ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
Influences to the global stiffness of ULMAAS were assessed parametrically on simplified models with  variation of free length of the springs, spring curvature as well as the size and shape of the fixation area. The change of the length of one spring will affect the general bent shape, stiffness, buckling strength and both motorization torque and end of deployment shock. Deployment Torque Analysis The folding and deployment of ULMAAS is a geometrically non-linear problem. The anisotropic material  properties of the composite make the FE calculation very complex. Convergence is sensitive both to the  material properties and the mesh. The corresponding motorization torque of the hinge was calculated  with an implicit code for a single pair (Figure 5b) of spring elements and with an explicit code for the  whole hinge (Figure 5c). In both cases, elastic instability points occur. In this regard, the calculation using ANSYS software (Figure 5b) was very challenging. The hinge model has a large rotational stiffness that is obtained by the chosen geometry of springs as  cylindrical sectors and the relative large distance between the spring rows. To come into regions of large bending rotations along with small changes for the rotational moment at the antenna side, one has to overcome a snap-through-instability of the longer spring. In order to match real behavior observed in  single spring models, the simulation used an arbitrary disturbance by applying a local displacement to the center of the long spring (Spring A in Figure 4) at the beginning of the movement. In real tests, a preflattening of the springs has to be done to ensure folding without damage to the tape so that essentially  the same disturbance as required for the mathematical models is necessary.  Most interesting in the simulation is the transition of the long spring behavior starting from an almost  symmetrical deformation state (three bending zones) to the final S-shape (two bending zones). It can be seen that the moment versus deployment angle curve in Figure 6 reaches a minimum, which can be identified as a snapping through between two qualitatively distinct deformation states (elastic instability).  angle [rad] moment  [N*mm]  Figure 6. ANSYS simulation – moment vs. angle  226
The moment versus angle relationship could be determined during the unfolding process. To do so, all  auxiliary forces or displacements were removed. But the unfolding analysis did not fully converge to the straight position. Furthermore the implicit code was very sensitive to material stiffness properties. A simulation with PamCrash (Figure 5c), which is an explicit solver, was more stable to model variations and allowed introduction of anisotropic material properties. The effects of coupling in the stiffness matrix for the anisotropic material were clearly visible in the out of plane behavior of the springs seen in Figure  7. The disadvantage one has to face using the explicit code is the rather long solution time (days rather  than hours), which limits iterations and variations in the parameters of the model. Both simulation and test showed that end of deployment shocks are quite high. A reduction of torque  cannot be achieved without reducing the deployed stiffness and although certain margin exists in this  regard (3.4 Hz vs. a 2 Hz requirement) it was decided to implement a damping system.  Figure 7. PamCrash simulations – hinge deployment  Thermal Investigations The objective was to predict thermo-mechanical behavior of the hinge. Using the existing FEM, an orbit  simulation with the I-DEAS TMG module was performed. Previous calculations revealed that even with low CTE materials an additional thermal insulation is necessary to meet the stringent requirement of pointing accuracy. A multi-layer insulation (MLI) was included in the simulation where a geosynchronous orbit was simulated. The results (Figure 8) showed a moderate temperature gradient within the hinge. Hot and cold spots appeared at the interface points. Temperature extremes were in the range of -60°C to  +20°C and corresponding deflections are moderate and stay within the pointing budget of the mechanism. Material Testing  The material testing was performed in order to obtain property values and re-run the analysis with them.  Tests were performed in the two orthogonal directions. The tests of the CFRP material covered included  following measurements:  - Density was measured to obtain the exact mass of the model in the simulation.  - Glass transition temperature  - Thickness allows assessing the manufacturing quality. Furthermore, the bending radius depends on  the thickness.  227
- Fiber, matrix and void fraction reveal a lot about the manufacturing quality but are not really  necessary for the analysis because strain/stress tests were also performed.  - Young’s modulus, maximum strain and ultimate stress in tension and compression at reference  temperature, minimum and maximum temperatures. - Coefficient of thermal expansion (CTE) tests are very important as one of the most stringent  requirements refer to pointing stability, which is driven mostly by the thermal characteristics of the  material. - Long-term stress relaxation to assess the change of deployment torque over time due to storage of  the hinge in folded configuration for a long period of time.  - Moisture pick-up  - Thermal conductivity  Figure 8. Thermo mechanical simulation – Thermal Gradients  Figure 9 shows the test machine used for the test and the failure mode in compression of a sample.  Figure 9. Test set-up and destroyed compression sample  For the long-term creep test, a pair of short tape springs were glued to fittings and then bent and mounted  to a test jig as in Figure 10. Accelerated testing at 60°C was performed in parallel to room temperature  tests.  228
Figure 10. Test Jig for creep test  Bread Board and Performance Testing  Bread Board Testing Basic investigations in terms of CFRP bendability, handling and component interface were performed  with a full-scale breadboard model as shown in Figure 11. The observations made were used to enhance the design of the hinge and identify weak points. Figure 11. Full-scale breadboard model  Several deployments were performed with the test gadget and an inertia dummy attached to observe the  unbuckling and latching of the springs. The breadboard made it easy to exchange the tape springs. This  allowed replacing tape springs quickly and made it possible to test different lay-ups. The different lay-ups  were assessed qualitatively regarding their robustness against defects. Figure 12 shows a side and a top view of the test gadget with a dummy inertia attached.  Performance Testing: Motorization The laminates are by their very nature asymmetric. To compensate for coupling effects, they were  mounted pair-wise in the test gadget. The test gadget has the same dimensions as the hinge except for  the width, which had been halved. The test rig is depicted in Figure 13. The measurement determined the  angle-moment relationship of the hinge and thus the motorization of the ECH. The test was performed by applying equal torque to both interfaces. The resultant angle was measured from 90° to approximately 5°.  229
Smaller angle a higher torque than the test rig could withstand, so the peak moments close to 0° could  not be determined during these tests. Figure 12. ECH and dummy inertia   Figure 13. Test rig during static tests  Figure 14 shows the results from 5° to 90° for different lay-ups. It shows that the motorization depends  mostly on the laminate thickness. In contrast to this, the spring behavior is driven mostly by the lay-up identifiable in the graph between 15° and 40° where the buckling shape of the long tape spring changes.  Furthermore, the test revealed the robustness and sensitivity to bending of the specimens. Therefore, the  selection of appropriate lay-up was not only based on the numerical data but also on the observations  recorded. 230
00.511.522.533.54 0 1 02 03 04 05 06 07 08 09 0 Rotation [deg]Moment [Nm][0,+45,-45,90] and [0,-45,+45,90] [0,+60,-60] and [0,-60,+60] [0,+45,-45] and [0,-45,+45] [90,+45,-45,0] and [90,-45,+45,0] Figure 14. Quasi static moment vs. angle   Performance Testing: Stiffness The stiffness and the dynamic test required a different test jig. The test facility used a jig to mount the  hinge, a dumbbell shaped dummy inertia to simulate the spacecraft appendage, a video camera to record  the dynamic behavior of the hinge during deployment and laser interferometers to monitor the static  deflection under a given load and the pointing accuracy after every deployment. Under the mounting  plate, hinge force transducers measured the latching shock. The set-up is shown in Figure 15.  To determine the bending stiffness of the hinge, weights were attached at both ends of the dummy  inertia. This led to a pure bending moment about the hinge axis. The distortion due to the applied load was monitored by laser interferometers pointing on a target fixed to the hinge. As the displacement is very small, it was assumed that the distance from interferometer to target is  constant at all time. A detailed view of the apparatus is given in Figure 16.  231
Figure 16. Test setup for stiffness measurement  Figure 17 shows that the deployed stiffness is mostly driven by geometric parameters (curvature, tape  spring width, but not thickness) rather than by the lay-up. Except for the thickness, all tested items had  identical dimensions. The lay-up is decisive for the motorization torque but not for the deployed stiffness.  0246810 0 0.02 0.04 0.06 0.08 0.1 0.12 0.14 Deflection [°]Moment [Nm][0,+60,-60] and [0,-60,+60], before deployment [0,+60,-60] and [0,-60,+60], after deployment [+/-45,0/90,0/90,+/-45], before deployment [+/-45,0/90,0/90,+/-45], after deployment Figure 17. Stiffness of the hinge  Performance Testing: Dynamic and Latching Shock The setup proved to be difficult, as the non-symmetric lay-ups tend to deform sideward when bent and  the clearance to the side walls of the rig had to be adjusted carefully. The force recording, which is depicted in Figure 18, showed that most of the travel of the appendage was accelerated with low spring torque, which corresponds to the quasi static measurements. A notable first peak can be interpreted as the huge stiffness increase of the tape springs close to zero degree deflection. A second peak, which was higher and in negative direction resulting from overtravel, is the one to be regarded by the maximum  233
shock criteria of the performance specification. The decay of the vibration and the periods of the  amplitudes allow an assessment of structural damping (15-20%) and natural frequency (3.85Hz).  -140-120-100-80-60-40-20020406080 0123456789 1 0 time (s)root moment (Nm)1st peak  2nd peak  Figure 18. Dynamic moment vs. time  The final lay-up design is based on to the best variant in sample testing either [0, 60,-60] or its mirrored  variant [0,-60, 60]. The thickness of the composite is approximately 0.33 mm. The support parts that are bridging the different length of the springs are made from the same laminate with a thickness of 2.5 mm. Conclusions The final design of ULMAAS, a light weight self-motorized deployment mechanism, was achieved through  iterative analysis and testing. The first analytical predictions led to a design with a very high deployment torque, which then had to be mitigated. Use of dedicated tests showed that the lay-up of the composite could be fine-tuned to reduce this torque without a substantial reduction in deployed stiffness. The analysis approach shows that prediction of characteristics is possible even for complex non-linear problems but in this case it goes hand in hand with testing, as the traditional approach of validating an analysis with a test will not work.  References 1. Nicolas Elie, Alain Lacombe, Stéphane Baril, Ultra-light reflectors: A high performance and industrial  concept for commercial telecom antennas , EADS SPACE Transportation  2. Sicre 1, Givois2, Emerit2,Application of “Maeva” Hinge to Myriade Microsatellites Deployments Needs , CNES1, Metravib2 3. Pellegrino S., et al, IUTAM-IASS Symposium on Deployable Structures: Theory and Applications , Drodrecht: Kluwer Academic Publishers, 2000  4. Pellegrino, Yee, Foldable Composite Structures , University of Cambridge, UK, 2003  234
Development of the Aquarius Antenna Deployment Mechanisms and  Spring/Damper Actuator  Joel A Johnson* Abstract  The Aquarius Instrument’s large radar reflector dish needed to be stowed for launch, and then deployed  on-orbit. The Deployment Subsystem consisted of a cantilevered boom structure and two single-axis  hinge mechanisms to accurately deplo y and position the reflector dish relative to the radar feed horns.  The cantilevered design demand ed high stiffness and accuracy from the deployment mechanism at the  root of the boom. A preload-generating end-of-travel latch was also required. To largely eliminate the  need for control systems, each deployment mechanism was actuated by a passive spring motor with viscous-fluid damping. Tough requirements and adaptation of a heritage actuator to the new application  resulted in numerous challenges. Fabrication, assembly, and testing encountered additional problems,  though ultimately the system was demonstrated very successfully. This paper revisits the development to  highlight which design concepts worked and the many important lessons learned.  Introduction  Aquarius Project Aquarius is the primary, NASA-funded instrument aboard SAC-D, a service platform designed, built, and  operated by the Argentinean Comision Nacional De Actividades Espaciales (CONAE). Launched in 2010  with a 3-year earth-orbiting science mission, Aquarius’s primary objective is to map sea-surface salinity to  better understand the interactions between the global water cycle, ocean circulation, and climate (salinity affects seawater density, which governs ocean circulation). Aquarius’s sensitive L-band radiometers  measure ocean brightness temperature while a scatterometer corrects for ocean roughness effects. The  channels share a 2.5 m x 2.8 m composite Reflector cantilevered from the Instrument Primary Structure.  The Reflector and support Boom must  be folded to fit within the 3m fairing of a Delta II (Fig. 1). The  Reflector pivots about a single-axis hinge-line bet ween it and the Boom, performed by the Upper  Deployment Mechanism (UDM). Similarly, the 2 m long Boom pivots at its root via the Lower Deployment  Mechanism (LDM). Once stowed, the Reflector is restrained against the Primary Structure via two pyrotechnic Separation Nuts. Upon reaching orbit, deployment occurs in two stages to prevent the  Reflector from dragging across the launch restraints (Fig 1). Staging is provided by a pyrotechnic Pin  Puller at the LDM, which carries no launch loads. Initial separation of the Reflector is assisted by Kick-off Springs. At the end of its deployment, each Deployment Mechanism latches into final position.  Figure 1. SAC-D Launch Configuration and Aquarius Deployment Sequence  *Jet Propulsion Laboratory, California Institute of Technology, Pasadena, CA  Proceedings of the 39th Aerospace Mechanisms Symposium, NASA Marshall Space Flight Center, May 7-9, 2008. Pin Puller firing  initiates LDM  Deployment and  Latchin g STOWED Reflector Sep Nut firing  initiates   UDM Deployment  and Latching UDM SAC-DLDMFeed Horns Boom Aquarius  DEPLOYED 235
Together, the Deployment Mechanisms and the Boom comprise the Deployment Subsystem. The overall  concept had been accepted at PDR, with actuator selection being an important aspect of the design.  “Passive” constant-torque spring motors [1], rate-limited with viscous-fluid rotary dampers, had been chosen to eliminate the need for motor control electronics [2]. JPL’s Spring/Damper Actuator had heritage  from Mariner, Viking, Galileo, NSCAT and other missions, but hadn’t been flown in over a decade [3].  The author was tasked with developing the Deployment Mechanisms from this post-PDR level, to  overseeing delivery of a single protoflight Mechanism/Boom assembly. The work was carried out at the  Jet Propulsion Laboratory, California Institute of Technology, under a contract with the National Aeronautics and Space Administration. The ensuing phases of Design, Fabrication, Assembly, and Test  all came with their unique set of challenges, succe sses, failures, and lessons learned.   Design Requirements The requirements that significantly drove  design were as follows [4]:  1. The 1st mode of the deployed Reflector shall be greater than 2 Hz (originally 3 Hz)  2. The Reflector I/F shall be positioned relative to the Primary Structure I/F in the nominal  deployed location within the following tolerance s: (Reflector pointing alignment)  a. As-delivered: ± 0.1 degree, ± 0.25 mm b. Deployment repeatability: ± 0.05 degree, ± 1 mm  c. Mechanical jitter: ± 0.005 degree, ± 0.25 mm  3. Temperature requirements: (Protoflight/design temps)  a. UDM Bearings: -75/+45°C Op, -95/+25°C NonOp  b. Damper: -5/+45°C Op, -50/+60°C NonOp  Requirements 1 and 2c meant that the deployed system needed to be extremely rigid. The rotational  inertia of the deployed Boom/UDM/Reflector about the LDM was approximately 200 kg*m 2. An LDM  rotational stiffness of about 340,000 N*m/rad (allocation) would be necessary to meet the original 3 Hz  requirement. At the same time, the system was tightly mass constrained. This presented a significant  challenge, taking months of iteration and optimization to try to develop an acceptable Hinge/Latch design.  When it started looking like it might not be possible, the Project relaxed the requirement to 2 Hz with  surprisingly little resistance. The lesson learned was that before spending too much time trying to meet a  requirement, find out how rigid (!) it is. Lower/Upper Deployment Mechanisms Each Deployment Mechanism was comprised of the Hinge (Axle/Clevis/Bearings), the Latch, the  Spring/Damper Actuator, the Cable Spo ol, and Telemetry Switches (Fig. 2). The LDM also included a Pin  Puller for staging.  Boom Lower End  FittingDEPLOYEDSTOWEDSpring/Damper  Actuator Figure 2. Lower Deployment Mechanism (LDM) Axle  Clevis  Pin PullerCable SpoolTelemetry  Switches  Latch  Assy  236
Hinge nge consisted of an Axle supported on bearings mounted in a structure called the Clevis (Fig. 3).  he PDR design consisted of a hinge-line created by two angular contact bearings preloaded across the  he use of duplex pairs meant the necessary stiffness and load capacity could be achieved in a bearing  he philosophy of minimizing interfaces was used throughout the design. This became especially The Hi The Spring/Damper deployed the Axle approximately 60 degr ees where it hit a Hardstop and a Latch  engaged. The force couple between the Bearings and the Hardstop/Latch resisted motion in the deployment direction. Ultimately, the system performance was driven by the stiffness of the structural  connection between these two elements, as well as the coupling of this stiffness to the mating structures.  Figure 3. Hinge  T entire mechanism. It had been noted that thermal differentials within the Hinge would make it difficult to  maintain the bearing preload, and it was recommended at PDR to switch to a back-to-back duplexed (DB) pair at one end, and a floating radial bearing at the other. Unfortunately, due to the stiffness and pointing  requirements, it became impractical to implement a bearing that floated axially. Instead, duplex pairs were  installed at each end. To minimize the over-constraints, face-to-face (DF) pairs were used, and a careful procedure of axial shimming was implemented. Temperature effects also had to be carefully considered.  T with nearly half the OD, which would be necessary to keep from doubling the drag. As a result, the Hinge  was reduced by about 40% in all three dimensions. However, the Hardstop/Latch was maintained at as great a distance from the hinge-line as practical, since latched stiffness would increase as roughly the  square of this distance. Effectively tying all these elements together became crucial for high stiffness,  resulting in a very compact design. As such, both the Hardstop and the Latch were made as integral with the Hinge as possible to reduce the structural inefficiencies associated with mechanical interfaces.  T important when it came to how the Hinge was to be assembled. Parting the Clevis for Axle installation  would have resulted in a massive, heavily bolted seam, as well as registration features and special  fabrication steps for bearing alignment. Parting the bearing hou sings would introduce the same issues, as  well as uneven bearing support. Instead, the Axle was parted in such a way as to have negligible effect  on Hinge stiffness, with none of the other complications (Fig. 3). Radial load on the Stub results in a shear  that is carried through the Stub’s press fit with the Axle, and a moment that is supported by the toe-heel  reaction from the axial Bearing clamping. It also turned out that the clamping could be athermalized  through simple selection of fastener length. The Stub Axle design, though initially viewed as  unconventional, was very successful in the end.   Axle (Ti) Clevis (Ti) DF Bearing Pairs  Spring/Damper (52100 Steel)  ShimStub (Ti) Outer Race  Clamp/Cover  237
Latch The Latch proposed at PDR was a hasp design like that used on Magellan, NSCAT, and MER. As the  Hinge was being designed, this type of latch was abandoned in favor of a wedging pawl latch (Fig. 4),  based on ability to create preload, stiffness, capture range, and form factor. The Pawl began engaging  one degree before the Axle reached  the Hardstop, eliminating the potential fo r rebounding and helping to  “drive it home”. The Latch Depressor was implemented so that the work of “cocking” the Latch was done  in the stowing process, and wouldn’t impede deployment. This al so simplified analyses and reduced test  cases. The Pawl was actuated by a redundant pair of nested helical compression springs, through a  simple linkage. The linkage was near its knuckle point in the stowed position, to minimize force and drag  on the Latch Depressor, with a snubber to prevent linkage over-center. By the same linkage, the torque applied to the Pawl increased to a maximum at the nominal latched position. A very detailed analysis was  performed to calculate the preload force between the Axle and Hardstop generated by the Latch (Fig. 5),  which was needed for maximum stiffness. The design resulted in only a 13% preload variability over a  large (0.5 mm) range of capture. The Hardstop an d Latch Strikeplates could be shimmed for adjustment  of the Axle’s deployed position, and to center the Latch in the middle of its capture range. The design also  had to account for disengaging the Latch after a test, requiring tooling to be developed at the same time.  Figure 4. Latch  050100150200250300350400450 0246810121416182022242628303234 Latch Angle [deg]Hardstop Preload [N]μ=0.1 μ=0.2 μ=0.4 Nominal Latching Range Nominal Latch Position Latch Spring ForceSTOWEDSnubber  Spring Canister  Pawl Hardstop Latch Depressor DEPLOYED Strikeplates  Figure 5. Hardstop Preload due to Latch, for various coefficients of friction  239