JP2012072795A - Vibration reducing device - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a device for actively reducing vibrations input to a vehicle body via an engine mount.SOLUTION: The device comprises a rod 11 that couples a first elastic body 112c on the engine side and a second elastic body 113c on the vehicle body side, an inertial mass 15 supported by the rod 11, an actuator 17 that reciprocates the inertial mass 15 in an axial direction of the rod 11, an acceleration sensor 21 that detects acceleration of vibration in the axial direction of the rod 11 and the vibration in the crankshaft rotating direction, and a control means 22 that controls the actuator 17 so as to generate a force proportional to the speed of the rod 11. The actuator 17 is arranged so that the force in the axial direction of the rod 11 and the moment in the crankshaft rotating direction act on the rod 11. The control means control the actuator 17 based on the detected value of the acceleration sensor 21 so as to reduce the amplitude level of the resonance frequency of the rigid body resonance in the axial direction of the rod 11 and the rigid body resonance in the crankshaft rotating direction.

Description

  The present invention relates to a vibration reducing device that reduces vibration transmitted from an internal combustion engine to a vehicle body.

  By connecting the diaphragm to an intermediate plate that divides the rubber part as an insulator into two parts in order to reduce vibration from the engine to the vehicle body, the resonance point is made different between the partitioned outer rubber part and inner rubber part. Patent Document 1 describes the effect of double vibration isolation.

Japanese Patent Laid-Open No. 9-273586

  By the way, in the case of obtaining the effect of double vibration isolation, the transmission force transmitted from the engine to the vehicle body becomes large at a frequency near the resonance point. Therefore, in order to further reduce the vibration from the engine to the vehicle body, the resonance itself must be reduced. It is necessary to suppress it.

  In this case, when the damping of the rubber part is increased, the transmission force at a frequency near the resonance point is reduced and the resonance itself is suppressed. However, in the high frequency range above the resonance frequency, the transmission force becomes larger than before the attenuation is increased, and the transmission characteristic to the vehicle body side member on the high frequency side is deteriorated. Further, if the damping of the rubber part is simply increased in order to suppress the resonance, the double anti-vibration effect is deteriorated.

  This problem can be solved by providing an actuator that can generate an excitation force on a rod that connects the vehicle body side insulator and the engine side insulator. However, the rigid resonance of the rod includes resonance in the axial direction of the rod and resonance in the direction in which the rod is rotated so that a straight line perpendicular to the axial direction of the rod is the center of rotation (hereinafter referred to as pitch direction). If an actuator is provided for each resonance direction, there is a problem that the cost increases.

  Therefore, an object of the present invention is to provide an apparatus that can suppress resonance itself using an actuator while suppressing the double vibration isolation effect, and can suppress an increase in cost due to the use of the actuator.

  The vibration reducing device of the present invention is supported by a first elastic body attached to the engine side, a second elastic body attached to the vehicle body side, a rod connecting the first elastic body and the second elastic body, and the rod. And an actuator that reciprocates the inertial mass in the axial direction of the rod. Furthermore, an acceleration sensor that detects acceleration due to vibration in the axial direction of the rod and vibration in the rotation direction of the crankshaft, and control means that controls the actuator to generate a force proportional to the speed of displacement of the rod.

  The actuator is arranged so that an axial force of the rod and a moment in the rotation direction of the crankshaft can act on the rod. Then, the control means controls the actuator based on the detection value of the acceleration sensor so as to suppress the amplitude level of the resonance frequency of the rigid resonance in the axial direction of the rod and the rigid resonance in the rotation direction of the crankshaft.

  According to the present invention, the actuator is disposed so that the axial force of the rod and the moment in the rotation direction of the crankshaft can act on the rod, and the acceleration of the rod due to the force and the moment can be detected by the acceleration sensor. Therefore, it is possible to suppress both the rigid body resonance in the axial direction and the pitch direction of the rod with one actuator, and the cost is not increased.

It is a schematic perspective view which shows a mode that 1st Embodiment of the vibration reduction apparatus by this invention was applied to the engine mount structure of the pendulum system. It is a top view which shows the torque rod from which the double vibration isolating effect is acquired. It is a frequency characteristic figure of the transmission force by the structure by which the effect of double vibration isolation is acquired. It is a diagram which shows the physical model of a torque rod. It is the top view which looked at the torque rod assembly of a vehicle-mounted state from upper direction. 2 is a block diagram functionally representing a voltage amplification circuit and an actuator. It is the diagram which expanded the deformation | transformation of the outer cylinder of a large end part when engine rigid body resonance has arisen in the large end part. It is a diagram which shows the physical model of the torque rod assembly of 1st Embodiment. It is a frequency characteristic figure of the transmission force by the torque rod assembly of a 1st embodiment. It is a characteristic view which shows the effect of the noise at the time of acceleration of 1st Embodiment. It is a diagram which shows an example of the map for setting the excitation force for the booming sound reduction of 1st Embodiment. It is a characteristic view which shows the effect of the booming sound of 1st Embodiment. It is a figure which shows the structure of the torque rod of 1st Embodiment. It is a figure which shows the structure of the torque rod of 2nd Embodiment.

  Hereinafter, embodiments of the present invention will be described with reference to the drawings.

(First embodiment)
FIG. 1 is a schematic perspective view showing a state in which a vibration reduction device according to a first embodiment of the present invention is applied to a Pendulum engine mount structure.

  The engine 1 is an engine in which an unbalanced inertia force does not act on the basic order of engine rotation, and only a reaction force of engine torque fluctuations acts mainly. Such engines include, for example, a 4-cylinder engine with a secondary balancer and a V-type 6-cylinder engine. The engine 1 is a horizontal type in which a crankshaft is placed in the left-right direction of the vehicle. In the present embodiment, the right side of the vehicle is the engine front.

  The structure that reduces vibration transmitted from the engine 1 is a part of the structure that supports the engine 1. The engine 1 is supported at two locations above the center of gravity by a right engine mount 3 and a left engine mount 4. The right engine mount 3 supports the engine 1 from the right side of the vehicle. The left engine mount 4 supports the engine 1 from the left side of the vehicle. Such a support method is called a pendulum method.

  In the Pendulum-type engine mount structure, the engine 1 is tilted about an axis connecting two mount points by a rotational inertia force during operation. In order to prevent this inclination, an upper rod 11-1 and a lower rod 11-2 are provided. The upper rod 11-1 is provided on the upper right side of the vehicle, and has one end connected to the engine 1 and the other end connected to the vehicle body 2. The upper rod 11-1 has a rod shaft 111 attached in parallel to the ground. The lower rod 11-2 is provided on the lower side of the vehicle, and has one end connected to the engine 1 and the other end connected to the vehicle body 2. The lower rod 11-2 also has the rod shaft portion 111 attached in parallel to the ground.

  For a four-cylinder engine with a secondary balancer or a V-6 engine, an unbalanced inertial force with the basic order of engine rotation (secondary engine rotation for an in-line four-cylinder engine and third rotation of an engine for a V-6 engine) Does not act, and only the reaction force of engine torque fluctuation mainly acts. Accordingly, the inventors of the present invention have found that in the basic order, the in-vehicle sound and the in-vehicle vibration are generated by the vibration transmitted to the vehicle body via the torque rod supporting the torque. Furthermore, the present inventors have found that the in-vehicle sound up to about 1000 Hz, which is mainly composed of a high-order basic order, is a problem for the occupant when the vehicle accelerates.

  Therefore, the present inventor has a configuration in which a double anti-vibration effect is obtained in order to reduce vibration transmitted from the engine 1 to the vehicle body via the upper rod 11-1 and the lower rod 11-2. We propose a new torque rod assembly with a structure that can reduce vibration.

  The basic structure of the upper rod 11-1 and the lower rod 11-2 is the same. Therefore, in the following, the rod 11 will be described when it is not necessary to distinguish between them.

(About the structure which can obtain the double anti-vibration effect)
FIG. 2 is a plan view showing a torque rod that does not include active vibration reducing means and that can obtain a conventional double vibration isolation effect.

  The torque rod 11 shown in FIG. 2 can be expected to have a certain degree of vibration isolation effect due to the double vibration isolation effect. This point will be described.

  In the torque rod 11, both ends of the rod shaft portion 111 are a large end portion 112 and a small end portion 113.

  The large end portion 112 includes an outer cylinder 112a, an inner cylinder 112b, and an elastic body 112c.

  The outer cylinder 112 a is welded to the rod shaft portion 111.

  The inner cylinder 112b is concentric with the outer cylinder 112a. As shown in FIG. 1, the inner cylinder 112 b is fixed to the engine 1 through the bolt 18.

  The elastic body 112c is interposed between the outer cylinder 112a and the inner cylinder 112b. The elastic body 112c is, for example, elastic rubber. The elastic body 112c has not only elasticity but also damping properties.

  The basic structure of the small end portion 113 is the same as that of the large end portion 112. That is, the small end portion 113 includes an outer cylinder 113a welded to the rod shaft portion 111, an inner cylinder 113b concentric with the outer cylinder 112a, and an elastic body 113c interposed between the outer cylinder 112a and the inner cylinder 112b. ,including.

  In this embodiment, the large end portion 112 and the small end portion 113 have different diameters of the outer cylinder and the inner cylinder. That is, the diameter of the outer cylinder 113 a of the small end portion 113 is smaller than the diameter of the outer cylinder 112 a of the large end portion 112. The diameter of the inner cylinder 113b of the small end portion 113 is smaller than the diameter of the inner cylinder 112b of the large end portion 112. Further, the rigidity of the elastic body 113 c of the small end portion 113 is larger than the rigidity of the elastic body 112 c of the large end portion 112.

  As described above, the large end outer cylinder 112a and the small end outer cylinder 113a are welded, that is, rigidly coupled to the rod shaft portion 111. Therefore, in the following, the rod end 111 having the large end outer cylinder 112a and the small end outer cylinder 113a welded to the rod shaft portion 111 will be appropriately referred to as a rod rigid body 110.

  In such a torque rod, two resonance points appear as shown by a solid line in FIG.

  One is engine rigid body resonance A. The engine rigid body is obtained by rigidly coupling the large end inner cylinder 112b to the engine. The resonance frequency of the engine rigid body resonance A is determined by the engine mass and the characteristics of the large end elastic body 112c.

  The other is the rod rigid body resonance B. The resonance frequency of the rod rigid body resonance B is determined by the mass of the rod rigid body 110, that is, the mass of the rod shaft portion 111, the large end outer cylinder 112a, and the small end outer cylinder 113a, and the characteristics of the small end elastic body 113c.

A general vehicle engine has an elastic primary resonance frequency f ₃ such as bending and twisting of about 280 Hz to 350 Hz. Therefore, the resonance frequency of the engine rigid body resonance A and the resonance frequency of the rod rigid body resonance B are large so as to be smaller than the resonance frequency (hereinafter referred to as engine elastic resonance frequency) f _{3 of} main elastic modes such as bending and torsion of the engine. The characteristics of the end elastic body 112c, the mass of the rod shaft 111, the large end outer cylinder 112a, and the small end outer cylinder 113a, and the characteristics of the small end elastic body 113c are set.

In the present embodiment, as shown in FIG. 3, the resonance frequency of the engine rigid body resonance A is adjusted to a frequency f ₁ [Hz] that is nearly zero. The resonance frequency of the rod rigid body resonance B is adjusted to a frequency f ₂ [Hz] close to 200 Hz.

  If adjusted in this way, resonance vibration of engine bending and torsion (hereinafter referred to as engine elastic resonance vibration) is first prevented by the first bush. Next, it is prevented by the second bush. Accordingly, the engine elastic resonance vibration is doubled and the transmission to the vehicle body is suppressed.

  Thus, even with the torque rod 11, a certain amount of vibration isolation effect is expected due to the double vibration isolation effect. However, it is difficult to obtain a further anti-vibration effect. This point will be described.

  In order to obtain a further anti-vibration effect with the torque rod 11, consider suppressing the rod rigid body resonance B. The engine rigid body resonance A is ignored. In order to suppress the rod rigid body resonance B, it is preferable to increase the attenuation term of the small end elastic body 113c.

  However, when the attenuation term of the small end elastic body 113c is increased, as shown by the broken line in FIG. 3, the transmission force is reduced in the vicinity of the rod rigid body resonance B and the rod rigid body resonance B itself is suppressed. On the other hand, the transmission power increases and the transmission characteristics deteriorate.

  This mechanism is as follows.

  FIG. 4 is a diagram showing a physical model of the torque rod 11.

  From the model shown, the equation of motion for the rod is:

  Further, the input Ft from the rod 11 to the vehicle body 2 is expressed by the following equation (2).

  The transmission characteristic of the torque rod 11 to the vehicle body 2 is expressed by the following equation (3) from the equations (1) and (2).

At frequencies near the rod rigid body resonance B, since the absolute value of m _r omega ² of the absolute value and k _r are approaching -m _r omega ² and k _r is canceled, the transfer characteristic of the vehicle body 2, the formula (3) and thus by the damping coefficient c _r in the denominator of the right side.

Therefore, if the damping coefficient _cr is increased, the transmission force decreases and the rod rigid body resonance B itself is suppressed in the vicinity of the rod rigid body resonance B as shown by the broken line in FIG.

Molecule on the right side of the equation (3) is determined by the stiffness coefficient k _r of the rod axis of the small end, and the damping coefficient c _r of the rod axis of the small end. In the attenuation to such an extent that a normal double anti-vibration effect is obtained, the attenuation coefficient _cr is small and the rigidity coefficient is dominant. However, if by increasing the damping coefficient c _r in the denominator attempts to suppress the rod rigid body resonance B, also interlocked damping coefficient c _r of the molecule. Then, as shown by a broken line in FIG. 3, the transmission force to the vehicle body 2 increases on the frequency range exceeding the resonance frequency f ₂ of the rod rigid body resonance B, and the transmission characteristic to the vehicle body 2 on the high frequency region side deteriorates. .

  Based on the above, with reference to FIG. 5, the configuration of a vibration reducing device that is a premise of the first embodiment of the present invention and that can provide further vibration isolation effects will be described. FIG. 5 is a plan view of the torque rod assembly in the vehicle-mounted state as viewed from above.

  The vibration reducing apparatus 100 includes a torque rod assembly 10, an acceleration sensor 21, a controller 22, and a voltage amplification circuit 23.

  The torque rod assembly 10 includes a torque rod 11, a leaf spring 16, an inertia mass 15, and an actuator 17. The torque rod assembly 10 has one end connected to the engine 1 and the other end connected to the vehicle body 2. The torque rod assembly 10 is attached so that the rod shaft portion 111 is parallel to the ground.

  The torque rod 11 includes a torque rod housing 19 and a torque rod shaft portion 111. The torque rod housing 19 has a large end 112 formed at one end and a small end 113 formed at the other end, and has a gap between the large end 112 and the small end 113. The torque rod shaft portion 111 is press-fitted along the line connecting the large end portion 112 and the small end portion 113 into the space of the torque rod housing 19 with the leaf spring 16, the inertia mass 15 and the actuator 17 provided. Is done.

  Two leaf springs 16 are provided on the engine side and the vehicle body side of the rod shaft portion 111. The leaf spring 16 is an elastic part. The leaf spring 16 has a relatively small rigidity.

  The inertia mass 15 is provided around the rod shaft portion 111. The inertia mass 15 is a rectangular tube type as shown in FIG. The inertia mass 15 is coaxial with the rod shaft portion 111. As shown in FIG. 5, the inertia mass 15 is fixed to both left and right ends of the leaf spring 16. A plate spring on the vehicle body side is fixed to the vehicle body side end of the side wall of the inertia mass 15. An engine-side leaf spring is fixed to the engine-side end of the side wall of the inertia mass 15. That is, the fixed portion between the leaf spring 16 and the inertia mass 15 extends from the front side to the back side. That is, it is parallel to the direction of gravity. Inertial mass 15 is a magnetic metal body. The cross section of the inertial mass 15 is bilaterally symmetric and vertically symmetric. The center of gravity of the inertia mass 15 coincides with the center of the rod 11. A part of the inner wall 15 a of the inertia mass 15 is convex toward the permanent magnet 17 c of the actuator 17.

  Since the inertial mass 15 is supported by a leaf spring 16 having relatively small rigidity, the resonance frequency in the rod axis direction (vertical direction in FIG. 5) is in a low range from 10 Hz to 100 Hz. Since the secondary vibration frequency of the idle speed of the 4-cylinder engine is about 20 Hz, if the resonance frequency of the inertial mass 15 is 10 Hz, the inertial mass 15 does not resonate regardless of the operating conditions of the engine 1. However, for the resonance frequency of the inertial mass 15 to be 10 Hz, the inertial mass 15 becomes very heavy. When it is difficult to make the inertial mass 15 heavy, if the resonance frequency of the inertial mass 15 is set lower than about half the frequency of the rod rigid resonance B (200 Hz in this configuration), the mutual resonance frequency is sufficient. The vibration transmission as described later is sufficiently suppressed. However, the frequency of the rod rigid resonance B should not be an integral multiple of the resonance frequency of the inertial mass 15 or a fraction of an integer. By doing so, the low-order and high-order excitation forces of the engine do not simultaneously excite the inertial mass resonance and the rod rigid body resonance, respectively, so that it is difficult for noise to occur.

  The actuator 17 is a linear motion type actuator that reciprocates the inertia mass 15 in the rod axis direction (vertical direction in FIG. 5). As will be described later, the actuator 17 generates a force based on the signal amplified by the voltage amplifier circuit 23 and having the opposite sign. As a result, speed feedback control for increasing the attenuation of the rod 11 to be controlled is performed.

  The actuator 17 is provided in a space between the inertia mass 15 and the torque rod shaft portion 111. The actuator 17 includes a core 17a, a coil 17b, and a permanent magnet 17c.

  The core 17a has a rectangular tube shape. The core 17a is fixed to the torque rod shaft portion 111. The core 17a is composed of a plurality of laminated steel plates. The core 17a constitutes a magnetic path of the coil 17b. In the core 17a, a steel plate is fixed to the periphery of the torque rod shaft portion 111 with an adhesive to form a core 17a having a square tube shape as a whole. The coil 17b is wound around the 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 reciprocates in the rod axis direction by reactance torque generated by the magnetic field generated by the coil 17b and the permanent magnet 17c.

  The acceleration sensor 21 detects the acceleration of vibration in the axial direction of the torque rod 11. The acceleration sensor 21 is attached to the side surface of the torque rod 11.

The controller 22 has a filter function of allowing a signal having a predetermined frequency to pass among signals input from the acceleration sensor 21 and cutting signals having other frequencies. Specifically, the controller 22 includes at least the resonance frequency f ₂ of the rod rigid body resonance B and passes the lower limit frequency f _{5 of the} vibration isolation region. The lower limit frequency of the vibration isolation region is a frequency at which the transmissibility becomes one time, and specifically, a frequency obtained by multiplying the resonance frequency f ₂ of the rod rigid body resonance B by a predetermined value (2 ^1/2 ). It is. More preferably, the controller 22 passes signals up to an upper limit (for example, 400 Hz) at which control does not diverge. In other words, the controller 22 does not pass a signal having a frequency exceeding an upper limit (for example, 400 Hz) at which control does not diverge.

  Further, the controller 22 has a filter function that allows a frequency equal to or higher than the resonance frequency of the inertia mass 15 in the rod axis direction to pass. In other words, the controller 22 does not pass a frequency lower than the resonance frequency of the inertial mass 15 in the rod axis direction. The resonance frequency of the inertial mass 15 in the rod axis direction is determined by the mass of the inertial mass 15 and the rigidity of the leaf spring 16 and is about 10 Hz to 100 Hz. As described above, since the secondary vibration frequency of the idle rotation speed of the four-cylinder engine is about 20 Hz, there is a possibility of coupling when the resonance frequency of the inertia mass 15 in the rod axis direction is 20 Hz. Therefore, it is more desirable to set the pass frequency of the controller 22 so as to avoid coupling.

  Thus, in the above configuration, control is not performed with an extra frequency. Therefore, the control stability is enhanced, and the transmission force can be reliably suppressed in the target frequency range while suppressing excessive power consumption.

  The voltage amplification circuit 23 amplifies the signal input from the controller 22. That is, the voltage amplification circuit 23 amplifies the velocity in the rod axis direction of vibration detected by the acceleration sensor 21. The voltage amplification circuit 23 applies voltage to the coil 17b of the actuator 17 to perform voltage control. The voltage amplifier circuit 23 is an operational amplifier, for example.

  This will be further described.

  FIG. 6 is a block diagram functionally representing the voltage amplification circuit 23 and the actuator 17.

The axial acceleration d ² x _r / dt ² of the rod 11 is detected by the acceleration sensor 21.

  The voltage amplification circuit 23 increases the drive voltage according to the gain G determined by the controller 22.

In the actuator 17, the coil 17b acts as an integrator. Therefore, the controller 22 outputs -G · d ² x _r / dt ² . As a result, the force Fa generated by the actuator 17 is proportional to dx _r / dt. That is, speed feedback control is performed to increase the attenuation of the torque rod 11 that is the control target.

  FIG. 7 is an enlarged diagram of the deformation of the outer cylinder 112a of the large end 112 when the engine rigid body resonance A occurs in the large end 112. FIG.

The resonance frequency f ₁ of the engine rigid body resonance A is close to zero as described above. In this case, the large end outer cylinder 112a is greatly deformed. The vibration of the torque rod 11 does not coincide with the vibration of the large end outer cylinder 112a.

  FIG. 8 is a diagram showing a physical model of the torque rod assembly.

  Here, considering that the rod rigid body resonance B is suppressed, the engine rigid body resonance A is ignored. In FIG. 5, the actual attachment points of the inertia mass 15 are the C point and the D point, but in the physical model of FIG. 8, the E point that is the average of the C point and the D point is “ It is treated as “the attachment point of the inertial mass 15”.

  From the model shown, the equation of motion for the torque rod 11 is the following equation (4).

  Further, the input Ft from the rod 11 to the vehicle body 2 is expressed by the following equation (5).

  The actuator 17 generates a force Fa expressed by the following formula (6).

As can be seen from equation (6), an actuator generating force Fa is first-order differential value of the axial displacement x _r of the torque rod 11, i.e. proportional to the axial velocity of the torque rod 11.

  Substituting equation (6) into equation (4) yields the following equation (7).

From equation (7), it can be seen that the damping term of the torque rod 11 is increased from c _r in c _r + G.

  As described above, the torque rod assembly 10 in which the inertia mass 15 and the actuator 17 are added to the torque rod 11 that can obtain the double vibration isolation effect is used. The controller 22 and the voltage amplification circuit 23 perform speed feedback control. The transmission characteristic to the vehicle body 2 at this time is expressed by the following equation (8) from the equations (5) and (7).

In Equation 8, the coefficient of the right side of the denominator of the damping term, while the c _r + G, since the coefficient of attenuation term of the right side of the molecule does not change a c _r, the influence of the increase in the damping coefficient of the denominator Not receive.

  By doing in this way, a damping coefficient can be increased so that it may affect only the input Fe from the engine 1 which transmits via the large end part 112, and a transmission force falls.

Therefore, by doing so, as shown by the one-dot chain line in FIG. 9, the rod rigid body resonance B can be suppressed, and a vibration isolation effect can be obtained even in a frequency range exceeding the resonance frequency f ₂ of the rod rigid body resonance B. it can.

Incidentally rod axis direction damping coefficient c _r of the small end, the extent to which the normal dual vibration damping effect is obtained, i.e., on the order of values that can sufficiently suppress the transmission power at a higher frequency range than the rod rigid body resonance B .

  Further, the attenuation of the rod rigid body resonance B can be improved in the frequency range that has passed through the controller 22. Thus, the gain G sufficiently reduces the transmission force in the vicinity of the frequency of the rod rigid body resonance B. In other words, the value is set such that the transmission force due to the rod rigid body resonance B does not increase.

  FIG. 10 is a diagram showing the total noise level of the in-vehicle sound from 200 Hz to 1000 Hz when the accelerator pedal is fully depressed under the condition that the engine speed is 3000 rpm and the vehicle is accelerated.

  As can be seen from FIG. 10, according to the above-described configuration, the noise level can be reduced as compared with the comparative example in which the double anti-vibration effect can be obtained.

  The above has been considered to reduce the vibration from the medium frequency range to the high frequency range mainly transmitted from the engine 1 to the vehicle body 2.

  Next, consider reducing vibrations in the low frequency range transmitted from the engine 1 to the vehicle body 2. Such vibration is transmitted as a booming sound.

  The booming noise is generated by engine vibration based on the basic order of engine rotation. The basic order of the four-cylinder engine is the secondary rotation. The basic order of the 6-cylinder engine is the rotational third order.

  The following countermeasures are taken against muffled noise. For example, in an in-line four-cylinder engine, a map illustrated in FIG. 11 is prepared for each engine speed. Then, this map is searched with the engine speed to obtain the magnitude and phase of the amplitude. Then, the optimum excitation force for the engine speed is set by the following equation (9).

  Then, the excitation force F of Expression (9) is applied to the generated force Fa of the actuator 17 of Expression (6).

  Thus, by adding the excitation force F of Expression (9) to the generated force Fa of the actuator 17, as shown in FIG. 12, when the engine speed is low in the in-line four-cylinder engine, Compared to the comparative example in which the excitation force F of the equation (9) is not added to the generated force Fa of the actuator 17, the booming noise (in-vehicle sound) can be reduced.

  In this way, according to the vibration reducing device having the above-described configuration, it is possible to greatly reduce the noise from the low-frequency range to the noise during acceleration.

  Here, the function and effect of the vibration reducing apparatus which is the premise of the above-described first embodiment of the present invention will be described.

  The rod 11 has a resonance frequency of the rod rigid body lower than the engine elastic resonance vibration frequency, and the actuator 17 generates a force proportional to the axial speed of the rod 11 to reciprocate the inertia mass 15 in the axial direction of the rod. Therefore, it is possible to increase the attenuation of the rod 11 while maintaining the attenuation characteristic of the elastic body 113c of the small end portion 113, and to achieve both suppression of the rod rigid body resonance B in the rod axis direction and double vibration isolation. it can.

Further, the filter allows a signal in a predetermined frequency range including at least the resonance frequency of the rod rigid body resonance among the acceleration signals (or velocity signals) in the axial direction of the rod to pass but does not pass a signal outside the range. The actuator generates a force proportional to the axial speed of the rod based on the signal passing through the filter. Since it did in this way, it is possible not to perform the control at the extra frequency, to improve the control stability, and to suppress the transmission force in the vicinity of the rod rigid resonance frequency f ₂ while suppressing the extra power consumption.

Furthermore, since the predetermined frequency range includes a frequency in a vibration isolation region (frequency range of frequency f ₅ or more shown in FIG. 5) existing on the higher frequency side than the frequency f ₂ of the rod rigid resonance B, the rod rigid resonance frequency. The transmission force can be suppressed in the frequency range from f ₂ to the vibration proof region.

Furthermore, since the predetermined frequency range includes the resonance frequency f _{2 in} the rod axis direction of the inertial mass 15 existing on the lower frequency side than the resonance frequency f ₂ of the rod rigid body resonance B, the resonance of the high frequency local deformation is performed. Since the control is not performed, the stability of the control can be improved.

Moreover, resilient elements (leaf spring 16), since the elastic coefficient is determined so that the resonance frequency of the inertial mass 15 is smaller than 1/2 of the rod rigid body resonance frequency f _2, the rod rigid body resonance frequency of the inertial mass 15 it can be separated sufficiently from the resonance frequency f _2.

Further, the rod rigid body includes a rod shaft portion 111, an outer cylinder 112a that is a component part of the engine mounting portion (large end portion 112) and is fixed to one end of the rod shaft portion, and a vehicle body mounting portion (small end portion 113). ) And an outer cylinder 113a fixed to the other end of the rod shaft portion, and the mass of the rod rigid body so that the resonance frequency of the rod rigid body is lower than the engine elastic resonance vibration, and Since the characteristic of the elastic body 113c that is a component of the vehicle body mounting portion and is provided inside the vehicle body mounting portion outer cylinder is set, the rod rigid body resonance frequency f ₂ suitable for double vibration isolation in the inner and outer cylinder bush structure is obtained. Can be set.

  Further, since it is attached to the engine 1 mounted by the pendulum system, it can be controlled mainly by a transmission path through which an input is input, so that a great vibration / noise reduction effect can be obtained.

  Further, the vibration reducing device 100 is mounted on the vehicle with the torque rod shaft portion 111 parallel to the ground. Therefore, the influence of gravity can be avoided when the actuator 17 moves the inertial mass 15. The fixed portion between the leaf spring 16 and the inertial mass 15 is parallel to the direction of gravity. This also makes it possible to avoid the influence of gravity when the actuator 17 moves the inertial mass 15.

By the way, in the configuration which is the premise of the above-described first embodiment of the present invention, in order to suppress in-vehicle sound / in-vehicle vibration generated by vibration transmitted to the vehicle body via the torque rod, the axial rod rigid resonance frequency f ₂ is suppressed. Is below the elastic resonance frequency of the engine. And about the rod rigid body resonance B, the resonance level is suppressed by the actuator 17 arrange | positioned in the torque rod axial part 111. FIG.

In order to set the axial rod rigid body resonance frequency f ₂ to be equal to or lower than the engine elastic resonance frequency, the rigidity of the small end portion 113 may be lowered. However, when the rigidity of the small end portion 113 is reduced, the rigid resonance in the left-right direction in FIG. 5 of the small end portion 113, that is, the pitch direction in which the rod is rotated (a straight line perpendicular to the axial direction of the rod is the center of rotation). May enter the normal range. When the rigid resonance in the pitch direction enters the normal range, there arises a problem that the booming noise increases.

  This problem can also be prevented by controlling the vibration by adding an actuator to the rigid body resonance in the pitch direction. However, with the increase in the number of actuators 17 to be used, there are problems such as an increase in cost and an increase in the size of the torque rod assembly. Arise.

  Therefore, in the first embodiment, the configuration described below suppresses rigid resonance in the axial direction and the pitch direction of the rod 11 while preventing an increase in cost and the like.

  FIG. 13 is a plan view of the torque rod assembly of the present embodiment as viewed from above, as in FIG.

  As shown in FIG. 1, the direction of the central axis of the inner cylinder / outer cylinder of the large end portion 112 and the small end portion 113 may be parallel to the ground (upper rod 11-1) or vertical ( There is also a lower rod 11-2). Therefore, the axial directions of the inner and outer cylinders of the large end portion 112 and the small end portion 113 are not necessarily orthogonal to or parallel to the offset direction of the inertia mass described later (in FIG. 13, they are not orthogonal). An example is shown).

  The large end portion 112 and the small end portion 113 of the first embodiment are composed of an inner cylinder and an outer cylinder with an elastic body interposed therebetween, as in the presupposed torque assembly (FIG. 5), and are members on the engine side or the vehicle body side. The bolt 18 in FIG. 1 is configured to have a predetermined length, that is, a predetermined thickness in the insertion direction. Assuming that the large end portion 112 and the small end portion 113 of the present embodiment are configured symmetrically in the horizontal direction of the paper surface and the depth direction of the paper surface, the centers of the elastic bodies of the large end portion 112 and the small end portion 113 are: It corresponds to the center of the inner and outer cylinders and the center in the thickness direction. Further, even when the configuration is not symmetrical, the point that the load generated between the torque rod mainly acts on the engine side or the vehicle body side member is the large end portion 112 and the small end portion 113, or It can be regarded as the center of those elastic bodies.

  The difference of the first embodiment from the presupposed configuration shown in FIG. 5 is mainly the position of the center of gravity of the inertial mass 15 and the mounting position of the acceleration sensor 21. In this embodiment, the position of the center of gravity of the inertial mass 15 coincides with the rod shaft part 111 (the inertial mass 15 is symmetrically arranged around the rod shaft part 111). You can also. The rod shaft portion 111 is offset from the axis line connecting the large end portion 112 and the small end portion 113 within a plane passing through (including) the straight line connecting the large end portion 112 and the small end portion 113 and orthogonal to the ground. Provided. Therefore, the reciprocating locus of the center of gravity of the inertial mass 15 is at a position offset by a predetermined amount from the axis of the rod.

  For example, the upper rod 11-1 is offset in either the left or right direction on the paper surface of FIG.

  As shown in FIG. 1, in this embodiment, the upper rod 11-1 and the lower rod 11-2 are in a posture in which the axial direction is the lateral direction of the engine, and the inertia mass 15 provided on them is It is attached to the vehicle body side and engine side members so as to be offset. Since the axial direction of each rod is transverse to the engine, for example, when the engine is in a condition where the unbalanced inertial force does not act at the basic order of engine rotation, only the reaction force of engine torque fluctuations acts mainly. Vibration due to the reaction force of torque fluctuation can be effectively suppressed.

  In addition, the moment input in the pitch direction of the rod obtained when the actuator 17 is driven is the pitch vibration of the rod (centered on a straight line parallel to the crankshaft axis direction) generated by the swing of the engine in the pendulum engine mount. The moment in the crankshaft rotation direction is input.

  In other words, in this embodiment in which the rod shaft portion 111 coincides with the center of the inertia mass 15, when the actuator 17 is driven by offsetting the rod shaft portion 111 in this way, the torque rod assembly 10 is moved from the actuator 17. On the other hand, a force in the rod axis direction and a moment in the pitch direction are input. That is, a single actuator 17 can input the torque in the engine torque support direction and the moment in the crankshaft rotation direction to the torque rod assembly 10. The acceleration sensor 21 detects an acceleration obtained by combining the acceleration in the rod axis direction and the acceleration in the pitch direction.

  The acceleration sensor 21 is attached to the side surface of the torque rod casing 19 on the side where the rod shaft portion 111 is offset. The acceleration sensor 21 is configured to be able to detect the vibration in the axial direction of the rod 11 in a plane including the axis of the rod 11 and the reciprocating locus of the center of gravity of the inertia mass 15 and away from the axis of the rod 11. The reciprocating locus of the center of gravity of the inertial mass 15 and the position of the vibration detected by the acceleration sensor 21 are in the same direction as viewed from the axis of the rod 11.

  In the case of the upper rod 11-1, the rod shaft portion 111 is offset to the vehicle lower side, and the acceleration sensor 21 is also arranged on the vehicle lower side. In order to ensure safety in the event of a collision, it is necessary to provide a predetermined distance between the engine hood and the lower engine, etc., but by placing the acceleration sensor 21 below the torque rod casing 19, The distance can be increased.

  On the other hand, in the case of the lower rod 11-2, the rod shaft portion 111 is offset to the upper side of the vehicle, and the acceleration sensor 21 is also offset to the upper side of the vehicle. Thereby, even if the acceleration sensor 21 is attached, the minimum ground height of the vehicle is not affected.

In the configuration shown in FIG. 13, when the inertia mass 15 is moved downward by driving the actuator 17, the force in the rod axis direction acts in the upward direction, and the moment in the pitch direction acts in the clockwise direction. When the acceleration sensor 21 is provided on the side surface of the torque rod housing 19 in the same direction as the offset direction of the rod shaft portion 111, the acceleration sensor 21 detects a vibration in which the vibration in the rod axis direction and the vibration in the pitch direction are combined. can do.

  The torque rod rigid body resonance can be suppressed by executing the same control routine as described in FIG. 5 based on the detection value of the acceleration sensor 21.

Here, a mechanism capable of suppressing vibrations in the rod axis direction and the pitch direction with one actuator 17 with the configuration of FIG. 13 will be described. Here, the “rod axis direction” means the direction of a straight line connecting the large end portion 112 and the small end portion 113 in a stationary state, and the vibration in the rod axis direction is x _s . In addition, when specific numerical values are used in the following description, the numbers are merely examples.

  Equation (10) shows the equation of motion when the torque rod is converted into modal coordinates. In Equation (10), M is a modal mass, K is a modal rigidity, and C is a modal damping. Fa is a force generated by the actuator 17, and the direction from the actuator point Xa of FIG. 13 toward the small end portion 113 is positive. ξ is the displacement of each mode in the mode coordinates.

  φa is a mode vector of the actuator point xa. On the other hand, the mode vector of the sensor point xs is represented by φs. These φa and φs are, for example, as shown in Expression (11).

  The first component of φa and φs is the torque rod rigid body resonance in the rod axis direction, and the second and third components are the rigid body resonances in the pitch direction or vertical direction of the rod 11. Therefore, in the configuration shown in FIG. 13, the second and third components are smaller than the first component.

  Further, the force Fa generated by the actuator 17 is expressed as in Expression (12).

  That is, the force generated by the actuator 17 is obtained by amplifying the speed signal detected by the acceleration sensor 21 with the gain G.

  When the actuator 17 is arranged coaxially with the rod shaft portion 111, the actuator 17 generates only an axial force, so the second and third components of φa are substantially zero. On the other hand, if the gravity center position (rod shaft portion 111) of the inertia mass 15 which is the action point of the generated force of the actuator 1 is offset from the axis line connecting the large end portion 112 and the small end portion 113 as shown in FIG. A moment for rotating the rod 11 in the pitch direction is generated. Therefore, the second and third components of φa are not zero. Furthermore, since the acceleration sensor 21 is also offset in the same manner, the rigid resonance in the pitch direction is measured in addition to the rigid resonance in the axial direction of the rod 11.

  Further, by disposing the acceleration sensor 21 on the same side as the offset direction of the actuator 17, the second and third components of φa and φs have the same sign.

  The result of speed feedback control using the actuator 17 and the acceleration sensor 21 is shown in Expression (13). Moreover, the attenuation | damping Ce as a result of performing speed feedback control is shown in Formula (14).

  As shown in the equation (14), all components of the attenuation Ce increase as the gain G increases, and decrease as the gain G decreases. That is, by operating the actuator 17, vibrations in the rod axis direction and the pitch direction can be attenuated.

  For comparison, consider the case where the acceleration sensor 21 is attached to the side surface of the torque rod casing 19 opposite to the offset direction of the rod shaft portion 111.

  In this case, the mode vector is φa and φs as shown in the equation (15), and the second and third components are reversed in polarity.

  And attenuation Cf at the time of performing speed feedback control becomes as shown in a formula (16).

  As shown in Expression (16), the attenuation Cf has a component that increases as the gain G increases and a component that decreases as the gain G increases. Therefore, if the acceleration sensor 21 is attached to the side surface of the torque rod housing 19 opposite to the offset direction of the rod shaft portion 111, the rigid resonance in the rod axis direction and the pitch direction cannot be effectively attenuated.

  As described above, in the first embodiment, the following effects can be obtained.

  (1) The actuator 17 is arranged so that a force in the rod axis direction and a moment in the crankshaft rotation direction can act on the rod 11. Then, the controller 22 controls the actuator 17 based on the detection value of the acceleration sensor 21 so as to suppress the amplitude level of the resonance frequency of the rigid body resonance in the rod axis direction and the rigid body resonance in the crankshaft rotation direction. Thereby, both the axial resonance of the rod and the rigid resonance in the pitch direction can be suppressed by one actuator, and the cost is not increased.

  (2) The actuator 17 is disposed at a position shifted from a straight line connecting the large end 112 and the small end 113 in a plane including a straight line connecting the large end 112 and the small end 113 and orthogonal to the ground. It is supported by the rod axial part 111 so that it may be. Thereby, the force of a rod axis direction and a pitch direction can be generated with one actuator 17.

  (3) The acceleration sensor 21 is disposed at a position shifted from a straight line connecting the large end 112 and the small end 113 in a plane including the straight line connecting the large end 112 and the small end 113 and orthogonal to the ground. Therefore, the acceleration in the rod axis direction and the pitch direction can be detected.

(Second Embodiment)
FIG. 14 is a plan view of the torque rod assembly of the present embodiment as viewed from above, as in FIG.

  The first difference from the configuration shown in FIG. 13 is the position where the acceleration sensor 21 is provided. In the present embodiment, the acceleration sensor 21 is provided on the wall surface of the gap provided between the large end portion 112 and the small end portion 113 and on the side opposite to the offset direction of the rod shaft portion 111. That is, the acceleration sensor 21 is provided in a space generated by offsetting the rod shaft portion 111.

  Accordingly, it is sufficient to provide a protective cover for waterproofing, dustproofing, and the like of the actuator 17 and the acceleration sensor 21 so as to close the gap provided between the large end portion 112 and the small end portion 113. That is, it is not necessary to provide separate protective covers for the actuator 17 and the acceleration sensor 21.

The second difference is that the acceleration sensor 21 detects acceleration in two directions. Specifically, the acceleration in the rod axis direction and the acceleration in the direction orthogonal to the rod axis direction are detected. The “rod axis direction” is the direction of a straight line connecting the large end portion 112 and the small end portion 113 in a stationary state, as in the first embodiment. Also, let y _s be the vibration in the direction orthogonal to the rod axis direction.

  In other words, the acceleration sensor 21 is in a direction perpendicular to the axial direction of the rod 11 in a plane that includes the axis of the rod 11 and the reciprocating locus of the center of gravity of the inertial mass 15 and is away from the axis of the rod 11. The vibration is configured to be detectable, and the reciprocation locus of the center of gravity of the inertial mass 15 and the position of the vibration in two directions detected by the acceleration sensor 21 are opposite to each other when viewed from the axis of the rod 11. It is comprised so that it may become.

  Note that the acceleration sensor 21 is not a single sensor, but includes two sensors including a second acceleration sensor, and the reciprocating locus of the center of gravity of the inertial mass 15 and the acceleration sensor 21 including these two sensors detect The position of the vibration to be performed may be configured to be opposite to each other when viewed from the axial direction of the rod 11.

  If the acceleration components in the two directions are separately detected in this way, even if the positive and negative of the mode vectors φa and φs are reversed, the single body 17 attenuates the rigid resonance in the rod axis direction and the pitch direction. Can do. The mechanism will be described below.

  The mode vector in the rod axis direction in FIG. 14 is shown in Expression (17) as in Expression (11).

  In addition, a mode vector φs2 in a direction orthogonal to the rod axis direction of the sensor attachment point is represented by Expression (18).

  The force Fa generated by the actuator 17 is obtained by amplifying the difference between the sensor signals with a gain G, and the control signal is as shown in Expression (19).

Then, the results of performing the speed feedback control and the attenuation C _{e2 of} the result of performing the speed feedback control are shown in Expression (20) and Expression (21).

All the components of the attenuation Ce ₂ increase as the gain G increases, and decrease as the gain G decreases. That is, by operating the actuator 17, vibrations in the rod axis direction and the pitch direction can be attenuated.

  In addition, the result of the speed feedback control described above is based on the premise that each component is positive as a result of subtracting φs2 from φs in equation (20). Therefore, it is necessary to install the acceleration sensor 21 so that the mode vector has a relationship as shown in the equations (17) and (18). That is, when the difference between the mode vector in the axial direction of the rod 11 and the mode vector in the direction orthogonal to the axial direction of the rod 11 is obtained, the center of gravity of the inertial mass 15 is set so that all elements have the same sign. The reciprocation locus and the position of the acceleration sensor 21 (that is, the position of vibration in two directions detected by the acceleration sensor 21) need to be set.

  As described above, in this embodiment, in addition to the same effects as those of the first embodiment, there is an effect that it is not necessary to separately provide a protective cover for waterproofing and dustproofing of the actuator 17 and the acceleration sensor 21. .

  The present invention is not limited to the above-described embodiments, and it goes without saying that various modifications can be made within the scope of the technical idea described in the claims.

DESCRIPTION OF SYMBOLS 1 Engine 2 Car body 3 Right side engine mount 4 Left side engine mount 10 Torque rod assembly 11 Rod 15 Inertial mass 16 Leaf spring 17 Actuator 19 Torque rod housing 21 Acceleration sensor 22 Controller 23 Voltage amplification circuit 100 Vibration reduction device 111 Rod shaft part 112 Large End 113 Small end

Claims (10)

A first elastic body attached to the engine side;
A second elastic body attached to the vehicle body side;
A rod connecting the first elastic body and the second elastic body;
An inertial mass supported by the rod;
An actuator for reciprocating the inertial mass in the axial direction of the rod;
An acceleration sensor that detects acceleration due to vibration including vibration in the axial direction of the rod and vibration in the rotational direction of the crankshaft;
Control means for controlling the actuator to generate a force proportional to the speed of displacement of the rod;
Comprising
In a vibration reduction device that reduces vibration transmitted from the engine to the vehicle body,
The actuator is arranged so that an axial force of the rod and a moment in a crankshaft rotation direction can act on the rod,
The control means controls the actuator so as to suppress the amplitude level of the resonance frequency of the rigid resonance in the axial direction of the rod and the rigid resonance in the rotation direction of the crankshaft based on the detection value of the acceleration sensor. Vibration reduction device.   The actuator is disposed on the rod so as to be disposed in a plane that includes a straight line connecting the centers of the first elastic body and the second elastic body and that is perpendicular to the ground and is deviated from the straight line. The vibration reduction device according to claim 1, which is supported.   3. The vibration reduction device according to claim 1, wherein the acceleration sensor is arranged in a plane that includes the straight line and is orthogonal to the ground, and is deviated from the straight line. The acceleration sensor detects only the axial acceleration of the rod,
The vibration reduction device according to claim 2, wherein the actuator and the acceleration sensor are arranged so as to be shifted in the same direction with respect to the straight line. It has a Pendulum engine mount structure,
The first elastic body is attached above the axis connecting the two mounting points,
The vibration reduction apparatus according to claim 4, wherein the acceleration sensor is arranged to be shifted downward with respect to the straight line. It has a Pendulum engine mount structure,
The first elastic body is attached below a shaft connecting two mounting points,
The vibration reduction device according to claim 4, wherein the acceleration sensor is arranged to be shifted upward with respect to the straight line. The acceleration sensor detects the acceleration in the axial direction of the rod and the acceleration in the rotation direction of the crankshaft,
The vibration reduction device according to claim 2, wherein the actuator and the acceleration sensor are arranged so as to be shifted in opposite directions with the straight line interposed therebetween. A rod housing that connects the first elastic body and the second elastic body;
A rod shaft portion fixed in parallel to the straight line in a gap provided between the first elastic body and the second elastic body of the rod housing and supporting the inertial mass;
The rod shaft portion is arranged at a position deviated from the straight line in a plane including the straight line and orthogonal to the ground,
The vibration reducing apparatus according to claim 7, wherein the acceleration sensor is disposed on a wall surface of the gap.   The vibration reducing apparatus according to any one of claims 1 to 8, wherein a rigid resonance frequency in the axial direction of the rod is lower than a frequency of a main elastic mode of the engine. A first elastic body attached to the engine side;
A second elastic body attached to the vehicle body side;
A rod connecting the first elastic body and the second elastic body;
An inertial mass supported by the rod;
An actuator for reciprocating the inertial mass in the axial direction of the rod;
A sensor configured to detect vibration in the axial direction of the rod;
Control means for controlling the actuator to generate a force proportional to the speed of displacement of the rod based on the detected vibration;
In a vibration reduction device that reduces vibration transmitted from the engine to the vehicle body,
The inertial mass is supported by the rod so that its center of gravity is offset by a predetermined amount from the axis of the rod, and the reciprocating locus of the center of gravity of the inertial mass is at a position offset by the predetermined amount from the axis of the rod. ,
The sensor is configured to be capable of detecting vibration in the axial direction of the rod in a plane including the axis of the rod and a reciprocating locus of the center of gravity of the inertia mass and away from the axis of the rod. Vibration reduction device.

Images