A High Frequency Stabilization System for UAS Imaging Payloads

Transcription

1 Proceedings of the IMAC-XXVIII February 1 4, 2010, Jacksonville, Florida USA 2010 Society for Experimental Mechanics Inc. A High Frequency Stabilization System for UAS Imaging Payloads ABSTRACT Katie J. Stuckel, William H. Semke, Nicolai Baer, and Richard R. Schultz Unmanned Aircraft Systems Engineering Laboratory School of Engineering and Mines University of North Dakota Grand Forks, ND A high frequency stabilization mount to compensate for small attitude fluctuations is developed for enhanced imaging and pointing systems on Unmanned Aircraft Systems (UAS). This system consists of a custom camera mount, piezoelectric actuators, and a digital controller to actively control flight vibrations. Payload designs that acquire views of the Earth surface and stationary or moving targets require stable cameras for precision viewing. Placing cameras onboard these small aircraft are vital to the Intelligence, Surveillance, and Reconnaissance (ISR) mission of many payload designs. It is necessary to have real time precision viewing imaging systems while in flight, but it is increasingly difficult as the small planes reach higher altitudes. A slight change in the camera angle at high altitudes results in a large shift from the designated target. This project focuses on high frequency analysis for small oscillations rather than large attitude changes that are accomplished with a gimbal. The two systems work together to handle both the high frequency oscillations due to engine vibration and turbulence as well as large low frequency attitude changes. Results will include laboratory testing and simulation data to further prove the effectiveness of this specialized stabilization system. NOMENCLATURE α T = Coefficient of Thermal Expansion in per Kelvin ( /K) α = rotation about x-axis of stabilization system β = rotation about y-axis of stabilization system θ = pitch φ = roll E = Modulus of Elasticity in Pascal (Pa) = displacement in the z-direction, usually pertaining to actuators δ = displacement due to expansion/contraction of actuator in meters (m) or micrometers (µm) F = Force in Newton (N) L = length in meters (m) x = x axis (or coordinate) of the stabilization mount y = y axis (or coordinate) of the stabilization mount z = z axis (or coordinate) of the stabilization mount X a = x axis of the plane coordinate system Y a = y axis of the plane coordinate system Z a = z axis of the plane coordinate system X c = x axis of the camera coordinate system Y c = y axis of the camera coordinate system Z c = z axis of the camera coordinate system ω n = natural frequency in radians/second 2 (rad/sec 2 ) f = frequency in Hertz (Hz) k = stiffness of spring in N/m k eq = equivalent k component for natural frequency calculation m = mass in kilograms (kg) V app = applied voltage = maximum voltage V max

2 Acceleration (m/s 2 ) I. Introduction A high frequency stabilization mount is presented for use with unmanned aerial systems (UAS) to enhance imaging and pointing systems during flight. The active vibration control mount uses piezoelectric stack actuators to move the camera mount as needed. As the plane vibrates, the sensors onboard detect the slight change in tilt or displacement, send a signal to the actuators, causing them to move in the opposite direction for stability. It is the goal of the vibration compensator to steady any imaging or pointing system used during flight. This allows for precision viewing and higher quality systems onboard. Figure 1 shows the platform and the prototype of the stabilization system which will be flown. The Bruce Tharpe Engineering (BTE) Super Hauler is a custom built small unmanned aerial vehicle (UAV) designed for multiple types of payloads and flown by the Unmanned Aircraft Systems Engineering (UASE) laboratory. One of the primary payload development areas of the UASE is target imaging and tracking for use in Intelligence, Surveillance, and Reconnaissance (ISR) missions [1,2]. The active vibration control mount enhances these categories and fits into the body of the plane with the camera nadir pointing towards Earth. Figure 1. Unmanned Aerial Vehicle High and Frequency Vibration Mount Precision pointing is vital to certain missions flown on UAVs for real time viewing and post analysis. The high frequency vibrations from the engine, as well as many other sources including wind buffeting, prove to be a challenge during flight. Image streams become grainy and unfocused the more the UAV shakes. The camera also moves and tilts, sometimes losing track of the target. This becomes more common as the UAV gains altitude and speed. Because of the slight tilt in the camera, the projected view has a large shift in the pointing of the camera. When zoom is also included in the system, the shift may be enough lose the target in the field of view completely, causing a failure in the mission. It is the goal of the high frequency stabilization mount to rectify this problem and help improve imaging and pointing capabilities. To begin development of the stabilization mount, it was necessary to analyze accelerometer data from a UAV during flight. Figure 2 shows example data from the Piccolo II autopilot, by Cloud Cap Technologies in which a range of the acceleration in the vertical direction is observed. The data is centered about Earth s gravitational constant at a value of m/s 2 and at any given time during flight data analyzed, the typical magnitude of acceleration is less than one g (9.806 m/s 2 )[3]. Based on a FFT analysis of the data, one can see the need for a high frequency stabilization system as the plane shows an unstable platform in the high frequency range. This vibration may interfere with other systems onboard or cause complications in imagery. 0-2 July 31, 2008 August 1, Time (sec) Figure 2: Sample acceleration data in the vertical direction comparison between Jul 31 and Aug 01 flights

3 The designed system with its coordinate system, angles and orientations is shown in Figure 3. The active vibration control mount is defined by aligning the coordinate system of the mount with that of the airplane. This simplifies the algorithms used by keeping the coordinate systems aligned and translating the axis of the vibration mount along the X a -axis of the plane coordinate system. The X a -axis is the roll axis of the plane while the Y a -axis is the pitch axis of the plane. The Z a -axis then becomes the direction of interest in which the actuators move. The vibration mount works in tandem with a gimbal system for the large angle rotations which the small, while high frequency angles are controlled with the active vibration control mount. x y z Figure 3. Stabilization mount and aircraft coordinate system The intent for this small, lightweight system includes high voltage DC-DC converters, software written in C code, a digital controller and displacement sensors to complete the active vibration control stabilization mount. As the displacement sensors detect a change in position, signals will be set through the controller to the actuators instantaneously to counteract this movement. In depth laboratory testing and simulation is used to support the analysis of the high frequency stabilization mount. Accuracy of the system is based on the measure of current deflections and rotations, with a future measure of the stability of the support ring. Both laboratory and simulation testing results are compared to hand calculations to demonstrate the effectiveness of the system. II. Description and Design of Model The high frequency stabilization mount is made from 6061-T6 aluminum for stiffness and weight attributes. Since the platform for use of the mount is a UAV, it is necessary to make the stabilization mount as light as possible. Table 1 also shows the dimensions of the piezoelectric compensators to show the use of active vibration control on a small scale design. A primary use will be for mounting cameras up to 2 kg. Mathematically, a plane can be defined by three points, which allows the system to properly orientate to any rotation angle. From this plane, a normal vector can be found which we will define as the line of sight vector. This is placed on the z-axis of the viewing angle of the camera. The goal of the system is to move the actuators in the z direction on the stabilization mount to keep the camera z c -axis pointing steadily at the target [4]. The actuators used in this system are purchased from APC International LTD with specified maximum stroke of 20 µm with 0 to +150 V, as shown in Table 1. These actuators are able to supply and receive a load up to 800 Newton and are vital in the high frequency stabilization mount design. The three actuators located at equal distances around the ring provide the necessary movement for prescribed rotations to help actively control the vibration. A static deflection test was performed on four actuators to verify the specifications. From this test, it was determined that the actual stroke of the actuators is greater than the specified deflection. Figure 4 shows the deflection of the actuators versus the amount of voltage applied. The predicted value is also plotted on this graph. The first three actuators tested were very similar in response, while the fourth actuator showed a greater distance traveled. Because of this, the three similar actuators are used in the stabilization mount to increase the reliability of performance. The updated actuator curve is used in the analytical calculations and ANSYS models to better match the experimental results. This is further explained in the Analytical Calculations and Mathematical Modeling section.

4 Displacement (µm) Table 1. Mechanical and Electrical Specifications for APC International Piezoelectric Stack Actuators Mechanical and Electrical Specifications Length 3.5 mm Width 3.5 mm Height 18 mm Voltage Range 0 to +150 V Max Stroke 20 µm Capacitance 200 nf Resonance Frequency 50 khz Stiffness 25 N/µm Blocking Force 800 N Max Load Force 800 N µm 1 Predicted Value µm 2 µm 3 µm Voltage Input (V) Figure 4. Comparison of actuators by user voltage input A ring was chosen for the mount to help simplify the design and reduce material. It is necessary to have a hollow shape to mount the camera without blocking the field of view, but also necessary to have sufficient material for mounting and allow for ease of manufacturing. It is critical to have spring/actuator pairings such that the ring remains firmly in place against the actuator so no impulsive loads are transmitted and the ring and actuator motion remain in phase. The rods next to each spring and actuator set allows for the system to move up and down freely while constraining the motion to the vertical direction. The springs have a stiffness of 63,550 N/m. The stiffness for the springs was chosen based on the systems natural frequency. Data is received from the sensors and fed into the system at 100 Hz, so it is important to have the natural frequency of the system, as well as the natural frequency of each set of spring and actuator greater than this to ensure the appropriate response. The natural frequency can be found using Eq. 1 and the conversion from ω n (rad/s 2 ) into Hz. (1) The k eq required needed to be stiffer than 27,240 N/m. With the chosen springs based upon availability and acceptable dimensions, the natural frequency of the system is calculated at Hz.

5 III. Analytical Calculations and Mathematical Modeling To better understand what is happening with the high frequency stabilization mount, it is necessary to perform analytical calculations to compare with simulated and experimental results. To calculate the deflection of a single actuator, the ratio of the voltages simply needs to be multiplied by the maximum deflection, Eq. 2. (2) For calculations of the deflection at the actuator of the system including the spring, it is necessary to taken into consideration the stiffness of the actuator as well. As shown in Table 1, the stiffness of the actuators is 25 N/µm. This in combination with the deflection of the actuator at the designated voltage and the deflection equation for a spring (Eq. 3) will give the final deflection as the actuator pushes against the spring. Displacement of actuator including stiffness is shown in Eq. 4. (3) (4) Combining these two equations and solving for the force will give the final results for the deflection. By setting spring = act, we are able to achieve this. Since the spring and actuator are working against each other, equilibrium will be reached and the force will be the same in each. Putting the found force back into Equation 3 or 4, the final deflection can be found using Eq. 5. (5) At 100 V, the calculated force is N, which gives a final deflection of µm for each actuator. For simulation and experimental results, this will be the comparison standard. IV. Finite Element Analysis A finite element analysis and simulation was done to verify the equivalent rotation and displacements of the spring, as well as complete a modal analysis to confirm the natural frequency of the system. The ring, actuators and springs were all modeled in ANSYS, with the appropriate geometry, material definitions, properties and element settings were all set according to each item. The model was meshed and the necessary constraints applied for each analysis. The analyses performed included both static and a modal analysis. To begin, the ring mount was modeled using SOLID 45 elements, which is an 8-node 3-dimensional structure. The spring was modeled using the COMBIN 14 element with a spring constant of 63,5000 N/m and a damping constant of 0. The piezoelectric actuators were modeled using LINK 1. To allow for the actuators to act as a piezoelectric material with a voltage applied, the analysis in ANSYS was performed using a thermal analysis, comparing the rise in temperature to the rise in voltage. Equations 6 and 7 were to find the necessary constants, E and α T, to simulate the piezoelectric expansion. (6) (7) Table 2 has complete details on material element types, real constants and material properties. The ring was modeled using keypoints 120 at equal radius apart from the origin. Lines were created from the keypoints, areas from lines and volumes from areas. Meshing was completed of the ring with 60 elements on each arc, 20 elements in the width and 5 elements in the height. The meshing used was a tetrahedral, 4 node mesh. Displacement constraints were applied in all directions and all rotation to the tops of the springs and bottom of the actuators, allowing the ring to move freely up and down, simulating what is happening in the actual system. Figure 5 shows the meshed structural and boundary conditions.

6 Table 2: ANSYS properties and element types Element Type Real Constants Material Properties SOLID 45 N/A Density: 2700 kg/m 2 Modulus of Elasticity: 69 GPa Poisson s Ratio:.36 COMBIN 14 Spring Constant: N/m N/A LINK 1 Cross Sectional Area: Modulus of Elasticity: GPA x 10-5 m 2 Poisson s Ratio:.33 Thermal Expansion Coefficient: x 10-6 /K Figure 5. ANSYS constraints and meshing for high frequency stabilization mount Since this simulation is used to compare to the experimental and hand calculations, only one actuator was activated in ANSYS, and was done so at 100 V. Tests were performed both with the springs attached and without. The results from the final ANSYS simulation with the springs is in Figure 6 and a summary of results are listed in Table 3. The first tests without the springs showed very promising results with percent errors of 0% and 0.07%. The predicted results with the springs calculate µm to be the deflection at the actuator with the springs and ANSYS gives µm. Because of the nature of the system, it was not possible to measure the displacement at the exact location of the actuator, so measurements had to be taken at the edge of the ring, also shown in Figure 6. The parallel triangle theorem was used while the ring was assumed to rotate about the fixed axis of the other two actuators. This gives a displacement of µm at the actuator. These similar deflections, 0.0% and 0.27% errors also show promising results for the tilt compensator. Figure 6. ANSYS results with one actuator activated and experimental set-up (laser measurement point highlighted red)

7 Table 3. Summary of Simulated, Experimental and Hand Calculation Deflection of Stabilization Mount (µm) Displacement ANSYS Experimental Calculations % Difference % Difference w/o Spring % 0.07% w/spring % 0.0% V. Modal Analysis The next analysis performed was the ANSYS modal analysis to compare the numerical natural frequencies to those predicted. In this analysis, only the ring and springs are included in the model. Using Equation 1, the analytical calculations showed the natural frequency to be Hz. As shown in Figure 7, ANSYS calculates the axial natural frequency to be Hz. This gives an error of 1.3%, which is determined to be acceptable. Table 4 shows a list of the first eight natural frequencies of the system calculated, listed from 0 to 5000 Hz. Modes 1 and 2, 4 and 5, and 6 and 7 are paired symmetric natural frequencies. The second mode in each pair is an orthogonal rotation of the first mode in each pair. The lower two natural frequencies show a tilt of the support ring and the axial natural frequency is the third mode shape. Higher mode shapes show a ring deflection for the shape. Figure 7 illustrates the first four mode shapes showing only one of each set of modes. Table 4: Natural frequencies for the stabilization mount Mode Frequency (Hz) a. b. c. d. Figure 7: a. Mode 1, f= Hz b. Mode 3, f= Hz c. Mode 4, f= Hz d. Mode 6, f= Hz

8 VI. Laboratory Testing The previous static laboratory testing data presented on the ring resulted from supplying voltage to the actuators using a Hewlett-Packard Lab Power Supply and measuring the deflection of the spring with a Keyence Laser Displacement Sensor (LDS). A more in depth system used for testing included the use of a LabVIEW program and a National Instruments USB-6251 Data Acquisition (DAQ) Board for communication with the actuators. This DAQ board allows a range of 0 to +10 V and communicates through USB. As the LDS measures the displacement of the ring and sends the data through the DAQ board, the LabVIEW program corrects for this displacement and supplies a voltage through a TREK Piezo Amplifier to the actuators. The gain on the amplifier is set to 1:100 V. The actuators then expand or contract to the necessary height, seemingly that the stabilization mount never moved. The first round of static testing was performed by activating a single actuator. It was very important to isolate the system and separate the ring mount and laser head from the other devices, which cause noise in the system. For example, the fan in the laptop computer and the piezoelectric amplifier cause small motions in the table, creating an unstable reading on the laser head. Figure 8 shows the set up of the experiment. Figure 8. Static experimental set-up: isolated mount and separate power supply on the left, computer, amplifier, DAQ and laser head controller on right A static test was performed by deflecting the base of the high frequency stabilization mount and running the LabVIEW program such that the control system corrected for the known displacement. An open loop controller was used to perform a single calculated correction to the system. In this manner, the entire system was exercised including both hardware and software components. The LabVIEW block diagram is shown in Figure 9. The open loop control system allowed for end-to-end testing and communication between all hardware, as well as demonstrating the ability of the LabVIEW program. The initial value was entered manually into the program in the Initial Value on the block diagram. After the base of the vibration mount was given a deflection, the inputs were read from the laser with a DAQ board through the DAQ Assistant vi. The constant is entered into the program to calibrate the sensor. The data is then processed through the system by calculating the difference in the laser voltage readings and multiplying that value by , a constant gain that is the ratio from µm to V for the piezoelectric actuators. That value is then added to 0.2 V, which was selected as the starting voltage for the experiment. The starting voltage partially activates the actuators at 20 V, which is equivalent to 2.96 µm, to allow for motion in the positive and negative directions. Finally, the voltage, with a safety limit of 1.4 V, is fed into the DAQ board (140 V from the amplifier to the actuator). The check is done because the maximum voltage allowed per actuator is +150 V. This test was repeated five times for verification purposes and the results are shown in Table 5. The average deflection of the base is micrometers, with an average error of 2.66%. The low percent error for this experiment shows that this experiment was a success. The actuators proved to respond appropriately, as well as demonstrated a fully operational LabVIEW program. Table 5. Laser readings for static deflection test results Initial Value (mm) Deflection Value (mm) Return Value (mm) % Difference Average % Difference % 2.66% % % % %

9 Figure 9. Open loop LabVIEW program for stabilization mount VII. Conclusion The data obtained in these experiments demonstrates the effectiveness of the high frequency stabilization mount prototype. Analytical calculations, computer simulations and static laboratory experiments have all shown the reliability of the system. The modal analysis in ANSYS simulates the various modes and frequencies for the system and gives a range of usable frequencies. Future work for the stabilization mount includes the addition of a Proportional, Integral, Derivative (PID) closed loop controller to eliminate the high frequency dynamic oscillations. The finalized high frequency stabilization mount will be a miniaturized version of this prototype, implementing micro DC-DC amplifiers and a programmed digital controller to produce a small lightweight system that will be operated autonomously onboard a UAS. VIII. Acknowledgements This research was supported in part by Department of Defense contract number FA C-C006 Unmanned Aerial System Remote Sense and Avoid System and Advanced Payload Analysis and Investigation, and the North Dakota Department of Commerce, UND Center of Excellence for UAV and Simulation Applications. The authors would like to also acknowledge the contributions of the Unmanned Aircraft Systems Laboratory team at UND. IX. References [1] Semke, W., Schultz, R., Dvorak, D., Tandem, S., Berseth, B., and Lendway, M., Utilizing UAV Payload Design by Undergraduate Researchers for Educational and Research Development, Proc. of 2007 ASME International Mechanical Engineering Congress and Exposition, IMECE , November [2] Lendway, M., Berseth, B., Tandem, S., Schultz, R., and Semke, W., Integration and Flight of a University- Designed UAV Payload in an Industry-Designed Airframe, Proc. of the AUVSI, [3] Semke, W., Stuckel K., Anderson K., Spitsberg R., Kubat B., Mkrtchyan A., Schultz R., Dynamic Flight Characteristic Data Capture for Small Unmanned Aircraft, 26 th annual IMAC SEM Conference, Feb 08. [4] Buisker, M., Statistically Significant Factors that Affect the Pointing Accuracy of Airborne Remote Sensing Payloads, M.S. Mechanical Engineering, University of North Dakota, May 2007.