ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
References 1. McClendon, M., “NIRSpec MSS Magnet Actuator Life Test Unit Wear Particle Evaluation”, no  report number, document obtained by Krantz, T., May 16, 2011 2. Authors unstated, “Micro Shutter Subsystem (MSS) Qualification Unit Test Report”, JWST-RPT013819, Rev. A, June 2010. 3. Krantz, T., Shareef, I, “Wear of Steel and Ti6Al4V Rollers in Vacuum”, proceedings of the 41st Aerospace Mechanisms Symposium, also NASA TM-2012-217610, 2012. 4. Pellicciotti, J. /g11/g86/g88/g69/g80/g76/g87/g87/g72/g71/g3/g69/g92/g12 , “JWST NIRSpec Micro Shutter Subsystem”, NESC-RP-11-00701,  2012. 5. Nishimura, Makoto, and Mineo Suzuki. "Solid-lubricated ball bearings for use in a vacuum—stateof-the-art." Tribology International 32, no. 11, 1999. 6. Pepper, S., “Research Note-Characterization of the Test Environment of JWST Roller Wear  Evaluation at NASA-GRC”, Aug. 1, 2011.  7. Johnson, K.L., Contact Mechanics , Cambridge University Press, 1985. 8. Kalker, J.J., “Rolling contact phenomena: linear elasticity”, Rolling Contact Phenomena CISM  Courses and Lectures, Issue 411, Spinger-Verlag, 2000. 9. McGinness, H., “Lateral forces induced by a misaligned roller”, DSN Progress Report 42-45,  March and April 1978, Jet Propulsion Laboratory, Pasadena, Calif., 1978. 10. Krantz, T., DellaCorte, C., Dube, M., “Experimental Investigation of Forces Produced by  Misaligned Steel Rollers”, proceedings of the 40thAerospace Mechanisms Symposium,  NASA/CP-2010-216272, also NASA/TM-2010-216741, 2010. 11. Antoine, J. F., et al. "Approximate Analytical Model for Hertzian Elliptical Contact Problems."  Journal of Tribology Vol. 128, 2006. 12. Archard, J. F., &Hirst, W. “The wear of metals under unlubricated conditions.”, Proceedings of the  Royal Society of London. Series A. Mathematical and Physical Sciences, Vol. 236, 1956 13. Roberts, E. W. "Space tribology: its role in spacecraft mechanisms." Journal of Physics D:  Applied Physics 45, no. 50 , 2012. 14. Scharf, T. W., and S. V. Prasad. "Solid lubricants: a review." Journal of Materials Science 48,  no. 2, 2013. 15. Authors not stated, "Properties of Du Pont VESPEL® Parts", Brochure Number H-15724-1, Du  Pont Corporation, 1993. 16. Wheeler, J. M., C. A. Collier, J. M. Paillard, and J. A. Curran. "Evaluation of micromechanical  behaviour of plasma electrolytic oxidation (PEO) coatings on Ti-6Al-4V." Surface and Coatings  Technology 204, no. 21, 2010..  17. Santos, Lucia V., Vladimir J. Trava-Airoldi, Evaldo J. Corat, JadirNogueira, and Nélia F. Leite.  "DLC cold welding prevention films on a Ti6Al4V alloy for space applications." Surface and Coatings Technology 200, no. 8, 2006. 18. Eckels, M., M. N. Kotzalas, and G. L. Doll. "Attaining High Levels of Bearing Performance with a  Nanocomposite Diamond-Like Carbon Coating." Tribology Transactions 56, no. 3, 2013. 19. Scharf, T. W., P. G. Kotula, and S. V. Prasad. "Friction and wear mechanisms in  MoS2/Sb2O3/Au nanocomposite coatings." Acta Materialia 58, no. 12, 2010. 101
Wear Potential due to Low EHD Films during Elevated Temperatures Alan Leveille*and Peter Ward** Abstract An earlier study showed that EHD films could be accurately measured in a running bearing and that the  EHD film eventually runs-in to a steady state value [1]. In the present paper, we report on additional tests  conducted on bearings with more lubricants, wider speeds, and higher temperatures. The new results  consistently show that all lubricants tested, including MAC-based lubricants have EHD film levels that are lower than model predictions in some situations. In addition, the MAC lubricants studied have lower film  thickness than traditional hydrocarbons. Figure 1 is taken from [1] and shows room temperature data of  MAC oil and Corey 100 oil, illustrating the smaller EHD film results when using this MAC oil.  Since higher temperatures produce lower films by changing the viscosity, the concern we have is that the  EHD films may be too small to prevent ball/race metal contact and resulting wear at lower speeds. Best bearing practices would have the EHD film thickness be at least three (3) times the composite surface roughness. In this paper, we will present measured EHD thicknesses of lubricant films at speeds up to  several thousand RPM for bearing bore sizes from as low as 6 mm (0.2 in) to as large as 35 mm (1.4 in)  using MAC, Corey and KG-80. Ambient temperatures from room temperature to 52 /g113C (125/g113F) are used.  Testing was done with the base oils as well as formulated greases. Greases eventually ran in to the same EHD values as the base oil but took longer times to get there.  The results clearly indicate that wear is very possible in all steel bearings when using MAC lubricants and  that this condition worsens with higher temperatures and smaller bearing size. Figure 1.  Total Film thickness (in microinches) as a function of cumulative time operating at 6000  rpm with neat base oils formulated with TCP. Open symbols represent Corey and filled symbols  represent MAC. Triangles and circles are the res ults from two independent measurement techniques *Independent Consultant, Las Vegas, NV **The Aerospace Corporation, El Segundo, CA /g51/g85/g82/g70/g72/g72/g71/g76/g81/g74/g86/g3/g82/g73/g3/g87/g75/g72/g3/g23/g21/g81/g71/g3/g36/g72/g85/g82/g86/g83/g68/g70/g72/g3/g48/g72/g70/g75/g68/g81/g76/g86/g80/g86/g3/g54/g92/g80/g83/g82/g86/g76/g88/g80/g15/g3/g49/g36/g54/g36/g3/g42/g82/g71/g71/g68/g85 /g71/g3/g54/g83/g68/g70/g72/g3/g41/g79/g76/g74/g75/g87/g3/g38/g72/g81/g87/g72/g85/g15/g3/g48/g68/g92/g3/g20/g23/g16/g20/g25/g15/g3/g21/g19/g20/g23/g30 20 40 60 80 100 12005101520253035Total Film Thickness (uin) Time (hrs) 103
Introduction Coincident with the use of MAC-based lubricants in the 90’s was the observation of wear magnitudes in  all-steel components that struck most bearing engineers as larger than previously seen and had expected. In 2008, Ward, Frantz, and Leveille presented the first publication [1] of a new instrument that  could measure EHD films in real bearings over hundreds of hours. These initial but limited tests showed  that MAC-based lubricants produced EHD films about 50% smaller than previous hydrocarbon-based  lubricants such as KG-80 and Corey 100. Additional testing done over the past few years has confirmed  that initial observation. We also observed that all of the EHD film thicknesses were thinner than predicted  by calculations using the method developed by Hamrock-Dawson [2] with the Coy-Zaretsky [3] starvation  factors that we have been using in our bearing modeling tools. The smaller EHD films raised the concern of how thin do the lubricant films become at elevated  temperatures? Will the EHD films be too thin to prevent wear due to ball/race metal contact? We will  present measured EHD films at speeds up to several thousand RPM for bearing bore sizes from as low  as 6 mm (0.2 in) to as large as 35 mm (1.4 in) using MAC, Corey and KG-80. All the test data was at  ambient of 24 /g113C (75/g113F) except for the 6-mm (0.2-in) bearing size for which data was obtained up to 52/g113C (125/g113F). We picked the 6-mm (0.2-in) size for the higher temperature testing because a smaller bearing  would inherently have a smaller film at any given speed. The results are that we see the potential for EHD films that may be bordering on marginal or a Lambda of 1 and therefore, may be a cause of wear that could affect longevity in the presence of wear acceleration factors.  Results from Earlier Testing at Room Temperature A review of data from our EHD test apparatus for actual bearings over the years of use for various  problems was done. It showed a general trend that MAC lubricants were about 50% or lower in EHD film  thickness than the old hydrocarbon lubricants such as KG80 and Corey 100. Table 1 summarizes some  of this data over the years. Table 1 is a summary at 3000 or 6000 RPM at room temperature. Table 1. Summary of Results at Room Temperature LUBRICANTBEARING  SIZE (mm) [in]RPMTOTAL MEASURED FILM  THICKNESS (μm) [μin]MULTIPLIER TO MATCH  H-D + C-Z  Corey 100  12 [0.47] 3,000 0.66 [26] 0.46 Corey 100  20 [0.79] 6,000 0.51 [20] 0.13 Corey 100   25 [0.98] 6,000 0.71 [28] 0.25            KG80  35 [1.4] 6,600 0.46 [18] 0.43            MAC oil  6 [0.24] 3,000 0.2 [7] 0.375 MAC oil  12 [0.47] 3,000 0.38 [15] 0.52 MAC oil 20 [0.79] 6,000 0.2 [9] 0.11 MAC oil  25 [0.98] 6,000 0.38 [15] 0.23            MAC grease 25 [0.98] 6,000 0.25 [10] 0.14 MAC grease 25 [0.98] 6,000 0.30 [12] 0.2 NOTE: The EHD values listed in Table 1 is the sum of the EHD film at the inner race plus the EHD film at  the outer race. H-D+C-Z is the DYBA model calculation of film thickness using Hamrock-Dowson with  Coy-Zaretsky reduction factors. 104
The data in Table 1 seem to illustrate around 50% less EHD measured values than modeling predictions  for the older oils. Additionally, the EHD film thicknesses for MAC oil are lower than the older oils.  Testing up to 52/g113/g113C (125/g113F) Further work was executed recently on a 6-mm (0.2-in) bore size bearing while using increased capability  of the EHD test apparatus. Specifically, the test ambient temperature was varied using a heat lamp. The  bearings and the entire test housing are allowed to reach a steady state ambient before starting the EHD  test. Thermocouples measure the test bearing outer ring temperature rise above ambient to obtain  accurate “actual average lubricant temperature” for modeling purposes. EHD data was taken for various speeds from 3000 rpm to less than 1000 rpm to show the speed effect on  EHD thickness. The actual EHD data was gathered, first on a bearing pair using just a small amount of  MAC oil, which is Case 1. Each bearing had just a small meniscus of oil at the ball/race interface when viewed at rest (see Figure 2). Weight checks showed 30 mg of lubricant. The phenolic cage was properly impregnated with the oil prior to test. The intent was to mimic adequate but minimal amount of lubricant oil. The EHD measurement is taken within a few seconds of a spin-motor stop, allowing the endplay  changes associated with EHD film collapse at 0 rpm to be captured isothermally. Figure 2.  An impregnated phenolic  cage with just an adequa te amount of oil to show a  small meniscus at start. Next, another bearing pair was assembled with the same impregnated phenolic cage. Then typical  cleaning sloshes with an oil/heptane mixture were done to assure bearing cleanliness followed by the addition of 25 milligrams of MAC grease. This Case 2 test shows a different lubricant procedure that represents methods known to be commonly used in industry, i.e., the use of grease lubricants. The two lubricant cases were then subjected to a moderate amount of time at 52 /g113C (125°F) running  temperature while at 1000 rpm, approximating a beginning “run-in” process.  EHD data was then taken at various speeds from 3000 to less than 1000 rpm. Two running temperatures  are shown in Figures 3 and 4, room temperature and 52/g113C (125°F). The EHD data is shown as a function  of temperature and speed for the two lubricant cases. The overall intent is to show general operating lubricant amounts, coupled with running effects theoretically known to thin an operating lubricant film. We  then correlate actual changes in EHD film thickness due to running conditions, with the modeling of EHD  film thickness using the 50% reduction factor, over and above the standard Coy/Zaretsky reduction factor.  105
Results and Discussion The data presented in Figure 3 is on the 6-mm (0.2-in), all-steel bearing pair. It is the first lubricant case  with a barely adequate amount of oil. As noted before, this data is collected after a brief “run-in” to distribute the lubricant and establish a steady-state condition. Then, the hot ambient tests are executed at different speeds followed by room temperature ambient tests after achieving the new steady state. This  approach is used in an attempt to keep the lubricant condition in the bearing as identical as possible . Figure 3. Room Temperature and 52/g113/g113C (125/g113F) EHD total values for the 6-mm (0.2-in) bore bearing lubricated with oil (Case 1).  The data clearly shows the reduction of EHD film thickness due to the actual bearing running temperature  and the accompanying viscosity change.   Next, Figure 4 shows data from the second lubricant case of tests with oil and grease. Again, the bearing  is “Run–In” first at 1000 rpm in an elevated temperature condition, then data is taken.  As before, the chamber temperature was elevated to 52/g113C (125/g113F) steady state before starting bearing  rotation. The bearing temperature during rotation was approximately 54/g113C (130/g113F). Note at the high  temperature and 1000 RPM, the total EHD film thickness has decreased to less than 1 micro-inch (0.03  micro-meters). This bearing is definitely starved.  The difference between this lubricant condition with oil and grease versus the previous test with only oil is  rather minimal and close to the EHD measurement error. The results indicate that the extra grease did not provide better protection in the first few hundred hours of operation. However, the grease/oil mixture  may act differently from pure oil as operational time continues. 106
Figure 4. The “oil plus grease” bearing (Case  2) pair at room temperature and 52/g113/g113C (125°F).  In the next few figures, the bearing race surfaces from these tests are depicted. It is evident some nonEHD conditions were established as there is evidence of beginning wear present. Figure 5.  6-mm (0.2-in) inner ring race of an original  test with a hybrid bearing of silicon nitride  ceramic balls and steel rings after the “Run-In”. Note there is no evidence of race wear with this dissimilar ball material on the steel ring.  107
Figure 6.  6-mm (0.2-in) race that generated the da ta shown in Figure 3, oil - only test. Note the worn ball path where hone lines have been elimin ated when all-steel balls and  rings/races are used. Figure 7. Inner ring that ge nerated the data in “Oil and Grea se Test” shown in Figure 4. Again,  there is wear in the ball path of the raceways that were tested against steel balls.  108
Figure 8. A higher magnification view of the race in Figure 7, clearly showing the worn off hone  lines and darkened grease in upper portion of the photomicrograph. Modelling Estimates of Inner Race (IR) EHD Va lues over a Range of Speed and Temperatures The measured EHD data was the total EHD at the inner ring plus at the outer ring.  That is:                                        EHDir + EHDor = EHDtotal                                                 (Eq. 1 ) The Hamrock-Dowson equations estimate that the inner ring lubricant film thickness is about 40% of the  total. While not exact, assuming that the inner ring EHD is 40% of the total EHD measured should be  sufficient for illustrating the inner ring behavior over a range of speed and temperature versus bearing size.   All the modeling estimates of EHD films shown in this paper were based on the software “DYBA” [4]. The  MAC lubricants “multiplier” factors needed to match the Hamrock-Dowson + Coy-Zaretsky starvation  factors are plotted in Figure 9. Applying a linear fit, we see the reasonably expected reduction as bore size increases, i.e., the larger bearing would have higher starvation at a given speed. We candidly admit  that the correlation coefficient is not good but we would need significantly more testing to assure that we  have really good control on the test variables. Nevertheless, we will use the multipliers as deduced by  Table 1 to make comparative estimates of EHDir vs. speed at 24 /g113C (75/g113F) and 52/g113C (125°F) from 200 to  6000 RPM. Conservative practice to prevent ball to race contact and wear is to have a value of /g540/g3> 3 where:   /g540/g3/g32 (EHD film thickness)/ (Composite ball and raceway surface roughness)          [Eq. 2] Using a composite bearing race/ball roughness of 2.0, a condition of 3/g540 would require an EHD film of 6  micro-inches (0.2 micro-meters). The test bearings are at least this good.  109
Figure 9. Multiplier Factors from Table 1, to Match Hamrock-Dawson +  Coy-Zaretsky EHD Predictions From Figure 9, the multiplier factors used in the analysis were:  6 mm (0.2 in) = 0.42, 12 mm (0.47 in) = 0.35, 20 mm (0.8 in) = 0.25 and 25 mm (0.98 in) = 0.18.  Figure 10 shows plots of the analytical estimates for the EHD film thickness at the inner race for the 6-mm (0.2-in) and 12-mm (0.47-in) bore bearings. Figure 11 shows similar plots for the 20-mm (0.8-in) and 25mm (0.98-in) bore bearing estimates. To illustrate what might be an adequate lambda, we have placed an arbitrary line at 4 micro-inches (0.1 micro-meters) in Figures 10 & 11. Figure 10. Estimated IR EHD for 6-mm (0.2-i n) and 12-mm (0.47-in) bore bearings at 24/g113/g113C and 50/g113C y = -0.0127x + 0.4992  R² = 0.6019  0 0.05 0.1 0.15 0.2 0.25 0.3 0.35 0.4 0.45 0.5  0 5 10 15 20 25 30 MULTIPLIER TO MATCH H-D + C-z BORE SIZE - MM FROM TABLE 1 - MUL TIPLIER TO MATCH HD + C-ZFOR MAC LUBRICANTS  110
Figure 11. Estimated IR EHD for 20-mm (0.8-in) and 25-mm (0.98-in) bore bearings  at 24/g113/g113C and 50/g113C  The results shown in Figures 10 and 11 illustrate that MAC-based lubricants do not have large margins to  separate the balls and raceways and that elevated temperatures and/or smaller bearing sizes increase that danger. Conclusions and Observa tions for Future Work MAC lubricants have thinner EHD films at comparable bearing running conditions than older natural  hydrocarbon lubricants. Both lubricant types, using EHD predictions with approximately 25% to 50% reduction, seem to agree  with actual data on films present.  The thinner films react predictably to raising bearing operating temperatures and speeds, and the  resulting film thicknesses at high temperatures and at low speeds in actual applications need to be  approached with caution in terms of resulting ball to race contact and possible wear. The low film thicknesses do seem to allow ball metal to race metal wear to start. Both tests presented showed  polished wear bands in the race and darkening lubricant with metallic fines in the lubricant. When silicon  nitride balls are substituted, the wear is totally mitigated. Varying the lubricant quantity and examining long term run-in effects on EHD films are obvious next  directions for future experiments. In this preliminary study of two lubrication cases, oil and oil with grease added, the EHD thickness values appear to be equivalent. However, long term effects have not been  studied.  111
References 1. Ward, P. C., Leveille, A. R., ”Measuring the EHD Film Thickness in a Rotating Ball Bearing” , /g51/g85/g82/g70/g72/g72/g71/g76/g81/g74/g86/g3 /g82/g73/g3/g87/g75/g72/g3/g22/g28/g87/g75/g3/g36/g72/g85/g82/g86/g83/g68/g70/g72/g3/g48/g72/g70/g75/g68/g81/g76/g86/g80/g86/g3/g54/g92/g80/g83/g82/g86/g76/g88/g80/g15/g3/g49/g36/g54/g36/g3/g48/g68/g85/g86/g75/g68/g79/g79/g3/g54/g83/g68/g70/g72/g3/g41/g79/g76/g74/g75/g87/g3/g38/g72/g81/g87/g72/g85/g15/g3/g48/g68/g92/g3/g26/g16/g28/g15/g3/g21/g19/g19/g27    2. Hamrock, B.J. and Dowson, D. “Isothermal Elastohyrodynamic Lubrication of point contacts, Part IV,  Starvation Results”  /g36/g54/g48/g40/g3/g45/g17/g3/g47/g88/g69/g85/g17/g3/g55/g72/g70/g75/g81/g82/g79/g17 99 (1977), 15-23 3. Coy, J.J., and Zaretsky “Some Limitations in Applying Classical EHD Film Thickness Formulas to a HighSpeed Bearing” /g36/g54/g48/g40/g3/g45/g17/g3/g47/g88/g69/g85/g17/g3/g55/g72/g70/g75/g81/g82/g79/g17 103 (1981), 295-301 4. Leveille, A. R> “DYBA – A New Dynamic Bearing Analysis Software Tool” International Symposium on  Rolling Element Bearings, Baja Hotel, San Diego, CA May 8, 2007 112
2.0 Test Set Up A test facility at The Aerospace Corporation was used to measure thermal conductance under controlled thermal and mechanical conditions [1]. The test rig allows the measurement of thermal conductance across a single bearing under controlled axial loads and speeds, and provides the ability to vary and  monitor the thermal boundaries, all under a vacuum environment. The speed refers to the angular velocity  of the inner race, where the shaft is driven by a motor. The other race is held stationary. A single test  bearing is held under a constant load applied by the axial load device. Photos of the test facility used to measure the larger bearing’s thermal conductance and its fixture are shown in Figures 2.0.1 and 2.0.2 respectively. A similar test set up was created to measure small and medium size bearing thermal  conductance.

Figure 2.0.1 Test Facility, Including the  Electronics, and Part of the Open Vacuum  Chamber  Figure 2.0.2 Bearing Test Fixture The sizes, geometry, and material information for the bearings that were tested are listed in Table 2.0.1.  They are all conventional angular contact bearings with inner ring piloted phenolic retainers. Bearing sizes range from 28 mm to 62 mm outer diameter (OD). None of the test bearings are sealed or shielded.   Table 2.0.1 Bearing sizes, materials, and lubricant quantities Sample DescriptionOD  (mm)Race  material Ball materialDry  Bearing  weight (g)Initial  lubricant  weight (mg) Large bearing, steel 62 52100 steel 52100 steel 228.54 130 Large bearing, hybrid 62 52100 steel Silicon Nitride 188.05 130 Medium bearing, steel 47 52100 steel 52100 steel 63.4076 96.9 Small bearing, steel 28 52100 steel 52100 steel 19.8058 49.9 115
The relationship between the major and minor axis dimensional geometries and load, P, is documented in  the reference by Yovanovich [3]. A simplified summary: 3/1P an/g118 (3.5) 3/1P bn/g118 (3.6)  Conductance is the inverse of the resulting calculated Yovanovich bearing thermal resistance equations.  /g1833=/g2869 /g3019/g3277(3.7)  Simplification is required to gain a basic understanding of the relationship between conductance and axial  load and area. Substituting and simplifying with the following assumptions: o iaa/g124 (3.8) o ibb/g124 (3.9) We obtain a relationship between bearing thermal conductance and load. /g1833/g1503/g1842/g2869//g2871(3.10)  The Hertzian contact area is also defined as: 3/2 4P ba An n n /g118/g152/g152/g32/g83(3.11)  Thus, /g1827/g3406 /g1827 /g3041/g1503/g1842/g2870//g2871(3.12)     and /g1833/g1503 /g958A (3.13)  Equation 3.13 also means that the dependence of conductance on contact area weakens as the contact  area increases. As the contact area grows, conductance becomes less sensitive to further increases. Equation 3.10 shows a similar, but stronger attenuation for the effect of increasing preload on  conductance.  The relationship between a bearing’s thermal conductance and axial load is demonstrated in Figure 3.2.1. This figure compares measured and calculated conductance values for the small (Table 2.0.1), dry (non-lubricated), static (non-moving) 52100 steel bearing. The conductance analysis utilized the Yovanovich method. Note the conductance vs. axial loads relationship demonstrates the P 1/3 curve, where low values  of axial loads result in high conductance sensitivity, but with higher loads the sensitivity gradually diminishes, evidenced by a decreased slope. A comparison between the calculated and measured values  of thermal conductance indicates that the trends correlate well. 117