JPS6030838A - Vibration damper - Google Patents

Abstract

PURPOSE:To effect reliable damping of vibration during resonance and to perform reliable absorption of vibration outside a resonance frequency range, by a method wherein a force, which is exerted in the reverse direction to that of a vibration and has amplitude approximately proportional to the velocity of vibration, is exerted on an elastic retainer part only within a resonance frequency range. CONSTITUTION:An actuator part, which serves to generate a force across an elastic retainer part 11 for retaining a vibration body, is formed with a chamber 14, a line 22, a servo valve 23, and hydraulic source 24. Only components, in a specified frequency range, of signals from a speed detector 25 and a pressure detector 26, are selected by low pass filters 27 and 28 and sent to a controller part 31, and the servo valve 23 is controlled to actuate the actuator part. A force, which is generated in the reverse direction to that of a vibration and has amplitude approximately proportional to the velocity of vibration, is exerted on the elastic retainer part within the specified (resonance) frequency range to increase a dynamic spring constant. Vibration is absorbed, outside the specified (resonance) frequency range, by means of the constant of the elastic retainer part 11.

Description

[Detailed description of the invention]

The present invention relates to a vibration damping device used in an automobile engine mount or the like. Conventionally, there is a vibration damping device as shown in FIG. 5, and this vibration damping device is equipped with a tip-shaped rubber member 1, the opening of which is closed by a metal member 3 to which a bolt 2 is connected, and a bottom portion of the damping device is closed. A metal member 5 to which a bolt 4 is connected is fixed. A diaphragm 6 and a partition plate 7 are interposed between the rubber member 1 and the metal member 8, and an orifice 7a is formed in the partition plate 7. A chamber 8 between the metal member 3 and the diaphragm 6 is a sealed air chamber, and a chamber 9 is located between the bottom of the rubber member 1 and the diaphragm 6 and partitioned by a partition plate 7. l. is filled with liquid. Such a vibration damping device is put into practical use by attaching the opening of the rubber member 1 to a vibrating body (for example, an engine) with bolts 2, and the bottom of the rubber member 1 to a support body (for example, a vehicle body frame) with bolts 4. It functions like this. That is, the vibration of the vibrating body elastically deforms the rubber member 1, and the sealed liquid flows through the orifice 7atl-10 between the chambers 9 and 10 whose volumes change.
After that, they come and go under restrictions. The flow resistance at this time damps vibrations, and a predetermined vibration damping effect is obtained. In addition, during this action, chamber 9
．． 10 total volume (change in chamber 8 via diaphragm 6
The air inside is absorbed by being compressed or expanded and does not interfere with the above action. The operating characteristics of the above vibration damping device are as follows: loss factor (static spring constant = 1 damping coefficient = 0) for the input vibration frequency
If the one-angle gradual number is ω, then the change characteristic of tan y is shown as the solid line a in FIG. 6. By the way, the vibration damping device is used for the purpose of quickly damping the free vibration of the vibrating body when it starts to vibrate. The dynamic spring constant kd (kd = k" + (expressed as ccc+-1) is made thicker in the middle f. and below), and the dynamic spring constant k is
Ideally, an example in which cl becomes almost zero would have operating characteristics as shown by the dashed dotted line in FIG. However, in the conventional vibration damping device, if the dynamic spring constant k (1) in the resonance frequency range is made large as required above, the 6th
As is clear from the characteristic a in the figure, the dynamic spring constant kd outside the frequency range of the image also increases, and vibrations from the vibrating body are directly transmitted to the support body. When used, engine vibrations outside the resonant frequency range cannot be absorbed completely, causing vibrations in the vehicle body and increasing noise inside the vehicle. There are also vibration damping devices that change the rigidity of the damping device according to engine rotation, such as the vibration damping device shown in Utility Model Application Publication No. 57-422, but such damping devices only provide vibration damping primarily. Because the system only changed the rigidity of the device, it was not possible to always prevent noise from occurring inside the vehicle. The present invention provides an actuator section that generates a force between both ends of an elastic support section that supports a vibrating body on a support body, and operates this actuator section in a specified (resonant) frequency range to produce vibrations in a direction opposite to the vibration direction and at a vibration speed. A force that is approximately proportional to the elastic support is applied to the elastic support, and the force is not applied to the elastic support in other frequency ranges, so that the vibration is absorbed by its static spring constant. For example, it is possible to solve the above-mentioned problem that vibrations from the vibrating body are directly transmitted to the support body outside the resonance frequency range, while increasing the dynamic spring constant as required in the resonance frequency range. From this point of view, we are attempting to provide a vibration damping device that embodies this idea. Hereinafter, the present invention will be explained in detail using solid line examples shown in the drawings. FIG. 1 shows an embodiment of the device rj of the present invention, and FIG. 2 shows its conceptual diagram. In these figures, reference numeral 11 denotes an elastic support portion for attaching the vibrating body to the support body. The elastic support part 11 includes a dish-shaped elastic body 12 made of oil-resistant rubber or the like, and its opening is closed by a metal part 'JfA13 to define a chamber 14.

[Undertone of the sexual body 12] Fix the metal part 15 to 5. A frame body 16 that limits deformation of the elastic body 12 in the frame direction is supported by the metal part 15 via an elastic stay 17. Metal department profit 13
．． By inserting bolts 18 and 1.9 into each of the ends of the elastic support section 11 and attaching the metal parts vJ' 13 and 15 to the vibrating body side bracket 20 and the support body side bracket 2 respectively through these bolts, the elastic support part 11 can be attached to the vibrating body. The chamber 14 for supporting the elastic support part 1 is connected via a conduit 22 to a servo valve 23, which in turn is connected to a hydraulic pressure #24.
1 constitutes an actuator section that generates force. The servo valve 23 supplies the pressure of the hydraulic pressure source 24 into the chamber 14 through the pipe line 22 to increase the internal pressure when the servo valve 23 is open, and returns the pressure inside the chamber 14 to the hydraulic source 24 when the servo valve 23 is closed. It lowers the internal pressure of the valve, and the opening and closing are controlled by the following electronic mouth traces. That is, the frame body 16 is provided with a speed detector 25 for detecting the relative movement speed of both ends of the elastic support section 11, and the chamber 14. A pressure detector 26 for detecting the internal pressure is provided inside the interior, and signals from these detectors 25.2G are passed through a low-pass filter 27, respectively.
．． 28 and amplifiers 29 and 30, and then input to the control section 81. The control section 3 controls the operation of the servo valve 23 as described later based on the calculation results of these input signals. Note that the low-pass filters 27 and 28 are used for detection 5z52. z6
Specified frequency range component of the signal from
. Only the following frequency range components are selected and passed.Next, the principle of the device of the present invention will be explained. The vibrating body Kasuke 1 is m1 elastic body 1
Intersect the spring constant of 2 with the speed at which the relative movement h1 at both ends is detected by the X1 detector 25 (intersection = dX/at)
, the proportionality constant of the force proportional to this speed is C, then the interlocking culm equation for a vibration system with l degrees of freedom is m52 + cx + kx
= 6, and by multiplying this equation by the intersection, integrating it over time, and looking at the energy balance, we get 3. imx2+3/, kx2=
The formula O-cf woman'-dt (where 0 is a constant) is obtained.If the second term on the right side of this formula, that is, C, is always positive, the energy exchanged between the vibrating body and the elastic body 12 It can be seen that a predetermined vibration damping effect can be obtained by gradually decreasing the above-mentioned velocity. Therefore, in this example, the pressure fluctuation ΔP is generated in the chamber 14 by switching the servo valve 2a to generate the force. The formula is expressed as %mM+ΔP・A+1 (X=0, where A is the effective pressure-receiving area, but since the pressure-receiving area A is approximately constant, there is a candy in the chamber 14 with an opposite phase to the speed and proportional to its size. Pressure fluctuation Δ
The control section 31 controls the operation of the servo valve 23 so as to cause P to satisfy the above principle. Specifically, as is clear from FIG. 2, the control unit 81 converts the phase and magnitude of the speed detected by the speed detector 25 into the above-mentioned speed and the pressure within the chamber 14 detected by the pressure sensor 26. The servo valve 23 is controlled so that the ratio is compared with the opposite phase component and the ratio is constant. This control is carried out by the low-pass filters 27 and 28 at f in FIG. 6, respectively. Since only the following frequency range components are passed, this I? , l wavenumber range only, and the above-mentioned control is not performed in other frequency ranges, and the required characteristics shown in FIG. 6 can be obtained. Therefore, when the control section 31 is constituted by a microcomputer, it executes the program shown in FIG. 3 to perform the above-mentioned control. That is, first, in step 40, the detector 2
It is determined whether the speed detected by 5 and the pressure fluctuation detected by the detector 26 are larger than a set value. If so, that is, if the vibration is large enough to be damped,
Control proceeds to step 41, otherwise no further control is executed. In step 41, the phase difference ψ between velocity and pressure fluctuations is measured, and in the next step 42, the phase difference ?
Based on this, the fluctuating pressure xtanP is calculated and the velocity component that caused the pressure fluctuation is found. Next, the control proceeds to step 413, where the above-mentioned variable pressure xtan (the calculated value of p is divided by the speed p
After that, the divided value is the target value, for example S in Figure 6. Determine whether it is greater than, less than, or equal to. Target value S. If it is larger, that is, if the loss factor is too large, the control proceeds to step 44 where t(
Output a signal. 4. The internal pressure will not decrease,
Loss factor to target value S. can be approached. The above division value is the target value S. If it is smaller, that is, the loss factor is too small, ff++ control is performed in step 4.
Proceeding to step 5, the servo valve 23 outputs a signal such as 17Li<. The pressure inside Shisou 14 is increased, and the loss factor is set to the target value S. In this case, the loss factor is the target. Therefore, the servo valve 23 is left as it is. By repeating the control described above, the vibration damping device of the present invention can set the loss factor to the target value S in the specified frequency range (f. or less) as shown by the dashed line in FIG. It is possible to obtain ideal operating characteristics that keep the loss factor at zero in other frequency ranges. By the way, to describe the case where the device of the present invention is used for an engine mount, the resonance wave number f. is 6H2% amplitude is 5
The elastic support part 11 has almost the same dimensions as the conventional device shown in Fig. 5, and the pressure receiving area A is 46 cm! (diameter approximately 40 mm), static spring constant approximately 13 to 9 f/,,
. It was confirmed that when the loss factor So is set to 0.3, the pressure fluctuation ΔP can be as small as about 0-5 kgf/cm2. In the above embodiment, hydraulic pressure was used to apply force between both ends of the elastic support part 11, but as shown in FIG. An electromagnet 33 is also connected between these magnets so as to generate the above-mentioned attractive force or repulsive force with a value proportional to the magnitude of the magnetic force and the phase opposite to that of the electromagnet according to the signal from the control unit 3 or the speed detector 25. 1. By controlling, the same effect as in the example described in 611 can be obtained. Thus, as described above, the vibration damping device of the present invention is provided with the actuator sections 14, 22 to 24 (32, 33) that generate force between both ends of the elastic support section 11 that supports the vibrating body, and the actuator sections 14, 22 to 24 (32, 33) generate force in a specified frequency range. is actuated to apply a force opposite to the vibration direction and approximately proportional to the vibration speed to the elastic support part 11, and
Since the force is not applied in the l-wavenumber range and vibrations are absorbed by the static spring constant of the elastic support 11, vibrations from the vibrating body are reliably absorbed outside the specified frequency range while maintaining the ζ-1 constant. The dynamic spring constant can be increased as required in the frequency range to ensure the vibration damping effect.

[Brief explanation of the drawing]

Fig. 1 is a system diagram showing an embodiment of the present invention, Fig. 2 is a conceptual diagram thereof, Fig. 8 is a flowchart of a control program executed by the control unit of the embodiment, and Fig. 4 is a system diagram showing an embodiment of the present invention. FIG. 5 is a sectional view showing a conventional vibration damping device. FIG. 6 is a diagram showing a loss factor change characteristic of the device of the present invention in comparison with that of a conventional device. 11...-Elastic support part 12... Elastic body 13, 15
... Metal member 14 ... Chamber 16 ... Frame 17.
・・Elastic stay 18.19・Mounting bolt 20・・
- Vibrating body side bracket 21...Support side bracket 22...Pipe line 23...Servo valve 24...Hydraulic pressure source 25...M degree detector 26...I
E, l detector 27, 28...Low pass filter 29
．． 30...Amplifier 81...Control unit 32...Permanent magnet 83...Electromagnet Fig. 4

Claims (1)

[Claims] 1. In a vibrating body support structure that supports a vibrating body on a support body via an elastic support part, a detector is attached to the vibrating body and the elastic support part to detect the relative movement speed and direction of movement between both ends of the elastic support part; a filter that selects and outputs only a specified frequency range component from the signal from the detector; an actuator unit that generates a force between both ends of the elastic support unit;
The actuator section is configured to receive the output from the filter and generate a force in the opposite direction to the moving direction with a magnitude substantially proportional to the relative moving speed when the relative moving speed in the specified frequency range is equal to or higher than a set value. A vibration damping device comprising: a control unit that operates a vibration damping device;