GB2189573A - Vibration absorber - Google Patents

Abstract

A vibration suppression device for machinery 140 comprises a vibration absorber having means to adjust the stiffness; means to monitor the running speed of the machinery; means to monitor the absorber stiffness; and control means to adjust the stiffness to provide the maximum vibration suppression at the running speed of the machinery. The vibration absorber comprises a mass 130 supported between a pair of gas bellows 131, 132. A gas control circuit 146 simultaneously connects the pneumatic bellows to a pressurised gas supply 148, with a pressure transducer 145 provided at one of the bellows to monitor the absorber stiffness. An open loop control system is provided to regulate the pressure within the bellows, by positioning solenoid operated control valves 149, and 151-153 in the gas control circuit 146, in response to alterations in rotational speed of the machine to which the device is attached. A digital controller 147 is provided to vary the gas pressure simultaneously within both of the pneumatic springs in response to a detected alteration in said rotational machine speed. <IMAGE>

Description

SPECIFICATION A self tuning vibration absorber The invention relates to a self-tuning vibration absorber and in particular, though not exclusively, to an anti-vibration device for rotating machinery.

Vibration monitoring may be defined as the technique of observing the vibrations which are inherent in, or caused by, machinery, engines, installations etc. The purpose of vibration monitoring is to protect these components from excessive vibration levels. In one particular aspect of the problems caused by vibration in Naval machinery large displacements in the mounts of machinery leads to a high radiated underwater noise signature. This provides a tactical advantage to any potential aggressor. One of the main causes of vibration of Naval equipment is due to the imbalance in rotating machinery which is directly speed related.

The conventional mass/spring dynamic vibration absorber offers good vibration reduction but only over a limited frequency range. The effective range of the absorber can be extended, at the expense of effectiveness, by the introduction of damping. Therefore if the disturbing force varies over a wide frequency range the ideal vibration absorber should possess the ability to provide substantial vibration reduction over an equally wide frequency range.

The object of the invention is to provide a vibration absorber which can be automatically tuned to the excitation frequency.

The invention provides: a vibration suppression device for machinery comprising: a) a vibration absorber having means to adjust the stiffness; b) means to monitor the running speed of the machinery; c) means to monitor the absorber stiffness: and d) control means responsive to b) and c) to adjust the stiffness to provide the maximum vibration suppression at the running speed of the machinery.

In an advantageous arrangement the vibration absorber comprises a mass supported by gas bellows, there being provided means to vary the pressure of the gas to thereby vary the stiffness and pressure monitoring means to thereby monitor the absorber stiffness.

In a preferred arrangement the mass is supported between similar bellows units. In this arrangement for use with rotational machinery the device comprises: a pair of pneumatic springs spaced apart from one another by a mass free to move along the axis of the spring elements, the device having a natural frequency governed by the spring rates of the said pneumatic springs and the magnitude of the mass; a gas pipeline circuit connected to said pneumatic springs and connected to a pressurised gas supply supplying both pneumatic springs simultaneously; and an open loop control system to regulate the pressure within the said pneumatic springs, by positioning solenoid operated control valves in the said gas pipeline circuit, in response to alterations in rotational speed of the machine to which the device is attached.

Advantageously there is also provided 9 digital controller which varies the gas pressure simultaneously within both of the said pneumatic springs in response to a detected alteration in said rotational machine speed. Thus the natural frequency of the suppression device may be maintained automatically, requiring no manual input, at a value equal to the running frequency of the machine.

Fine gas pressure control of the pressure within the pneumatic springs may be effected by pulsing the solenoid operated control valves in response to an output signal from the digital controller.

The vibration suppression device may be attached to the machinery via sliding runners such that the position of the device on the machinery can be varied.

The invention will now be described by way of example only with reference to the attached Drawings of which: Figure 1 illustrates a vibrating system with one degree of freedom; Figure 2 shows the frequency response curve for the Fig. 1 arrangement; Figure 3 illustrates the addition of an absorber to the Fig. 1 arrangement; Figure 4 shows the response curve of the Fig. 3 arrangement; Figure 5 shows the effect of including a damper in a fixed stiffness or passive absorber; Figure 6 shows the effect of including an active spring in the absorber; Figure 7 is a schematic representation of a damper with variable stiffness as used in the present invention; Figure 8 shows a more detailed view of the variable damper of Fig. 7; Figure 9 is a graph of the variation of the spring stiffness k of the bellows with pressure;; Figure 10 illustrates an effect of varying the ratio of mass of absorber to mass of machinery; Figure ii shows vibration test measurements both with and without the absorber according to the invention over a speed range of 900 rpm; GB 2 189 573A Figure 12 shows a control system for controlling the vibration absorber of Figs. 7 and 8; Figure 13 shows the vibration suppression device of Figs. 7,8 in greater detail; and Figure 14 is a block diagram illustrating a control system for the Fig. 12 device.

The theory of the dynamic vibration absorber in all its various forms is well known (for example Omondroyd & Hartog "The Theory of the Dynamic Vibration Absorber" Transactions of the ASME, Vol 50, 1928, APM-50-7) and has been used to great effect to decrease or eliminate unwanted vibration in synchronous machinery.

Consider a single degree of freedom system subjected to a rectilinear exciting force as shown in Fig. 1. As the frequency of excitation w, increases up to and beyond the natural frequency of the system, the response will be as shown in Fig. 2, where x, is the displacement of the mass m,. The large response in the region of the natural frequency 21 may be reduced by the addition of a second spring/mass system, 31, normally termed the absorber as shown in Fig. 3.

The effect of adding the absorber 31 is clearly to add another degree of freedom to the system, the response of which is shown in Fig. 4.

It can be shown that:

If we define: w12=k,/m, =natural circular frequency2 of the main system alone and: w22=k2/m2 =natural frequency2 of the absorber alone A= m2/m, mass ratio k2/k1 =LW22/W12 then (1) and (2) become:

It is evident from Equation (3) that when w=w2 the motion of the main mass ceases entirely, ie when the absorber is tuned to the excitation or running frequency of the main system, the displacement of the main mass is zero. This further means that the transmissibility of force to the ground is zero. However, it is only when w=w2 that it is effective and hence it is only applicable to constant speed machinery.The frequency range of operation of the passive absorber may be increased at the expense of the main system displacement, and hence transmissibility, by including a damper in the absorber system. Fig. 5 shows the response of both damped and undamped systems.

The objections referred to above may be overcome by making the absorber spring an active element such that: W22= k2(variable)m2 = kv/m2 (5) Again it is evident from Equation (3) that for zero damping if w2 is tuned to the running speed (w) then x1, and hence transmissibility, will be zero over a range of w determined only by the range of achievable variance of k, as shown in Fig. 6. Fig. 7 shows a schematic representation in which pneumatic bellows 71 are used to vary the absorber's spring stiffness (kv). The mass m2 is supported between the bellows 71 on the side of the mounting frame 72 which is attached near to the opposed to the machinery mass m,.

Following the nomenclature shown in Fig. 8, which shows a more detailed view of the antivibration mount, Cavanaugh shows ("Air Suspension and Servo-controlled Isolation Systems" Shock and Vibration Handbook Ch 33 Harris & Crede) that for a double acting pneumatic spring:

where c4 is the capillary flow resistance of the supply pipework and n is the ratio of the specific heats of the gas in the bellows. When operating under adiabatic compression and expansion Cavanaugh shows that Equation (6) may be simplified to:

Sainsbury ("Air Suspension for Road Vehicles" Proc Inst Mech Eng 1957/58 Pt 3) indicates a method of evaluating the effective area of double convolution bellows (as used in the present absorber) based on the concept of an effective diameter.Further, if the displacement x2 is small in relation to the full stroke of the bellows, then the term dA/dx2=O, thereby simplifying Equation (7).

To ensure that the motion of the main system is brought to a minimum when the absorber is tuned to the excitation frequency the damping should be minimised. A theoretical evaluation of the damping is given in Bacharach & Riven "Analysis of a Damped Pneumatic Spring" Journal of Sound and Vibration 1983 Vol 86 No 2, which shows:

This, in conjunction with Equation (7), demonstrates the importance of the relationship between the volumes of working (Vb) and non-working air (Vp). Hirtreiter "Air Springs" Machine Design" April 1965 justifies the use of the r=0 viscously damped transmissivity curve for air springs.Therefore by mounting the absorber on a shaker and providing base excitation the damping of the absorber could be evaluated from:

which when w=wn yields:

This concept is employed to optimise the design of the absorber by conducting a series of trials to determine the configuration which gives the lowest damping/highest transmissivity (x2/x,) over the widest frequency range. The working/non-working air volume ratio (Vb/vp) was chosen to optimise the damping (Equ (8)) accepting the resulting change in the absorber's natural frequency (Equ (7)).

When the optimum configuration had been determined, the variation of maximum transmissivity (x2/x1) with frequency was obtained, again by base excitation, and this gave the intended absorber operating characteristic. Thus the absorber would always operate at the maximum transmissivity as the forcing frequency varied, thereby minimising x, and the force transmission to the ground over a wide frequency range. The absorber's operating characteristic is shown in Fig. 9 which shows the achievable variance in kv.

To verify the results obtained from these tests, a variable speed electric motor was mounted on a beam whose length was chosen to provide the first natural frequency within the operating range of the absorber. Initial trials were conducted to verify the operating characteristic (Fig. 9), and then to determine the effectiveness of the absorber system. As expected from Figs. 2 and 4 the addition of the absorber introduces a resonant frequency either side of the running speed (w) and the separation of these two frequencies was found to be dependent on the mass ratio Cu=m,lm,) as shown in Fig. 10.

The effectiveness of the absorber can clearly be seen in Fig. 11, which shows the response x, of the main system (beam and motor) both with (111) and without (112) the absorber over a speed range of 900 rpm. Without the absorber the main system exhibited a torsional mode of vibration at 1230 rpm and the first natural frequency was easily distinguishable at 1890 rpm.

The effect of the absorber was to reduce the main system displacement by about 95% over its operating range.

Fig. 12 shows the Fig. 7/8 arrangement provided with a control system to automatically tune the absorber to provide the maximum vibration reduction at any frequency within the operating range. Within this range the absorber is tuned by varying the entrapped air pressure in the bellows 121. The control system 122 is open loop employing a digital controller 123 receiving tacho and pressure signals at respective inputs 124 and 125 to compute the desired bellows air pressure for optimum vibration reduction. The controller 123 then activates the control unit 126 to pressurise or vent the bellows accordingly.

Figs. 13 and 14 show in detail the vibration damper and pneumatic control system. The absorber mass 130 is attached by bolts between the adjacent ends of colinear pneumatic bellows 131,132. The opposed ends of the bellows are attached to fixed respective containment plates 133,134 connected together by four containment struts 135. The assembly is connected by bolts 135 and spacers 137 to an attachment plate 138. The attachment plate 138 has holes 139 provided such that it can be secured to the surface 140 of a structure being vibrated by operating machinery. Rails (not shown) may be provided on the surface 140 such that the position of the damper can be adjusted. Each pneumatic bellows is provided with an end connector 141,142 for connection to respective gas pipes 143 and 144. The connector 141 is also provided with a threaded portion for connecting a pressure transducer 145.The stiffness of the pneumatic springs 131 and 132 is adjusted by varying the pressure of gas connected thereto via the output pipes 143 and 144 from a gas control unit 146. Pressure in the pneumatic springs is monitored by a digital controller 147 which receives signals from the pressure transducer 145.

The gas control unit 146 has an input pipe 148 connected to a regulated gas supply (not shown). The input pipe 148 is connected to an input solenoid operated valve 149, operation of which allows the system to be vented to air. The input valve 149 is connected via a manual restrictor valve 150 and a second solenoid operated vale 151 to a pair of output solenoid operated valves 152,153 which control gas to the output pipes 143,144 to the respective pneumatic springs 131 and 132. Each of the solenoid operated valves 149, 151-153 is controlled by the digital controller 147. In addition to monitoring the gas pressure in the pneumatic springs the digital controller also receives a signal via line 154 representative of the excitation of the surface 140 of the structure to be damped.

The pressure transducer 145 is arranged to provide digital pressure readings for the controller 147. The controller is arranged to operate such that the valves controlling the gas pressure in the pneumatic springs are simultaneously altered in response to changes in the speed (e.g.

machine rotational speed) of the machinery attached to the structure 140, the spring pressure being monitored by the transducer 145. The spring pressure in the two pneumatic springs is thus simultaneously increased by admitting more air or reduced by venting to air through valve 149. The pressure transducer is connected to only one of the pneumatic springs and to combat minor depressurisation of the spring not connected thereto the solenoid operated valves may be arranged so as to connect together the springs during intervals determined by the digital controller.

The damping system thus employs an open loop control system requiring a speed input from the machine to be damped to vary the spring pressure. The change in spring pressure is such as to adjust the natural frequency of the suppression device to the running frequency of the machine. Minor depressurisation of the pneumatic springs is overcome by pulse operating the two valves 152 and 153 so as to interconnect the two springs at intervals determined by the controller 147. Fine pressure control of the gas pressure within the pneumatic springs is also provided by pulsing the appropriate valves 149 and 151-153 again under the direction of the controller 147.

The present invention employs an open loop control system based on the machine rpm to yield spring pressure; the pressure monitor signal from the transducer 145 is employed to determine the solenoid valve pulse length only and does not affect the open nature of the control circuit.

Claims (10)

1. A vibration suppression device for machinery comprising: a) a vibration absorber having means to adjust the stiffness; b) means to monitor the running speed of the machinery; c) means to monitor the absorber stiffness; and d) control means responsive to b) and c) to adjust the stiffness to provide the maximum vibration suppression at the running speed of the machinery.

2. A device as claimed in claim 1 wherein thereis included a mass supported by a gas bellows, there being provided means to vary the pressure of the gas to thereby vary the stiffness and pressure monitoring means to thereby monitor the absorber stiffness.

3. A device as claimed in claim 2 wherein the mass is supported between similar bellows units.

4. A device as claimed in claim 3 wherein the device comprises: a) a pair of pneumatic springs spaced apart from one another by a mass free to move along the axis of the spring elements, the device having a natural frequency governed by the spring rates of the said pneumatic springs and the magnitude of the mass; a gas pipeline circuit connected to said pneumatic springs and donnected to a pressurised gas supply supplying both pneumatic springs simultaneously; and an open loop control system to regulate the pressure within the said pneumatic springs, by adjusting solenoid operated control valves in the said gas pipeline circuit, in response to alterations in rotational speed of the machine to which the device is attached.

5. A device as claimed in claim 4 wherein there is also provided a digital controller which varies the gas pressure simultaneously within both of the said pneumatic springs in response to a detected alteration in said rotational machine speed.

6. A device as claimed in claim 5 wherein fine gas pressure control of the pressure within the pneumatic springs is effected by pulsing the solenoid operated control valves in response to an output signal from the digital controller.

7. A device as claimed in any one preceding claim wherein there is provided means for attachment of the device to machinery via sliding runners such that the position of the device on the machinery can be varied.

8. A device as claimed in any one of claims 5 to 7 including means to monitor the rotational speed of the machinery and a means to transfer a signal input representative of the rotational speed into a form required by the digital controller.

9. A device as claimed in any one of claims 5 to 8 wherein the means to monitor the stiffness of the springs comprises a means for detecting the gas pressure entrapped within one of the pneumatic springs and a means to transform a signal input representative of the gas pressure measured thereby into a form required by the digital controller.

10. A device as claimed in claims 9 wherein there is provided a means to pulse two interconnecting solenoid operated valves within the gas pipeline circuit, the valves being situated so as to connect the two pneumatic springs together at intervals determined by the digital controller to thereby alleviate minor depressurisation of the spring not connected to the gas pressure detecting means.