JP4713355B2 - Active vibration control device - Google Patents

Description

  The present invention relates to an active vibration control device that generates a control force between a base and a mass body by an actuator and suppresses vibration, and more particularly to an active vibration control device that is suitable for suppressing vibration of a base that vibrates aperiodically. .

  Skyhook damper theory is well known as a method for calculating a control force in a conventional active vibration control device. The skyhook damper theory is a theory that controls damping so that the damping force acts on the absolute velocity of the mass body to be controlled, and the vibration is converged on the absolute static space instead of the base standard. In practice, the absolute velocity of the mass body is calculated by integrating the absolute acceleration of the mass body measured by the acceleration sensor, and the control force obtained by multiplying the absolute velocity by the damping coefficient is the inverse of the absolute velocity for the mass body. Add in the direction. This control theory is effective in reducing the absolute vibration of the mass body.

  By the way, if the calculation method of the control force in the active vibration control device based on the Skyhook damper theory described above is used, the relative displacement between the base and the mass body will increase conversely. In general, it is not preferable that the relative displacement between the two increases. In acceleration measurement, high-frequency noise and zero point deviation are unavoidable. In particular, the zero point deviation is controlled in order to diverge the absolute velocity in the process of calculating the absolute velocity by integrating the absolute acceleration. It becomes the cause which falls into an impossible state.

  In order to deal with this problem of zero point deviation, a method is considered in which frequency weighting is applied to the damping force calculated by the Skyhook damper theory so that control is not applied to the zero point deviation. (For example, refer to Patent Document 1). In order to perform frequency weighting, a high-pass filter (HPF) process for attenuating a low-frequency signal to the calculated damping force is performed. When this HPF process is performed by an analog circuit, an IIR (Infinite Impulse Response) type filter is used. On the other hand, when the filter processing is performed after the acceleration is digitized by A / D conversion, an FIR (Finite Impulse Response) type filter can be used in addition to the IIR type. However, when a FIR filter is used, a waste time corresponding to 1/2 of the filter coefficient length occurs in the filter output, so the FIR filter can mainly synchronize the dead time with periodic vibration. (For example, when periodic vibration such as engine vibration is a target). Therefore, when the target vibration is aperiodic, an IIR filter is generally used.

Japanese Patent No. 3060809

  For example, when the cutoff frequency is 1 [Hz], the HPF processing (low frequency removal) using the IIR filter has gain / phase characteristics as shown in FIG. dB] (about −30 [%]) and a phase advance of π / 2 [rad]. This phase advance significantly deteriorates the controllability as described in Patent Document 1, and this point will be described in detail.

  FIG. 1 shows a simple vibration model. The mass body 1 is supported by a spring 3 and a damper 4 on the base 2 in an anti-vibration manner, and an actuator 5 that generates a control force is provided between the mass body 1 and the base 2. An acceleration sensor 6 is attached to the mass body 1, and the absolute velocity of the mass body 1 is calculated from the absolute acceleration measured by the acceleration sensor 6 by the control means 7. Then, a reverse control force is applied to the mass body 1 by the actuator 5 in proportion to the calculated absolute speed. Random vibration is given to the base 2, and the mass body 1 is forcibly excited by the base vibration.

  FIG. 3 shows an acceleration transmission rate from the base 2 to the mass body 1 in the vibration model of FIG. Fig. 3 shows a passive model that does not perform active vibration control (indicated by a thin solid line in the figure) and an ideal active vibration control that does not perform IIR HPF processing using true absolute acceleration that does not include low-frequency noise. And a realistic active vibration control model (indicated by a thick solid line in the figure) that has been subjected to IIR type HPF processing.

  From FIG. 3, it can be seen that the realistic model subjected to the IIR type HPF process has a larger resonance magnification in the vicinity of 1.5 [Hz] than the ideal model. This is an influence resulting from the phase advance of the high-pass filter described above.

Furthermore, a mechanism in which the phase advance by the IIR HPF leads to an increase in the resonance magnification will be described with reference to FIG. As shown in FIG. 4, the vibration component of the frequency f [Hz] can be expressed as a vector having real and imaginary parts as components in the complex plane rotates at f [Hz]. It is assumed that the acceleration of the mass body 1 is at the position of the vector [1, 0] at a certain moment. At this time, the velocity is phase −π / 2 [rad], the amplitude is 1 / (2πf), that is, the vector [0, −1 / (2πf)], and the displacement is the vector [−1 / (2πf) ² , 0]. Therefore, the ideal damping force in the Skyhook damper theory is the vector [1, ca / (2πf)] multiplied by the active damping coefficient ca in the direction opposite to the speed (advanced phase π). On the other hand, when the absolute velocity is calculated from the absolute acceleration subjected to the IIR type HPF process, the phase is advanced, so that the phase relationship is as shown in FIG.

  Here, with reference to FIG. 5, the state of the vibration system will be described based on the relationship between the displacement and speed of the mass body 1 and the force acting on the mass body 1. In FIG. 5, only the direction of the vector is considered, and the size is not considered. Let the displacement of the mass body 1 at a certain frequency on the complex plane be a vector [-1, 0] (for the sake of clarity, the phase of the description of the skyhook damper theory in FIG. 5 is matched). At this time, the velocity is a vector [0, -1]. On the other hand, when the external force acting on the mass body 1 is a vector [1, 0] having an opposite phase to the displacement, the force is called a restoring force. The restoring force is a so-called spring force. Further, when the external force is a vector [0, 1] having a phase opposite to that of the speed, the force is called a damping force. The damping force is a so-called viscous damping force (damper force). In the passive vibration system, the external force is composed of a combination of a restoring force and a damping force, and the mass body 1 vibrates accordingly. When the external force vector is between the restoring force vector and the damping force vector as in the passive vibration system, the vibration system is stable. However, when the external force vector deviates from between the restoring force vector and the damping force vector, for example, when it has the same vector [-1, 0] as the displacement vector, the external force becomes a propulsive force. Furthermore, when it has the same vector [0, −1] as the velocity vector, the external force becomes an oscillation force. That is, when the external force vector deviates from between the restoring force vector and the decaying force vector, the vibration system is unstable. Therefore, in order to perform stable active vibration control, the sum of the external forces acting on the mass body 1 must be between the restoring force vector and the damping force vector. Even if the external force vector is in a stable region, it becomes more vibrational as it approaches the restoring force vector, so it is preferable that the external force vector be as close to the damping force vector as possible.

  Returning to the description of FIG. 4 based on the above description, the image control force vector when the IIR type HPF processing is performed is out of the stable region, and therefore the vibration system becomes unstable. That is, this instability causes an increase in vibration transmissibility in the vicinity of 1.5 [Hz] in the active vibration control model in which the IIR type HPF process is performed.

  As described above, the active vibration control based on the skyhook damper theory is realistic in order to reduce the relative displacement between the base 2 and the mass body 1 at a low frequency and to remove the low frequency noise caused by the acceleration measurement. Therefore, it is necessary to perform the IIR type HPF process on the absolute acceleration measured in the above, which causes a problem that the controllability is deteriorated.

  An object of the present invention is to provide an active vibration control device with improved controllability.

(1) In order to achieve the above object, the present invention calculates a control force from a base that vibrates aperiodically, a mass installed on the base, and motion information of the base or the mass. In an active vibration control apparatus having a control means and an actuator that generates a control force between the base and the mass body in response to a command from the control means, the control means transmits vibration information of a future mass body the acceleration and velocity as the initial value of the mass, with the speed is calculated assuming converging free vibration manner, from the vibration information of the exercise information of past mass body and future mass body, FIR digital filter processing Is used to calculate vibration information with a predetermined frequency weighting, and to calculate a predetermined actuator control force based on the vibration information.
With this configuration, controllability in active vibration control can be improved.

( 2 ) In the above ( 1 ), preferably, the control means is such that the natural frequency of the free vibration is updated as needed in accordance with past vibration information of the mass body.

( 3 ) In the above ( 2 ), preferably, the control means uses the absolute acceleration frequency measured immediately before as the natural frequency of the free vibration.

  According to the present invention, the controllability in the active vibration control device can be improved.

Hereinafter, the configuration and operation of the active vibration control device according to the embodiment of the present invention will be described with reference to FIGS. 1 and 6 to 12.
First, the configuration of the active vibration control device according to the present embodiment will be described with reference to FIG.
FIG. 1 is an explanatory diagram of past and future vibrations used in an active vibration control apparatus according to an embodiment of the present invention.

  The active vibration control apparatus according to the present embodiment is a damped vibration system in which a base 2 and a mass body 1 are coupled by a spring 3 and a damper 4, and an actuator 5 is provided in parallel with the spring 3 and the damper 4. The mass body 1 is provided with an acceleration sensor 6, and the absolute acceleration of the mass body 1 measured by the acceleration sensor 6 is A / D converted by the control means 7 and taken into the control program as a digital value. Further, the control means 7 outputs a control command to the actuator 5, and the actuator 5 receives the control command and applies a control force to the mass body 1.

  The active vibration control device according to the present embodiment is used for a work machine such as a wheel loader, a dump truck, a power shovel, and a vibration control of a vehicle such as an automobile. In particular, it is used for a work machine such as a wheel loader, a dump truck, a power shovel. It is suitable for. That is, in a working machine such as a wheel loader, a dump truck, or a power shovel, the traveling road surface is often rough. The vibration from the road surface such as rough terrain has a low frequency, while the active vibration control device of the present embodiment removes the low frequency component and suppresses the vibration of the higher frequency component. Therefore, when the active vibration control device of the present embodiment is used, vibration from rough terrain cannot be suppressed. Rather, vibration from the traveling road surface is transmitted to the driver of the work vehicle. The driving feeling that the foot is on the ground is obtained, which is preferable. On the other hand, vibrations caused by bucket work in a wheel loader, vibrations associated with dump truck mount operations, vibrations associated with power shovel boom operations, and the like can be effectively suppressed by the present embodiment.

Next, a specific configuration of the control means 7 used in the active vibration control apparatus according to the present embodiment will be described with reference to FIG.
FIG. 6 is a block diagram showing a specific configuration of the control means used in the active vibration control apparatus according to the embodiment of the present invention.

  The control means 7 includes past vibration storage means 7A, future vibration prediction means 7B, FIR type HPF processing means 7C, and control force calculation means 7D. The past vibration storage means 7A calculates and holds a past absolute velocity from the absolute acceleration of the mass body 1 measured by the acceleration sensor 6. The future vibration prediction means 7B estimates the future absolute speed based on the past absolute speed and the latest measured value held in the past vibration storage means 7A. The FIR type HPF processing means 7C performs FIR type HPF processing on the past absolute speed held in the past vibration storage means 7A and the future absolute speed estimated by the future vibration prediction means 7B, thereby reducing unnecessary low speed. Calculate absolute velocity with frequency components removed. The control force calculating means 7D obtains the active control force by inverting the sign by multiplying the absolute velocity obtained by the FIR type HPF processing means 7C by the active damping coefficient. The control force thus obtained is output as a control command to the actuator 5, and the actuator 5 causes the control force to act on the mass body 1.

Here, past and future vibrations of the mass body 1 used in the active vibration control apparatus according to the present embodiment will be described with reference to FIG.
FIG. 7 is an explanatory diagram of past and future vibrations used in the active vibration control apparatus according to the embodiment of the present invention.

The absolute acceleration of the mass body 1 actually measured is assumed to be a (i) [m / s ² ]. Here, i = −128 to i = −1 means a past sample, and i = 0 means a current (latest) sample. The noise n (i) [m / s ² ] is mixed in a (i), and there is a relationship expressed by the following formula (1) with the true absolute acceleration ar (i).

a (i) = ar (i) + n (i) (1)

The absolute velocity ν (i) [m / s] of the mass body 1 is obtained by integrating the measured absolute acceleration a (i) by the Euler approximation shown in the following equation (2). T [s] is a sampling interval.

ν (i + 1) = ν (i) + T · a (i) (2)

The past vibration storage means 7A holds the absolute speed for the past. The future vibration prediction means 7B calculates the future absolute acceleration a ′ (i) [m / s ² ] and the absolute velocity ν ′ (i) [m / s] (i = 1 to i = 127) of the mass body 1. The following equations (3) to (7) are used for calculation. These equations have the property that the absolute velocity ν ′ (i) converges to 0 by free damped oscillation. The coefficients that determine the convergence of the absolute velocity ν ′ (i) are κ [Ns / m] and c [Ns2 / m], which contribute to the vibration period and the attenuation, respectively. Here, in order to match the absolute velocity ν ′ (i) with the absolute velocity generated in the future as much as possible, it is necessary to consider the characteristics of the vibrations currently occurring.
a ′ (0) = a (0) (3)
v ′ (0) = v (0) (4)
j ′ (i) = − c · a (i) −κ · ν ′ (i) (5)
a ′ (i + 1) = a ′ (i) + T · j ′ (i) (6)
ν ′ (i + 1) = ν ′ (i) + T · a ′ (i) (7)

Therefore, in this embodiment, as shown in FIG. 8, the frequency f [Hz] of the absolute acceleration a (i) measured immediately before is detected, and the coefficient κ having the frequency f as a natural vibration is expressed as follows. It is calculated at any time according to the equation (8), and ν ′ (i) adapted to the natural frequency of the vibration system and the periodic vibration component existing in the non-periodic base vibration is calculated. The coefficient c is calculated by the following equation (9) with the damping ratio ζ being a fixed value, and is linked to the coefficient κ.

κ = (2 · π · f) ² (8)
c = ζ · 2√κ (9)

The FIR type HPF processing means 7C receives the past absolute velocity ν (i) and the estimated future absolute velocity ν ′ (i) calculated as described above, and performs the FIR type HPF processing according to the following equation (10). The absolute velocity ν / (0) performed can be determined.

In equation (10), h (i) is an FIR filter coefficient having a coefficient length of 256 points (i = −128 to i = 127).

  The calculation algorithm described above is executed every time the absolute acceleration a (0) is read by the control means 7.

Next, the FIR filter coefficient h (i) used in the active vibration control apparatus according to the present embodiment and its frequency response characteristic will be described with reference to FIG.
FIG. 9 is an explanatory diagram of FIR filter coefficients and their frequency response characteristics used in the active vibration control apparatus according to the embodiment of the present invention.

  FIG. 9 shows the FIR filter coefficient h (i) used in this embodiment and its frequency response characteristic. This frequency response characteristic is not a theoretical value, and is obtained by actually performing FIR type HPF processing on white noise using the filter coefficient h (i) and calculating input / output frequency responses. Normally, when the FIR type HPF filter is applied to the past input, the output causes a dead time (phase delay) that is ½ of the coefficient length. However, in this embodiment, ideal filter processing without phase shift can be performed by canceling the dead time using future data corresponding to ½ of the coefficient length.

The control force calculating means 7D obtains the active control force by multiplying the absolute velocity obtained by the FIR type HPF processing means 7C by the active damping coefficient and inverting the sign. That is, using the absolute velocity vf (0) obtained by the above process, the control force F (n) of the actuator is calculated by the following equation (l1).

F = −ca · νf (0) (11)

Note that ca [Ns / m] is an active attenuation coefficient.

  In order to confirm the operation of this control system, a vibration calculation model of the mass body 1 was constructed as follows. The control force F (n) of the actuator is calculated by equation (10). Then, the mass of the mass body 1 is m [kg], the spring constant is κp [N / m], cp [Ns / m] is the damping coefficient, and the true absolute acceleration of the mass body 1 is ar (i). , The true absolute velocity is νr (i) [m / s], the true absolute displacement is χr (i) [m], and the true absolute velocity of the substrate 2 is νrb (i) [m / s]. Assuming that the true absolute displacement is χrb (i) [m], the vibration of the mass body 1 is calculated by the following equations (12) to (14).

Here, the absolute displacement χrb (i) of the foundation 2 is assumed to be complete forced vibration (random vibration) that is not subjected to vibration of the mass body 1.

  Using the above vibration calculation model, the following four types of vibration characteristics were compared.

1) PASS (Active PASS: HPF is not used for measurement acceleration)
2) IIR1 (Active IIR1: IIR type HPF with a cutoff of 0.1 [Hz] is used to remove the DC component of measured acceleration)
3) IIR2 (active IIR2: IIR type HPF with a cut-off of 1.0 [Hz] is used to suppress relative displacement at low frequencies)
4) FIR (Active FIR: FIR type HPF with a cutoff of 1.5 [Hz] is used to suppress relative displacement at low frequencies)
Here, type 4) is a model of this embodiment.

Here, the true absolute acceleration ar (i) and the measured absolute acceleration a (i) of the mass body 1 in the active vibration control apparatus according to the present embodiment will be described with reference to FIG.
FIG. 10 is an explanatory diagram of absolute acceleration used in the active vibration control device according to the embodiment of the present invention.

  In FIG. 10, the solid line indicates the true absolute acceleration ar (i) of the mass body 1, and the dotted line indicates the measured absolute acceleration a (i). White noise and offset noise are intentionally mixed in the absolute acceleration a (i).

  Next, in order to compare the controllability of the active vibration control using FIG. 11, the control force Fr calculated from the true absolute acceleration ar (i) of the mass body 1 and the measured acceleration a (i) The result of having compared the frequency characteristic F / Fr of the control force F calculated by performing a filter process is shown.

  First, as shown in FIG. 11A, in the PASS of type 1), the displacement of the mass body 1 diverges due to noise in acceleration measurement, and an abnormally large control force is generated (a large steady state). Deviation has occurred). From this, it can be confirmed that in the active vibration suppression control using the skyhook damper theory, it is indispensable to perform HPF processing on the measured acceleration.

  Next, as shown in FIG. 11B, in IIR1 of type 2), the amplitude ratio is approximately 1 and the phase is approximately 0, and almost ideal control is achieved. From this, it is understood that an HPF having an extremely low cut-off frequency may be used for damping control (reduction of absolute vibration of the mass body 1).

  Next, IIR2 of type 3) shown in FIG. 11 (C), the amplitude ratio of the low frequency is small, so the control force for the low frequency is small, but the phase around 1 [HZ]. Is advanced by π / 2 [rad], so that it has a characteristic of oscillation.

  Finally, as shown in FIG. 11D, in the FIR of this embodiment of type 4), the phase advance is also suppressed to about π / 8 [rad] while the amplitude ratio of the low frequency is suppressed to be small. Therefore, active vibration control that suppresses the relative displacement between the low-frequency substrate 2 and the mass body 1, which has been a problem of the prior art, can be stably performed.

  Next, the result of comparing the vibration model (excluding PASS) damping performance and the low frequency relative displacement with the passive model will be described with reference to FIG.

  As shown in FIG. 12 (A), in the FIR of the present embodiment, the absolute acceleration peak at the natural frequency is reduced as compared with the passive model, and the relative displacement at the low frequency is the same as shown in FIG. 12 (B). It turns out that it is suppressed to the extent.

  As described above, according to the present embodiment, vibration information that has been subjected to predetermined frequency weighting by FIR digital filter processing is calculated from the motion information of the past base or mass body and the vibration information of the future base or mass body. Since a predetermined actuator control force is calculated based on the vibration information, filter processing with a small phase shift with respect to the ideal control force is possible, and controllability at low frequencies is improved.

  In addition, by calculating the vibration information of the future mass body assuming that the current mass body acceleration and velocity are the initial values and assuming that the velocity converges freely, it will cause the future mass body vibration. Considering the influence of the latter of the base vibration and the residual free vibration of the mass body, the accuracy of the filter processing is improved and the controllability is improved.

  In addition, the natural frequency of free vibration is updated as needed to adapt to the vibration information of the past mass body. Therefore, it is possible to estimate an appropriate residual free vibration of the mass body, so that the accuracy of the filtering process is improved and the controllability is improved.

It is a conceptual diagram of a vibration model. It is explanatory drawing of the gain and phase characteristic of HPF processing using an IIR type filter. It is explanatory drawing of the acceleration transmissibility from the base | substrate to the mass body in the vibration model of FIG. It is explanatory drawing of the phase relationship of mass body vibration and control force in the skyhook damper theory. It is explanatory drawing of the stability of control in the relationship between a vibrating body and excitation force. It is a block diagram which shows the specific structure of the control means 7 used for the active vibration control apparatus by one Embodiment of this invention. It is explanatory drawing of the vibration of the past used in the active vibration control apparatus by one Embodiment of this invention, and the future. It is explanatory drawing of the actual measurement acceleration containing the noise used in the active vibration control apparatus by one Embodiment of this invention. It is explanatory drawing of the FIR filter coefficient used in the active vibration control apparatus by one Embodiment of this invention, and its frequency response characteristic. Frequency characteristics (F / Fr) of ideal control force and actual control force It is explanatory drawing of the absolute acceleration used in the active vibration control apparatus by one Embodiment of this invention. It is comparison explanatory drawing of the controllability of the active vibration control by the active vibration control apparatus by one Embodiment of this invention. It is comparison explanatory drawing of the damping performance and low frequency relative displacement by the active vibration control apparatus by one Embodiment of this invention.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 ... Mass body 2 ... Base | substrate 3 ... Spring 4 ... Damper 5 ... Actuator 6 ... Acceleration sensor 7 ... Control means 7A ... Past vibration storage means 7B ... Future vibration prediction means 7C ... FIR type HPF processing means 7D ... Control force calculation means

Claims (3)

A base that vibrates aperiodically, a mass body installed on the base, control means for calculating a control force from motion information of the base or the mass body, and the base and the base according to a command from the control means In an active vibration control device having an actuator that generates a control force between mass bodies,
The control means, the vibration information of future mass, the acceleration and velocity of the current of the mass as an initial value, with speed calculated assuming converging free vibration manner, motion information of the past mass body active vibration from the vibration information of future mass body, characterized in that calculates the vibration information given frequency has been weighted by the FIR digital filter processing, it calculates a predetermined actuator control force based on the vibration information Control device. The active vibration control device according to claim 1,
The active vibration control apparatus according to claim 1, wherein the control means updates the natural frequency of the free vibration as needed in accordance with past vibration information of the mass body. The active vibration control device according to claim 2,
The active vibration control apparatus characterized in that the control means uses the absolute acceleration frequency measured immediately before as the natural frequency of the free vibration.

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