ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
106 References  

1. W.J. Derner and E.E. Pfaffenberger: “Rolling Element Bearings ,” in CRC Handbook of Lubrication,  volume II, pages 495- 537, edited by E.R. Booser, CRC Press, Boca Raton, 1983.   2. J.L. Radovich: “Gears”, in CRC Handbook of Lubrication, volu me II, pages 539- 564, edited by   E.R. Booser, CRC Press, Boca Raton, 1983.   3. C. DellaCorte: “Novel Superelastic Materials for Advanced Bearing Applications”, Presented at 13th  International Ceramics Conference of CIMTECH 2014, Montecatini, Italy, June 8- 13. Published in  Advances in Science and Technology, volume 89 (2014), pp 1-9, 2014.   4. C. DellaCorte, S. V. Pepper, R. D. Noebe, D.R. Hull, an d G. Glennon: “Intermetallic Nickel -Titanium  Alloys for Oil -Lubrication Bearing Applications,” NASA TM- 2009- 215646, March 2009.   5. S. V. Pepper, C. DellaCorte, R. D. Noebe, D.R. Hull, and G. Glennon: “Nitinol 60 as a Material for  Spacecraft Triboelements,” ES MATS 13 Conference, Vienna, Austria, September 2009.   6. C. DellaCorte, R.D. Noebe, M.K. Stanford, and S.A. Padula: “Resilient and Corrosion- Proof Rolling  Element Bearings Made From Superelastic Ni -Ti Alloys for Aerospace Mechanism Applications,”  NASA TM -2011- 217105, September 2012.   7. D.L. Carter: “Status of the Rengenerative ECLSS Water Recovery System”, Paper number M10-0762, AIAA International Conference on Environmental Systems; 11- 16 July, Barcelona, Spain, 2010.   8. D.W. Holder and C.F. Hutchens: “ Development S tatus of the International Space Station Urine  Processor Assembly,” SAE 2003- 01-2690, 33rd International Conference on Environmental Systems,  July 7 -10, Vancouver British Columbia, 2003.   9. C. DellaCorte, and W.A. Wozniak: “Design and Manufacturing Considerat ions for Shockproof and  Corrosion- Immune Superelastic Nickel -Titanium Bearings for a Space Station Application”, NASA TM  2012- 216015, presented at the 41st Aerospace Mechanisms Symposium, Pasadena, CA, May 2012.   10. M.K. Stanford, F. Thomas, and C. DellaCorte:  “Processing Issues for Preliminary Melts of the  Intermetallic Compound 60- NITINOL,” NASA TM -2012 -216044, November 2012.   11. M.K. Stanford, W.A. Wozniak, and T.R. McCue: “Addressing Machining Issues for the Intermetallic Compound 60- NITINOL,” NASA TM -2012- 216027, August 2012.   12. Barden Super Precision Ball Bearings Catalogue, D/SPC/1/USA/113/T, Danbury, CT, 2013.   13. C. DellaCort e, M.K. Stanford, and T.R. Jett : “Rolling Contact Fatigue of Superelastic Intermetallic  Materials (SIM) for Use as Resilient and Corrosion Resistant Bearings,” Tribology Letters, (2015) 57:26, January 23 rd, 2015.
109   Figure 1.  Bearing nomenclature for equivalent raceway diameter  

Capabilities  and Features  

Brief overview  of ORBIS   The ORBIS main graphical interface is shown in Figure 2. This interface is used to configure/setup the  bearing system to be analyzed. ORBIS implements a database approach for definition of key common   elements of the analysis, thereby making the setup, and subsequent attainment of results, a fairly quick  process. Database features are provided for  definition of bearings, materials and lubricants. Once  these  elements have been defined within the database,  the user only needs to assign them to their system from  the main interface.     Another noteworthy feature of the main interface is the system sketch (shown in the upper -right quadrant  of the interface). Th is sketch provides a useful  means  for checking that the setup is as intended. Within the  sketch is a drawing of all assigned rows  along with their associated contac t angle orientation.  The bearing  rows are labeled with their database name and coordinate position along the shaft. All assigned load points  are also included  with their defined coordinate positions. Additionally, rows that float on a preload spring  are identified with a spring symbol adjacent to the bearing row.    Analytical r esults are displayed in a s eparate output window as shown in F igure 3. In addition to displaying  tabulated text for the complete set of results,  the window contains  features to help recognize common  results quickly. At the top of the window is a quick reference summary table of maximum values for mean  stress, truncation, mounted preload and torque at each bearing row. The bottom region of the results window contains a control panel that allows the user to filter which results to display. Another feature, which  is illustrated in the figure, is that  all elements with st resses exceeding a predefined threshold are highlighted  in red. Additionally, any element  found to have truncati on is always  highlighted.   

A more thorough description of the ORBIS software can be found at [2] .
112 a Aluminum Parameter Value I.D., mm (in) 50.8 (2.000) O.D., mm (in) 67.36 (2.652) Ball Dia., mm (in) 3.968 (0.15625) No. balls per row 34 Free Contact Angle (deg) 22.5 Inner and Outer Curvature 0.53 Free Preload, N (lbf) 178 (40) Ring and Ball Material 440C

Figure 4.   Example problem  

Using the  sensitivity utility  in ORBIS  allows  us to perform parametric studies of inner ring fits and bulk  system temperature.  This utility  takes the user defined system and allows perturbations to an input variable,  such as inner ring fit up  or bulk system temperature, to be plotted against selected output variables, such  as mounted preload and maximu m stre ss.   Figure 5 shows two pairs of plots illustrating the effects of inner ring fit up  and bulk temperature of the  system on the mounted preload and mean Hertzian stress of the bearings. Figures 5a & 5b illustrate  that  small amounts of interference at the inner diameter of the bearing cause a rapid increase in the mounted preload. This occurs because the expansion of the inner ring reduces the free contact angle of the bearing ;  thereby increasing the manufactured preload gap, also referred to as the ring stick-out, between the  abutting inner ring faces. With a hard- preloaded  system,  this gap is forced closed and the balls must react  the increased strain.    Figures 5c & 5d show the effects of the aluminum expansion rates on the bearings . The plots illus trate that  as the temperature is either increased or decreased from ambient the mounted preload and stress will increase.  Initial fits for this example are line- to-line on the shaft and 5 μm (0.0002 in) clearance on the  housing. Note that the influence to cold temperature is more pronounced than when the system is heated  above ambient. At first glance,  this may not seem intuitive since both the shaft and housing are made from  aluminum. Hence,  as you heat the system above ambient the inner ring fit gets tighter and the outer ring fit  gets looser and, adversely , as you cool the system the outer ring fit gets tighter while the inner ring fit gets  looser. In either case one ring of the bearing is driven by its boundary condition, however  for the bearing to  react  preload the ball must have equal loads at both raceways. Therefore, the rate of preloading will be a  function of the loose ring’s hoop stiffness. Since the inner ring has a higher hoop stiffness than the outer  ring it will react loads quicker and thus the cold case is more sensitive. Another contribution to why the cold  case is worse is the fact that the diameter of the outer ring is larger than the inner ring and,  for a given  temperature delta from ambient , the magnitude of diametral change for the outer ring will slightly exceed  that of the inner ring.
113   Figure 5.  Effects of inner ring fit up on (a) mounted preload and (b) mean Hertzian stress and  effects of shaft and housing bulk temperature on (c) mounted preload and (d) max mean   Hertzian stress  

One other noteworthy model ing assumption, which is inherent in th e Jones model, is that the contact area   between the rolling elements and raceways are always fully contained. When the contact ellipse is not fully  contained,  it is said to be truncated. A truncated contact area will invalidate the reported stress but also   incurs errors to the resolved internal load distribution since the contact stiffness is a function of the contact  area. There are some simple methods to de- rate the stress for truncated elements but not the stiffness .  Hence, it becomes important to, at a minimum, know when a system has truncation.   

(a)   (b)  (c)   (d)
114 Model Validation  

Core Jones Model   Since ORBIS is based on theories published by A.B. Jones, validation of the core solver can be achieved  by comparing identical analysis cases with the A. B. Jones High Speed B all and Roller Bearing Analysis  Program . Numerous test cases were designed to demonstrate that ORBIS maintains accuracy for wide  variety of different system geometry . Key results tracked  for comparison included maximum mean Hertzian  contact stresses, beari ng row displacements, row reaction forces, and row stiffness components. These  result parameters validate that the nonlinear elastic system solver is correctly determining the internal load  distributions  and Hertzian contact mechanics  for the bearing system .   Twenty -nine test case s were developed for this comparison and are shown in Table 2 . The characteristic  system consisted of a back -to-back  pair of bearings with a single central load point. Note that contact  angles, raceway curvatures and material properties were held constant for all test cases. These parameters will be more thoroughly studied  in the flexible ring validation cases  discussed later. Independent axial, radial  and moment loading was studied to verify principal load component accuracy. Combined, or  simultaneous  loading (all 5 directions) , was also studied and included variations in system preload, bearing size, number  of balls, ball diameter and bearing row separation.    As shown i n Table 1, which summarizes the percent differences between ORBIS and Jones software for  all twenty -nine test cases, ORBIS  results correlate to the A.B. Jones program within a fraction of one  percent. These differences are deemed insignificant and likely r elate to either differences in convergence  criteria for the Newton- Raphson solver  or rounding errors between the two programs (note that ORBIS  maintains 64- bit floating point precision while it is believed that the Jones code might only hold 16- bit  precisi on). Figure  6 depicts the Hertzian stress comparisons for each of the test cases.     

Table 1.  Comparison of all results   Comparison Parameter  % Difference   Min  Mean  Max   Bearing Row Stiffness Components  -0.1%  0.0%  0.2%   Inner Ring Deflections  -0.3%  0.1%  0.6%   Bearing Row Reaction Forces  -0.2%  0.0%  0.8%   Max Mean Hertzian Contact Stress  -0.5%  -0.3%  0.0%
115 Table 2.  Test Case Parameters   Case  Bore   (mm) O.D.  (mm ) #  Balls  Ball  Dia.  (mm ) Row  Span  (mm)  Preload  (N) Fx (N) Fy (N) Fz (N) My  (Nm) Mz  (Nm)  1 12.7 19.05  21 1.59 38.10  89 89 89 89 2.26 2.26  2 19.05  25.4 30 1.59 50.8 178 178 178 178 4.52 4.52  3 25.4 34.93  28 2.38 69.85  267 267 267 267 226 226  4 50.8 66.68  34 3.97 133.35  356 445 890 890 452 452  5 76.2 92.08  50 3.97 184.15  445 890 1779  1779  904 904  6 101.6  114.3  80 3.18 228.6  667 1334  3559  3559  1695  1695   7 127 165.1  36 9.53 330.2  890 1779  4448  4448  2825 2825  8 152.4  190.5  42 9.53 381 1112  2669  8896  8896  5085  5085   9 127 165.1  36 9.53 254 890 - - - - -  10 127 165.1 36 9.53 254 890 2224  - - - -  11 127 165.1 36 9.53 254 890 4448  - - - -  12 127 165.1 36 9.53 254 890 8896  - - - -  13 127 165.1 36 9.53 254 890 1.78E4  - - - -  14 127 165.1 36 9.53 254 890 3.56E4  - - - -  15 127 165.1 36 9.53 254 890 7.12E4  267 - - -  16 127 165.1 36 9.53 254 890 - 2224  - - -  17 127 165.1 36 9.53 254 890 - 4448  - - -  18 127 165.1 36 9.53 254 890 - 8896  - - -  19 127 165.1 36 9.53 254 890 - 1.78E4  - - -  20 127 165.1 36 9.53 254 890 - 3.56E4  - - -  21 127 165.1 36 9.53 254 890 - 7.12E4  - - -  22 127 165.1 36 9.53 254 890 - - - 113 -  23 127 165.1 36 9.53 254 890 - - - 226 -  24 127 165.1 36 9.53 254 890 - - - 452 -  25 127 165.1 36 9.53 254 890 - - - 904 -  26 127 165.1 36 9.53 254 890 - - - 1808  -  27 127 165.1 36 9.53 254 890 - - - 3616  -  28 127 165.1 36 9.53 254 890 - - - 7232  -  29 127 165.1 36 9.53 254 890 - - - 14464  -
116   Figure 6 .  Comparison of maximum mean Hertzian contact stresses for various loading. ORBIS  results denoted with circles and A. B. Jones results shown with solid line. (a) shows axial loading,  (b) shows radial loading, (c) shows moment loading and (d) shows 8 combi ned load cases.  

Ring Compliance Model  Validation   Since the ring compliance model is used to establish the mounted, or operating, conditions we shall first  consider a simple duplex pair of bearings against a well -known code that contains both ring compliance  and considers both thermal expansions and ring clamping. The configuration consists of a duplex pair of  angular contact bearings oriented back -to-back  with a hard preload. The shaft and housing are made from  titanium and the bearing rings and balls are 440C. Variations of the free contact angle, raceway curvatures,  shaft and housing temperatures, I.D. and O.D. fits, and pre- ground preload are studied. Comparison of the  predicted mounted preload are made against two different versions of  A. Leveille’s cod e: BRGS10 and  BRGS14.   (a)  (c)   (d)  (b)
118 constructed as a function of gradient along with a few points from Loewenthal’s plot.  As shown, ORBIS  correlates well with Loewenthal’s  predictions for most of the gradient spectrum considered (note: original  units have been converted to SI in accordance with AMS policy) . The biggest difference between the two  predictions occurs at - 20°F  (-29°C)  case, where ORBIS predicts about twice the stress.  Note that this region  of bearing stress is highly  nonlinear because the contact area is approaching a theoretical point  (near zero  ball loads) . 

Figure 7.  Comparison of thermal gradient stresses to Loewenthal's paper  

The next example is from an analysis of the space s tation centrifuge rotor bearing s ystem  [4]. In this study,   the rotor shaft assembly is analyzed for both bulk temperature loading and thermal gradients. The rotor  shaft assembly is comprised of a three- bearing sys tem: a duplex face- to-face pair and deep- groove  (Conrad) bearing. The housing and shaft are aluminum 7075- T73 and all bearing rings and balls are 440C  stainless. Figure 8 shows the cross section of the rotor shaft assembly . As shown, the system has a fairly  large span (~600 mm) and preload springs are located on rows 2 and 3 (where row 1 is the leftmost row in the figure).  Due to both the bearing span and fact that the housing and shaft are made from aluminum there  are obvious concerns regarding how this bearing will react to both bulk temperature and t emperature  gradients.    Figure 8 shows the results from ORBIS plotted against the original publication. Here, we have chosen to  plot onl y rows 2 and 3 because row 1 does not differ from  row 2  appreciably  (row 1 and 2 being the faceto-face pair). As shown, ORBIS matches very closely with the original author for both bulk temperature  loading and temperature gradients. Row 3, which is the deep- groove bearing, clearly has the most  sensitivity to temperature fluctuations. It should be noted that the ambient shaf t fit for row 3 has a fairly  profound effect on the contact angle and the publication only mentions a range, or tolerance, for this parameter. The results shown assumed this fit was at its minimum value (0.022 mm). Also, as shown in Figure 8 (right) , row 3  has a maximum predicted contact angle that is approximately 2 degrees higher  from  the publication when the gradient is 20°C.  All other points, which are located at both higher and lower  gradients, have a much smaller deviation (less than 1° average) . This  one point appears a bit out of family  compared with the other points.    0.00.51.01.52.02.53.03.5 -20.00 -10.00 0.00 10.00 20.00 30.00 40.00 50.00 60.00Max Contact Stress, Gpa Thermal Gradient (Tir - Tor, °F)ORBIS Loewenthal
120 Conclusion  

Many aerospace designs contain bearing rings have been thinned down to minimize weight . They may also  utilize  dissimilar materials for the  bearing support structure.  Such systems require ring compliance  considerations to accurately determine the internal load distribution of the  bearings and, ultimately, the  proper performance assessment of the bearings.  The ORBIS software tool has been introduced and was  developed to account for these effects . Various  key technical analysis considerations, including relevant   assumptions and a description of the compliant ring modeling, has been discussed. Numerous validation  cases, including a few published bearing analysis studies,  were performed and all results suggest the code  closely correlates to other trusted predictions . In addition to providing reliable results,  the tool offers a  significant upgrade from the earlier MS-DOS based programs. Forethought has been taken to help ensure  ORBIS will provide  long- term dependability and robust compatibility with various computer o perating  systems and machine architectures  by using the Java programming language.  Moreover, ORBIS strives to  eliminate  many hand calculations often required with other  programs and provides iterative design utilities  that allow the user to quickly underst and and asses the performance of their bearing system . 

References  

[1]  A. B. Jones, "The Mathematical Theory of Rolling -Element Bearings," in Mechanical Design and  Systems Handbook , New York, McGraw -Hill, 1964.   [2]  J. D. Halpin, "Halpin Engineering - Software," 2016. [Online]. Available:  http://www.halpinengineeringllc.com/software .  [3]  S. H. Loewenthal, "Two Gimbal Bearing Case Stdies: Some Lessons Learn ed," in 22nd Aerospace  Mechanisms Symposium , Hampton, 1988.   [4]  J. V. Poplawski, S. H. Loewenthal, F. B. Oswald, E. V. Zaretski, W. Morales and K. W. Street,  "NASA/TP -20142217906 Analysis of Space Station Centrifuge Rotor Bearing Systems: A Case  Study,"  NASA, Cleveland, 2014.
© 2016 Sierra Nevada Corporation   122 x 1012 psi-cycles. Both the 1X and 2X life requirements for the APS  program exceed the published  lubricant test data from NASA/CR -2009- 215681, which showed acceptable life test results up to 88 x 1012  bp•psi. Since the life of the lubricated SMMA rotor bearings could not be verified by comparative test  data, a life test w as necessary to validate the design against the specific program requirements.  

Figure 1.  SMMA Actuator  

   Lubrication Film Thickness and Lambda Ratio  

A valid bearing life test needs to operate within the same lubricant regime as the on- orbit conditions  to  ensure the wear phenomena are similar . As detailed in NASA Technical Memorandum 88875 [3] , the  lubrication regime can be estimated from the Lubrication Film Parameter  (λ), which is defined as the ratio  of the lubrication film  thickness divided by the composite roughness of the contacting surfaces . Per STLE  SP-34 [4], Lambda ratios less than 1.0 result in Boundary lubrication characterized by a high likelihood of  surface asperity contact and resultant wear to the contacting sur faces . This bearing was predicted to  have an on- orbit operating temperature primarily in the range of 15°C to 20°C. As shown in Figure 2, the  lubrication film parameter  at 40.69 RPM is in the boundary lubrication regime with λ < 1 for this predicted  operating temperature range and above.    With the predicted bearing operation on- orbit occurring in Boundary lubrication, it was determined to keep  the predicted Lubrication Film Parameter (λ) below 1 for the life test in order to stay within that same lubrication regime and replicate the expected wear conditions. Additionally, there was a goal to complete the 1X portion of the life test within 20 months so that the l ife test results would be available to the  program to support modification of on- orbit operating parameters if required. Accomplishing 64.2 million  revolutions within 20 months required a life test speed of at least 73.24 RPM . Figure 3 shows that at   74 RPM, the bearings will stay in Boundary lubrication with λ < 1 at temperatures of 25°C and above. Lubrication viscous losses increase with bearing speed, generating high er friction torques within the  lubricant at 74 RPM, and results in an increase in power and temperature. Since the vacuum chamber  was without thermal control, it was predicted that the operating bearing temperature due to frictional self - heating would offs et cooling induced by the vacuum chamber cryopump and result in a net steady state  bearing temperature of at least 25°C. Any further temperature increase over 25°C would reduce the  viscosity of the lubricant and thin the lubricant film, further ensuring th at the life test bearings would  operate within the Boundary lubrication regime. After evaluating the steady state operating temperature of the bearings in the life test setup, a maximum test temperature limit of 36°C was imposed to limit  significant temper ature rise while allowing for some environmental temperature fluctuation.
© 2016 Sierra Nevada Corporation   123   Figure 2.  Lubrication Film Parameter  (λ) vs. Temperature @ 40.69 RPM  

Figure 3.  Lubrication Film Parameter (λ)  vs. Temperature @ 74 RPM  00.511.522.533.5 0 5 10 15 20 25 30 35 40 45Lubrication Film Parameter (λ ) Temperature ( °C)Lubrication Film Parameter ( λ) vs. Temperature @ 40.69 RPM Hydrodynamic Elastohydrodynamic Mixed BoundaryOn orbit operating temperature Acceptance Test Temperature 00.511.522.533.544.55 0 5 10 15 20 25 30 35 40 45Specific Film Thickness ( λ) Temperature ( °C)Lubrication Film Parameter ( λ) vs. Temperature @ 74 RPM Hydrodynamic Elastohydrodynamic Mixed BoundaryAcceptance Test Temperature
© 2016 Sierra Nevada Corporation   126 Life Test Setup  

As can be seen in Figure 4, the life test is performed inside a vacuum chamber . The selected drive motor  is not vacuum compatible and is located outside the chamber . The drive motor is a h igh quality  commercial brush DC motor  with a planetary gear output ; and drives the bearing shaft through a  ferrofluidic feedthrough in the chamber wall . The electrical feedthrough brings out the signals to monitor  the load cells and the bearing housing thermocouples . The cryopump maint ains the required pressure  level inside the chamber as monitored by the vacuum gauge.  

Figure 4.  Vacuum Chamber Setup  

Inside the vacuum chamber is the bearing life test fixture,  as shown  in Figure 5. Each bearing is mounted  and clamped within its own housing and hub in the same manner and with the same tolerances as the flight application. The six bearing housing assemblies are then loaded onto the drive shaft and clamped in  place. The drive shaft is supported by redundant fixture bearings which are spring preloaded. A flexible  coupling connects the shaft to the ferrofluidic feedthrough in the chamber wall . The bearing housings  each have a resistive arm mounted to the housing outside diameter . The resistive arm contacts the load  cells through a ball -in-cup arrangement . Based on the force recorded in the load cell and the moment arm  length to the spherical contact, the friction torque withi n the bearing can be monitored. On the opposite  side of the bearing housing a counterbalance is mounted to eliminate torque offset  due to the weight of  the resistive arm . Thermocouples are mounted to each bearing housing to monitor temperature  throughout the test.  If one bearing experienced increased torque that did not impact the testing of the  remaining bearings, then either all bearings could continue to operate without disturbing the setup or the  Cryopump   Vacuum Gauge   Electrical   Feedthrough   Drive   Motor   Roughing  Pump Hose   Manual Vent   Ferrofluidic   Feedthrough