ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
demonstrate negligible change in spring/hinge pin assembly performance from 20°C to -100°C as shown  in Figure 10. This is consistent with design expectations, as no wet lubricants are used in the spring  assembly design and adequate clearance has been incorporated between all materials with dissimilar  coefficients of thermal expansion internal to the mechanism. Furthermore, the flexural modulus of 17-7  CH900 springs changes negligibly from 20°C to -100°C. Similar results were observed between room  temperature and the maximum flight design temperature for both projects (+105°C).  Figure 10. SWOT Mid-Hinge Spring Assembly Output Torque at Hinge Closure at Various Temperatures  Rotary Viscous Damper Design Overview  Rotary viscous dampers have been used in dozens of spaceflight missions. Significant challenges have  been encountered in the development and qualification of these dampers and several noteworthy  breakthroughs in the design and assembly process, including vacuum fluid degassing and fill methods as  well as gasket sealing designs, have been established as standard practice to ensure acceptably consistent  performance and reliable operation in flight [2].   Building upon the state of the art for rotary viscous dampers, JPL partnered with D.E.B. Manufacturing Inc.  to qualify an improved spring/piston rotary viscous damper design. The rotary viscous damper design used  in this program is a variant of the heritage D.E.B. Manufacturing Inc. 1025 model damper. New design  features include the following items:  1. Higher strength steel single wing vane shaft 2. Updated custom ball bearing shields to reduce unintended leak paths from the damping chamber to the expansion chamber through the ball bearings 3. Reduced taper angle on the valve adjustment surface for precision damping rate setting 4. An additional fastener holding the case and abutment together for added strength 5. A new compression spring and gasketed piston arrangement to replace the heritage elastomeric diaphragm used to pressurize the damping fluid expansion chamber as shown in Figure 11. Note this basic functionality of the spring/piston arrangement is similar to that of the spring/piston damper temperature compensator used on Mariner, Viking, Galileo, NSCAT and other missions [3].01020304050 -110 -90 -70 -50 -30 -10 10 30Torque (in ●lbf) Temperature (°C)Spring Cartridge & Hinge Pin Assembly  Net Output Torque at Hinge Closure Angle vs Temperature  NASA/CP—2018-219887 409
Figure 11. Heritage (Left) and New (Right) Viscous Dampers (Image Credit D.E.B. Manufacturing Inc.)  Initial prototype test results for this new D.E.B. Manufacturing Inc. damper design suggest several  noteworthy performance advantages over heritage 1025 model dampers: reduced internal backlash,  improved damping rate consistency, lower fluid recovery time, and greater structural robustness to transient  impulse torque application (such as from deployable separation kickoff events). The sections that follow  provide data supporting each of these observed advantages.  Empirical Data-Based Damper Analytical Model  As-tested damper performance for viscous dampers in this program did not conform well to ideal damper  performance theory. This result is consistent with previous viscous damper characterization test program  observations [3].  The ideal viscous damper is a rate-dependent torque device that conforms to the following simple equation:   =∙θ̇ +  (1)  where ‘c’ is the damping rate constant, ‘ θ̇’ is the angular velocity of the damper shaft, and ‘  ’ is rateindependent friction torque loss of the damper. Damper testing demonstrates that the damping rate for any  given device is not constant, varying significantly with both temperature and with applied torque. Data from  prototype damper thermal testing was used to generate functions for damping rate versus applied torque  at qualification temperature limits as well as the nominal expected flight deployment temperature. These  curve-fit functions are shown in Figure 12.     Figure 12. Measured Damping Rate vs Applied Torque at Different Temperatures  Graph Equations At -15ᵒC:   y = 4e6 *x-1.315 At 20ᵒC:   y = 7.96025e5 *x-1.111 At 40ᵒC:   y = 5.60677e5 *x-1.161 NASA/CP—2018-219887 410
As shown in Figure 12, at a given applied torque, the damping rate varies by a factor of approximately four  to five between qualification cold and qualification hot temperature extremes listed in Figure 12 for the low  torque range shown. In order to comply with project deployment duration requirement limits in all allowable  flight thermal environments, in spite of this significant damping rate variability with temperature,  thermostatically-controlled film heaters have been added to the design baseline for all dampers on both the  SWOT and NISAR spacecraft to actively control damper temperature.   Damper deployment simulations were conducted using this test-informed analytical model in order to  calculate peak damper reaction loads and to generate preliminary deployment duration estimates for worst  case flight environments and deployment configurations.  Damper Variable Torque and Temperature Testing  In order to reduce the risk of boom hinge deploy duration requirements non-compliances surfacing in the  damper qualification program, prototype damper thermal testing was conducted with the highest and lowest  net torque profiles applied to the damper throughout deployment at qualification temperature limits. Linearly  decreasing torque was applied to the damper throughout the test using a chain-on-pulley configuration in  which the chain piles on the floor, mimicking the near-linear reduction in torque available from flight  deployment torsion springs. A linearly-decreasing maximum torque profile that bounds the worst-case flight  torque profile was applied in the qualification hot temperature environment over the minimum hinge deploy  angle to generate the shortest deployment duration result. Similarly, a linearly-decreasing minimum torque  profile that bounds the worst case low flight torque profile was applied in the qualification cold temperature  environment over the maximum hinge deploy angle to generate the longest deployment duration result.  See test results in Figure 13.  Figure 13. Worst-Case Long (Left) and Worst-Case Short (Right) Deploy Duration Test Data  When these test results are compared to elastomeric diaphragm test results from past projects, the  spring/piston damper exhibited more consistent damping rate performance under constant external torque  application, as shown in Figure 14. The suspected explanation is that the spring recovers to achieve  constant damping chamber pressure more quickly than the elastomeric diaphragm. Note there was a 9%  difference in constant applied torque between the two plots in Figure 14. While the lower applied torque in  the spring/piston damper test may have influenced the damping rate consistency result, this is not expected  to be a significant factor.   NASA/CP—2018-219887 411
Figure 14. Comparison of Damping Performance with Constant Applied Torque  Damper Impulse Torque Testing  At the time of SWOT and NISAR deployable boom launch restraint release, significant separation forces  act on the boom from explicit kickoff springs as well as thermal-strain induced structural forces. These large  forces are applied at the separation interfaces, which are a significant distance from the hinge lines. The  result is high torque in excess of 678 N•m  (6000 in •lbf) applied about the hinge line. The damper is  intentionally isolated from this torque with rotational backlash in the torque transmission hardware and by  pre-setting the damper shaft rotation angle to the “backseat” of this backlash prior to launch.   A damper kickoff simulation test was implemented to represent this critical load case as shown in  Figure 15. The maximum achievable kickoff torque for the test setup was implemented at approximately  775 N•m  (6858 in •lbf), applied over the full stroke angle of the mast separation kickoff spring plungers. This  peak torque was applied to rotational inertia simulators that bounded the relevant inertia of the flight hinges.  Flight-like damper torque transmission backlash was represented in the test. Damper temperature was  controlled to the extremes of the qualification temperature range, and damper torque response was  measured (see Figure 15). Average damping rate over this brief duration was calculated and used to  generate functions for damping rate versus applied torque in the high torque regime (see Analytical  Modeling section).  The plot in Figure 16 shows significant initial damper backlash from the heritage elastomeric diaphragm  damper design with significantly lower nominal damping rate of 101.7 N •m/(rad/s) (900 in• lbf/(rad/s)) versus  565 N •m/(rad/s) (5000 in •lbf/(rad/s)) with the current design. No damper backlash has been observed in  prototype testing of the spring/piston damper.  With high enough torque applied, D.E.B. Manufacturing Inc. reported they have permanently deformed the  elastomeric diaphragm in test dampers, resulting in cavitation issues and in some cases external gas  entrainment into the expansion chamber and damping chamber (see Figure 17). Prototype high-torque  impulse testing of the spring/piston damper has demonstrated the robustness of this design against  catastrophic damage due to short duration high torque events, as can be generated by kickoff springs and  strain energy in flight deployable separations. The elastomeric diaphragm can also permanently deform  from hot and cold temperature deployments and thermal vacuum cycling. The spring/piston damper also  reduces the list of non-metallic, permeable materials within the damper to only gaskets and O-rings,  reducing the potential for unintended fluid seepage through the diaphragm (also reported by D.E.B  Manufacturing Inc. in testing).   NASA/CP—2018-219887 412
Figure 15. Prototype Damper Impulse Torque Test  Figure 16. SMAP Project Damper Test Data Showing Typical Diaphragm Viscous Damper Backlash  Figure 17. New Diaphragm (Left) and Inelastically Deformed Test Diaphragm (Right)  0246810 0 0.1 0.2 0.3 0.4 0.5 0.6Angle (Deg) Time (seconds)Prototype Damper Impulse Test Response 1060 inlb Applied for 6.3 deg, -15degC, 1  Atm NASA/CP—2018-219887 413
Due to intermittent issues with cavitation and prolonged recovery time in developmental testing for the Soil  Moisture Active Passive (SMAP) Project with the elastomeric diaphragm damper, the SMAP Project  instituted a two hour wait for damping fluid recovery prior to subsequent damper usage. Prototype testing  with the spring/piston damper has demonstrated that 30 minutes is adequate recovery time (note: less time  may be acceptable but has not been tested). The slow process of the elastomeric diaphragm recovering its  original shape is the suspected root cause of the prolonged recovery time for the elastomeric diaphragm  damper. It is important to note the elastomeric diaphragm may never recover to its original shape and can  be permanently deformed resulting in cavitation issues and in some cases external gas entrainment into  the expansion chamber and damping chamber. Damper Vibe Testing  In the SWOT and NISAR deployable reflector boom mechanism, the damper shaft must be mechanically  isolated from the high kickoff torque of the boom. These high torques are applied to the boom over a  maximum angle of 0.8 degree. In the ground stowing process, the damper shaft is positioned in the  “backseat” of the 4-degree minimum backlash. It is critical that the shaft not rotate within this backlash in  the launch vibe environment. To demonstrate that the damper shaft does not move in a bounding vibratory  environment simulating launch vibe, a random vibration test was conducted with the damper mounted on  its flight-like mounting bracket. Tests were conducted in 3 orthogonal axes as shown in Figure 18. The  initial damper shaft clocking was marked on the housing. It was confirmed that the shaft did not rotate from  the initial clocking after each 2.5-minute test. Test levels are shown in Figure 18.  Figure 18. Damper Vibration Test (X-, Y-, & Z-Axes from Left to Right)  Hinge Latching Mechanism  The hinge latching mechanism preloads precision alignment features on each hinge half to each other after  hinge closure. There is a separate latching mechanism located at each hinge. Each latching mechanism  consists of an actuator with torque transmission hardware, and a flexured hook/roller assembly (see  Figure 19). Noteworthy design, analysis, and test information for each of these are described in the sections  that follow.  NASA/CP—2018-219887 414
Figure 19. Hinge Latching Hardware Overview  Latching Actuator Thermal Testing   The harmonic gear set used in the actuator is not capable of transmitting more than 22.6 N•m (200 in•lbf)  of torque without structural damage. Stall torque was adjusted to just below this threshold in the worst case  hot (highest output torque) environment for actuator operation. A significant reduction in actuator output  torque of about 43% was observed in prototype testing from this maximum at the worst-case hot  environment (75°C) to the worst-case cold environment (-30°C) due to Braycote 601 lubricant thickening  and reduction in motor winding resistance at cold temperature as shown in Figure 20. This reduced output  torque performance in the worst case cold environment drives the minimum latching mechanism torque  margin.  Figure 20. Actuator Speed versus Torque Performance at Various Temperatures  Dynamometer testing of a prototype actuator demonstrated a slight improvement in actuator output torque  performance in the worst case cold condition by adding a few drops of Brayco 815Z to the Braycote 601  grease on the internal motor ball bearings when compared to the same bearings with only Braycote 601  added (see Figure 21). This lubrication approach is being used for both the SWOT and NISAR Projects.   00.10.20.30.40.50.60.7 0 20 40 60 80 100 120 140 160 180 200Speed (RPM) Torque (in•lbf)Standard Dual Drive Actuator Output Speed vs Torque at Various Temperatures -50ᵒC -40ᵒC -30ᵒC -20ᵒC 0ᵒC 20ᵒC 40ᵒC 75ᵒC NASA/CP—2018-219887 415
Figure 21. Actuator Performance with and without Brayco 815Z added to Braycote 601 on Motor Bearings  Flexured Hook and Roller Assembly  The flexured hook and roller assembly is designed to achieve the target joint preload of 1641 N (369 lbf)  while requiring the minimum peak torque to latch within available volume constraints. A modified parabolic  hook ramp profile was designed to maintain nearly constant torque load during the latching process,  minimizing peak torque required to approximately 5.65 N•m (50 in•lbf) and thereby maximizing mechanism  minimum torque margin.   The significant deflection of the flexures in the assembly of approximately 2.03 mm (0.080 in) under preload  ensures that there is not significant change in preload due to mismatch in the coefficients of thermal  expansion between latch component materials and alignment critical structure materials. A three-blade oneaxis flexure was designed to achieve the required stiffness and stroke in the limited volume available.   Conclusions  The following key mechanism conclusions have been drawn based on the NISAR and SWOT RF Boom  Hinge deploy and latch mechanism developmental testing and analysis results referenced in this document:  1. Mechanical decoupling of mechanism hardware from alignment critical structures is a robust and effective approach to avoid mechanism-induced thermal distortions. 2. High-aspect-ratio rectangular wire torsion spring arms must be adequately supported to prevent undesirable twist along the wire axis when used at significant angles of deflection. 3. The new spring/piston viscous damper design described in this document provides performance benefits over heritage elastomeric diaphragm dampers. 4. The addition of Brayco 815Z to Braycote 601 lubricant provides measurable improvement in cold temperature ball bearing performance. References  1. NASA-STD-5017A, “Design and Development Requirements for Mechanisms”, 2015-7-31, 62. 2. Stewart, Alphonso, Charles Powers, and Ron Lyons. “Improvements for Rotary Viscous Dampers Used in Spacecraft Deployment Mechanisms.” Proceedings of the 32nd Aerospace Mechanism Symposium, 115-124. 3. Harper, Jack. “Viscous Rotary Vane Actuator/Damper.” Proceedings of the 10th Aerospace Mechanisms Symposium, 198-207. Acknowledgement  The research described above was carried out at the Jet Propulsion Laboratory, California Institute of  Technology and at D.E.B. Manufacturing, Inc. under a contract with the National Aeronautics and Space  Administration. © 2018. All rights reserved. 00.050.10.150.20.25 0 20 40 60 80 100 120Output Speed (RPM) Output Torque (in•lbf)Standard Dual Drive Actuator Output Speed vs Torque -30ᵒC, With/WIthout Brayco 815Z Added to Motor Brearings Brayco 815Z Added to Motor Bearings No Brayco 815Z Added to Motor Bearings NASA/CP—2018-219887 416
Figure 1. Deployment precision degrees of freedom  Deployment precision measurements of the JWST Development Optical Telescope Assembly (DOTA)  demonstrated that the structure has 20 µm of piston precision error, 30 µm of axial precision error and 44  µrad of pitch precision error.10 JWST has been designed with a complex metrology system and 7 degreesof-freedom of actuation for each primary mirror segment so that the deployment position error of the optical  components only needs to be a few millimeters. Placing the precision alignment burden on the metrology  system and shape correcting actuators leads to a more complicated and heavy telescope system than  would be necessary with a higher precision structural deployment. Other precision deployable mechanisms  and frame concepts have been developed and tested for deployment precision. A summary of previous  deployment precision measurements for optical deployable hinges and structures is shown in Table 1. As  can be seen from these data, while micron-level deployment precision has been measured, only one  measurement of angular precision has been reported in the literature. To better understand the  performance of optical precision deployable structures in all of these degrees-of-freedom and help mature  the technology for future missions, more development is needed.  Table 1. Previous optical deployable structure deployment precision measurements  Test Article Piston (µm) Axial (µm) Roll (µrad) Pitch (µrad)  JWST DOTA1020 30 - 44 IFH Specimen 2120.85 0.31 - - EDSM112.97 2.97 - - Lidar Prototype118.59 0.81 - - MIT Lincoln Laboratory (MITLL) is currently working on technologies to simplify deployable space  telescopes including precision deployable structures, precision actuation and image based metrology  techniques. In this paper, we present recent technology developments in the area of precision deployable  structures. First, deployment precision measurements for a newly developed hinge mechanism using thin,  flexible composite materials are presented. Second, the design of a prototype deployable frame using this  new hinge mechanism is presented along with deployment precision testing results for this deployable  frame.  High-Strain Composite Hinges  High-Strain Composite (HSC) materials are thin, flexible composite laminates that can undergo strains  much larger than metals with very little plastic deformation. This enables new ways of rolling or folding large  structures without the use of traditional revolute joints. Furthermore, HSC-based mechanisms can serve as  NASA/CP—2018-219887 418
parallelism between the tip and base. This ensured that a consistent tape free edge length was achieved  for all of the test specimens  Table 2. Test specimen materials and layups  Specimen Mandrel  Diameter Material Angles (deg) Total Ply  Count  4-1 10.2 cm IM7/PMT-F7 PW (a), IM10/PMT-F7 UD (b) 45(a)/0(b)/45(a) 3  4-2 IM7/PMT-F7 PW (a), IM10/PMT-F7 UD (b) 45(a)/0(b)/0(b)/45(a) 4  4-3 IM10/PMT-F7 PW 45/45 2  4-4 IM10/PMT-F7 PW 45/0/45 3  4-5 AstroQuartz (a), IM10/PMT-F7 UD (b) 45(a)/0(b)/45(a) 3  4-6 AstroQuartz (a), IM10/PMT-F7 UD (b) 45(a)/0(b)/0(b)/45(a) 3  6-1 15.2 cm IM7/PMT-F7 PW (a), IM10/PMT-F7 UD (b) 45(a)/0(b)/45(a) 3  6-2 IM7/PMT-F7 PW (a), IM10/PMT-F7 UD (b) 45(a)/0(b)/0(b)/45(a) 4  6-3 IM10/PMT-F7 PW 45/45 2  6-4 IM10/PMT-F7 PW 45/0/45 3  6-5 AstroQuartz (a), IM10/PMT-F7 UD (b) 45(a)/0(b)/45(a) 3  6-6 AstroQuartz (a), IM10/PMT-F7 UD (b) 45(a)/0(b)/0(b)/45(a) 3  HSC Hinge Deployment Precision Testing  When HSC hinges are folded, they are under higher strain than they were in the deployed state. Under  stress, the fibers in the composite can shift slightly within the matrix. In the folded position, the hinge is  under stress and the fiber shifting results in a relaxation of the stress. This phenomenon is known as stress  relaxation. When unfolded, over time the fibers shift back some amount to a new minimum stress state in  a process known as creep recovery. In this new minimum stress state, the hinge has slightly shifted. The  shape change is the deployed position error. When this fold and deploy process is repeated, the standard  deviation of the test-to-test shape change is referred to as the deployment precision of the hinge in this  paper.  The creep recovery behavior of the hinge specimens is related to the specimens’ material properties,  magnitude of applied strain, duration of applied strain and temperature. The current tests were designed to  focus on the effect of stow duration and strain level on creep recovery for a set of composite layups. The  applied strain level is a function of the stowed fold angle, the specimen geometry and the specimen material.  Therefore, the only metrics varied in the test were stow duration, specimen material and specimen  geometry. The folded hinges were strained by folding to a set angle and clamped in place, as shown in  Figure 3.  Figure 3. Hinge in deployment precision measurement fixture deployed (left) and stowed (right)  NASA/CP—2018-219887 420