ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
there was no shaft. The outer rings also required very high tolerances that resulted in a hybrid superduplex  configuration.  Cartridge Design: To Flange or Not to Flange?  The preliminary design of the RT and HAM were based on a simplified loads assessment of standard 108H  angular contact bearings, preloaded as a duplex back-to-back (DB) pair with a 10.16-cm (4.00-in) row  separation. By working with the bearing manufacturer to minimize radial and axial runout, a new cartridge  design configuration was developed. The OCI bearing cartridge concept is similar to x-ray bearing  cartridges, wherein the inner raceways are ground directly into a shaft and the preload is provided by fitting  a spacer between wide outer rings. The spacer is essentially in a “C” shape to facilitate assembly. This  entire preloaded cartridge can be fitted into a housing structure and clamped into position. The bearing  manufactuer has experience building this type of bearing cartridge.  The detailed design included selection of contact angle, material, and fits. Material selection was dominated  by the launch load demands and other qualitative considerations. Two contact angles were considered: 15°  and 25°. The 25° contact angle offered lower contact stress due to preload and better axial capability with  a slight decrease in radial capability and no truncation issues. Though a tighter raceway curvature could be  used to decrease operating contact stress and increase the load capacity, a looser raceway curvature  provides a lower running torque given the smaller contact area.  The initial target preloads evaluated were 111 N (25 lbf), 267 N (60 lbf), and 445 N (100 lbf). The 267-N  preload offered the best balance between protection during launch and operating contact stress. The asbuilt bearing cartridge is shown in Figure 3.  . Figure 3. Bearing Cartridge Showing Captured  Titanium Mounting Collar (Floating)Figure 4. Flanged Sleeve While material selection will be discussed next, that selection process did indeed have a direct effect on  the design. Disregarding material concerns for the moment, the original design concept was to implement  a mounting flange on the forward (shaft flanged end) outer ring. The original material consideration of M62  brought with it a concern and a reality of limited supply stock size options. That reason, combined with  requiring different designs for the two outer rings (and the increased cost of such), drove the design away  from a flanged cartridge. This required the use of the sleeve to mount the cartridge into the mechanism  housing. When the material choice later led to 440C, this decision was not revisited. The result of the flange  question was for the mechanism team to implement “sleeving”. The sleeve provides a method for mounting  the cartridge into the aluminum mechanism housing, which also provides structural support for the encoder  read heads and motor stator. In order for the sleeve to provide maximum performance enhancement for  the RT and HAM assemblies, it must: 1) be a relatively close coefficient of thermal expansion (CTE) match;  NASA/CP—2018-219887 99
2) be relatively easy to thermally control; 3) have precision interfaces (fits) to the cartridge and the housing; and 4) must completely capture the cartridge. The cartridge basic outer diameter and sleeve bore are 67.75 mm (2.6772 in). The sleeve bore runout is  specified as 5.08 µm (0.0002 in). Cartridge to sleeve fit up design values are given in Table 2. Fit ups for  the sleeve to housing were similar.  Table 2. Cartridge to Sleeve Fit Up  Diametric Tolerance Total Clearance  Cartridge +0.00 µm -2.54 µm(+0.0000 in)  (-0.00001 in) Line-to-line, min  7.62 µm (0.0003 in), max Sleeve +5.08 µm -0.00 µm(+0.0002 in)  (-0.0000 in)  Thermal control of the cartridge is accomplished by mounting a flexible, wrap-around Kapton® heating  element on the exterior mid-section of the sleeve. Test thermocouples and flight thermistors were also  installed for heater control and telemetry. The sleeve is shown in cross-section in Figure 4. Capturing the  cartridge in the sleeve was completed with a customer-supplied titanium mounting collar that was “captured”  between the rotating shafts and fixed outer forward outer races during clean assembly. At the aft end of the  sleeve, a tabbed collar was mounted such that there was axial compression providing a nominal 3.89-kN  (875-lbf) clamping force. Sleeve material selection for CTE and machinability will be described in the next  section.  Material Choices  Materials considered for the bearing rings included M62 tool steel and 440C stainless steel. The M62 tool  steel has a higher allowable contact stress than 440C (Table 3), which potentially allows the use of a smaller  bearing. A smaller bearing would lower the running torque, weight, and stiffness. For the same geometry  and load, the M62 bearing would operate at a higher contact stress with a correspondingly smaller contact  area. The 440C stainless steel is more available and is a common bearing steel, which lowers risks  associated with supply and specialized processing. Ultimately, the 440C was able to provide adequate load  capacity with the initially determined bearing size and M62 was not used. This choice assured material  availability and reduced risks associated with specialized material processing at the cost of reduced  stiffness and lower load capacity.  Table 3. Bearing Component Material Properties M62 440C Si3N4  Allowable Max Contact Stress 3790 MPa  (550 ksi) 2310 MPa  (335 ksi) n/a  Modulus 235 GPa  (34.1e6 psi) 205 GPa  (29.7e6 psi) 320 GPa  (46.4e6 psi)  Poisson's Ratio 0.29 0.283 0.26  Density / Specific Gravity 8.17 g/cc  (0.295 lbm/in3) 7.8 g/cc  (0.282 lbm/in3) 3.16 g/cc  (0.114 lbm/in3)  Thermal Expansion Coefficient 10.66e-6 1/C 10.2e-6 1/C 2.9e-6 1/C  Type Tool Steel Stainless Steel Ceramic  Expected Hardness RC 65 RC 61 HV 1550  The ball materials considered were 440C stainless steel or silicon nitride ceramic (Si 3N4). The silicon nitride  ball offers several benefits of enhanced lubricant life, lighter weight, and higher stiffness at the expense of  significantly different coefficient of thermal expansion compared to the steel rings [3, 4, 5, 6]. The additional  NASA/CP—2018-219887 100
Clamping effect on preload  The bearing integration concept calls for the bearing cartridge, which consists of the outer rings preloaded  with the C-spacer and shaft, to be clamped into the bearing sleeve. The clamping rings squeeze the outer  rings together, compressing the outer rings and the C-spacer. Since the outer rings are being compressed,  the outer raceways move closer together which lowers the preload. The compression also increases the  diameter of the raceways due to Poisson’s effect. This axial displacement of the raceway and increase in  diameter of the rings are included in the bearing mechanical analysis via a change in the initial preload and  change in fit.  C-Spacer effect on raceway parallelism The circumferential gap in the C-spacer caused an uneven distribution of the preload on the inner face of  the outer rings. Although the clamping ring was contacting the entire face of the outer ring, the C-spacer  only supported a section of the outer ring; this caused a variation in axial displacement of the outer raceway.  The concern was a loss of parallelism between the outer raceways. Finite Element Analysis (FEA) was  used to evaluate raceway parallelism affected by the C-spacer as shown in Figure 5. The worst-case  difference in axial displacement due to clamping of any point on the raceway was predicted by the FEA to  be around 525 µm (0.000021 in).  Figure 5. FEA Results Showing Deformation of the Outer Rings Due to Clamping Load Sleeve fitting effect on preload  As it is well known in the bearing field, an interference fit of the outer rings in the sleeve increases preload  by compressing the rings and decreasing the radial play. This decrease in radial play is seen by the  preloaded bearing as a change in preload and contact angle.  Mechanism housing fit effect on sleeve fit  The interference fit of the sleeve into the mechanism housing affects the sleeve diameter in turn also affects  the bearing/sleeve fit.  Analysis process  The BRGS10C software was used for the bearing internal mechanical analysis and Python for scripting and  cartridge mounting mechanical analysis. The process had to take into account the effects mentioned above.  The bearing mechanical analysis software BRGS10C was originally created for analyzing directly adjacent  preloaded pairs with conventional housings, shafts, and clamping. BRSG10C cannot on its own properly  analyze the complex mounting of a cartridge such as the OCI concept. The Python scripting was able to  account for the unique mounting concept as well as provide a means of iterating through the entire design  space quickly. The ability to iterate through all design configurations allowed a better understanding of each  parameter’s effect on the final configuration. The adopted analysis process includes the following steps:  NASA/CP—2018-219887 102
1. Apply a temperature •Survival cold, cold, room temp, hot, survival hot •Calculate new part geometries 2. Apply a sleeve clamping load (based on allowable C-spacer deformation) •3.34, 3.89 & 4.45 kN (750, 875, & 1000 lbf) •Find axial displacement of raceway due to compression (at temperature) •Find Poisson’s effect on outer diameter (at temperature) 3. Apply housing-sleeve fit •0.0000 to 5.08 µm (0.0002 in) interference •Calculate change in fit between sleeve & bearing 4. Apply sleeve-bearing interference fit •-7.62 µm (-0.0003 in) to 0.0000 clearance •Adjust fit due to Poisson’s effect from clamping & housing fit 5. Adjust preload due to clamping compression (based on axial stiffness) 6. Run BRGS10C with adjusted preload & sleeve-bearing fit 7. Report results: •Final preload •Maximum mean stress •Stiffness / Displacement •Truncation This process was followed for each combination of design parameters to determine their effect on final  preload, contact stress, and stiffness. The information was plotted in order to examine the entire design  space, as well as to determine each parameter’s effect on the result. In this way, the best decision could  be made regarding the final fit, clamping load, and initial preload.  Analysis results  The analysis process provided results for each combination of detailed parameters. Figure 6 shows  maximum mean contact stress results for the worst-case fits and clamping configuration given a  combination of axial and radial loads.  Figure 7 shows the final preload variation for all temperature, sleeve fit, housing fit, and clamping  combinations considered. At colder temperatures, the shrinking sleeve and housing fits, along with the  silicon nitride ball, cause a rise in preload. Figure 8 shows the maximum mean contact stress for launch  and operational cases. Though a lower operating contact stress is desirable, the changes seen given the  mounting parameters considered are no cause for significant alarm.  In order to allow a better understanding of each parameter’s effect on the results, series of cases were run  where certain parameters were held constant, and only one varied. The effect of the varied parameter can  then be directly determined by examining the obtained result. Figure 9 shows the effect of housing fit on  final mounted preload given a loose sleeve fit and 3.89-kN (875-lbf) clamping force. Only at colder  temperatures does the housing fit affect preload.  Figure 10 shows the change in preload due to varying sleeve fit over temperature given a nominal housing  fit. As the figure shows, the sleeve fit has a greater influence on the final preload than the housing fit. This  made sense because the sleeve is closer to the raceways and is made of a stiffer material.  Figure 11 shows the effect of clamping on the final preload. The effect is fairly consistent over the  temperature range and is a smaller effect than the sleeve of housing fits.  NASA/CP—2018-219887 103
Figure 10. Change in Final Preload Due to Sleeve  Fit, Given a Loose Housing FitFigure 11. Change in Final Preload Due to Outer  Ring Clamping  Figure 12. Axial Load Displacement Figure 13. Radial Load Displacement Figure 12  and Figure 13 show the load-displacement relationship for axial and radial loads, respectively.  The change in the shape of the curve clearly shows where the front row unloads for the axial load case.  Given the numerous design parameters and the likelihood of achieving such close tolerances on each of  those parameters, the ability to create permutations through each range, and evaluate the effect of each  tolerance gave increased confidence in the performance of the fully assembled bearing.   Lubrication  The bearing cartridge is assembled after being lightly lubricated with Nye Lubricants 2001 synthetic oil. Nye  2001 has proven itself to be a reliable space lubricant as long as enough lubricant is available to the bearing.  NyeBar®-P oil barrier films are added to each end of the cartridge to reduce oil migration from the bearing.  A Nyesorb ™ lubricant reservoir in each outer ring ensures lubricant is available throughout the lifetime of  the mechanism. Cages and reservoirs are vacuum impregnated with lubricant and weighed prior to  assembly to ensure oil take-up. A life test will be conducted to ensure adequate lubricant life for the OCI  design and operating conditions.  NASA/CP—2018-219887 105
the analysis. The viscous torque estimates can be found in Table 5, for BRSG10 and ORBIS. The ORBIS  values are lower because of its use of dynamic rather than kinematic viscosity, and because of BRGS10’s  addition of a ball factor, which is the number of balls divided by 10. When corrected for these differences,  the estimation differs by less than 2 percent.  In addition, a pair of off-the-shelf 108H bearings were disassembled and cleaned, then reassembled with  Nye 2001A oil-impregnated phenolic retainers and ceramic balls. These bearings were then preloaded in a  back-to-back configuration and verified with an in-situ load cell. Here, the lubricant condition was duplicated  to what was to be the actual cartridge condition. The oil meniscus between ball and race in this test was  visually identical to final assembled cartridges. Once assembled into a torque test rig, starting and running  torque were measured as a function of preload.   The 108H bearings were subjected to Coulomb (starting) torque measurements at various preloading  values, as shown in Figure 14, with the slope and y-intercept identified. Notice that if the negligible yintercept value is ignored, the starting torque is the slope of the curve multiplied by the preload force. As  such, the slope is an alternative form for the friction coefficient, in this case, µ = 0.0152.  Figure 14. 108 Pair Starting Torque vs Preload (Note: units are in lb and oz•in)  Running torque as a function of time can be seen in Figure 15. This data shows that the viscous torque is  clearly associated with speed, as shown in the data where the speed was changed and the torque changed  respectively. Here the effect of run in can be dramatically seen; after 47 minutes of running the torque  reduced from 56 mN•m (8 oz•in) to 39 mN•m (5.5 oz•in). This data also shows that the viscous torque is  only weakly associated with preload, as shown with the changes in preload but only minor changes in  torque given the same speed.  NASA/CP—2018-219887 107
Figure 15. Torque vs. Time Test Results (Note: units are oz•in)  As-Built Cartridge Torque Performance  Prior to shipping from the manufacturer, the bearings were torque tested on a Vibrac machine. Torque data  is typically captured during 0.5 RPM to 2 RPM rotations. Example results can be seen in Figure 16 and  summarized in Table 6. This data shows that the low-speed torque for the cartridge is about 50 percent  higher than expected from analysis or off-the-shelf testing. This means that 1) the preload is higher than  expected, 2) the coefficient of friction is higher than expected, or 3) the viscous contribution at low speed is  higher than expected. Torque testing is necessary to ground / anchor the torque estimation given the many  empirical factors included in torque estimation methods.  Figure 16. VIBRAC Running Torque Test on SN 006, Typical (Note: units are oz•in)  Table 6. SN 006 Torque Test Summary CW CCW Units  Peak Torque -14.17 (-2.007) 13.76 (1.949) mN •m (oz•in) Average Running Torque -11.21 (-1.588) 11.21 (1.588) mN •m (oz•in) Maximum Hash Width 4.44 (0.629) 2.75 (0.389) mN •m (oz•in) Average Hash Width 2.27 (0.322) 0.71 (0.101) mN •m (oz•in) Spindle Assembly and Thermal Conditioning  Spindle assembly appeared simpler “on paper” but was more difficult in the practice due to the tight  tolerances. Pre-assembly steps included precision cleaning of the sleeve, housing and flexible clamp as  well as installation of the thermal control hardware onto the sleeve. The sleeve bore was carefully inspected  for internal surface abnormalities. The sleeve bore was lubricated with a light film of Nye 2001 oil with a  clean swab and then the heater leads were connected to a 30-V power supply to expand the sleeve for  installation. Since the assembly was taking place in the swift moving, cool air of a flow bench (ISO 146441 Class 5), the sleeve in its support stand was covered to reduce convection losses. During the heating  process, the exterior of the bearing cartridge was also coated with a thin film of Nye oil and placed in a  NASA/CP—2018-219887 108
flange-up orientation to allow any excess oil to pool on the lead in edges of the cartridge. Once the sleeve  temperature, as measured by a thermistor at the flange end, reached 50-55°C the cartridge was manually  suspended over the sleeve at the lead in and dropped in. Being that the cartridge and sleeve are the same  material (and CTE) and have a very tight tolerance, binding was always a concern. While the first of five  assemblies resulted in a cartridge seizing with only a small fraction of the cartridge passed the lead in, the  remaining four cartridges dropped in flawlessly, guided be the expert hands of our assembly technicians.  The partially installed cartridge was bagged along with the sleeve and support stand and pressed into  position on a tension testing machine with approximately 534-N (120-lbf) load. The collar fasteners were  installed and torqued to fully seat the cartridge. After that, the flex clamp was installed and torqued. Figure  17 shows the thermal control hardware mounted to the sleeve prior to assembly and the cartridge installed  and fully seated in the sleeve.  Figure 17. Flanged sleeve with and without bearing cartridge Upon completion of the sleeve installation, the housing bore was inspected and lightly lubricated with Nye  oil. It was then placed on a ceramic electric heating element until a thermistor registered 65-70°C. During  heating, the sleeve external lands were similarly lubricated with Nye oil. The sleeve was dropped into the  aluminum housing and registered with its alignment pin without any difficulty. The completed housing  assembly is shown being installed in its support stand in Figure 18.  Figure 18. Assembled spindle (hardware and CAD model) In order to hold repeatable relative position of the encoder code disk and read head, the completed spindles  were exposed to thermal conditioning prior to shipment for installing the encoder components.   NASA/CP—2018-219887 109
The spindles were exposed to 12 ½ thermal cycles at ambient pressure over the survival temperature range  of 0°C to 40°C. The cycles assure that there is no residual assembly shear stress present due to differential  offsets from the shrink fits.  Spindle Performance  Upon thermal settling cycles, spindle performance testing was undertaken to assure that the assembly  mounting had not adversely affected the preload, starting torque, running torque or runout. Drive electronics  to be used for spindle testing and the basis for life testing electronics were developed. The mechanism  drive electronics (MDE) is capable of commutating a zero-cogging torque motor mounted to the flanged  end of the spindle based on a 23,600-line optical encoder also mounted on the flanged end. The MDE can  accelerate to 360 RPM with the same preliminary profile as is expected for flight. It then runs for a fixed  time in order to capture any torque decrease/change as a function of running in and then similarly  decelerate. The MDE can operate in low-speed dynamometer mode to measure Coulomb friction at 0.5 to  2 RPM.  Figure 19 shows typical low-speed torque data for a 5-minute test. The red and blue traces represent motor  current captured at 10 Hz and 500-point median filtered data via a MATLAB script. Figure 20 shows the  entire torque test results using red and blue traces as described. The data shows a decrease in running  torque at 360 RPM from 47.67 mN•m (6.75 oz•in) to 26.83 mN•m (3.8 oz•in) after 1 hour of ambient  laboratory operation.  Figure 19. Spindle Coulomb friction measurement (16.7 mN •m [2.36 oz•in], average; units are oz•in)  Two spindle assemblies were delivered to BEI Precision for high-resolution encoder installation. Their  installation process usually requires the grinding of the flange mounting interface to the encoder disk to  minimize runout and achieve the required accuracy. In this case, BEI responded with confirmation of 0.380.51•10-6 m (15-20•10-6 inch) runout, axial, including the flange surface finish. When coupled to their  accuracy tester, 0.33-0.53•10-6 m (13-21•10-6 inch) runout, axial, and 0.84•10-6 m (33•10-6 inch), radial, and  do not include surface finish. The cartridge performance was “without question, the best bearing assemblies  we’ve ever seen” and therefore required no pre-installation grinding.  Torque [in-oz] NASA/CP—2018-219887 110
Figure 20. Spindle torque measurement, ±360 RPM running, ±2 RPM Coulomb (units are oz•in)  Life Testing Plan  A plan has been created to document details of the life testing. The testing will be in two parts to address  separate life-related risks. The first tests will be of bearings only, abbreviated BOLT (for bearings only life  test) and will address the risk of lubricant life and potential degradation. For this test, 6 bearing cartridges  have been selected as representative of the lot. They will be installed into 440C sleeves with an appropriate  fit, representing flight unit fits, and then installed into stove pipe fixturing representing the mechanism  housing, again, with an appropriate fit. Heaters and thermistors will be applied as in the flight configuration  allowing us to control the bearings’ bulk temperatures within OCI’s operational range. A life test with multiple  units under test (UUTs) allows us to gain the same confidence with fewer cycles than with only one UUT.  The UUTs will be fitted with a mass simulator prior to random vibration and sine sweep testing. The mass  simulators will be removed and the UUTs will then be installed in pairs into three thermal vacuum (TVAC)  chambers sharing a vacuum pump and chiller. The vast majority of testing will consist of operation at 360  RPM as torque is “self-sensed” using the current to a cogless brushless motor. Occasional spin downs to  180 RPM and 2 RPM for Coulomb friction measurements will be performed over the course of the nonaccelerated testing. Spin down time testing may also be used to estimate any changes in friction over time.  Figure 21. BOLT Test Configuration  The second test will be of the RT and HAM mechanisms working in synchronization and will address the  risk that changes over time will affect the ability of the control electronics to synchronize the relative RT and  HAM position and velocity such that the image passes through the instrument optical slit and illuminates  the detectors with the required accuracy. These two mechanisms will be configured with an engineering  test unit telescope and half-angle mirror and installed on a surrogate optical bench. The UUTs will be  Torque [in-oz] NASA/CP—2018-219887 111