ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
Results  Krvtox 1 43ACTM  The results of four tests of Krytox 143ACTM run in each of the five specified environments are presented  in Figure 4. The shortest lifetimes are exhibited by tests run in ultrahigh vacuum (UHV). The longest  lifetimes are observed by running in the presence of water vapor, with increasing lifetime corresponding  to increasing water vapor concentration in the test environment. Only 16 ppm water vapor in an  atmosphere of nitrogen, corresponding to a relative humidity of c0.07%, is sufficient to extend the lifetime  by an order of magnitude relative to testing in vacuum. Pure water vapor itself, without its being present in  nitrogen gas, also leads to lifetimes much longer than testing in vacuum. Testing in one atmosphere of  ‘‘dw nitrogen, in which the hygrometer indicated no additional water vapor above its lowest reading of  Td= -99.9% (0.014 ppm), gives lifetimes greater than the tests in vacuum, but only marginally so. Testing  in 0.21 3 kPa (1.6 Torr) dry oxygen, corresponding to -0.2% trace contamination of oxygen in a nitrogen  atmosphere, also has little effect on lifetime relative to that in vacuum. The results presented here thus  indicate that water vapor in the test environment exercises a “protective” effect on the PFPE lubricant,  permitting a longer life before failure. Finally, the coefficient of friction of 0.134 was independent of the  test environment and thus of the lifetime.  700  600  500 a  u) s  e c .-  400 .- f  E c  J  300 .- - m  z 200  100  0 (ppm H20 in N2) + (16) 0  (Total pressure  H20, milliTorr)  (4) 0  (14) 0  (5) (.W  i (.3)  8  0 - - - 8 -  UHV 1.6 Torr O2 1 atm ”dry” N2. 1 atm N2 wlH20 Pure H20  Figure 4. Normalized lifetimes of Krytox 143ACTM when testing in different environments  Pennzane@ 2001 A  Friction traces for Pennzane@ 2001A are plotted in Figure 5 for testing in three environments - ultrahigh  vacuum, an atmosphere of nitrogen with a minimum amount of water vapor and an atmosphere of  nitrogen with a greater concentration of water vapor. The normalized lifetimes for the tests are indicated  above the individual traces. The minimum concentration of water vapor in an atmosphere of nitrogen was  achieved simply by closing the gate valve between the chamber and the turbomolecular pump and  allowing the water vapor to accumulate in a nitrogen atmosphere as shown in Figure 2. The test at 75  286
ppm water vapor was achieved by deliberately admitting water vapor to the isolated test chamber from  the degassed water vapor source and then backfilling with dry nitrogen. The test in ultrahigh vacuum  lasted much longer than that of the test of the PFPE in ultrahigh vacuum as demonstrated in [4]. The  tribodegradation rate of the MAC is much lower than that of the PFPE. The lifetime for the test in the  minimum concentration of water vapor in the nitrogen atmosphere - whose evolution of concentration is  shown in Figure 2 - is greater than the lifetime for the test in ultrahigh vacuum. The lifetime for the test in  75 ppm water vapor is greater still. Thus water vapor in the test environment extends the lifetime of the  MAC as it does for the PFPE. As with the PFPE, the initial coefficient of friction of 0.08 is the same for all  test environments, regardless of lifetime.  0.3  0.2  LL 6  0.1  0.0 2925  Vacuum Normalized Lifetimes, orbitdpg  7902 18,180  1: . Run in 1 atmosphere nitrogen  ;/-  0 100000 200000 300000 400000  Orbit Number  Figure 5. Friction traces of Pennzand P2001 A in different environments  Discussion  Contact Conditions  The low constant friction coefficient exhibited in the tests prior to failure is an indication that the contact is  lubricated. However, the state of the lubricant in the contact is not known. Cann and coworkers [12] have  studied similar systems under a state of reduced availability of liquid lubricant using optical EHL. Although  the lubricant charge used in the SOT is three orders of magnitude smaller than used in their  investigations, the SOT system still appears to be in a state of lubrication. Gann, et al, presented an  expression for the film thickness in the contact in the fully starved regime. An evaluation of their  expression for the present conditions yields a film thickness that is unphysically small (<<I nm),  confirming that the contact regime here is fully starved or parched [13]. The specific concentration of  lubricant molecules in the contact is not known, except for the fact that there are enough of them to  287
European Space Mechanisms and Tribology Symposium (ESMATS 2005), Lucerne, Switzerland,  September 21 -23,2005.  8. Carre, D.J., “Lead Naphthenate Tribochemistry Under Vacuum and Gaseous Nitrogen Test  Conditions”, Tribology Letters, 16 (2004), 207-21 4.  9. Carre, D.J., “Effect of Test Atmosphere on Moving Mechanical Assembly Test Performance”,  Proceedings of the 3fh Aerospace Mechanisms Symposium, Johnson Space Center, May 19-21,  2004.  10. John, P.J., Cutler, J.N. and Sanders. J.H., “Tribological Behavior of a Multialkylated Cyclopentane  Oil Under Ultrahigh Vacuum Conditions”, Tribology Letters, 9 (2000), 167-1 73.  11. Pepper, S.V. and Kingsbury, E.P., “Spiral Orbit Tribometry - Part I: Description of the  T ri born e te r” , Tribology Transactions, 46 (2003), 57-64.  12. Cann, P.M.E., Damiens, B. and Lubrecht, A.A., “The Transition between Fully Flooded and  Starved Regimes in EHL”, Tribology International, 37 (2004), 859-864.  13. Kingsbury, E. P., “Parched Elastohydrodynamic Lubrication”, Transactions ASME, J. Tribology,  14. Spikes, H.A., “Comparison of Krytox 1 43ABTM and 1 43ACTM in Prolonged Mixed Sliding/Rolling”,  Report TS03 1/99 of Tribology Section, Department of Mechanical Engineering, lmperial College,  London.  15. Israelachvili,, J.N., Intermolecular and Surface Forces, 2“d Ed., Academic Press, New York, 1992  and Gregg, S.J. and Sing, K.S.W., Adsorption, Surface Area and Porosity, 2“d Ed., Academic  Press, New York, 1982, Chapter 5.5.  16. Kasai, P.H., “Perfluoropolyethers: Intramolecular Disproportionation”, Macromolecules, 25 (1 992), 107 (1 985), 229-233.  6791 -6799.  290
Nomenclature  G - conductance across the bearing  N - number of balls in the bearing  Ri -thermal resistance at the contact between the inner race and one ball  R, -thermal resistance at the contact between the outer race and one ball  Rb - thermal resistance across inner to outer race of one ball  k, - thermal conductivity of the inner race material  kb - thermal conductivity of the ball material  k, - thermal conductivity of the outer race material  a - major axis of a Hertzian contact area  q - major axis of the Hertzian contact area between the inner race and one ball  a, - major axis of the Hertzian contact area between the outer race and one ball  b - minor axis of a Hertzian contact area  b, - minor axis of the Hertzian contact area between the inner race and one ball  bo - minor axis of the Hertzian contact area between the outer race and one ball  Experiment  Reference [l 11 provides a detailed description of the experimentaj setup and measurement tecQniques. All  tests were conducted in vacuum environments of at least 1 xl0 Torr (approximately 1.3~10- Pa). As a  summary, the experimental design matrix is outlined as follows: .  1. Vary rotational speeds between 0 and 6,000 RPM. The maximum speed of 6,000 RPM is typical  of a spacecraft control moment gyroscope (CMG) or momentum wheel.  2. Apply a constant pure axial load ranging from 40 to 129 N.  3. Accommodate bearings of different sizes, namely the 101 and 204-size ball bearings.  4. Test dry or oil lubricated bearings.  5. Vary the average bearing temperature.  Oil lubricated bearings were tested in either the virgin or fully run-in states. A fully run-in bearing was  established by continually running the bearing at a constant speed until heat generation, torque, and  thermal conductance remain unchanged (about a week of continual operation).  Test Bearings  A 521 00 steel bearing and a hybrid bearing, consisting of silicon nitride balls and 521 00 steel races, were  tested. Table 1 provides the specifications of the two different angular contact ball bearing sizes used.  Table 2 summarizes the relevant material properties.  Table 1. Ball Bearing Specifications  292
Table 2. Ball Bearing Material Properties  8 0.06-  E  3 0.04- aJ V  u I m  u 0.02 - oisson's Ratio 0.27  hermal Conductivi  Yovanovich (theory)  G = 0.276 * P'" Results  A Yovanovich Following the experimental matrix, tests were performed to establish the influence of rotational speed,  axial load, and temperature on bearing thermal conductance. While results for oil-lubricated bearings  prove most relevant to a spacecraft application, the influence of axial load on the thermal conductance of  a static dry bearing is already well known analytically [l-21. Thus, the experimental results for a static dry  bearing provide an excellent means of verifying the experimental technique. Further testing extends the  results to static and dynamic (constant rotational speed) lubricated bearings, and reveals drastic shifts in  the magnitude and trend of bearing thermal conductance relative to axial load and temperature.  Drv (Non-Lubricated) Static Bearinas  The analytical relationship between thermal conductance and axial load in a non-lubricated, non-rotating  bearing is well known. Yovanovich [l -21 developed the best-known analytical method, deriving the  equation for bearing thermal resistance. A comparison with this model was used to verify the experimental  approach, before proceeding to the more complex cases of lubrication and motion.  Figure 1 provides a comparison of experimental and theoretical bearing thermal conductance as a  function of axial load at 20 "C. Both experiment and analysis agree closely in both magnitude and trend.  The Figure shows that thermal conductance of the dry static bearing responds to axial load to the 1/3  power.  0.08 1  o Experiment  Axial Load (N)  Figure 1. Comparison of Analytical Calculation of Thermal Conductance with Experiment for a  Static Dry 204-Size Hybrid Bearing at 20 "C  293
Yovanovich’s analysis provides an explanation for the observed trend. His model assumes the  mechanism of heat transfer through a non-lubricated static bearing is pure conductance, from the hotter  race, through the ball, and to the cooler race (Figure 2) and the driving source of resistance across the  bearing arose from the thermal constriction region between the ball and the race contacts. This  recognition presumes perfect contact (no asperities), and negligible resistance within the ball and both  races.  Q A Cooler  Race  Rball = Rin + Rout  Gball = 1 /Rball Hertzian  Contact  Hotter Race t Rin  Figure 2. Thermal Resistance Across One Bearing Ball  As such, Yovanovich calculated the thermal resistance across a dry static bearing by modeling the ball  and races as semi-infinite half-planes with the Hertzian contact area modeled as the thermal constriction  region. The basic equations to calculate thermal resistance across each of the ball to race contacts are:  inner race to ball thermal resistance  ball to outer race thermal resistance  total thermal resistance across the ball  Where Y is a non-dimensional geometric factor defined as:  \I12 where n = i or o 17  “-b“sin2 a: (4)  The Hertzian contact ellipse, and associated major and minor axes (a and b), can be calculated by a  number of existing programs based on classical Hertzian theory, such as BRGSlOC [12]. Here, the major  and minor axes are related to the applied axial load to the 113 power. Thus, the total Hertzian contact area  is related to the applied axial load to the 213 power.  a.b  4 A , =n.- Hertzian  Once the resistance across each ball is known, the conductance is determined by taking its inverse.  Scaling this result by the number of balls in the bearing yields the total bearing conductance (Equation 6).  Equation 6 assumes that each ball provides an equivalent thermal pathway as a result of pure axial load, a  condition consistent with our experimental setup. As the influence of axial load cancels out in the non-  294
dimensional parameter Y, only the influence on the Hertzian contact ellipse major axis remains in the  thermal conductance equations of 1-3. Thus, analytically, conductance is sensitive to axial load to the 1/3  power. Both the magnitude and trend predicted by the analysis compared well with the experimental  measurements.  0.070 -  0.060 -  @ E 0.050 -  E 3 0.030 -  a 8 0.040-  a  g 0.020 - Figure 3 shows the effect of the average bearing temperature (average between the inner and outer  races) on thermal conductance of a dry (non-lubricated) 204-size hybrid bearing for three different axial  loads. The graph shows that the conductance of the dry bearing is not strongly responsive to  temperatures, especially when compared to the effect of axial load. This is attributed to the weak  dependencies of the relevant material properties (thermal conductivity and Young's modulus) of the steel  races and ceramic balls to temperature.  8-4  Axial Loads  +129 N  1 u  0.01 0 e84 N  +40 N  0.000 I I 1 I I  0 10 20 30 40  Temperature (C)  Figure 3. Effect of Average Temperature on Thermal Conductance of a Static  Dry 204-Size Hybrid Bearing  Oil Lubricated Static Bearinas  Yovanovich's model is a good approximation for some applications, such as dry lubricated bearings with  little motion, but once lubrication is introduced into the system, bearing thermal conductance can change  significantly. Figure 4 plots the static thermal conductance of a 204-size hybrid bearing in three lubrication  states; dry, virgin oil lubricated, and oil lubricated after run-in. The exercised bearing was fully run-in at  6000 RPM, then brought back to 0 RPM for testing. Thermal conductance was measured for three axial  loads, over a range of temperatures for each lubrication state.  The significantly higher thermal conductance of the oil-lubricated bearings, in comparison with the dry,  warrants attention. To explain the difference between dry and oil lubricated bearings, Figure 5 depicts one  of the ball-to-race contact regions. The mechanism of heat transfer is still pure conductance, but the large  increase suggests that the lubricant meniscus surrounding the ball contributes a significant heat transfer  path by increasing the constriction area at the ball to race interface. Lubricant could also potentially reduce  the thermal resistance due to asperity contact at the metal-to-metal contact region; however, the close  agreement between the Yovanovich model and the experiment for dry bearings suggest that this  represents a secondary effect.  The large increase in thermal conductance implies that the heat path provided by the lubricant meniscus  ultimately dominates, masking the influence of the Hertzian contact area. While the thermal conductance  across dry bearings are driven by the size of the Hertzian contact ellipse, and thus proves sensitive to  295
axial load, the masking of the Hertzian contact area by the meniscus explains the minimal influence of  axial loading upon the conductance of the oil-lubricated bearing. The dominance of the meniscus also  explains the observation that oil lubricated bearings were found to be more sensitive to temperature, a  result attributable to the higher temperature dependencies of the lubricant material properties.  0.25 0.20 1  @I 0.15  Q)  $ 0.10 -  a a  u 0.05 - +129 N, Dry  -a-84.5 N, Dry  4 129 N, Virgin, Lubricated  +84.5 N, Virgin, Lubricated  +40 N, Virgin, Lubricated + 129 N, Run-In, Lubricated  +84.5 N, Run-In, Lubricated  4-40 N, Run-In, Lubricated 440 N, Dry  0.00 I I I I I I  -1 0 0 10 20 30 40  Temperature (C)  Figure 4. Effect of Average Temperature and Axial Load on Conductance of a Static 204-Size  Hybrid Bearing, for Dry and Oil Lubricated Bearings  Lubricant  meniscus  erkian contact area  Figure 5. Lubricant Meniscus Adds to Thermal Pathway Between Ball and Race  The conductance values of the virgin bearing are slightly higher than those of the fully run-in bearing, as  apparent in Figure 4. This observation was attributed to lubricant loss during the run-in process, as the ball  pushed excess oil out of its pathway and centrifugal forces displaced lubricant from the ball. This means  that the degree of difference will be dependent upon the initial amount of lubrication and the maximum  run-in speed of the bearing.  For space applications, bearings are typically lubricated once, and this lubricant may deplete over the  mission duration due to various reasons, such as lubricant migration or run-in. Knowing this, one may  construct bounds on static or slow moving conductance values expected throughout the spacecraft  mission life. Yovanovich establishes the lower extreme bound, and test measurement of a virgin bearing  establishes the upper bound. Lubricant loss can occur throughout the life of the bearing for various  reasons including run-in and lubricant migration, meaning the end-of-life conductance would be  somewhere between these bounds. This observation, however, does not hold for dynamic bearings.  296
Oil Lubricated Dvnamic Bearinas  Figures 6-8 show the influence of average bearing temperature on conductance at different rotational  speeds. Each figure represents a different axial load applied to an oil-lubricated bearing that had been  run-in at 6,000 RPM.  0.20 -  @ 0.18 -  E 0.16 -  1 0.22 I  +ORPM  -ff 3000 RPM  -A- 4500 RPM  -8- 6000 RPM 40N Axial Load  0.08 ! I I I I I  -1 0 0 10 20 30 40 50  Average Bearing Temperature (C)  Figure 6. Thermal Conductance of a 101-Size Steel Bearing for a 40-N Axial Load  0.22 I  0.20 -  a 0.18 -  0.16 -  1 0.14 - iE  0  'F) E 6 0.12 -  0.10 - +O RPM 84.5N Axial Load  43- 3000 RPM  k- 4500 RPM  I '3000RPM P  0.08 f I I I I I  -1 0 50 0 10 20 30 40  Average Bearing Temperature (C)  Figure 7. Thermal Conductance of 101-Size Steel Bearing for 84.5-N Axial Load  297
0.22 I  0.20 -  0.18 -  0.16 -  y 0.14 -  u 0.12 -  0.10 - @  i  a  0.15 -  @ 0.14 -  E 0.13 - +ORPM 129N Axial Load  -Ef 3000 RPM  4500 RPM  1000 RPM  3  8 0.11 - 3 0.12 -  m  u 0.10 -  0.09 -  0.08 t  I I I d  0.08 1 I I I I I  -1 0 0 10 20 30  Average Bearing Temperaturn (C) 40 50  Figure 8. Thermal Conductance of 101-Size Steel Bearing for 129-N Axial Load  Several observations arise from these results. First, a distinct difference exists between static and  dynamic bearings. Once in motion, the thermal conductance increased with temperature in a linear  manner over the data range explored. Moreover, the slope and magnitude increased with axial load. To  separate the response to each individual variable, Figure 9, shows the influence of axial load on the  bearing at 6000 RPM and at a temperature of 20 "C. Unlike static lubricated bearings that are insensitive  to axial loads, the dynamic lubricated bearing responds in a linear manner over the test range.  P  Figure 9. Effect of Axial Load on a Lubricated 101-Size Steel Bearing at 20 "C  298
Figure 10 re-plots the data to investigate the influence of rotational speed at a constant temperature  (20°C). There is a distinct difference between 0 RPM and motion, which becomes more pronounced with  increasing axial loads. Furthermore, contrary to intuition, at a given temperature and axial load the thermal  conductance proved insensitive to rotational speeds below the run-in speed. The difference between static  and dynamic bearings reflects a difference in the predominant heat transfer mechanism.  0.16 I  0 40N Axial Load  0 84.5N Axial Load A 0.15 1  5 0.13  0  0 A  0  0 1  I  Static  0.09 1 Test Data Dynamic Test Data  0 1000 2000 3000 4000 5000 6000  Speed (RPM)  Figure 10. Effect of Speed on a Lubricated 101-Size Steel Bearing at 20 "C  For moving bearings, conductance through the bearing is not the primary mechanism of heat transfer.  Figure 11 illustrates a ball with a lubricant film and a meniscus at the ball to race contact. The dominant  mechanism of heat transfer is most likely mass transport, where the lubricant at the meniscus of the hotter  race picks up heat, transports it with the ball as it rotates, then deposits heat at the cooler race. The film  thickness on the ball becomes a dominant player as it determines the amount of heat transport that  occurs. The ball film thickness should not be confused with the elastohydrodynamic (EHD) film thickness,  which in the context of our argument has no influence on heat transport.  Lubricant film I)- thickness on ball -  EHD film thickness Meniscus  Figure 11. Lubricant Distribution on Bearing Ball  As the minimum rotational speed tested was 3000 RPM, it stands to reason that there may be a transition  region between 0 RPM and 3000 RPM, where speed does influence bearing conductance. Looking  through existing literature, Stevens and Todd from ESTL [3] explored the influence of speed on bearing  thermal conductance at low speeds. An example of their findings, depicted in Figure 12, indicate an initial  decline in thermal conductance, followed by an increase at higher speeds.  299
0.1  0.08  @ 5 a  0.06  8  8 a  u 0.04  0.02 + 1OOrevs  0 15,OOOrevs (Speed Increasing)  A 100,OOOrevs (Speed Decreasing)  + 100,OOOrevs (Speed Increasing)  100,OOOrevs (Speed Decreasing) .. .. .: :  ... 1  0.1 1 10 100 1000  Figure 12. Effect of Low Speeds on a 42-mm OD Bearing Lubricated with 11.4 mg of BP135 Oil  with 40-N Load; Study by Stevens and Todd [3]  Of additional relevance, Stevens and Todd [3] conducted tests multiple times with consecutively  increasing or decreasing speeds. The result was a reduction in conductance occurring with additional  revolutions. This is an effect that we also observed when comparing virgin and fully run-in bearings, again  most likely due to the effect of excess lubricant being displaced over time. Our observations, and the  literature results, indicate that this effect eventually diminishes and tests become repeatable when the  bearing reaches a fully run-in state.  Discussion  The objective of this research was to acquire a fundamental understanding of bearing thermal  conductance, to establish which variables influence that property, and to determine the dominant  mechanisms of heat transfer. Research results indicate that bearing thermal conductance was influenced  by a number of interdependent variables, underpinned by the lubrication and the dynamic state of the  bearing. Figure 13 provides a global example by plotting the influence of axial load on a 101-size steel  bearing at 20 "C. The large differences in trend and magnitude between the dry and lubricated, static and  dynamic bearings illustrate that the thermal analyst needs to recognize the assumptions underlying  experimental data or analytical models of bearing thermal conductance.  Metal-to-metal contact, through the Hertzian contact ellipses, provided the heat transfer mechanism for  dry static bearings. Experimentally, dry static bearings were found to be sensitive to axial load to the 113  power, as predicted by the Yovanovich analytical model, and relatively insensitive to temperature.  Yovanovich's method establishes a lower bound for any bearing.  The presence of lubrication can drastically change the bearing thermal conductance. The menisci  contributed an additional thermal pathway that overshadowed the conductance through the metal-to-metal  Hertzian contact area. In contrast to dry bearings, the influence of axial load was negligible for static  lubricated bearings, but the effect of temperature proved significant.  300
Mass transport of oil was considered the dominant mode of heat transfer in the lubricated dynamic  bearing. Both axial load and temperature all affected this mechanism of heat transfer, and ultimately the  bearing thermal conductance.  0.1 6  0.1 5  0.14  0.1 3  0.1 2  0.1 1  0.09 9  s Y u  E a a 0.08  0.07  0.06  0.05  0.04  0.03  0.02  0 Dynamic, Lubricated  4 Static, Lubricated  -0- Static, Non-Lubricated  50 100  Axial Load (N) 150  Figure 13. Effect of Axial Load on Thermal Conductance of a 101-Size Steel Bearing at 20 "C  30 1