ai-aerospace/ams_data_full_2000-2020 huggingface dataset:
61  Distortion remained above the 2% level for all of the above modifications  Some of the lamination styles that were tried to reduce the harmonic distortion are shown in Figure 11.  Recognize that this is a time consuming process to build the stacks, EDM the laminations, machine, and coat with an insulator as well as wind and insert the coils. 

Figure 11.   Pole shoe shaping, tooth gap, and magnetic span variations were tried.  FEA analysis alone was not considered a viable approach alone in that the difference of distortion was only about 1%. Shapes tried were:  wide boot height to narrow boot height, wider tooth widths, wide tooth to tooth gap to narrow tooth to tooth gap, wide magnet to magnet spacing to narrow magnet to magnet spacing - all provided no change in harmonic distortion to bring values below the 2% limit. There were also models made with the edges of the magnets rounded.  At some point defeat must be acknowledged and a call for help initiated, be it self imposed or externally  suggested. The suggestion was to call in help from an recommended and established consultant. This was accomplished and was followed by an impressive 14-page report from the consultant. The report offered a solution. The proposed solution was so deceptively simple that it was welcomed with great enthusiasm. At the same time, the report gratifyingly duplicated the non-sinusoidal BEMF results we were observing, as well as duplicating other motor parameters, so high expectations were anticipated.  The solution based on a detailed FEA analysis (using a different analysis tool than the ones we were  using) was to make the magnets as flat rectangles rather than the having the normal surface curvature mirroring the stator curvature as in traditional designs.  A test motor was then wound, magnets purchased, and tested with the wonderful results that all of the  parameters, including harmonic distortion were now within specification. A bonus feature was the detent was reduced from 18 to 9 N-cm (25 to 13 in-oz) giving a good margin.  What then followed was the production manufacture and testing of the modified rotor design. Total  disbelief ….the motor performed as poorly as it had before the rotor magnet change. What had  happened? Our thoughts went back to the CP49 and the M-15 materials. We always felt (just  experience), despite the FEA analysis showing the reasonable flux paths within the two materials (Figure 10) that the M-15 was the proper way to go given past experience with the CP49 Hi nickel materials annealing variations.  What had happened is that the flat magnets were mistakenly tested with the original M-15 stator that, as  you recall, was originally made but never used. The M-15 lamination stack had gotten mixed up with the CP49 lamination stack. To prove that the model that was successful with the flat magnets was using M-15, a known M-15 stator stack was then wound and tested. Back to the good results tested!
62 Tests with older curved rotors using the M-15 stator also showed the same excellent results indicating  that the issue was with the CP49 material and not anything else. The lesson learned here is that FEA analysis is a path towards good designs but is not the end all. Practical machines with practical materials must ultimately be the final say. We stayed with the flat magnets because we had purchased the entire programs supply. In retrospect, had we had been able to have gone with the M-15 material (usually our preferred material for this type of motor) most of these delays and perturbations relative to the harmonics would not have occurred or been necessary. A photo of the G motor is shown in Figure 13.  The data in Figure 12 shows the distortion improvement with the M-15 material from 2.79% to 1.57%.  Fundamental 18.4 volts; 5 th Harmonic at 0.26 volts   Harmonic distortion = 0.260/18.4 = 0.0157 = 1.57% 

Figure 12.  M-15 Harmonic Distortion Data for the M-15 Material 

Figure 13.  Photo of the G Motor - Rotor is encapsulated in titanium ring 

Major Motor Parameters for H Motor   The H motor, also an inside out design, presented with a whole different set of issues. The initial design  using CP49 and Samarium Cobalt magnets resulted so enormous a difference between performance and requirements that an initial drastic step needed to be taken before any iteration would be possible.
63   Figure 14.  Geometry of the H motor  The Kt, detent and distortion below represent the performance requirements:   * Kt > 56 N-cm/amp (80 in-oz/amp)   * R = 0.84 ohms within 7%, L = 2.28 mH within 30%   * Friction torque < 19 N-cm (27 in-oz) at 272 RPM   * Detent torque < 15 N-cm (21 in-oz)  * % Harmonic Distortion < 8%  The initial lamination stack was made with CP49 material. The waveform and detent were much worse  than seen of the G motor and the decision to use M-15 at the outset was immediately made. The result for the first CP49 H motor is shown in Figure15.  The Kt was measured at 183  N-cm/amp (259 in-oz/amp) and the detent was over 141 N-cm (200 in-oz). 

Figure 15.  BEMF Curve with CP 49 and 0.635-mm (0.025-inch) air gap  The first attempt using M-15 material resulted in a Kt of about 79.8 N-cm/amp (113 in-oz/amp) with the detent  being approximately 64 N-cm (91 in-oz). The decision to skew the stator was made to reduce the detent  and smooth the waveform with the results shown in Figure 16.
64   Figure 16.  BEMF for H motor with M-15 Material  Again, exploring the flux density differences between the FEA analyses of the materials (Figure 17), there  again did not seem to be much difference to cause this great disparity of performance for the H motor due to material as seen between Figures 15 and 16. 

Figure 17.  Comparison of lamination material for the H motor  This motor was most critical in terms of balancing the flux carrying capability of the laminations with the  amount of copper area available affecting Kt, inductance and resistance. The motor needed a full complement of 40 turns (6 of #28 and 2 of #29 AWG wire) to conform to the specification requirements (Kt, resistance and inductance). A bonus of the skew was that the detent torque was reduced to about  1/3  of specified allowable which was well received.  The wire would not fit within the area needed for the lamination design. (Several tooth widths were tried to  arrive at the final configuration). Not a quick task when the programming, EDM process, lamination coating and winding time factors are considered. As a result, the motor could not be wound and inserted in the normal production manner. Normally, a string of coils per phase is wound on a mandrel, loosely tied to hold each coil of 40 turns together and then inserted into the stator as whole coils placed wire by wire into each stator slot. In this case the slot fill factor is so high that this would not allow the coil bundles to fit into the slots, just missing by a few percent. The stator needed to be hand wound, turn by turn, into the slots allowing each turn to be tightly fit into the slot maximizing the space available with just the right length of wire to make the resistance specified. Under these conditions it is necessary to have a very patient and amiable technician in your employ!
65   Figure 18.  Photograph of the H motor with the lamination skew 

Summary  With the above trades and compromises, all of the motors styles ultimately made their way into successful  production cycles.  In summary, what was needed to be changed from the 5 initial restrictions (which were not specific to the  performance requirements) in order to meet the performance criteria is as follows:  1 - Magnet material must be Samarium Cobalt - In one case (G motor) higher energy product Neodymium  magnets needed to be substituted.   2 - Lamination material must be Carpenter 49 (High Nickel Steel) – In two cases silicon steel was needed  to replace the high nickel steel. 3 - Primary and Redundant windings to be mechanically spaced at 120 degree intervals for each of the  three phases alternating every 60 degree increment - Preserved   4 - Stators were not to have a skew – In one case (H motor) a partial skew was required for detent and waveform conformance.  5 - Parameters must meet the tight specification values – In one case some relief was required for the  inductance values  6 - 0.75 slot per pole configuration with 18 slots and 24 poles was required. - Preserved   The major lesson learned from this project is to have give and take discussions between parties to  determine which of the specification requirements are truly pertinent to the performance as opposed to those which would be nice to have based on existing hardware.
67 Development of the Vibration Isolation System for the   Advanced Resistive Exercise Device 

Jason H. Niebuhr* and Richard A. Hagen** 

Abstract   This paper describes the development of the Vibration Isolation System for the Advanced Resistive Exercise Device from conceptual design to lessons learned. Maintaining a micro-g environment on the International Space Station requir es that experiment racks and major vibration sources be isolated. The  challenge in characterizing exercise loads and testing in the presence of gravity led to a decision to qualify the system by analysis. Available data suggests that the system is successful in attenuating loads, yet there has been a major component failure and several procedural issues during its 3 years of operational use.   Introduction   Resistive exercise (weightlifting) is prescribed for crew of the International Space Station (ISS) to maintain muscular strength and bone density during long duration space flight. However, the repetitive nature of exercise induces vibratory loads that can degrade the sensitive ISS micro-g environment and reduce its structural fatigue life. The Interim Resistive Exercise Device (iRED) was flown aboard the ISS from 2002 through 2008 as the primary means for resistive exercise. Ongoing maintenance issues, increased performance requirements, and the need for a Vibration Isolation System (VIS) eventually drove the need for a new clean-sheet exercise machine, called the Advanced Resistive Exercise Device (ARED) [1]. ARED creates resistance with a pair of vacuum cylinders that are connected through a variable length lever to an adjustable height bar as shown in Figure 1. The exerciser can stand, sit, or lie on the platform or bench (not shown) to perform a variety of exercises.                                                     * Formerly with NASA JSC, now with Apogee Engineering at the US Air Force Academy, Colo. Spgs., CO  ** NASA Johnson Space Center, Houston, TX 

Proceedings of the 41st Aerospace Mechanisms Symposium, Jet Propulsion Laboratory, May 16-18, 2012  Vacuum Cylinders Figure 1.  The Advanced Resistive Exercise Device Lever Arm Length  Determines Load Pivot PointAdjustable Height Ba r Platform
68 The primary micro-g requirement is a frequency-dependent acceleration limit at the US Lab and  Columbus module rack interfaces. Structural dynamics transfer functions are available to quantify the acceleration at those racks as a function of input load at the ARED location within the ISS. With the assumption of equal loading in all axes, the acceleration limit and transfer functions can be taken together to consider the requirement in terms of allowable load. A similar frequency-dependent load limit is required to preserve ISS fatigue life. The composite allowable load enveloping both micro-g and structural fatigue requirements is shown in Figure 2. This is a simplified and conservative view of the requirements,  yet highly insightful. The graph highlights the most significant challenge to the design; a major ISS  structural mode exists near 0.24 Hz which can be seen as a steady state allowable load of only 1.7 N (0.4 lbf) between 0.18 Hz and 0.28 Hz. As shown by the shaded area of the graph, this is well within the 0.09 Hz – 1.18 Hz range of exercise frequencies recorded during the ARED man-in-the-loop test (MILT) [2]. Analysis of the ARED in a non-isolated configuration suggests that exercise loads can be as a high as 67 N (15 lbf) at 0.24 Hz, requiring attenuation by more than an order of magnitude.    Analytical Approach   A rigid body dynamic model was created in the motion analysis software visualNastran 4D from MSC.Software Corporation. This software was chosen over general numerical computing codes for its native 3D visualization capabilities and ease of use. This made it possible to examine many different design concepts and better understand how they worked. Though somewhat controversial, it was felt that the cost and complexity of an actively-controlled isolation system wasn’t justified. Being able to visually simulate system dynamics was essential in convincing both technical and non-technical stakeholders alike that a passive isolation system would be sufficient. Model development focused on three distinct elements: a geometrical and mass model of the ARED, a geometrical and mass model of the exercising crew member, and a mathematical model of the VIS.   Development of the ARED model was straightforward as solid geometry was imported from computer-aided design (CAD) software after being simplified into 12 major subcomponents by suppressing details like fasteners, fillets, and lightening pockets. Detailed mass properties of the subcomponents were retained from the original CAD model. Appropriate constraints were applied to the model to simulate fixed, rotational, and sliding joints. Capturing this level of detail kept the computing requirements manageable.    The exercise model consists of a CAD model human constrained to simulate exercise motion when driven by displacement. Figure 3 illustrates how the human model and subcomponents of a non-isolated ARED  Figure 2.  Allowable Loads Requirement and Exercise Frequency Range Exercise Frequency Range 0.09 Hz – 1.18 Hz
69 both move throughout the stroke of a squat exercise. The y-axis is parallel with the long axis of ARED,  the z-axis positive toward the exerciser, and the x-axis completes the right-handed Cartesian system. Exercise variables include type, stroke, crew size, and frequency. The exercise types and stroke lengths range from a squat moving nearly 90% of the body’s mass through 0.70 m (27.5 in) of stroke to a neck flexion that moves less than 6% of the body’s mass through less than 0.23 m (9 in) of stroke [3]. However, the squat, dead lift, straight leg dead lift, and heel raise are worst-case and envelope all other exercises in the analysis because they move the most amount of body mass over the greatest distance. The 95 th percentile American male represents the heaviest crew analyzed while the 5th percentile  Japanese female is the lightest [4]. Data from the ARED MILT defines exercise stroke profiles and statistically quantifies exercise frequency. Figure 4 shows a box plot of the exercise frequencies for 3 of the worst-case exercises. The middle dash represents the average frequency, the box envelopes  ±1 standard deviation, and the whiskers represent the extreme value recorded. Either idealized sinusoidal  or real stroke profile data derived from testing can be used to drive the human exercise model. Use of  idealized sinusoidal data tends to be more severe because spectral energy is focused into a singular frequency, whereas real stroke profiles capture the inherent variability in human exercise. The idealized sinusoidal approximation becomes worse at lower exercise frequencies that are typically distinguished by longer pauses between exercise cycles.  The VIS model consists of mathematical definitions of spring and damping rates. Many different concepts were examined and several lessons learned:   1) The enveloping exercises only excited the system in 3 degrees of freedom (DOF) planar motion:  Translation along the y- and z- axes, and rotation about the x-axis. Trading mounting rigidity for Figure 3.  Non-Isolated ARED and Human Model Motion During Squat Exercise  Y  Z  Figure 4.  Box Plot of Frequency Statistics for 3 Exercises
70 isolation by reducing stiffness in only these 3 DOF is both sufficient and provides a certain level of  safety when suspending nearly 567 kg (1250 lbm) in a micro-g environment.    2) Each enveloping exercise excites the 3 DOF differently. Squats and heel raises primarily excite  the y-axis and rotation about the x-axis. Dead lifts, particularly the straight leg variant, tend to excite all 3 axes.   3) Decoupling the VIS DOF by lining them up with the ISS axes made the design of the system  much easier. Changes to spring or damping rates in one axis only affect that axis.   4) Analysis lessons learned 2 and 3 combined with the knowledge that structural dynamics transfer  function response varied in each axis allowed the stiffness and damping of the system to be optimized against the micro-g and structural fatigue requirements.   A key principle in vibration isolation is choosing isolator properties such that the natural frequency of the system is below the forcing frequency. In this case the forcing frequency is the aforementioned exercise  frequency range from 0.09 Hz to 1.18 Hz. The natural frequency,  fn, is given by: 

௡=1 ߨ݇ ݉)ݖܪ(		   Where  k is the spring rate and m is the mass. Additional damping was added to more quickly dissipate  energy in the system. Transmissibility, T, is a measure of the amplification of the isolation system or ratio  of output to input. It is given by: 

݂ ௡൰ଶ ݂ ௡൰ଶ ݂ ௡൰ଶ ቇଶ 

Where f is the forcing frequency and ζ is the ratio of damping to critical damping. Values of T > 1 signify  an amplification of input load, whereas values of T< 1 signify an attenuation of input load. The chosen  spring rates, percent damping, resultant natural frequencies, and amplification at various frequencies are  shown in Table 1 for each of the 3 DOF.  

Table 1.  VIS Properties and Isolation Qualities 

Axis Spring Rate Damping Natural  Frequency Amplification, T  (0.09 Hz  ) Amplification, T  (0.24 Hz) Amplification, T  (1.18 Hz)  Y 70 N/m  (0.4 lbf/in) 11% 0.06 Hz 0.7 0.08 0.01  Z 175 N/m  (1 lbf/in) 10% 0.09 Hz 1.9* 0.19 0.02  X  (rotation) 0.6 N•m/deg  (5.2 in•lbf/deg) 17% 0.05 Hz 0.4 0.08 0.01 

* Amplification based on 0.11 Hz minimum deadlift exercise frequency since the squat exercise creates minimal  excitation in z-axis
71 The spring rates, particularly in the y-axis, had to be extremely low to keep the natural frequency of the  system from overlapping the exercise frequency range and to meet the micro-g and fatigue requirements. The z-axis motion is generally only a fraction of what is seen in either of the other two axes which is why the spring rate and the resultant natural frequency was increased to nearly encroach upon resonance in a worst-case scenario. This was necessary to minimize the motion envelope; one of several other issues that developed as a result of the low spring rates. Inertial accelerations applied to ARED and friction forces also became issues. While ARED is allowed to move in response to exercise, the inside of the ISS is a relatively limited space and there was a possibility that a crew member could hit their head while exercising. Travel limits were setup to prevent this from occurring. Another consideration was the effect of inertial accelerations applied to the ARED system as a result of events like reboost and docking. These events are enveloped by a 0.4 g load factor applied in any direction and cause the ARED to move across the motion envelope and collide with the end of travel limits. This potentially overloads the interface and exceeds a transient acceleration micro-g requirement. Much larger dampers, called snubbers, were sized and located at the travel limits to attenuate those loads. They were selected by calculating the velocity of ARED due to the inertial acceleration and sizing them to efficiently dissipate the energy. Friction had to be considered because, if too high, it had the effect of causing the entire system to “inch-worm” until it reached one end of the motion envelope, potentially reacting repeatedly against the end of travel limits and not re-centering. A special effort was made during the mechanical design phase to select components with low friction as well as to include its effects within the analytical simulation.   Mechanical Design 

To stay within cost and schedule constraints many commercial off-the-shelf (COTS) components are used in the design of the VIS. It is a nearly symmetrical assembly composed of 2 plate assemblies connected by a beam as shown in Figure 5.    Figure 5.  The Vibration Isolation System VISARED Y Plate Z Plate  Rotation HousingInterface to ARED Rotation Arm
72 Each side has half of the linear and rotational bearings that constrain motion to 3 DOF and the isolation  components that attenuate exercise loads. Figure 6 shows that the Y Plate contains linear rails, springs, and dashpots that connect to the Z Plate, allowing motion in the y-axis. The Z Plate has the same components that allow the Rotation Housing to move in the z-axis. Lastly the Rotation Housing has a bearing to allow rotational motion and an arm to convert linear spring and damper forces into torque about the x-axis. The allowable deflection between the left and right plates was specified as ±0.051 mm (±0.002 in) to reduce the possibility of the VIS binding under load. This placed a burden on manufacturing since interface joints and bearing rail surfaces were held to extremely tight tolerances. To minimize misalignment due to accumulated tolerances, there are only 11 VIS structural components from one side of ARED to the other. The VIS was assembled in a fixture to verify alignment and set the spacing of the two halves.  The linear bearings that constrain motion were chosen from amongst several candidates. The DualVee guide wheel design exhibited low friction, robustness, and ease of maintenance. Rails are made from 420C stainless steel while the bearings are made from 440C stainless steel, simplifying the material certification process. The low bearing friction is achieved by using a bearing with shields instead of seals, high quality Rheolube 2000 grease, and just enough bearing preload to maintain linear rigidity. The preload in the bearings is adjusted until the maximum load needed to move a given plate is 1.3 N  (0.3 lbf). The double row angular contact design of the DualVee provides a robust bearing that can handle large static loads. Finally, by separating the bearing elements from the track as shown in Figure 7, maintenance is significantly reduced since dirt and debris attracted to the rails does not come in contact with the ball bearings. They also have a better tolerance for bearing misalignment than other linear bearing systems tested.   Linear rails allow the Z plates  to move in the y-axisLinear rails allow the Rotation  Housin gto move in the z-axis A rotational bearing allows  motion about the x-axisRails Dashpots Springs  Figure 6.  The Left Plate of the Vibration Isolation System Motion MotionMotion  Figure 7.  DualVee Bearing and Rail