AGMA 10FTM12-2010 Flank Load Carrying Capacity and Power Loss Reduction by Minimized Lubrication《通过最小润滑减少侧面承载能力和功率损耗》.pdf

《AGMA 10FTM12-2010 Flank Load Carrying Capacity and Power Loss Reduction by Minimized Lubrication《通过最小润滑减少侧面承载能力和功率损耗》.pdf》由会员分享，可在线阅读，更多相关《AGMA 10FTM12-2010 Flank Load Carrying Capacity and Power Loss Reduction by Minimized Lubrication《通过最小润滑减少侧面承载能力和功率损耗》.pdf（15页珍藏版）》请在麦多课文档分享上搜索。

1、10FTM12AGMA Technical PaperFlank Load CarryingCapacity and PowerLoss Reduction byMinimized LubricationBy B.-R. Hhn, K. Michaelis, andH.-P. Otto, Gear Research Centre(FZG) of the TechnischeUniversitt Mnchen (TUM)Flank Load Carrying Capacity and Power Loss Reduction byMinimized LubricationB.- R. Hhn,

2、K. Michaelis, and H.- P. Otto, Gear Research Centre (FZG) of theTechnische Universitt Mnchen (TUM)The 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.AbstractThe lubrication of

3、 gears has two major functions: Reducing friction and wear as well as dissipating heat. Thepower losses, especially the no-load losses, decrease with decreasing immersion depth using diplubrication. The load-dependent gear power losses are nearly unaffected by minimized lubrication. However,the gear

4、 bulk temperatures rise dramatically by using minimized lubrication due to a lack of heat dissipation.With minimized lubrication the scuffing load carrying capacity decreased by up to more than 60% compared torich lubrication conditions. The dominating influence of the bulk temperature is therefore

5、very clear. Starvedlubrication leads to more frequent metal - to contact and the generation of high local flash temperatures mustbe considered. An additional factor for the scuffing load carrying capacity calculation in case of minimizedlubrication conditions is proposed.Concerning pitting damage te

6、st runs showed that by lowering the oil level the load cycles without pittingdamage decreased by approximately 50% up to 75% for minimized lubrication compared to the results withrich lubrication conditions. The allowable contact stress is clearly reduced (up to 30%) by minimizedlubrication. A reduc

7、ed oil film thickness as a consequence of increased bulk temperatures results in morefrequent metal-to-metal contacts causing a higher surface shear stress. In combination with a decreasedmaterial strength due to a possible tempering effect at high bulk temperatures the failure risk of pitting damag

8、eis clearly increased. The common pitting load carrying capacity calculation algorithms according to DIN/ISOare only valid for moderate oil temperatures and rich lubrication conditions. For increased thermal conditions,the reduction of the pitting endurance level at increased gear bulk temperatures

9、can be approximated with themethod of Knauer (FZG TU Mnchen, 1988). An advanced calculation algorithm for pitting load carryingcapacity calculation at high gear bulk temperatures (valid for high oil temperatures as well as for minimizedlubrication) is proposed.The micropitting risk was increased by

10、low oil levels, especially at high loads and during the endurance test.The micropitting damage is caused by poor lubrication conditions which are characterized by a too lowrelative oil film thickness due to high bulk temperatures. Again, the actual bulk temperatures are of majorsignificance for calc

11、ulation of the micropitting load carrying capacity.The wear rate of the gears is almost unaffected by the oil level. Only a slight increase of wear could beobserved with minimized lubrication. This increase can be explained by the higher bulk temperature of thegears running under minimized lubricati

12、on conditions. The investigations showed that there exists a naturallimitation for lowering the oil quantity in transmissions without detrimental influence on the load carryingcapacity. Knowing these limitations enables the user to determine the possible potential benefits of reducedoil lubrication.

13、 The correct prediction of the actual gear bulk temperatures is of major importance in thiscontext. A method for the estimation of the gear bulk temperature at reduced immersion depth respectivelypoor lubrication conditions is proposed.Copyright 2010American Gear Manufacturers Association1001 N. Fai

14、rfax Street, Suite 500Alexandria, Virginia, 22314October 2010ISBN: 978-1-55589-987-53Flank Load Carrying Capacity and Power Loss Reductionby minimized LubricationB.-R. Hhn; K. Michaelis; and H.-P. Otto, Gear Research Centre (FZG)of the Technische Universitt Mnchen (TUM)IntroductionGears are machine

15、components which determinethe capability and reliability of many technicalproducts. Continuous demand for higher efficiencyand reliability, increased load carrying capacity andendurance life, smaller size, lower noise andvibrations, prolonged service intervals, low environ-mental impact and low cost

16、s will remain the maindriving forces in the development of gear drives inthe future. The fatigue of contacting surfaces ingears is often the life-limiting factor in transmis-sions which should, ideally, operate reliably fortwenty years or more.Low oil levels and thus reduced oil quantities aresometi

17、mes used for the reduction of no-load lossesin dip lubricated automotive and industrial transmis-sions. Specifically, a higher efficiency can beobtained by reducing no-load power losses such assqueezing, splashing and ventilation losses. Theselosses can be reduced by lowering the oil volume,namely t

18、he oil level in dip lubricated transmissionsand the oil flow rate with oil injection lubrication. Inthese cases the oil amount required for lubricationmay be sufficient, but there may be a lack of coolingoil. This leads to high gear bulk temperatures result-ing in thin separating films with higher f

19、riction andwear on the mating surfaces and therefore anincreased risk of gear failures such as scuffing,pitting, micropitting and low speed wear.Therefore the objective of this study was toinvestigate the limits concerning possible reductionof lubricant quantity in gears which could be toler-ated wi

20、thout detrimental effects on their loadcarrying capacity.Test apparatusTest gearsTable 1 shows the main geometrical values of thetwo gear types used in this investigation.For all investigations on the flank load carryingcapacity under minimized in comparison to rich lub-rication conditions standard

21、test gears (Table 1)were used. For the investigations on pitting andwear gear type C-PT (“PT” stands for pitting) wasused. For evaluation of the micropitting load carry-ing capacity gear type C-GF (“GF” stands forGrauflecken the German word for micropitting)according to 4 was used. The gear type C h

22、as aclose to practical design with a well balanced slidingspeed distribution along the path of contact. For theconducted scuffing tests gear type A according to3 was used. Gear Type A has an uneven slidingspeed distribution along the path of contact in orderto the increase the scuffing risk especial

23、ly towardsthe tip of the pinion. The ground flanks of the testgears had a mean roughness of Ra= 0.2 mm0.4 mm (gear type C-PT and A) respectivelyRa = 0.4mm0.6mm (gear type C-GF). The higherroughness of gear type C-GF increases the risk forthe gear failure mode micropitting. The meansurface roughness

24、is the only difference betweenC-PT and C-GF.Table 1. Main geometrical data of test gearsParameter Symbol Unit C ACenter distance a mm 91.5 91.5Number of teeth pinion z1- - 16 16Number of teeth gear z2- - 24 24Normal module mnmm 4.5 4.5Normal pressure angle n 20 20Face width b mm 14 20Profile shift f

25、actor pinion x1- - 0.182 0.864Profile shift factor gear x2- - 0.172 -0.500Material - - 16MnCr5 20MnCr54Type of lubricationFor the investigations of power loss and bulktemperature for dip lubrication seven different im-mersion depths at standstill were chosen(Figure 1).For the main scuffing, pitting,

26、 micropitting and weartests an immersion depth of one and three times themodule of the gear was chosen.An oil/air lubrication device (Figure 2) was used tolubricate the gears with an oil-air mixture and thebearings with plain oil.It is a lubrication system in which small measuredquantities of oil fr

27、om the reservoir are introducedinto an air/oil mixing device which is connected to alubricant supply line to the gear set. The air velocitytransports the oil along the interior walls of thelubricant line to the point of application. Thecontinuous air stream to the gear set is only thetransporting me

28、dium for very small oil quantities only some milliliters per hour.Oil/air lubrication is used for tests under severestarved lubrication conditions where the gear meshis lubricated in order to decrease friction and wear,but no heat dissipation by cooling oil is available dueto the very low oil quanti

29、ties.LubricantsFor the tests at different immersion depths andcorresponding oil quantities the FVA reference oilsFVA 3 pure, FVA2A and FVA3A (with a high EP-additive content in order to prevent scuffing duringthe efficiency and bulk temperature experiments),M220 (a mixture of FVA3A and FVA4A) and th

30、eCEC reference oil RL 133, were used (Table 2).Figure 1. Oil levels at standstill for testing for dip lubricationFigure 2. Oil/air lubrication device5Test rigFor all durability tests (pitting, micropitting andwear) and the scuffing tests, a standard FZG back-to-back test rig 5 (Figure 3) with a pitc

31、h linevelocity of 8.3 m/s was used. For the wear tests at0.05 m/s a speed reducer was installed between themotor and drive gear. For the test program at a pitchline velocity of 30 m/s one of the FZG test rigs wasequipped with a variable speed three-phaseasynchronous motor and a speed increaser.The m

32、ain elements of the test rig are the twoconnected gear boxes and an engine behind thedrive gear. The test gear box has an easilyremovable front and top plate in order to enable aquick change of the test gears. The two spur gearunits (test and drive gear set) are connected by twoparallel shafts.In or

33、der to put a specific load on the gear flanks oneof the shafts can be divided into two separate partsconnected with the load clutch. With the load leverand different weights a defined torsion can be gen-erated. Table 3 shows for the different load stagesLS in the FZG scuffing, pitting and micropitti

34、ng testthe corresponding pinion torque T1and the Hertzianstress at the pitch point pC.Table 2. Properties of used lubricantsName RL 133FVA2 +4% A99FVA 3 +6.5% A99FVA 3pureM220 +4% A99Abbreviation RL133 FVA2A FVA3A FVA3 M220AISO VG 100 32 100 100 200Kinematic viscosity at 40C(mm2/s)40101.4 29.8 95 95

35、0 210Kinematic viscosity at 100C(mm2/s)10013.45 5.2 10.7 10.7 19Density (at 15C) (kg/m3) 15892 871 872 872 895Figure 3. FZG back-to-back test rig6Table 3. Load stage LS, pinion torque T1, and Hertzian stress at pitch point pcLoadstage,LS, (-)FZG scuffing test FZG pitting test FZG micropitting testTo

36、rque,T1,(Nm)Hertzian stressat pitch point,pc,(N/mm2)Torque,T1,(Nm)Hertzian stressat pitch point,pc,(N/mm2)Torque,T1,(Nm)Hertzian stressat pitch point,pc,(N/mm2)1 3.3 146 3.3 172 3.3 1722 13.7 295 13.7 346 13.7 3463 35.3 474 35.3 565 28.8 5104 60.8 621 60.8 741 46.6 6495 94.1 773 94.1 922 70.0 7956 1

37、35.3 927 135.3 1105 98.9 9457 183.4 1080 183.4 1287 132.5 10948 239.3 1232 239.3 1470 171.6 12459 302.0 1386 302.0 1651 215.6 139510 372.6 1538 372.6 1834 265.1 154711 450.1 1691 450.1 2016 319.3 169812 534.5 1841 534.5 2197 378.3 1848The great advantage of the closed power loop is thatthe engine on

38、ly has to provide the power loss of thetwo gear boxes.Therefore the back-to-back test rig can also beused for efficiency tests by installing a torque andspeed measuring device on the input shaft.One tooth of the pinion and one tooth of the gear canbe additionally equipped with a bulk temperaturesens

39、or in order to measure the actual gear bulktemperatures for different operational conditions.Power loss and bulk temperature withminimized lubricationLoad-dependent and no-load lossesUsing either dip lubrication with high to very low oillevels or oil/air lubrication at very low oil quantitiesshowed

40、no significant change in the load dependentlosses (Figure 4) and is in quite good accordance tothe theoretical values which were calculatedaccording to Ohlendorf 10.Figure 4. Load dependent torque losses with minimized dip and oil/air lubrication 127The friction regime is obviously not changed byusi

41、ng minimized lubrication. Even small oil quantit-ies (less than 30 ml per hour) are able to reducefriction in the gear mesh quite the same as withabundant lubrication (oil level up to centerline of thegears).Tests with oil/air lubrication compared to tests withoil moistened gears showed the importan

42、ce of acontinuous lubrication otherwise an immediatescuffing damage is the fatal consequence.The no-load losses, which are dominated by thegear splashing and squeezing losses, decreasedsignificantly with decreasing immersion depth(Figure 5).Using oil/air lubrication with oil quantities below 100ml/h

43、 results in very low no-load losses compared todip lubrication with high, medium and very low im-mersion depths. No splashing and almost nosqueezing effect can be observed in this case.Gear bulk temperaturesThe bulk temperature of gears has a stronginfluence on the load carrying lubricant film.Incre

44、asing temperatures result in decreasing oilviscosity which produces thinner oil films underconstant load and speed conditions.According to Oster 11 the bulk temperature Mofthe gears can be calculated based on the idea thatthe generated heat in the tooth contact is dissipatedby the tooth surface area

45、 to the surrounding oilwhich works as a cooling fluid with the followingempirical equation 1:M= L+ 7400 PVZPab0.72XS1.2 XCA(1)whereMgear bulk temperature, C;Lno-load oil temperature, C;PVZPload dependent gear losses, W;a center distance, mm;b tooth width, mm;XSlubrication coefficient;XCaprofile modi

46、fication coefficient.The factor XSissetto1.0inthecaseofdiplubrication and to 1.2 in the case of oil injection lub-rication. The equation for the bulk temperaturedoes not differentiate between pinion and gear nordoes it take account of different immersion depthsfor dip lubrication or the rotational d

47、irection.(Figure 6)Figure 5. No-load torque losses with minimized dip and oil/air lubrication 128Figure 6. Pinion bulk temperature with minimized dip and oil/air lubrication 12From the experimental data a simple equation 2 canbe derived which allows calculation of a mean XS-factor for different imme

48、rsion depths and rotationaldirections for gear type A and C in the FZG test rig:0.3 XS= 0.35 edaD 3.7 (2)wheree immersion depth of gear, mm;datip diameter of gear, mm;D parameter for rotational direction.With D = 0.75 for the standard (meaning longdistance between oil sump and gear mesh) rotation-al

49、 direction and D = 0.5 for the reversed rotationaldirection, meaning direct transportation of coolingoil to the gear mesh.For oil/air lubrication a constant XS= 3.7 can beused which indicates starved lubrication without anyheat dissipation by cooling oil but only by convectionand conduction to the surrounding metalcomponents.Several negative consequences arise from highbulk temperatures:- decreasing oil film thickness;- shift from mixed lubrication conditions toboundary lubrication