AGMA 91FTM13-1991 Separation of Lubrication and Cooling in Oil-Jet Lubricated Gears《油喷式润滑齿轮的润滑和冷却的分离》.pdf

《AGMA 91FTM13-1991 Separation of Lubrication and Cooling in Oil-Jet Lubricated Gears《油喷式润滑齿轮的润滑和冷却的分离》.pdf》由会员分享，可在线阅读，更多相关《AGMA 91FTM13-1991 Separation of Lubrication and Cooling in Oil-Jet Lubricated Gears《油喷式润滑齿轮的润滑和冷却的分离》.pdf（19页珍藏版）》请在麦多课文档分享上搜索。

1、91 FTM 13wSeparation of Lubrication and Cooling inOil-Jet Lubricated Gearsby: J. Greiner, Cincinnati Gear andK. Langenbeck, University of Stuttgart,!iiiAmerican Gear Manufacturers AssociationTECHNICAL PAPERASeparation of Lubrication and Cooling in Oil-Jet Lubricated GearsJ. Greiner, Cincinnati Gear

2、and K. Langenbeck, University of StuttgartTheStatements andopinionscontainedherein are thoseofthe authorandshould notbe construed asan official actionoropinion of the American Gear Manufacturers Association.ABSTRACT:Results _om the tests on a high-speed back-to-back stand (vt= 70 m/see = 13725 ft/mi

3、n) show the influence of theseparation of lubrication and coolingoil supply on the geartemperatures (scuffing load capacity) and efficiency. Thegear mesh is only given a minimum oil flow rate necessaryfor lubrication of the working tooth flanks. Cooling isprovided by spraying oil onto the inner surf

4、aces of the rims. This leads to a reduction of the total oil flow rate of up to60% compared to the currently recommended flow rates. In spite of the reduced total oil flow rate the temperature levelof the gears can be kept low and the efficiency slightly increases due to reducedhydraulic losses in t

5、he gear mesh. -.-.Copyright 1991American Gear Manufacturers Association1500 King Street, Suite 201Alexandria, Virgini_ 22314October, 1991ISBN: 1-55589-610-3wlSeparation of Lubrication and Coolingin Oil-Jet Lubricated GearsJ. GreinerCincinnati Gear CompanyK. LangenbeckUniversity of Stuttgart, Germany

6、1 Introduction Oil jet lubrication is usually used for gearswith pitch line velocities of v_ 30 m/sec (= 6000Theoretical analyses and experimental tests /2,3/ .ft/mzn). The oll Jet lubrlcat_on variants common inwith oil-jet lubricated gears have shown that the research and industry basically vary in

7、 position ofactual delivered oil flow rate is considerably the nozzles and their spray direction into the gearhigher than necessary to maintain a load carrying, mesh. As shown in Fig. 1 they can be classified aselasto-hydrodynamic oil film. Excessive oil flow follows:rates cause a reduction in effic

8、iency, especially inA high-speed gears with a pitch line velocity of vt = Variant A: Oil-jet is directed towards the pitch200 m/sec (= 40000 ft/min) the load-independent point perpendicular to the center line.losses (“hydraulic losses“ due to lubricant, see Variant B: Oil-jet is directed towards the

9、 gear/14/ rise considerably. Too low oil flow rates on center.the other hand cause an unacceptable temperature Variant C: Oil-jet is directed towards the gearrise of the working tooth flanks, since the fric- mesh respectively towards the face pa-tional heat produced in the gear mesh is insuf- rallel

10、 to the axis of the gears.ficiently carried away bythe oil. Variant D: Centrifugal lubrication;The considerations of reducing the oil flow rate Fling-off cooling.and keeping constant gear mesh temperatures lead toa separation of the oil supply into a lubricating For gears for industrial usage basica

11、lly variantsand a cooling oil supply. In doing so, the mesh is A1, A2, AI+A2 as well as BI, B2, BI+B2 are used. Inonly given a minimum oil flow rate necessary for literature /5/ you generally find the statement thatlubrication of the working tooth flanks. The gears as to the lubrication “Spraying in

12、to the engagingare cooled outside the mesh area by oil that is mesh“ (AI,EI) and as to cooling “Spraying into thesprayed onto the inner surfaces of the gear rims on disengaging mesh“ (A2,E2) should be preferred.both sides of the gears. However, several experimental investigations withThe goals of th

13、e investigations in /i/ consisted pitch line velocities up to vt = I00 m/sec (= 20000of determining the influences of the separation of ft/min) have shown, that the gear bulk temperatureslubricating and cooling oil supply on the pinion and are considerably higher (lower scuffing load capa-wheel temp

14、eratures, on the power losses, and on the city) when spraying into the disengaging mesh. Onwear of a single-stage helical gearset as a function the other hand temperature measurements at veloci-of load and pitch line velocity. It was to be shown ties in the range of I00 to 200 m/sec (= 20000 tohow a

15、nd to what extent the total oil supply can be 40000 ft/min) support the statement that A2 is withreduced, regard to cooling more appropriate than A1.At pitch line velocities of more than I00 m/sec(= 20000 ft/min) frequently both variants are2 Background combined (AI+A2,BI+B2). The oil flow sprayed i

16、ntothe engaging mesh has to be kept as low as possibleAn overview of the actual situation concerning because of hydraulic losses, so that only the lubri-“Separation of lubricating and cooling oil supply of cation is guaranteed, while much more oil is sprayedhigh duty gears“ is given by a literature

17、research into the disengaging mesh for cooling. It is a tom-/4/ as well as the explanations in /I/. It was mon fear that at these velocities the oil sprayedproved that up to now no experimental investigations into the disengaging mesh will be immediatelyhave been done as to this subject and that the

18、 rejected by the approaching tooth flanks. The oiltheoretical knowledge, too, is insufficient, film thickness on the tooth flanks will be reduceddue to the extremely high centrifugal forces tillthe next tooth engagement occurs, which can lead toinsufficient lubrication conditions. It is a generalopi

19、nion that the cooling effect is much better whenProf.or.-I_. K. La_enbeck isHead of the “Institutf_rNasch_nen- oil is sprayed into the disengaging mesh, becausek_strukti_und 6etriebehau“attheUniversity_ Stuttgart,Germny. there is a higher temperature difference between theDr.-l_.J. Greinerwas Sci_ti

20、ficEmployeeat this Istitute a_ is “hot“ flanks coming out of the mesh and the oil thatpres_tly_rki_ as StaffEngi_er at the CincinnatiGear Company, is sprayed in.Cincinnati,_io.1The extremely fast running but relatively lowAi B oi _ C cl loaded flanks in aviation gears can (according toC _ _ /9/) eve

21、n be operated without “direct“ lubrication -of the gear mesh, because the oil mist caused by theI_ i_ horizontal spraying of oil on web and rim is suffi- _cient for lubrication. This, however has to be que-stioned for very wide gears.c2 With a separate cooling oil supply the question_ _ _ _ of the m

22、inimum lubricating oil flow necessary for anoperational lubrication becomes very important. In_i_ _ industry the determination of the lubricating oil flow rate is done by taking into consideration the/ expected power losses and the heat balance of thegear box /8,9,12/, the face width /8,10,11/ and t

23、heEHD-Theory /2,13/. To be able to compare the dif-ferent recommendations as to the amount of theA_C ,1.c2 B*C B_-c_ CI_C2 lubricating oil flow rate, the operating conditions_ _i_ _c_ of a single-helical gearset were set up in Table 1.c Table 2 shows the calculated results.- The reco-_endati_arding

24、the heat balance ofthe gear box show considerably higher oil flowrates. This is mainly due to the very small per-mitted temperature difference between oil inlet and_W_=_ i_ oil outlet. A doubling of the allowed temperature_ I difference would cause a halving of the oil flow/_ rate and thus less hydr

25、aulic losses._ _ cI When calculating the lubricating oil flow rate byC regarding the face width one is on the safe sidebecause of experiences made with slowly runninggears.The according to /2,13/ calculated minimum oilflow rates cannot be considered as recommendations,A2+C2_2.c2 since the frictional

26、 heat in the gear mesh can nolonger be dissipated. The idea of cooling the gearsAt*A2 A1 SlY-B2 B_ D by spraying oil on the faces requires the deter- _ “i_“ ruination of the optimum of sufficient lubricationwhile minimizing the hydraulic losses and dissi-_ pation of the frictional heat. 3 Experiment

27、al Equipment_ 3.1 Back-to-Back Test Stand, Test GearsFig. 1: Injection Variants A,B,C,D with and without Fig. 3 schematically shows the back-to-back test ._separated cooling oil supply stand with the experimental arrangement for thetemperature and efficiency investigations. Fig. 4For separate supply

28、 of lubricating and cooling shows the operable test stand in the gear lab.oil, combinations of injection variants At, A2 and Basically the test stand corresponds to a designC2 as well as Bl, B2 and C2 would be well suited which has been reported several times in literature(see Fig. i). For lubricati

29、on the oil is sprayed /1,2/. Table 3 shows the technical data of the twoeither directly into the mesh or radially onto the different test gearsets. Profile B is designed forgears. Cooling is provided from oil sprayed onto the balanced sliding at tooth root and tip, while theinner surfaces of the gea

30、r rims. sliding velocity at the tip of the pinion of profileDudley/Richardz /9/ mention in their final con- E is approximately 16 % higher than at the wheel.clusions that spraying of cooling oil onto the faces since the tests did not show significant differ-of the gears is a procedure especially use

31、d for avi- ences in the following only the results with profileation gears with thin rims and webs (see Fig. 2). B will be discussed.3.2 Positions of NozzlesFig. 5 shows the view of the test gear box withthe pipework for the separate supply of lubricatingand cooling oil without gear cover. At the en

32、d ofthe pipe of the lubricating oil supply of the gearGearBody_ mesh a spray-head with 3 solid stream nozzles isUghl,.wtlg_lgn mounted. It distributes the oil equally along thewhole width of the gear mesh. The axes of thenozzles in the spray-head are directed towards thepitch point perpendicular to

33、the center line.Two oil distributors outside the housing qua-_ _ j rantee an equal distribution of the cooling oil ontothe inner surfaces of the gear rims. The axes ofthese solid stream nozzles point on the filletradius between the web and the rim. The two in-jection variants A1+C2 and A2+C2 are rea

34、lized bycorresponding choice of the rotation direction.pi_ w,h 3.3 Measuring F_uipmentNm_iIn the back-to-back test stand two torque-metersare mounted for measuring the transmitted power andthe power losses (see Fig. 3). The temperatures ofFig. 2: Lubrication of aviation gears acc. to /9/ the gears a

35、nd the oil temperatures in the housingTransmitted Power p kW i000Permitted temperature differencebetween oil inlet and oil outlet A_ K 15Outside surface of housing A m2 0.5Number of teeth on pinion/wheel zl/z 2 - 31/47Face width b _ 40Pitch line velocity vt m/sec 60Helix angle _ deg. i0Tooth flank s

36、urface roughness R z _m 2.5Characteristic Values of Mineral Oil:Specific heat capacity c = 2000 Wsec_(kgK)Density p 850 kg/m _Table 1: Operating conditions of single-helical gearsetReference ; Source Equation; Recomm. CommentsReference Value Standard Value for V.in i/_in- Niemann /12/Gear Mesh Losse

37、s P_ VZ = 40,8.(P_/A_) 10.2 - P_ = 3.76 kW ; CZ for full_._ Temperature Difference wlth compensation of P_ by the ollbetween Oil Inlet and P_=P.(O.1/Zl/COS_)+O.O3/(vt+2)Oil Outlet _Niemann ; Winter /8/ oI. Face Width b Vzmin = 0.080.1 i/(min.mm 3.2-4.0 - Estimation2. Power Losses PV CZ = (P_ - PKG )

38、/(cpA_) 17.4 - at complete dissipation of P_Heat Dissipation PKG by the oildue to Convection PKG = 0.8.1.2 kW/m 2 = 0.6 kWP_ = 0.8 % of P = 8 kWDudley ; Richardz /9/_Z = 1.27 i/min per I00 kW of P 12.7 - generous supply A_ up to I0 CTransmitted Power P _Z = 0.64 i/min per I00 kW of P 6.4 - normal su

39、pply n_ up to 24 CYz = 0.42 i/min per I00 kW of P 4.2 - lean supply =_ up to 38 CV z = 0.32 i/min per I00 kW of P 3.2 - starvation supply _ 38 CShell /10/I. Transmitted Power P - Diagram - 23.0 - friction coeff, g_7 = 0.06 ;allow, temp. rise“-A_= 8 C2. Face Width b YZ = 0.55 . 0.8 i/(min.cm) 2.2-3.2

40、 - for high-speed gearsVZ = 1.6 i/(min.cm) 6.4Mobil /Ii/Face Width b CZ = 0.5 . 1 i/(min.cm) 2.0-4.0Ohlendorf /13/Tooth Flank Roughness Vzmin = 2“bvt(_1+R_ )/I04 1.44 - _I + R_ = 2“(0.6“Rz) = 3 gmSchober /2/EHD-Theory VZm_, = 2“bvthi/104 0.09 - hi 0.18 gmTable 2: Minimum oil flow rate for sufficient

41、 lubricationJ3J Gear Temperatures Test Gears SlaveGears TvPZ PKR! _ t“ PwerLss D.C. Motor III Revolu_i0n VK _ cov.ncer I Control BoxD Torquerh SpeedI t Control Panel, _ TernperatuTeHeaterOiJunitSuppyIt _ f Device JD = Manometric S_ 1 “_F_.z = TharmometmrV,._ = Valve _Cooling WaterFig. 3: Schematic v

42、iew of test arrangement for temperature and power loss measurementsFig. 4: Back-to-Back Test Stand in the Gear LabFig. 5: Positions of nozzles at separation of lu-brication and cooling oil supply (Injectionvariants AI,AI+C2,A2,A2+C2)Profile B ECenter distance a mm 180Face width b mm 40Helix angle _

43、deg. i0Normal module mn mm 4.5Number of teeth zI - 31z2 - 47Addendum modification x I - 0.190 0.413coefficient- x2 - 0.223 0.000Normal pressure angle =n deg. 20Transv. pressure angle =_ deg. 21.775Outside diameter dal mm 152.2 154.2da2 mm 225.6 223.6Length of action of recess g_1 mm 10.580 12.589g_a

44、2 mm 10.547 8.285Sliding velocity at the vpl m/sec 0.246.v t 0.293-v ttipI Vg_ m/seo 0.245.v t 0.193-v tI Quality - - 5 acc. to DIN 3962_ Surface Roughness - - Ground_=2.5_Material and - - 16 Mn Cr 5Heat Treatment Carburized and HardenedTable 3= Test gear data - Profile B and Eare measured by thermo

45、couples of Fe-CuNi. The 4 Test Program, Test Conditionstransmission of the tooth center, rim, web and hubtemperature data of the gears is done by mercury In /i/ a lot of tests have been conducted as toslip ring transmitters with 8 channels each. The conventional injection variants (AI,A2,BI,B2,Cl,C2

46、)transmitter mercury-cells work without sliding as well as to the variants with separate cooling oilcontacts. Therefore they eliminate the various dis- supply (AI+C2,A2+C2). Here only the results of testsadvantages of the usual contact transmitters such as with separation of lubrication and cooling

47、will betransitional and isolatory resistances as well as provided.disturbant thermovoltages. Table 4 shows in detail the load and injectionTemperature-measuring positions in the gears (in conditions of the tests. The tests were done undertransverse plane): practical conditions. The torque was determ

48、inedaccording to the findings and results of SchoberPinion (Ri) / Wheel (Ra) /2/. In 4 load-stages the torque of the pinion hasbeen increased in step_ from TI = 1100 Nm (Hertzian in the tooth centers (_;0_a Pressure aHc = 989 N/ram _) to up _o 1700 Nm (aHc = 1202 in the rim and webs (_SR_;_Ra;_SiRa)

49、 N/ram2). The pitch line velocity was mostly increased in the hubs (_HRi;_HRa) in steps of 10 m/sec from v_ = 20 m/see to vt = 70m/see. A number of these _tests was done at ain the middle of the gears at half face width (see constant pitch line velocity of vt = 60 m/sec (=Fig. 6). 12000 ft/min).The thermocouples are embedded into the holes At a maximum pitch line velocity of v = 70 m/secdesignated for the temperature measurements with a (= 14000 ft/min) the power range limit (_ 1664 kW)thermoconducting two component metallic loctite, of t