Active vibration isolation system

Abstract: An active force actuated vibration isolation system uses a velocity geophone to sense payload velocity and modifies the geophone signal to control the force applies to oppose the vibration. The modified signal effectively reduces the inherent resonant frequency of the geophone and avoids instability.

Description: This invention relates to a vibration isolation system as indicated in the precharacterizing part of claim 1 In such a prior art (FR-A-2 200 457) system the motion-responsive means for producing the output signal is a accelerometer which requires an integrator in a sensor signal modifier to convert the sensor signal which measures acceleration of the payload mass to a signal which is proportional in magnitude and sign to the absolute velocity of the payload mass. The combination of passive and active vibration control systems has been analyzed in a paper by Guntur and Sankar entitled "Fail-Safe Vibration Control Using Active Force Generators" published by the American Society of Mechanical Engineers in the transactions, Journal of Mechanical Design, Paper No. 81-DET-85. The generalized form of a combined active and passive system is illustrated in Fig. 3 of this article and the transmissibility characteristic of a velocity analog force feedback system is shown in fig.4 for various values of gain. The description is with reference to the theoretical "sky-hook" damper without reference to the sensor for derivation of the velocity signal or the requirements for maintaining stability throughout the frequency range of operation. EP-A-032 158 discloses a vibration controller for a vehicle, which incorporates cylinders for suppressing the vertical vibration of an automotive body. Detection signals from vibration acceleration detectors are compensated by compensating circuits, the outcome signals of which are further amplified through amplifiers and then applied to servo valves, which control the internal pressure of the cylinders. The compensating circuits have transfer functions forming the product of an integrating element, a primary leading element and a primary lagging element in order to avoid any peak over a wide frequency range in the response of the vibration acceleration of the automotive body. It is the object of the invention to improve the vibration isolation system as indicated in the precharacterizing part of claim 1 such that a high degree of stability over the required frequency range is maintained. This object is solved by the feature as claimed in the characterizing part claim 1. The present invention provides a vibration isolating system that operates in response to a velocity sensor providing a signal representative of velocity to an electronic circuit which is adapted to provide extremely high gain over all frequencies above the very low frequencies near zero frequency with provision to avoid instability at all frequencies by reducing the gain value to substantially less than unity at phase angles of ±180°. The system operates in the preferred form relative to a passive vibration isolation system thereby minimizing the requirements for substantial force application components and the attendant bulk and expense of power amplifiers for their actuation. The system isolates the payload from the residual vibration that is transmitted by a conventional passive system with a consequent absolute vibration value far below that which is applied to the system by transmission of base motion or applied inertial forces. The velocity feedback force actuated vibration isolation control system provides a vibration transmission attenuation (i.e. vibration isolation) at substantially all frequencies and the absence of any substantial resonance response which results from the achievement of a high degree of stability at all frequencies such that base vibration transmission to the payload and oscillatory response to impulse loading are reduced by an order of magnitude or better relative to a passive isolation system of the same resonant frequency. The term "passive isolation system" as used herein refers to a conventional system designed specifically to remove seismic vibrations including industrial environment induced vibrations. Embodiments of the invention are explained in detail with reference to the drawings which include: Fig. 1 is a schematic representation of a combined passive and velocity feedback force actuated vibration isolation system. Fig. 2 is a partial schematic of a signal feedback circuit used in the system of Fig. 1. Fig. 3 is a schematic diagram of a complete feedback circuit for the system of Fig. 1. Fig. 4 is a representation of a mass supported for motion with six degrees of freedom with corresponding active force application for vibration control. Fig. 5 is a top plan view of a platform having six degrees of freedom stabilized in accordance with the invention. Fig. 6 is a elevation view partly in section of Fig. 6. Fig. 7 is a transmissivity curve showing the passive resonant response and the response obtained with the active system of this invention in the embodiment of Fig. 6. Fig. 8 is a response curve for an impact loading of the payload in the system of Fig. 5. Fig. 9 is a sectional view showing the arrangements of parts in an electromagnetic force transducer. Fig. 10 is a schematic diagram of the application of the invention to a pneumatic force servo system. Referring to Fig. 1 a single degree of freedom velocity feedback force actuated isolation control system will be described. In accordance with conventional representation for vibration isolation analyses the system shown in Fig. 1 has a payload mass M, 11, supported with reference to a support structure 12 by a resilient spring 13 having a spring constant K and a viscous damper 14 having a damping coefficent C. The system is subject to disturbances which include: (a) base motion transmitted typically through support structure 12 to the payload mass 11, and (b) external forces applied directly to the payload mass 11. Also shown in Fig. 1 is a position feedback system which is generally responsive with slow response to the changes in the position of support 12 designated by quantity U and positions of the payload 11 designated by the quantity X, all in the vertical direction, which is the assumed direction of the degree of freedom involved, such that the output of a position servo 15 maintains spring 13 at approximately a predetermined compression level. For example, the system as thus far described could represent a work table isolated by passive isolation means 13-14 and maintained at a working level that is adjusted for changes in gross weight or the like. In accordance with the present invention the payload mass 11 supported by support system 15 and passively damped by elements 13 and 14 is further isolated from vibrations whether they occur due to vibration transmitted from the support surface 12 or from impact or inertial forces that operate directly on the payload mass 11. This further isolation is achieved by a velocity feedback circuit stabilized to provide a counteracting force application applied to the payload mass 11 without the instability associated with oscillation, ie. a realizable system. The basic system of the invention comprises a velocity sensor 21 for developing an electrical signal representing an approximation of the velocity of payload mass 11 in the direction of the degree of freedom of motion. The velocity signal is applied through an electronic circuit 22 providing the necessary transfer function for controlling a force application unit 23 which applies a force F v to the payload mass M. The feedback from sensor 21 through circuit 22 to force generating unit 23 is negative such that the application of the force F v tends to cancel the velocity component which produces the output signal from transducer 21. The selection of the velocity sensor 21 can be made from various known devices such as displacement, velocity and acceleration sensors per se where the displacement measuring transducer would require that the measured quantity be differentiated to obtain velocity and the accelerometer would require integration to obtain the velocity quantity. Both of the sensors requiring this extra processing step have other limitations in terms of sensitivity and threshold response. In accordance with the invention the preferred form of transducer 21 is a unit known as the velocity sensor geophone. Velocity sensitive geophones measure the relative velocity between the unit base or case of the sensor and a suspended mass called the proof mass. The sensor consists of a coil of wire containing many turns of very fine wire suspended on very linear and compliant springs within a magnetic field provided by a core and permanent magnet assembly. The motion to be measured is the velocity of the payload 11. It was discovered however, that the electro-magnetic sensor actually measures the relative velocity of the proof mass relative to the case. The relative motion of the proof mass and case are determined by the dynamics of the suspended mass. The geophone output voltage E v is proportional to the relative velocity between the geophone case and the geophone coil (proof mass). As will be readily understood by one skilled in this art, the sensor's electrical voltage response can be approximated mathematically, using Laplacian Notation, as follows: (I) E v (S) = β S (X-Z) = βS S 2 + (SX) 2ξg w g S + w g 2 where β ≅ gain in units of volts per cm per second and is derived by the geophone calibration. S ≅ Laplace Operator X ≅ displacement of geophone case. Z ≅ displacement of proof mass S (X-Z) ≅ relative velocity of proof mass to geophone case ξg ≅ Geophone fraction of critical damping w g ≅ proof mass resonant frequency The geophone output response vector magnitude and associated phase angle  are related to the actual velocity of the geophone case which is to be measured, SX. Converting Equation (I) to the frequency plane, the actual, geophone output can be stated for our purposes as (where SX represents X ˙ in the time domain) (II) E v (jw) ≅ β X · w 4 (w g 2 - w 2 ) 2 + (2ξ g w g w 2 ) where w ≅ frequency X ˙ ≅ velocity to be measured Equation (II) yields the output's vector magnitude and Equation (III) yields the associated phase angle. The above frequency response equations indicate that the geophone response output signal can peak at the resonant frequency, w g , if the fraction of critical damping is small. The fraction of critical damping is controlled by the magnitude of viscous damping. The viscous damping provided for the control of the output response at the resonant frequency comes from two sources. First, the proof mass coil suspension has some natural and inherent viscous damping caused by eddy currents flowing as the coil moves in the magnetic field. The inherent damping is not small and typically gives rise to a fraction of critical damping of about 0.3. Additional damping is easily added by placing a resistor across the output of the geophone. The presence of the damping resistor causes current to flow through the coil of wire as voltages are induced by coil movement in the magnetic field. The magnitude of this additional damping can be quite large. Preferable, the damping resistor is such that overall fraction of critical damping is unity, or ξ g = 1.0. With this unity value the output voltage will not show a resonant peak and will exhibit a smooth curve transition into the operation region above the frequency w = w g . In addition, for ξ g = 1.0. the transfer function given in Equation (I) may be factored into two poles as shown below: Importantly, the geophone output response modeled by this transfer function of Equation (IV) is not, as one can easily see, equal to the product of the gain and the measured velocity of the case. This product is further multiplied by the bracketed term to yield a more accurate and usable approximation of the actual geophone output. This transfer function in its factored form, can be used to design an electrical processing network whose output can be fed back to provide active vibration isolation. Such a network will be described below. Commercially available geophones include a wide variety of units with a wide range of signal capability, frequency response and cost. Typically the lowest frequency models are the most expensive. For applications in accordance with the invention it is desirable to have both small size and relatively low cost. Geophones for sensing velocity are commerically available which have response characteristics as described mathematically in Equation (IV). One suitable geophone is the model HS-1 manufactured by Geo Space Corporation, Houston, Texas. The other transducer required for force actuator systems is the force generating unit 23 shown in Fig. 1. One practical form of this actuator is a modified high-fidelity loud speaker assembly chosen for its characteristics and availability to operate from a control system circuit using conventional integrated circuit elements. With these requirements in mind an assembly such as that illustrated in Fig. 9 was used. Fig. 9 shows a conventional speaker magnet voice coil assembly comprising a steel core 91, a permanent ferrite magnet 92 and an annular steel yoke 93 providing with the upright extension of core 91 and an air gap 94. Mounted for movement within the air gap 94 is a non-metallic cylinder 95 on which a multiturn coil 96 is wound. The construction should provide: 1) force output that is proportional to an input current. 2) substantially frictionless motion in all axes of movement. 3) little added spring stiffness to the spring rates provided by the passive vibration isolation system. An intergrated circuit power amplifier can meet the linear response requirement, number (1) above. These integrated circuit power amplifiers are small and inexpensive and very desirable to use when possible as final power output stages of a control system's electronic circuitry. For power level above twelve watts, or two amps output current, discrete power amplifiers must be made up from individual power transistors which greatly adds complexity and cost to the electrical control system circuitry. Requirement No.2 is quite important, since friction force levels can act as force noise sources leading to a higher payload output response level. Perhaps this can be best understood by way of example, consider, using a 45.4 Kg (100 pound) payload and an actuator for force generation having a friction force of say 1/45.4(100) of one Kg (pound), could not control vibration acceleration response levels at the system output below the acceleration level in G units of A g = F/W = 0.01/100 or 10 -4 g units. However, the desired response level of the isolated payload is one Micro-G or less. To obtain a vibration level of less than one Micro-G, the force actuator friction level must be less then 1 x 10 -6 Kg (pounds) per 45.4 Kg (100 pounds) of payload weight that the actuator acts upon. Thus the friction requirement for the force actuator is extreme, essentially the actuator must be frictionless. The most effective manner to produce a substantially frictionless force transducer is to use a magnet and coil assembly from a high-quality loud speaker with a wider than normal gap for the coil to move within. One example, which worked acceptably was a magnet used with a 25.4 cm (10 inch) high fidelity woofer loud speaker. One skilled in the art should understand that a commercially processed loud speaker may have to be adapted for use in accordance with the present invention. For example, the existing gap width might have to be increased. After magnetization a voice coil was placed in the field to obtain a calibration scale factor for the force actuator. The calibration scale factor is an important design parameter yielding the overall loop gain and maximum force capabilities for the actuator. Typically, such actuators are capable of providing 0.9 Kg (two pounds) of force per ampere of input current. In the assembly of a velocity feedback, force actuated vibration isolation control system either the magnet or the coil may be rigidly attached to the payload with the remaining component rigidly attached to the isolation system base or cabinet structure. Having discussed the sensing and actuating means, it is now timely to discuss the operation of the vibration isolation system. Critical to desired operation is a realizable system. This term commonly used in servo-mechanical analysis, denotes the condition wherein the system is stable. To determine system stability, it is necessary to analyze and modify the system open loop transfer function. In accordance with the present invention the stability of the velocity signal feedback loop is assured by selection of the transfer function of the network 22 connecting the geophone 21 to the force transducer 23. Stability in such systems can be analyzed by opening the loop and determining the output voltage E o in response to a given input voltage E in as shown in Fig. 1, the ratio of which being called the open loop transfer function. The switch 24 is shown in Fig. 1 for purposes of this analyses and is not ordinarily required in an actual system. If the network 22 has a transfer function which merely represents a high gain amplifier, instability results. Such instability is generally associated with excessive gain and ±180° phase shift at some frequency of operation. The velocity sensed by the velocity sensor 21 will include the amplified vibration due to the resonance point of the passive vibration of the isolation system 13-14 and this frequency generally is near the resonant frequency of the geophone 21 itself. Finding the proper transfer function for the network 22 requires compensating for this effect. Analyzing the phase and gain relationships shows that the amplification and phase shift at resonance of the passive vibration isolation system occurs at a frequency slightly lower than the geophone natural frequency. The geophone is very highly damped, as is explained above, with ξ= 1.0, and thus has no resonance amplification. If the geophone resonant frequency could be placed well below, say by a decade, the passive system resonance then the stability state will be vastly different. With a geophone resonance one decade or so lower in frequency than the passive system resonance, the open loop transfer function goes through nearly 180° of the phase lag prior to the frequencies at which gains greater than unity exist, thereby allowing the open loop transfer function to pass through the ±180° phase point with a gain substantially less than one. Typically, the gain at said point has been maintained at approximately 0.3. For a low resonance frequency of the geophone relative to the passive system resonance, much more gain may be applied and still have a stable system. In general, the higher the velocity feedback gain, the higher the degree of vibration isolation. The highest possible gain is sought that will still yield a stable and well behaved active vibration isolation control system. The process of closing the loop and obtaining a higher gain is called statilization. To do this requires acting on the geophone electrical signal by linear or non-linear electrical networks as will now be described. With reference again to fig. 1. the transfer function of the geophone, F(S), may be factored as shown below for the condition of critical damping ξ g = 1.0 (V) F(S) = S 2 (S + w g ) (S + w g ) Let the term w g or the geophone resonant frequency be expressed as a time period called a time constant such that T g = 1 / wg. Placing this term into the geophone transfer function and rearranging terms gives the expression: (VI) F(S) = (T g ) 2 S 2 (1 + T g S) (1 + T g S) A modified transfer function may be obtained by feeding the geopone electrical signal through an electronic network such as shown in fig. 2. This network is two lag-lead networks using, for purposes of example, two operational amplifiers 25, 26 with feedback networks comprised of the two resistors R f and R g and the capicator C. The transfer function g(S) = E out /E in for the circuit in fig. 2 will now be determined. In this circuit R f , R g , and C may be chosen such that the value of R p C (where R p is the parallel resistance of both the feedback. R f , and ground, R g resistors) is equal to the geophone time constant T g . In addition, the value of R f is selected such that R f C has a value ten times greater than R p C. With these values, the transfer function for the electronic compensator circuit 22 can be approximated as: Now placing the above transfer function g(S) in series with the geophone output gives the compensated geophone transfer function shown below: which simplifies by pole, zero cancellation to: (VIII) F(S) g(S) = 100(T g ) 2 S 2 (1 + 10 T g S) (1 + 10 T g S) The compensated response of the geophone is now exactly equivalent to that of a geopone having a resonant frequency that is ten times lower than the actual value. With the addition of the geophone compensator, the phase shift of the feedback signal takes place prior to the resonant frequency and motion amplification of the passive system, thus the entire control system now has low gain when the phase angle is close to ± 180°, and high gain where the phase angle is close to 0°. The low gain, for example, has been maintained at less than 0.35 and the high gain at greater than 10. The feedback circuit as analyzed for fig. 2 was applied to the velocity force feedback systems of fig. 1 with further modifications as now will be described with reference to fig. 3. In fig. 3 the geophone 21 and the force actuator 23 are identified as the elements previoulsy described in fig. 1. The coil of the geophone 21 is connected across a damping resistor 31 as previously described. The damped output of the geophone 21 is one input to an operational amplifier 32 the output of which is applied to a potentiometer the tap of which supplies the other input of operational amplifier 32 thereby controlling the gain of the velocity signal. The output of the gain control operational amplifier 32 is applied through a first-lag network 33 to a first lag-lead circuit 35 corresponding to circuit 25 of Fig. 2. The output of amplifier 35 is applied to a first lead network 34 which supplies an input to an operational amplifier 36 corresponding to amplifier 26 of Fig. 2. The output of amplifier 35 is passed through a second lead network 37 to a third lag-lead circuit 38(similar to 35 and 36) the output of which passes through a second lag network 39. The output signal from network 39 is applied as the input to a power amplifier 40 which in turns drives the coil of the force transducer 23. As can be seen in Fig. 3, the preferred feedback circuit includes three phase rotation elements. While the addition of geophone compensation shown in Fig. 2, provides for much more velocity feedback gain, it was experimentally found that the addition of a third lag-lead compensation stage 38 in Fig. 3 allows even more gain to be applied. Importantly, the additional phase rotation element maintains the overall control system resonance transmissibility to even smaller values, about 1.10 (as shown in Fig. 7), compared to the two element circuit. The loop gain used in the control system is adjustable by means of potentiometer 42 and can be set such that the gain at a phase shift of ±180° , called the gain margin, is 0.35, and the phase shift at unity gain called the phase margin, is ±45° from ±180° . For this condition the basic system stability state is excellent. At high frequencies where minor resonances in some part of the entire system can cause gain amplification and near unstable conditions the lag networks 33 and 39 in the feedback loop lower the loop gain above 200 cycles per second. In addition the high DC gain associated with the lag-lead networks, which causes large DC offset voltages due to errors in the DC operating level of the operational amplifiers is avoided by the two lead networks 34 and 37 placed in series in the feedback loop to block DC and very low frequency signals. Referring to Fig. 4 a mass used as a platform 41 for vibration isolation has six degrees of freedom. These six degrees of freedom are shown in Fig. 4 as motion corresponding to vertical, roll, pitch, horizontal long axes, horizonal short axis, and yaw. All possible motions of the mass can be described by motions in these six degrees of freedom through translation along and rotation about the three mutually perpendicular axes. Achievement of low vibration level for the platform 41 using an active vibration isolation system requires active vibration isolation controlling all six degrees of freedom. This may be done by using three independent translatory control systems and three independent rotational control systems. However, rotational control systems are hard to implement mechanically both from a sensor and actuator point of view. Thus, translatory control systems only are used to obtain the same goal of control of all six degrees of freedom. In Fig. 4 typical points of application for six linear force actuations, 1,2,3,4,5 and 6 are shown. Each force actuation results from an actuator and sensor combination and corresponds to a channel. In Fig. 4 channels 1 and 2 control part of the vertical axis, and all of the pitch axis. Channels 1,2 and 3 control all of the vertical axis and all of the roll axis. Channel 4 controls all of the horizontal long axis and channels 5 and 6 combined control all of the horizontal short axis and all of the yaw axis. Ideally, motions in any one axis are completely independent of other motions in other axes, such as for example if a displacement occurs on the vertical axes 1,2, 3 combined, no other motions or rotations occur on any other axis. This independence of degrees of freedom however, is nearly impossible to achieve in a real world system. In an actual system constructed, the three vertical channels are nearly independent in that a combined motion of channels 1, 2 and 3 generate almost pure vertical translation with little horizontal or rotational motion. Control of pitch and roll is also quite good, in that the combined motions of channels 1,2 and 3 to generate a pitch or roll motion only will generate little motion in yaw or the two horizontal directions. However, motion generated by the actuators on channels 4, 5 and 6 generate axis coupling in that motion of channel 4 generates some pitch and motion by either channel 5 or 6 generates some roll as well as yaw motion. The location of the force actuators in the horizontal plane was selected to act at points on the payload mass that minimized the coupling effect. Another form of coupling is caused by gravity and exists only on the horizontal channels. When a horizontal HS-1 geophone is exactly level with respect to gravity, there is no gravitional force acting in the horizontal direction. If, however, a roll or pitch motion takes place an acceleration component occurs in the nearly horizontal plane of the geophone sensitive axis. For small angles of rotation A of the geophone sensitive axis, the sine of the angle A is nearly equal to the angle A itself or A=sinA. Thus the output of the geophone is dependent on the velocity motion sensed in its own plane plus some signals generated by gravity when rotations of a horizontal plane occur. This coupling effect causes stability problems for the horizontal channels because control motions of a vertical channel couple to horizontal sensors, when pitch or roll occurs, and causes false signals. Thus, an interaction between vertical and horizontal channels exists due only to gravity, but which nevertheless must be dealt with. The solution chosen was to limit the low frequency response of the horizontal channels 4, 5 and 6 by the use of higher frequency DC blocking lead networks, 43,44 of fig. 3, having a higher resonant frequency, as well as a lower overall loop gain, as compared to the vertical channels, as will be understood by those skilled in the art. Referring now to Figs. 5 and 6 the plan and elevation views of an embodiment of the invention with six degrees of freedom is shown. The payload is a table work surface 61 supported in a frame 62 by four passive but height regulated vibration isolation elements 63 located generally near the corners of the work table 61. Six velocity sensing geophones 64, 65, 66, 67, 68 and 69 are mounted on the table 61 with their axes of motion sensing aligned to sense the individual components of the six degrees of freedom. Mounted between the frame 62 and the table 61 are a plurality of force actuators of the type described with reference to Fig. 9. Thus as shown in Fig. 5 there are two actuators 71 and 72 positioned to control the short axis Y translation and yaw. There are four vertical force actuators 73, 74, 75 and 76 for controlling vertical motion and pitch and roll and there is an actuator 77 positioned to control long axis Y translation. It will be noted that there are seven force actuators in Figs, 5 and 6, one more than that described with reference to Fig. 4. This merely represents allocating the signal response from the third channel motion between the two actuators 73 and 75 instead of the pivot motion indicated in Fig.4. Thus actuators 73 and 75 operate jointly on the same signal which is modified relative to the signals described with the circuit of Fig. 5 in a manner which will be apparent to those skilled in the art. Referring to Fig. 7 the transmissivity curve of the embodiment of the invention shown in Fig. 6 as measured is shown in curve 81 with a comparison curve 82 representing the transmissivity curve of the passive isolation system alone. As will be noted the curve 82 exhibits a natural resonance at approximately three Hz where the transmitted motion is actually amplified by a factor of almost 10. It is this phenomenon universal with passive systems which the present invention so effectively avoids. As seen in the curve 81, the resonance of the system has been reduced in frequency by approximately a factor of 10 occuring somewhere near 0.4 Hz and its amplitude is greatly reduced being approximately 1.1. Thereafter the attenuation provided by the active force velocity feedback system of the invention reduces the percent transmission well below 1.0 and in fact it is approximately 0.1 at the frequency of resonance of the passive system alone. The curve 81 represents the response with the passive and active systems both present and hence is the actual result achieved by the invention. As noted curve, 81 continues to reduce the percent transmission until it merges with the isolation provided by the passive system alone as indicated in curve 82 somewhere in the region of 30 Hz. Referring to Fig. 8 a typical example of the results achieved with the embodiment of the invention described in Fig. 6 is shown. The top curve 83 shown in Fig. 8 represents the response of the combined passive and active system of the invention to a 1.7 cm-Kg (1.5 inch-pound) impact load such as drop of a rubber ball on a 22.5 Kg (50 pound) payload of Fig. 5. The curve 84 shown in Fig. 8 represents the reponse of the passive system alone to the same impact loading. Note that the initial deflection is approximately the same for curves 83 and 84 but that the recovery time with the active system operating as indicated in curve 83 is much shorter and of substantially lower amplitude than the oscillatory deflection shown in curve 84. As can be seen from curve 83, a vibration isolation system made in accordance with the present invention exhibits a fast settling time resulting from the large damping forces generated by the feedback. The increased system damping not only results in superior vibration isolation compared with typical passive systems, but also greatly reduces the system payload response motions resulting from external forces applied to the payload. Whenever a payload is of such character that it requires a force of more than approximately 0.9 Kg (2 pounds) to counteract externally applied forces, the use of electro-magnetic force transducers such as described with rference to Fig. 9 become expensive and ultimately impractical. Other types of force actuation means can be substituted without departing from the scope of invention. In Fig. 10 a pneumatic system is shown wherein a payload 101 is supported from a piston 102 which is freely suspended from diaphragms 103 in an upper and lower pressure cylinders 104, 105. The pressure within the cylinders 104, 105 operates on the equal upper and lower area surfaces of the piston 102 to apply the motion of the piston 102 to the payload 101. This pressure is controlled by a flow control servovalve 106 which applies differential pressure to the piston 102 to move same. The flow control servovalve 106 is controlled by a servo system comprising a force command signal input at line 107 corresponding to the final velocity sensor signal which is compared in a differential summing device 108 with a signal representing the differential pressure in the cylinders 104 and 105 obtained from a differential pressure transducer 109. Any deviation in position called for by the velocity signal on line 107 relative to the position established by the differential pressure sensed by unit 109 results in an error signal output which is applied to a transfer function device 111 and pneumatic amplifier 112 to control the servovalve 106. The system of Fig. 10 thus applies force from the piston 102 to the payload 101 in accordance with the requirements of the velocity signal on line 107 to achieve the results described in the present invention where higher values of force are required.

Claims: A vibration isolation system comprising: a payload mass (11); a support (12); a passive isolation system (13,14) having a certain resonance frequency and supporting said payload mass (11) on said support (12); motion-responsive means (21) having a resonance frequency for sensing vibrational movement of said payload mass (11) and for producing an output signal from which it is possible to derive a value for the velocity of said payload mass (11) as it undergoes vibrational motion, control circuit network means (22) responsive to said output signal for modifying said output signal; a force transducer (23) for applying a force to said payload mass (11) in response to an input signal; and signal applying means for applying said modified output signal as an input signal to said force transducer (23) so as to cause said force transducer (23) to apply a force to said payload mass (11) in a direction opposite to the direction of movement of said payload mass (11) as it vibrates; characterized in that said control network means (22) has a transfer function which is selected so that the effective resonance frequency of said motion-responsive means (21) is substantially lower than the resonant frequency of said passive isolation system (13,14). A vibration isolation system according to claim 1 characterized in that said motion-responsive means (21) comprises a velocity transducer. A vibration isolation system according to claim 1 or 2 characterized in that the effective resonance frequency of said motion-responsive means (21) is lower than the resonant frequency of said passive isolation system (13,14) by at least one decade. A vibration isolation system according to any of claims 1 to 3 characterized in that said control network means (22) and said signal applying means form a feedback loop and have a transfer function with a feedback gain of substantially zero at DC. A vibration isolation system according to any preceding claim characterized in that said motion-responsive means (21) is a geophone. A vibration isolation system according to any of claims 1 to 4 characterized in that said motion responsive means (21) is a geophone and said signal-applying means is adapted to apply said geophone output signal as an input signal to said force transducer (23).

Inventors: SCHUBERT DALE W, SCHUBERT, DALE W.
Applicants: WRIGHT BARRY CORP, BARRY WRIGHT CORPORATION
Application year: 1985
Main classification: F16F15
IPC codes: F16F15, F16F9, G01V1, G05D19