1、16FTM14 AGMA Technical Paper Impact of Surface Condition and Lubricant on Effective Gear Tooth Friction Coefficient By Aaron C. Isaacson, Matthew E. Wagner Ra and Rz were then computed by analyzing 0.76-mm long cross-sections of the data. This results in an effective cutoff wavelength of 0.76 mm. Th
2、e results of the surface characterizations are summarized in Table 3 and are considered to be consistent with what is normally obtained in industry. Surface roughness (Ra) on the tooth flank ranged from 0.06 to 0.29 m. Table 3: Surface Roughness Data of Test Gears Notes: 1. Surface roughness measure
3、ments are reported in micrometers. 2. Measurements reported are directional, taken orthogonal to grinding direction. 3. Averages are computed based upon two interferometry measurements taken at the half height on four teeth for each gear (eight total interferometry measurements per gear). Four cross
4、 sections of each interferometry measurement are analyzed to generate average Ra and Rz values. BEARING AND LIP SEAL LOSSES In order to establish a value of the bearing losses (TB) and lip seal losses (TL), the four-square loop was operated at all three test speeds with one gear removed from the tes
5、t box and oil preheated to 46 C. In this configuration, the only torque present at transducers T1 and T2 was due to frictional losses in the test box bearings and seals. Seal losses are independent of load and are therefore accounted for in these no-load tests. The engineering group of SKF Bearings
6、provided results of a proprietary computer simulation to account for load-dependent increases in bearing losses present during actual testing. The bearing 6 16FTM14 model included frictional effects from rolling, sliding, and drag. Bearings used in the test box were all rated to operate at the speed
7、s tested in this program, however data was not available regarding the expected bearing model accuracy. The aerodynamic windage loss (Tw) due to spinning gears was not computed for several reasons. Several windage loss models were examined 12, 13, 14, all of which were validated on gears with substa
8、ntially larger face widths and pitch diameters than those used in this effort. Also, discussions with the Computational Mechanics group in the authors organization indicated that past unpublished research to model this gear geometry showed windage was not significant due to the narrow face width and
9、 small pitch diameter of these gears. Lastly, while the no-load bearing and seal losses were being measured, one gear was installed in the test box. Therefore, the windage effect from one gear is included in the no-load TB + TL measurements, so the only unknown windage component is from one remainin
10、g test gear. Since the windage loss due to one test gear is believed to be insignificant, it was decided to neglect this value rather than include it, using models which had not been validated for a gear of this size. EXPERIMENTAL RESULTS Utilizing the experimental arrangement detailed in Figure 1,
11、measurements of losses in the test gear box were conducted under the test conditions outlined in Table 1. Results of all tests can be found in Tables 46 in chronological order for each test group. Where possible, tests were repeated to establish the accuracy of the test method. An example of typical
12、 data collected is illustrated in Figure 3, which shows data obtained for Mobil SHC 626, with ground gears at 5,000 rpm and 96 N-m torque. All tests were run for an initial break-in period until the total torque loss reached a steady state value. After steady state conditions were reached, an averag
13、e value for total torque loss was taken and recorded as TF1. The computed bearing and seal losses (TB + TL) were then subtracted from the total torque loss TF1 to yield the gear mesh loss TF. An average measured value of the applied loop torque T1 was also taken over the same time period. This avera
14、ge T1 value was then used with TF to compute eff. Thermocouple data was used to record an actual value of the test box oil temperature. An average temperature value was recorded over the same time period used to record TF1 and T1. A nominal test temperature of 46 C was used for all tests, and the re
15、corded temperature during the measured period ranged from 43.6 C to 51.1 C. Oil was delivered to the mesh at 345 kPa using three jets; two 1.27-mm jets on the out-of-mesh side aimed into the contact surfaces and one 0.46-mm diameter jet aimed straight into the mesh. The configuration is shown in Fig
16、ure 2. Test gear geometry, speed, load, and post-test roughness measurements were used with ANSI/AGMA 2001-C95 Annex A to compute the flash temperature for each test condition 11. The flash temperature and recorded test box oil temperature for each test were then used to compute an initial estimate
17、of bulk temperature, per ANSI/AGMA 2001-C95 Annex A. Using data available from a prior testing program that utilized thermocouple instrumented test gears, a correction factor was established to relate the AGMA bulk temperature estimate to actual recorded bulk temperature measurements. It should be n
18、oted that the prior test program used for thermocouple data was conducted with gears of identical geometry and on the same test rig as this program. The corrected bulk temperature for each test is shown in Tables 46. Bulk temperature was then used with test gear geometry, speed, load, and post-test
19、roughness measurements to compute the specific film thickness () according to the Mobil EHL Guide Book 9. was computed at points along the active profile and averaged to produce the values in Tables 46. A plot of vs. eff is shown in Figure 4. Figure 5 shows the average eff values for all speeds, tor
20、ques, and repetitions for each gear type and oil. The standard error for each test group is also shown. It should be noted that AGMA 925-A03 15 supersedes AGMA 2001-C95 Annex A. Bulk temperatures were computed following the method detailed in AGMA 925-A03 and resulting values differed by a maximum o
21、f 0.02. These results and methods are considered identical for the purposes of this research. 7 16FTM14 Table 4. Experimental Results for As-Ground Tests Table 5. Experimental Results for ISF + W-DLC Tests 8 16FTM14 Table 6. Experimental Results for ISF Tests Figure 3: Torque Data Plot for Test 1a 9
22、 16FTM14 Figure 4: Effective Friction Coefficient vs. Specific Film Thickness () Figure 5: Average Values of Effective Friction Coefficient (eff) 10 16FTM14 DISCUSSIONS AND CONCLUSIONS The following observations are made when results from Tables 46 are averaged over all loads and speeds: The average
23、 percentage improvement in effective coefficient of friction of W-DLC and ISF / ISF vs. ground / ground gears when operating in Mobil Jet Oil II is 54%. The average percentage improvement in effective coefficient of friction of ISF / ISF vs. ground / ground gears when operating in Mobil Jet Oil II i
24、s 47%. The average percentage improvement in effective coefficient of friction of W-DLC and ISF / ISF vs. ground / ground gears when operating in Mobil SHC 626 is 44%. The average percentage improvement in effective coefficient of friction of ISF / ISF vs. ground / ground gears when operating in Mob
25、il SHC 626 is 33%. The average percentage improvement in effective coefficient of friction of ground / ground gears in Mobil SHC 626 vs. identical gears running in Mobil Jet Oil II is 7%. The average percentage improvement in effective coefficient of friction of W-DLC and ISF / ISF gears in Mobil Je
26、t Oil II vs. identical gears running in Mobil SHC 626 is 10%. The average percentage improvement in effective coefficient of friction of ISF / ISF gears in Mobil Jet Oil II vs. identical gears running in Mobil SHC 626 is 15%. There are three issues that warrant further discussion. First, the magnitu
27、de of the estimated values of the effective coefficient of friction is lower than typical values for dynamic coefficient of friction reported elsewhere. Second, for ground gears, the data suggests that the effective coefficient of friction depends on torque in the four-square loop. Third, the ground
28、 gears on average have lower effective coefficient of friction values when operating in Mobil SHC 626 vs. Mobil Jet Oil II. It must be emphasized that the parameter, termed “effective coefficient of friction”, that is being computed in this work is not the fundamental coefficient of friction between
29、 the sliding surfaces of the teeth. However, it is a parameter that can be utilized by a gear box designer to estimate the friction losses in a gear box and the amount of heat that will need to be absorbed in the operation of that gear box. Further, while there is significant confidence in the actua
30、l measured losses in the test gear box, the load-dependent increases in bearing losses present during actual testing are computed and are therefore estimates. The dependency of the effective coefficient of friction on torque appears in the case of the ground gears operating in both oils. This is con
31、tradictory to friction theory, where the coefficient of friction is considered independent of the load. Two gear teeth in mesh are a significant departure from two flat surfaces sliding on each other; effects from tooth bending, load sharing, and tip relief are likely contributors to this discrepanc
32、y. All of the gears utilized in this effort were manufactured to the same specification, however the ground gears were manufactured in a different lot than the superfinished and coated gears. It was found that the ground gears had tip relief starting at 22 degrees of roll angle, while the superfinis
33、hed and coated gears had tip relief starting at 26 degrees of roll angle. Diab, et al. 6 discuss the influence of tip relief on power loss and indicate that tip relief may be a non-negligible parameter for optimizing gear efficiency. It is likely that if the ground gears had tip relief identical to
34、the superfinished and coated gears, the losses at higher torques would be greater than those measured. The ground gears on average have lower effective coefficient of friction values when operating in Mobil SHC 626 vs. Mobil Jet Oil II. It is likely that the efficiency improvement due to the lower v
35、iscosity oil is offset by the increased asperity interaction of the ground surfaces. This effect has been reported elsewhere in the literature for ground gears 4. The superfinished and coated tests both showed efficiency improvement when operating in the lower viscosity oil. This effort clearly prov
36、ides a gear box designer with a parameter, “effective coefficient of friction”, that will be useful in computing frictional losses in a gear box. This parameter appears to behave as anticipated with the variables investigated in this study. Further investigation into the effect of gear geometry and
37、lubricant temperature would be logical next steps to continue this work. 11 16FTM14 ACKNOWLEDGEMENT The first three authors would like to thank REM Surface Engineering and the Aerospace Bloc of the Gear Research Institute for supporting this experimental effort. Special thanks are also extended to J
38、oe Bitner for his exceptional effort during test rig assembly and testing. REFERENCES 1 Rao, S. et al, “The Impact of Surface Condition and Lubricant on Gear Tooth Friction”, AGMA FTM 2014. 2 Yoshizaki, M., et al., “Study of Frictional Loss of Spur Gears”, STLE Tribology Transactions, Vol. 34, 1991,
39、 1, pp.138146. 3 Britton, R. D., et al., “Effect of Surface Finish on Gear Tooth Friction”, ASME Transactions Journal of Tribology, Vol. 122, January 2000, pp. 354360. 4 Petry-Johnson, T. T. et al., “Experimental Investigations of Spur Gear Efficiency”, Proceedings of ASME, paper no. DETC2007-35045.
40、 5 Martins, R., et al., “Friction Coefficient in FZG Gears Lubricated with Industrial Grade Oils: Biodegradable Ester vs. Mineral Oil”, Tribology Transactions, Vol. 39, 2006, pp. 512521. 6 Diab, Y., et al., “Prediction of Power Losses due to Tooth Friction in Gears”, Tribology Transaction, Vol. 49,
41、2006, pp. 260270. 7 Mihailidis, A., et al., “Prediction of Friction Coefficient of Spur Gears” VDI-Berichte, NR. 1665, pp. 705719. 8 Xu, et al., “Prediction of Mechanical Efficiency of Parallel-Axes Gear Pairs”, ASME Transactions Journal of Mechanical Design, Vol. 129, 2007, pp. 5868. 9 Mobil EHL Gu
42、idebook, 4th edition, Mobil Oil Corporation Technical Publications, TBK0092007. 10 Rao, S. B. and McPherson, D. R., “Gear Tooth Temperature Measurements”, Gear Solutions, August 2009, pp. 3441. 11 ANSI/AGMA Standard 2001-C95, Fundamental Rating Factors and Calculation Methods for Involute Spur and H
43、elical Gear Teeth. 12 Dawson, P. H., “Windage Loss in Larger High-Speed Gears”, Proceedings for the Institution of Mechanical Engineers, Vol 198A, No 1, 1984, pp. 5159. 13 Diab, Y., et al., “Windage Losses in High Speed Gears Preliminary Experimental and Theoretical Results”, ASME Design Engineering
44、 Technical Conferences and Computers and Information in Engineering Conference, Chicago, IL, USA, Sept. 26, 2003, DETC2003/PTG-48115. 14 Diab, Y., et al., “Investigations on Power Losses in High-Speed Gears”, Proceedings for the Institution of Mechanical Engineers: Journal of Engineering Tribology, Vol 220, 2006, pp. 191198. 15 ANSI/AGMA Standard 925-A03, Effect of Llubrication on Gear Surface Distress.
