ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
References  1. Roberts, E. W. "Towards an Optimised Sputtered MoS 2 Lubricant Film" Proc. 20th Aerospace Mechanisms Symposium, NASA CP-2423 (1986) 2. Hampson, M. et al. “Towards the Effective Solid Lubrication of Ball Bearings Operating at High Temperature” Proc. 12th ESMATS , ESA SP-653 (2007) 3. Hampson, M. & Roberts, E. “Cryogenic-Vacuum Assessment of Bearings Fitted with PGM cages and Lubricated with MoS 2 or Dicronite” ESA-ESTL-TM-0055 01-  (2008) 4. Cunningham, J. “Effects of Ground Testing on the Torque Performance of MoS 2 Lubricated Ball Bearings Fitted with Duroid Cages” ESA-TM-193  (1998) 5. Buttery, M. & Roberts E. “Impact of Environment on MoS 2” ESA-ESTL-TM-0118 01-  (2014) 6. Buttery, M. & Cropper, M. “Effect of Air Exposure and In-Vacuo Dwell on Lubricant Coatings of Molybdenum Disulphide and Lead” ESA-ESTL-TM-0071 01-  (2010) 7. Buttery, M. “Effect of Curing Temperature on Solid Lubricant Films” ESA-ESTL-TM-0081 01- (2011) 8. Jones Jr, W. “The Properties of Perfluoropolyether Used for Space Applications” NASA TM106275  (1993) 9. Conley, P. & Bohner, J. “Experience with synthetic fluorinated fluid lubricants” Proc. 24th Aerospace Mechanisms Symposium, NASA CP-3062 (1990) 10. Mori, S. & Morales, W. “Tribological Reactions of Perfluoroalkyl Polyether Oils with Stainless Steel Under Ultrahigh Vacuum Conditions at Room Temperature” Wear 132(1) , pp. 111 (1989) 11. Buttery, M. “Fomblin Z25: A New Method for its Degradation Assessment & Proposal for Safe Operation in Space” Proc. 15th ESMATS  (2013) 12. Carre, D. “Perfluoropolyalkylether Oil Degradation: Inference of FeF 3 Formation on Steel Surfaces under Boundary Conditions” SLE Transactions , 29(2), pp. 121 (1986) 13. Pepper, S. “Effect of Test Environment on Lifetime of Two Vacuum Lubricants Determined by Spiral Orbit Tribometry” Proc. 38thAerospace Mechanisms Symposium, (2006) 14. Kent, A. “Life of Space Greases in Air” ESA-ESTL-TM-0190 01- , (2016) 15. Pepper, S. & Kingsbury, E. “Spiral Orbit Tribometry – Part I: Description of the Tribometer” PepperI.JanTT03, (2002) 16. Buttery, M. “An Evaluation of Liquid, Solid and Grease Lubricants for Space Mechanisms Using a Spiral Orbit Tribometer ” Proc. 40th Aerospace Mechanisms Symposium, (2010) 17. Buttery, M. “Hybrid Lubrication ” ESA-ESTL-TM-0110 01-  (2012) 18. Vortselas, A. “Preliminary Evaluation of Solid, Fluid and Hybrid Lubricated Gears” ESA-ESTL-TM0202 01-  (2018 in prep) 19. Colas, G. et. al. “Dry Lubrication Efficiency: From Ground to Space” Proc. 15th ESMATS , (2013) NASA/CP—2018-219887 164
experimental modal analysis and finite element simulations, the compensation ring was identified as the  main cause of the 300-Hz mode, from here on referred to as CBEND. The mode shape presented in  Figure 3 shows a deformation of the ring geometry and lateral shift of the ring on its soft serial-kinematic  suspension. Mirror motion results from this coupled mode.   Figure 2. Schematic 120-degree section cut of the Tilt Chopping Mechanism showing the main moving  components. Flexure elements are used throughout the whole system to reduce friction, backlash, and  wear out.  Figure 3. Mode Shape of the 300-Hz resonance (CBEND).   Left: Deformation of the compensation ring (cyan) and the coupled tip-tilt motion of the mirror (magenta).  Right: A lateral shift of the compensation ring is the result of the soft serial-kinematic mount and the soft  compensation ring. The compensation blade flexure (green) and the lever pivots (cyan) show  deformations.  NASA/CP—2018-219887 167
Solving the 300-Hz Resonance Problem  Concepts have been developed to constrain the 300-Hz resonance. The goal is to find a minimal invasive  solution providing maximal performance gain to not interfere with the ongoing science campaigns. One  approach that has been followed is the implementation of a parallel kinematic suspension to stiffen the nonoperational degrees of freedom of the compensation ring as shown in Figure 4 (left). The original ring is  supported by a serial kinematic suspension as shown in Figure 3. Here the loads are transferred from the  ring over the compensation blades, the lever arm, and the lever pivots to the FCM interface. Serial kinematic  suspensions are to be avoided when seeking high stiffness since the flexibility and positioning error of each  element adds up. The other approach is to stiffen the compensation ring itself and to concentrate mass at  the suspension points (nodes of mode shape). Taking strain energy out of the mode will reduce its  magnitude. Please note that the inertia tensor of the ring has to remain unchanged to retain its original  functionality.   Figure 4. Two design concepts to eliminate the 300-Hz structural resonance. Left: Parallel kinematic  suspension to constrain in plane deformations of the compensation ring as well as in plane translational  motion of the ring on top of the suspension. Right: Concentrating mass at the suspension points (nodes of  mode shape) will take strain energy out of the mode. Selecting stiffer martials will additionally push the  resonance to higher frequencies.  Both concepts, the new suspension architecture and the compensation ring with optimized mass distribution  (material combination) have been analyzed in a finite element model to predict the new system response.  The results are shown in Figure 5 where the system response is plotted over the frequency range of interest.  The dominant resonance (CBEND) is clearly visible for the original configuration. Another structural  resonance is present at about 430 Hz (SWAY), a sideways motion of the mirror coupled with a deformation  of the compensation ring. Different material combinations are shown combined with the serial kinematic  suspension (SKS) and the parallel kinematic suspension (PKS). A clear improvement is visible for both  concepts. The CBEND magnitude is decreasing and the mode is pushed to higher frequencies. Applying  the parallel kinematic suspension while keeping the original aluminum ring results in a magnitude drop of  about 35%. Applying an extreme mass concentration at the suspension points using a material combination  of tungsten and aluminum while keeping the SKS results in a CBEND magnitude drop of about 90%.   NASA/CP—2018-219887 168
Figure 5. Open Loop frequency response for different compensation ring designs named by the  suspension type (SKS = Serial Kinematic Suspension, PKS = Parallel Kinematic Suspension) and the  material combination. The parallel kinematic suspension shows a reduced CBEND amplitude compared  to the original configuration. However, the improvement by concentrating mass at the suspension points  (diverse material combination) is dominant and leads to a reduced amplitude and a frequency shift toward  higher values.  Combining both concepts can basically result in an elimination of the CBEND mode but promotes the SWAY  resonance as shown for PKS W-AL. Using stiffer material combinations (less damping) results in a  moderate reduction of the CBEND mode without amplifying SWAY (see PKS W-SiC and SKS W-SiC). In  general, it can be observed that the more strain energy is taken out of the ring by concentrating mass at  the suspension points, the lower the positive effect of the parallel kinematic mount becomes.  Even though a general improvement for most configurations is visible, predicting the new system  performance from the open-loop frequency response is not trivial. Does a reduced magnitude or a displaced  resonance frequency result in a larger actuation bandwidth? Therefore, these transfer functions were  implemented in a control simulation developed in [3] and [4]. It was determined that the parallel kinematic  suspensions in combination with an extreme mass redistribution (PKS W-SiC and PKS W-AL) show the  largest performance improvement of about 95%. Silicon carbide benefits from its high specific stiffness  while aluminum benefits from its good damping behavior. When combining the extreme mass redistribution  with the original serial kinematic suspension (SKS W-SiC and SKS W-AL), there is still an improvement of  around 80%. Considering the hardware changes required to implement the parallel kinematic suspension  to the extra 15% performance gain, the decision was made to retain the original suspension. A detailed  description of the end-to-end simulation used to predict the new closed-loop mechanism performance is  covered in [7].  Manufacturing of a Prototype Even if the combination of tungsten and silicon carbide showed a slightly better performance, the SKS WAL is selected for manufacturing a prototype. Making a complex structure from pure silicon carbide is very  challenging and expensive. Furthermore, once installed to the aircraft, the compensation ring is  inaccessible for periodic inspections so very fragile materials should be avoided.  NASA/CP—2018-219887 169
Nevertheless, joining two very divergent materials such as tungsten and aluminum can be challenging as  well. Due to their diverse material properties, classical brazing or welding is not an option. Consequently, a  mechanical joint has to be designed that is stiff and able to compensate the divergent thermal expansions.  Figure 6 shows an approach of using flexure elements for the thermal expansion compensation. The heavy  tungsten compensation mass (material at the node location) and the lightweight aluminum bridge element  (material at the anti-node location) are bolted at the lower joint. A tempered steel flexure connects the two  bodies on the upper joint allowing vertical thermal expansion. All other degrees of freedom are blocked  keeping a high stiffness of the compensation ring assembly.  Figure 6. CAD model of the new compensation ring showing the segmentation. To compensate the  vertical thermal expansion, the bridge element and the compensation mass are connected by a flexure  joint shown on the right. A heavy tungsten alloy (97%W, 2.1%Ni, 0.9%Fe) is used for the compensation masses shown in  Figure 7a. Along with its high density (18.5 g/cm3) this tungsten alloy comes with a very high Young’s  modulus (365 GPa) providing stiffness at three of the weak cross sections of the ring. Solid carbide tools  with polychrome coating were used to machine this tough material. A dynamic trochoidal milling approach  was applied to reduce the tool wear and machining loads. To reach a very precise planarity and parallelism  of the contact surfaces, they have been precision ground.  NASA/CP—2018-219887 170
Figure 7. Compensation Ring Assembly. a: Heavy tungsten compensation mass. b: Steel flexure for  thermal expansion compensation. c: Lightweight MMC bridge element. d: Assembled compensation ring.  e: Compensation ring integrated into the SMA test bench.  To maintain a high stiffness but at a low density in between the tungsten parts, the light weighted bridge  elements (Figure 7c) are made from an aluminum-based metal matrix composite (MMC Al70-SiC30). This  material has a low density of 2.9 g/cm3 and provides a high stiffness of about 120 GPa. It combines the  advantages of aluminum (fracture toughness, damping behavior, density) and silicon carbide (stiffness,  density) at a manageable machinability. To reduce secondary machining, the main geometry has been  cast. A prototype casting approach was selected based on the thin wall geometry, the deep pockets, and  the small quantity ordered (10 pieces). A wax model was 3D printed and invested into a ceramic mold (see  Figure 8a-c). After burning out the wax, liquid AlSiC is poured into the resultant cavity forming the final part  (Figure 8d). Hot Isostatic Pressing is used to reduce the porosity of the cast as well as to improve the  chemical and mechanical bonding of the SiC phase in the metal matrix. This leads to improved material  properties and prevents the harder SiC from breaking off during machining and use. Due to the high ceramic  content of the material, post machining of the joints has been performed using diamond-coated tools  (Figure 8e). Complex fixtures were prepared to enable a steady and defined positioning of the cast part on  the milling machine.   To compensate the thermal expansion in the operational temperature range between -60°C to 30°C, flexure  elements made of high strength steel are used, shown in Figure 7b. The flexure geometry is optimized in a  NASA/CP—2018-219887 171
way to minimize local stress concentrations during cool down. The assembled compensation ring is shown  in Figure 7d.  Figure 8. Manufacturing process of the MMC parts. a: 3D printed wax model. b: Additional wax features  are added to the printed wax model forming the spruing, riser and pouring cup of the mold. c: Mold  building. d: Casting part after removing the mold. e: Post machining of the critical surfaces, threads   and pin holes.  Performance Testing  To prove the new design concept and to verify the resulting in-flight performance, extensive testing has  been performed on a mockup of the Secondary Mirror Assembly (test bench) (Figure 7e). The test bench  is a true scale rebuild of the TCM plus FCM suspension mounted on a heavy optical table. Flight electronics  (controller boards, amplifier) were used to operate the test bench for better comparison with the actual  SMA. For more information on the test bench please refer to [7]. Figure 9 shows the open-loop system  response of the original and the new compensation ring tested on the testbench. The system was driven  by applying a white noise excitation to the voice coil actuators while measuring the mirror movements with  the eddy current sensors.  Compared to the original aluminum ring, a significant improvement is visible. The CBEND amplitude  dropped by approximately 80% while the frequency increased by about 33 Hz. As predicted by the finite  element simulations, the frequency of the SWAY mode drops. Fortunately, an increase in SWAY amplitude  is absent, which can be credited to the SiC phase in the MMC bridge element when comparing to the SKS  W-AL and SKS W-SiC curve in Figure 5. Please note that due to the stiff suspension of the test bench NASA/CP—2018-219887 172
(mounted on an optical table), the resonance of higher frequency modes (CBEND, SWAY) is amplified  compared to the spider arm suspended system of the flight hardware (applied in FE simulations).  Figure 9. Open-loop frequency response of the original and the new compensation ring mounted on the  test bench. A noticeable deviation between the new hardware and the simulation result is visible, which is  mainly caused by the stiffer suspension of test bench (optical table vs. spider arms of the telescope) and  by simplifications of the contact stiffness in the FE model.  Adjusting the controller to the new open-loop system response leads to the closed-loop response shown in  Figure 10. Again, the system was driven by a white noise actuator input while measuring the mirror position.  The controller gains are changed in a way to adjust the controller to its new stability limit. Based on the  -90° cutoff frequency, the actuation bandwidth was improved by 80% (from 50 Hz to 90 Hz). Figure 10 additionally shows a closed-loop simulation result from [8] using the material properties of Al70-SiC30 for the bridge elements. Here a closed-loop bandwith of about 95 Hz (90%) is predicted. The deviation of 5 Hz can be explained by simplifications in the finite element model such as assuming a rigid behavior of the frictional contacts. Also the stiff suspension of the test bench (optical table) as described above is expected to have an influence on the closed-loop bandwidth. NASA/CP—2018-219887 173
Figure 10. Closed-loop frequency response of the original and the new compensation ring mounted on  the test bench. Zoom of magnitude and phase plot presenting the bandwidth improvement based on the  -3 dB and the -90° cutoff frequency. Conclusion  As a conclusion, the 300-Hz structural resonance of the SOFIA secondary mirror mechanism has been  solved by redesigning the ring-shaped compensation mass in a way that its mass is concentrated at the  suspension points while keeping the same inertia tensor. Tungsten and AlSiC ring segments are combined  to achive the extreme mass redistribution. The high elastic modulus of the tungsten and the SiC phase of  the metal matrix composite additionally provide an improved overall stiffness. Eliminating the resonance  enables a more aggressive controller design. Adjusting the control gains to the new stability limit, a  significant improvement in actuation bandwith of 80% was determined  from testing the prototype on a fullscale system mockup. Based on this prototype design, two flight units will be manufactured for future SOFIA  missions. The improved actuation bandwith will provide faster mirror steering capability for image  stabilization and infrared square wave chopping.  NASA/CP—2018-219887 174
References  1. L. Zago, P. M. Genequand, and J. Moerschell, “Extremely compact secondary mirror unit for the SOFIA telescope capable of 6-degree-of-freedom alignment plus chopping,” in Advanced Technology Optical IR Telescopes VI, L. M. Stepp, ed., Proc. SPIE 3352, pp. 666–674, 1998. 2. Alfred Krabbe, "SOFIA telescope", Proc. SPIE 4014, Airborne Telescope Systems, (20 June 2000); doi: 10.1117/12.389103 3. Andreas Reinacher, "Inflight Commissioning and Performance Improvement of the SOFIA Secondary Mirror Mechanism" Dissertation , 1 (2014), University of Stuttgart. 4. Andreas Reinacher and Hans-Peter Roeser, “Modeling of the SOFIA secondary mirror controller," Proc. SPIE  8336, 83360S-83360S-10 (2011). 5. Andreas Reinacher, Yannick Lammen and Hans-Peter Roeser, “SOFIA’s Secondary Mirror Assembly: In-Flight Performance and Control Approach”, (2016) Proc. SPIE 9908, doi: 10.1117/12.2232851 6. Yannick Lammen, Andreas Reinacher, Ivar Kjelberg, Serge Droz, Holger Jakob, Friederike Graf, Michael Lachenmann, Alfred Krabbe, ”SOFIA secondary mirror mechanism heavy maintenance and improvements”, Proc. SPIE 9906, Ground-based and Airborne Telescopes VI, 99061R (27 July 2016); doi: 10.1117/12.2231352 7. Yannick Lammen, Andreas Reinacher, Rick Brewster, Benjamin Greiner, Friederike Graf, Alfred Krabbe, "A new test environment for the SOFIA secondary mirror assembly to reduce the required time for in-flight testing", Proc. SPIE 9906, Ground-based and Airborne Telescopes VI, 99064T (27 July 2016); doi: 10.1117/12.2232152 8. Yannick Lammen, Andreas Reinacher, Alfred Krabbe, “Bandwidth increase of SOFIA’s tilt chop mechanism by hardware modifications and subsequent controller adjustments to achieve better image quality and improve observation efficiency” (2018), Proc. SPIE, (in preparation) SOFIA, the ”Stratospheric Observatory For Infrared Astronomy” is a joint project of the Deutsches Zentrum  fuer Luft- und Raumfahrt e.V. (DLR; German Aerospace Centre, grant: 50OK0901) and the National  Aeronautics and Space Administration (NASA). It is funded on behalf of DLR by the Federal Ministry of  Economics and Technology based on legislation by the German Parliament, the state of BadenWuerttemberg and the Universität Stuttgart. Scientific operation for Germany is coordinated by the German  SOFIA-Institute (DSI) of the Universität Stuttgart, in the USA by the Universities Space Research  Association (USRA).  NASA/CP—2018-219887 175
(1)  (2)  (3)  (4)to the square of the multiplication ratio of the gear (n2) [11]. Theoretical principles are discussed,  and two  demonstrators of the technology were designed, manufactured and tested. An overview of the tests results  of the demonstration of a high- temperature integrated eddy current damper (Z -Damper) able to operate up  to 200ºC and a mass reduced tuned vibration absorber (Z -TVA) are presented in this paper .   Op erational Principle Description   The  basic principle behind the Z -Damper technology is simple: a linear magnetic gear is used to multipl y  the input vibration displacement amplitude, speed and acceleration. The magnetic gear is composed of  three stages: a stator  (1), a n input or slow stage (2) and an output or  fast-moving stage (3) as shown in Fig.  1.   Fi gure 1. Working principle of a linear magnetic gear.   The gear  ratio is defined by the number of magnetic elements involved in the gear unit as [11] :  𝒏𝒏=𝑵𝑵𝒐𝒐−𝑵𝑵𝒔𝒔 𝑵𝑵𝒐𝒐 where  No is the number of soft magnetic teeth per gear unit  (Input stage)   Ns is the number of stator magnets per gear unit   Fo r an ideal gear, the gear ratio can be defined as  shown in E q. 2:  𝒏𝒏=𝑭𝑭𝒊𝒊𝒏𝒏𝒊𝒊𝒊𝒊𝒊𝒊 𝑭𝑭𝒇𝒇𝒇𝒇𝒔𝒔𝒊𝒊=𝑽𝑽𝒇𝒇𝒇𝒇𝒔𝒔𝒊𝒊 𝑽𝑽𝒊𝒊𝒏𝒏𝒊𝒊𝒊𝒊𝒊𝒊 where  𝑭𝑭𝒊𝒊𝒏𝒏𝒊𝒊𝒊𝒊𝒊𝒊 and 𝑭𝑭𝒇𝒇𝒇𝒇𝒔𝒔𝒊𝒊 are the input and fast -moving stage forces respectively  and 𝑽𝑽𝒊𝒊𝒏𝒏𝒊𝒊𝒊𝒊𝒊𝒊 and 𝑽𝑽𝒇𝒇𝒇𝒇𝒔𝒔𝒊𝒊 are the input and fast stage speed.   The im pedance matching between S tages 2 and 3 can then be described as:   𝒁𝒁𝒊𝒊𝒏𝒏𝒊𝒊𝒊𝒊𝒊𝒊 𝒁𝒁𝒇𝒇𝒇𝒇𝒔𝒔𝒊𝒊=𝑭𝑭𝒊𝒊𝒏𝒏𝒊𝒊𝒊𝒊𝒊𝒊·𝑽𝑽𝒇𝒇𝒇𝒇𝒔𝒔𝒊𝒊 𝑭𝑭𝒇𝒇𝒇𝒇𝒔𝒔𝒊𝒊·𝑽𝑽𝒊𝒊𝒏𝒏𝒊𝒊𝒊𝒊𝒊𝒊 Ther efore, by combining both E quations  2 and 3 into Equation  4:  𝒁𝒁𝒊𝒊𝒏𝒏𝒊𝒊𝒊𝒊𝒊𝒊 𝒁𝒁𝒇𝒇𝒇𝒇𝒔𝒔𝒊𝒊=𝒏𝒏𝟐𝟐  However, when dynamic effects of the internal masses and damping forces are considered, a more complex model is required [12] to define the vibration isolation performance of the Z -Damper. Fig.2 shows an internal  diagram of the Z -damper and the diagram of the Z -Damper connected to a vibrating mass.   NASA/CP—2018-219887 178
(5)  (6) Fi gure 2. Internal diagram of the Z -Damper (right) and diagram of the Z -Damper when connected to a  vibrating mass (left).   The black body represent ed in Fi g.2 (right) is the input or slow -moving stage of the magnetic gear , which  is rigidly connected to the vibrating mass. The blue body represents the stator of the Z -Damper , rigidly  connected to the ground. A virtual intermediate stage  is defined inside the gearbox to assist the theoretical  description of the device and the interpretation of the proposed model. This “virtual stage” presents a perfect  displacement multiplication with regard to the input stage displacement ( n·x). An inertial mas s (𝒎𝒎𝒇𝒇) is  magnet ically coupled to the “virtual stage” by an equivalent spring of stiffness ( 𝒌𝒌𝒇𝒇). T he magnetic equivalent  stiffness will be only acting when a relative displacement from the virtual stage and the inertial mass is  induced by internal inertial or damping forces.   The m ain contribution to damping will be generated by eddy current dissipation  inside the device. An  equivalent viscous damping coefficient ( 𝒄𝒄𝒇𝒇) can be  defined related to the speed of the inertial mass with  regard to the stator of the device.   It  has to be noted that those couplings or connection between stages inside the Z -Damper  are based on  contactless magnetic interactions and the representation of 𝒌𝒌𝒇𝒇 and 𝒄𝒄𝒇𝒇 are on ly intended to assist the  interpretation of the phenomena. No physical contact is required between the moving stages.   Based o n the previous model, the motion equation for the inertial mass of the Z -Damper  (𝒎𝒎𝒇𝒇) can be

defined as:   The dis placement ( 𝒙𝒙𝒇𝒇) can be related to the input displacement ( 𝒙𝒙) by the equation:   where   −𝒎𝒎𝒇𝒇𝒙𝒙̈𝒇𝒇−𝒌𝒌𝒇𝒇(𝒙𝒙𝒇𝒇−𝒏𝒏𝒙𝒙)−𝒄𝒄𝒇𝒇𝒙𝒙̇𝒇𝒇=𝟎𝟎  𝒙𝒙𝒇𝒇=𝒏𝒏𝒙𝒙·𝒌𝒌𝒇𝒇 (𝒌𝒌𝒇𝒇−𝒎𝒎𝒇𝒇𝝎𝝎𝟐𝟐)+𝒋𝒋𝒄𝒄𝒇𝒇𝝎𝝎=𝒏𝒏𝒙𝒙·𝟏𝟏 (𝟏𝟏−𝜴𝜴𝟐𝟐)+𝟐𝟐𝝃𝝃𝒇𝒇𝜴𝜴  NASA/CP—2018-219887 179