ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
85   Figure 3. TO Isolator Strut Cross-Sectional View  

During assembly, each spring is pulled back 4.123 cm (1.625 in) from its free length. The reactive force of this tensile load on the spring compresses the Center Rod onto the front Delrin® pad and preloads the assembly. During launch, when a large enough tension load is applied to overcome the preload, the center rod will unseat. After unseating as the spring extends further under still higher tensile launch loads, the relative motion of the Center Rod may travel 1.59 cm (0.625 in) until it hits another Delrin® pad stop on the Flanged Sleeve. The Delrin® pads were used so that titanium-on-titanium contact would not occur during impacts, preventing galling of the contacting parts. The expanded diameter head on the front end of the Center Rod limits its motion to this 1.59 cm (0.625 in) travel space. The Ares first stage forcing function has been computed to apply tensile loads resulting in head travel within this 1.59 cm (0.625 in) range, effectively isolating the pallet from the rest of the crew module in the X direction.    All of the load bearing machined parts and the machined spring are made out of titanium to reduce the weight of the assemblies. Each of the machined springs were machined from a single piece of titanium and have identical bolted interfaces on both ends. Due to the method of construction and the bolted interfaces, the springs can react both compressive and tensile loads. The 105 kN/m (600 lbf/in) Isolator spring assemblies each have a mass of 5.08 kg (11.2 lbm) and the 210.2 kN/m (1200 lbf/in) Isolator spring assemblies each have a mass of 6.21 kg (13.7 lbm). The total mass for all four Isolator spring assemblies is 22.6 kg (49.8 lbm) which does not include the existing CM portion of each strut.   Damping  To provide for the low-level damping required to maintain stability in a LA scenario, the motion of the Isolator requires a damping component. Due to the late addition into the design, a passive, friction-based damping system was devised. Two Delrin® components clamp on the Center Rod; fasteners squeeze the components together, providing a friction force against the Center Rod that can be controlled based on the fastener torque. They are designed to provide between 111 – 156 N (25 – 35 lbf) of frictional force.  Lockout Mechanism  In order to satisfy the locking requirement for landing loads, a feature was added to prevent any stroking of the Isolators. For ease of prototype testing and assembly, quick release (insertion) pins were inserted behind the head of the Center Rod preventing it from moving and creating a rigid load path that bypasses the machined spring.  For flight, a quick-reacting lock out mechanism would be required during landing. A concept for a Non-Explosive Actuator (NEA) driven blade was added to the design. The concept uses an NEA to hold a spring-loaded Plunger that is threaded into the NEA nut; when actuated by a 4-amp electrical signal, the NEA releases the Plunger which wedges behind the Center Rod, securing the Rod in place and creating
86 a single shear load path that bypasses the spring creating a strut stiffness equal to that of the CM  baseline strut. A second blade 180° from the first would be needed for redundancy. The concept used for testing incorporated an NEA mockup on one of the four Isolators to characterize its static performance. Refer to Figures 4 and 5 for images of the Isolator and NEA Lockout Mechanism.  

Figure 4. TO Isolator Strut with NEA Lockout Mechanism  

Figure 5. NEA Lockout Mechanism Cross-Sectional View      Test and Evaluation Program 

The test and evaluation program aimed to characterize the effectiveness of the Isolation springs in attenuating the launch dynamic accelerations and loads imparted to the crew pallet. In addition, the test program aimed to understand the effect the Isolators had on the wire bender stroke during landing impact and better understand the correlation of model predictions to test results.    The test program included: Isolator Proof Testing, Isolator Stiffness and Damping Characteristics, Strut Impact Testing, System Drop Testing, Modal Testing of the launch configuration, and Vibration Testing to characterize performance.      NEA Lockout  Mechanism (x2)  Plunger Plunger drops  behind Center Rod  flange, locking Center Rod and  directing load path  through internal  components Internal nut  separates  when NEA actuated, releasing Plunger NEA Center  Rod Isolation  Spring Coulomb  Damper
87 TO Isolator Proof Testing / Stiffness and Damping Characteristics  Each Isolator assembly was tested in tension (locked and unlocked) and in compression to prove strength  margins and verify spring stiffnesses. Maximum loading expected on the strut occurs during drop testing and is 4.448 kN (10,000 lbf). A test factor of 1.25 was included in proof tests leading to testing in compression and tension at 5.560 kN (12,500 lbf). All proof tests were performed successfully with no signs of failure or yielding. Spring stiffnesses and friction forces were found to be in acceptable ranges.  Strut Sub-System Impact Testing  Impact testing aimed to characterize TO vibration isolation springs in a dynamic environment and assess their effect on the behavior of the wire bender landing attenuation struts. In addition, an accurate LS-DYNA® model of the vibration isolation spring and attenuation strut was desired so that a reliable analytical model is available for subsequent landing simulations performed with the Orion vehicle under simulated landing conditions. 

The tests were conducted using the 711th Human Performance Wing’s Vertical Deceleration Tower (VDT)  at Wright Patterson Air Force Base in Dayton, Ohio. The VDT consists of an 18 m (60 ft) vertical steel tower that allows a carriage to enter a free-fall state (guided by rails) from a pre-determined drop height. The plunger mounted on the rear of the carriage is guided into the hydraulic deceleration device (cylinder filled with water located at the base and between the vertical rails), producing an impact deceleration pulse, see Figure 6.   

       Figure 6. Attenuation Struts in VDT Setup  

The testing showed that the transmitted acceleration peaks are similar regardless of the presence of the isolation spring. Testing and simulations for the strut with both the TO vibration isolation spring and the wire bender incorporated into the strut showed that the isolation spring has the effect of reducing the Brinkley injury criteria at the expense of increasing the strut stroke.   

System Drop Testing  Orion CIAS impact tests were performed at the NASA Langley Landing and Impact Dynamic Research Facility to provide a demonstration of strut system performance, evaluate the performance differences  VDT Carriage  with  Strut  Wire Bender Isolator
89     Figure 7. Attenuation Strut in Assembly (Left); Drop-Test Fixture (Right)  

System Modal Testing  As part of the pre-test planning for the base-drive vibration testing of the CIAS isolation system, a series of modal tests were performed at NASA Langley to assess the pre-test finite element model (FEM) predictions. The objectives of the modal test were: (1) to investigate potential fixture modes in the frequency range of the vibration test (0 - 20 Hz); and (2) to verify FEM predicted modes for the isolated crew pallet.   Two Isolator test configurations were evaluated: (1) locked and (2) unlocked. In the locked configuration, the isolated struts perform like rigid elements. For the unlocked configuration, the Isolators were shimmed to position with the Isolators in their active stroke range. Accelerometers measured the Frequency Response Functions (FRFs), calculated as the ratio of the acceleration response to the input force. Modal parameters (natural frequencies, damping factors, and mode shapes) were then estimated from the FRFs. Base-drive data acquired during the vibration testing was also used to obtain modal estimates with and without the friction dampers, see Figure 8.    

Figure 8. LS-DYNA® Model (Left); System-Level CM Mockup (Right)    Due to low-level inputs, at certain times it was difficult to overcome the friction in the damper, which created difficulties in test execution and data collection. Eventually the friction dampers were removed, allowing for more ideal performance of the Isolators; as a result all three pallet modes were identified.  Crew  Pallet  Mass Mockup Wire  Bender &  Isolator Strut  Crushable  Honeycomb Support Structure CEV CM  Structure  Mockup
90 There is good agreement between the measured frequency estimates and FEM predictions with the  largest frequency difference at 8% for the first mode. Modal results are presented in Table 1.    Table 1. Modal Estimates for the CIAS Pallet Modes   Mode Predicted  (Hz) Modal with   friction dampers Base-Drive with friction  dampers Base-Drive without  friction dampers  Freq (Hz) Damp  (%) Freq (Hz) Damp  (%) Freq  (Hz) Damp  (%)  Rocking-Y:+Z  struts 3.77 4.1 8.8 - - 4.1 3.2  Twist about X 4.37 - - 4.5 7.9 4.7 4.1  Rocking-Y: -Z  struts 5.35 5.3 6.0 5.3 7.1 5.4 3.5   System Vibration  The objective of the system-level vibration test was to demonstrate the effectiveness of the pallet-to-strut isolation springs for reducing crew loads during a TO event occurring at approximately 12 Hz. The Orion wire bender Vibration Testing Unit (VTU) was subjected to dwell tests at frequencies of 10 and 12 Hz with varying amplitudes to represent a range of possible launch oscillation loads. Testing occurred at the Naval Surface Warfare Center (NSWC) Dahlgren Division Vibration Test Facility. See Figure 9 for a picture of the test setup.  The test consisted of three configurations:    Test Sequence 1: rigid, Isolators locked out   Test Sequence 2: Isolators released, friction dampers installed   Test Sequence 3: Isolators released, friction dampers removed 

Each configuration was subjected to three sine sweeps at varying input levels to test linearity. After the  sine sweeps, each configuration underwent dwell testing at the TO frequencies with three different input levels: 0.2, 0.35, 0.5g.   Conducting the signature sine sweeps verified the linearity of the system for Test Sequences 1 and 2. Because dampers were removed for Test Sequence 3, acceleration limits were reached as the pallet isolation frequency coupled with the input. As a result, only one successful sine sweep at 0.025g was completed over the range 2 – 8 Hz for Test Sequence 3. The test was modified to only sweep across 8 – 20Hz for levels higher than 0.025g (0.075 and 0.1g) which verified the linearity of the system over the range 8 – 20 Hz. However, the linearity of Test Sequence 3 over the range 2-8 Hz cannot be verified. Any slight deviation from linearity in Test Sequence 1, Test Sequence 2, and Test Sequence 3 (8-20 Hz only) can be attributed to the noise introduced by the hydraulic shakers.   Figures 10 and 11 display the reduction in accelerations attributed to the Isolators at the 10 Hz dwell and 12 Hz dwell. Tables 2 and 3 display the reduction in acceleration for the 10 Hz and 12 Hz dwells (note that a negative reduction indicates an increase in acceleration).
91   Figure 9. Orion VTU Mounted on Hydraulic Shakers in the NSWC Vibration Facility 

Figure 10. Comparison of Pallet 10 Hz Dwell Test Results for Each Test Configuration 

00.10.20.30.40.50.60.70.8 0.05 0.2 0.35 0.5 0.65Acceleration Output (g) Acceleration Input (g)Accel 1 - TS1 (Rigid) Accel 2 - TS1 (Rigid) Accel 3 - TS1 (Rigid) Accel 4 - TS1 (Rigid) Accel 1 - TS2 (Isolated w/ Damper) Accel 2 - TS2 (Isolatedw/ Damper) Accel 3 - TS2 (Isolatedw/ Damper) Accel 4 - TS2 (Isolatedw/ Damper)10 Hz Dwell Tests Pallet Comparison Plot
92   Figure 11. Comparison of Pallet 12 Hz Dwell Test Results for Each Test Configuration 

Table 2. Average Pallet Accelerations for 10 Hz Dwell Tests   G Input Test  Sequence Input as  Measured  (g) Average  Pallet   Response  (g) Average Pallet  Transmissibility Average Pallet %  Reduction  Compared to  Input* Average Pallet  % Reduction  Compared to  TS1 (Rigid)  0.2 TS1 0.203 0.258 1.27 -27.09 N/A  TS2 0.2 0.092 0.46 54.00 64.34  TS3 0.198 0.06 0.30 69.70 76.74  0.35 TS1 0.348 0.443 1.27 -27.30 N/A  TS2 0.347 0.133 0.38 61.67 69.98  TS3 0.355 0.112 0.32 68.45 74.72  0.5 TS1 0.501 0.642 1.28 -28.14 N/A  TS2 0.506 0.18 0.36 64.43 71.96  TS3 0.497 0.154 0.31 69.01 76.01  *Note: Negative reduction values indicate increase in input.    00.10.20.30.40.50.60.70.80.9 0.05 0.2 0.35 0.5 0.65Acceleration Output (g) Acceleration Input (g)Accel 1 - TS1 (Rigid) Accel 2 - TS1 (Rigid) Accel 3 - TS1 (Rigid) Accel 4 - TS1 (Rigid) Accel 1 - TS2 (Isolated w/ Damper) Accel 2 - TS2 (Isolatedw/ Damper) Accel 3 - TS2 (Isolatedw/ Damper) Accel 4 - TS2 (Isolatedw/ Damper)12 Hz Dwell Tests Pallet
95 Acknowledgements 

The authors would like to thank Chuck Lawrence, Greg Walsh, Steve Mark, Steve Hendricks, Justin  Templeton, and Don Jarosz for their work and support on this project and acknowledge the invaluable contributions from the test teams at Wright Patterson AFB Impact Test facility, Dahlgren Naval Facility, as well as NASA Langley's Landing and Impact Research Facility and Vibration/Modal Laboratory. Their support during this project was essential to its success.  

References 

1. Lawrence, Charles et al.  NASA/TM-2008 215198. The Use of a Vehicle Acceleration Exposure Limit  Model and a Finite Element Crash Test Dummy Model to Evaluate the Risk of Injuries During Orion Crew Module Landings. April 2008.
97 The Damper Spring Unit of the Sentinel 1 Solar Array 

Frans Doejaaren* and Marcel Ellenbroek* 

Abstract 

The Damper Spring Unit (DSU, see Figure 1) has been designed to provide the damping required to control the deployment speed of the spring driven solar array deployment in an ARA Mk3 or FRED based Solar Array in situations where the standard application of a damper at the root-hinge is not feasible. The unit consists of four major parts: a main bracket, an eddy current damper, a spring unit, an actuation pulley which is coupled via Kevlar cables to a synchro-pulley of a hinge. The damper slows down the deployment speed and prevents deployment shocks at deployment completion. The spring unit includes 4 springs which overcome the resistances of the damper and the specific DSU control cable loop. This means it can be added to any spring driven deployment system without major modifications of that system. Engineering models of the Sentinel 1 solar array wing have been built to identify the deployment behavior, and to help to determine the optimal pulley ratios of the solar array and to finalize the DSU design. During the functional tests, the behavior proved to be very sensitive for the alignment of the DSU. This was therefore monitored carefully during the qualification program, especially prior to the TV cold testing. During TV “Cold” testing the measured retarding torque exceeded the max. required value:  284 N-mm versus the required 247 N-mm. Although this requirement was not met, the torque balance analysis shows that the 284 N-mm can be accepted, because the spring unit can provide 1.5 times more torque than required. Some functional tests of the DSU have been performed without the eddy current damper attached. It provided input data for the ADAMS solar array wing model. Simulation of the Sentinel-1 deployment (including DSU) in ADAMS allowed the actual wing deployment tests to be limited in both complexity and number of tests. The DSU for the Sentinel-1 solar array was successfully qualified and the flight models are in production. 

Figure 1:  Damper Spring Unit. 

                                                     * Dutch Space BV, Leiden, The Netherlands 

Proceedings of the 41st Aerospace Mechanisms Symposium, Jet Propulsion Laboratory, May 16-18, 2012
98 Introduction 

The Damper Spring Unit (DSU) is part of the deployment system of the Sentinel 1 Solar Array (SA), an ARA Mk3 product family solar array from Dutch Space. Sentinel-1 is an imaging radar satellite aimed at providing continuous all-weather, day-and-night imagery for monitoring sea ice zones and the arctic environment, surveillance of marine environment and monitoring land surface motion risks, mapping of land surfaces: forest, water and soil, and provide mapping in support of humanitarian aid in crisis situations. The function of the deployment system is to deploy the Sentinel 1 SA wing from the folded panel stack (stowed configuration) mounted on the S/C side wall to the deployed configuration (see Figure 2). The DSU is designed to provide the damping required to control the deployment speed of the spring driven solar array in situations where the standard application of a damper at the root-hinge is not feasible. The large number of power harness cables of the Sentinel 1 SA prevents the location of the damper at the root hinge. This paper gives a description of the damper spring unit and its functioning in the Sentinel 1 deployment system . The qualification program of the DSU is discussed, the torque balance analysis and deployment analysis in ADAMS, which is required to qualify the use of the DSU in the Sentinel 1 deployment system. 

Figure 2: Sentinel 1 solar array wing (stowed and deployed configuration). 

Sentinel 1 

The DSU will be first applied in the Sentinel 1 solar array wing. The wing consists of 6 panels; 5 solar  panels and a yoke panel. It is fixed to the spacecraft via 4 hold down stacks, each containing 2 Thermal Knives to cut the hold down cable inside the hold down stack (see Figure 2 and Figure 3). The deployment system of the Sentinel 1 SA consists of:    Twelve (12) panel hinges per wing (i.e. two hinges per hinge line). Each hinge has a pretensioned spring to deliver the hinge line torque for deployment and a locking device.   A deployment synchronization system which consists of pulleys, guide-blocks and cables to  control the deployment. A system of Kevlar synchronization cables are connected via pulleys to the panel hinge assemblies. These cables synchronize the panel motion between all panels.    The damper-spring-unit, mounted on the yoke panel front side (at sun side in deployed condition  and between yoke panel and spacecraft side wall in stowed condition) and coupled via Kevlar cables to the synchro-pulley of a dedicated DSU hinge at hingeline 2 (see Figure 4).   Kick-off-spring-unit attached on the yoke panel and connected with the root hinge synchronization  pulley to support a compact first stage 90° deployment.
99  

Figure 3: Sentinel 1 stowed solar array wing (side view). 

The Sentinel 1 solar array deviates slightly from standard ARA Mk3 solar array designs mainly in respect with the deployment movement from stowed to the deployed configuration. The root hinge must deploy over 180 degrees, but more important, the trajectories of the C.O.G.’s of all panels have to be confined to a limited area which requires a two stage deployment movement of the panels which is shown in Figure 5. The synchronization system controls these 2 phases in a continuous way avoiding interference with S/C structure. The Sentinel 1 deployment design is a passive system as the traditional deployment system with the same design of hinges, motoring springs, synchro-cables and eddy-current deployment damper. The root hinges pulleys are, however, specifically designed eccentric and the routing of the cables is adapted. The deployment of the panels starts immediately after the last hold down cable has been cut. The deployment motion is driven by the kick-off spring on the yoke panel and the deployment springs mounted at the root hinge and at the panel hinges. The eccentric roll-on / roll-off pulleys on the root hinges realize a gradually increasing synchro-cable movement. The cable payout up to 90 degrees hinge line rotation is small, thus the panels will stay approximately stacked up to about 90º opening of the root hinge line (phase 1). In phase 2 the root hinge opens from about 90 to 180 degrees, while all other hinges open in a synchronized manner up to deployment completion (from about 5 to 180 degrees). This sequence ensures that the CoG travel and MoI shift – and therefore the disturbances on the spacecraft’s attitude - are minimized. At the end of the deployment each individual hinge will latch to assure a stiff deployed configuration.   

Figure 4: The DSU position at the Sentinel 1 solar array.
100 Sentinel 1 Engineering Model 

Using engineering models helped us to identify the behavior of the two stage deployment system and  how the DSU functions within this system (see Figure 5). Due to the severe thermal environment, some flexibility of the synchronization system is required. This flexibility also introduces unwanted non-synchrony. Design modifications are introduced to minimize these effects. As the DSU is not directly integrated in the synchronization system, its pulley can be increased or reduced in diameter. This change can introduce a more favorable ratio between the various pulleys and thus limit the non-synchrony effects of the flexibility. Unfortunately these effects cannot be eliminated completely, which means it has to be accounted for in the deployment analysis of the system. After the final pulley ratio was identified, the DSU design could be finalized. 

Figure 5: The deployment of the Sentinel 1 engineering model.
101 Damper Spring Unit description 

The Damper Spring Unit consists of 4 major parts (see Figure 6):    The main bracket, which supports all other parts of the DSU and allows fixation to the Solar  Array. The main bracket interface with the solar array is dedicated per project. The Sentinel 1 DSU is fixed with a M4 into a blind insert and a M8 bolt which is interfaces with an adapted cup-cone bracket, see Figure 7.   The spring unit, which includes 4 standard ARA Mk3 root hinge actuation springs, to counter the  retarding torque of the specific DSU control cable loop and the eddy current damper start-up torque with sufficient factor of safety.   The actuation pulley, which is coupled via Kevlar cables to a synchro-pulley of a hinge of the  solar array (see Figure 9).    The damper is an eddy current damper supplied by RUAG Space AG, Zurich, Switzerland. The  damper consists of three basic modules:  o A high efficiency damper unit which is a self-contained module with its own shaft  supported by its own set of bearings. The damping is generated with a high purity copper disc rotating within a highly concentrated magnetic field. The field is provided by 12 pairs of samarium-cobalt magnets. The damping rate can be varied. The different damping rate levels depend on the relative orientation of the magnet pairs which can easily be set by rotating the ECD end cover.  o An intermediate gear-head is employed based on a standard planetary unit. This unit  provides the first stage of torque amplification.   o The input stage is the second torque amplification stage and provides a further  amplification ratio. This stage is designed to accommodate high torque levels (up to 100Nm).  The damper is currently qualified to provide a damping rate between 500 and 1700  N-m-s/rad at  operational temperatures of -55 to +100°C. A dedicated attachment plate is required to attach the damper to the main bracket.  

Figure 6:  Damper Spring Unit parts.
102  

Figure 7:  Damper Spring Unit interface. 

Due to the heavy cable loads typical for a DSU application a new cable end fitting has been designed and  qualified. The Kevlar cable is a standard ARAMk3 / FRED cable type of which the fitting is changed from a single knot in a cylindrical bush into a spike in a conical receptacle (see Figure 8). Both DSU cables contain a preload spring. The design of the preload spring depends on the routing of the synchro-cable. For the Sentinel 1 DSU cable route a preload spring with a preload of 5N has been adopted. This value is an important aspect in the torque balance of the unit as a whole and should be analyzed carefully. The torque balance analysis is described later in this paper. Figure 9 shows the interfaces of the DSU cables.  

Figure 8: DSU control cable design.                  Figure 9:  DSU control cable interfaces.