WO2013161544A1

WO2013161544A1 - Vibration-damping device for vehicle

Info

Publication number: WO2013161544A1
Application number: PCT/JP2013/060464
Authority: WO
Inventors: 金堂　雅彦; 裕介佐藤; 圭一郎松本
Original assignee: 日産自動車株式会社
Priority date: 2012-04-27
Filing date: 2013-04-05
Publication date: 2013-10-31

Abstract

A vibration-damping device for a vehicle is provided with: a rigid rod (11) having one end mounted to the engine (1) side and the other end mounted to the vehicle body side; and bushes (12, 13) provided at both ends of the rigid rod, that is, between the engine side and the rigid rod, and between the vehicle body side and the rigid rod. If the mass of the rigid rod is m, the moment of inertia of the rigid rod is I, a first distance, which is the distance from the center of gravity of the rigid rod to the center of mounting on the vehicle body side, is a, and a second distance, which is the distance from the center of gravity of the rigid rod to the center of mounting on the engine side, is b, the expression of I/mab ≤ 1 ± 0.1 is satisfied.

Description

Vibration isolator for vehicle

The present invention relates to a vibration isolator for a vehicle that suppresses vibration transmitted from an engine that is a vibration source to a vehicle body side.

In a vibration isolator that connects an engine and a vehicle body side, supports an inertial mass on a torque rod having insulators at both ends, and reciprocates the inertial mass in the axial direction of the torque rod by an actuator. By setting the actuator lower than the engine bending / torsion resonance frequency and generating a force proportional to the speed of the axial displacement of the torque rod in the actuator, the damping of the torque rod can be increased while maintaining the damping characteristics of the insulator. An anti-vibration device has been proposed in which torque rod rigid body resonance suppression in the torque rod axial direction and double anti-vibration are compatible (Patent Document 1).

That is, the rigid resonance frequency in the axial direction of the torque rod is set to a frequency lower than the main elastic resonance frequency of the engine, and the vibration is amplified by the engine elastic resonance at a frequency higher than the rigid resonance frequency. Is not transmitted to the vehicle, and at the same time, the rigid resonance, which is a trade-off with this, is to suppress the resonance with the actuator of the torque rod.

JP 2011-12757 A

However, in the conventional torque rod, since the rigid resonance frequency in the axial direction, which is the longitudinal direction of the vehicle, is greatly reduced, the rigid resonance frequency in the pitch direction of the torque rod is also greatly reduced, thereby causing vibrations in the vertical direction of the engine. However, this increases the amplitude of the torque rod in the pitch direction. As a result, vibration in the vertical direction is transmitted to the vehicle body, which mainly increases the booming noise.

Also, in such a conventional torque rod, if the vibration detection sensor mounting position for detecting vibration and feeding back to the axial rigid body resonance suppression control of the torque rod deviates vertically from the shaft supporting the torque, the pitch Large-amplitude vibration is detected, and a signal in the pitch direction unnecessary for axial resonance suppression control is output from the actuator. As a result, the vehicle longitudinal input (rod axial vibration input) is increased. As a result, there is a problem that the sound inside the vehicle deteriorates and the power consumption increases.

For this reason, the applicant proposed in Japanese Patent Application No. 2011-166467, which is a patent application filed by the applicant, to place the sensor mounting position on the axis, but this is a solution for the vehicle longitudinal direction input. However, since the above-mentioned vehicle up-down direction input is not reduced, it is not a solution to the increase in the booming noise. Even if the position of the center of gravity of the vibration isolator is set on the line of force, only the axial direction is controlled because the rod axial direction (vehicle longitudinal direction) vibration and pitching vibration are coupled during actuator control. An actuator that cannot be used cannot control the vertical vibration input to the vehicle as described above. Therefore, it is not sufficient to place the center of gravity on the line of action in order to cut off the force transmitted by the pitching vibration of the torque rod (input in the vertical direction to the vehicle body) to the vehicle body.

In particular, in order to satisfy the durability of the insulator for supporting a high torque engine, the rigidity distribution of the engine-side insulator and the vehicle body-side insulator is adjusted to increase the rigidity of the vehicle-side insulator or increase the displacement. If the engine's vertical vibration increases and the input to the torque rod increases, the vertical input to the vehicle body due to the pitching vibration of the torque rod greatly increases due to the mass characteristics of the torque rod and the center of gravity. I found out. That is, if various requirements are satisfied, the dynamic spring characteristic in the vertical direction of the vehicle (the ratio of the engine vibration displacement input to the torque rod and the resulting transmission force to the vehicle) cannot be reduced, and the interior noise is increased. It may end up.

The problem to be solved by the present invention is to provide a vibration isolator for a vehicle capable of suppressing in-vehicle sound generated by transmission from an engine.

The present invention relates to a rigid rod having bushes at both ends, one end attached to the engine side and the other end attached to the vehicle body side, wherein the mass of the rigid rod is m, the inertia moment of the rigid rod is I, the rigid rod When the distance from the center of gravity of the vehicle body to the mounting center on the vehicle body side is a, and the distance from the center of gravity of the rigid rod to the mounting center on the engine side is b, I / mab ≦ 1 ± 0.1 is set. To solve the above-mentioned problem.

The dynamic spring characteristics of the rigid rod for cutting off the force transmitted from the engine to the vehicle body are summarized as follows: f _z / z ₀ · cosωt = k _z1 k _z2 (I-mab) ω ² / α, and when I-mab = 0 (I / mab = 1), the dynamic spring constant becomes extremely small. The mass m of the rigid rod, the inertia moment I of the rigid rod, and the distances a and b from the center of gravity of the rigid rod to both ends can be set to I / mab ≦ 1 by appropriately setting the mass and shape of the rigid rod. It can be ± 0.1. As a result, it is possible to suppress in-vehicle sound generated by transmission from the engine.

1 is a front view illustrating an example in which a vibration isolator according to an embodiment of the present invention is applied to an engine of a vehicle. It is a top view of FIG. 1A. It is a disassembled perspective view of FIG. 1A and FIG. 1B. It is sectional drawing which shows the basic structure of the upper torque rod of FIG. 1B. It is a front view which shows the specific structure of the upper torque rod of FIG. 1B. It is a graph which shows the relationship of the transmission force level with respect to the vibration frequency of the rod axial direction of the upper torque rod of FIG. 1B. It is a figure which shows the model of the upper torque rod of FIG. 1B. It is a graph which shows the relationship of the transmission force level with respect to the vibration frequency of the up-down direction of the upper torque rod of FIG. 1B. It is a side view of FIG. 1A. It is a graph which shows the relationship of a, b, I of the upper torque rod of FIG. 1B. It is a graph which shows the relationship of a displacement when a preload acts on the upper torque rod of FIG. FIG. 9 is a graph showing the relationship between the vibration frequency and the dynamic spring constant when I / mab ≦ 1 ± 0.1 under the condition that the preload acting on the upper torque rod of FIG. FIG. 9 is a graph showing the relationship between the vibration frequency and the dynamic spring constant when I / mab ≦ 1 ± 0.1 under the condition that the preload acting on the upper torque rod in FIG. 8 is full load. It is a front view which shows the relationship between a and b of the upper torque rod of FIG. 1B. It is a graph which shows the relationship between a and mab in FIG. 12A. FIG. 6 is a perspective view and a front view showing still another specific structure of the upper torque rod of FIG. 1B. It is the front view and side view which show the other specific structure of the upper torque rod of FIG. 1B. It is a front view which shows the other specific structure of the upper torque rod of FIG. 1B. FIG. 16 is a perspective view showing the upper torque rod of FIG. 15. FIG. 10 is a cross-sectional view of a main part showing still another specific structure of the upper torque rod of FIG. 1B. FIG. 18 is a cross-sectional view of a main part showing another specific structure of the upper torque rod of FIG. 17. FIG. 10 is a cross-sectional view of a main part showing still another specific structure of the upper torque rod of FIG. 1B.

First, a so-called pendulum engine 1 to which a vehicle vibration isolator according to an embodiment of the present invention can be applied will be described. As shown in FIGS. 1A and 1B, the support structure of the engine 1 by the pendulum system supports the engine 1 with respect to a so-called horizontal engine 1 in which the inertia main shaft L of the engine 1 is arranged as illustrated. The two support points P1 and P2 are positioned on opposite sides of the center of gravity G on the inertial main axis L of the engine 1 in the plan view of FIG. 1B, and P1 in the side view of FIG. 1A. Is a support structure provided on the inertial main axis L and P2 is positioned above the inertial main axis L in the vehicle. The two support points P1, P2 are constituted by left and right engine mounts 3, 4 as shown in FIG.

The support structure of the pendulum type engine is a torque rod attached to the vehicle body with an engine center of gravity G swinging around a straight line connecting the support points P1 and P2 while supporting the engine 1 like a pendulum. It is configured to be restrained by a rod-like member such as the assemblies 5 and 6 (hereinafter also referred to as the upper torque rod 5 and the lower torque rod 6), and has the merit that the vibration damping effect similar to the conventional one can be obtained with a small number of parts. . That is, in the engine 1 mounted in the pendulum system, the engine 1 is tilted around the axis connecting the two support points P1 and P2 by the rotational inertia force when the engine 1 is operated. In order to prevent the inclination and support the engine 1, the upper torque rod 5 that connects the substantially upper half of the engine 1 and the vehicle body side member, and the lower torque that connects the remaining lower half of the engine 1 and the vehicle body side member. Rod 6. The upper torque rod 5 is connected to the engine 1 from the upper right side of the vehicle, and the other lower torque rod 6 is connected to the engine 1 from the lower side of the vehicle. The two

torque rods

5, 6 cause the pendulum engine 1 to tilt. To prevent.

The engine 1 is, for example, an in-line 4-cylinder engine. In particular, an engine with a relatively large displacement (2L or more) is often equipped with a balance shaft. In this case, since the unbalanced inertia force is small at the basic order (secondary component) of the engine rotation, Reaction force of engine torque fluctuation acts on the engine 1. Accordingly, it has been found by the present inventor that, in the basic order of engine rotation, in-vehicle sound and in-vehicle vibration are mainly generated by input from the two

torque rods

5 and 6 that support torque. (This is also the case with the multi-link engine proposed by the applicant.) Furthermore, when the vehicle is mainly accelerated, the in-vehicle sound up to about 1000 Hz, which is composed of a high order basic order, is a problem for passengers. It is known to be.

As described above, the vibration isolator for a vehicle of this example includes two

torque rods

5 and 6. The upper torque rod 5 is mounted between the upper part of the engine 1 and the vehicle body as shown in FIG. 1B. On the other hand, the lower torque rod 6 is mounted between the lower portion of the engine 1 and the subframe 2 as shown in FIGS. 1A, 1B and 2. Since the basic configuration of the upper torque rod 5 and the lower torque rod 6 in this example is the same, the configuration of the upper torque rod 5 will be described, and the description of the configuration of the lower torque rod 6 will be omitted by using this.

FIG. 3 is a cross-sectional view of the main part showing the basic structure of the upper torque rod 5 according to this example, and FIG. 4 is a front view showing the specific structure of the upper torque rod. 3 illustrates the basic structure of the upper torque rod 5 of this example, the portion connecting the

bushes

12 and 13 is constituted by a shaft-like rod 11, and the actuator 17 (including the inertia mass 15) is the rod 11. The thing of the form surrounding the circumference of was shown. On the other hand, FIG. 4 shows a configuration in which a portion connecting the

bushes

12 and 13 includes a housing 20 that houses the actuator unit 25 in order to explain a specific structure of the upper torque rod 5 of this example. The housing 20 in FIG. 4 functionally corresponds to the rod 11 in FIG. 3 as a member for connecting the

bushes

12 and 13 and supporting the actuator unit 25. In FIG. 3, the housing 20 may be regarded as an illustration omitted. In any example, the actuator is positioned relative to the rod 11 or the housing 20 so as to have a center of gravity on the axis of the upper torque rod 5 (a line connecting the centers of the cylinders in the thickness direction center of the bushes 12 and 13). It is attached to be coaxial. Hereinafter, the basic structure of the upper torque rod 5 will be described with reference to FIG. 3, and then the specific structure of the upper torque rod 5 will be described with reference to FIG. 4.

As shown in FIG. 3, in the upper torque rod 5 of this example, a pair of

bushes

12 and 13 are fixed to both ends of a rod-shaped rod 11 (corresponding to the housing 20 in the case of FIG. 4) by welding. The bush 12 fixed to the engine side includes a cylindrical outer cylinder 12a, a cylindrical inner cylinder 12b concentric with the outer cylinder 12a, and an elastic body (vibration isolation material) that connects the outer cylinder 12a and the inner cylinder 12b. ) 12c. The bush 12 is fixed to the engine 1 by a bolt 18 (see FIG. 2) that is inserted into the inner cylinder 12b in a direction orthogonal to the paper surface in FIG.

On the other hand, the bush 13 fixed to the vehicle body also has a cylindrical outer cylinder 13a, a cylindrical inner cylinder 13b concentric with the outer cylinder 13a, the outer cylinder 13a and the inner cylinder 13b, like the bush 12. And an elastic body (vibration isolation material) 13c for connecting the two. The bush 13 is fixed to a member on the vehicle body side by a bolt 19 (see FIG. 2) that is inserted into the inner cylinder 13b in a direction orthogonal to the paper surface in FIG. When expressed simply as a rod, it means the entire rod rigid body including the outer cylinder 12a of the bush on the engine side, the outer cylinder 13a of the bush on the vehicle body side, and the rod 11 or the housing 20 connecting them.

In the illustrated embodiment, the bush 12 is fixed to the engine 1 and the bush 13 is fixed to the vehicle body side. However, the embodiment is not limited thereto, and the bush 12 is fixed to the vehicle body side and the bush 13 is fixed to the engine 1. It may be fixed. 3 shows an example in which two

bolts

18 and 19 inserted through the

inner cylinders

12b and 13b of the

bushes

12 and 13 are arranged in parallel, as shown in FIGS. The upper torque rod 5 shows an example in which two

bolts

18 and 19 inserted through

inner cylinders

12b and 13b of

bushes

12 and 13 are arranged in directions orthogonal to each other. These directions can be changed as appropriate according to the shapes of the vehicle-side fixing portion and the engine fixing portion.

The elastic bodies (vibration-proofing materials) 12c and 13c of this example are members having both a spring and a damping function, and for example, elastic rubber can be used.

In the upper torque rod 5 of this example, the diameters of the outer cylinder and the inner cylinder of the

bushes

12 and 13 are different. That is, the diameters of the outer cylinder 13a and the inner cylinder 13b of the bush 13 are made relatively smaller than the diameters of the corresponding outer cylinder 12a and the inner cylinder 12b of the bush 12, and the rigidity of the elastic body 13c of the bush 13 is further increased. The rigidity of the elastic body 12c of the bush 12 is set to be relatively larger. Thereby, engine rigid body resonance and rod rigid body resonance in the rod axis direction suitable for double vibration isolation can be generated by setting the rigidity of the

elastic bodies

12c and 13c of the pair of

bushes

12 and 13.

If the engine rigid body resonance and the rod rigid body resonance are described based on a spring mass system that is extremely simplified for easy understanding, the engine rigid body resonance A is the engine mass and the rigidity of the elastic body 12c of the bush 12. The rod rigid body resonance B is determined by (spring constant), and the mass of the rod 11 (and the outer cylinder 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 bush 13 It is determined by the rigidity (spring constant) of the elastic body 13c. By setting the diameter and rigidity of the

bushes

12 and 13 as described above, as shown in FIG. 5, the engine rigid body resonance A in the rod axis direction determined from the rigidity of the elastic body 12c of the bush 12 has a frequency f1 close to 10 Hz. The rod rigid body resonance B in the rod axis direction which occurs at [Hz] and is determined from the rigidity of the elastic body 13c of the bush 13 occurs at a frequency f2 [Hz] close to 200 Hz.

Since the primary resonance frequency of bending and twisting of the engine 1 alone is about 280 Hz to 350 Hz in a general vehicle engine, the engine rigid body resonance A is about 10 Hz and the rod rigid body resonance B is about 200 Hz as in this example. Then, the transmission of resonance vibrations of bending and torsion of the engine 1 to the vehicle body is effectively suppressed (double vibration isolation) on the high frequency side (within the vibration isolation region).

From the above, the rigidity (spring constant) of the elastic body 12c of the bush 12 and the elastic body 12c of the bush 12 are set so that the engine rigid body resonance A and the rod rigid body resonance B have a frequency lower than the resonance frequency of bending and torsion of the engine. And the mass of the rod 11 (the actuator 17 and the outer cylinder portion of each bush), which is the mass between the elastic body 13c of the bush 13 and the rigidity (spring constant) of the elastic body 13c of the bush 13. As described above, the engine rigid body resonance A and the rod rigid body resonance B are generated at two different frequencies, that is, at the frequency f1 in the low frequency region and the frequency f2 in the middle frequency region, and transmitted from the engine 1 to the vehicle body side. It is the effect of the double vibration isolation that the effect of preventing the generated vibration is obtained. However, in the vibration isolator of the present invention, it is not essential that the diameters of the outer cylinder and the inner cylinder of the

bushes

12 and 13 are different, and the

bushes

12 and 13 may have the same structure.

3, the upper torque rod 5 of this example includes an inertial mass 15 made of a magnetic metal or the like, an actuator 17, an acceleration sensor 21, a band-pass filter 22, and a voltage amplification circuit 23.

The inertial mass 15 is provided around the rod 11 coaxially with the rod 11. The section of the inertial mass 15 viewed in the axial direction of the rod 11 has a point-symmetric shape with the center (center of gravity) of the rod 11 as the center, and the center of gravity of the inertial mass 15 coincides with the center of the rod 11. The inertia mass 15 has 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 leaf spring having a relatively small rigidity. A part of the inner wall 15a of the inertia mass 15 is protruded toward a permanent magnet 17c of the actuator 17 described later.

In the upper torque rod 5 of this example, an actuator 17 is provided in the space between the inertia mass 15 and the rod 11 as shown in FIG. The actuator 17 is a linear type (linear motion type) actuator including a square cylindrical core 17a, a coil 17b, and a permanent magnet 17c, and reciprocates the inertia mass 15 in the axial direction of the rod 11.

The core 17a constituting the magnetic path of the coil is made of laminated steel plate and fixed to the rod 11. The core 17a is divided into a plurality of members before the assembly of the upper torque rod 5, and the plurality of members are bonded to the periphery of the rod-shaped rod 11 with an adhesive, thereby forming a rectangular tube as a whole. A core 17a is formed. The coil 17b is wound around the square cylindrical core 17a. The permanent magnet 17c is provided on the outer peripheral surface of the core 17a.

Since the actuator 17 has such a configuration, the inertial mass 15 is linearly moved by the reluctance torque generated by the magnetic field generated by the coil 17b and the permanent magnet 17c, that is, the inertial mass 15 is reciprocated in the axial direction of the rod 11. Will be driven.

As shown in FIG. 4, the specific structure of the upper torque rod 5 of this example includes a housing 20 that accommodates the actuator unit 25, which is omitted in FIGS. 2 and 3, and the housing 20 includes the bush 12 and the bush 13. They are rigidly connected. The specific structure of the actuator unit 25 is the same as that shown in FIG. 3, and includes an inertia mass 15, an elastic support spring 16, and an actuator 17. A shaft fixed to the housing 20 is provided inside the actuator unit 25 instead of the rod 11. ing. Here, the torque support shaft of the torque rod 5 can be defined as the line of action of the axial force transmitted to the torque rod due to the rotational inertia force (torque) of the engine. , 13 has a uniform cylindrical shape with almost no change in the thickness direction, and the axis of the torque rod obtained by connecting between the cylindrical center of the bush 12 and the cylindrical center of the bush 13 at the center of each bush thickness direction. It can be considered the same as the center (the center of the rod or the shaft center of the actuator 17).

As described above, the rod rigid body resonance must be a frequency smaller than the resonance frequency of the bending and torsion of the engine, and the rigidity of the

bushes

12 and 13 is relatively low (substantially soft compared to a general one). Therefore, even when the position of the center of gravity relative to the axis of the housing 20 is slight, the vibration in the axial direction of the actuator is likely to be coupled with the vibration in the pitch direction, and the controllability of the vibration control is deteriorated. As in this example, with the torque rod 5 having a symmetrical cross section including the actuator 17 (inertia mass 15) or the actuator unit 25, the center of gravity of the entire torque rod is brought closer to the axis of the rod 11 or the housing 20. It is possible to suppress the vibration in the axial direction from exciting the pitch vibration, and as a result, it is possible to suppress the deterioration of controllability.

As indicated by a dotted line in FIG. 4, on the horizontal plane between the

bushes

12 and 13 and passing through the axis of the rod 11, the acceleration of the axial vibration at the substantially axial position of the rod 11 is from the engine 1. An acceleration sensor 21 that detects the acceleration of vibration transmitted to the rod 11 is attached. The rod axis direction acceleration signal from the acceleration sensor 21 is input to the voltage amplification circuit 23 via the bandpass filter 22, and the signal amplified by the voltage amplification circuit 23 is applied to the coil 17 b of the actuator unit 25. (Voltage control is performed). The voltage amplifier circuit 23 can be composed of, for example, an operational amplifier. The acceleration sensor 21 is located in the vicinity of the rotation center axis of the pitching of the rod 11 from the front side of the paper surface to the depth direction of the paper surface, and is hardly affected by the pitching. That is, since the excitation force (vibration) of the engine is relatively large in the vertical direction, in FIG. 4 in which the upper side in the paper is the upper side in the gravity direction, the pitching vibration in the direction of shaking in the paper is larger than the other directions. Therefore, according to the arrangement of the sensor, there is an effect that the pitching vibration of the engine 1 can be detected without including the displacement in the rod axis direction due to the pitching of the rod 11 as much as possible.

The inertial mass 15 is supported by a relatively soft leaf spring (elastic support spring 16). For example, resonance of the inertial mass 15 with respect to the rod 11 in the rod axis direction occurs at a low frequency from 10 Hz to 100 Hz. For example, since the secondary vibration frequency of the idle speed of a four-cylinder engine is about 20 Hz, if the resonance frequency of the inertial mass 15 can be set to 10 Hz, the inertial mass 15 resonates regardless of the operating conditions of the engine 1. Can be suppressed.

On the other hand, if it is attempted to set the resonance frequency of the inertial mass 15 to such a low frequency such as 10 Hz, if the inertial mass 15 becomes too large and such setting is difficult, the rod rigidity resonance B ( If the frequency is set lower than about 1/2 of 200 Hz) in the embodiment, the resonance frequencies are sufficiently separated from each other, and vibration transmission is sufficiently suppressed.

Further, by passing the acceleration signal detected by the acceleration sensor 21 through the band-pass filter 22, it is possible not to perform control at an extra frequency, thereby improving control stability and suppressing the extra power consumption. It is possible to reliably suppress the transmission force in the range. Then, in order to perform speed feedback control that increases the attenuation of the rod to be controlled, in the frequency band passing by the band pass filter 22, it is substantially proportional to the speed in the rod axis direction of the vibration detected by the acceleration sensor 21. A force having the opposite sign is generated from the actuator unit 25.

Next, the relationship between the rod 11 shown in FIG. 3 or the housing 20 shown in FIG. 4 and the

bushes

12 and 13 will be described.

In the upper torque rod 5 of this example, as shown in the dynamic model diagram of FIG. 6, the mass of the rod 11 or the housing 20 (hereinafter also simply referred to as the rigid rod 11) that is a rigid body is m (inertia mass 15 and actuator). 17), the inertia moment around the center of gravity G1 of the rigid rod 11 is I, the first distance from the center of gravity G1 of the rigid rod to the mounting center O2 on the vehicle body side is a, and the mounting center on the engine side from the center of gravity G1 of the rigid rod 11 is The mass, shape, and the like of the rigid rod 11 are set so that the relationship of I / mab ≦ 1 ± 0.1 is established when the second distance to O1 is b. That is, in the rigid rod 11 of this example, the relationship of 90% ≦ I / mab ≦ 110% is established. Hereinafter, this principle will be described.

FIG. 6 is a diagram showing a dynamic model of the upper torque rod 5 shown in FIGS. 3 and 4, in which the right side of the figure shows the engine as a vibration source and the left side shows the vehicle body on the vibration transmission side. In FIG. 6, the rigidity (dynamic spring coefficient) k _x1 of the bush 12 in the x-axis direction, the rigidity (dynamic spring coefficient) k _z1 in the z-axis direction, the damping coefficient c ₁ , and the rigidity (dynamic spring coefficient) of the bush 13 in the x-axis direction. ) K _x2 , stiffness in the z-axis direction (dynamic spring coefficient) k _z2 , damping coefficient c ₂ , here, first, without considering the damping coefficients c ₁ and c ₂ of the

bushes

12 and 13, rigidity (dynamic spring coefficient) Only consider.

As shown in the figure, from the right side of the engine and the x-axis direction vibration and the z-axis direction of the x ₀ cos .omega.t against the rigid rod 11 and the vibration of z ₀ cos .omega.t inputted, z-axis direction (the rigid rod 11 As for the vibration (excitation force) fz transmitted to the vehicle body side in the pitching direction of the vehicle body and the vertical vibration of the vehicle body, the following equation 1 is established from the equation of motion related to the center of gravity G1 of the rigid rod 11.

Further, since the vibration (excitation force) fz in the z-axis direction is also affected by the rotational moment of the rigid rod 11, the inertia moment around the center of gravity G 1 of the rigid rod 11 is I, the displacement in the rotational direction is θ, and the bush 12 Assuming that the rigidity in the rotational direction (dynamic spring coefficient) k _θ1 and the rigidity in the rotational direction of the bush 13 (dynamic spring coefficient) k _θ2 , the following equation 2 is established.

When the

above formulas

1 and 2 are arranged, the following

formulas

3 and 4 are obtained.

Solving Equation 3 with respect to θ yields Equation 5 below, and when Equation 4 is organized with θ, Equation 6 follows.

Substituting Equation 5 into θ in Equation 6 and organizing it with z yields Equation 7 below.

Here, when the multiplier of z on the left side of the above equation 7 is set as α as in the following equation 8, the above equation 7 is arranged as in the following equation 9.

Here, if the ratio K between the transmission force fz from the engine to the vehicle body and the displacement excitation z ₀ cos ωt from the engine is expressed by the following equation 10,

Substituting the above θ and z into Equation 10 and rearranging results in Equation 11 below.

In Equation 11, if the direction of rotation of the rigid (dynamic spring coefficient) k _.theta.2 direction of rotation of the rigid (dynamic spring coefficient) k _.theta.1 and the bush 13 of the bush 12 is negligible relative to other elements (k _{.theta.1 =} k _{When θ2} = 0) and I-mab = 0 on the right side, the transmission force fz in the z-axis direction from the engine to the vehicle body becomes 0, and vibration in the vertical direction of the vehicle body is suppressed.

In the case where the rotational direction of the stiffness (dynamic spring coefficient) k _.theta.2 direction of rotation of the rigid (dynamic spring coefficient) k _.theta.1 and the bush 13 of the bush 12 is negligible compared to the other elements present embodiment as described above, I The mass, shape, etc. of the rigid rod 11 are set so that −mab = 0, thereby suppressing the vibration in the z-axis direction transmitted to the vehicle body, but I−mab = 0 (I = mab, I / mab) = Mab / I = 1), but not limited to I / mab ≦ 1 ± 0.1 or I / mab ≦ 1 ± 0.06 or I / mab ≦ 1 ± 0.02. Also good. In other words, in addition to I / mab = 1, 90% ≦ I / mab ≦ 110%, 94% ≦ I / mab ≦ 106%, or 98% ≦ I / mab ≦ 102%.

The value of this I / mab can be appropriately set according to the required vehicle interior noise level, and the present inventors have confirmed that the level of transmission force (dB) from the engine to the vehicle body is higher than the original performance. If the vehicle interior noise level falls within the allowable range if it falls by -6 dB, it may be set to I / mab ≦ 1 ± 0.1, and if it falls by −10 dB, I / mab ≦ 1 ± if it falls within the acceptable range. What is necessary is just to set to 0.06, and if it falls in -20dB and if it will enter into an allowable range, I / mab <= 1 +/- 0.02 should just be set. FIG. 7 shows the measurement results of the transmission force with respect to the vibration frequency of the example in which I / mab = 99% and the comparative example in which I / mab = 87%, and the comparative example in which I / mab = 99% is set. It was confirmed that the transmission force level was reduced by 20 dB or more compared to the above.

Incidentally, in FIG. 6, since the rod 11 and the housing 20 are rigid bodies, their center of gravity G1 does not change, but the first distance a from the center of gravity G1 to the fixed center O2 of the bush 13 to the vehicle body and the center of gravity G1 to the bush 12 Since the

bushes

12 and 13 are elastic bodies, the second distance b to the fixed center O1 to the engine varies depending on the torque generated by the engine and the preload acting on the bush of the torque rod. For example, as shown in the side view of FIG. 8, torque Te centered on the crankshaft is generated in the engine 1 in accordance with the operating conditions such as torque, thereby preloading the upper torque rod 5 in its axial direction. P _L is generated. For example, although the preload P _L when the load of the engine 1 is low is 0, the more the preload P _L the larger load increases.

Therefore, in FIG. 6, when the preload increases, the engine 1 falls to the rigid rod 11 side. Therefore, both the fixed center point O1 between the bush 12 and the engine 1 and the fixed center point O2 between the bush 13 and the vehicle body are shown in FIG. Displace to the left. For this reason, although both the first distance a and the second distance b are reduced, the displacement amount of the first distance a and the displacement amount of the second distance b are different due to the rigidity of the

bushes

12 and 13.

FIG. 10 is a graph obtained by measuring the amount of displacement at the center point O1 of the bush 12 when the bush 13 side is fixed (the amount of displacement is zero) and the preload indicated by the arrow in FIG. The larger the value, the larger the displacement. In the vehicle vibration isolator of this example, the main purpose is to suppress the vehicle interior noise up to about 1000 Hz, which is mainly composed of a high order of the basic order during acceleration of the vehicle, which is transmitted to the vehicle interior. The rotation speed of the engine corresponding to the vibration frequency region targeted for vibration suppression, that is, the first distance a and the second distance b when a preload corresponding to the rotation speed is applied is I / mab ≦ 1 ± 0. It is preferable to configure the rigid rod 11 so as to satisfy the relational expression 1.

Figure 11A is a vibration frequency and dynamic spring constant when the preload P _L acting on the rigid rod 11 is set to I ≒ mab for the first distance a and a second distance b in the case of zero (correlated with transmission force dB vibration to) a graph showing the relationship, FIG. 11B, the vibration frequency when preload P _L acting on the rigid rod 11 is set to I ≒ mab for the first distance a and a second distance b when at full load It is a graph which shows the relationship with a dynamic spring constant (it correlates with the transmission force dB of vibration).

In the rigid rod 11 shown in FIG. 11A, the dynamic spring constant is small when the preload of the engine 1 is small. However, as the engine approaches the full load, the dynamic spring constant increases, and as a result, the level of vehicle interior noise increases. . On the other hand, in the rigid rod 11 shown in FIG. 11B, the dynamic spring constant increases when the preload of the engine 1 is small, but the engine 1 is not fully loaded as in the case where the vehicle accelerates and the rotational speed increases. The closer to, the smaller the dynamic spring constant, and as a result, the level of vehicle interior noise does not increase. Thus, by using the first distance a and the value of the second distance b when the preload P _L corresponding to the load area to be damping purposes is applied, it is desirable to apply the I / mab ≦ 1 ± 0.1 It can be said. Generally, since it is desired to reduce the full load noise, the setting of FIG. 11B is desirable.

FIG. 9 is a graph showing an iso-inertia moment I line where I = mab is established when the first distance a is 65 to 80 mm and the second distance b is 40 to 55 mm. In the figure, the hatched portion of the torque rod A is a combination example in which the first distance a is set to a relatively small value, and shows a combination when the preload increases in the direction of the arrow. On the other hand, the hatched portion of the torque rod B is a combination example in which the first distance a is set to a relatively large value, and shows a combination when the preload increases in the direction of the arrow. As described above, it is desirable to set a combination of the first distance a, the second distance b, and the moment of inertia I when the preload PL is in the full load region.

Now, as shown in FIG. 12A, if the distance between the fixed center point O1 on the engine side of the upper torque rod 5 and the fixed center point O2 on the vehicle body side is L, the first distance a and the second distance b are obtained. On the other hand, since L = a + b, a combination of a = b, a <b or a> b is conceivable depending on the setting of the position of the center of gravity G1 of the rigid rod 11 (corresponding to the housing 20 in the example in the figure). Since the first distance a is 0 <a <L, if the first distance a takes an arbitrary value within this range, mab = ma (La) = ma ² −mLa, as shown in FIG. 12B. The value of mab takes a maximum value L ² m / 4 when a = L / 2, ie, a = b. Therefore, when the mass, shape, and the like of the rigid rod 11 are designed so that a = b, the mab, that is, the moment of inertia I is maximized. As a result, it is necessary to increase the mass of the rigid rod 11 itself or to increase the shape. In order to avoid this, in the upper torque rod 5 of this example, the mass of the rigid rod 11 is set so that the first distance a and the second distance b are not equal, that is, a <b or a> b is established. Set the shape.

Incidentally, in order for the rigid rod 11 to satisfy I / mab ≦ 1 ± 0.1, the mass m of the rigid rod 11, the position of the center of gravity G1 that affects the first distance a and the second distance b, The inertia moment I may be adjusted. Of these, the inertia moment I can be adjusted, for example, by providing an adjustment weight 24 shown in FIGS. 13 and 14. In the example of the upper torque rod 5 shown in FIG. 13, an adjustment weight portion 24 is provided at the end of the housing 20 on the bush 12 side that is away from the axis connecting the centers O ₁ and O ₂ of the

bushes

12 and 13. Since the inertia moment I around the center of gravity G1 of the housing 20 which is a rigid rod is smaller as the distance from the center of gravity G1 is larger, the weight 24 for adjustment is removed from the center of gravity G1 in order to reduce the weight of the entire rigid rod 11. This is because it is preferable to separate them as much as possible. In addition, you may provide the weight part 24 for adjustment shown in the figure not the bush 12 side but the bush 13 side.

On the other hand, in the example of the upper torque rod 5 shown in FIG. 14, the material constituting the housing 20 at the end of the housing 20 on the bush 12 side on the axis connecting the centers O ₁ and O ₂ of the

bushes

12 and 13. The weight part 24 for adjustment comprised with the material of specific gravity larger than these specific gravity is provided. That is, the adjustment weight portion 24 is made of a material different from the housing 20 and made of a high specific gravity material, so that the overall length of the housing 20 that is a rigid rod can be suppressed and the size can be reduced. If the adjustment weight 24 is provided at a position separated from the axis connecting the centers O ₁ and O ₂ of the

bushes

12 and 13 as in the example of FIG. 13, the adjustment weight 24 is further reduced. can do.

In the above-described embodiment, in the rigid rod 11 (including the housing 20) including the inertia mass 15 and the actuator 17, the rigidity (dynamic spring coefficient) _kθ1 of the bush 12 and the rigidity of the bush 13 in the rotation direction. The present inventors have confirmed that (dynamic spring coefficient) _kθ2 is often negligible compared to other elements. However, as shown in FIGS. 15 and 16, an inertia mass 15 and an actuator 17 are provided. If the torque rod 5 that is not used is set so as to satisfy the relationship of I / mab ≦ 1 ± 0.1, the same effects as the above-described vibration damping action can be obtained.

However, even when the direction of rotation of the rigid (dynamic spring coefficient) k _.theta.2 direction of rotation of the rigid (dynamic spring coefficient) k _.theta.1 and the bush 13 of the bush 12 is not negligible as compared with other elements, in the right side of expression 11 described above The rigidity (dynamic spring coefficient) _{kθ1 in} the rotational direction of the bush 12 and the rigidity (dynamic spring coefficient) _{kθ2 in} the rotational direction of the bush 13 are set to 0 as _much as possible, and I / mab ≦ 1 ± 0.1. If set so as to satisfy the relationship, the transmission force fz in the z-axis direction from the engine to the vehicle body approximates zero. Hereinafter, embodiments will be described in the case where the rotational direction of the stiffness (dynamic spring coefficient) k _.theta.2 direction of rotation of the rigid (dynamic spring coefficient) k _.theta.1 and the bush 13 of the bush 12 can not be ignored.

In the upper torque rod 5 shown in FIGS. 15 and 16, a pair of

bushes

12 and 13 are disposed at both ends of the housing 20 constituting the rod-like rigid rod 11. The bush 12 fixed to the engine side includes a cylindrical outer cylinder 12a, a cylindrical inner cylinder 12b concentric with the outer cylinder 12a, and an elastic body 12c that connects the outer cylinder 12a and the inner cylinder 12b. . The bush 12 is fixed to the engine 1 by bolts 18 (see FIG. 2) that are inserted into the inner cylinder 12b in a direction orthogonal to the paper surface of FIG. On the other hand, the bush 13 fixed to the vehicle body also has a cylindrical outer cylinder 13a, a cylindrical inner cylinder 13b concentric with the outer cylinder 13a, the outer cylinder 13a and the inner cylinder 13b, like the bush 12. And an elastic body 13c for connecting the two. The bush 13 is fixed to the vehicle body member by a bolt 19 (see FIG. 2) that is inserted into the inner cylinder 13b in a direction parallel to the paper surface of FIG.

In the upper torque rod 5 of this example, as shown in FIG. 16, when the axial direction of the rod 11 is the X axis, the axial direction of at least one of the

bushes

12 and 13 is the first in the pitch direction of the engine 1. It is set along the Y-axis direction corresponding to one axis. In the example shown in FIGS. 15 and 16, the fixed axis Ly of the bush 12 on the engine 1 side is set parallel or substantially parallel to the Y-axis direction, and the fixed axis Lz of the bush 13 on the vehicle body side is parallel or substantially parallel to the Z-axis direction. It is set in parallel. In the

bushes

12 and 13 constituted by the

outer cylinders

12a and 13a, the

inner cylinders

12b and 13b, and the

elastic bodies

12c and 13c, the dynamic spring coefficient in the shearing direction is smaller than the dynamic spring coefficient in the compression direction. rotational direction of the stiffness of the

bushes

12 and 13 because the (dynamic spring coefficient) k _.theta.1 and k _.theta.2 can be reduced. In this sense, there is no restriction on the layout of the engine 1, and it can be said that it is most desirable to set the fixed shafts of both

bushes

12 and 13 parallel or substantially parallel to the Y-axis direction.

In order to set the fixed shafts of the

bushes

12 and 13 along the first axis in the pitch direction of the engine 1, it is most preferable to set them in parallel, but more preferably parallel ± 15 °, preferably parallel ± 30. °.

Like the bush 12 shown in FIGS. 15 and 16, the dynamic spring coefficient k _θ1 in the rotational direction can be reduced by setting the fixed axis Ly of the bush 12 parallel or substantially parallel to the Y-axis direction. In order to further reduce the dynamic spring coefficient in the rotation direction, for example, a configuration as shown in FIG. FIG. 17 is a cross-sectional view showing a bush 12 according to another example corresponding to the bush 12 of FIG. 15, and in this example, the inner cylinder 12b shown in FIG. 15 is composed of two inner cylinders 12b1 and 12b2. Are provided with a rolling bearing 12d. 18 is a cross-sectional view of the main part showing a bush 12 according to still another example corresponding to the inner cylinders 12b1 and 12b2 shown in FIG. 17, and in this example, a sliding bearing is provided between the two inner cylinders 12b1 and 12b2. 12e is provided. In any of the examples, the bolt 18 shown in FIG. 2 is inserted through the inner cylinder 12b1. Thus, by providing the rolling bearing 12d or the sliding bearing 12e between the inner cylinders 12b1 and 12b2, the dynamic spring constant in the rotational direction of the bush 12 can be reduced to almost zero.

In the example shown in FIGS. 17 and 1, a bearing is provided between the two inner cylinders 12b1 and 12b2 to reduce the dynamic spring coefficient in the rotational direction. However, the inner cylinder 12b shown in FIGS. A bearing may be provided between the elastic body 12c and between the elastic body 12c and the outer cylinder 12a, or the outer cylinder 12a may be constituted by two outer cylinders and a bearing may be provided between them. In short, what is necessary is just to provide a bearing in at least any one between the inner cylinder 12b fixed with the volt | bolt 18 shown in FIG. 2, and the outer cylinder 12a.

On the other hand, the bush 13 shown in FIGS. 15 and 16 cannot be set along the Y axis corresponding to the first axis in the pitch direction of the engine 1 due to, for example, the layout restriction of the engine 1, as shown in FIG. The fixed axis Lz of the bush 13 is set along the Z-axis direction as a second axis that is perpendicular to the first axis in the pitch direction. Therefore, with respect to the vibration in the pitch direction of the engine 1, a force in the compression and tension directions acts on the elastic body 13c constituting the bush 13, and the dynamic spring coefficient becomes larger than the force in the shear direction. However, in the bush 13 of this example, the thickness t1 of the upper end and the lower end of the elastic body 13c is made thicker than the thickness t2 of the center as shown in FIG. The dynamic spring coefficient is made small.

In order to make the thickness t1 of the upper end portion and the lower end portion of the elastic body 13c of the bush 13 thicker than the thickness t2 of the center portion, as shown in FIG. In addition to scraping, the inner surface of the upper end portion and the lower end portion of the outer cylinder 12a may be scraped instead of or together with this. Further, in order to reduce the dynamic spring coefficient in the pitch direction of the engine 1, it is most desirable to increase the thickness of both the upper end portion and the lower end portion of the elastic body 13c. Also good.

As described above, in the vehicle vibration isolator of this example, the structure of the torque rod 5 satisfies the relationship of I / mab ≦ 1 ± 0.1, preferably 1 ± 0.06, more preferably 1 ± 0.02. Therefore, the excitation force transmitted to the vehicle body can be significantly reduced by the vertical vibration and pitch vibration of the torque rod 5 caused by the vertical vibration of the engine 1 (ideally, it can be reduced to zero). it can). Therefore, it is possible to provide a comfortable vehicle interior acoustic space by suppressing deterioration of vehicle interior noise caused by vibration of the torque rod.

Further, when the vibration eigenvalue of the torque rod 5 matches the vibration eigenvalue of the vehicle body, the noise level increases. However, the vibration eigenvalue of the torque rod changes as the

bushes

12 and 13 are preloaded as the rotational speed of the engine 1 increases. However, in order to make the car body correspond to it, the weight is greatly increased by improving the rigidity of the skeleton. However, in this example, even if such a fluctuation of the vibration eigenvalue occurs, the structure is difficult to transmit the vibration, which contributes to the weight reduction of the vehicle body and improves the design freedom of the vehicle.

Furthermore, when designing the vibration eigenvalue of the torque rod 5 itself, it is sufficient to design with emphasis on the vibration eigenvalue of the torque rod 5 in the axial direction, and design without worrying about the dynamic spring constant in the direction perpendicular to the axis. Can do.

In addition, the upper torque rod 5 of this example is configured to receive engine torque in the direction of an action line connecting the fixed center O1 of the bush 12 on the engine side and the fixed center O2 of the bush 13 on the vehicle body side. Since 13 is bent in the rod axis direction by preloading, the first distance a and the second distance b change. For this reason, by constructing the engine 1 so as to satisfy the relationship of I / mab ≦ 1 ± 0.1 with the a and b dimensions when the excitation force of the engine 1 is large and the influence on the vehicle body is large, Can be improved more.

Further, the upper torque rod 5 of this example is a so-called active control type torque rod 5 in which the inertia mass 15 is vibration-controlled by an actuator 17, and the axial dynamic rigidity of the bush 13 is reduced so that the torque rod doubles. In order to exhibit the anti-vibration effect, the vibration eigenvalue in the axial direction is set as small as about 200 Hz. As a result, the pitch vibration in the vertical direction enters the normal range (70 Hz to 150 Hz, which corresponds to a rotational speed of about 2000 rpm to 4500 rpm for the secondary of the four cylinders), and noise such as a booming noise increases. If the vibration eigenvalue of the torque rod is set to a high frequency in order to eliminate this influence, control cannot be performed in conjunction with the original axial rod resonance, or the control current increases and the power consumption increases. Problems arise. However, by designing a rigid rod that satisfies the relationship of I / mab ≦ 1 ± 0.1 as in this example, even if the torque rod vibrates in the vertical direction, the excitation force to the vehicle body in the vertical direction can be reduced. Since it can be greatly reduced, the vehicle interior noise is greatly reduced.

In the upper torque rod 5 of this example, the first distance a and the second distance b are set to be unequal distances. If the first distance a and the second distance b are set to be equal, the mab becomes the maximum value, and if the inertia moment I is set to it, the rigid rod itself becomes large or increases in weight, but as in this example This problem can be avoided by making the first distance a and the second distance b unequal. In particular, by taking as large as possible a first distance a from the fixed center O ₂ of the body-side bush 13 to the center of gravity G1, as the torque rod 5 can be reduced in size and weight.

In the upper torque rod 5 of this example, the adjustment weight portion 24 is provided at the end of the housing 20 on the bush 12 side away from the axis connecting the centers O ₁ and O ₂ of the

bushes

12 and 13, so that the weight is small. The inertia moment I can be adjusted, and the weight of the rigid rod 11 as a whole can be reduced.

In the upper torque rod 5 of this example, a material having a specific gravity larger than the specific gravity of the material constituting the housing 20 is formed at the end of the housing 20 on the bush 12 side on the axis connecting the centers O ₁ and O ₂ of the

bushes

12 and 13. Since the configured adjustment weight portion 24 is provided, the overall length of the housing 20 that is a rigid rod can be suppressed and downsizing can be achieved.

In the upper torque rod 5 of the present example, the fixed axis Ly of the bush 12 on at least one of the engine side and the vehicle body side is set along the Y axis in the pitch direction of the engine. The dynamic spring coefficient k _{θ of the} vibration is reduced, and as a result, even if the torque rod vibrates in the vertical direction, the excitation force to the vehicle body in the vertical direction can be greatly reduced, so that the vehicle interior noise is greatly reduced. .

DESCRIPTION OF SYMBOLS 1 ... Engine 2 ...

Sub-frame

3, 4 ... Engine mount P1, P2 ... Supporting point 5 ... Upper torque rod 6 ... Lower torque rod 11 ...

Rod

12, 13 ... Bush 15 ... Inertial mass 17 ...

Actuator

18, 19 ... Bolt 20 ... Housing 21 ... Acceleration sensor 22 ... Band pass filter 23 ... Voltage amplification circuit 24 ... Weight part for adjustment 25 ... Actuator unit

Claims

A rigid rod with one end attached to the engine side and the other end attached to the vehicle body side;
In the vehicle vibration isolator, comprising both ends of the rigid rod, and bushes provided between the engine side and the rigid rod and between the vehicle body side and the rigid rod,
The mass of the rigid rod is m, the moment of inertia of the rigid rod is I, the first distance from the center of gravity of the rigid rod to the mounting center on the vehicle body side is a, the center of gravity of the rigid rod to the mounting center on the engine side When the second distance of b is b,
I / mab ≦ 1 ± 0.1
Is a vehicle vibration isolator for which
2. The vehicle vibration isolator according to claim 1, wherein the I / mab ≦ 1 ± 0.1 is established with respect to the first distance a and the second distance b in the engine operating region for vibration isolation purposes. .
The rigid rod is
An inertial mass supported by the rigid rod;
The vibration isolator for vehicles according to claim 1 or 2 including an actuator which reciprocates said inertial mass in the direction of an axis of said rigid rod.
The vehicle vibration isolator according to any one of claims 1 to 3, wherein the first distance a and the second distance b are set to different distances.
The vehicle weight according to any one of claims 1 to 4, wherein an adjustment weight portion for setting an inertia moment I of the rigid rod is provided apart from an axis connecting the centers of the two bushes. Anti-vibration device.
6. The vehicle vibration isolator according to claim 5, wherein the adjustment weight section is a member different from the rigid rod and is made of a material having a higher specific gravity than the rigid rod.
The vehicular vibration isolation according to any one of claims 1 to 6, wherein a fixed shaft of at least one bush on the engine side and the vehicle body side is set along a first axis in a pitch direction of the engine. apparatus.
The bush with the fixed shaft set along the first shaft includes an inner cylinder portion fixed to the engine or the vehicle body, an outer cylinder portion fixed to the rigid rod, the inner cylinder portion and the outer cylinder The vehicle vibration isolator according to claim 7, further comprising a bearing portion provided between the cylinder portion and the bearing portion.
The bush in which the fixed shaft is set along the second axis perpendicular to the first shaft includes an inner cylinder portion fixed to the engine or the vehicle body, and an outer cylinder portion fixed to the rigid rod. And an elastic body provided between the inner cylinder part and the outer cylinder part,
The vibration isolator for a vehicle according to claim 7 or 8, wherein the elastic body is formed such that a thickness of at least one of both end portions in the second axial direction is larger than a thickness of a central portion.