcharacteristics that

17、 were within the tolerance limits.Figure 2. Crankcase vibration measurement at 1500 rpmFigure 3. Analytical inspection of gear A, WOW 15 12FTM19Figure 4. Analytical inspection of gear B, WOW 1The legacy gear drawings specified a K-chart as the profile tolerance, but they did not specify an ideallyde

18、sired trace (target) within that K-chart, see Figure 5. The K-chart did not explicitly require tip and root re-liefs. However, the nature of a K-chart tolerance with a narrow band in the middle and awider bandat thetipand root implies that tip and root reliefs are not only allowed, but desirable.K-c

19、harttolerancecancreateaconditionforasignificantpart-to-partvariationintheprofiledeviationfromthepure involute. On one extreme end of the K-chart tolerance is a profile absent of tip and rootreliefs whiletheother extreme is a tooth that has a significant tip and root reliefs. Both of these conditions

20、 are consideredwithin tolerance. Another challenge with this K-chart tolerance specification was that it hindered aquantitat-ive comparison of parts and a measurement of the process capability.Figure 5. Tooth profile tolerance for gear A6 12FTM19GearsfromanadditionalWOWenginewerealsoinspected,seeFig

21、ure 6,Figure 7. andFigure 8. Profilesofthesegearswerewithinthetolerancelimit. Yet,thegearshadrootreliefsthatwereclosetothetolerancelimit.Inspection of the BOB geartrain showed a somewhat lesser deviation of the profile form from the pure invol-ute. Gear A, Figure 9, had a profile slope error that wa

22、s consistent through the entire tooth. Gear B,Figure 10, had a part of the profile that was straight. Lastly, Gear C, Figure 11, had a profile closer to a pureinvolute. All three gears had profile and lead errors that were within the K-chart specifications. Indexcharacteristics were also within the

23、tolerance limits.While most inspected gears had profiles that were within spec, there were dramatic variations in tip and rootreliefs - both in form and amount. There was further confirmation of excessive profile variability whilereviewing random production inspection records.Figure 6. Analytical in

24、spection of gear A, WOW 2Figure 7. Analytical inspection of gear B, WOW 2Figure 8. Analytical inspection of gear C, WOW 27 12FTM19Figure 9. Analytical inspection of gear A, BOBFigure 10. Analytical inspection of gear B, BOBFigure 11. Analytical inspection of gear C, BOBDesign and validation planThed

25、ecisionwasmadetoa)determineandvalidateatoothprofileformforimprovedfirstordermeshvibrationcharacteristics, b) re-target the tooth profile tolerance specification accordingly, and c) improve the processcapability for a more consistent tooth profile and thus vibration performance.The plan also included

26、 a statistical validation of re-targeted profile. A statistical comparison was necessarybecauseanaccuratetestingandassessmentofhowanindividualgeometriccharacteristicaffectsgearmeshvibrationcanbecomplicatedbysomanydifferentmodifyingvariablesinthesystemincluding:a)alignmentofgearaxesresultingfromastac

27、kupofmanufacturingerrorsofshafts,bearings,andhousings;b)manufactur-ing variations of all mating gears; c) variations in structure response to excitation; d) variation in the bearingsupport and rigidity.8 12FTM19The decision was made to test at least 30 gear samples with a new profile specification t

28、o mitigate otherinfluencing factors and provide more reliable results.Development of a new tooth profile targetInaperfect world,unloaded involutegears withconjugate toothsurfaces wouldtransmit motionat aconstantangular velocity thus zero transmission error. Gear imperfections, which create velocity

29、variation, statictransmission error, and vibration, are universally defined by three major characteristics - profile, lead, andindex. These characteristics overlap in their effects on the static transmission error that in turn will result ingear mesh vibration. For several reasons, however, the plan

30、 was to retarget the tolerance for only one toothcharacteristic, tooth profile.Both sets of BOB and WOW inspected gears had very small index (tooth-to-tooth) errors, typically0.002-0.004mm. Incontrast,theprofilemodificationvalues(whicharepurposelyinducederrors - deviationsfroma pureinvolute) weremuc

31、h larger,up to0.012mm fortip reliefand 0.024mmfor rootrelief, seeFigure 5.Lead errors (including crown as it can be considered a necessary error) could also adversely affect statictransmission error, especially for helical gears. However, compared with profile modifications, both WOWand BOB gear set

32、s had relatively small lead errors and a small crown with an average measurement of0.009mm.With lead and index errors considered to have minimal effects, the profile characteristic seemed to be thedominant factor responsible for a non-uniform transmission that induced first order gear mesh vibration

33、.A plan was created to test a pure involute profile (no modifications except for a short tip relief/chamfer). Thehypothesis for elimination of profile modification was that, in this engine application, a pure involute profilewould provide a better vibration characteristic as compared to the existing

34、 profile specification that allowedsignificant tip and root reliefs (up to 0.012mm for tip relief and 0.024mm for root relief). There were twoarguments in support of that premise.Load conditionProfile modifications (tip and root reliefs) are purposely induced deviations from a pure involute and typi

35、callydesigned for only one specific tooth loading. However, in diesel engines, the instantaneous forces appliedtogears are continuously oscillating. Torsional effects create conditions that not only do these instantaneousgear forces change their values, they can also reverse direction. At times, eve

36、n tooth separation can occur.Figure 12 and Figure 13 illustrate thecyclicalnatureof torquethat istransmitted throughthe crankshaftgear(driver) during an engine cold test (compression only) and typical engine firing respectively. X-axis repres-entsthecrankshaftrotationalangleand y-axisrepresents torq

37、ue. Thusit waspossible thatduring mostloadconditions, the tip and root reliefs were inducing non-uniform transmission resulting in vibration.Figure 12. Compression only torque in crank angle domain9 12FTM19Figure 13. Engine firing torque in crank angle domainContact ratioOne approach for specifying

38、profile modifications (tip and root reliefs) is based on the highest and lowestpointsofasingletoothcontactarea56. Asthecontactpointmovesalongthepathofaction(Figure 14andFigure 15, ANSI/AGMA 1012-G05), the load applied to a gear tooth is changing depending on how manypairsofteethareengagedatthesameti

39、me. Whenacontactratiois betweenone andtwo (Figure 16,AGMA915-1-A02),thechangeintoothbendingwithinthecontactpathcanbesignificant. Specifically,atoothbend-ingforcecanbedoubledwhentwoteethpairengagementtransitstoasinglepair. Asaresult,toothdeflectionmayvaryconsiderablyduringthetransitionfromadualtoasin

40、glepaircontact. Atoothprofilemodification(tipandrootreliefs)istypicallyintroducedtomitigatetheproblemofthetransitionaltoothdeflection. However,thegearsinthisengine applicationhad ahigh contactratio - transverse contactratio was2.13. That meantthat,atanygiventime,therewereatleasttwopairsofteethincont

41、actinthetransverseplane. Inaddition,theaxialcontactratiowasgreaterthanone(1.12)thatfurtherhelpeddistributetheloadevenly. Inreality,therewereatleastthreepairsofgearteethcarryingtheloadatanygiventime. Thus,fromadurabilityperspective,thelargecontact ratio negated a need for a profile modification.There

42、fore, the plan was to re-target the gear profile for a pure involute form. The only exception was a shorttip relief/chamfer for a smooth entry and exit of the tooth.Path of actionPoint of contactPitch pointFigure 14. Line of action 110 12FTM19Base circleBase circleLine of actionThelineofactionisthep

43、athofactionfor involutegears. It isthe straightlinepassing through the pitch point and tangent to both base circles.Figure 15. Path of action 1tiptiprootrootmaster gearprofile deviation diagramsproduct geardirection of paper feed1 2 3123master geartangential compositedeviationprofilecomponentproduct

44、 gearpbpbgpb= base pitchg= length of path of contactstylusEffect of contact transfer on the profile component in a tangential composite deviation diagram(spur gears)Figure 16. Transverse contact ratio 211 12FTM19Prototype gearsAfter developing a profile target, the team collaborated with both the ge

45、ar supplier and their tool supplier tomanufacture30controlledsetsofgears. Thesamples weremanufactured withno rootrelief anda veryshort(orno)tiprelief. Figure 17,Figure 18,andFigure 19illustratetypicalchartsoftheprototypegearsA,B,andCrespectively. These 30 sets of gears were serialized and 100% inspe

46、cted on an analytical gear checker.Figure 17. Analytical inspection of gear A, 30 piece sampleFigure 18. Analytical inspection of gear B, 30 piece sampleFigure 19. Analytical inspection of gear C, 30 piece sample12 12FTM19Vibration measurement and data analysisA production technique was developed to

47、 acquire gear-related vibration measurements on partially builtengines near the end of the assembly line. The engine is first loaded into the test station on a conveyer andthen accelerometers on spring-loaded arms are robotically placed on the cylinder head of the engine. Thecrankshaft of the engine

48、is thencoupled toa motorto spinthe engine. Nocombustion takesplace duringthistest. Theengineissweptthroughavarietyofspeeds. Whenaspeedsensormeasuresacrankshaftspeedof1500 RPM, the vibration measurement begins.Anorderanalysisalgorithmwasutilizedtodeterminetheamountofgearmeshrelatedvibrationpresentint

49、heaccelerometer signal. Order analysis is a method of deconstructing a time signal measured on a piece ofrotating machinery to determine the amount of energy produced at specific rates relative to rotational speed.Thealgorithmusedontheengineteststationcollectsmultiplerevolutionsworthofdatafromtheacceleromet-er and speed signals. To separate the first gear mesh order (48th engine order) from surrounding orders, aresolution width of 1/16th order was used.The vibration level from multiple crankshaft revolutions was averaged and stored.