ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
157   Figure  2.  Current NEA  Scout configuration      Therefore, large power and data wire bundles requir ed passage through the AMT . This wire bundle  consist ed of an estimated 40 power and data leads and 3 coax cables . Furthermore, they require d flexibility  and resistivity to radiation and extreme temperatures. The AMT was  required to translate ±8 cm by ±2 cm .  The wire bundles had  to pass through the AMT as it translate d between any two point s within the translation  envelope. Not only d id the dynamic envelope house the wire and cable bundle service loop, but also the  connector locations , which had to be positioned so that there was  no potential interference or tangling.  Thanks to the oversight of many experienced mechanical designers , the feedthrough bundle received much  attention  in early design stages . The AMT designer  was presented with m any technical papers regarding  the perform ance and degradation of flexible wire bundles in moving systems, including a previous AMS  publication “Lessons Learned to Avoid Coax Cable Failure in Moving Mechanical Mechanisms” , by Sheah  Pirnack. This paper suggested what types of coax cable designs wer e more resistive to cold welding  between conductor layers and provided suggestions to cable lots for workmanship and contamination, as  these parameters influenced cold welding. Furthermore, the paper discussed  when cold welding tends to  occur and how to design and test cables accordingly3. In turn, the design was modified to include ribbon  cables for data and power and accommodate for the largest possible bend radii for t he three coax cables .  Currently, o ptions are being investigated to route the coax cables outside of the AMT and reduce the  probability of cold welding altogether.  

Locking Mechanism   As with many translating or articulating mechanical systems, locking features are required for launch and  some orbiting maneuvers . The AMT was no different . The system  present ed a clear weak point and  introduces more failure modes to the vehicle. During the launch phase of the mission, a locking mechanism  was required to create a load path, reduce stress in fasteners, and constrain the translating components of  the AMT during the deployment phase of the mission. NEA  Scout’s interface to the CubeSat Dispenser is  a rail, and due to the discontinuity across the AMT, the interface was held to tight tolerances . The locking  mechanism will help control this tolerance. Although this mechanism w ill operate independently of AMT, it  was included as a subcomponent of AMT and therefore be contained within all AMT requirements . This w ill  prove exceptionally difficult as the locking mechanism w ill inherently reduce the translation ability of the  AMT.         RCS   AMT  Solar Sail  Deployer   Avionics  Volume   Approximate CP   Calculated  CM  Direction of  Flight
159 The main design trade completed during the AMT design was for the driver motors . Stepper motors proved  to be the best choice for a number of reasons . First, steppers have flight heritage on many spacecraft  deployment mechanisms and solar array gimbals . These heritage sys tems have similar duty cycles,  translation speeds and environments to NEA  Scout . Secondly, the stepper motor design allows for  microstepping, which when combined with gearboxes and lead screw  drive systems, can produce very  small, precise steps . Lastly, stepper motors can track motion by counting steps, so long as the detent torque  is not exceeded, causing skips . This step counting could replace a positional sensing system as long as  the motor is properly sized for the environment and loads4. 

In order to quantify the required driving torques needed for each of the two motors, the AMT designer  used  static and dynamic equations to calculate the friction forces and torques in the drive system. These equations combined  inertial loads from the driven mass; fri ction from the bearings, lead screw and nut;  motor efficiency; lead screw efficiency; and acceleration. The result produced  a torque requirement for a  particular set of motors, lead screws, nuts and bearings . These results were then documented and  compared to the torque margins required by the organization .    Although the torque outputs for stepper motors and torque requirements for the AMT drive systems were   relatively simple to calculate, there were many other elements  of the calculations that were assumed   (workmanship, degradation, feedthrough wire bending resistance, etc. ). Therefore, a drive system test  fixture was manufactured to simulate a single axis of the AMT, inertial loads and wire bend resistance. The  fixture was then paired with a custom driver  board to fine tune motor speed and current supplies . The test  assembly is shown in Figure 4 . 

Figure 4.  AMT test assembly   The test assembly was developed using stock ¼”  (6.35- mm) aluminum plate and 3D -printed components  to replicate the X and Y transl ation axes (the left and right side of the assembly, respectively) of the AMT .  The 3D -printed components were designed and manufactured to fit all configurations outlined in Figure 5.  The adjustability of this setup allowed for quick interchanging and cont rolled variations between tests.
162   References  

1. "Near Earth Asteroid Scout (NEAScout)." Jet Propulsion Laboratory . Web. 19 Jan. 2016.  <http://www.jpl.nasa.gov/cubesat/missions/neascout.php>.   2. Mehrparvar, Arash (February 20, 2014).  "CubeSat Design Specification"  (PDF).  The CubeSat  Program, CalPoly SLO . The CubeSat Program, CalPoly SLO . Retrieved December  2015.   3. Pirnack, S. (2012). Lessons lear ned to avoid coax cable failure in moving mechanical mechanisms . In  Proceedings of the 41st Aerospace Mechanisms Symposium, NASA JPL, Pasadena, CA, USA.   4. USA. NASA. Glenn Research Center. NASA Space Mechanisms Handbook . Ed. Robert L. Fusaro.  Linthicum Heights: NASA Center for Aerospace Information, 1999. Print.
163 Use of Volute Spring s in Space Applications  

Chris Floyd *, Ian Bournelis * and Cary Clark* 

    Abstract  

A volute spring was de signed and tested for use in launch restraint assemblies (LRA) on  satellite  deployables . The volute spring type was chosen because it offers a very compact stowed profile and a  relatively long spring stroke compared to conventional wire compression springs . Test results  showed that  the output force was very predictable with the developed analysis  and consistent with negligible change  following exposure to thermal cycling, vibration environment, and low cycle life testing. Although the  volute  spring was ideal for this  application, there are several lessons learned that should be considered before  using this spring type,  includin g the limited number of vendors willing/able to manufacture custom volute  springs , high relative cost per unit , nonlinearity of the spring output force, high weight relative to wire  compression springs,  the criticality of the spring end configuration (unaltered vs. ground vs. close d) and the  how it affected coil-to-coil frictional losses/build -up, and spring lubrication selection and migration during  handling.  

Introduction  

Springs are a simple and diverse component in many different types of mechanic al systems. They can be  used as mechanical energy storage devices, dampers, release devices and come in many different  materials, shapes, and sizes. Each spring is designed to suit the needs of its system. One particular application for satellites , LRAs , requires a spring with a very compact compressed height  and a very long  stroke length . The most suitable candidate for this application was a volute spring but the supporting  analysis and design background were not available.  

Volute Spring Overview  

The v olute spring  is manufactured from  a relatively  wide and thin bar or strip  of spring material . The bar or  strip is coiled around a mandrel such that each coil overlaps the adjacent coil forming a volute or cone. This  process forms a tiered inner diameter, shown in Figure 1, which allows each coil to slide past the adjacent  coil when the spring is compressed. The result is a spring that can be compressed to the initial width of the spring material regardless of the number of coils or stroke length the spring has . This is very convenient for  space applications because it allows for significantly more output force and stroke versatility for a given volume than any leaf, coil, or torsion bar spring. Due to the overlapping strip material configuration, the  volute spring also offers more stability than other spring types.  

Figure 1.  Volute Spring                                                     * Lockheed Martin Space Systems Company, Littleton, CO  

Proceedings of the 43rd Aerospace Mechanisms Symposium, NASA Ames Research Center, May 4-6, 2016
165   Figure 3.  Key Parameters and Variable Definition  [1] 

Figure 4.  Volute Spring Load (P), Deflection (δ), and Shear Stress (τ) Equations  [1]  Figure 5 is a theoretical load versus deflection curve for a volute spring. Initially, the spring has a linear  force constant as the coils slide past each other under axial compression. At the point when the largest coil  reaches the Initial Bottoming Load, the spring constant becomes exponential. This is because each  following coil bottoms out in sequence, increasing the stiffness after each one. Finally, the smallest coil bottoms out and the spring i s fully compressed.

Figure 5.  General Load -Deflection Characteristics of Volute S prings  [1]
166 Test Results   A number of different tests were performed to determine the whether the volute springs were both compliant  to drawing requirements and acceptable for use in the spacecraft applications.   Baseline Spring Force vs. Deflection Test   An Instron fixture with a motorized head ( Figure 6) was used to generate the baseline spring force vs.  deflection curve. The spring was set up in a simulated containment fixture and compressed to the stowed condition. The spring was then manipulated to relieve any frictional build- up to simulate a post -vibration  deployment . The Instron head was then programmed to deploy the spring at a rate of 0 .1 in/sec   (0.25 cm/sec) to evaluate slip- sticking that may occur during the deployment. Following the deployment ,  the Instron head then compressed the spring at the same rate to the fully stowed condition. The spring  force was recorded throughout this sequence. This test was repeated two more times to evaluate  repeatability.  As shown in Figure 7, the results proved to be very repeatable.  

    Figure 6.  Instron Test Fixture for Load vs. Deflection Curves   Vibration Testing   The spring was then compressed and subjected to a random  vibration environment. Following the vibration  environment , the force vs. deflection test was repeated three more times to evaluate the effect of vibration  on the spring force.  As shown in Figure 7, the results proved to be very repeatable and in family with the  baseline test results .  Thermal  Testing  The spring was then compressed and subjected to ten thermal cycles from - 100°C to +95°C . Following the  thermal cycles,  the force vs. deflection test was repeated three more times to evaluate the effect of thermal  environment  the spring force. As shown in Figure 7, the results proved to be very repeatable and in family  with the baseline and post -vibe test results . 

Figure 7.  Volute Spring Force vs. Deflection Data   01234567 0 10 20 30 40 50SPRING FORCE (KG) SPRING DEFLECTION (CM)Run 1 - Baseline Run 2 - Baseline Run 3 - Baseline Run 4 - Post Vibe 1 Run 5 - Post Vibe 2 Run 6 - Post Vibe 3 Run 7 - Post Thermal 1 Run 8 - Post Thermal 2 Run 9 - Post Thermal 3
168 The second configuration tested had ground ends which removed the rectangular points that were causing  issues with the previous test . This configuration was successful at mitiga ting the deployment  stability issues   as the spring now deployed straight and kept the parts oriented properly  in the final position. However,  while it was im proved slightly, the ground- end spring still exhibited an unacceptable frictional build- up by  the solid height due to the expanding inner coil.   

The final configuration tested had closed ends with a full dead coil at either end ( Figure 8). This configuration  successfully solved all the issues  discovered during testing of the unaltered and ground  variants. Ultimately  it provided a much more stable and repeatable deployment envelope and there were no indications of  frictional changes at any point during the stow or deploy process . The reliability of the output force  measurement significantly improved as well .  

Figure 8.  Unaltered vs. Closed Spring Ends    Based on the testing performed, it is highly recommended that  all future volute spring designs specify that  the ends be closed with a single dead coil on either end on the engineering drawing.   

Spring Lubrication   Two prim ary concerns from the beginning of the design processes were how the coil -to-coil friction would  affect the overall output force of the spring and whether the coils of the spring could cold weld together when the spring is held compressed in a vacuum envir onment . A trade study was performed to determine  the best options for lubricating the spring to reduce the impact and likelihood these concerns. Ultimately a  sputtered dry film lubricant  was selected since there was less concern of migration associated with wet  lubrication and no concern of cracking/chipping/flaking during spring operation associated with hard coatings .  

Instron testing was performed to determine deflection vs. spring force curves for both  lubricated and  unlubricated springs . The resulting data suggests that the lubrication had a minimal effect on the curve  under ambient temperature and pressure conditions . Both curves followed their analytically predicted  curves and did not show signs of sticking. The unlubricated spring was also functionally tested in a TVAC  chamber to evaluate the risk of cold welding. Following thermal cycling from - 100°C to +95°C in the  compressed state, the spring was released at each temperature ext reme. The following inspections showed  no signs  of cold welding.  

It is recommend that dry film lubrication be applied on springs to be used for space applications . The testing  performed was not comprehensive enough to eliminate the need for lubrication. I t did not evaluate the  frictional effects on the output force at the temperature, pressure, and humidity of the space environment and it did not evaluate the likelihood of cold welding if the spring was held compressed for vibration/launch then held in a v acuum environment with no release in between.
169 Conclusion  

While the volute spring may not be widely used terrestrially, they are ideal for space applications that require  a spring with a very compact compressed height and a very long stroke length. The equations  outlined in  this paper , which were validated by test data, provide the means to design and analyze this spring and to  verify force and stress margins. Testing also verified the volute spring perform s consistent ly after exposure  to hostile space environments.  For future designs , leverage the lessons learned and recommendations  herein to reduce the unit cost and schedule risk and increase the stability and reliability of the spring.  

References  

1. Wahl, A . M. “Mechanical Springs” Penton Publishing Company  (1944) , Page 359 -377.
172 and release clamp (based on CBK PAN idea). A n additional common disadvantageous feature of the  existing mechanisms is  their permanent, maximum stroke setting  without gradation of the stroke value. In  the proposed electromagnetic drive, equipping it with the power settings function has been realized  through electrical power supply.  

Preliminary Tests [3] 

The concept’s verification test -stand was developed and a single electromagnetic drive section (one coil)  was tested on this test -stand. Geometrical and mass proportions are retained in accordance with the  actual full design. Velocity of the hammer was measured using a linear encoder and magnetic tape. For  the purpose of the tests, the Counter Mass remained fixed, and as the result full stroke energy was  transferred to the Hammer Assembly.  

Goals of the t est campaign on this test -stand  were following:   • to see how coil material influences  efficiency of the system,   • to determine the energy fit of the  system (meaning at which stored ener gy the  system achieves the highest efficiency),   • to optimize balance between pair of  parameters: capacitance and voltage on the capacitor to achieve the most energetic stroke,   • to determine optimal initial gap  between hammer and counter -mass for the  fixed st ored energy value,   • and finally to prove the concept of  implementing a coil with ‘foot’ to the mole -type  design and see the outcome of it in comparison to the previously tested concepts.   For these reasons, a single coil with copper wire of 0.4- mm diameter ( 147 coil turns) was  tested to cover all the above mentioned system configurations.  

Figure  2.  The system’s efficiency as a function  of the capacitor’s energy for Iron and  Permendur 49 coils. Dashed circle indicates  optimal capacitor energy.   Figure 3.  Optimization of Hammer travel  gap. Measured hammer kinetic energy for  Permendur 49 coil for 4.3- J capacitor energy  level. Dashed circle indicates optimal gap.   Figure 1.  Verification test -stand, in the top  right the actual stand  with control electronics.
173 The results and analysis show :  • It is clear that using Permendur 49 can increase the efficiency of the system by 13- 20% with respect  to the electromagnet made of Iron coil, see Fig. 2.  • There are no radical changes in the resulting hammer’s kinetic energy when taking into account  different capacitors settings. Nevertheless, a certain optimal setting can be distinguished for  Permendur 49 (see Fig. 3 .), which is for 100 V/860  uF.  • When it comes to adjusting the energy level to achieve the most efficient system, it is clear from the test resul ts that by lowering energy on the capacitor we get higher efficiency (at least when  approaching down to energy of c.a 3.6 J). Nevertheless, it is energy of 4.3 J that is selected as the  optimal .  • Optimal travel gap for hammer was also researched for various  capacitors’ settings. Fig. 3  shows  plots for constant capacitor energy of 4.3 J. For most of the cases , the optimal gap is in the range of  4.5-6.5 mm (with full travel freedom of 8 mm). The selected capacitor settings (100 V/860  uF) hav e  an optimal travel  gap of 5.5 mm.  

Design Features  

Mechanical design description   The drive  consist s of 5 electromagnetic driving sections in which coils,  ferromagnetic cores and magnetic separators are concurring to the counter -mass, while armatures of electromagnets with a connection shaft  are comprising the hamm er. All electromagnetic circuit  elements are made  of Permendur 49, whereas the magnetic separators are of a tungsten alloy and the connection shaft is made of titanium alloy. The propulsion (counter -mass + hammer)  is enclosed in the outer casing assembly and consists of a long, thin wall, titanium sleeve and hardened  stainless steel tip.    The hammer is guided by two sets of  3 bearing balls inside the counter - mass which provides low  friction  losses, while the counter -mass is  sliding inside the outer -casing tube.  As in all the previous penetrators made by CBK PAN, while the hammers starting position (before hit) is set up by a hammering spring, the position of the counter -mass is  determined by a brake spring designed in a way that allows the penetrator to work in a non- gravity  environment w ithout any support . A  cross- section of a CAD model of  EMOLE is shown in Fig. 4.   The EMOLE is  25 mm in diameter  and 250 mm length. The mass of the overall assembly is about   700 g, where the counter -mass mass amounts  to 520 g and the  hammer and the outer c asing are of equal weight of about 90 g.   Figure 4.  EMOLE CAD model  cross- section.  

Figure 5.  Outer casing,  driving sections and  return spring just  before final  assemblage.
174 Electrical design description [4]  

The electrical system consists of three parts: control unit (controller), DC/DC converters unit , and  electromagnetic drive placed in penetrator. All three parts are designed to operate separately. Control  unit, currently in form of a control panel, is used to set signals to turn on the desired mode of operation of  the drive electronic system and therefore to control the penetrators behavior . The control unit’s interface  is able to  turn on and off each of the two DC/DC converters as well as to select one of several “gears” –  penetrator stroke energy and actuation frequency.    The DC/DC converter  unit consists of two DC/DC converters, both in flyback topology. Each DC/DC  converter has an overvoltage protection mechanism ensuring safe operation of the device.  High Energy  DC/DC (DC/DC HE) converter charges large capacitor, stores the energy, which i s then released to the  electromagnetic drive when it reaches a certain level, set by device operator. The higher the charge (the energy stored), the lower the actuation frequency.    High Frequency DC/DC (DC/DC HF) converter, operates in a similar way as lat ter DC/DC converter but  operates on the energy level ~200 times  lower and two order s of magnitude higher frequency. It is  galvanical ly separated from the system ground to operate totally independently from the state of DC/DC  HE. 

Tests [5] 

Actual velocity  (and kinetic energy) speed measurements on the finalized assembly   The actual velocity of the hammer with a blocked counter -mass was measured on a test stand. For this  purpose,  the counter -mass was clamped in the test stand (without the outer casing)  and the movement of  the hammer was captured on a high- speed camera. The measurement was repeated for all five high  energy power settings that the electromagnetic drive possesses. The summary for the achieved velocities, resulting energies and expected eff iciencies f or each PS are presented in Table  1. 

Table 1.  Summary of results   Power setting  Hit velocity [m/s]  Hit energy [J]  Efficiency [%]   PS-1 3.352  0.49 10.72   PS-2 4.621  0.94 10.62   PS-3 5.774  1.47 11.22   PS-4 6.503  1.86 10.62   PS-5 7.096  2.22 10.28    These results show that the efficiency varies insignificantly between each power setting. The highest  efficiency was achieved for the third power setting.  Furthermore, this test proved that the stroke energy   changes quite equally with the number of the drivin g sections. It means  the amount of sections can be  raised or decreased with respect to the direct mission mass or power requirements, and the stroke energy will raise or decrease equally.  

Tests in Syar r egolith analogue   After determination of the energy generated by five -coils driving system, the EMOLE penetrator was  tested several times, on different power settings, in Syar regolith up to depth of about 50 – 60 cm. The  summary of the se tests is presented in Figures 6, 7 and 8.