ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
the power amplifiers driving the actuators. The update rate is slow enough that the controller is of the  look-shoot-look-shoot type and is thus inherently stable.  0.000.100.200.300.400.500.600.700.800.901.00 0 1 02 03 04 05 06 07 08 09 0 1 0 0 1 1 0 1 2 0 1 3 0 1 4 0 1 5 0 1 6 0 Piston absolute pressure, psiaVertical suspension frequency (Hz) In atmosphere In vacuumPayload (lbs) 0.01000.2000.3000.4000.5000. Figure 3  Calculated vertical suspension frequency and payload. The air spring using a frictionless piston behaves essentially as an ideal regulated air spring. Its stiffness  is purely a function of the gas properties of air, the piston area, and the total pressurized volume. The  relationship, derived from basic thermodynamics assuming linearity and is entropic compression and  expansion, is k = /g74 A2 po/Vo where k is the stiffness, /g74is the specific heat ratio of the gas (1.40 for air), A  is the piston area, p o is the mean absolute pressure and V o is the mean volume. The dependence of  stiffness on pressure means that the suspension frequency is essentially independent of payload mass2. Further, it means that an air spring can always be brought back to a zero sag condition under weight load  by adjusting the mass of air in the cylinder.  Implementation via a frictionless piston rather than a flexible air bag has the advantage that an air bag adds significant stiffness, which sets a lower limit to the  suspension frequency. The suspension frequency of a piston air spring can be reduced without limit by  increasing the external volume.  Figure 3 shows the calculated suspension frequency and payload for a payload suspended by a Type A  device with 227-l (60-gallon) accumulators. There is some loss of payload capacity and increase in  suspension frequency because of the vacuum itself. As will be shown, the moving element of the  suspension devices is subject to vacuum force, which slightly decreases the payload capacity and  increases the suspension frequency. The increase in frequency is also caused by flexible metal bellows seals needed for vacuum compatibility.  Design of Suspension Devices  Figure 4 shows two views of a simplified solid model of a Type A suspension device. Missing from the  model is the “bell jar” vacuum housing that encloses the part of the mechanism above the baseplate. The  housing recaptures all the exhaust air from the air bearings and pistons such that it can be routed via a  tube out through the wall of the va cuum chamber. This and the bello ws seals around the carriage rails  that extend downwards through the baseplate are the essential features that make the devices operable  in vacuum.  2 This independence requires that the absolute pressure be at least several times the ambient pressure acting on the  top of the piston.  319
Figure 4  Conceptual design of suspension (type A with active augmentation).  The device consists of two main subassemblies: a moving carriage shown in shades of red and a fixed  frame shown in shades of blue. The carriage moves vertically on four journal air bearings, two in each of  the two cylindrical housings as shown. The carriage lifts the payload via a cable connected to its lower crossmember through a load cell as shown. In the present application, the load cells are used to monitor  the force distribution among the devices when “parked”, i.e. when the air springs are deflated and the  carriages rest against their lower travel stops. Uplift force against the upper crossmember of the carriage is supplied by the frictionless air piston through a connecting rod. The piston is actually a combination of a  journal air bearing and a piston. A very thin, stable air film is produced around the piston skirt such that it  never actually touches the cylinder, thus eliminating all friction. A small amount of air leakage is allowed  with make-up air being supplied by the precision pressure regulator supplying the external accumulator  tank. As noted earlier, the carriage lower crossmember is sealed against the bottom side of the baseplate  by two flexible metal bellows seals which surround the two carriage rails as shown. Type A devices use four voice coil actuators operating in parallel with the air piston. Type P devices do not have voice coil  actuators but have dual air cylinders for increased lifting capacity. Both Type A and Type P devices  include LVDT displacement transducers to sense the carriage position, high-resolution pressure transducers for monitoring piston pressure, and load cells for monitoring the load distribution.  Pistons are 146.0 mm (5.75 inches) in diameter for a bore area of 16,753 mm 2 (25.96 in2). This produces  a gross uplift force of 13,890 N (3115 lbf) at a gauge piston pressure of 8.28 bar (120 psig). Type A  devices are nominally rated at 11,147 N (2500 lbf) in vacuum which accounts for the weight of the moving  carriage and the vacuum force pulling down on the carriage (the interior of the bell jar housing is at  atmospheric pressure). The Type P devices use two pistons of this size and are rated at 24,524 N (5500  lbf) in vacuum. Vertical stroke between the travel bumpers is 3.8 cm (1.50 in) for both types.  Voice coil actuators are driven by current-control analog power amplifiers using 16-bit digital commands.  Force resolution of the active subsystem in a Type A device is approximately 0.36 N (0.08 lbf).  The magnetic pendulum dampers (Figure 1) are built from the magnet bodies of the voice coil actuators  with the wire coil replaced by a solid copper annulus. A flexure suspension is used to allow the conductor to move axially in the magnet body while remaining concentric to it. Details are in the next section.  320
Hardware Description  Figure 5  Suspension devices.  Type A, augmented (left), Type P, passive only (right). Figure 5 shows some of the delivered  hardware. The Type A suspension  device on the left is identified by its  single piston and four voice coil actuators. The Type P on the right has  dual pistons and no voice coils. The  circular groove in the top side of the baseplates is for the O-ring that seals  the bell jar flange to the baseplate. All  air and electrical feedthroughs are  located in the baseplate where the  interfaces can be accurately machined. The devices are shown mounted on  assembly and shipping stands.     Figure  6 shows a bell jar being trial-fitted onto a Type A suspension device.  Figure 7 shows a Type A during  assembly. The carriage is assembled  into the frame initially without the  bellows seals. After the plumbing, wiring, and various small subassemblies are added, the lower crossmember is removed, the seals are mounted into place, and the crossmember is reinstalled. This  allows the carriage motion to be checked for friction before the bellows are added and also protects the  bellows, which are somewhat fragile, from damage during most of the assembly process.       Figure 6  Trial-fitting of vacuum housing onto        suspension device. 321
An example of the bellows seal is shown in  Figure 8. It is formed bellows rather than  edge-welded. While the latter type would  have had the advantage of lower added stiffness, it would have been much more  costly and would have had an unacceptably  long delivery time. Since analysis had shown that the target suspension frequency  could be met with formed bellows, that was  the type used. The end flanges are machined with standard O-ring grooves as  shown and then tungsten-inert gas welded  to the formed bellows. The bellows are  fabricated from 321 stainless steel because  of its high ductility. The flanges are made from the same material for welding  compatibility. Inside diameter of the bellows  is 93.7 mm (3.69 in) and flange-to-flange length in the unstrained condition is   107.4 mm (4.23 in). Stiffness added by the  bellows is about 2.90 N/mm (16.5 lbf/inch) per bellows.  Figure 7  Partially assembled Type A device showing  the carriage rails and lower crossmember.  Bellows  seals are installed later.  An important consideration for internally  pressurized, formed bellows is lateral  (inchworm) buckling. Because the critical  buckling pressure is always proportional to axial stiffness divided by length, there is a certain  stiffness that must be tolerated. As usual, this  controlled the design since the critical pressure had to be at least 1 atm. Measured buckling  pressure was found to be in good agreement  with theoretical predictions.  Figure 9 shows two views of the passive  magnetic pendulum dampers. The body of each  damper was mounted to a temporary stanchion  fixed to the floor of the vacuum chamber. The moving element of the damper was then  connected to the paylo ad by a “stinger” having a  flex joint at either end.  The resemblance of the magnet bodies of the dampers to those of the  voice coil actuator (Figure 5) is clear. While  damping performance could have been improved by a custom design, the ad-hoc design  based on the actuators was adequate so the  ever-present schedule pressure dictated its use.  Working stroke of the dampers was about 12.7 mm (0.5  inch). Overall length of the assembly as shown is 248.7 mm (9.79 in) and the envelope diameter is 279.4  mm (11.00 in). Weight is approximately 11.4 kg (25 lb) per damper.  Figure 8  Flexible metal bellows seal. Experience in modal testing has shown that pendulum modes have practically zero inherent damping.   The basic reason is that the potential energy is stored in geometric stiffness rather than elastic stiffness of  a material. Hence, there is no material loss factor to dissipate energy. While relatively simple compared to  the suspension devices, the passive pendulum dampers were essential to the overall success of the  suspension system.  322
Figure 9  Passive magnetic dampers for payload pendulum modes Testing The aggressive development schedule for LSIS left time for only the most essential tests prior to shipping  the hardware. These were friction and proof load tes ts of the suspension devices and leak rate tests of  the devices and their accumulator tanks. Dynamic tests were performed after the fact on a spare damper  for engineering purposes. This section reports methods and results for these tests.  Friction tests of suspension devices. Figure 11 shows the setup for friction tests. A  custom test stand (the white structure in the  figure) was designed and built for the purpose.  It provided a mounting location for the device under test and a means of moving the carriage  up and down through a small stroke while the  air spring and accumulator tank were fully pressurized to their maximum expected  operating pressure. Force applied to the  carriage to produce the motion and resulting  displacement were transd uced and recorded.  If friction was present, it would show up as a  step in the force at the instant the carriage motion reversed direction. The amplitude of  the step would be twice the friction force.   Figure 10 shows a detail of the actuation  device for moving the carriage. Initially a  variable speed, DC elec tric motor was to be  used to turn a crankshaft with a small, 1.27-    Figure 10  Actuation mechanism for friction tests   of suspension devices. 323
Figure 11  Test rig for measuring friction in suspension devices mm (0.050-inch) stroke. The motor speed was reduced to about 120 RPM at the crank by a two-stage  timing belt drive. However, in initial trials, it was found that the SCR motor controller would not control the  speed smoothly as the crank passed TDC and BDC and the motor transitioned from driving to retarding the load. Large spikes in the motor torque occurred which would have obscured the small transients due  to friction, had such been present. The solution was actually a simplification to the rig. A hand lever was  fabricated and mounted directly on the crankshaft to replace the motor drive. With practice, it was possible to use the lever to rotate the crank back and forth through about 180 degrees with very smooth  transitions in rotational direction. This was the system used for the actual tests although the DC motor  and most of the belt drive are still present in the photographs.  Figure 12 and Figure 13 show typical results. Figure 13 is data obtained with a device operating correctly.  The force trace, which has had its DC component removed, shows no sudden steps at the points where the displacement hits its extreme values and the direction of travel reverses. Figure 12 shows a case  where there was friction. A step of about 5.3 N (1.2 lbf) occurs as the carriage reverses direction. This  data was obtained in testing of a Type  A device so the friction force (half the step amplitude) amounted to  only about 0.019% of the average force. However, this exce eded the nominal specification for smaller  devices of this type (<0.005%) so the offending part (the piston) was rejected. It was replaced with a  spare unit that passed the test. Experience has shown that friction, if not due to some obvious rubbing of a visible moving part against a fixed part, is almost always due to the piston. Of the 22 pistons and  cylinders fabricated, only one failed the friction test. This is actually a better yield that is typically obtained  with smaller piston/cylinder sets. It appears that frictionless pistons are actually somewhat easier to build  in larger sizes than in smaller.  The force transducer used for the tests was a piezoelectric, charge mode type with a charge amplifier  having a time constant of about 10 seconds. The extremely large dynamic range of this measurement  system allowed detection of force transients well under 1 N in the presence of DC forces of 25,000 N.  324
Proof load tests of suspension devices Each of the 14 devices was proof tested to twice  its rated load capacity. A modified version of the  friction test rig was used for this purpose. Shown in   Figure  17, it substituted a hydraulic cylinder  for the crankshaft mechanism of the friction  tester. Force was determined from the hydraulic pressure applied by a hand pump. The proof  loads produced severe  deflection of the  elastomeric travel stops built into the suspension devices (Figure 4) but they all survived and  returned to normal after the load was removed.  No failures occurred during proof load tests.  Dynamic testing of dampers One of the passive magnetic dashpots was  tested to verify its damping properties. Figure 18  shows the test rig. The body of the damper containing the magnets and back iron (the shiny  object in the photo) is mounted to a bulkhead in  the center of the aluminum frame of the rig. The conductor is mounted on a carriage rail that forms part of the rig. The rail runs on two journal air bearings  to eliminate any friction. One end of the rail (the right end in the photo) protrudes through the end  bulkhead of the test rig frame. It is driven axially by an electrodynamic shaker working through a piezoelectric load cell. The other end of the carriage rail mounts an aluminum disk that serves as a target  for a non-contact eddy current displa cement sensor. The eddy current probe mounts to the left-end  bulkhead of the test rig frame. The shaker is driven with a band-limited random current and the signals from the force and displacement sensors are input to a multi-channel digital Fourier analysis system. It  computes a frequency response having carriage velocity as the input (denominator) quantity and force as  the response (numerator) quantity. In the process, it removes from the force signal the portion that is due to the known inertia of the moving carriage. What remains is the drag force from the magnetic damper  only. The result is a complex-valued function of frequency whose magnitude repre sents the damping  force per unit of velocity (often called the dashpot constant, although it is anything but constant in this  case). The phase of the complex  function represents the phase of the damping force relative to the  velocity.   Figure 17  Setup for proof load test of suspension    device Figure 18  Test rig for measuring complex stiffness of magnetic dampers.  328
Figure 19  Measured force/velocity ratio for the magnetic damper Figure 19 shows the magnitude of the measured frequency response function for three configurations of  the conductor of the damper. The uppermost (red) curve is for the dampers as delivered and installed  with the suspension system. These used a single, solid-copper annulus (also called a single shorted turn)  as the conductor. The middle (green) curve is for a conductor formed from the same wire-wound coil used  in the voice-coil actuators but with the wire ends shorted together. The blue curve is for the aluminum coil form (also called the bobbin) used in the voice-coil actuators but with no wire windings. Not surprisingly,  the force/velocity ratio increases as the resistance in the induced current path is decreased, just as theory  predicts.  The back iron assembly of the voice-coil actuator (identical to that of the damper) was of a design that  produced relatively high inductance in the coil. For the actuator, this was not important since it is used essentially only at zero frequency. For the damper, it was expected that this would cause the induced  currents in the damper conductor (which cause the damping force) to drop off with increasing frequency.  This would be valuable for the present application since it would allow the magnetic dashpots to damp the very low frequency pendulum mode (less than 1 Hz) but then disappear at higher frequency so as not to  transmit disturbing forces into the payload at higher frequencies. The red curve of Figure 19 shows that  the design was quite successful. At frequencies over about 50 Hz where the resonances of the payload begin, the force/velocity ratio is reduced by about an order of magnitude from its DC value.  Conclusion Figure 20 shows eight of the fourteen suspension devices mounted in a framework, ready to be  suspended from the roof of the vacuum chamber. These are the eight that support the Body C of Figure  1. While schedule constraints did not allow systematic tests, informal measurements of the vertical  frequency by the end user indicated that it was, in fact, 0.50 Hz as designed. Most important, the degree of vibration isolation provided was sufficient to allow the functional tests of the spacecraft to be performed  without limitations from ambient disturbances. The LSIS did its job.  329
The primary conclusion from the development was that the very low stiffness pneumatic suspension  devices are scalable. In fact, anecdotal evidence suggests that, at least up to the largest sizes built to  date, obtaining the critical zero-friction behavior is actually easier in larger devices (piston bores over  about 76.2 mm (3 inches)) than in smaller ones. A system is now in preliminary design that will use multiple devices with piston bores over 203.2 mm (8 inches) to float a payload of approximately 45,454 kg  (100,000 lb) in vacuum.  Figure 20  Eight suspension devices with accumulator tanks mounted in frame for supporting  Body C.  Size is evident from the person in the lower right of picture.  References  1. Kienholz, D.A.  “Simulation of the Zero-Gravity Environment for Dynamic Testing of Structures,”                Proc. 19th IEST Space Simulation Conference, Oct 28-31, 1998, Baltimore, MD  2. Sillls,Jr.J.W., Voorhees,C.R. “Characterization and Application of Pneumatic Suspension Devices               for Vibration Disturbance Testing,” Pro ceedings of the 20th International Modal Analysis               Conference, February, 2002.  330
the USMs and BDMs are ignited, a pyrotechnic joint at the forward end of the interstage initiates and the  vehicle begins to separate. Figure 1 shows the Ares I configuration with the BDMs mounted on the interstage. In the most recent configuration, they are relocated to the aft skirt of the first stage. The J-2X nozzle exit plane is 7.1 meters aft of the separation plane. With the current arrangement separation system, it takes approximately 1.7 seconds for the nozzle to pass the forward end of the interstage.   Nozzle Clearance During Fly-Out Considerations   Many factors affect the amount of radial clearance between the engine nozzle and the interstage wall during the fly-out. The most significant factor contributing to clearance issues for BDM separation is asymmetric plume impingement force on the first stage that can occur if one motor fails to fire. Secondly, since the first stage has 178 kN of residual thrust at the time of separation, significant pitching and yawing loads may be imposed on the stack before separation and on the first stage after separation due to thrust vector pointing uncertainties. With one BDM out, a worst-on-worst analysis of the separation shows contact between the interstage and the engine nozzle during fly-out. Monte Carlo analysis of this scenario  shows that nozzle clearance can only be demonstrated to a 2.5-sigma level.   The proposed MAG Strut system uses eight gas-charged struts mounted inside the interstage to force the two stages apart. The struts essentially act as alignment guides during separation. Figure 2 shows the relative position of the struts on the interstage to the USMs and the BDMs they will replace.  

Figure 2 - Interstage showing BDMs and Struts   Although the struts extend above the separation plane, they provide superior clearance, even with one strut out. The primary reason for this superior pe rformance is that the mass-moment-of-inertia of the Ares  I upper stage/crew vehicle is approximately ½ that of the mass-moment-of-inertia of the expended first stage, while the distance from the upper stage/crew vehicle center-of-gravity to the J2 nozzle exit plane is approximately ½ the distance of the center-of-gravity of the first stage to the separation plane. Figure 3 shows the relative positions of the centers-of-gravity of the separated stages to the nozzle exit plane and first stage separation plane. With a strut system, any disturbance force, regardless of its origin, is compensated for by the struts, forcing the separated stages to rotate in the opposite directions. The rate-of-rotation, W, induced on the two bodies in always close to 2/1 with the upper stage/crew vehicle rotating at twice the rate of that of the first stage. The rate of rotation of each body is small with the gas strut system. Distance D3 is considerably larger than distance D4 so some of the disturbance force coming from the first stage results in translating the upper stage in the same direction the interstage is moving. This translation effect, though beneficial, is not as significant as the rotational compensation.    332
Figure 3 - Comparison of Gas Strut Separation and BDM Separation 

Figure 4 - Ares I Separation Clearance using Gas Strut System  Figure 4 shows the preliminary clearance results for the Ares I upper stage engine nozzle with one strut  out. The WOW*1.5 curve represents a worst-on-worst assessment of the radial clearance with a margin of 50% added to account for unknowns in the analysis. Even in this conservative case, the nozzle clears  the extended end of the strut by 45.7 cm. The dash lines represent WOW case clearances for different  failed struts with different disturbance scenarios. Two seals must fail on the same strut to result in a 100%  333