ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
"I  Outer Rina Press Fit  The operating temperature range of the gimbal is between 120 K and ambient temperature. A  compression fit between the outer bearing ring and the housing, and the inner bearing and the shaft is  desired at any temperature.  The material of the housing is Ti-6AI-4V-ELI. The bearing steel, Cronidur 30, has a higher coefficient of  thermal expansion than Ti-6AI-4V-ELI. This leads to a decrease of compression at lower temperatures as  the bearing ring shrinks more than the housing. The compensation of this effect is done by a proper  dimensioned compression fit between the bearing housing and the bearing outer ring.  The dimensioning of this compression fit must also take into account the worst-case manufacturing  tolerances; sufficient compression must still be granted even with the smallest outer ring diameter and the  largest housing bore diameter. Combining the greatest outer ring diameter with the smallest bore  diameter, the integration force of the outer bearing ring will increase. To keep this increase as small as  possible, tight tolerances are given: P5 (9 pm) on the bearing outer ring and IT5 (13 pm) (in accordance  with DIN 7151) on the housing bore.  The housing bore diameter was chosen to still have compression even with the smallest outer ring  diameter and the largest housing bore diameter at the lowest expected temperature. The seat  dimensioning provides a compression of 8...30 pm at ambient temperature and 2...24 pm at the operating  temperature of 120K.  Inner Rina Press Fit  A compression fit between the bearing inner ring and the shaft is desired. The material of the shaft is  lnconel 718. lnconel 718 has a higher coefficient of thermal expansion than Cronidur 30. So the same  thermal expansion situation as occurs between the bearing outer rings and the yokes also occurs for the  inner bearing ring seat on the shaft of the gimbal cross. Also, for the inner ring fit, a compensation using a  compression seat with tight tolerances is used: P5 (8 pm) on the bearing inner ring and IT4 (7 pm) on the  shaft.  The inner ring bore diameter is chosen to still have compression even with the smallest shaft diameter  and the largest bearing bore diameter at the lowest expected temperature. The seat dimensioning  272
Figure 3. Bearing Support (located in the gimbal yokes)  Optimization of the Needle Roller Design  The gimbal transmits the thrust of the rocket engine to the stage. This force bends the gimbal cross. The  bearing inner rings are attached to the deflecting gimbal cross whereas the bearing outer rings that are  attached on the gimbal yokes do not perform a corresponding deflection. This leads to a considerable  angular misalignment between the axis of the bearing inner ring and the axis of the bearing outer ring.  The bearing is not pre-loaded but has radial clearance. For this reason, the load distribution along the  loaded bearing needle rollers is not constant but approximately triangular shaped. It was decided to study  the load distribution in detail and to optimize the shape of the needle rollers to minimize the Hertzian  contact stress.  The load distribution was analyzed in a detailed FEM analysis of the bearings, which included a model of  each needle roller within the complete gimbal assembly so that also the impact of the stiffness of the  adjacent structure was adequately modelled. In this model, each needle roller was divided into sub-  segments. Each of these sub-segments has adequate stiff ness and is equipped with gap-elements to  simulate the bearing clearance.  Figures 4 and 5 present a typical FE model of the needle rollers attached to the bearing inner ring, and the  resulting load distribution within the bearings under worst case loads. This model was used to optimize the  load distribution and reduce the peak values of Hertzian stress on the needle roller by local reductions of  the stiff ness of the bearing support structure, and to predict the load distribution of the most loaded needle  roller of the bearing.  The resulting stress distribution was used as input for the optimization of the needle roller shape. An  adequate crowning of the needle was defined by the bearing supplier FAG.  Thus the bearing allows angular misalignment between the inner ring and the outer ring while transmitting  the operational loads with permissible Hertzian stresses.  274
Figures 4 and 5. Finite element model of the needle roller and analysis result for the needle loads  Dimensioning of Clearance  To reduce the bearing friction torque it was required that the bearings have positive radial clearance  during the whole operation cycle. A mathematical model was established that considered:  0  0  0  0  0  0 The reduction'of the outer ring race diameter by compression of the bearing outer ring due to its  compression fit in the housing. As the compression depends on the temperature, this value had to be  evaluated for the whole operating temperature range of the gimbal.  The increase of the inner ring race diameter by tension of the bearing inner ring due to its  compression fit on the shaft. Also this value had to be evaluated as temperature dependent.  The thermal elongation of the bearing steel, the Titanium alloy of the housing, and the lnconel 718 of  the shaft.  The manufacturing tolerances of the fit diameters of the outer ring and the housing as well as the  inner ring and the shaft.  The roundness tolerances of these diameters.  The elastic deformation of the outer bearing ring in the housing due to the unequal stiffness  distribution of the housing as described above.  The clearance was evaluated for the whole mission profile, using the thermal model results to apply the  appropriate thermal gradients. Figure 6 shows the bearing clearance results for the whole mission profile.  Integration of the Bearing  As highly pre-loaded compression fits are used for the bearings, integration by a press tool is mandatory.  The gimbal yokes and the gimbal cross are fixed in a special tool. The complete bearing assembly,  comprising of the inner ring, the outer ring, and 37 needle rollers is pressed axially into the gimbal  assembly by a spindle press. To ease the press process and to prevent cold welding of the parts, all  contact areas are greased with Fomblin Y 225 by Solvay Solexis, Italy. The residue is removed by  cleaning after the press process.  During integration, the integration force is measured by a load cell, and recorded. The press process  stops when the measured integration force increases rapidly, which is an indication that the shoulders on  the bearing rings are in contact with the according counter-faces of the housing and of the gimbal cross.  Figure 7 shows a typical load trace as experienced during bearing integration:  275
Time  Figure 6. Bearing clearance results for the whole mission profile  Figure 7. Bearing integration force versus insertion depth trace  The trace shows a continuous increase in insertion force as the bearing slides into its seat. It is overlaid by  vibrations of the setup. At a sudden point, the integration force increases rapidly signalling the  completeness of the integration process.  Development Test Setup  Testing of the bearing on bearing level was not considered necessary. The development testing of the  bearing was part of the development test campaign of the gimbal assembly. For the development testing  of the gimbal assembly, a dedicated test rig was developed and manufactured that allows applying the  following loads simultaneously:  0  0  Dry Nitrogen atmosphere  0 Gimbal movement of ~k6O Mechanical load, consisting of forces in longitudinal (X), lateral (Y, Z) direction and moment load,  Thermal load, applied as temperature gradient between the upper and the lower gimbal interface  plane,  276
During application of these loads the following measurements are continuously recorded:  0  0  0  0 Forces and moment load  0  0 Friction torque of the bearings  Deformation and clearance of the gimbal assembly  Gimbal angle and gimbal count  Temperature of the interface planes and of 8 measurement points on the gimbal  Strain of the gimbal structure  To allow measurement of the bearing friction torque under load and under low. temperature, a special  design of this test rig was developed. The gimbal assembly forms one joint of a pendulum support. The  second joint is a large spherical, case-hardened contact that has a very low friction torque compared to  the bearing friction torque. The pendulum is equipped with a quartz measurement plate that measures the  bending moment within the pendulum. Rotation of the gimbal assembly around a gimbal axis causes  friction torque that induces bending moment in the pendulum. By measuring the bending moment, the  friction torque can be determined. As the location of the quartz measurement plate is distanced from the  gimbal assembly and it is thermally isolated, it is possible to operate the gimbal at low temperature while  the quartz measurement platform is at ambient temperature, which is necessary for high accuracy friction  torque measurement. Accuracy is 10 Nm, measurement range is 300 Nm.  The mechanical loads, which are forces and a moment load, are applied by a computer-controlled  hydraulic cylinder system consisting of three cylinders arranged on the test rig structure, which allows  applying all the required load combinations. The loads are continuously measured by separate load cells.  The gimbal motion is applied by a crank driven by an electrical motor that is supplied by a frequency  converter to allow adjustment of the gimbal frequency.  The individual temperature of the interface plates is controlled by two separate temperature controllers  that control the flow of liquid nitrogen through pipes embedded in the interface plates. These interface  plates are thermally isolated from the test rig structure. The gimbal assembly is placed in a housing that is  purged with dry nitrogen during low temperature testing. Figure 8 shows a schematic view of the test rig.  Figure 8. Gimbal Test Rig: Schematically View and Photograph  277
By testing the bearings at the gimbal assembly level, the bearings could be tested with representative  integration such as the compression and the tension from the press fits, and with representative  deflections of the rings and between the inner ring and the outer ring caused by the deflections of the  gimbal structure under load.  Development Test Resu Its  The development tests comprised the environmental tests and the life test of the bearing. After these  tests, the performance was verified as end-of life characteristic. An overload test concluded the bearing  development test campaign.  Environmental Tests  Before the mechanical tests, environmental tests were performed to simulate on-ground and flight  environmental loads such as dry heat, damp heat, thermal cycles in damp air, thermal shock and salt fog.  The tests were performed in a climatic chamber.  No visual degradation of the bearings was observed. The selected bearing steel Cronidur 30 did not show  any signs of corrosion. The lubricant (sputtered lead) formed lead salt as expected. Particles of this lead  salt escaped from the bearing as can be seen in Figure 9.  A  Figure 9. Formation of lead salt on the lubricated needle rollers, escaping lead salt  The formation of the lead salt as a consequence of the environmental testing was observed before during  component level testing of the lubricants. These results were presented at the 37th Aerospace  Mechanisms Symposium [3]. It is considered acceptable as the lead salt acts as protective layer which  encapsulates the underlying lead. The lead salt has lubricating properties and acts as a lubricant itself.  Functional performance was verified during the subsequent performance tests.  Load Test  The bearings were exposed to the mechanical design loads at ambient and at low temperature. No  structural damage occurred.  Friction Toraue Test  The friction torque of the gimbal assembly around one axis was measured. The result is the friction torque  sum of the engaged pair of bearings. A lateral support transmits the lateral loads as the bearings support  radial loads only. The friction torque of this radial support is also included in the results; however it is  supposed to be negligible.  The friction torque was measured at room temperature, and at low temperature, and result was well below  the allowable friction torque. This means that the lubricant performed as expected even after the harsh  environmental tests.  Life Cycle Test  The life cycle test was intended as a tribological test to verify the lifetime at worst-case loads and  temperature. A thermal gradient simulating the worst-case operating interface temperatures was applied.  278
Figure 10 presents the friction traces during the test (friction torque peaks and cycle count vs. load  cycles). Gimbal operations with the worst case non operating load applied and additional gimbal  operations under the worst-case operating loads were performed. The bearing friction torque was  recorded during this test. The friction torque keeps approximately constant during the load cycles. No  significant increase in bearing torque could be observed.  Figure 10. Bearing friction trace during the life cycle test under operational load  Clearance Test  The clearance was measured after the life cycle test as an end of life value. The measured clearance was  within the clearance requirements and was according to the test prediction.  Stiffness Test  The stiffness of the gimbal assembly was measured. It is mainly determined by the bearings' stiffnesses.  The stiffness was measured in all load directions: compression of the gimbal, tension of the gimbal, lateral  force, and moment load on the gimbal. As the bearing stiffness is determined by the deformation of the  contact between the races and the needle rollers, which is a Hertzian line contact, the spring rate is not  constant. For this reason, the stiffness was predicted and evaluated for 50% and 100% of the maximum  loads. Figure 11 shows a typical deformation versus load graph of the gimbal assembly.  Flex Y-100% RT  Figure 11. Deformation versus load graph  The graph shows the deformation versus the load in both load directions, thus the hysteresis is also  shown in the graph. While recording this graph, the load direction was swapped so that during at the zero  crossing of the load the bearing clearance can be evaluated from the graph.  As a result of the test, it turned out that the stiffness in longitudinal direction and in lateral direction met the  predictions of the FEM analysis. However, the measured torsional stiffness was noticeably smaller than  predicted.  279
The reason for this difference was investigated in detail. It turned out that the torsional stiffness strongly  depends on the crowning of the needle rollers. The local rotation of the bearing housing is determined by  the contact point of the needle rollers which is the point of torque introduction into the bearing housing.  un-deformed shape  Cardan  cross  connected  with bracket  Figure 12. Simulated deformation of the bearing housing and structural analysis model  Figure 12 shows on the left hand side the locally rotated bearing housing, and its un-deformed shape. On  the right hand side, a model detail of the structural analysis model is shown. Each needle roller is  simulated by a row of 7 rod elements and 7 gap elements connected to each other. The rod element  simulates the stiffness of the racehollerhace contact and the roller whereas the gap element simulates the  radial clearance of the bearing.  The load introduction point is as initially resulting of the structural analysis model of the gimbal assembly is  marked with a full arrow. This analysis model assumed cylindrical rollers. The dashed arrow presents an  expected point of load introduction into the bearing housing as can be assumed from the stiffness test  result because of the crowning of the rollers.  A refined analysis model was established to evaluate the impact of the roller crowning by variation of  clearance along the needle similar to the real bearing needle shape instead using a cylindrical model for  the needle rollers. It turned out that the calculated stiffness decreases due to this refined simulation of the  needles. This could explain the test result.  The needle shape is optimized for stress distribution and not for optimized stiff ness.  Life Cvcle Marain Test  To verify the friction and coating behavior under higher load, the life cycle test was continued. The load  was increased up to 150% of the qualification load in steps of 10%. 500 gimbal operations were  performed with each load step.  Figure 13 shows the friction envelope traces in red and blue color, the counter in green line, and the  compression forces (negative) in black color. The graph shows the envelope curves of the measurement  platform torque which are not identical with the friction torque of the gimbal bearings themselves, however  they are proportional. It showed that the friction torque increases nearly proportional with the increasingly  applied test load. The friction torque kept far below the specified value even at 150% of the qualification  load. The hardware under test showed no visible damage.  280
Thu. 25.AwUsl2W5  - H(lku).m - HO*; *own' - ZuDan  -2- -uM - FWTX  Figure 13. Torque traces of the gimbal assembly during the life cycle margin test  Summary and Conclusion  It was demonstrated that the gimbal bearings were adequately designed and dimensioned and that the  tribological system is capable to sustain the environmental and mechanical loads and meets the related  performance requirements.  b  b  1.  2.  3. The sputtered lead coating on the Cronidur 30 substrate provides excellent tribological behavior in  terms of friction and life even after being exposed to harsh environmental loads. The expected but  unintentionally produced lead salt acts also as a lubricant.  The un-conventional bearing mount design using press fit mounts proved successful however, a  careful design was necessary as dimensional changes of the bearing rings occur due to elastic  deformation and due to summation of manufacturing tolerances.  The stiffness of needle roller bearing strongly depends on their axial alignment between inner and  outer ring. The crowning of the rollers must be included in the structural analysis model to allow  accurate prediction of the stiffness behavior.  References  Trojahn, W. et al. "Progress in B:aring Performance of Advanced Nitrogen Alloyed Stainless Steel,  Cronidur 30." Presented at the 5' International ASTM Symposium on Bearing Steels, New Orleans,  Louisiana, USA - Nov. 19-21, 1996.  N.N. "Space Tribology Handbook, 3rd Edition." AEA Technology plc - ESTL Cheshire, UK, 2002.  Neugebauer, C. and Falkner M. ,,Lessons Lhearnt on Cryogenic Rocket Engine's Gimbal Bearing  Lubrication Selection." Proceedings of the 3? Aerospace Mechanisms Symposium, (May 2004), pp.  137-1 41.  Remark & Acknowledgement  Ariane 5 is an ESA program, managed by CNES, prime contractor of VlNCl is SNECMA. The gimbal is  part of the VlNCl Thrust Chamber with prime contractor EADS-ST. The VlNCl gimbal has been developed  under contract to EADS-ST of Ottobrunn, Germany.  Austrian Aerospace gratefully acknowledges the continuing support of all parties.  281
Guide Plate  Load I- Force Transducer  Figure 1. Schematic of the Spiral Orbit Tribometer  The 12.7-mm (.5-in) diameter ball and the plate specimens were 440C steel. All tests were run at room  temperature at a mean Hertz pressure of 1.5 GPa and a ball orbit rate of 30 rpm (rolling velocity 0.071  m/s). The stainless steel, metal-gasketed test chamber was evacuated to ~0.133~1 0-8 kPa (1 x108 Torr)  with a turbomolecular vacuum pump. A gate valve between the test chamber and the turbomolecular  pump could be closed to permit the evacuated chamber to be backfilled with a particular gas. The  nitrogen and oxygen supply bottles were specified to have a water vapor content of c 2 ppm. The gasses  were admitted into the test chamber through an evacuated and baked stainless steel transfer line and a  variable leak valve. Pure water vapor was admitted into the test chamber through a variable leak valve  from a water supply that had been thoroughly degassed by the freeze-pump-thaw method. Water vapor  concentration in the nitrogen atmosphere test environment was determined with a thin film hygrometer  (Kahn Cermet II) inserted directly into the test chamber. The hygrometer's readout in dewpoint  temperature, Td, could be converted to either partial pressure of water vapor or Concentration of water  vapor in ppm through the Magnus formula and then to relative huyidity. Test :hamber total pressure P  was determined by a cold cathode ionization gauge for P-d.064~10- kPa (8x1 0- Torr), a Pirani gauge for  1.06~1 0-3 kPa (8x1 0-3T~rr)<P<0.266 kPa (2 Torr) and a diaphragm gauge for P>0.266 kPa (2 Torr).  The lowest reading of the hygrometer is Td = -99.9"C which corresponds to a water concentration of .014  ppm. This reading is achieved in vacuum and initially upon backfilling the chamber to an atmosphere of  nitrogen. Eventually, however, water vapor accumulates in the valved-off chamber as shown in Figure 2.  The accumulation is detectable after about 2x1 O4 orbits (-1 1 hours) and continues to increase thereafter.  This water vapor is due to desorption from the chamber's unbaked interior surfaces. The lowest rate of  accumulation is observed after the chamber had been evacuated for many days and then exposed to  room air for a minimal time - approximately 10 minutes - during insertion of test specimens.  284
0 40000 aoooo 120000 160000 200000  Orbit Number  0.25 - Figure 2. Accumulation of water vapor in an atmosphere of nitrogen  0 A lubricant charge of -25 pg for the PFPE and -20 pg for the MAC was deposited only on the ball from a  gas-tight syringe containing a dilute lubricant solution in a volatile solvent. This lubricant amount was  consumed in a reasonable time - from a few hours to a few days. The speed and temperature of the test  are similar to values that would be seen in spacecraft bearings, with test acceleration achieved through  limited lubricant. Friction is recorded as a function of ball orbit and results in a ‘friction trace’, shown in  Figure 3.The trace exhibits constant coefficient of friction (0.134) until an abrupt increase at -1230 orbits.  A friction coefficient exceeding 0.2 is defined as test failure, the point when the lubricant is totally  consumed. Normalized lifetime is obtained by dividing the orbits to failure by the initial lubricant charge.  In this test, failure occurred at 1287 orbits and the initial lubricant charge was 25 pg, resulting in a  normalized lifetime of 52 orbitdpg. Normalized lifetime is inversely proportional to the lubricant  destruction rate.  I i 25 pg Krytox 143ACTM, run in 1.6 Torr O2  I  1287 orbits/ 25 pg = 52 orbits/p.g  ‘Lo 0.15 i  -- 1 0 I.  285