CN102900804B - Vibration isolating unit for vehicle - Google Patents

Abstract

The invention provides a vibration isolating unit for vehicles. Driving power of an actuator can be reduced by the device. The vibration isolating unit for vehicles comprise a rod (11) with one end (12) thereof fixed on an engine (1) and the other end (13) fixed on the vehicle body; an inertia block (15) supported on the rod (11); an actuator (17) for driving the inertia block to move backward and forward in an axis direction of the rod; an electric control component (26) for driving the actuator; and an electric power storage component (26) for supplying electric power to the electric control component; wherein the actuator does not drive, the electric control component converts vibration of the inertia block into electric power and charge the electric power into the electric power storage component.

Description

Vibration isolation device for vehicle

Technical Field

The present invention relates to a vibration damping device for a vehicle for damping vibration transmitted from an engine as a vibration source to a vehicle body side.

Background

As a vibration damping device for suppressing vibration transmitted from an engine to a vehicle body side, there has been proposed one including: the rigid body resonance frequency of the torsion bar is set lower than the resonance frequency of the engine, and a force proportional to the speed of axial displacement of the torsion bar is generated in the actuator (patent document 1).

Patent document 1: japanese patent laid-open publication No. 2011-

However, in the conventional vibration damping device described above, since the driving power of the actuator is supplied from a battery mounted on the vehicle, there is a problem that the fuel consumption of the vehicle increases by an amount corresponding to the driving power.

Disclosure of Invention

The present invention has been made to solve the above-described problems, and an object of the present invention is to provide a vibration damping device for a vehicle, which can reduce driving power supplied from outside a torsion bar to an actuator.

The present invention solves the above problems by the following methods: when the actuator is not driven, the vibration of the inertia mass is converted into electric power by the actuator and stored in the power storage unit, and when the actuator is driven, the inertia mass is driven and controlled by the electric power stored in the power storage unit.

Since the vibration of the engine also vibrates the inertial mass, the vibration energy of the inertial mass can be converted into electric power by the actuator and stored while the vibration isolation function is stopped. Further, when the vibration damping function is required, the actuator can be driven by the stored electric power, and therefore, the driving electric power supplied from the outside of the torsion bar to the actuator can be reduced by only the stored amount.

Drawings

Fig. 1A is a front view showing an example in which the vibration damping device according to the embodiment of the present invention is applied to a vehicle engine.

Fig. 1B is a top view of fig. 1A.

Fig. 2 is an exploded perspective view of fig. 1A and 1B.

Fig. 3 is a sectional view showing the upper torsion bar of fig. 1B.

Fig. 4A is a perspective view illustrating the upper torsion bar of fig. 1B.

Fig. 4B is a perspective view of the upper torsion bar of fig. 4A viewed from the opposite side.

Fig. 5A is a four-view (front view, left view, right view, top view) showing the upper torsion bar of fig. 4A and 4B.

Fig. 5B is a perspective view and a front view showing a state in which the base plate is removed from the upper torsion bar of fig. 5A.

Fig. 5C is a perspective view and a front view showing the substrate of fig. 5A.

Fig. 6A is a plan view showing an example of mounting the upper torque rod of fig. 4A and 4B to the engine.

Fig. 6B is a plan view showing another example of attachment of the upper torque rod shown in fig. 4A and 4B to the engine.

Fig. 7 is a diagram for explaining a vibration state of the engine.

Fig. 8A is a graph showing vibration characteristics when the upper torsion bar is driven and when it is not driven.

Fig. 8B is a diagram showing an example of switching between the control area and the charging area of the vehicular vibration damping device according to the embodiment of the present invention.

Fig. 8C is a graph showing the vibration energy of the inertia mass of the upper torsion bar in the idle operation state and the power consumption of the driver in the control mode.

Fig. 9 is a frequency characteristic diagram of transmission force of a structure capable of obtaining a double vibration damping effect.

Fig. 10 is a graph showing an example of setting the bushing rigidity of the torsion beam.

Fig. 11 is a graph for explaining an example of calculation of the engine speed using an acceleration sensor of the torsion bar.

Fig. 12 is a graph for explaining another example of calculating the engine speed using the acceleration sensor of the torsion bar.

Detailed Description

First, a so-called Pendulum (Pendulum) engine to which the vibration damping device for a vehicle according to the embodiment of the present invention is applied will be described. As shown in fig. 1A and 1B, the support structure of the engine 1 using the pendulum type is: in a so-called transverse engine 1 in which a principal axis of inertia L of the engine 1 is arranged in parallel with a width direction of the vehicle (a direction orthogonal to a traveling direction, also referred to as a vehicle right-left direction), two support points P1, P2 that support the engine 1 are located near the principal axis of inertia L of the engine 1 in a plan view of fig. 1B and at positions on axially opposite sides from each other across a center of gravity G, and both the two support points P1, P2 are located above the vehicle with respect to the principal axis of inertia L in a side view of fig. 1A. As shown in fig. 2, the two support points P1 and P2 are formed by the engine mounts 3 and 4 on the left and right, respectively.

The support structure of the pendulum engine has the advantages that: the engine 1 is suspended and supported like a pendulum, and the operation of the engine center of gravity G swinging around a straight line connecting both support points P1 and P2 is suppressed by a rod-like member such as the torque rods 5 and 6 attached to the vehicle body, so that the same vibration damping effect as in the conventional art can be obtained with a small number of members. That is, in the engine 1 supported by the pendulum method, the engine 1 is inclined about the axis connecting the two 2 support points P1 and P2 by the rotational inertia force when the engine 1 is operating. In order to support the engine 1 while preventing this inclination, there are provided a 1 st torque rod (upper torque rod) 5 for coupling the substantially half portion of the engine 1 and the vehicle body-side member, and a 2 nd torque rod (lower torque rod) 6 for coupling the remaining lower half portion of the engine 1 and the vehicle body-side member. The upper torque rod 5 is connected to the engine 1 from the upper right of the vehicle, and the other lower torque rod 6 is connected to the engine 1 from the lower side of the vehicle, and the tilt of the pendulum engine 1 can be prevented by these two torque rods 5, 6.

The engine 1 is, for example, an inline 4-cylinder, V-6-cylinder engine with a two-stage balancer. In the inline 4-cylinder engine and the V-type 6-cylinder engine with the two-step balancer, the unbalanced inertial force is small in the basic order of the engine rotation (japanese: basic order), and therefore the reaction force mainly of the engine torque variation acts on the engine 1. Thus, the present inventors have arrived at the finding that: the input from the two torsion bars 5, 6 supporting the torque mainly causes noise and vibration in the vehicle at the basic order of engine rotation. It is also known that, mainly when the vehicle is accelerated, the noise in the vehicle, which is made up of high-order fundamental orders and reaches about 1000Hz, becomes a problem for the passengers.

As described above, the vehicular vibration damping device of the present example includes the two torsion bars 5 and 6. As shown in fig. 1B, the upper torsion bar 5 is mounted between the upper portion of the engine 1 and the vehicle body. In contrast, as shown in fig. 1A, 1B, and 2, the lower torque rod 6 is mounted between the lower portion of the engine 1 and the subframe 2. Since the basic structures of the upper torque rod 5 and the lower torque rod 6 in this example are the same, the structure of the upper torque rod 5 will be described, and the structure of the lower torque rod 6 will be referred to and omitted.

The upper torque rod 5 shown in fig. 2 and 3 is shown with the housing 20 shown in fig. 4A and the like removed for the purpose of explaining the internal structure thereof, but as shown in fig. 4A to 6B, the upper torque rod 5 actually has a structure including the housing 20 and the like. As shown in fig. 2 and 3, the upper torsion bar 5 includes: a rod 11 having a bush 12 at one end fixed to an upper portion of the engine 1 and a bush 13 at the other end fixed to a vehicle body; an inertia mass 15 supported by the rod 11; and a driver 17 that reciprocates the inertial mass 15 in the axial direction of the rod 11.

Fig. 3 is a main part sectional view of the upper torque rod 5, and a pair of bushes 12, 13 are fixed to both ends of a rod-like rod 11 by welding. The engine-side bush 12 is composed of a cylindrical outer cylinder 12a, a cylindrical inner cylinder 12b coaxial with the outer cylinder 12a, and an elastic body (sound insulating material) 12c connecting the outer cylinder 12a and the inner cylinder 12 b. The bush 12 is fixed to the engine 1 by a bolt (not shown) that penetrates the inner tube 12b in a direction perpendicular to the plane of the drawing of fig. 3.

On the other hand, the bush 13 fixed to the vehicle body side is also composed of a cylindrical outer cylinder 13a, a cylindrical inner cylinder 13b coaxial with the outer cylinder 13a, and an elastic body (sound insulating material) 13c connecting the outer cylinder 13a and the inner cylinder 13b, similarly to the bush 12 described above. The bush 13 is fixed to a member on the vehicle body side by a bolt (not shown) that penetrates the inner cylinder 13b in a direction orthogonal to the sheet surface of fig. 3.

In the embodiment shown in fig. 3, the bush 12 is fixed to the engine 1 and the bush 13 is fixed to the vehicle body side, but the present invention is not limited to this, and the bush 12 may be fixed to the vehicle body side and the bush 13 may be fixed to the engine 1. Further, while the upper torque rod 5 shown in fig. 3 is an example in which two bolts penetrating the inner cylinders 12B, 13B of the bushings 12, 13 are arranged in parallel, the upper torque rod 5 shown in fig. 2 and 4A to 6B is an example in which two bolts 18, 19 penetrating the inner cylinders 12B, 13B of the bushings 12, 13 are arranged in directions orthogonal to each other. The fixing direction of the bushes 12, 13 can be changed as appropriate according to the shapes of the fixing portion on the vehicle body side and the fixing portion of the engine 1.

The elastic bodies (sound insulating materials) 12c and 13c of the present example are members having both a spring function and a damping function, and for example, elastic rubber can be used.

In the upper torque rod 5 of the present example, the diameters of the outer cylinder and the inner cylinder of the bushings 12, 13 are made different. That is, the diameters of the outer tube 13a and the inner tube 13b of the bush 13 are made relatively smaller than the diameters of the outer tube 12a and the inner tube 12b of the corresponding bush 12, and the rigidity of the elastic body 13c of the bush 13 is also made relatively larger than the rigidity of the elastic body 12c of the bush 12. By making the rigidity of the elastic bodies 12c, 13c of the pair of bushes 12, 13 different, engine rigid resonance and rod rigid resonance in the rod axial direction suitable for double vibration isolation are generated at two different frequencies.

That is, as shown by the solid line in fig. 9, the engine rigid resonance a in the rod axial direction determined by the rigidity of the elastic body 12c of the bush 12 occurs at a frequency f1[ Hz ] substantially close to 0, and the rod rigid resonance B in the rod axial direction determined by the rigidity of the elastic body 13c of the bush 13 occurs at a frequency f2[ Hz ] close to 200 Hz. For ease of understanding, the description will be made based on a spring block system in which the engine stiffness resonance and the rod stiffness resonance are extremely simplified, and the engine stiffness resonance a is determined by the engine mass and the stiffness (spring constant) of the elastic body 12c of the bush 12, and the rod stiffness resonance B is determined by the mass between the elastic body 12c of the bush 12 and the elastic body 13c of the bush 13, that is, the mass of the rod 11 (and the outer cylinder portion of each bush) and the stiffness (spring constant) of the elastic body 13c of the bush 13.

The resonance frequency f3 of the engine 1 alone, which is bent and twisted 1 time, is about 280Hz to 350Hz in the case of a normal vehicle engine, and if the engine rigid resonance a is set to substantially zero and the rod rigid resonance B is set to about 200Hz as in this example, transmission of the bending and twisting resonance vibrations of the engine 1 to the vehicle body can be effectively suppressed on the high frequency side (in the vibration isolation region) (double vibration isolation).

From the above, in order to make the frequencies of the engine stiffness resonance a and the rod stiffness resonance B smaller than the resonance frequency f3 of the bending and torsion of the engine, the stiffness (spring constant) of the elastic body 12c of the bush 12, the mass of the rod 11 (and the outer cylindrical portion of each bush), which is the mass between the elastic body 12c of the bush 12 and the elastic body 13c of the bush 13, and the stiffness (spring constant) of the elastic body 13c of the bush 13 may be determined. Thus, there is a double vibration-damping effect in which the engine rigid resonance a and the rod rigid resonance B are caused to occur at two different frequencies, i.e., at two frequencies f1 in the low frequency range and f2 in the intermediate frequency range, thereby providing an effect of preventing vibration transmitted from the engine 1 to the vehicle body side. However, in the vibration damping device according to the present invention, it is not necessary to make the diameters of the outer cylinder and the inner cylinder of the bushings 12, 13 different, and the bushings 12, 13 may be configured to be the same.

Returning to fig. 3, the upper torsion bar 5 of the present example includes an inertial mass 15 made of a metal having magnetism or the like, an actuator 17, an acceleration sensor 21, a band-pass filter 22, and a voltage amplifier circuit 23.

The inertial mass 15 is arranged coaxially with the rod 11 around the rod 11. The cross section of the inertia mass 15 viewed in the axial direction of the rod 11 is a point-symmetric shape centered on the center (center of gravity) of the rod 11, and the center of gravity of the inertia mass 15 coincides with the center of the rod 11. As shown in fig. 2 and 5C, the inertia mass 15 is formed in a rectangular tube shape, and both ends (upper and lower ends in fig. 3) of the inertia mass 15 in the rod axis direction are connected to the rod 11 via elastic support springs 16, respectively. The elastic support spring 16 is, for example, a plate spring having a relatively small rigidity. A part of the inner wall 15a of the inertial mass 15 is provided to protrude toward a permanent magnet 17c of a driver 17 described later.

As shown in fig. 3, the upper torsion bar 5 of the present example is provided with an actuator 17 in a space between the inertia mass 15 and the rod 11. The actuator 17 is a linear (linear motion type) actuator including a square cylindrical iron core 17a, a coil 17b, and a permanent magnet 17c, and is used to reciprocate the inertial mass 15 in the axial direction of the rod 11.

The core 17a constituting the coil magnetic circuit is made of laminated steel plates and is fixed to the rod 11. The core 17a is divided into a plurality of members before the torque rod 5 is assembled, and the plurality of members are bonded around the rod-like rod 11 with an adhesive, thereby forming a square cylindrical core 17a as a whole. The coil 17b is wound around the square cylindrical iron core 17 a. The permanent magnet 17c is provided on the outer peripheral surface of the iron core 17 a.

Since the actuator 17 has such a configuration, the inertial mass 15 is driven by the reactive torque generated by the magnetic field generated by the coil 17b and the permanent magnet 17c, and the inertial mass 15 is linearly moved, that is, the inertial mass 15 is reciprocated in the axial direction of the rod 11. In contrast, when the vibration of the engine 1 is transmitted and the inertia mass 15 reciprocates in the axial direction of the rod 11, an alternating current is generated in the coil 17b by the electromagnetic induction action. That is, since the actuator 17 also functions as a generator, the actuator 17 is driven by the generated electric power in the vibration damping device for a vehicle of this example. The details of which will be described later.

An acceleration sensor 21 is attached between the bushes 12, 13 on a plane parallel to a horizontal plane passing through the axial center of the rod 11, and this acceleration sensor 21 is used to detect the axial vibration acceleration of the rod 11 at a substantially axial center position as the acceleration of the vibration transmitted from the engine 1 to the rod 11. Specifically, as shown in fig. 5C, the acceleration sensor 21 is mounted on the substrate 24, and the substrate 24 is mounted on the opening 20A of the housing 20. Then, a signal of the axial acceleration of the rod from the acceleration sensor 21 is input to the voltage amplification circuit 23 through the band-pass filter 22, and the signal amplified by the voltage amplification circuit 23 is applied to the coil 17b (control voltage) of the driver 17. The voltage amplifier circuit 23 can be formed of, for example, an operational amplifier. As shown in fig. 5C, the band-pass filter 22 and the voltage amplifier circuit 23 are also mounted on the substrate 24 attached to the opening 20A of the housing 20.

The inertial mass 15 is supported by a relatively flexible plate spring (elastic support spring 16), and the inertial mass 15 resonates in the axial direction of the lever with respect to the lever 11 at a relatively low frequency of, for example, 10Hz to 100 Hz. For example, since the two-step vibration frequency of the idle rotation speed of a 4-cylinder engine is about 20Hz, the resonance of the inertia mass 15 can be suppressed regardless of the operating conditions of the engine 1 as long as the resonance frequency of the inertia mass 15 can be set to 10 Hz.

On the other hand, if the resonant frequency of the inertial mass 15 is set to a low frequency such as 10Hz, if the inertial mass 15 is too large to make the setting difficult, the resonant frequencies of the inertial mass 15 are sufficiently apart by setting the resonant frequency to be lower than about 1/2 of the rod rigid resonance B (200 Hz in the embodiment) to be suppressed, and the transmission of vibration can be sufficiently suppressed.

Further, by passing the acceleration signal detected by the acceleration sensor 21 to the band-pass filter 22, control is not performed at an excessive frequency, control stability is improved, and it is possible to suppress excessive power consumption and reliably suppress transmission power in a target frequency range. As shown in fig. 9, the vibration isolation region for the rod rigid resonance B is obtained by multiplying the resonance frequency f2 of the rod rigid resonance B by a predetermined valueFor the frequency range of the frequency f5 or more obtained, a filter that passes a signal in a frequency range from the resonance frequency to the vibration isolation region where the rod is rigidly resonated B, that is, a filter that passes a signal up to the upper limit (for example, 400 Hz) of a range where control in the vibration isolation region does not diverge, is selected as the band-pass filter 22.

In order to perform speed feedback control for increasing the attenuation of the rod 11 to be controlled, the force obtained by adding a negative sign to a force substantially proportional to the rod axial speed of the vibration detected by the acceleration sensor 21 is generated from the driver 17 in the frequency band passed through the band-pass filter 22.

Next, the case 20 and the substrate 24 will be described.

As shown in fig. 4A to 6B, the housing 20 of the present example is formed of a rigid body fixed to or integrally formed with the outer cylinders 12a and 13a of the bushes 12 and 13, and the vibration of the rod 11 in the axial direction and the pitch direction is transmitted in an equivalent manner. An opening 20A is formed in the casing 20 at a position between the bushes 12 and 13, and the base plate 24 is attached so as to hermetically or watertightly seal the opening 20A. The inertial mass 15 and the actuator 17 shown in fig. 3 are housed in the case 20, and are protected from external water or the like by the substrate 24.

As shown in fig. 5C, the acceleration sensor 21, a control circuit 25 including the band-pass filter 22 and the voltage amplification circuit 23, and a secondary battery 26 including a power conversion circuit are mounted on the main surface of the substrate 24.

Among them, the acceleration sensor 21 is mounted on the base plate 24 so as to be positioned between the bushes 12 and 13, that is, on a plane parallel to a horizontal plane passing through the shaft center (the axis supporting the torque) of the rod 11. As shown in fig. 7, when the vibration generated by the unbalanced inertial force acting in the vertical direction is generated in a 4-cylinder engine or the like, and the acceleration sensor 21 is disposed at a position displaced upward with respect to the axial direction of the torque of the support rod 11, the vibration in the pitch direction is generated in the torsion bar due to the vertical vibration of the engine 1, but in this example, the sensitivity is lowered in the vibration in the pitch direction because the acceleration sensor 21 is disposed on a plane parallel to the horizontal plane passing through the torque support shaft. That is, the axial vibration detection accuracy is increased. As a result, as shown in fig. 10, even when the rigid resonance in the axial direction of the rod 11 is greatly reduced, since the noise of the rigid resonance in the pitch direction is hardly detected, the acceleration sensor 21 detects that the rigid resonance in the pitch direction is reduced to the normal region as in the conventional case, and thus, it is possible to suppress the problem of increasing the control power.

In particular, since the acceleration sensor 21 is disposed between the bushes 12, 13, the sensitivity in the pitch direction is made smaller by disposing the acceleration sensor 21 in the region where the node of the rigid resonance in the pitch direction of the rod 11 is present.

Then, when the upper torque rod 5 having the above-described structure is mounted between the engine 1 and the vehicle body and the engine is driven at an engine rotation speed of 2000rpm to 6000rpm, the vibration state in the vehicle longitudinal direction is observed with the driver 17 of the upper torque rod 5 being driven and controlled and with the driver 17 of the upper torque rod 5 not being driven, and the result is shown in fig. 8A. From this result, the following can be obtained: in the region where the engine speed is high and 3500rpm or more in the result shown in the figure, the vibration damping effect is large in the case where the actuator 17 of the upper torsion bar 5 is driven and controlled, as compared with the case where the actuator 17 of the upper torsion bar 5 is not driven and controlled, but in the region where the engine speed is low and 3500rpm or less in the result shown in the figure, the magnitude of vibration does not change much regardless of whether the actuator 17 of the upper torsion bar 5 is driven and controlled.

Therefore, in the vehicle vibration damping device of the present example, as shown in fig. 8B, in the operating state in which the engine speed is less than 3500rpm, the supply of electric power to the actuator 17 of the upper torsion bar 5 is not driven, the vibration energy of the inertia mass 15 is converted into ac power by the actuator 17, the ac power is converted into dc power by a power conversion circuit (an inverter circuit or the like) included in the secondary battery 26, and the dc power is charged into the secondary battery 26.

On the other hand, in the operating state where the engine speed is 3500rpm or more, the dc power charged in the secondary battery 26 is converted into ac power by the power conversion circuit, and then supplied to the actuator 17 of the upper torsion bar 5, so that the inertia mass 15 is controlled to vibrate as described above, thereby exhibiting the vibration damping function. Fig. 8C shows the result of measuring the vibration energy of the inertia mass 15 when the engine 1 is in the idling state (left histogram) and the result of measuring the power consumption when performing control for supplying electric power to the actuator 17 of the upper torque rod 5 to suppress resonance in the vehicle longitudinal direction (right histogram), both of which are substantially equal electric powers. Therefore, in the normal operation state, the self-sufficiency can be achieved only by the vibration energy of the inertia mass 15 without supplying power from the outside.

The engine speed as a threshold value for causing the actuator 17 to function as a generator or as a vibration isolator may be a speed at which a detection signal from a speed sensor provided in the engine 1 can be obtained. However, in the upper torsion beam 5 of the present example, since no wiring connected to the outside is required in addition, calculation can be performed inside the upper torsion beam 5 using the acceleration sensor 21. For example, an IC circuit for performing a fourier transform operation is provided on the substrate 24, and the IC circuit performs a fourier transform operation on the detection signal of the acceleration sensor 21 shown in the left diagram of fig. 11, and performs a fourier transform operation on the detection signal, thereby detecting a frequency at the maximum level as shown in the right diagram of the diagram. The engine speed can be obtained by multiplying the frequency by 60 and dividing the frequency by the basic order of rotation of the engine (if the engine is a 4-cylinder engine, the basic order is 2).

Alternatively, as shown in fig. 12, the magnitude of the signal obtained by applying band-pass filtering in the control region of 3500rpm to 6000rpm may be set as a threshold value for causing the actuator 17 to function as a generator or a vibration isolator to the detection signal of the acceleration sensor 21.

As described above, in the vehicle vibration damping device of the present example, the vibration energy of the inertia mass 15 is converted into electric power and charged to the secondary battery 26 when the actuator 17 is not driven, and the vibration damping control is performed using the electric power charged to the secondary battery 26 when the actuator 17 is driven. Further, since the secondary battery 26 and the actuator 17 are disposed close to each other in the torsion bar, a voltage drop due to the wiring is small, and thus the power loss can be reduced. Further, since the control circuit 25 and the secondary battery 26 are provided in the torsion bar, they also function as the inertia mass 15, and the rigid body resonance frequency of the torsion bar in the vehicle longitudinal direction can be reduced.

In particular, the vibration damping device for a vehicle according to the present example controls the rigid body resonance frequency of the torsion bars 5 and 6, and it is necessary to perform vibration damping control under operating conditions of a fundamental order close to that in the high rotation region of the engine, while vibration transmitted from the engine 1 to the support portion of the engine becomes large in the low rotation region of the engine 1. Therefore, the charging mode is set when the engine 1 is operated in the low rotation speed region, and the control mode is set when the engine is operated in the high rotation speed region, whereby the self-sufficiency of the electric power can be realized.

In the vehicular vibration damping device of the present example, the low rotation speed region of the engine 1 preferably includes at least an idle rotation state or a charging state when the vehicle is a hybrid vehicle (for example, a state in which the engine is operated to charge a battery during a stop of the vehicle). In such an operating state of the engine 1, the vibration transmitted to the support portion of the engine 1 becomes large, and therefore, the amount of charge to the secondary battery 26 can be increased.

It is preferable that the charging mode is set at least when the fundamental order of rotation of the engine 1 coincides with the natural frequency of the inertia mass 15 in the axial direction. When the fundamental order of rotation of the engine 1 coincides with the natural frequency of the inertia mass 15, the displacement of the inertia mass 15 increases, and therefore the amount of charge to the secondary battery 26 increases, and efficient charging is possible.

On the contrary, it is preferable that the control mode is set at least when the resonant frequency of the torsion bars 5 and 6 in the front-rear direction is substantially equal to the rotation order of the engine 1. Since the rigid resonance of the torsion bars 5, 6, which deteriorates the vehicle interior noise, can be suppressed, the vehicle interior noise can be reduced.

Further, in order to efficiently perform charging, the vibration of the engine 1 is transmitted to the torque rods 5 and 6 as much as possible, and in order to ensure quietness of the vehicle, the vibration of the engine 1 is not transmitted to the vehicle. Therefore, it is preferable that the rigidity of the bushes 12, 13 be set higher in the torque supporting direction than in the torque axial direction of the bush 13 fixed to the vehicle body side, at least in the charge mode, in the bush 12 fixed to the engine side. Since the rigidity of the engine-side bush 12 is high, vibration of the engine 1 is transmitted to the torsion bars 5 and 6, and charging can be performed efficiently, and since the rigidity of the vehicle-side bush 13 is low, transmission of vibration to the vehicle can be blocked, and the vehicle can be kept quiet.

In the vehicle vibration damping device of the present example, since the acceleration sensor 21 and the control circuit 25 including the band-pass filter 22 and the voltage amplification circuit 23 are mounted on the substrate 24, the work such as wiring arrangement is not required, and the cost can be reduced.

As shown in fig. 6B, the base plate 24 of the present example can be fixed to the surface of the bracket 1a of the engine 1 to which the bush 12 is fixed by the bolt 18, on the engine 1 side. However, as shown in fig. 6A, it is more preferable that the base plate 24 is fixed to a side surface of the bracket 1a of the engine 1 on a side away from the engine 1 opposite to the engine 1.

In the vibration damping device for a vehicle of the present example, as shown in fig. 10, the rigidity of the vehicle body side bushing 12 of the upper torsion bar 5 is significantly reduced as compared with the conventional upper torsion bar, and therefore, for example, when the vehicle turns, the upper torsion bar is greatly swung in the lateral direction of the vehicle by the acceleration of the upper torsion bar itself. Therefore, it is necessary to set the clearance C between the engine 1 and the upper torque rod 5 to be large (see fig. 6A and 6B).

On the other hand, the force generated by the unbalanced inertial force acts on the front side of the center of gravity G of the engine 1, and therefore a moment is generated. Therefore, the vibration of the engine due to the vertical displacement of the engine 1 at the front end thereof becomes large. Therefore, as in the example shown in fig. 6, by arranging the position of the acceleration sensor 21 on the surface of the housing 20 of the upper torsion beam 5 on the side away from the engine 1 on the side opposite to the engine 1, the clearance C between the upper torsion beam 5 and the engine 1 can be shortened, and the vertical vibration of the engine 1 transmitted to the upper torsion beam 5 can be reduced. Similarly, since the clearance C between the upper torque rod 5 and the engine 1 can be shortened, the size of the components connected to the torque rod on the engine 1 side can be reduced, and the characteristic value of the components connected to the torque rod can be improved.

In the vehicle vibration damping device of the present example, the upper torsion bar 5 includes the actuator 17 as a heat source, and heat transfer to the acceleration sensor 21 is problematic, but the acceleration sensor 21 can be disposed at a position opposite to the engine 1 where wind current from the front of the vehicle blows, and therefore the heat radiation performance is also advantageous.

The secondary battery 26 corresponds to a power storage device and a power control device of the present invention, and the acceleration sensor 21 corresponds to a vibration detection device of the present invention.

Description of the reference numerals

1. An engine; 2. an auxiliary frame; 3. 4, suspending an engine; p1, P2, bearing points; 5. an upper torsion bar; 6. a lower torsion bar; 11. a rod; 12. 13, a bushing; 15. an inertial mass; 17. a driver; 18. 19, bolts; 20. a housing; 20A, an opening part; 21. an acceleration sensor; 22. a band-pass filter; 23. a voltage amplifying circuit; 24. a substrate; 25. a control circuit; 26. a secondary battery is provided.

Claims (9)

1. A vibration isolation device for a vehicle, comprising:

a rod having one end fixed to the engine and the other end fixed to the vehicle body;

an inertia mass supported by the rod;

a driver for reciprocating the inertial mass in an axial direction of the rod;

a power control unit for driving the driver;

a power storage unit for supplying electric power to the electric power control unit; wherein,

the power control means converts the vibration of the inertial mass into electric power and charges the electric power into the power storage means when the actuator is not driven,

the vibration isolation device for a vehicle further includes:

a vibration detecting member for detecting vibration of the rod in an axial direction;

the vibration detection member is attached to the base plate together with the electric power control member and the power storage member, and is attached to the rod.

2. The vibration isolation device for a vehicle according to claim 1,

the base plate is disposed on a portion of the lever on a side away from the engine.

3. The vibration isolation device for a vehicle according to claim 1,

the vibration isolation device for a vehicle further includes:

a housing for covering the lever, one end portion and the other end portion of the lever, and accommodating the inertial mass and the actuator;

the substrate is attached to an opening of the housing, and the actuator is housed in the housing in an airtight or watertight manner.

4. The vibration isolating device for a vehicle according to claim 1,

the vibration detection member is disposed between the one end portion and the other end portion of the lever, and is positioned on a surface parallel to a horizontal plane that supports the shaft through the torque of the lever.

5. The vibration isolating device for a vehicle according to claim 1,

the power control device charges the power storage member when the engine is at a low rotation speed lower than a predetermined value, and supplies the power charged in the power storage member to the actuator when the engine is at a high rotation speed equal to or higher than the predetermined value.

6. The vibration isolation device for a vehicle according to claim 5,

the electric power storage device is charged when the engine is in an idle rotation state or an engine charging state in the hybrid vehicle.

7. The vibration isolation device for a vehicle according to claim 5,

and charging the power storage member when at least a basic rotation order of the engine matches an axial natural frequency of the inertia mass.

8. The vibration isolation device for a vehicle according to claim 5,

the electric power charged in the power storage member is supplied to the actuator at least when the resonance frequency of the rod in the front-rear direction is equal to the basic rotation order of the engine.

9. The vibration isolation device for a vehicle according to claim 1,

when the power storage device is charged, the rigidity of the rod in the torque support direction, which is fixed to one end of the rod on the engine side, is higher than the rigidity of the rod in the torque axial direction, which is fixed to the other end of the rod on the vehicle body side.