Patent Publication Number: US-2017349022-A1

Title: Suspension device and suspension control unit

Description:
TECHNICAL FIELD 
     The present invention relates to a suspension device and a suspension control unit. 
     BACKGROUND ART 
     As such a type of the suspension device, for example, there is known a suspension device provided with a hydraulic cylinder interposed between a sprung member as a vehicle chassis and an unsprung member as a traveling wheel to actively exert a thrust force and a controller for controlling the hydraulic cylinder, the suspension device serving as an active suspension device. 
     The controller obtains a vertical velocity of the sprung member by processing a vertical acceleration of the sprung member detected by an acceleration sensor with a low-pass filter and obtains a required thrust force of the hydraulic cylinder necessary to suppress a vibration in the sprung member by multiplying the vertical velocity by a gain. In addition, the controller processes a vertical acceleration of the unsprung member with a band-pass filter and obtains a required thrust force of the hydraulic cylinder necessary to suppress a vibration in the unsprung member by multiplying the vertical acceleration by a gain. A resultant force between these thrust forces is set as a final target thrust force (for example, refer to JP 63-258207 A). 
     By obtaining the target thrust force in this manner, it is possible to suppress a vibration in the sprung member and the unsprung member using the hydraulic cylinder. 
     SUMMARY OF INVENTION 
     The suspension device performs control for suppressing a vibration in the unsprung member. The unsprung member typically has a resonance frequency of around 10 Hz. For this reason, the hydraulic cylinder is required to generate a thrust force capable of suppressing a vibration in a frequency band of around 10 Hz, and the suspension device is required to have a very high response characteristic. 
     In general, as a control valve for controlling the thrust force of the hydraulic cylinder, an electromagnetic valve provided with a valve body driven by a solenoid is employed. The solenoid has a response delay, so that it is difficult to control the electromagnetic valve with high accuracy at a frequency band of around 10 Hz. For this reason, using a suspension device provided with an electromagnetic valve, it is difficult to improve a vehicle ride quality. 
     The present invention provides a suspension device and a suspension control unit capable of improving a vehicle ride quality without using a high responsiveness device. 
     According to one aspect of the present invention, a suspension device includes: an actuator interposed between a sprung member and an unsprung member of a vehicle and capable of generating a thrust force; and a controller configured to control the actuator. The controller includes: a first vibration suppression force computation unit configured to obtain a first vibration suppression force from a vertical velocity of the sprung member, a second vibration suppression force computation unit configured to obtain a second vibration suppression force from a vertical velocity of the unsprung member or a relative velocity between the sprung member and the unsprung member, a low-pass filter having a breakpoint frequency between a sprung resonance frequency and an unsprung resonance frequency and processing a signal in a course of obtaining the second vibration suppression force in the second vibration suppression force computation unit, and a target thrust force computation unit configured to obtain a target thrust force of the actuator on the basis of the first vibration suppression force and the second vibration suppression force. 
     According to another aspect of the present invention, a suspension control unit configured to control an actuator interposed between a sprung member and an unsprung member of a vehicle, the actuator being capable of generating a thrust force, the suspension control unit includes: a first vibration suppression force computation unit configured to obtain a first vibration suppression force from a vertical velocity of the sprung member; a second vibration suppression force computation unit configured to obtain a second vibration suppression force from a vertical velocity of the unsprung member or a relative velocity between the sprung member and the unsprung member; a low-pass filter having a breakpoint frequency between a sprung resonance frequency and an unsprung resonance frequency and processing a signal in a course of obtaining the second vibration suppression force in the second vibration suppression force computation unit; and a target thrust force computation unit configured to obtain a target thrust force of the actuator on the basis of the first vibration suppression force and the second vibration suppression force. 
    
    
     
       BRIEF DESCRIPTION OF DRAWINGS 
         FIG. 1  is a diagram illustrating a suspension device and a suspension control unit according to a first embodiment; 
         FIG. 2  is a diagram illustrating a frequency characteristic of a low-pass filter; 
         FIG. 3  is a flowchart illustrating a processing flow for obtaining a target thrust force in a suspension control unit according to the first embodiment; 
         FIG. 4  is a diagram for describing motions of the suspension device and the vehicle according to the first embodiment in a dynamic sense; 
         FIG. 5A  is a diagram illustrating a frequency characteristic of a vibration transfer rate of an unsprung member for a road surface input; 
         FIG. 5B  is a diagram illustrating a frequency characteristic of a vibration transfer rate of a sprung member for a road surface input; 
         FIG. 6  is a diagram illustrating a suspension device and a suspension control unit according to a first modification of the first embodiment; 
         FIG. 7  is a flowchart illustrating a processing flow for obtaining a target thrust force in a suspension control unit according to a first modification of the first embodiment; 
         FIG. 8  is a diagram illustrating a suspension device and a suspension control unit according to a second modification of the first embodiment; 
         FIG. 9  is a flowchart illustrating a processing flow for obtaining a target thrust force in a suspension control unit according to a second modification of the first embodiment; 
         FIG. 10  is a diagram illustrating a suspension device and a suspension control unit according to a second embodiment; 
         FIG. 11  is a flowchart illustrating a processing flow for obtaining a target thrust force in a suspension control unit according to a second embodiment; 
         FIG. 12  is a diagram illustrating a suspension device and a suspension control unit according to a first modification of the second embodiment; 
         FIG. 13  is a flowchart illustrating a processing flow for obtaining a target thrust force in a suspension control unit according to a first modification of the second embodiment; 
         FIG. 14  is a diagram illustrating a suspension device and a suspension control unit according to a second modification of the second embodiment; 
         FIG. 15  is a flowchart illustrating a processing flow for obtaining a target thrust force in a suspension control unit according to a second modification of the second embodiment; and 
         FIG. 16  is a diagram illustrating an exemplary configuration of an actuator preferably employed in the suspension device. 
     
    
    
     DESCRIPTION OF EMBODIMENTS 
     Embodiments of the present invention will now be described with reference to the accompanying drawings. 
     First Embodiment 
     A suspension device  51  according to a first embodiment of the present invention will be described with reference to  FIG. 1 . 
     The suspension device  51  includes an actuator A interposed between a sprung member B as a vehicle chassis and an unsprung member W as a traveling wheel to generate a thrust force, a passive damper D interposed between the sprung member B and the unsprung member W in parallel with the actuator A, and a controller C 1  as a suspension control unit for controlling the actuator A. 
     The actuator A includes an extensible/contractible body E provided with a cylinder (not shown), a piston movably inserted into the cylinder to partition the cylinder into an extension-side chamber and a contraction-side chamber, and a rod movably inserted into the cylinder and connected to the piston, and a hydraulic pressure unit H that supplies and distributes fluid to the extension-side chamber and the contraction-side chamber to extensibly or contractibly drive the extensible/contractible body E. 
     As illustrated in  FIG. 1 , the extensible/contractible body E is interposed between the sprung member B and the unsprung member W of a vehicle. In  FIG. 1 , a vehicle is illustrated schematically, in which a suspension spring SP is installed between the unsprung member W and the sprung member B in parallel with the actuator A. In addition, a tire T installed in the traveling wheel serves as a spring provided between the road surface and the unsprung member W. In addition, a damper D is provided between the unsprung member W and the sprung member B in parallel with the actuator A. Although not shown specifically, the damper D is an extensible/contractible passive damper that exerts a damping force for suppressing extension or contraction when an external force is applied. 
     Although not shown specifically, the hydraulic pressure unit H has a hydraulic pressure source and a switching member capable of selectively supplying the fluid supplied from the hydraulic pressure source to any one of the extension-side chamber R 1  and the contraction-side chamber R 2  of the extensible/contractible body E. The switching member and the hydraulic pressure source of the hydraulic pressure unit H are driven by an electric current supplied from the controller C 1 . As the hydraulic pressure source and the switching member are driven, the fluid is supplied to the extension-side chamber R 1  or the contraction-side chamber R 2  of the extensible/contractible body E to extensibly or contractibly drive the extensible/contractible body E. 
     Meanwhile, the controller C 1  calculates a target thrust force Fref to be generated in the actuator A. That is, the controller C 1  supplies an electric current to the hydraulic pressure source and the switching member to allow the actuator A to exert the target thrust force Fref. In this manner, the actuator A is controlled by the controller C 1 . As the hydraulic pressure source, a pump driven by an accumulator or an engine of a vehicle may also be employed. In this case, if a control valve for controlling a pressure of the fluid supplied from the hydraulic pressure source such as a pressure control valve and the like is provided, the controller C 1  can control the pressure of the fluid supplied from the hydraulic pressure source using this control valve. Therefore, the controller C 1  is not necessary to directly control driving of the hydraulic pressure source. 
     The controller C 1  receives a vertical acceleration Gb of the sprung member B detected by an acceleration sensor  4  installed in the sprung member B and a vertical acceleration Gw of the unsprung member W detected by an acceleration sensor  5  installed in the unsprung member W. The controller C 1  processes such accelerations Gb and Gw and outputs an electric current for controlling the actuator A to the hydraulic pressure unit H. 
     The controller C 1  has a low-pass filter L 1  for filtering the vertical velocity Vw of the unsprung member W and obtains a target thrust force Fref of the actuator A on the basis of a first vibration suppression force F 1  obtained from the vertical velocity Vb of the sprung member B and a second vibration suppression force F 2  obtained from a velocity Vw processed by the low-pass filter L 1 . 
     Specifically, the controller C 1  includes an integrator  10  for integrating the acceleration Gb of the sprung member B input from the acceleration sensor  4  to obtain the vertical velocity Vb of the sprung member B, an integrator  11  for integrating the acceleration Gw of the unsprung member W input from the acceleration sensor  5  to obtain the vertical velocity Vw of the unsprung member W, a first vibration suppression force computation unit  12  for multiplying the velocity Vb output from the integrator  10  by a gain Cb to obtain the first vibration suppression force Fl, a second vibration suppression force computation unit  13  having a multiplier  14  for multiplying the velocity Vw by a gain Cw and processing a signal output from the multiplier  14  using the low-pass filter L 1  to obtain the second vibration suppression force F 2 , a target thrust force computation unit  15  for adding the first vibration suppression force F 1  and the second vibration suppression force F 2  to obtain the target thrust force Fref to be generated by the actuator A, a control command generator  16  configured to generate a control command transmitted to the switching member and the hydraulic pressure source of the hydraulic pressure unit H on the basis of the target thrust force Fref, and a driver  17  configured to output an electric current for driving the switching member and the hydraulic pressure source of the hydraulic pressure unit H by receiving the control command from the control command generator  16 . 
     The integrator  10  obtains the velocity Vb by integrating the acceleration Gb of the sprung member B. The integrator  10  may be, for example, a low-pass filter capable of simulatively integrating the acceleration Gb. Similarly, the integrator  11  may be a low-pass filter capable of simulatively integrating the acceleration Gw of the unsprung member W. 
     The first vibration suppression force computation unit  12  obtains the first vibration suppression force F 1  by multiplying the velocity Vb output from the integrator  10  by a gain Cb. The gain Cb is a gain multiplied to the velocity Vb in order to obtain the first vibration suppression force F 1  for predominantly suppressing a vibration in the sprung member B. For this reason, the gain Cb is set on the basis of a weight of the sprung member B and the like. The first vibration suppression force computation unit  12  obtains the first vibration suppression force F 1  by multiplying the velocity Vb by the gain Cb. Note that, if the first vibration suppression force F 1  is not linearly changed relative to the velocity Vb and has a characteristic difficult to be expressed using a numerical function, a relationship between the velocity Vb and the first vibration suppression force F 1  may be mapped, and the first vibration suppression force F 1  may be obtained from the velocity Vb using a map. 
     The multiplier  14  obtains a signal Fw in the course of obtaining the second vibration suppression force F 2  by multiplying the velocity Vw of the unsprung member W output from the integrator  11  by a gain Cw. The gain Cw is a gain multiplied to the velocity Vw in order to obtain the second vibration suppression force F 2  for predominantly suppressing a vibration in the unsprung member W. For this reason, the gain Cw is set on the basis of a weight of the unsprung member W and the like. 
     As illustrated in  FIG. 2 , the low-pass filter L 1  removes, from the frequency components of the signal Fw, a frequency component of a frequency band of the unsprung resonance frequency cow as a resonance frequency of the unsprung member W and passes a frequency component of a frequency band the sprung resonance frequency cob as a resonance frequency of the sprung member B. For this reason, the low-pass filter L 1  has a frequency characteristic in which a breakpoint frequency ωc is provided between the sprung resonance frequency ωb and the unsprung resonance frequency cow. The breakpoint frequency we may be set arbitrarily between the sprung resonance frequency cob and the unsprung resonance frequency ωw. The functionality obtained by the low-pass filter L 1  as described above is to remove a frequency component of a frequency band of the unsprung resonance frequency cow from the frequency components of the velocity Vw and pass a frequency component of a frequency band of the sprung resonance frequency ωb. Therefore, the breakpoint frequency we may be set around a median value between the sprung resonance frequency ωb and the unsprung resonance frequency ωw. In a typical vehicle, the sprung resonance frequency ωb is set to approximately 1 Hz, and the unsprung resonance frequency ωw is set to approximately 10 Hz. Therefore, the breakpoint frequency ωc as a frequency characteristic of the low-pass filter L 1  may be set to a range of, for example, 4 Hz or higher and 7 Hz or lower. 
     The second vibration suppression force F 2  is obtained as the signal Fw output from the multiplier  14  is processed by the low-pass filter L 1 . That is, the second vibration suppression force computation unit  13  includes the multiplier  14  and the low-pass filter L 1 . The second vibration suppression force computation unit  13  obtains the signal Fw in the course of obtaining the second vibration suppression force F 2  by multiplying the velocity Vw by the gain Cw. Alternatively, for example, if the second vibration suppression force F 2  is not linearly changed relative to the velocity Vw, and has a characteristic difficult to be expressed using a numerical function, a relationship between the velocity Vw and the signal Fw may be mapped, and the signal Fw may be obtained from the velocity Vw using a map. In addition, the signal Fw obtained from the multiplier  14  is filtered using the low-pass filter L 1 . Alternatively, after the filtering of the vertical velocity Vw of the unsprung member W, the second vibration suppression force F 2  may be obtained by multiplying the filtering result by the gain Cw in the multiplier  14 . In this manner, any signal in the course of computing the second vibration suppression force F 2  from the vertical velocity Vw of the unsprung member W may be processed by the low-pass filter L 1 . For this reason, when the processing using the low-pass filter L 1  is performed may be determined arbitrarily. 
     The target thrust force computation unit  15  adds the first vibration suppression force F 1  and the second vibration suppression force F 2  to obtain the target thrust force Fref to be generated by the actuator A. The second vibration suppression force F 2  is set to a very small value because a vibration frequency of the velocity Vw set around the unsprung resonance frequency ωw does not easily pass through the low-pass filter L 1 . Similarly, for the first vibration suppression force F 1 , since the velocity Vb is a vertical velocity of the sprung member B, a frequency component of the velocity Vb increases around the sprung resonance frequency ωb, but decreases in a band exceeding the sprung resonance frequency ωb or its vicinity. Therefore, the target thrust force Fref becomes a very small value in a high frequency area around the unsprung resonance frequency ωw and higher. 
     The control command generator  16  generates a control command transmitted to the switching member and the hydraulic pressure source of the hydraulic pressure unit H on the basis of the target thrust force Fref obtained by the target thrust force computation unit  15 . Specifically, a control command transmitted to the switching member is generated depending on a direction of the target thrust force Fref, that is, a direction of the thrust force generated by the actuator A, and a control command for instructing an electric current flowing to the hydraulic pressure source is generated from the magnitude of the target thrust force Fref. 
     If the switching member is, for example, an electromagnetic type direction switching valve and connects any one of the extension-side chamber R 1  and the contraction-side chamber R 2  to the hydraulic pressure source depending on whether or not an electric current is supplied to the solenoid of the direction switching valve, the control command of the control command generator  16  for driving the switching member may be a control command for instructing whether or not an electric current is applied to the solenoid. For example, if the hydraulic pressure source is a pump driven by a motor, the control command of the current command generator  16  for driving the motor may be a control command for instructing an electric current amount applied to the motor. In this manner, the control command generator  16  may generate the control command depending on a driving unit necessary to control extension or contraction of the actuator A. As described above, if the hydraulic pressure unit H has a pressure control valve, and the control of the supplied pressure is performed by the pressure control valve, the control command generator  16  may generate the control command for instructing an electric current amount applied to the solenoid of the pressure control valve. 
     The driver  17  outputs an electric current applied to the driving unit necessary to control extension/contraction of the actuator A, that is, in this case, each of the switching member and the hydraulic pressure source of the hydraulic pressure unit H in response to the control command input from the control command generator  16 . 
     The driver  17  has a driving circuit for driving the motor and the solenoid in a pulse-width modulation (PWM) method, if the hydraulic pressure source is a pump driven by a motor and the switching member is a direction switching valve driven by the solenoid. As the driver  17  receives the control command from the control command generator  16 , the electric current is supplied to the solenoid and the motor in response to the command. Alternatively, each driving circuit of the driver  17  may be any driving circuit other than the PWM-type driving circuit. 
     Next, a series of processing flow performed by the controller C 1  will be described. As illustrated in the flowchart of  FIG. 3 , the controller C 1  reads the vertical acceleration Gb of the sprung member B and the vertical acceleration Gw of the unsprung member W (step  501 ). Subsequently, the velocities Vb and Vw are obtained by integrating the accelerations Gb and Gw (step  502 ). Then, the controller C 1  obtains the first vibration suppression force Fl by multiplying the velocity Vb by the gain Cb (step  503 ). In addition, the controller C 1  multiplies the velocity Vw by the gain Cw to obtain a signal Fw (step  504 ), and performs filtering for the obtained signal Fw using the low-pass filter L 1  to obtain the second vibration suppression force F 2  by removing a frequency component exceeding the unsprung resonance frequency ωw from the signal Fw (step  505 ). 
     The controller C 1  adds the first vibration suppression force F 1  and the second vibration suppression force F 2  to obtain the target thrust force Fref (step  506 ). In addition, the controller C 1  generates a control command from the target thrust force Fref (step  507 ) and supplies an electric current from the driver  17  to the switching member and the hydraulic pressure source of the hydraulic pressure unit H (step  508 ). By repeating the aforementioned process, the controller C 1  controls the actuator A. Note that the aforementioned series of processing flows are just for illustrative purposes and may be changed appropriately. 
     In the suspension device  51  and the controller C 1  as a suspension control unit configured as described above, the target thrust force Fref of the actuator A is obtained on the basis of the first vibration suppression force F 1  obtained from the vertical velocity Vb of the sprung member B and the second vibration suppression force F 2  processed by the low-pass filter L 1 . For this reason, for vibrations of the sprung member B and the unsprung member W in a frequency area around the unsprung resonance frequency ωw and higher, the target thrust force Fref is significantly reduced, and the thrust force generated by the actuator A is also significantly reduced. 
     In this manner, if the sprung member B and the unsprung member W are vibrated at a frequency band of the unsprung resonance frequency ωw, the thrust force generated by the actuator A of the suspension device S 1  is reduced, so that the vibrations of the sprung member B and the unsprung member W are suppressed by a passive damping force of the damper D. As a result, a vehicle ride quality can be improved. 
     If the sprung member B and the unsprung member W are vibrated at a frequency band of the unsprung resonance frequency ωw, the target thrust force Fref becomes a small value. Therefore, even when there is a response delay in the switching member or the hydraulic pressure source, a vehicle ride quality is not degraded. 
     From the aforementioned description, using the suspension device S 1  and the controller C 1  as a suspension control unit, it is possible to improve a vehicle ride quality without using a high-responsiveness device. 
     If the unsprung member W is vibrated due to unevenness on a vehicle traveling road surface, this vibration is transmitted to the sprung member B by virtue of a repulsive force of the suspension spring SP and a damping force of the damper D, so that the sprung member B is vibrated. 
     As illustrated in  FIG. 4 , assuming “M 1 ” denotes a mass of the unsprung member W, “M 2 ” denotes a mass of the sprung member B, “X 0 ” denotes a road surface displacement, “X 1 ” denotes a vertical displacement of the unsprung member W, “X 2 ” denotes a vertical displacement of the sprung member B, each upward direction is set to “positive”, “F” denotes a thrust force of the actuator A, a contracting direction of the extensible/contractible body E is set to “positive,” “Ks” denotes a spring constant of the suspension spring SP, and “Cp” denotes a damping coefficient of the damper D, a motion equation of the sprung member B can be expressed as the following Formula (1). 
       [Formula 1] 
         M   2   X   2   ″=−C   p   X   2   ′+C   p   X   1   ′−K ( X   2   −X   1 )− F    (1)
 
     Since the factor “−C p X 2 ”&#39; in the right side of Formula (1) refers to a force opposite to a motion of the sprung member B, it acts to suppress a vibration of the sprung member B at all times and exhibits an effect of preventing a vibration of the sprung member B. Meanwhile, the factor “C p X 1 ” applies an effect of exciting a vibration of the sprung member B or in contrast, suppressing a vibration of the sprung member B depending on a sign of the value “X 1 ′.” Here, since the target thrust force Fref is obtained by adding the first vibration suppression force F 1  and the second vibration suppression force F 2 , the following Formula (2) is established. 
       [Formula 2] 
         F=C   2   X   2   ′+C   1   X   1 ′  (2)
 
     If this Formula (2) is substituted with Formula (1), the following Formula (3) can be obtained. 
       [Formula 3] 
         M   2   X   2 ″=−( C   p   +C   2 ) X   2 ′+( C   p   −C   1 ) X   1 ′−K s ( X   2   −X   1 )   (3)
 
     Focusing on the first factor of the right side of Formula ( 3 ), the first vibration suppression force F 1  is a force proportional to the velocity Vb of the sprung member B, and, similar to the force of the damper D for suppressing a vibration of the sprung member B, serves as a force for suppressing and damping a vibration of the sprung member B at all times. Meanwhile, focusing on the second factor of the right side of Formula (3), the second vibration suppression force F 2  serves as a force for canceling the force changing depending on a sign indicating whether a vibration of the sprung member B is excited or suppressed. That is, if the gains are equal (Cw=Cp), a parenthesized value of the second factor of the right side becomes zero. Therefore, the velocity Vw of the unsprung member W does not affect a vibration of the sprung member B. That is, if the second vibration suppression force F 2  is obtained by multiplying the velocity Vw of the unsprung member W by the gain Cw equal to the damping coefficient of the damper D, the second factor of the right side of Formula (3) becomes zero. Therefore, a force of exciting a vibration of the sprung member B is not exerted, and a mode of exciting a vibration of the sprung member B does not occur. As a result, it is possible to effectively suppress a vibration of the sprung member B. 
     Note that the thrust force exerted by the actuator A acts to reduce a damping effect for a vibration of the unsprung member W. Therefore, if the unsprung member W vibrates at the unsprung resonance frequency ωw, the vibration of the unsprung member W is excited. However, according to the first embodiment of the invention, using the low-pass filter L 1  having a characteristic capable of establishing the breakpoint frequency ωc as a cut-off frequency between the sprung resonance frequency ωb and the unsprung resonance frequency ωw, filtering is performed for the signal in the course of obtaining the second vibration suppression force F 2 . For this reason, the value of the second vibration suppression force F 2  becomes very small for a vibration at a frequency band of the unsprung resonance frequency ωw, so that the vibration of the unsprung member W can be suppressed by the damping force of the damper D. Meanwhile, the value of the second vibration suppression force F 2  becomes large for a frequency area lower than the unsprung resonance frequency ωw, so that it is possible to suppress the vibration of the sprung member B from being excited by the vibration of the unsprung member W. As a result, it is possible to obtain an excellent vibration suppression effect of the sprung member B. Therefore, for a vibration at a frequency band of the unsprung resonance frequency ωw of the unsprung member W, the vibration of the unsprung member W is not excited by the actuator A. Therefore, it is possible to suppress the vibration of the unsprung member W using the damping force of the damper D. 
     In this manner, using the suspension device Si and the suspension control unit, it is possible to reduce a vibration of the unsprung member W at a frequency band of the unsprung resonance frequency ωw and reduce a vibration of the sprung member B at a frequency band of the sprung resonance frequency ωb as indicated by the solid line in  FIGS. 5A and 5B , compared to a case where a thrust force control of the actuator A is performed by obtaining the second vibration suppression force without using the low-pass filter L 1  (as indicated by the dotted line in  FIGS. 5A and 5B ). 
     Next, a first modification of the first embodiment will be described. 
     In order to obtain the second vibration suppression force F 2 , as illustrated in  FIG. 6 , the velocity Vw of the unsprung member W may be obtained by obtaining a stroke velocity Vs of the extensible/contractible body E as a vertical relative velocity between the sprung member B and the unsprung member W and subtracting the stroke velocity Vs from the velocity Vb of the sprung member B. 
     Specifically, compared to a configuration of the controller C 1  of the suspension device S 1  of  FIG. 1 , a controller C 2  of a suspension device S 2  of  FIG. 6  detects a stroke displacement Xs of the extensible/contractible body E by preparing a stroke sensor  6  instead of the acceleration sensor  5  that detects an acceleration Gw of the unsprung member W. In addition, instead of the integrator  11 , a differentiator  18  is provided to obtain the stroke velocity Vs by differentiating the stroke displacement Xs. Furthermore, an unsprung velocity computation unit  19  is provided to obtain the vertical velocity Vw of the unsprung member W by subtracting the stroke velocity Vs from the velocity Vb of the sprung member B obtained by the integrator  10  in the unsprung velocity computation unit  19 . 
     The extensible/contractible body E is connected to the sprung member B and the unsprung member W. For this reason, by installing the stroke sensor  6  in the extensible/contractible body E, it is possible to detect a vertical relative displacement between the sprung member B and the unsprung member W. By differentiating the detected relative displacement, it is possible to obtain the relative velocity. The stroke sensor  6  may be provided integrally with the extensible/contractible body E or may be a separate member. 
     Next, a series of processing flows performed by the controller C 2  will be described. As illustrated in the flowchart of  FIG. 7 , the controller C 2  reads the vertical acceleration Gb of the sprung member B and the stroke displacement Xs of the extensible/contractible body E (step  601 ). Subsequently, the velocity Vb is obtained by integrating the acceleration Gb, and the stroke velocity Vs as a relative velocity between the sprung member B and the unsprung member W is obtained by differentiating the stroke displacement Xs (step  602 ). Then, the controller C 2  obtains the vertical velocity Vw of the unsprung member W by subtracting the stroke velocity Vs from the velocity Vb (step  603 ). The controller C 2  obtains the first vibration suppression force F 1  by multiplying the velocity Vb by a gain Cb (step  604 ). In addition, the controller C 2  obtains a signal Fw by multiplying the velocity Vw by a gain Cw (step  605 ), and performs filtering for the obtained signal Fw using the low-pass filter L 1  to obtain the second vibration suppression force F 2  by removing a frequency component exceeding the unsprung resonance frequency ωw from the signal Fw (step  606 ). 
     The controller C 2  obtains the target thrust force Fref by adding the first vibration suppression force F 1  and the second vibration suppression force F 2  (step  607 ). In addition, the controller C 2  generates a control command from the target thrust force Fref (step  608 ) and supplies an electric current to the switching member and the hydraulic pressure source of the hydraulic pressure unit H from the driver  17  (step  609 ). By repeating the aforementioned process, the controller C 2  controls the actuator A. Note that the aforementioned processing flow is just for illustrative purposes and may be changed appropriately. 
     Similarly, the suspension device S 2  and the controller C 2  as a suspension control unit obtain the target thrust force Fref of the actuator A on the basis of the second vibration suppression force F 2  processed by the low-pass filter L 1 . For this reason, for vibrations of the sprung member B and the unsprung member W in a frequency area around the unsprung resonance frequency ωw and higher, the thrust force generated from the actuator A becomes very small. Therefore, in this case, the vibrations of the sprung member B and the unsprung member W are suppressed by the passive damping force of the damper D. 
     Therefore, using the suspension device S 2  and the controller C 2  as a suspension control unit, similar to the suspension device Si and the controller C 1 , it is possible to improve a vehicle ride quality without using a high-responsiveness device. 
     In a vehicle having a vehicle height adjustment capability such as an air suspension, typically, a relative distance between the sprung member B and the unsprung member W is measured. For this reason, in the suspension device S 2  and the controller C 2  as a suspension control unit, this measurement value is used as the stroke displacement Xs. Therefore, it is possible to control the actuator A without separately providing a sensor for detecting the stroke displacement Xs. 
     Next, a second modification of the first embodiment will be described. 
     In the suspension devices Si and S 2  described above, the damper D is provided in parallel with the actuator A. Alternatively, as illustrated in  FIG. 8 , the damper D may be omitted, and the actuator A 1  may exert the damping force to be generated by the damper. 
     Specifically, in the suspension device S 3 , the controller C 3  additionally has a third vibration suppression force computation unit  20  that obtains a third vibration suppression force F 3  by multiplying the stroke velocity Vs by a gain Cp corresponding to the damping coefficient of the damper, compared to the controller C 2  of the suspension device S 2  of  FIG. 6 . The third vibration suppression force F 3  is a force corresponding to the damping force generated by the omitted damper D. 
     In the target thrust force computation unit  15 , the target thrust force Fref is obtained by adding the first vibration suppression force F 1 , the second vibration suppression force F 2 , and the third vibration suppression force F 3 . In this manner, since the third vibration suppression force F 3  is added to the target thrust force Fref, the actuator A 1  exerts the damping force instead of the omitted damper D. Even in this case, it is possible to improve a vehicle ride quality using the suspension device S 3  and the controller C 3  as a suspension control unit. 
     Next, a series of processing flows performed by the controller C 3  will be described. As illustrated in the flowchart of  FIG. 9 , the controller C 3  reads the vertical acceleration Gb of the sprung member B and the stroke displacement Xs of the extensible/contractible body E (step  700 ). Subsequently, the velocity Vb is obtained by integrating the acceleration Gb, and the stroke velocity Vs is obtained by differentiating the stroke displacement Xs (step  701 ). Then, the controller C 3  obtains the vertical velocity Vw of the unsprung member W by subtracting the stroke velocity Vs from the velocity Vb (step  702 ). The controller C 3  obtains the first vibration suppression force F 1  by multiplying the velocity Vb by the gain Cb (step  703 ). In addition, the controller C 3  obtains the signal Fw by multiplying the velocity Vw by the gain Cw (step  704 ). A frequency component exceeding the unsprung resonance frequency ωw is removed from the signal Fw by performing filtering using the low-pass filter L 1  for the obtained signal Fw to obtain the second vibration suppression force F 2  (step  705 ). 
     The controller C 3  obtains the third vibration suppression force F 3  by multiplying the stroke velocity Vs by the gain Cp (step  706 ). In addition, the controller C 3  obtains the target thrust force Fref by adding the first vibration suppression force F 1 , the second vibration suppression force F 2 , and the third vibration suppression force F 3  (step  707 ). In addition, the controller C 3  generates the control command from the target thrust force Fref (step  708 ) and supplies an electric current to the actuator A 1  from the driver  17  (step  709 ). By repeating the aforementioned process, the controller C 3  controls the actuator A 1 . Note that the aforementioned series of processing flows are just for illustrative purposes and may be changed appropriately. 
     Even in the suspension device S 3  and the controller C 3  as a suspension control unit, the target thrust force Fref of the actuator A 1  is obtained on the basis of the second vibration suppression force F 2  processed by the low-pass filter L 1 . In addition, the third vibration suppression force F 3  obtained from the stroke velocity Vs is added to target thrust force Fref. Therefore, it is possible to generate the damping force using the actuator A 1  instead of the damper. 
     For this reason, for vibrations of the sprung member B and the unsprung member W in the frequency area around the unsprung resonance frequency ωw and higher, the first vibration suppression force F 1  and the second vibration suppression force F 2  actively generated by the actuator A 1  as an actuator become very small, so that the vibrations of the sprung member B and the unsprung member W are suppressed by the third vibration suppression force F 3  (damping force) generated by the actuator A 1  as a damper. 
     Since the damping force to be generated by the damper is generated by the actuator A 1 , control responsiveness is obtained up to a frequency band of the unsprung resonance frequency ωw. In the suspension device S 3  and the controller C 3  as a suspension control unit, an electromagnetic actuator that does not use a hydraulic pressure is employed as the actuator A 1 . Therefore, it is possible to obtain control responsiveness up to a frequency band of the unsprung resonance frequency ωw and improve a vehicle ride quality. 
     Since the damping force to be generated by the damper can be generated by the actuator A 1 , it is possible to omit the damper D. For this reason, it is possible to reduce the manufacturing cost of the suspension device S 3 . 
     Second Embodiment 
     Next, a suspension device S 4  and a controller C 4  as a suspension control unit according to a second embodiment of the invention will be described with reference to  FIG. 10 . In the following description, differences from the first embodiment will be focused. In addition, like reference numerals denote like elements as in the first embodiment, and they will not be described repeatedly. 
     As illustrated in  FIG. 10 , the suspension device S 4  includes an actuator A interposed between a sprung member B as a vehicle chassis and an unsprung member W as a traveling wheel to generate a thrust force, a damper D interposed between the sprung member B and the unsprung member W in parallel with the actuator A, and a controller C 4  as a suspension control unit for controlling the actuator A. 
     Similar to the first embodiment, the actuator A includes an extensible/contractible body E provided with a cylinder (not shown), a piston movably inserted into the cylinder to partition the cylinder into an extension-side chamber and a contraction-side chamber, and a rod movably inserted into the cylinder and connected to the piston, and a hydraulic pressure unit H that supplies and distributes a fluid to the extension-side chamber and the contraction-side chamber to extensibly or contractibly drive the extensible/contractible body E. 
     Similar to the first embodiment, the controller C 4  obtains a target thrust force Fref to be generated by the actuator A and supplies an electric current to the switching member and the hydraulic pressure source of the hydraulic pressure unit H to allow the actuator A to exert the target thrust force Fref. 
     The controller C 4  receives a vertical acceleration Gb of the sprung member B detected by an acceleration sensor  4  installed in the sprung member B and a stroke displacement Xs of the extensible/contractible body E detected by a stroke sensor  6  installed in the extensible/contractible body E as a vertical relative displacement between the sprung member B and the unsprung member W. The controller C 4  processes the acceleration Gb and the stroke displacement Xs and outputs an electric current for controlling the actuator A to the hydraulic pressure unit H. 
     The controller C 4  has a low-pass filter L 2  for filtering a signal Fd in the course of obtaining the second vibration suppression force F 2  from the stroke velocity Vs obtained by differentiating the stroke displacement Xs and obtains the target thrust force Fref of the actuator A on the basis of the first vibration suppression force F 1  obtained from the vertical velocity Vb of the sprung member B and the second vibration suppression force F 2  processed by the low-pass filter L 2 . 
     Specifically, the controller C 4  includes an integrator  30  for integrating the acceleration Gb of the sprung member B input from the acceleration sensor  4  to obtain the vertical velocity Vb of the sprung member B, a differentiator  31  for differentiating the stroke displacement Xs of the extensible/contractible body E input from the stroke sensor  6  to obtain a stroke velocity Vs of the extensible/contractible body E as a vertical relative velocity between the sprung member B and the unsprung member W, a first vibration suppression force computation unit  32  for multiplying the velocity Vb output from the integrator  30  by a gain Cb to obtain the first vibration suppression force F 1 , a second vibration suppression force computation unit  33  having a multiplier  34  for multiplying the stroke velocity Vs by a gain Cs and processing a signal output from the multiplier  34  using a low-pass filter L 2  to obtain the second vibration suppression force F 2 , a target thrust force computation unit  35  for adding the first vibration suppression force F 1  and the second vibration suppression force F 2  to obtain a target thrust force Fref to be generated by the actuator A to cancel the damping force generated by the damper D, a control command generator  36  that generates a control command transmitted to the switching member and the hydraulic pressure source of the hydraulic pressure unit H on the basis of the target thrust force Fref, and a driver  37  that outputs an electric current for driving the switching member and the hydraulic pressure source of the hydraulic pressure unit H by receiving the control command from the control command generator  36 . 
     The integrator  30  obtains the velocity Vb by integrating the acceleration Gb of the sprung member B. The integrator  30 , for example, may be a low-pass filter capable of simulatively integrating the acceleration Gb. Similarly, the integrator  31  may be a high-pass filter capable of simulatively integrating the stroke displacement Xs. 
     The first vibration suppression force computation unit  32  obtains the first vibration suppression force F 1  by multiplying the velocity Vb output from the integrator  30  by a gain Cb. The gain Cb is a gain multiplied to the velocity Vb in order to obtain the first vibration suppression force F 1  for predominantly suppressing a vibration in the sprung member B. For this reason, the gain Cb is set on the basis of a weight of the sprung member B and the like. The first vibration suppression force computation unit  32  obtains the first vibration suppression force F 1  by multiplying the velocity Vb by the gain Cb. Note that, if the first vibration suppression force F 1  is not linearly changed relative to the velocity Vb and has a characteristic difficult to be expressed using a numerical function, a relationship between the velocity Vb and the first vibration suppression force F 1  may be mapped, and the first vibration suppression force F 1  may be obtained from the velocity Vb using a map. 
     The multiplier  34  obtains a signal Fd in the course of obtaining the second vibration suppression force F 2  by multiplying the stroke velocity Vs as a vertical relative velocity between the sprung member B and the unsprung member W output from the differentiator  31  by a gain Cs. The gain Cs is a gain multiplied to the stroke velocity Vs in order to obtain the second vibration suppression force F 2  for suppressing a relative movement of the sprung member B and the unsprung member W. 
     Similar to the low-pass filter L 1  of the first embodiment, as illustrated in  FIG. 2 , the low-pass filter L 2  removes a frequency component of the unsprung resonance frequency ωw as a resonance frequency of the unsprung member W from the frequency components of the signal Fd and passes a frequency component of the sprung resonance frequency ωb as a resonance frequency of the sprung member B. For this reason, the low-pass filter L 1  has a frequency characteristic in which a breakpoint frequency we is provided between the sprung resonance frequency ωb and the unsprung resonance frequency ωw. The breakpoint frequency ωc of the low-pass filter L 2  may be set to a range of, for example, 4 Hz or higher and 7 Hz or lower. 
     The second vibration suppression force F 2  is obtained as the signal Fd output from the multiplier  34  is processed by the low-pass filter L 2 . That is, the second vibration suppression force computation unit  33  includes the multiplier  34  and the low-pass filter L 2 . The second vibration suppression force computation unit  33  obtains the signal Fd in the course of obtaining the second vibration suppression force F 2  by multiplying the stroke velocity Vs by the gain Cs. Alternatively, for example, if the second vibration suppression force F 2  is not linearly changed relative to the stroke velocity Vs, and has a characteristic difficult to be expressed using a numerical function, a relationship between the stroke velocity Vs and the signal Fd may be mapped, and the signal Fd may be obtained from the stroke velocity Vs using a map. In addition, the signal Fd obtained from the multiplier  34  is filtered using the low-pass filter L 2 . Alternatively, after the filtering of the stroke velocity Vs, the second vibration suppression force F 2  may be obtained by multiplying the filtering result by the gain Cs using the multiplier  34 . In this manner, any signal in the course of computing the second vibration suppression force F 2  from the stroke velocity VS may be processed by the low-pass filter L 2 . For this reason, when the processing using the low-pass filter L 2  is performed may be determined arbitrarily. 
     The target thrust force computation unit  35  adds the first vibration suppression force F 1  and the second vibration suppression force F 2  to obtain a target thrust force Fref to be generated by the actuator A such that the damping force generated by the damper D is cancelled. The second vibration suppression force F 2  is set to a very small value because a vibration frequency of the stroke velocity Vs set around the unsprung resonance frequency ωw does not easily pass through the low-pass filter L 2 . Similarly, for the first vibration suppression force F 1 , since the velocity Vb is a vertical velocity of the sprung member B, a frequency component of the velocity Vb increases around the sprung resonance frequency ωb, but decreases in a band exceeding the sprung resonance frequency ωb or its vicinity. Therefore, the target thrust force Fref becomes a very small value in a high frequency area around the unsprung resonance frequency ωw and higher. 
     Similar to the first embodiment, the control command generator  36  generates a control command transmitted to the switching member and the hydraulic pressure source of the hydraulic pressure unit H on the basis of the target thrust force Fref obtained by the target thrust force computation unit  35 . Specifically, a control command transmitted to the switching member is generated depending on a direction of the target thrust force Fref, that is, a direction of the thrust force generated by the actuator A, and a control command for instructing an electric current flowing to the hydraulic pressure source is generated from the magnitude of the target thrust force Fref. 
     The driver  37  outputs an electric current applied to a driving unit necessary to control extension/contraction of the actuator A, that is, in this case, each of the switching member and the hydraulic pressure source of the hydraulic pressure unit H in response to the control command input from the control command generator  36 . 
     Next, a series of processing flow performed by the controller C 4  will be described. As illustrated in the flowchart of  FIG. 11 , the controller C 4  reads the vertical acceleration Gb of the sprung member B and the stroke displacement Xs (step  801 ). Subsequently, the velocity Vb is obtained by integrating the acceleration Gb, and the stroke velocity Vs is obtained by differentiating the stroke displacement Xs (step  802 ). Then, the controller C 4  obtains the first vibration suppression force F 1  by multiplying the velocity Vb by the gain Cb (step  803 ). In addition, the controller C 4  multiplies the stroke velocity Vs by the gain Cs to obtain the signal Fd (step  804 ), and performs filtering for the obtained signal Fd using the low-pass filter L 2  to obtain the second vibration suppression force F 2  by removing a frequency component exceeding the unsprung resonance frequency ωw of the signal Fd (step  805 ). 
     The controller C 4  adds the first vibration suppression force F 1  and the second vibration suppression force F 2  to obtain the target thrust force Fref such that the damping force generated by the damper D is cancelled (step  806 ). In addition, the controller C 4  generates a control command from the target thrust force Fref (step  807 ) and supplies an electric current from the driver  37  to the switching member and the hydraulic pressure source of the hydraulic pressure unit H (step  808 ). By repeating the aforementioned process, the controller C 4  controls the actuator A. Note that the aforementioned series of processing flows are just for illustrative purposes and may be changed appropriately. 
     In the suspension device S 4  and the controller C 4  as a suspension control unit configured as described above, the target thrust force Fref of the actuator A is obtained on the basis of the first vibration suppression force F 1  obtained from the vertical velocity Vb of the sprung member B and the second vibration suppression force F 2  processed by the low-pass filter L 2 . For this reason, for vibrations of the sprung member B and the unsprung member W in a frequency area around the unsprung resonance frequency ωw and higher, the target thrust force Fref is significantly reduced, and the thrust force generated by the actuator A is also significantly reduced. 
     In this manner, if the sprung member B and the unsprung member W are vibrated at a frequency band of the unsprung resonance frequency ωw, the thrust force generated by the actuator A of the suspension device S 4  is reduced, so that the vibrations of the sprung member B and the unsprung member W are suppressed by the passive damping force of the damper D. As a result, a vehicle ride quality can be improved. 
     If the sprung member B and the unsprung member W are vibrated at a frequency band of the unsprung resonance frequency ωw, the target thrust force Fref becomes a small value. Therefore, even when there is a response delay in the switching member or the hydraulic pressure source, a vehicle ride quality is not degraded. 
     From the aforementioned description, using the suspension device S 4  and the controller C 4  as a suspension control unit, it is possible to improve a vehicle ride quality without using a high-responsiveness device. 
     In a vehicle having a vehicle height adjustment capability such as an air suspension, typically, a relative distance between the sprung member B and the unsprung member W is measured. For this reason, in the suspension device S 4  and the controller C 4  as a suspension control unit, this measurement value is used as the stroke displacement Xs. Therefore, it is possible to control the actuator A without separately providing a sensor for detecting the stroke displacement Xs. 
     As illustrated in  FIG. 4 , assuming “M 1 ” denotes a mass of the unsprung member W, “M 2 ” denotes a mass of the sprung member B, “X 0 ” denotes a road surface displacement, “X 1 ” denotes a vertical displacement of the unsprung member W, “X 2 ” denotes a positive ascending vertical displacement of the sprung member B, “F” denotes a thrust force of the actuator A, a contracting direction of the extensible/contractible body E is set to “positive,” “Ks” denotes a spring constant of the suspension spring SP, and “Cp” denotes a damping coefficient of the damper D, a motion equation of the sprung member B can be expressed as the following Formula (4). 
       [Formula 4] 
         M   2   X   2   ″=−C ( X   2   ′−X   1 ′)− K   5 ( X   2   −X   1 )− F    (4)
 
     The factor “−Cp(X 2 ′−X 1 ′)” in the right side of Formula ( 4 ) refers to a force exerted by the damper D and exhibits an effect of exciting a vibration of the sprung member B or damping a vibration of the sprung member B depending on a sign of the value “X 1 ′.” 
     Here, since the target thrust force Fref is obtained by adding the first vibration suppression force F 1  and the second vibration suppression force F 2 , the following Formula (5) is established. 
       [Formula 5] 
         F=C   2   X   2   ′−C   1 ( X   2   ′−X   1 ′)   (5)
 
     If this Formula (5) is substituted with Formula (4), the following Formula (6) can be obtained. 
       [Formula 6] 
         M   2   X   2   ″=−C   2   X   2 ′−C p ( X   2   ′−X   2 )+ C   1 ( X   2   ′−X   1 ′)− K   5 ( X   2   −X   1 )   (6)
 
     Focusing on the first factor of the right side of Formula ( 6 ), the factor “−C 2 X 2 ′” is applied opposite to a motion of the sprung member B and suppresses a vibration of the sprung member B at all times, so that an effect of controlling a vibration of the sprung member B can be obtained. Meanwhile, the factor “+C 1 (X 2 ′−X 1 &#39;)” in the right side of Formula (6) is applied so as to cancel the factor “−C p ( X   2 ′−X 1 ′)” in the right side of Formula ( 6 ). Therefore, it applies an effect of reducing a force of exciting a vibration of the sprung member B. As recognized from Formula (6), the force exerted by the damper D is perfectly cancelled when “C 1 =C p ” so that a vibration suppression effect is applied to the sprung member B at all times by virtue of the force “−C 2 X 2 ′.” 
     Note that the thrust force exerted by the actuator A acts to reduce a damping effect for a vibration of the unsprung member W. Therefore, if the unsprung member W vibrates at the unsprung resonance frequency ωw, the vibration of the unsprung member W is excited. However, according to the second embodiment of the invention, using the low-pass filter L 2  having a characteristic capable of establishing the breakpoint frequency ωc as a cut-off frequency between the sprung resonance frequency ωb and the unsprung resonance frequency ωw, filtering is performed for the signal in the course of obtaining the second vibration suppression force F 2 . For this reason, the value of the second vibration suppression force F 2  becomes very small for a vibration at a frequency band of the unsprung resonance frequency ωw, so that the vibration of the unsprung member W can be suppressed by the damping force of the damper D. Meanwhile, the value of the second vibration suppression force F 2  becomes large for a frequency area lower than the unsprung resonance frequency ωw, so that it is possible to suppress the vibration of the sprung member B from being excited by the vibration of the unsprung member W. As a result, it is possible to obtain an excellent vibration suppression effect of the sprung member B. Therefore, for a vibration at a frequency band of the unsprung resonance frequency ωw of the unsprung member W, the vibration of the unsprung member W is not excited by the actuator A. Therefore, it is possible to suppress the vibration of the unsprung member W using the damping force of the damper D. 
     In this manner, similar to the first embodiment, using the suspension device S 4  and the suspension control unit, it is possible to reduce a vibration of the unsprung member W at a frequency band of the unsprung resonance frequency ωw and reduce a vibration of the sprung member B at a frequency band of the sprung resonance frequency ωb as indicated by the solid line in  FIGS. 5A and 5B , compared to a case where a thrust force control of the actuator A is performed by obtaining the second vibration suppression force without using the low-pass filter L 2  (as indicated by the dotted line in  FIGS. 5A and 5B ). 
     Next, a first modification of the second embodiment will be described. 
     In order to obtain the second vibration suppression force F 2 , the stroke velocity Vs is necessary. However, as illustrated in  FIG. 12 , the stroke velocity Vs may be obtained by subtracting the vertical velocity Vw of the unsprung member W from the vertical velocity Vb of the sprung member B. 
     Specifically, compared to a configuration of the controller C 4  of the suspension device S 4  of  FIG. 10 , a controller C 5  of a suspension device S 5  of  FIG. 12  detects an acceleration Gw of the unsprung member W by preparing an acceleration sensor  5  in the unsprung member W instead of the stroke sensor  6 . In addition, instead of the differentiator  21 , an integrator  38  is provided to obtain a velocity Vw of the unsprung member W by integrating the acceleration Gw of the unsprung member W. Furthermore, a stroke velocity computation unit  39  is provided to obtain the stroke velocity Vs as a vertical relative velocity between the sprung member B and the unsprung member W by subtracting the velocity Vw of the unsprung member W obtained by the integrator  38  from the velocity Vb of the sprung member B obtained by the integrator  30  using the stroke velocity computation unit  39 . 
     Next, a series of processing flows performed by the controller C 5  will be described. As illustrated in the flowchart of  FIG. 13 , the controller C 5  reads the vertical acceleration Gb of the sprung member B and the acceleration Gw of the unsprung member W (step  901 ). Subsequently, the velocities Vb and the Vw are obtained by integrating the accelerations Gb and Gw (step  902 ). Then, the controller C 5  obtains the stroke velocity Vs by subtracting the velocity Vw from the velocity Vb (step  903 ). The controller C 5  obtains the first vibration suppression force F 1  by multiplying the velocity Vb by the gain Cb (step  904 ). In addition, the controller C 5  multiplies the stroke velocity Vs by the gain Cs to obtain the signal Fd (step  905 ), and performs filtering for the obtained signal Fd using the low-pass filter L 2  to obtain the second vibration suppression force F 2  by removing a frequency component exceeding the frequency band of the unsprung resonance frequency ωw from the signal Fd (step  906 ). 
     The controller C 5  adds the first vibration suppression force F 1  and the second vibration suppression force F 2  to obtain the target thrust force Fref (step  907 ). In addition, the controller C 5  generates a control command from the target thrust force Fref (step  908 ) and supplies an electric current from the driver  37  to the switching member and the hydraulic pressure source of the hydraulic pressure unit H (step  909 ). By repeating the aforementioned process, the controller C 5  controls the actuator A. Note that the aforementioned series of processing flows are just for illustrative purposes and may be changed appropriately. 
     Similarly, in the suspension device S 5  and the controller C 5  as a suspension control unit described above, the target thrust force Fref of the actuator A is obtained on the basis of the second vibration suppression force F 2  processed by the low-pass filter L 2 . For this reason, for vibrations of the sprung member B and the unsprung member W in a frequency area around the unsprung resonance frequency ωw and higher, the thrust force Fref generated by the actuator A is significantly reduced. Therefore, in this case, the vibrations of the sprung member B and the unsprung member W are suppressed by the passive damping force of the damper D. 
     Therefore, similar to the suspension device S 4  and the controller C 4 , using the suspension device S 5  and the controller C 5  as a suspension control unit, it is possible to improve a vehicle ride quality without using a high-responsiveness device. 
     Next, a second modification of the second embodiment will be described. 
     In the suspension devices S 4  and S 5  described above, the damper D is provided in parallel with the actuator A. Alternatively, as illustrated in  FIG. 14 , the damper D may be omitted, and the actuator A 1  may exert a damping force to be generated by the damper. 
     Specifically, in the suspension device S 6 , the controller C 6  additionally has a third vibration suppression force computation unit  40  that obtains a third vibration suppression force F 3  by multiplying the stroke velocity Vs by a gain Cp corresponding to the damping coefficient of the damper, compared to the controller C 4  of the suspension device S 4  of  FIG. 10 . The third vibration suppression force F 3  is a force corresponding to the damping force generated by the omitted damper D. 
     In the target thrust force computation unit  35 , the target thrust force Fref is obtained by adding the first vibration suppression force F 1 , the second vibration suppression force F 2 , and the third vibration suppression force F 3 . In this manner, since the third vibration suppression force F 3  is added to the target thrust force Fref, the actuator A 1  exerts the damping force instead of the omitted damper D. Even in this case, it is possible to improve a vehicle ride quality using the suspension device S 6  and the controller C 6  as a suspension control unit. 
     Next, a series of processing flows performed by the controller C 6  will be described. As illustrated in the flowchart of  FIG. 15 , the controller C 6  reads the vertical acceleration Gb of the sprung member B and the stroke displacement Xs of the extensible/contractible body E (step  1000 ). Subsequently, the velocity Vb is obtained by integrating the acceleration Gb, and the stroke velocity Vs is obtained by differentiating the stroke displacement Xs (step  1001 ). Then, the controller C 6  obtains the first vibration suppression force F 1  by multiplying the velocity Vb by the gain Cb (step  1002 ). In addition, the controller C 6  obtains the signal Fd by multiplying the velocity Vs by the gain Cs (step  1003 ), and filtering is performed for the obtained signal Fd using the low-pass filter L 2 , so that the second vibration suppression force F 2  is obtained by removing a frequency component exceeding the unsprung resonance frequency ωw from the signal Fd (step  1004 ). 
     The controller C 6  obtains the third vibration suppression force F 3  by multiplying the stroke velocity Vs by the gain Cp (step  1005 ). In addition, the controller C 6  obtains the target thrust force Fref by adding the first vibration suppression force F 1 , the second vibration suppression force F 2 , and the third vibration suppression force F 3  (step  1006 ). In addition, the controller C 6  generates the control command from the target thrust force Fref (step  1007 ) and supplies an electric current to the actuator A 1  from the driver  37  (step  1008 ). By repeating this process, the controller C 6  controls the actuator A 1 . Note that the aforementioned processing flow is just for illustrative purposes, and may be changed appropriately. 
     Even in the suspension device S 6  and the controller C 6  as a suspension control unit, the target thrust force Fref of the actuator A 1  is obtained on the basis of the second vibration suppression force F 2  processed by the low-pass filter L 2 . In addition, the third vibration suppression force F 3  obtained from the stroke velocity Vs is added to the target thrust force Fref. Therefore, it is possible to generate the damping force using the actuator A 1  instead of the damper. 
     For this reason, for vibrations of the sprung member B and the unsprung member W in the frequency area around the unsprung resonance frequency ωw and higher, the first vibration suppression force F 1  and the second vibration suppression force F 2  actively generated by the actuator A 1  as an actuator become very small, so that the vibrations of the sprung member B and the unsprung member W are suppressed by the third vibration suppression force F 3  (damping force) generated by the actuator A 1  as a damper. 
     In order to generate the damping force to be generated by the damper by the actuator A 1 , control responsiveness is obtained up to a frequency band of the unsprung resonance frequency ωw. In the suspension device S 6  and the controller C 6  as a suspension control unit, an electromagnetic actuator that does not use a hydraulic pressure is employed as the actuator A 1 . Therefore, it is possible to obtain control responsiveness up to a frequency band of the unsprung resonance frequency ωw and improve a vehicle ride quality. 
     Since the damping force to be generated by the damper can be generated by the actuator A 1 , it is possible to omit the damper D. For this reason, it is possible to reduce the manufacturing cost of the suspension device S 6 . 
     Next, an exemplary configuration of the actuator A suitable for the suspension devices S 1 , S 2 , S 4 , and S 5  will be described in details with reference to  FIG. 16 . 
     The actuator A includes an extensible/contractible body E provided with a cylinder  1 , a piston  2  movably inserted into the cylinder  1  to partition the cylinder  1  into an extension-side chamber R 1  and a contraction-side chamber R 2 , and a rod  3  movably inserted into the cylinder  1  and connected to the piston  2 , and a hydraulic pressure unit H that supplies and distributes a fluid to the extension-side chamber R 1  and the contraction-side chamber R 2  to extensibly or contractibly drive the extensible/contractible body E. The extension-side chamber R 1  is a chamber compressed during an extension stroke, and the contraction-side chamber R 2  is a chamber compressed during a contraction stroke. 
     The hydraulic pressure unit H includes a pump P, a reservoir R connected to a suction side of the pump P, and a hydraulic circuit HC provided between the extensible/contractible body E and an assembly of the pump P and the reservoir R. 
     The hydraulic circuit HC includes a supply channel  51  connected to a discharge side of the pump P, a discharge channel  52  connected to the reservoir R, an extension-side passage  53  connected to the extension-side chamber R 1 , a contraction-side passage  54  connected to the contraction-side chamber R 2 , a direction switching valve  55  as a switching member that selectively connects one of the extension-side passage  53  and the contraction-side passage  54  to the supply channel  51  and connects the other one of the extension-side passage  53  and the contraction-side passage  54  to the discharge channel  52 , an extension-side damping element  56  provided in the extension-side passage  53  to apply resistance to a flow direct from the extension-side chamber R 1  to the direction switching valve  55  and allows an opposite flow, a contraction-side damping element  57  provided in the contraction-side passage  54  to apply resistance to a flow from the contraction-side chamber R 2  to the direction switching valve  55  and allows an opposite flow, a control valve  58  capable of adjusting a pressure of the supply channel  51  in response to the supplied electric current, a suction passage  59  which connect the supply channel  51  and the discharge channel  52 , a suction check valve  60  provided in the middle of the suction passage  59  to allow only a flow of the fluid direct from the discharge channel  52  to the supply channel  51 , and a supply-side check valve  61  provided between the control valve  58  and the pump P in the middle of the supply channel  51  to allow only a flow directed from the pump P side to the control valve  58  side. 
     The extension-side chamber R 1  and the contraction-side chamber R 2  is filled with a liquid such as hydraulic oil as a hydraulic fluid, and the reservoir R is filled with a liquid and a gas. The liquid filled in the extension-side chamber R 1 , the contraction-side chamber R 2 , and the reservoir R may include, for example, water or a water solution as well as the hydraulic oil. 
     The pump P is a unidirectional discharge type that receives a fluid from the suction side and discharges the fluid to the discharge side and is driven by the motor  62 . The motor  62  may be either a DC or AC motor, and various types of motors such as a brushless motor, an induction motor, or a synchronous motor may be employed. 
     The suction side of the pump P is connected to the reservoir R through the pump passage  63 , and the discharge side of the pump P is connected to the supply channel  51 . Therefore, as the pump P is driven by the motor  62 , the fluid is sucked from the reservoir R and is discharged to the supply channel  51 . 
     The direction switching valve  55  is a 4-port 2-position electromagnetic switching valve and selectively switches a state in which the supply channel  51  communicates with the extension-side passage  53 , and the discharge channel  52  communicates with the contraction-side passage  54 , and a state in which the supply channel  51  communicates with the contraction-side passage  54 , and the discharge channel  52  communicates with the extension-side passage  53 . By switching the direction switching valve  55 , the fluid supplied from the pump P can be supplied to any one of the extension-side chamber R 1  and the contraction-side chamber R 2 . If the pump  4  is driven while the supply channel  51  communicates with the extension-side passage  53 , and the discharge channel  52  communicates with the contraction-side passage  54 , the fluid is supplied to the extension-side chamber R 1 , and the fluid is discharged from the contraction-side chamber R 2  to the reservoir R, so that the actuator body A is contracted. Meanwhile, if the pump  4  is driven while the supply channel  51  communicates with the contraction-side passage  54 , and the discharge channel  52  communicates with the extension-side passage  53 , the fluid is supplied to the contraction-side chamber R 2 , and the fluid is discharged from the extension-side chamber R 1  to the reservoir R, so that the actuator body A is extended. 
     The extension-side damping element  56  includes an extension-side damping valve  56   a  that applies resistance to a flow from the extension-side chamber R 1  to the direction switching valve  55 , and an extension-side check valve  56   b  provided in parallel with the extension-side damping valve  56   a  to allow only a flow from the direction switching valve  55  to the extension-side chamber R 1 . Therefore, since the extension-side check valve  56   b  is maintained in a closed state for a flow of the fluid moving from the extension-side chamber R 1  to the direction switching valve  55 , the fluid passes through only the extension-side damping valve  56   a  and flows to the direction switching valve  55  side. In contrast, for a flow of the fluid moving from the direction switching valve  55  to the extension-side chamber R 1 , the extension-side check valve  56   b  is opened. Since the extension-side check valve  56   b  applies less resistance compared to the extension-side damping valve  56   a,  the fluid preferentially passes through the extension-side check valve  56   a  and flows to the extension-side chamber R 1  side. The extension-side damping valve  56   a  may be a throttling valve that allows a bidirectional flow or a damping valve such as a leaf valve or a poppet valve that allows only a flow directed from the extension-side chamber R 1  to the direction switching valve  55 . 
     The contraction-side damping element  57  includes a contraction-side damping valve  57   a  that applies resistance to a flow from the contraction-side chamber R 2  to the direction switching valve  55 , and a contraction-side check valve  57   b  provided in parallel with the contraction-side damping valve  57   a  to allow only a flow directed from the direction switching valve  55  to the contraction-side chamber R 2 . Therefore, since the contraction-side check valve  57   b  is maintained in a closed state for a flow of the fluid moving from the contraction-side chamber R 2  to the direction switching valve  55 , the fluid passes through only the contraction-side damping valve  57   a  and flows to the direction switching valve  55  side. In contrast, for a flow of the fluid moving from the direction switching valve  55  to the contraction-side chamber R 2 , the contraction-side check valve  57   b  is opened. Since the contraction-side check valve  57   b  applies less resistance to the flow compared to the contraction-side damping valve  57 a, the fluid preferentially passes through the contraction-side check valve  57   b  and flows to the contraction-side chamber R 2  side. The contraction-side damping valve  57   a  may be a throttling valve that allows a bidirectional flow or a damping valve such as a leaf valve or a poppet valve that allows only a flow directed from the contraction-side chamber R 2  to the direction switching valve  55 . 
     The control valve  58  is an electromagnetic valve and is provided in the middle of a control passage  64  that connects the supply channel  51  and the discharge channel  52  in parallel with the suction passage  59 . A pressure of the supply channel  51  in the upstream side of the control valve  58  can be controlled by adjusting a valve release pressure of the control valve  58 . The valve release pressure of the control valve  58  changes in proportion to the supplied electric current amount. As the supplied current amount increases, the valve release pressure increases. If no electric current is supplied, the valve release pressure is minimized. In addition, the control valve  58  has a characteristic having no pressure override in which a pressure loss increases in proportion to a flow rate in a practical area of the suspension devices S 1 , S 2 , S 4 , and S 5 . Note that the “practical area” refers to, for example, an area where the extensible/contractible body E is extended or contracted within a range of a velocity of 1 m/ sec, and the “characteristic having no pressure override in which a pressure loss increases in proportion to a flow rate of the control valve  58 ” in this practical area means that the pressure override is negligible for the flow rate passing through the control valve  58  when the extensible/contractible body E is extended or contracted within a range of a velocity of  1  m/sec. In addition, the valve release pressure of the control valve  58  in a non-conduction state is very small, and nearly no resistance is applied to the flow of the fluid passing in the non-conduction state. 
     Since the supply-side check valve  61  is provided between the control valve  58  and the pump P in the middle of the supply channel  51 , it is possible to prevent the fluid from reversely flowing to the pump P side by closing the supply-side check valve  61  even when the pressure of the direction switching valve  55  side is higher than the discharge pressure of the pump P. 
     Using the actuator A configured as described above, it is possible to actively extend or contract the extensible/contractible body E by driving the pump P using the motor  62 , supplying the fluid discharged from the pump P to one of the extension-side chamber R 1  and the contraction-side chamber R 2  connected to the pump P, and allowing the other one to communicate with the reservoir R through the discharge channel  52  by using the direction switching valve  55 . 
     If the extensible/contractible body E is extended or contracted by an external force, the liquid discharged from the extension-side chamber R 1  as the extension-side chamber R 1  is compressed passes through the extension-side damping valve  56   a  and then reaches the reservoir R through the control valve  58  or reaches the reservoir R without passing through the control valve  58  depending on a switching state of the direction switching valve  55 . In any case, the liquid discharged from the extension-side chamber R 1  inevitably passes through the extension-side damping valve  56 a. Therefore, a damping force that hinders extension of the extensible/contractible body E is applied. 
     Meanwhile, if the extensible/contractible body E is extended or contracted by an external force, the liquid discharged from the contraction-side chamber R 2  as the contraction-side chamber R 2  is compressed passes through the contraction-side damping valve  57   a  and then reaches the reservoir R through the control valve  58  or reaches the reservoir R without passing through the control valve  58  depending on a switching state of the direction switching valve  55 . In any case, the liquid discharged from the contraction-side chamber R 2  inevitably passes through the contraction-side damping valve  57   a.  Therefore, a damping force that hinders contraction of the extensible/contractible body E is applied. 
     That is, the actuator A has a functionality of generating the thrust force for actively extending and contracting the extensible/contractible body E and serves as a passive damper for a vibration input caused by an external force. 
     In this manner, the actuator A serves as both the actuator and the damper. Therefore, if the actuator A is employed in the suspension devices S 1 , S 2 , S 4 , and S 5 , the actuator A can serve as a damper for a vibration input having a frequency equal to or higher than the unsprung resonance frequency cow. For this reason, it is not necessary to separately provide the damper D in addition to the actuator A. Therefore, it is possible to reduce the manufacturing cost for the suspension devices S 1 , S 2 , S 4 , and S 5 . Note that the actuator A serving as a damper D is not limited to the actuator A having the aforementioned structure. Instead, the actuator A may have any other structure. 
     Embodiments of the present invention were described above, but the above embodiments are merely examples of applications of the present invention, and the technical scope of the present invention is not limited to the specific constitutions of the above embodiments. 
     This application claims priority based on Japanese Patent Application No. 2014-226735 filed with the Japan Patent Office on Nov. 7, 2014, the entire contents of which are incorporated into this specification.