HK1127795B - Damper for automobiles for reducing vibration of automobile body - Google Patents

Description

Vibration damper for vehicle for reducing vibration of vehicle body

Technical Field

The present invention relates to an automotive vibration damping device and a vibration damping control method for performing vibration damping control of an automobile.

The present application claims priority based on Japanese patent application No. 2006-.

Background

In order to improve the riding comfort and comfort of the automobile, a method for preventing vibration of an engine (engine) from being transmitted into a cab is required. Heretofore, a technique of adding a vibration absorbing function to a mount (mount) mechanism for supporting an engine or a technique of reducing vibration of a vehicle body by forcibly exciting the mount with an actuator (for example, refer to japanese patent laid-open nos. 61-220925 and 64-83742) has been proposed.

In the related art, in order to control an actuator, a sensor is used to detect a relative displacement, a relative velocity, and a relative acceleration between a movable portion and a fixed portion of the actuator. However, the sensor itself must be mounted around the engine exposed to a high-temperature environment, and there is a problem of poor reliability.

Further, when a leaf spring or the like is used in order to ensure durability of the movable portion supporting member as the actuator, a resonance system based on the movable portion mass and the leaf spring constant is constituted. However, when the resonance magnification is high in addition to the control of the vibration of the automobile, if the resonance frequency is slightly changed due to a temperature change, an aging change, or the like, the response of the actuator to the command signal is greatly changed, or the like, which causes a problem that the control of the vibration suppression has a bad influence.

Further, there is known a vibration suppression device that detects vibration of an object to be vibration-damped, generates a vibration waveform in which vibration is cancelled due to interference with vibration of the object to be vibration-damped by passing the detection signal through a filter, and applies a signal based on the vibration waveform to an actuator, thereby actively reducing vibration of the object to be vibration-damped over a wide frequency band (see, for example, japanese patent application laid-open No. h 03-219140).

In recent vehicles, control is performed to prevent an increase in fuel consumption by driving the engine in a small number of cylinders (for example, 3 cylinders) in which a cylinder pause of a 6-cylinder engine is performed, as needed. When the engine is driven by the cylinder halt, vibration may be increased as compared with when the 6-cylinder operation is performed. In order to solve such a problem, a vibration suppressing device that actively reduces vibration in a wide band is effective as described in japanese unexamined patent application publication No. 03-219140.

However, since the conventional vibration damping device is only a control for suppressing the generated vibration, in an automobile in which a control for stopping a predetermined number of cylinders is performed in a 6-cylinder engine, all the vibrations are suppressed, and it is difficult to feel that the engine is being driven, so that there is a problem that a driver feels uncomfortable. It is desirable to suppress the vibration so that the driver does not feel the shift from the 6-cylinder drive to the cylinder suspension operation, thereby causing no discomfort.

Further, there is known a vibration control device for a vehicle using an actuator that generates a vibration damping force corresponding to an engine speed by a reaction force for driving a movable portion (see, for example, japanese patent laid-open No. s 61-220925). According to this configuration, since the engine speed can be predicted from the vehicle body vibration and the force applied from the engine to the vehicle body can be cancelled by the actuator, the vibration of the vehicle body can be reduced. Such a damper device uses a linear actuator that reciprocates, and vibrates an auxiliary mass to reduce vibration of a vibration damping target. On the other hand, as a linear actuator, there is known a linear actuator in which a movable element is held at a fixed position by an elastic support unit (leaf spring) and is supported by being elastically deformed (for example, refer to japanese patent application laid-open No. 2004-343964). Since the linear actuator does not generate friction loss and sliding resistance (sliding resistance) in the movable element, the accuracy of the shaft support does not decrease even after a long-term use, high reliability can be obtained, and the improvement of the performance can be realized without loss of power consumption due to the sliding resistance. Further, since the stator (stator) supports the elastic support means at a position farther than the coil while avoiding interference with the coil and using the movable element as a reference point, the coil and the elastic support means, which are bulky, can be disposed in close proximity to each other, and therefore, the linear actuator can be downsized.

Further, in order to optimize vibration damping control, a vibration damping device is known in which a plurality of data maps for vibration damping and phase data are prepared in advance in accordance with the operating state of the vehicle, and a signal for driving an actuator for damping vibration is generated based on amplitude phase data extracted from the data maps in accordance with the operating state (see, for example, japanese unexamined patent application publication No. 11-259147). Further, as a method of reducing vibration by following a change in the state of a vehicle, there are known an adaptive filter realized in the time domain (see, for example, japanese unexamined patent application publication nos. 10-49204 and 2001 51703), and an adaptive filter realized in the frequency domain (see, for example, applicatorAdapting to material edge フイルタTechnical, small, Japanese, sensitivity testLearn 48 volume No. 7P 520). The methods using the adaptive filter perform control by finding an amplitude phase for suppressing vibration by itself based on an error signal (for example, an acceleration signal) at a specific observation point.

However, since the adaptive filter requires a long processing time, the vibration reduction effect is deteriorated when the engine rotation speed greatly varies, and the method implemented in the frequency domain in particular has a problem of requiring a long processing time. Further, if there is a characteristic variation or an aging variation such as a variation in a transfer function from a command value to an actuator to a signal (acceleration) at an observation point, there is a problem that vibration damping characteristics deteriorate. On the other hand, the method of performing control with reference to the map data can shorten the processing time and thus improve the responsiveness, but has a problem that the vibration damping performance is deteriorated due to individual differences or aging variations of the actuators used for control or the engine to be subjected to vibration damping.

When vibration damping control for suppressing vibration of a target device is performed by adding an auxiliary mass (weight) to a linear actuator and using a reaction force generated when the auxiliary mass is vibrated, an amplitude command value and a frequency command value are obtained based on a vibration state value of the target device to be controlled, and a current value applied to the linear actuator is controlled in accordance with the amplitude command value and the frequency command value. By mounting such a vibration damping device to the body of an automobile, a force applied from the engine of the automobile to the body can be cancelled, so that vibration of the body can be reduced.

However, when an external force (external disturbance) close to the natural frequency determined by the auxiliary mass fixed to the movable element and the plate spring holding the movable element acts or a drive command value close to the natural frequency is input, an excessive amplitude is generated due to resonance, a force equal to or greater than a required reaction force for vibration damping is generated, and appropriate vibration suppression control cannot be performed.

Further, since a stopper (stopper) for the movable element is structurally provided to limit the movable range of the movable element, in a case where a change in the trajectory of the automobile is severe due to rapid acceleration of the automobile or bad road running, an excessive amplitude is generated due to an external force acting on the auxiliary mass, and a phenomenon in which the movable element collides with the stopper is caused. Further, in a case where the trajectory of the automobile is changed drastically, the current for driving the linear actuator is also increased in proportion to the change, and the change becomes large, which causes a phenomenon that the movable element collides with the stopper. When the movable element collides with the stopper, the collision sound is generated as abnormal noise. Further, if collision between the movable element and the stopper occurs frequently, there is a problem that the life of the member constituting the linear actuator is likely to be shortened.

Disclosure of Invention

Problems to be solved by the invention

The present invention has been made in view of such circumstances, and an object thereof is to provide a vibration damping device for an automobile and a control method thereof, which can secure high reliability and obtain a good vibration damping effect with a simple structure.

It is another object of the present invention to provide a vibration damping device and a vibration suppression method that can maintain the state of vibration in the same state as before the change even when the state of the excitation source has changed.

It is another object of the present invention to provide a vibration damping device for an automobile and a vibration damping control method that can reduce adverse effects on vibration damping performance due to individual differences or aging changes and can maintain good vibration damping performance even with a large variation in engine speed.

Another object of the present invention is to provide a vibration damping device and a method for controlling the vibration damping device, which can suppress a resonance phenomenon and thereby achieve ideal suppression of vibration by setting the vibration width of the auxiliary mass within an appropriate range.

Further, it is another object to provide a vibration damping device and a method for controlling the vibration damping device that can suppress generation of abnormal noise and the like by limiting the vibration amplitude of the auxiliary mass to an appropriate range.

Means for solving the problems

The present invention is a vibration damping device for an automobile for reducing vibration of an automobile body, comprising: an actuator mounted on the vehicle body to drive the auxiliary mass; a current detector that detects a current flowing through an armature of the actuator; a detection unit that detects a terminal voltage applied to the actuator; an arithmetic circuit that calculates an induced voltage of the actuator based on the current detected by the current detector and the terminal voltage, and calculates at least one of a relative velocity, a relative displacement, and a relative acceleration of the actuator; and a control circuit that performs drive control of the actuator based on at least one of the relative velocity, the relative displacement, or the relative acceleration of the actuator calculated by the arithmetic circuit.

According to the present invention, a current flowing through an armature of an actuator that is mounted on the vehicle body and drives the auxiliary mass is detected, a terminal voltage applied to the actuator is detected, an induced voltage of the actuator is calculated based on the detected current and the terminal voltage, at least one of a relative velocity, a relative displacement, and a relative acceleration of the actuator is calculated, and the actuator is drive-controlled based on the calculated at least one of the relative velocity, the relative displacement, and the relative acceleration of the actuator. Therefore, it is not necessary to use a sensor itself disposed in a high-temperature environment in order to detect the relative displacement, the relative velocity, and the relative acceleration of the movable portion and the fixed portion of the actuator, and high reliability can be ensured. By using displacement information such as relative velocity, relative displacement, or relative acceleration of the actuator, spring effect (spring effect) can be obtained. Further, by performing feedback control using the calculated speed information (relative speed), the resonance characteristic of the actuator can be made gentle, and even if the resonance frequency of the actuator changes, the gain characteristic or the phase characteristic can be made gentle, so that the variation in response to the command signal can be reduced, and the influence on the control performance can be reduced.

The present invention provides a vibration damping device for suppressing unnecessary vibration and generating predetermined vibration as required, comprising: an excitation member that vibrates the auxiliary mass supported by the linear actuator to excite the vibration reduction target; a frequency detection unit that detects a frequency of an excitation source that vibrates the vibration control target; a vibration detecting member that detects vibration at the measuring point; an arithmetic unit that calculates a command value of vibration to be suppressed and a command value of vibration to be generated, based on the frequency of the excitation source and the vibration detected at the measurement point; and a control signal output unit configured to output a control signal obtained by superimposing the command value of the vibration to be suppressed and the command value of the vibration to be generated, to the excitation unit.

According to the present invention, since unnecessary vibration can be controlled and predetermined vibration can be generated as needed, an effect of preventing discomfort due to vibration control can be obtained.

The invention is a vibration damping device for an automobile, characterized by comprising: a vibration exciting part for exciting the vibration assisting mass; a state information acquiring unit that acquires information indicating an operating state of the vehicle; a mapping control unit configured to read an excitation force instruction value corresponding to the operation state information acquired by the state information acquisition unit from a vibration damping information table in which the operation state information and an instruction value for generating an excitation force by the excitation unit are associated with each other, and control the excitation unit based on the excitation force instruction value; a vibration detection unit that detects a vibration state value indicating a vibration state of the vibration control object at an observation point; an adaptive control unit that calculates an excitation force instruction value using an adaptive filter according to the vibration state detected by the vibration detection unit, and controls the excitation unit based on the excitation force instruction value; and a control switching unit that switches to control of the excitation unit by the adaptive control unit when the vibration state value detected by the vibration detection unit exceeds a predetermined value in the control of the excitation unit by the mapping control unit.

According to the present invention, when the damping control performance of the map control deteriorates due to the influence of individual differences or aging changes, the adaptive filter is switched to, and therefore the damping performance can be improved. Further, since the map data of the map control is updated in accordance with the control of the adaptive filter, the vibration damping performance of the map control can be recovered. Further, when the adaptive filter is implemented, the mode is switched to the mode according to the rotation speed variation rate, so that it is possible to select an appropriate adaptive filter and perform vibration damping control when the engine rotation speed varies. Further, the transfer function required in the time-domain adaptive filter is updated using the calculation process of the frequency-domain adaptive filter, so it is possible to prevent the characteristic degradation in the time-domain adaptive filter caused by the variation of the transfer function.

The present invention is a vibration damping device including an actuator for driving an auxiliary mass held by an elastic element to a vibration damping target device, the vibration damping device being configured to suppress vibration of the vibration damping target device by a reaction force generated when the auxiliary mass is driven, the vibration damping device further including: a resonance suppression component of the actuator based on an ideal actuator inverse characteristic using a transfer function of relative vibration velocity to excitation force of a vibration system of the actuator.

According to the present invention, since the resonance suppressing means for the actuator based on the ideal actuator inverse characteristic using the transfer function of the relative vibration speed of the excitation force with respect to the vibration system of the actuator is included, the characteristic of the actuator can be adjusted to an arbitrary characteristic by setting the ideal actuator inverse characteristic based on the predetermined characteristic. Thus, by increasing the damping characteristic of the desired characteristic, the resonance of the movable portion of the actuator can be set to a characteristic in which the resonance is not easily generated by the external force acting on the actuator main body, so that an appropriate reaction force can be generated to achieve a desired vibration suppression. Further, since the apparent natural frequency of the actuator can be reduced by reducing the natural frequency of the desired characteristic, stable vibration damping control can be realized even in the vicinity of the natural frequency of the actual actuator without being affected by the spring characteristic or the like. Further, since the movable range of the movable element of the actuator can be maintained within an appropriate range, the movable element and the stopper do not collide with each other, and generation of collision noise can be suppressed.

The invention is a vibration damping device comprising: an auxiliary mass member supported by the elastic element; an actuator that vibrates the auxiliary mass member; and a control unit that controls a current applied to the actuator so as to suppress vibration of a vibration control target using a reaction force when the auxiliary mass member is vibrated by the actuator, characterized in that the vibration damping device further includes: and an amplitude amount control unit that performs control of limiting a current value applied to the actuator so that the vibration amplitude of the auxiliary mass member does not exceed a predetermined value when the control unit controls the current applied to the actuator based on an amplitude command value and a frequency command value of the vibration to be generated.

According to the present invention, when controlling the current applied to the actuator based on the amplitude command value and the frequency command value of the vibration to be generated, the current value applied to the actuator is limited so that the vibration amplitude of the auxiliary mass member does not exceed a predetermined value, and therefore, an effect is obtained in which the movable element of the actuator can be driven within an appropriate movable range at all times. Thus, the movable element and the stopper do not collide with each other, and generation of collision sound can be suppressed.

Drawings

Fig. 1 is a block diagram showing the structure of an automotive vibration damping device according to embodiment 1 of the present invention.

Fig. 2 is a block diagram showing a configuration of a modification of embodiment 1 shown in fig. 1.

Fig. 3 is a block diagram showing the structure of an automotive vibration damping device according to embodiment 2 of the present invention.

Fig. 4 is a conceptual diagram illustrating a method of detecting an induced electromotive force of a linear actuator in embodiment 2.

Fig. 5 is a conceptual diagram illustrating a method of detecting an induced electromotive force of a linear actuator in embodiment 2.

Fig. 6 is a diagram showing gain characteristics and phase characteristics as an example of response of the linear actuator to a command signal (no induced voltage feedback).

Fig. 7 is a diagram showing gain characteristics and phase characteristics as an example of response of the linear actuator to a command signal (there is induced voltage feedback).

Fig. 8 is a conceptual diagram illustrating a modification of the method of detecting the induced electromotive force of the linear actuator shown in fig. 5.

Fig. 9 is a conceptual diagram illustrating a modification of the method of detecting the induced electromotive force of the linear actuator shown in fig. 5.

Fig. 10 is a block diagram showing the configuration of embodiment 3 of the present invention.

Fig. 11 is a block diagram showing the configuration of embodiment 4 of the present invention.

Fig. 12 is a block diagram showing the configuration of embodiment 5 of the present invention.

Fig. 13 is a state transition diagram showing an operation of the control switching unit 607 shown in fig. 12.

Fig. 14 is a diagram showing the configuration of the mapping control unit 604 shown in fig. 12.

Fig. 15 is a diagram showing the structure of the frequency domain adaptive filter 605 shown in fig. 12.

Fig. 16 is a diagram showing the configuration of the time-domain adaptive filter 606 shown in fig. 12.

Fig. 17 is a block diagram showing the structure of an automotive vibration damping device according to embodiment 6 of the present invention.

Fig. 18 is a block diagram showing a configuration of a modification of the automobile vibration damping device shown in fig. 17.

Fig. 19 is a schematic diagram showing the configuration of the excitation section 30 shown in fig. 17 and 18.

Fig. 20 is a block diagram showing the configuration of embodiment 7 of the present invention.

Fig. 21 is a block diagram showing a modification of the configuration of embodiment 7 shown in fig. 20.

Fig. 22 is a block diagram showing the configuration of embodiment 8 of the present invention.

Fig. 23 is a block diagram showing a modification of the configuration of embodiment 8 shown in fig. 22.

Fig. 24 is a block diagram showing the configuration of embodiment 9 of the present invention.

Fig. 25 is a block diagram showing a modification of the configuration of embodiment 9 shown in fig. 24.

Fig. 26 is a perspective view showing the structure of the linear actuator.

Description of the reference symbols

30.. an excitation unit, 31.. a linear actuator, 32.. an auxiliary mass, 33.. a relative speed sensor, 40.. an engine, 41.. a car body underframe, 43.. a vibration sensor, 50.. a high-level controller, 60.. a stability controller, 70.. a power supply circuit

Detailed Description

Hereinafter, preferred embodiments of the present invention will be described with reference to the accompanying drawings. However, the present invention is not limited to the following embodiments, and for example, the components of the embodiments may be appropriately combined.

[ embodiment 1 ]

Fig. 1 is a block diagram showing the structure of a vibration damping device according to embodiment 1 of the present invention. In the present embodiment, a case where the vibration damping device is applied to an automobile will be described as an example. In fig. 1, reference numeral 31 denotes a linear actuator that reciprocates an auxiliary mass 32, and the auxiliary mass 32 reciprocates in the same direction as the vibration direction to be suppressed. Reference numeral 33 is a relative velocity sensor that detects the relative velocity of the linear actuator 31 and the auxiliary mass 32. Reference numeral 40 denotes an engine of an automobile, and is fixed to a vehicle body frame 41. Reference numeral 42 is a wheel of an automobile. Reference numeral 43 denotes a vibration sensor (acceleration sensor) provided at a predetermined position of a seat (passenger seat)44 or the vehicle body under frame 41. Reference numeral 50 denotes a high-level controller that receives engine rotation information such as an ignition pulse, an accelerator opening (accelerator opening), and a fuel injection amount and an output of the vibration sensor 43 from a control device (not shown) provided in the engine 40 and outputs a drive command to the linear actuator 31 for damping vibration. The higher-level controller 50 generates and outputs a command signal (drive command) for controlling vibration of the vehicle body under frame 41 generated by rotation of the engine. Reference numeral 60 denotes a stabilization controller which inputs the relative speed signal output from the relative speed sensor 33 and the drive command output from the high-level controller 50, and thereby stably drives the linear actuator 31. Reference numeral 70 is a power supply circuit that outputs a drive current for the linear actuator 31 based on the stabilization drive command output from the stabilization controller 60.

The vibration damping device shown in fig. 1 suppresses vibration generated at a predetermined position of the vehicle body underframe 41 or the vehicle body by rotation of the engine 40, by using a reaction force when the auxiliary mass 32 attached to the linear actuator 31 is reciprocated. At this time, the stabilization controller 60 receives a relative speed signal between the main body of the linear actuator 31 fixed to the vehicle body 41 and the auxiliary mass 32 that reciprocates, and feeds back a drive command to generate a damping force in the linear actuator 31, thereby reducing sensitivity to disturbance vibration received by the vehicle body underframe 41 due to irregularities of the road surface. This can reduce the influence of disturbance vibration.

The relative speed sensor 33 may detect the relative speed by differentiating an output of a displacement sensor that detects the stroke (stroke) of the linear actuator 31. Further, the relative velocity sensor 33 may detect the relative velocity from a difference between integrated values of acceleration sensors provided in the linear actuator 31 and the auxiliary mass 32, respectively.

Here, a mechanism of a linear actuator (reciprocating motor) used in the present invention will be described with reference to fig. 26. As shown in fig. 26, the linear actuator includes a movable part 1, a fixed part 2 disposed around the movable part 1, and two sets of support members (elastic support means) 3 in which two leaf springs or a plurality of leaf springs are overlapped, and supports the movable part 1 by being elastically deformed by itself so that the movable part 1 can reciprocate with respect to the fixed part 2.

The movable part 1 is formed in a cylindrical shape having a female screw (female screw) unit 11a formed at the tip thereof, and includes a shaft (draft) 11 that reciprocates in the axial direction, and a movable element 12 as a movable magnetic pole that is fixed to the shaft 11 at a position halfway in the axial direction by inserting the shaft 11 into the shaft. A nut (nut)13 for fixing the shaft 11 to an object (not shown) to be driven is screwed to the female screw unit 11 a.

The fixing portion 2 includes: a yoke (yoke)21 having a rectangular outer shape when viewed in the axial direction of the shaft 11 and having an inner side hollowed out; and a pair of coils 22, 23 configured to sandwich the movable part 1 therebetween and fixed to the inside of the yoke 21. A bobbin 26 is attached to a magnetic pole unit 21a formed in the yoke 21 so as to project inward, and a wire 27 is wound multiple times around the bobbin 26 to form a coil 22. A bobbin 26 is similarly attached to the magnetic pole unit 21b formed at a position opposed to the magnetic pole unit 21a with the fixed portion 2 interposed therebetween, and a wire 27 is wound multiple times around the bobbin 26 to form the coil 23.

On the front end surface of the magnetic pole unit 21a facing the movable part 1, permanent magnets 24, 25 are aligned and fixed in the axial direction of the shaft 11. Permanent magnets 24, 25 are also aligned and fixed in the axial direction of the shaft 11 on the front end surface of the magnetic pole unit 21b facing the movable part 1. These permanent magnets 24 and 25 are made of, for example, a rare-earth magnet formed in a tile shape having the same diameter and the same length coaxially, and are arranged adjacent to each other in the axial direction. Here, the permanent magnets 24 and 25 are magnets having different radial directions in which magnetic poles are arranged in a direction orthogonal to the axial direction, and the arrangement of the magnetic poles is reversed. Specifically, the permanent magnet 24 has its N-pole disposed on the outer diameter side and S-pole disposed on the inner diameter side, and the other permanent magnet 25 has its N-pole disposed on the inner diameter side and S-pole disposed on the outer diameter side.

The two leaf springs 3 are disposed so as to be spaced apart in the axial direction of the shaft 11 with the yoke 21 interposed therebetween. The two plate springs 3 have the same shape, are formed by punching a metal plate having a uniform thickness, and form an "8" shape when viewed from the axial direction of the shaft 11. Through holes 3a are formed in the front end or the tip end of the support shaft 11 at positions corresponding to the portions where the center lines of "8" intersect. In addition, through holes 3b and 3c having a size enough to allow the coil 22 or 23 to pass through are formed in the inner side of the coil corresponding to "8". Further, small holes 3d and 3e for fixing the plate spring 3 to the yoke 21 are formed at the uppermost portion and the lowermost portion corresponding to "8", respectively.

Each leaf spring 3 supports the shaft 11 at a position halfway in the axial direction of the coil 22. More specifically, the leaf spring 3 supporting one side of the distal end of the shaft 11 is fixed to the yoke 21 by a screw passing through the small hole 3d and a screw passing through the small hole 3e at a position farther from the center of the shaft 11 than the coils 22 or 23, with the distal end side of the shaft 11 being fixed through the through hole 3 a. The other plate spring 3 supporting the end of the shaft 11 is fixed to the yoke 21 by a screw passing through the small holes 3d and 3e at a position farther from the center of the shaft 11 than the coils 22 and 23, while the end side of the shaft 11 is fixed through the through hole 3 a.

The plate spring 3 on the one hand projects the coil 22 from the through hole 3b toward the distal end side of the shaft 11 and projects the coil 23 from the through hole 3c toward the distal end side of the shaft 11, and the plate spring 3 on the other hand projects the coil 22 from the through hole 3b toward the distal end side of the shaft 11 and projects the coil 23 from the through hole 3c toward the distal end side of the shaft 11. The interval between the two plate springs 3 in the axial direction of the shaft 11 is narrower than the dimension of the coil 22 or 23 in the same direction, and the through-holes 3b and 3c function as "clearance (clearance)" for avoiding interference with the coil 23.

Each leaf spring 3 holds the movable portion 1 at two positions on the tip end side and the distal end side of the shaft 11, and elastically deforms itself to support the movable portion 1 so as to be movable in the axial direction of the shaft 11, instead of sliding the movable element so as to be supported so as to be movable in a reciprocating manner as in the conventional art. In addition, each leaf spring 3 is adjusted in advance such that the distance (not a linear distance but the length of the leaf spring itself) from the through hole 3a of the support shaft 11 to the small hole 3d or 3e is as long as possible or the thickness of the plate is made thin, so that the amount of deformation when the movable part 1 reciprocates is smaller than the amount of deformation that can be broken due to fatigue caused by forcibly repeating elastic deformation. However, when the entire linear actuator is viewed from the axial direction of the shaft 11, the outer shape of the plate spring 3 is not so large as to protrude from the outer shape of the yoke 21.

The operation method of the linear actuator configured as above will be described below. When an alternating current (sine wave current or rectangular wave current) flows through the coils 22 and 23, magnetic flux is guided from the S pole to the N pole by the permanent magnet 24 in a state where a current in a predetermined direction flows through the coils 22 and 23, thereby forming a magnetic flux loop (magnetic flux loop) that circulates in the order of the outer peripheral portion of the yoke 21, the magnetic pole unit 21a, the permanent magnet 24, the movable element 12, the shaft 11, and the outer peripheral portion of the yoke 21. As a result, a force acts on the movable portion 1 in the axial direction from the distal end of the shaft 11 to the distal end, and the movable portion 1 is pushed by the force and moves in the same direction. On the other hand, in a state where a current in the opposite direction to the predetermined direction flows through the coils 22 and 23, the magnetic flux is guided from the S pole to the N pole by the permanent magnet 25, thereby forming a magnetic flux loop circulating through the outer peripheral portion of the yoke 21, the magnetic pole unit 21a, the permanent magnet 25, the movable element 12, the shaft 11, and the outer peripheral portion of the yoke 21 in this order. As a result, a force acts on the movable portion 1 in the axial direction from the tip end to the tip end of the shaft 11, and the movable portion 1 is pushed by the force and moves in the same direction. The alternating current causes the flowing direction of the current in the coils 22 and 23 to change alternately, and the movable portion 1 repeats the above operation and reciprocates in the axial direction of the shaft 11 with respect to the fixed portion 2.

In the linear actuator described above, each leaf spring 3 holds the movable portion 1 at two points on the tip end side and the distal end side of the shaft 11, and elastically deforms itself to support the movable portion 1 so as to be movable back and forth in the axial direction of the shaft 11, instead of sliding the movable portion and being supported so as to be movable back and forth. Thus, no friction and no sliding resistance occur in the movable member 1. Therefore, even after a long time use, the accuracy of the shaft support is not lowered, and high reliability can be obtained. Further, the performance can be improved without loss of power consumption due to the sliding resistance. Further, in the above-described linear actuator, the fixed portion 2 supports the respective leaf springs 3 at a position farther than the coils with the movable element as a reference point while avoiding interference with the coils 22, 23. The bulky coils 22 and 23 can be disposed adjacent to the two leaf springs 3. Therefore, miniaturization of the linear actuator can be achieved.

Next, a modified example of the vibration damping device shown in fig. 1 will be described with reference to fig. 2. The apparatus shown in fig. 2 is different from the apparatus shown in fig. 1 in that a current detector 51 that detects a drive current is included instead of the relative speed sensor 33, and the apparatus is stabilized based on the current detected by the current detector 51. The stabilization controller 61 estimates an induced voltage generated by the linear actuator 31 from a coil current (drive current) of the linear actuator 31, a voltage command output from the power supply circuit 71, a terminal voltage, or the like, and estimates a relative speed of the linear actuator 31 and the auxiliary mass 32 based on this. By feeding back the estimated value, the linear actuator 31 generates a damping force. Thereby, the influence of the external disturbance vibration can be reduced.

The terminal voltage may be a signal obtained by multiplying a voltage command by a voltage amplifier gain in a voltage amplifier included in the power supply circuit 71.

Fig. 6 (a) and (b) are graphs showing gain characteristics and phase characteristics as examples of responses (no induced voltage feedback) of the linear actuator having the elastic element for supporting the auxiliary mass to the command signal. Fig. 7 (a) and (b) are graphs showing gain characteristics and phase characteristics as examples of responses of the linear actuator to the command signal (induced voltage feedback or relative velocity feedback).

When there is no induced voltage feedback, the change in the gain characteristic shown in fig. 6 (a) and the change in the phase characteristic shown in fig. 6 (b) are rapid with respect to the change in the resonance frequency of the actuator. In contrast, in the automotive vibration damping device according to the present embodiment, by performing induced voltage feedback (feedback control using speed information), even if the resonant frequency of the actuator changes, the gain characteristic shown in fig. 7 (a) and the phase characteristic shown in fig. 7 (b) are both gentle, and thus it is understood that the variation in response to the command signal is small and the influence on the control performance is small.

Further, feedback control is performed using the velocity information obtained from the calculation, so that the resonance characteristic of the actuator becomes gentle. Therefore, even if the resonant frequency of the actuator changes, the gain characteristic and the phase characteristic are both gentle, so that the variation in response to the command signal can be reduced, and the influence on the control performance can be reduced.

[ 2 nd embodiment ]

Next, the structure of the vibration damping device according to embodiment 2 of the present invention will be described. Fig. 3 is a block diagram showing the structure of the vibration damping device according to embodiment 2. In fig. 3, it is assumed that the vibration damping device is connected to a body underframe (main system mass) 41 of an automobile to be controlled, and controls vibration (vibration damping) in the vertical direction (gravity direction) generated in the body underframe 41.

The vibration damping device of the present embodiment is a so-called active dynamic vibration damper (active dynamic vibration absorber), including: a current detector 63 that detects a drive current to the linear actuator 31; a terminal voltage detector 64 that detects a terminal voltage of the linear actuator 31; and a linear actuator 31 driven based on the detection results of the current detector 63 and the terminal voltage detector 64. The vibration damping device drives the auxiliary mass 32 in the vertical direction (vibration direction to be damped) using the driving force of the linear actuator 31, and supplies the inertial force of the auxiliary mass including the auxiliary mass 32 to the primary system mass 41 as a reaction force, thereby suppressing the vibration of the primary system mass 41.

The current detector 63 shown in fig. 3 detects the current supplied to the linear actuator 31 and supplies it to the controller 62. Further, the terminal voltage detector 64 detects the terminal voltage applied to the linear actuator 31, and supplies it to the controller 62. When the linear actuator 31 is driven, the linear actuator 31 generates an induced electromotive force proportional to the velocity. By calculating the induced electromotive force, a velocity signal can be obtained. In addition, a vibration displacement signal can be obtained by integrating the induced electromotive force, and a vibration acceleration can be obtained by differentiating the induced electromotive force.

For example, as shown in fig. 4 and 5, the terminal voltage V and the current i are detected, and output as the induced voltage E through an amplifier circuit and a differential circuit. In this case, gains K1 and K2 corresponding to the winding (wire winding) resistance R and the winding inductance L need to be set. The setting is adjusted so that a current of a predetermined frequency flows in a state where the movable portion (movable element, auxiliary mass) of the linear actuator is restricted, and the output becomes zero. Since the relationship of E ═ V-R · i-L (di/dt) holds in the induced voltage E, the induced voltage E can be obtained by detecting the terminal voltage V and the current i.

In the case where a spring constant close to an optimum value with respect to the dynamic damper is obtained by a magnetic spring characteristic or a mechanical spring element, a high vibration damping effect can be obtained without providing energy for damping by adjusting the damping force generated by the linear actuator 31. The damping force can be adjusted by connecting a load resistor to both ends of the coil of the linear actuator 31 and changing the magnitude of the load resistor.

The controller 62 calculates the relative velocity, or relative displacement, or relative acceleration of the linear actuator 31 based on the induced voltage calculated from the current detected by the current detector 63 and the terminal voltage detector 64 and the terminal voltage, derives the optimum driving amount (control amount) of the linear actuator 31 so that the damping device obtains the optimum spring characteristic and damping characteristic for damping the main system mass 41, and outputs the derived result to the power amplifier 72 as a command signal. In addition, power is supplied to the power amplifier 72 through a power supply circuit 90. The power amplifier 72 drives the linear actuator 31 in accordance with the instruction signal of the controller 62, and the linear actuator 31 drives (vibrates) the auxiliary mass 32 in the up-down (gravity) direction, thereby damping the main system mass 41.

According to the above-described embodiment, it is not necessary to use a sensor that detects the relative displacement, the relative velocity, or the relative acceleration between the movable portion and the fixed portion of the actuator, but the relative velocity, or the relative displacement, or the relative acceleration of the linear actuator 31 is calculated based on the induced voltage calculated from the current of the linear actuator 31 and the terminal voltage, and the linear actuator 31 is controlled based on the relative velocity, or the relative displacement, or the relative acceleration, so that high reliability can be ensured.

Further, by using the displacement information obtained from the calculation, the elastic effect can be obtained. Further, by using the velocity information obtained from the calculation, a damping effect (damping effect) can be obtained.

The terminal voltage may be determined based on a command value of a voltage to be applied to the actuator.

Next, a modification of the method for detecting induced electromotive force shown in fig. 5 will be described with reference to fig. 8. The method of detecting induced electromotive force shown in fig. 8 is different from the method of detecting induced electromotive force shown in fig. 5 in that a band-pass filter (BPF) is provided in order to limit a control band for controlling resonance to the vicinity of the resonance frequency of the linear actuator 31 by feedback of the velocity estimation value. The band-pass filter is a filter for obtaining an attenuation effect only in the vicinity of the resonance frequency of the linear actuator 31 (a frequency close to the natural frequency), and the phase of the band-pass filter is set to 0 ° in the vicinity of the resonance frequency of the actuator.

By providing the band-pass filter, the phase can be adjusted while suppressing the noise component of the direct current, and therefore the accuracy of the estimated value of the induced voltage can be improved.

Next, a modification of the method for detecting induced electromotive force shown in fig. 8 will be described with reference to fig. 9. The method of detecting induced electromotive force shown in fig. 7 is different from the method of detecting induced electromotive force shown in fig. 8 in that two Low Pass Filters (LPF) are provided to suppress high frequency noise components. The cutoff frequency of the low-pass filter is set to a frequency higher than the resonance frequency (natural vibration frequency) of the linear actuator 31.

By providing the low-pass filter, the noise component of high frequency can be removed, and therefore, the generation of abnormal noise due to the noise component can be suppressed.

The vibration damper for an automobile according to the present invention is effective when it is mounted to an automobile component such as an under frame, an engine mount, a radiator, a lower part of a rear trunk, or a lower part of a trunk of an automobile.

The linear actuator 31 is an actuator utilizing electromagnetic force, and is effective when a reciprocating motor is used, for example.

As described above, in the actuator in which the movable portion is supported by the elastic element, even if an external excitation force due to irregularities of a road surface acts on the actuator during traveling of the automobile, generation of the external excitation force or excessive displacement due to a resonance phenomenon can be suppressed, so that generation of abnormal noise due to collision of the movable portion of the actuator with the limiter or the like can be prevented. Further, since the resonance phenomenon can be detected on the drive circuit side, it is not necessary to provide a sensor or the like in the actuator body, and the actuator body can be miniaturized. Further, even when an error occurs in the constant of the coil due to individual differences of the actuator, aging changes, or the like, it is possible to prevent the generation of a dc component of the current unnecessary for the vibration damping control. Further, since a noise component of a high frequency or the like is not amplified, a generation level of noise or abnormal noise can be reduced. Further, since the vibration velocity feedback is provided with a band-pass filter or a low-pass filter and is independent of the current feedback circuit, it is possible to prevent the influence on the responsiveness of the actuator to a high-frequency drive command.

[ embodiment 3 ]

A vibration damping device according to embodiment 3 of the present invention will be described below with reference to the drawings. Fig. 10 is a block diagram showing the configuration of this embodiment. Here, description will be given assuming that an engine for controlling the number of cylinders in an automobile is an excitation source. In the figure, reference numeral 41 denotes a body underframe of an automobile. Reference numeral 40 denotes an engine capable of controlling the number of cylinders in accordance with the operating state, and the engine 40 serves as a vibration generating source (excitation source). Reference numeral 44 denotes a driver's seat (hereinafter simply referred to as a seat), and the seat 44 is a vibration measurement point. Reference numeral 43 denotes an acceleration sensor attached to the seat 44, and detects the acceleration of the seat 44. Reference numeral 31 denotes a linear actuator (reciprocating motor) attached to a vehicle body frame 41, and suppresses vibration by vibrating the auxiliary mass 32 in order to suppress vibration generated by the engine 40. Reference numeral 52 is a control unit that controls the driving of the linear actuator 31 based on the vibration generated in the excitation source and the vibration detected in the measurement point.

Reference numeral 510 is a pulse IF (interface) to which an ignition pulse supplied to the engine 40 is input. Reference numeral 520 denotes a sensor IF (interface) which inputs the output of the acceleration sensor 43. Reference numeral 530 is a frequency detection unit that detects the frequency of the input ignition pulse. Reference numeral 540 denotes an FFT unit that performs FFT (Fast Fourier Transform), extracts how many frequency components are included in the output signal of the acceleration sensor 43, and outputs a phase/amplitude FB (feedback) signal of the 1-time vibration mode and a phase/amplitude FB signal of the 2-time vibration mode. Reference numeral 550 denotes a 1-time instruction ROM in which instruction values for generating vibrations in the 1-time vibration mode are stored in advance, and the instruction values corresponding to the frequencies detected by the frequency detection unit 530 are read out and output. Reference numeral 560 denotes a 2-time instruction ROM in which instruction values for generating vibrations in the 2-time vibration mode are stored in advance, and the instruction values corresponding to the frequencies detected by the frequency detection unit 530 are read out and output.

Reference numeral 570 denotes a 1-order frequency calculating unit which receives the 1-order vibration command value read from the 1-order command ROM550, and the 1-order vibration amplitude FB value and the 1-order vibration phase FB value output from the FFT unit 540, and calculates and outputs the 1-order vibration command value and the 1-order phase command value of the vibration to be excited. Reference numeral 580 denotes a sine wave generator which receives the 1 st order vibration command value output from the 1 st order frequency operation unit 570, the 1 st order phase command value, and the vibration frequency value output from the frequency detection unit 530, and outputs the 1 st order current command value. Reference numeral 590 denotes a 2-order frequency calculation unit which receives the 2-order vibration command value read from the 2-order command ROM560, and the 2-order vibration amplitude FB value and the 2-order vibration phase FB value output from the FFT unit 540, and calculates and outputs the 2-order vibration command value and the 2-order phase command value of the vibration to be excited. Reference numeral 600 is a sine wave generator that inputs the 2-time vibration command value output from the 2-time frequency operation unit 590, the 2-time phase command value, and the vibration frequency value output from the frequency detection unit 530, and outputs the 2-time current command value. Reference numeral 610 denotes a 0-time command output unit that outputs a 0-time current command value. Reference numeral 53 denotes a current amplifier that outputs the motor current flowing through the linear actuator 31 based on a superimposed current command value obtained by superimposing the current command value for 0, the current command value for 1, and the current command value for 2.

Next, with reference to fig. 10, an operation in which, among vibrations generated in the vehicle body underframe 41, only vibrations to be suppressed are suppressed and simultaneously vibrations to be superimposed and applied are generated will be described. In an automobile equipped with a 6-cylinder engine, when cylinder suspension control from 6 cylinders to 3 cylinders is performed, the 1 st-order frequency calculation means 570 calculates and outputs a command value for suppressing vibration newly generated by the 3-cylinder drive (vibration generated when the 3 cylinders are driven). On the other hand, the 2-order frequency calculating means 590 obtains and outputs a command value for regenerating the vibration that does not occur due to the 3-cylinder drive (vibration occurring during the 6-cylinder drive). When a superimposed current command value obtained by superimposing a command value for suppressing the vibration generated during the 3-cylinder drive and a command value for reproducing the vibration generated during the 6-cylinder drive is output to the current amplifier 53, the auxiliary mass 32 vibrates in the linear actuator 31 in order to suppress unnecessary vibration and generate vibration to be reproduced. Thus, even if the cylinder pause control from 6 cylinders to 3 cylinders is performed, the vibration generated during the driving of the 6 cylinders is continued, and the driver does not feel uncomfortable.

[ 4 th embodiment ]

Embodiment 4 is explained with reference to fig. 11. The vibration damping device shown in fig. 11 is different from the vibration damping device shown in fig. 10 in that a plurality of measurement points for detecting vibration are provided, and a plurality of linear actuators for performing excitation are provided. Then, the control unit 52 obtains command values of vibration to be suppressed and vibration to be generated at the measurement points 441, 442, 443, and 44n, and outputs the command values to the respective linear actuators 31, 301, 302, and 30 n. Thus, since the vibration to be suppressed can be reduced and the vibration to be emphasized can be increased, for example, the vibration of the engine can be suppressed and the bass vibration generated from the audio speaker during music reproduction can be increased based on the music signal during reproduction. Further, by suppressing the vibration of the engine, the sound of a muffler (sound) in the vehicle and the like can be reduced.

In this way, in order to suppress unnecessary vibrations and generate predetermined vibrations as needed, in a vibration damping device including a member for exciting a vehicle body underframe 41 by vibrating an auxiliary mass 32 supported by a linear actuator 31, the frequency of an engine 40 for vibrating the vehicle body underframe 41 and the vibrations in a seat 44 are detected, command values of the vibrations to be suppressed and the vibrations to be generated are obtained based on the frequency of the engine 40 and the vibrations in the seat 44, and a control signal in which these command values are superimposed is output to the linear actuator 31.

[ 5 th embodiment ]

A vibration damping device according to embodiment 5 of the present invention will be described below with reference to the drawings. Fig. 12 is a block diagram showing the configuration of this embodiment. In the figure, reference numeral 41 denotes a vehicle body underframe to which an engine such as an internal combustion engine is mounted, and a vibration system of the vehicle body is formed by rotational driving of the engine. Reference numeral 30 denotes an excitation unit that vibrates the auxiliary mass by a linear actuator to control vibration generated in the vehicle body of the vehicle body under frame 41. The excitation unit 30 may use a linear actuator such as a voice coil (voice coil) motor or a reciprocating motor. Reference numeral 70 denotes a power supply circuit which drives the excitation section 30. Reference numeral 604 denotes a map control unit that performs vibration damping control by referring to internal map data. Reference numeral 605 denotes a frequency-domain adaptive filter unit that performs damping control by the frequency-domain adaptive filter. Reference numeral 606 denotes a time-domain adaptive filter unit that performs damping control by the time-domain adaptive filter.

Frequency domain adaptive filter section 605 and time domain adaptive filter section 606 update the mapping data held in mapping control section 604 based on the result of the adaptive filter. Reference numeral 607 denotes a control switching means for selecting any one of the mapping control means 604, the frequency domain adaptive filter means 605 and the time domain adaptive filter means 606 in accordance with the state at the present time to perform vibration damping control, and switches the control to be applied based on the acceleration at the observation point of the predetermined position of the vehicle body under frame 41, the acceleration reference value, and the engine rotation speed change rate. Further, frequency-domain adaptive filter section 605 transmits transfer function G' (S) to time-domain adaptive filter section 606, thereby updating. The reciprocal of (S (n) -S (n-1))/(M (n) -M (n-1)) calculated in the frequency domain adaptive filter unit 605 corresponds to G' (S). Reference numeral 608 denotes an acceleration reference value table in which acceleration reference values corresponding to the rotational speeds are stored in advance for each operation state. Reference numeral 609 denotes a rotation speed change rate measuring unit that measures a change rate of the engine rotation speed from an engine pulse signal, which is updated for each engine pulse signal by calculating the engine rotation speed N and the rotation speed change rate dN/dt based on the generated time interval. The vehicle body under frame 41 includes an acceleration sensor that detects an acceleration a at an observation point and outputs an observation point acceleration signal, and a function (not shown) that outputs an operation state signal D0 indicating an operation state (gear position, air conditioner ON/OFF, accelerator opening degree) at the current time.

Here, the operation of each control unit will be described with reference to fig. 14 to 16. The control operation shown in fig. 14 to 16 is basically control based on the conventional technique, and therefore, a detailed description thereof is omitted here.

Fig. 14 is a diagram showing an operation of mapping control section 604 shown in fig. 12. Map control section 604 selects a control signal data map based on the operating state signal and the engine speed obtained from the engine pulse signal, reads out an amplitude command value and a phase command value defined in advance in the data map, and outputs the read out amplitude command value and phase command value to power supply circuit 70. The power supply circuit 70 controls the vibration of the excitation unit based on the command value, thereby reducing the vibration of the vibration control object (vehicle body).

Fig. 15 is a diagram showing an operation of time-domain adaptive filter section 606 shown in fig. 12. Time domain adaptive filter section 606 receives the observation point acceleration signal and the engine pulse signal as input, finds a sine wave excitation force command value (amplitude command value and phase command value) based on estimated transfer function g(s) of the signal transfer characteristic, and outputs the sine wave excitation force command value to power supply circuit 70. The power supply circuit 70 controls the vibration of the excitation unit based on the command value, thereby reducing the vibration of the vibration control object (vehicle body).

Fig. 16 is a diagram showing an operation of frequency domain adaptive filter section 605 shown in fig. 12. Frequency domain adaptive filter section 605 receives the observation point acceleration signal and the engine pulse signal as input, obtains a force command of a force to be generated based on the frequency component of the vibration damping target obtained by fourier transform, obtains a sinusoidal excitation force command value (amplitude command value and phase command value) obtained based on the force command, and outputs the sinusoidal excitation force command value to power supply circuit 70. The power supply circuit 70 controls the vibration of the excitation unit based on the command value, thereby reducing the vibration of the vibration control object (vehicle body).

Next, the operation of the damper device shown in fig. 12 will be described. First, when the automobile starts the engine, the switching unit 607 selects the mapping control unit 604. Thereby, mapping control is performed. In this state, the acceleration signal of the vehicle body under frame 41 at the observation point is compared with the acceleration reference value stored in the acceleration reference value table 608, and when the detected acceleration exceeds the acceleration reference value, the switching from the mapping control to the adaptive filter is performed. When switching from the mapping control to the adaptive filter is performed, control switching section 607 refers to the output of rotation speed change rate measuring section 609, and switches to time-domain adaptive filter section 606 when the change rate is large, and switches to frequency-domain adaptive filter section 605 when the change rate is small. When frequency-domain adaptive filter section 605 is operating, an estimated transfer function (S (n))/(M (n)) -M (n-1)) of signal transfer characteristics indispensable to time-domain adaptive filter section 606 is obtained by an adaptive filter calculation process. Since this estimated transfer function corresponds to 1/G '(S), estimated transfer function G' (S) of time-domain adaptive filter section 606 is updated based on the result.

Then, the control switching unit 607 switches from the adaptive filter to the map control at a time point when the acceleration reference value is lower than the acceleration reference value after the transition to the adaptive filter. At this time, since a sinusoidal excitation force command that can effectively reduce vibration is obtained by the operation of the adaptive filter in frequency domain adaptive filter section 605 or time domain adaptive filter section 606, frequency domain adaptive filter section 605 or time domain adaptive filter section 606 updates the mapping data held in mapping control section 604 based on the obtained excitation force command value. By this operation, since the map data is updated to the map data that is most suitable at the present time, it is possible to prevent the vibration damping performance from deteriorating due to the influence of individual differences or aging changes, and to maintain the execution of the appropriate vibration damping control process.

Next, the timing at which the control switching unit 607 switches each control will be described with reference to fig. 13. Fig. 13 is a diagram showing an operation of switching the control method based on the state value. In fig. 13, the reference value obtained by referring to the acceleration reference value table 608 is represented as a1(N, D0) or a2(N, D0) based on the engine speed N and the operating state value D0. A1(N, D0) is a reference value for transferring from mapping control to adaptive filter, A2(N, D0) is a reference value for transferring from adaptive filter to mapping control, and the relationship of A2(N, D0) ≦ A1(N, D0) is satisfied. Reference values for performing mode transition of the adaptive filter are represented as W1 to W4. W1 is a reference value for shifting from the frequency domain adaptive filter to the time domain adaptive filter based on the engine speed change rate dN/dt. W2 is a reference value for shifting from the time-domain adaptive filter to no control based on the engine speed change rate dN/dt. W3 is a reference value for shifting from the time-domain adaptive filter to the frequency-domain adaptive filter based on the engine speed change rate dN/dt. W4 is a reference value for shifting from no control to the time-domain adaptive filter based on the engine speed change rate dN/dt. The reference values W1-W4 satisfy the conditions that W1 is more than W2, W3 is more than W4, W1 is more than or equal to W3, and W2 is more than or equal to W4. Note that the mapping control state (initial state) is represented by C0 being 1, the control state of the frequency-domain adaptive filter is represented by C0 being 2, the control state of the time-domain adaptive filter is represented by C0 being 3, and the state without adaptive filter control is represented by C0 being 4. As shown in fig. 13, the optimal vibration damping control can be performed by selecting the control method most suitable for the current time based on the reference values a1 and a2 obtained from the engine speed N and the speed change rate dN/dt and the reference values W1 to W4 for adaptive filter switching.

The reference values a1 and a2 may be the same value, but may be set to satisfy a2< a1, and after the acceleration reference value is exceeded and the adaptive filter control is shifted to, hysteresis (hystersis) may be set when returning to the map control, and the reference value may be returned to the extent of the hysteresis below the acceleration reference value. This can further improve the vibration damping performance and can update the data to be better than the map data. Further, the map control may be returned to the map control with the acceleration reference value being lower, and the map control may be returned to the map control with the rotational speed change rate being higher than a predetermined value. This makes it possible to remain as long as possible in the adaptive filter control, and update the mapping data while improving the vibration damping performance.

In addition, when the map data is updated, only the data of the engine rotation speed at which the map control is shifted to the adaptive control may be updated, or all the data of the rotation speed lower than the acceleration reference value in the adaptive filter control may be updated. The timing of the map data update may be performed when the acceleration reference value is lower than a predetermined value in the adaptive filter control, in addition to the timing of returning to the map control. The update timing of the transfer function may be performed at every fixed time interval, or may be performed at a rotation speed that is a predetermined interval from the rotation speed updated last time. In addition, when the rotation speed variation rate is too large when the adaptive filter is implemented, the adaptive filter operation may be suspended (no control) in order to avoid adverse effects.

In this way, when the damping control performance of the map control is deteriorated due to individual differences, aging changes, or the like, the adaptive filter is switched to improve the damping performance, and the map data of the map control is updated, so that the damping performance by the map control can be restored. Further, when the adaptive filter is implemented, the frequency domain, the time domain, or the non-control is switched depending on the rotation speed variation rate and the responsiveness, so that it is possible to prevent the vibration from being increased by the adaptive filter when the rotation speed varies. Further, the transfer function required in the time-domain adaptive filter is updated using the calculation process of the frequency-domain adaptive filter, so it is possible to prevent the characteristic degradation in the time-domain adaptive filter caused by the variation of the transfer function.

[ 6 th embodiment ]

Next, a vibration damping device according to embodiment 6 of the present invention will be described with reference to the drawings. Fig. 17 is a block diagram showing the structure of the vibration damping device according to this embodiment. In fig. 17, reference numeral 30 denotes an excitation means for vibrating an auxiliary mass (weight) and suppressing vibration of a vibration damping target device such as an automobile by a reaction force thereof. Reference numeral 41 denotes a vehicle body underframe of the automobile to which the excitation unit 30 is attached. The excitation section 30 controls vibration in the vertical direction (gravity direction) generated in a vehicle body under frame (main system mass) 41. Reference numeral 72 denotes a power amplifier that supplies a current for driving the linear actuator included in the excitation unit 30. Reference numeral 65 denotes a current control unit which controls the current supplied to the linear actuator in accordance with the force to be generated in the excitation unit 30. Reference numeral 631 denotes a current detection unit that detects a current supplied to the excitation unit 30. Reference numeral 641 denotes an applied voltage detection unit that detects a voltage applied to the power amplifier 72. Reference numeral 66 denotes an actuator vibration velocity estimation unit that estimates the vibration velocity of the linear actuator included in the excitation unit 30 based on the outputs (current value and voltage value) of the current detection unit 631 and the applied voltage detection unit 641. Reference numeral 67 is an ideal actuator inverse characteristic unit that inputs the value of the vibration velocity output from the actuator vibration velocity estimation unit 66, and outputs a force command signal that the ideal actuator should output based on the ideal actuator inverse characteristic. Reference numeral 68 denotes a predetermined value output means for outputting a predetermined force command value.

Here, the detailed configuration of the excitation section 30 shown in fig. 17 will be described with reference to fig. 19. Fig. 19 is a diagram showing a detailed configuration of the excitation unit 30 shown in fig. 17. In the figure, reference numeral 32 denotes an auxiliary mass (weight) added to the vehicle body under frame 41. Reference numeral 34 denotes a stator constituting a linear actuator (reciprocating motor), and is fixed to the vehicle body frame 41. Reference numeral 12 denotes a movable element constituting a linear actuator (reciprocating motor), and the movable element performs, for example, reciprocating movement in the direction of gravity (vertical movement on the paper surface of fig. 3). The excitation unit 30 is fixed to the vehicle body frame 41 such that the direction of vibration to be suppressed by the vehicle body frame 41 coincides with the reciprocating direction (thrust direction) of the movable element 12. Reference numeral 3 denotes a leaf spring which supports the movable element 12 and the auxiliary mass 32 so as to be movable in the thrust direction. Reference numeral 11 denotes a shaft connecting the movable element 12 and the auxiliary mass 32, and is supported by the plate spring 3. Reference numeral 35 is a limiter that limits the movable range of the movable element 12, and limits the movable range in both ends (upper and lower limits in fig. 19) of the movable element 12. The excitation unit 30 constitutes a dynamic damper.

Next, the operation of the excitation section 30 shown in fig. 19 will be described. When an alternating current (sine wave current or rectangular wave current) flows through a coil (not shown) constituting a linear actuator (reciprocating motor), a magnetic flux is guided from an S pole to an N pole by a permanent magnet to form a magnetic flux loop in a state where a current flows through the coil in a predetermined direction. As a result, the movable element 12 moves in a direction (upward direction) opposite to the gravity. On the other hand, when a current in the direction opposite to the predetermined direction flows through the coil, the movable element 12 moves in the gravity direction (downward direction). The alternating current causes the direction of the current flowing through the coil to change alternately, and the movable element 12 repeats the above operation and reciprocates in the axial direction of the shaft 11 relative to the stator 34. Thereby, the auxiliary mass 32 connected to the shaft 11 vibrates in the vertical direction. The control force is adjusted by controlling the acceleration of the auxiliary mass 32 based on the control signal output from the current control unit 65, thereby reducing the vibration of the vehicle body under frame 41.

Next, the operation of the damper device shown in fig. 17 will be described with reference to fig. 17. First, the current detection unit 631 detects the current supplied to the excitation unit 30, and supplies the current to the current control unit 65 and the actuator vibration speed estimation unit 66. Further, the applied voltage detection unit 641 detects the voltage applied to the excitation unit 30 and supplies it to the actuator vibration speed estimation unit 66. When the linear actuator in the excitation section 30 is driven, the linear actuator generates an induced electromotive force proportional to the velocity. By calculating the induced electromotive force, the actuator vibration speed estimation unit 66 can obtain a vibration speed signal.

For example, the applied voltage V and the current i are detected and passed through an amplifying circuit and a differentiating circuit, thereby being output as the induced voltage E. In this case, gains K1 and K2 corresponding to the winding resistance R and the winding inductance L need to be set. The setting is adjusted so that a current of a predetermined frequency flows in a state where the movable portion (movable element, auxiliary mass) of the linear actuator is restrained, and the output becomes zero. Since the relation of E ═ V-R · i-L (di/dt) is established in the induced voltage E, the induced voltage E can be obtained by detecting the terminal voltage V and the current i. In addition, when a spring constant close to an optimum value is obtained in the excitation unit 30 by a magnetic spring characteristic or a mechanical spring element, a high vibration damping effect can be obtained without supplying energy for vibration damping by adjusting the damping force generated by the linear actuator. The damping force can be adjusted by connecting a load resistance to both ends of the coil of the linear actuator and changing the magnitude of the load resistance.

The output of the actuator vibration speed estimation unit 66 reflects the actual actuator output for the command value at the current time. If the linear actuator is assumed to be an ideal actuator, the force command signal required for the ideal actuator to output the vibration velocity estimated by the actuator vibration velocity estimation unit 66 is input to the current control unit 65, so the ideal actuator inverse characteristic unit 67 outputs the force command signal required for the ideal actuator to output the vibration velocity estimated by the actuator vibration velocity estimation unit 66. Equation (1) represents an example of a transfer function of the inverse characteristic of an ideal actuator.

[0125 Gi(s)＝(Mis2+Cis+Ki)/2 ...(1)

Wherein, Mi: auxiliary quality (ideal value), Ci: attenuation coefficient (ideal value), Ki: the spring constant (ideal value) is in the range of 1/100-100 times the attenuation coefficient of the transfer function of the critical attenuation (attenuation rate is 1).

The actual actuator can be made an ideal actuator by feeding back the differential value output between the actual command signal and the ideal actuator inverse characteristic unit 67 as the correction value of the command signal. The current control unit 65 derives an optimum driving amount (control amount) of the linear actuator so that the excitation unit 30 obtains an optimum spring characteristic and damping characteristic for controlling the vibration of the vehicle body under-frame 41, based on the current detected by the current detection unit 631 and the command signal output from the ideal actuator inverse characteristic unit 67, and outputs the derived result to the power amplifier 72 as a command signal. The power amplifier 72 drives the excitation unit 30 in accordance with the instruction signal of the current control unit 65, and the auxiliary mass 32 vibrates in the vertical (gravity) direction. The vibration generated in the vehicle body under frame 41 is suppressed by the reaction force caused by the vibration of the auxiliary mass 32.

In addition, a band-pass filter may be provided that limits the desired actuator inverse characteristics to a frequency band near the resonant frequency of the actuator.

In this way, by making the characteristics of the ideal actuator the same as those of the optimum dynamic damper, the active dynamic damper can be made the optimum dynamic damper, so that in the vibration damping device for an automobile, the resonance phenomenon is suppressed and the vibration width of the auxiliary mass can be set within an appropriate range, and therefore, the ideal vibration suppression can be realized and the vibration suppression performance can be improved.

Next, a modified example of the vibration damping device shown in fig. 17 will be described with reference to fig. 18. Fig. 18 is a block diagram showing a configuration of a modification of the vibration damping device shown in fig. 17. In this figure, the same parts as those of the apparatus shown in fig. 17 are assigned the same reference numerals, and the description thereof is omitted. The apparatus shown in fig. 18 is different from the apparatus shown in fig. 17 in that a control unit 69 is provided in place of the predetermined value output unit 68, and the control unit 69 acquires excitation source information (excitation timing, frequency, excitation force waveform, vehicle body vibration, and the like) and outputs a command value for suppressing vibration based on the excitation source information. The vibration suppression command value output by the control unit 69 is generated and output from the excitation force of the excitation source or the frequency information of the excitation force, and the vibration information or the excitation force information of the vehicle body underframe (excitation target equipment) 41. The excitation section 30 drives the linear actuator based on the vibration suppression command value. Other operations are the same as those described above, and therefore detailed description thereof is omitted here.

The vibration damping device for an automobile according to the present invention is effective when mounted on the under frame of the automobile body, the vicinity of the engine mount, the vicinity of the radiator, the lower part of the trunk or the lower part of the trunk in the rear part.

The linear actuator provided in the excitation section 30 is an actuator utilizing electromagnetic force, and is effective when a reciprocating motor is used, for example. The actuator provided in the excitation section 30 may be a piezoelectric actuator using an element that is displaced by application of a voltage.

As described above, since the resonance suppressing means of the actuator based on the ideal actuator inverse characteristic using the transfer function of the relative vibration speed of the excitation force with respect to the vibration system of the actuator is included, the characteristic of the actuator can be adjusted to an arbitrary characteristic by setting the ideal actuator inverse characteristic (invertebracteristic) based on the predetermined characteristic. Thus, by increasing the damping characteristic of the desired characteristic, the characteristic that the resonance of the movable portion of the actuator is less likely to be generated by the external force acting on the actuator main body can be achieved, and therefore, an appropriate reaction force can be generated to achieve a desired vibration suppression. Further, since the apparent natural frequency of the actuator can be reduced by reducing the natural frequency of the desired characteristic, stable vibration damping control can be realized even in the vicinity of the natural frequency of the actual actuator without being affected by the spring characteristic or the like.

[ 7 th embodiment ]

Next, a vibration damping device according to embodiment 7 of the present invention will be described with reference to the drawings. Fig. 20 is a block diagram showing the configuration of the same embodiment. In the figure, reference numeral 30 denotes an excitation unit which is fixed to a control target device 44 which is a target object of vibration damping control, and which suppresses vibration of the control target device 44 by driving an auxiliary mass member by a linear actuator (reciprocating motor) provided therein. The control target device 44 here refers to, for example, a body of an automobile.

Reference numeral 32 denotes an auxiliary mass (weight) added to the control target device 44. Reference numeral 34 denotes a stator constituting a reciprocating motor, and is fixed to the device to be controlled 44. Reference numeral 12 denotes a movable element constituting the reciprocating motor, and the movable element performs, for example, reciprocating movement in the direction of gravity (vertical movement on the paper surface of fig. 1). The excitation unit 30 is fixed to the control target device 44 such that the direction of vibration to be suppressed by the control target device 44 coincides with the reciprocating direction (thrust direction) of the movable element 12. Reference numeral 3 denotes a leaf spring which supports the movable element 12 and the auxiliary mass 32 so as to be movable in the thrust direction. Reference numeral 11 denotes a shaft connecting the movable element 12 and the auxiliary mass 32, and is supported by the plate spring 3. Reference numeral 35 is a limiter that limits the movable range of the movable element 12, and limits the movable range in both ends (upper and lower limits in fig. 20) of the movable element 12.

Reference numeral 620 denotes a command value generating means which receives a state value (for example, engine speed) of the device to be controlled 44, and calculates and outputs an amplitude command value and a frequency command value of vibration to be generated in the auxiliary mass. Reference numeral 621 is an amplitude upper limit clamp table defining an upper limit of an applicable current value determined according to the amplitude command value and the frequency command value output from the command value generation unit 620 for each frequency. Reference numeral 622 denotes an applied current generating unit that receives an amplitude command value and a frequency command value, performs correction for limiting the received amplitude command value to an appropriate movable range with reference to an amplitude upper limit clamp table 621, and obtains and outputs a command value for a current to be applied to the reciprocating motor based on the received frequency command and the corrected (limited) amplitude command value. Reference numeral 72 denotes a power amplifier which supplies a current to the stator 34 of the reciprocating motor constituting the excitation section 30 and controls the reciprocating movement of the movable element 12.

Next, the operation of the excitation section 30 shown in fig. 20 will be described. When an alternating current (sine wave current or rectangular wave current) flows through a coil (not shown) constituting the reciprocating motor, magnetic flux is guided from the S pole to the N pole by the permanent magnet to form a magnetic flux loop in a state where a current flows in a predetermined direction through the coil. As a result, the movable element 12 moves in a direction (upward direction) opposite to the gravity. On the other hand, when a current in the opposite direction to the predetermined direction flows through the coil, the movable element 12 moves in the gravity direction (downward direction). The alternating current causes the direction of the current flowing through the coil to change alternately, and the movable element 12 repeats the above operation and reciprocates in the axial direction of the shaft 11 relative to the stator 34. Thereby, the auxiliary mass 32 connected to the shaft 11 vibrates in the up-down direction. The control force is adjusted by controlling the acceleration of the auxiliary mass 32 based on the control signal output from the power amplifier 72, so that the vibration of the control target device 44 can be reduced.

In the linear actuator shown in fig. 20, the shaft 11 is not slidably supported so as to be movable back and forth, but the leaf springs 3 hold the movable element 12 at two positions on the upper end side and the lower end side of the shaft 11, and elastically deform themselves so that the movable element 12 is supported so as to be movable back and forth in the axial direction of the shaft 11. Thus, since the movable element 12 does not generate friction nor generate sliding resistance, the accuracy of the shaft support does not decrease even after a long-term use, high reliability is obtained, and the performance can be improved without loss of power consumption due to sliding resistance. However, as described above, even when the change in the trajectory of the automobile is strong due to rapid acceleration of the automobile or during traveling on a bad road, the variation in the current supplied to the stator 34 becomes large, and the movable element 12 collides with the stopper 35. When the excitation unit 30 is mounted as a vibration damping device in an automobile, it is desirable that there is no collision sound (abnormal noise) generated when the movable element 12 collides with the stopper 35.

Therefore, the upper limit value of the current which can be reapplied at the present time is obtained in advance for each current frequency applied to the stator 34, the relationship between the current frequency and the upper limit value of the current is replaced with the relationship between the amplitude command value and the frequency command value and stored in the amplitude upper limit clamp table 621, when the applied current generating unit 622 obtains a new applied current command value, the amplitude command value output from the command value generating unit 620 is corrected with reference to the amplitude upper limit clamp table 621, and the new applied current command value is obtained and output to the power amplifier 72 based on the corrected amplitude command value and the frequency command value output from the command value generating unit 620, so that the movable element 12 can be prevented from colliding with the limiter 35. Further, since the correction of the amplitude command value is performed by referring to the table, the amount of calculation in the applied current generating unit 622 can be reduced, and thus the processing can be speeded up, and a low-cost calculation device can be used to reduce the cost.

Next, a modified example of the vibration damping device shown in fig. 20 will be described with reference to fig. 21. In this figure, the same parts as those of the apparatus shown in fig. 20 are assigned the same reference numerals, and the description thereof is omitted. The apparatus shown in the figure is different from the apparatus shown in fig. 20 in that a current upper limit clamp table 623 is provided instead of the amplitude upper limit clamp table 621. The current upper limit clamp table 623 is a table in which an upper limit value of a current that can be reapplied at the present time is obtained in advance for each current frequency applied to the stator 34, and the relationship between the current frequency and the current upper limit value is replaced with the relationship between the applied current command value and the frequency command value, and is stored in advance. When the applied current generation means 622 determines a new applied current command value, the newly determined applied current command value is corrected based on the amplitude command value output from the command value generation means 620 and the frequency command value output from the command value generation means 620 with reference to the current upper limit clamp table 623, and is output to the power amplifier 72, whereby the movable element 12 can be prevented from colliding with the limiter 35.

[ 8 th embodiment ]

Next, a vibration damping device according to embodiment 8 of the present invention will be described with reference to fig. 22. Fig. 22 is a block diagram showing the configuration of this embodiment. In this figure, the same parts as those of the apparatus shown in fig. 20 are assigned the same reference numerals, and the description thereof is omitted. The apparatus shown in fig. 22 is different from the apparatus shown in fig. 20 in that a gradient limiting unit 625 is provided instead of the amplitude upper limit clamp table 621, and the applied current generating unit 624 obtains the applied current command value based on the amplitude command value in which the fluctuation gradient of the amplitude is limited by the gradient limiting unit 625. The gradient limiting unit 625 outputs the change gradient of the input amplitude command value as a gradual change. When applied current generating section 624 determines a new applied current command value, based on the amplitude command value limited by gradient limiting section 625 and the frequency command value output from command value generating section 620, the new applied current command value is determined and output to power amplifier 72, and the rapid fluctuation of the applied current can be suppressed, so that movable element 12 can be prevented from colliding against limiter 35. Further, by setting the limit only when the frequency variation is large, the delay of the response can be reduced.

Next, a modified example of the vibration damping device shown in fig. 22 will be described with reference to fig. 23. In this figure, the same parts as those of the apparatus shown in fig. 22 are assigned the same reference numerals, and the description thereof is omitted. The apparatus shown in the figure is different from the apparatus shown in fig. 22 in that the gradient limiting unit 625 is provided at a stage subsequent to the applied current generating unit 624. Gradient limiting section 625 has the same function as a low-pass filter, inputs the applied current command value obtained by applied current generating section 624, and outputs the input applied current command value with a fluctuation gradient that gradually fluctuates. Since the applied current command value newly obtained by the applied current generation means 624 is corrected so that the gradient of the fluctuation becomes gentle and is output to the power amplifier 72, the sudden fluctuation of the applied current can be suppressed, and the movable element 12 can be prevented from colliding with the limiter 35.

[ 9 th embodiment ]

Next, a vibration damping device according to embodiment 9 of the present invention will be described with reference to fig. 24. Fig. 24 is a block diagram showing the configuration of this embodiment. In this figure, the same parts as those of the apparatus shown in fig. 20 are assigned the same reference numerals, and the description thereof is omitted. The apparatus shown in fig. 24 is different from the apparatus shown in fig. 20 in that an amplitude suppression unit 627 and a fluctuation detection unit 629 are provided instead of the amplitude upper limit clamp table 621, and an applied current generation unit 628 obtains an applied current command value based on an amplitude command value whose amplitude is suppressed by the amplitude suppression unit 627. Amplitude suppressing section 627 suppresses the fluctuation of the amplitude command value in accordance with the fluctuation amount of the frequency command value detected by fluctuation detecting section 629. Fluctuation detecting section 629 constantly detects fluctuation of the frequency command value output from command value generating section 620, and notifies amplitude suppressing section 627 that the fluctuation amount exceeds a predetermined value when the fluctuation amount exceeds a predetermined value. When applied current generation section 628 determines a new applied current command value, a new applied current command value is determined based on the amplitude command value whose amplitude has passed amplitude suppression section 627 and has been suppressed based on the frequency fluctuation detected by fluctuation detection section 629 and the frequency command value output from command value generation section 620, and is output to power amplifier 72, whereby the rapid fluctuation of the applied current can be suppressed, and therefore, movable element 12 can be prevented from colliding with limiter 35. Further, by appropriately controlling the amount of suppression by the amplitude of the amplitude suppressing means 627, the driving can be continued to some extent even in the case of a rapid frequency variation.

Next, a modified example of the vibration damping device shown in fig. 24 will be described with reference to fig. 25. In this figure, the same parts as those of the apparatus shown in fig. 24 are assigned the same reference numerals, and the description thereof is omitted. The apparatus shown in the figure is different from the apparatus shown in fig. 24 in that a current suppressing unit 630 is provided instead of the amplitude suppressing unit 627. When the fluctuation amount of the frequency command value detected by the fluctuation detection unit 629 exceeds a predetermined value, the current suppression unit 630 suppresses the fluctuation of the applied current command value obtained by the applied current generation unit 628. When the variation of the frequency command value output from command value generating section 620 exceeds the predetermined value, the variation of the applied current command value newly obtained by applied current generating section 628 is corrected to be suppressed and output to power amplifier 72, whereby the rapid variation of the applied current can be suppressed, so that movable element 12 can be prevented from colliding with limiter 35.

As described above, when the current applied to the actuator (reciprocating motor) is controlled based on the amplitude command value and the frequency command value of the vibration to be generated, the current value applied to the actuator is limited so that the amplitude of the vibration of the auxiliary mass 32 does not exceed a predetermined value, and therefore the movable element of the actuator can be constantly driven within an appropriate movable range. This prevents the movable element 12 from colliding with the stopper 35, and therefore, the generation of collision noise can be suppressed. Further, since the movable element of the actuator can be driven within an appropriate movable range at all times by controlling the applied current value, the stopper 35 provided in the actuator (reciprocating motor) is not required, and the structure of the actuator can be simplified.

Further, the vibration suppression control may be performed by recording a program for realizing various functions in a computer-readable recording medium, and causing a computer system to read and execute the program recorded in the recording medium. The term "computer system" as used herein refers to a system including hardware such as an OS or peripheral devices. The "computer-readable recording medium" refers to a removable medium such as a flexible disk, a magneto-optical disk, a ROM, and a CD-ROM, and a storage device such as a hard disk built in a computer system. The "computer-readable recording medium" also includes a device that holds a program for a certain period of time, such as a volatile memory (RAM) in a computer system of a server or a client when the program is transmitted via a network such as the internet or a communication line such as a telephone line.

The program may be transmitted from a computer system that stores the program in a storage device or the like to another computer system via a transmission medium or by a transmission wave in the transmission medium. Here, the "transmission medium" for transmitting the program refers to a medium having a function of transmitting information, such as a network (communication network) such as the internet or a communication line (communication line) such as a telephone line. The program may be a program for realizing a part of the above-described functions. Further, the above-described functions may be realized by a combination with a program already recorded in a computer system, so-called differential documents (differential programs).

Industrial applicability of the invention

In the above description, the vibration damping target has been described as the body of an automobile, but the vibration damping target device of the present invention is not necessarily the body of an automobile, and may be an autonomous traveling carrier (autonomous traveling carrier), a robot arm (robot arm), or the like.

Claims (23)

1. A vibration damping device for an automobile for reducing vibration of an automobile body, characterized by comprising:

an actuator mounted on the vehicle body to drive the auxiliary mass;

a current detector that detects a current flowing through an armature of the actuator;

a detection unit that detects a terminal voltage applied to the actuator;

a power supply circuit outputting a voltage command;

an arithmetic circuit that calculates an induced voltage of the actuator based on the current detected by the current detector and the voltage command, and calculates at least one of a relative velocity, a relative displacement, and a relative acceleration of the actuator; and

and a control circuit configured to drive and control the actuator based on at least one of the relative velocity, the relative displacement, or the relative acceleration of the actuator calculated by the arithmetic circuit.

2. A vibration damping device for an automobile for reducing vibration of an automobile body, characterized by comprising:

an actuator mounted on the vehicle body to drive the auxiliary mass;

a spring element supporting the auxiliary mass so as to be movable in a driving direction of the actuator;

a current detector that detects a current flowing through an armature of the actuator;

a detection unit that detects a terminal voltage applied to the actuator;

a power supply circuit outputting a voltage command;

an arithmetic circuit that calculates an induced voltage of the actuator based on the current detected by the current detector and the voltage command, and calculates a relative speed of the actuator; and

and a control circuit that superimposes an attenuation characteristic on the actuator based on the relative velocity of the actuator calculated by the arithmetic circuit, and performs drive control on the actuator.

3. Vibration damping device for automobile according to claim 1 or 2,

the arithmetic circuit calculates the induced voltage in a frequency band defined around a natural vibration frequency of the actuator.

4. The vibration damping device for an automobile according to claim 3,

the arithmetic circuit includes a band-pass filter that limits a frequency band to a frequency band around a vibration frequency inherent to the actuator.

5. The vibration damping device for an automobile according to claim 4,

the band-pass filter is set to have a phase of 0 ° in the vicinity of a natural vibration frequency of the actuator.

6. The vibration damping device for an automobile according to claim 4,

the arithmetic circuit further includes a low-pass filter whose cutoff frequency is higher than the frequency of the natural vibration of the actuator.

7. Vibration damping device for automobile according to claim 1 or 2,

the terminal voltage is obtained from a command value of a voltage to be applied to the actuator.

8. Vibration damping device for automobile according to claim 1 or 2,

the vibration damping device for an automobile is mounted on a body underframe of the automobile.

9. Vibration damping device for automobile according to claim 1 or 2,

the vibration damping device for an automobile is mounted near an engine mount of the automobile.

10. Vibration damping device for automobile according to claim 1 or 2,

the vibration damping device for an automobile is installed near a radiator of the automobile.

11. Vibration damping device for automobile according to claim 1 or 2,

the vibration damping device for an automobile is installed at a lower portion of a cargo box or a lower portion of a trunk at the rear of the automobile.

12. Vibration damping device for automobile according to claim 1 or 2,

the actuator is an actuator utilizing electromagnetic force.

13. Vibration damping device for automobile according to claim 1 or 2,

the actuator is a reciprocating motor.

14. A control method for reducing vibration of a vehicle body, characterized by comprising:

detecting a current flowing through an armature of an actuator mounted on the vehicle body that drives an auxiliary mass;

a step of detecting a terminal voltage applied to the actuator;

calculating an induced voltage of the actuator based on the detected current and a voltage command output from a power supply circuit, and calculating at least one of a relative velocity, a relative displacement, or a relative acceleration of the actuator; and

and performing drive control of the actuator based on at least one of the calculated relative velocity, relative displacement, or relative acceleration of the actuator.

15. A control method for reducing vibration damping of a vehicle body, characterized by comprising:

detecting a current flowing through an armature of an actuator mounted on the vehicle body that drives an auxiliary mass;

a step of detecting a terminal voltage applied to the actuator;

calculating an induced voltage of the actuator based on the detected current and a voltage command output from a power supply circuit, and calculating a relative speed of the actuator; and

superimposing an attenuation characteristic based on the calculated relative velocity of the actuator, and performing drive control of the actuator.

16. The control method according to claim 14 or 15,

the induced voltage is calculated at a frequency band defined as around a natural vibration frequency of the actuator.

17. The vibration damping device according to any one of claims 1 to 13, which suppresses unnecessary vibration while generating prescribed vibration as needed, comprising:

an excitation member that vibrates the auxiliary mass supported by the linear actuator to excite the vibration reduction target;

a frequency detection unit that detects a frequency of an excitation source that vibrates the vibration control object;

a vibration detecting member that detects vibration at the measuring point;

an arithmetic unit that calculates a command value of vibration to be suppressed and a command value of vibration to be generated, based on the frequency of the excitation source and the vibration detected at the measurement point; and

and a control signal output unit configured to output a control signal obtained by superimposing the command value of the vibration to be suppressed and the command value of the vibration to be generated, to the excitation unit.

18. The vibration damping device of claim 17,

the exciting member vibrates the auxiliary mass supported by the linear actuator so as to be in a vibration state to be generated.

19. The vibration damping device of claim 17,

the excitation member includes a plurality of linear actuators having different vibration directions of the auxiliary mass,

the vibration detecting member is disposed at a position where vibration should be suppressed or vibration should be generated,

the calculation means obtains a command value of vibration to be suppressed and vibration to be generated at a position where the vibration detection means is provided.

20. The vibration damping device of claim 17,

the linear actuator is a reciprocating motor.

21. The vibration damping device of claim 17,

the vibration damping device is provided in a motor vehicle,

the arithmetic unit increases the amplitude of the bass vibration based on the music signal being reproduced in the vehicle based on the command value of the vibration to be generated.

22. The vibration damping device of claim 17,

the vibration damping device is provided in a motor vehicle,

the computing unit obtains a command value for suppressing vibration of engine vibration and reduces noise in the vehicle interior caused by the engine vibration.

23. The vibration suppression method according to claim 14 or 15, which is used for a vibration suppression device including an excitation member that excites a vibration suppression target by vibrating a supplementary mass supported by a linear actuator, so that the vibration suppression device suppresses unnecessary vibration and generates predetermined vibration as necessary, the vibration suppression method comprising:

a frequency detection step of detecting a frequency of an excitation source that vibrates the vibration reduction object;

a vibration detection step of detecting vibration at the measurement point;

an operation step of calculating a command value of vibration to be suppressed and vibration to be generated based on the frequency of the excitation source and the vibration detected at the measurement point; and

and a control signal output step of outputting a control signal in which the command value of the vibration to be suppressed and the command value of the vibration to be generated are superimposed to the excitation member.