JPH08310214A - Suspension control device for vehicle - Google Patents

Abstract

PURPOSE: To satisfy both improvement of wheel grounding performance and improvement of comfortableness. CONSTITUTION: A suspension control device is provided with a shock absorber 3 provided with a damping coefficient changing means 50, a means 51 to detect vertical acceleration of a spring-weight car body, a means 52 to detect unspring- weight acceleration, a control means 53 to drive the damping coefficient changing means 50, a road surface-unspring weight relative speed estimation means 54 to esmitate a road surface-unspring weight relative speed HV0 based on the spring-weight acceleration and the unspring-weight acceleration, and a target doping coefficient computing means 55 which calculate a target damping coefficient Cu corresponding to force in proportion to the road surface-unspring weight relative speed HV0 . The device is driven to control that the damping coefficient of the damping coefficient changing means 50 coincides with the target damping coefficient Cu.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a suspension control device capable of changing a damping coefficient of a vehicle shock absorber.

[0002]

2. Description of the Related Art Conventionally, in a vehicle equipped with a shock absorber whose damping coefficient can be adjusted, an actuator is driven by a command from a controller to improve the ground contact property of the wheels and the riding comfort, and the damping coefficient of the shock absorber is adjusted. Variable control is known, and for example, there is a device disclosed in Japanese Patent Laid-Open No. 5-155222.

In order to improve the grounding property of the wheel, the damping coefficient is controlled so that the damping force generated in the shock absorber becomes a force proportional to the unsprung absolute velocity, and at the unsprung resonance frequency. Suppresses unsprung free vibration to improve the wheel's ground contact.

[0004]

However, in the above-mentioned conventional suspension control device, since the damping force generated in the shock absorber is a force proportional to the unsprung absolute speed, the suppression of unsprung free vibration. Is accompanied by deformation of the tire, and there is a problem that the amount of deformation of the tire increases when passing over uneven road surface and the ground contact property deteriorates.As with the shock absorber with a fixed damping coefficient,
The suppression of unsprung vibration and the reduction of unsprung vibration in the frequency range from the unsprung resonance frequency to the unsprung resonance frequency are contrary to each other.
That is, as shown in FIG. 21, when the unsprung displacement X ₁ is fixed by controlling the unsprung as shown by the broken line in the figure when passing through uneven road surface (solid line in the figure). However, since the relative displacement (X ₀ −X ₁ ) of the road surface displacement X ₀ and the unsprung displacement X ₁ is larger than that with a fixed damping coefficient, the deformation amount of the tire also increases. , The ground contact of the wheels will be reduced.

Further, as shown in FIG. 22, when unsprung vibration is suppressed to the same level as that of a shock absorber having a fixed damping coefficient, sprung vibration (X _{2 in the} figure) in a frequency range from the sprung resonance frequency to the unsprung resonance frequency is suppressed. / X ₀ ) is equivalent to that with a fixed damping coefficient, and there is a problem that improvement in riding comfort cannot be obtained.

The present invention has been made in view of the above problems, and an object of the present invention is to provide a suspension control device for a vehicle that can improve the riding comfort while improving the ground contact of the wheels.

[0007]

The first invention, as shown in FIG. 23, comprises damping coefficient changing means 50 which is interposed between the sprung and unsprung portions of each wheel to change the damping coefficient. A shock absorber 3, a sprung acceleration detecting means 51 for detecting acceleration of each wheel in the vertical direction of the vehicle body on the spring, and an unsprung acceleration detecting means 52 for detecting acceleration of each wheel in the vertical direction of the vehicle body, The target damping coefficient of the shock absorber 3 is calculated on the basis of the detection values of the sprung and unsprung acceleration detecting means, and the damping coefficient changing means 50 is operated so that the damping coefficient of the shock absorber 3 matches the target damping coefficient. In a vehicle suspension control device having a control means 53 for driving, the control means 53 uses the sprung mass acceleration and the unsprung mass acceleration to determine a relative speed H between a road surface and an unsprung mass. Road surface estimating the ₀ - comprises the unsprung relative speed estimating unit 54, a target damping coefficient calculation means 55 for calculating a target damping coefficient Cu _i corresponding to the force proportional to the road surface and the relative velocity of the unsprung.

The second invention, as shown in FIG. 23, is a shock absorber provided with damping coefficient changing means 50 which is interposed between the sprung and unsprung portions of each wheel to change the damping coefficient. 3, sprung acceleration detection means 51 for detecting acceleration of each wheel in the vertical direction of the vehicle body, unsprung acceleration detection means 52 for detecting acceleration of each wheel in the vertical direction of the vehicle body, and these sprung and The control means 53 calculates the target damping coefficient of the shock absorber 3 based on the detection value of the unsprung acceleration detecting means, and drives the damping coefficient changing means 50 so that the damping coefficient of the shock absorber matches the target damping coefficient. In a suspension control device for a vehicle having: a traveling state detection means 60 for detecting a traveling state of the vehicle; and the control means 53, the sprung acceleration and the spring. Road surface estimating the relative velocity HV ₀ under the road surface and the spring from the acceleration - unsprung relative speed estimating unit 54
And a target damping coefficient calculating means 55 for calculating a target damping coefficient Cu _i corresponding to a force proportional to the relative speed between the road surface and the unsprung portion, based on the detected control gain according to the traveling state of the vehicle. .

The third aspect of the present invention is the traveling state detecting means 60 according to the second aspect of the invention, as shown in FIG.
Is constituted by a vehicle speed detecting means 61 for detecting a vehicle speed, and the target damping coefficient calculating means 55 changes the control gain Cu according to the vehicle speed.

A fourth aspect of the present invention is the traveling state detecting means 60 according to the second aspect of the invention, as shown in FIG.
However, the target damping coefficient calculation means 55 changes the control gain Cu according to the turning state of the vehicle.

A fifth aspect of the present invention is the traveling state detecting means 60 according to the second aspect of the invention, as shown in FIG.
Is composed of a road surface condition detecting means 63 for detecting the condition of the road surface, and the target damping coefficient calculating means 55 changes the control gain Cu according to the road surface condition.

A sixth aspect of the invention is the road surface condition detecting means 63 in the fifth aspect of the invention, as shown in FIG.
Is a means 52 for detecting the acceleration of the unsprung body in the vertical direction.
The target damping coefficient calculation means 55 changes the control gain Cu according to the unsprung acceleration.

A seventh aspect of the present invention is the road surface condition detecting means 63 according to the fifth aspect of the invention, as shown in FIG.
However, means 6 for detecting the stroke of the shock absorber
5, the target damping coefficient calculation means 55 changes the control gain Cu according to this stroke.

The eighth invention is the first to seventh inventions.
23. As shown in FIG. 23, the road surface-unsprung relative velocity estimating means 54 calculates a relative displacement H ₀ between the road surface and the unsprung portion from the sprung acceleration and the unsprung acceleration, as shown in FIG. 23. The calculating means 56 and the differentiating means 57 for differentiating the relative displacement H ₀ are provided.

A ninth aspect of the present invention is the first to eighth aspects of the invention.
23, the target damping coefficient calculation means 55 includes a sprung speed calculation means 58 for calculating the speed ZV of the sprung body in the vertical direction based on the sprung acceleration. The target damping coefficient Cu _i corresponding to the resultant force of the force proportional to the relative velocity HV ₀ between the road surface and the unsprung portion and the force proportional to the sprung velocity ZV is calculated.

[0016]

Therefore, according to the first aspect of the present invention, the force proportional to the relative velocity HV ₀ between the road surface and the unsprung portion estimated based on the sprung mass acceleration and the unsprung mass acceleration of each wheel = the target damping coefficient Cu corresponding to the damping force. _i is calculated, the damping coefficient changing means is driven so that the damping coefficient of the shock absorber matches the target damping coefficient Cu _i , and excessive fluctuation of the relative displacement between the road surface and the unsprung portion is suppressed, so that the wheel groundability is improved. It is possible to suppress the vibration on the spring from the sprung resonance frequency to the unsprung resonance frequency while improving the above.

The second aspect of the invention is a target damping coefficient Cu corresponding to a force proportional to the road surface-unsprung relative velocity HV ₀ estimated based on the sprung mass acceleration and the unsprung mass acceleration of each wheel = damping force. _i can be calculated based on the detected control gain Cu corresponding to the traveling state, and the grounding property of the wheel can be improved while generating the damping force according to the traveling state.

A third aspect of the present invention is the control gain Cu.
Can be changed according to the detected vehicle speed to reduce the damping coefficient at low speeds and increase the damping coefficient at high speeds, thereby ensuring both ride comfort at low speeds and improved grounding of wheels at high speeds.

In a fourth aspect of the present invention, the control gain Cu is changed in accordance with the detected turning state, the damping coefficient is small when going straight, and the damping coefficient is increased when turning to ensure riding comfort when going straight. It is possible to both improve the grounding property of the wheel during turning.

In the fifth aspect of the invention, the control gain Cu is changed according to the detected road surface condition, and the damping coefficient is changed according to whether the road surface condition is good or not, so that ride comfort and wheel grounding are improved. Can be compatible.

According to the sixth aspect of the invention, the control gain Cu is changed in accordance with the detected unsprung acceleration to suppress the generation of unnecessary damping force when the unsprung acceleration is small and the input from the road surface is small. When the downward acceleration is large and the input from the road surface is large, unsprung vibration is controlled to ensure both riding comfort and improved grounding of the wheels.

According to the seventh aspect of the present invention, the control gain Cu is changed according to the detected stroke of the shock absorber to suppress the generation of unnecessary damping force when the stroke is small and the input from the road surface is small, while the stroke is reduced. When the input from the road surface is large, unsprung vibration is controlled to ensure both riding comfort and improved grounding of the wheels.

In the eighth invention, the relative velocity HV ₀ between the road surface and the unsprung portion can be obtained by differentiating the relative displacement H ₀ between the road surface and the unsprung portion obtained from the sprung mass acceleration and the unsprung mass acceleration.

The ninth aspect of the present invention is the target damping coefficient Cu _i
Is calculated so as to correspond to the resultant force of the force proportional to the relative velocity HV ₀ between the road surface and the unsprung portion and the force proportional to the sprung portion velocity ZV. Both the sprung vibration at the frequency and the unsprung vibration at the unsprung resonance frequency can be suppressed to improve the riding comfort.

[0025]

Embodiments of the present invention will be described below with reference to the accompanying drawings.

As shown in FIGS. 1 and 2, each wheel 2FL ...
Shock absorber 3FR between 2RL and vehicle body 1
3RL and a spring 4 are respectively interposed, and the vehicle body 1 constitutes a sprung portion and the wheels 2FR to 2RL constitute an unsprung portion. In addition, F
R indicates the right front wheel, FL the left front wheel, RR the right rear wheel, RL the left rear wheel, and so on.

This shock absorber 3FR to 3RL
The actuator 7FR is provided with a damping force adjusting mechanism driven by actuators 7FR to 7RL as a means for changing the damping coefficient as will be described later, and is driven in response to a command from the controller 10 mainly composed of the microcomputer 100. .About.7RL sets the shock absorbers 3FR to 3RL to target damping coefficients.

The controller 10 detects the vertical acceleration of the vehicle body 1, that is, the acceleration sensors 6FR to 6R for detecting the acceleration on the spring, and the vehicle height sensor 5FR for detecting the relative displacement between the vehicle body 1 and the wheels 2FR to 2RL. .About.5RL, the steering angle sensor 8 for detecting the steering angle .theta. Of the steering wheel, the speed sensor 9 for detecting the speed V of the vehicle, and the brake sensor 11 for detecting the ON / OFF of the brake pedal. The target values of the damping coefficients of the shock absorbers 3FR to RL are calculated based on the damper control, and the control signals are output to the actuators 7FR to 7RL to change the damping coefficients of the shock absorbers 3FR to 3RL.

As shown in FIG. 2, vehicle height sensors 5FR to 5FR.
RL, acceleration sensors 6FR to 6R, steering angle sensor 8, vehicle speed sensor 9, steering angle θ from brake sensor 11, vehicle speed V
Signals such as are input interface circuit 111, A / D
It is input to the microcomputer 100 after being converted into a digital signal via the converter 112.

As will be described later, the microcomputer 100 calculates the target damping coefficient of each shock absorber 3FR to RL based on the skyhook damper control, and the damping coefficient of each wheel calculated from this target damping coefficient is calculated as D / A.
The shock absorber 3FR is converted into an analog signal through the converter 113 and the driver circuit 114 and then amplified.
It is output as a control signal to the actuators 7FR to 7RL of 3RL to 3RL.

The detection means of each signal input to the controller 10 will be described in detail below, and then the damping force adjusting mechanism and control operation of the shock absorbers 3FR to 3RL will be described in this order.

[Acceleration Sensor] The acceleration sensors 6FR to 6R for detecting the acceleration on the spring in the vertical direction of the vehicle body are as shown in FIG.
As shown in FIG. 4, the base end is fixed to the vehicle body 1 side, and a mass 61 is provided at the free end of the semiconductor piezo element 60 arranged in a substantially horizontal direction. Since the semiconductor piezo element 60 is distorted in accordance with the above, the magnitude of the acceleration applied in the vertical direction of the vehicle body 1 is converted into a voltage.

As shown in FIG. 4, 2.5 V is output at 0 G, and in FIG. 3, when the upward acceleration in the figure is 1 G, it is 4.0 V, and similarly, when the downward acceleration is -1 G, it is 1 V. It outputs 0.0V.

Here, as shown in FIG. 5, the acceleration sensors 6FR to 6R are arranged at predetermined three positions of the vehicle body 1, and the acceleration sensor 6FR and the left front wheel 2FL are located near the right front wheel 2FR. The acceleration sensor 6FL is located near the rear right wheel RR.
It is disposed an acceleration sensor 6R is in the vicinity of, and the acceleration sensor 6FR, 6FL is provided substantially parallel to the front axle, to the detection to acceleration of the acceleration sensor 6FR~6R and ZG _1, ZG _2, ZG ₃ respectively .

The acceleration on the spring is determined by the wheels 2FR to 2RL.
The three acceleration sensors 6
Since the positions of the FRs to 6Rs are known, the controller 10 calculates the accelerations ZG _FR generated on the springs of the wheels 2FR to 2RL according to the following equation from the _three accelerations ZG _{1 to} ZG 3 generated on the vehicle body 1. Calculate ZG _RL .

ZG _FR = (a ₁ × ZG ₁ + b ₁ × ZG ₂ + c ₁ × ZG ₃ ) / d (1) ZG _FL = (a ₂ × ZG ₁ + b ₂ × ZG ₂ + c ₂ × ZG ₃ ) / _{d ... (2) ZG RR =} (a 3 × ZG 1 + b 3 × ZG 2 + c 3 × ZG 3) / d ... (3) ZG RL = (a 4 × ZG 1 + b 4 × ZG 2 + c 4 × ZG 3 ) / d ... (4) _{However, a 1 = -L 2 L 4} - (L 1 + L 3) L 6 -L 1 L 2 b 1 = L 2 L 4 + L 3 L 6 c 1 = L 1 L 6 a ₂ = L ₂ L _5- (L ₁ + L ₃ ) L ₆ -L ₁ L ₂ b ₂ = -L ₂ L ₅ + L ₃ L ₆ c ₂ = L ₁ L ₆ a ₃ = -L ₂ L ₄ + (L _{1 + L 3) L 7 -L} 1 L 2 b 3 = L 2 L 4 -L 3 L 7 c 3 = -L 1 L 7 a 4 = L 2 L 5 + (L 1 + L 3) L 7 -L 1 _{L 2 b 4 = -L 2 L} 5 -L 3 L 7 c 4 = -L 1 L 7 d = -L 1 L 2 L 1; acceleration sensor 6FR and 6FL The width direction of the distance L _2; acceleration sensor 6FR and distance full length body direction 6R L _3; acceleration sensor 6FR and the distance in the vehicle width direction of the 6R L _4; right front and rear wheels 2FR, from the axis passing through 2RR to acceleration sensors 6FR In the vehicle width direction L ₅ ; Distance in the vehicle width direction from the axis passing through the left and right front wheels 2FL, 2RL to the acceleration sensor 6FR L ₆ ; Distance in the entire vehicle body direction from the front axle to the acceleration sensor 6FR L ₇ ; Rear Distance from the axle to the acceleration sensor 6FR in the overall vehicle body length direction From the above equations (1) to (4), three accelerations ZG ₁ ,
Acceleration ZG on the spring corresponding to each wheel 2FR to 2RL from the arrangement positions of ZG ₂ and ZG ₃ and acceleration sensors 6FR to 6R.
_FR , ZG _FL , ZG _RR , ZG _RL can be obtained.

[Vehicle Height Sensor] The vehicle height sensors 5FR to 5RL for detecting relative displacement between the sprung portion and the unsprung portion, that is, the relative displacement between the vehicle body 1 and the wheels 2FR to 2RL are each wheel 2FR to.
Corresponding to 2RL, they are respectively arranged at predetermined positions of the vehicle body 1, and the signals of these sensors are input to a controller 10 mainly composed of a microcomputer.

These vehicle height sensors 5FR to 5RL are composed of, for example, potentiometers, etc., and the base ends of the connecting rods 22B are connected to the shafts of the vehicle height sensors 5FR to 5RL as shown in FIG. The free end of the connecting rod 22B is connected via the connecting rod 22A to the middle of the arm 21 for freely swinging and supporting the wheels 2FR to 2RL and the vehicle body 1 in the vertical relative displacement as shown in FIG. Is captured as a voltage change according to the angle change.

[Damping Force Adjusting Mechanism] FIGS. 8 to 11 show the damping force adjusting mechanisms of the shock absorbers 3FR to 3RL.
The actuators 7FR to 7RL provided on the vehicle body 1 side of the shock absorbers 3FR to 3RL have their damping coefficients changed as described below by rotating the control rod 30.
The RL is composed of, for example, a step motor or the like.

The piston 32 constituting the shock absorbers 3FR to 3RL is housed in the inner circumference of the cylinder 31 connected to the wheel side, and the inner circumference of the piston 32 has a cylindrical stud 33 and further the stud 33. A cylindrical spool 34 is coaxially accommodated on the circumference, and the spool 34 is connected to the control rod 30 to connect the actuators 7FR to 7FR.
While being rotatably supported by the RL, the stud 33
Is integrally fixed to the inner circumference of the piston 32, and the spool 34 coupled to the control rod 30 is rotatable relative to the piston 32 and the stud 33.

When the compression side stroke of the shock absorber is applied to the piston 32 and the stud 33 (the contracting direction of the spring 4)
Is formed with pressure side oil passages 35 and 36 through which hydraulic oil passes, and extension side oil passages 37 and 38 through which hydraulic oil passes during the extension side stroke of the shock absorber (the extension direction of the spring 4). Pressure side valve 35 for generating damping force
A and 35B and extension side valves 37A and 37B are arranged. Further, the stud 33 has a through hole 33A for communicating the upper surface of the piston 32 with the inner circumference of the stud 33, and the piston 3
A through hole 33B for communicating the lower surface of 2 with the inner circumference of the stud 33,
A through hole 33C is formed to connect the extension side oil passage 38 and the inner circumference of the stud 33.

Here, the spool 34 is provided with a circular oil passage 34A and an elliptical oil passage 34B formed as recesses at predetermined positions on the outer circumference, and the circular oil passage 34A is arranged so as to face the pressure side oil passage 36. At the same time, it communicates with the inner circumference 34C of the spool 34 and the oil chamber 31A via the through hole 34D. On the other hand, the elliptical oil passage 34B is arranged at a position where it can face the through holes 33A to 33C of the stud 33, and these through holes 33A to 33C.
Are configured to be able to communicate with each other.

Here, the shock absorbers 3FR to 3R
The damping coefficient of L is, as shown in FIGS.
Actuator 7FR via control rod 30
Determined according to the rotational position of the spool 34 driven by 7 RL, and in the case of the pressure side oil passage 36, when the spool 34 is rotated in the direction of the arrow in the figure as shown in FIG. By rotating 34A to 34A 'and overlapping the pressure side oil passage 36 and the circular oil passage 34A, the area of the communicating portion 300 communicating with each other is enlarged, and the damping coefficient is changed to the smaller one. The damping coefficient can be changed to an arbitrary value according to the area change of the communication portion 300, and the setting of the damping coefficient can be changed with almost no step and with high responsiveness. However, the same applies to the elliptical oil passage 34B.

Such shock absorbers 3FR to 3
As shown in FIGS. 11 (A) to 11 (C), the damping coefficient of RL can be set for each of the compression side and the extension side, that is, the flow path of the hydraulic oil varies depending on the movement direction of the shock absorbers 3FR to 3RL. The switching is performed as follows and the damping coefficient is also set.

When the extension side damping coefficient is large (hard);
As shown in FIG. 11 (A), the rotation of the spool 34 displaces the elliptical oil passage 34B to a position where it does not face the through holes 33A to 33C of the stud 33, and the hydraulic oil flowing from the oil chamber 31B to 31A is the piston 32. By passing through the inflow portion 37C between the pressure side valve 35A and the pressure side valve 35A, the extension side oil passage 37, and the extension side valve 37A sequentially, a large damping force is generated by a small flow passage cross-sectional area.

When the extension side damping coefficient is small (soft);
As shown in FIG. 11B, the elliptical oil passage 34B has the stud 33.
The hydraulic oil flowing from the oil chamber 31B into the oil chamber 31A is rotated by rotating the spool 34 to a position facing the through holes 33A to 33C.
In addition to the case where the damping coefficient is large, by sequentially passing through the inflow portion 37C on the upper surface of the piston 32, the through hole 33A, the elliptical oil passage 34B, the extension side oil passage 38, and the extension side valve 37B,
A small damping force is generated by increasing the flow passage cross-sectional area.

In addition to the oil passage 37, the extension side damping coefficient is determined by the oil passage 3
8 are selectively communicated with each other, and the overlapping area of the through holes 33A to 33C and the elliptical oil passage 34B is set to the actuator 7F.
By adjusting R to RL, it is possible to set an arbitrary damping coefficient.

When the damping coefficient on the compression side is large (hard): As shown in FIG. 11C, the spool 34 is rotated to a position where the circular oil passage 34A does not face the oil passage 36 of the stud 33.
The hydraulic oil flowing from the oil chambers 31A to 31B is supplied to the piston 3
By successively passing through the second pressure side oil passage 35 and the pressure side valve 35A, a large damping force is generated with a small flow passage cross-sectional area.

When the damping coefficient on the compression side is small (soft); the spool 34 is rotated to a position where the circular oil passage 34A faces the oil passage 35 of the stud 33 as shown in FIG. In addition to the case where the damping coefficient is large, the hydraulic oil flowing into 31B has an inner circumference 34C of the spool 34, a through hole 3
4D, circular oil passage 34A, pressure side oil passage 36, pressure side valve 35
By successively passing through B, the flow passage cross-sectional area is increased and a small damping force is generated.

In addition to the oil passage 35, the damping coefficient on the compression side has an oil passage 3
6 are selectively communicated, and the oil passage 36 and the circular oil passage 34 are further connected.
By adjusting the overlapping area of A with the actuators 7FR to RL, it is possible to set an arbitrary damping coefficient.

[Control Operation] The controller 10 controls the target value of the damping coefficient according to the relative velocity between the road surface and the unsprung surface calculated based on the vertical acceleration and relative displacement of the sprung and unsprung parts detected by the above-mentioned sensors. And a control signal is output to the actuators 7FR to RL according to the target damping coefficient to change the damping coefficient so that the shock absorbers 3FR to 3RL generate a predetermined damping force.

FIG. 12 is a flowchart showing an example of the control performed by the controller 10. This is executed at predetermined time intervals by a timer interrupt or the like, and will be described in detail below with reference to these flowcharts.

First, in step S1, the acceleration sensor 6FR
~ 6RL detected vertical acceleration ZG _{1 of} the vehicle body ₁ ~
The steering angle θ and the vehicle speed V detected by ZG ₃ , the steering angle sensor 8 and the vehicle speed sensor 9 are read respectively, and step S2
Then, the relative displacement between the vehicle body 1 and the wheels 2FR to 2RL detected by the vehicle height sensors 5FR to 5RL, that is, the strokes of the shock absorbers 3FR to 3RL are read as relative displacements H _{FR to} H _RL .

[0054] In step S3, from the acceleration ZG ₁ ~ZG ₃ read in step S1, the above (1) to (4)
The accelerations ZG _FR to ZG _RL generated on the springs of the wheels 2FR to 2RL are calculated based on the formulas, and in step S4, the absolute speeds ZV _{FR to} ZV on the springs are integrated by integrating these accelerations ZG _{FR to} ZG _RL. Calculate _RL respectively.

On the other hand, in step S5, the above step S
Differentiating the relative displacements H _FR 〜 H _RL between the sprung and unsprung portions read in 2, the relative velocity HV _FR 〜 between the sprung and unsprung _portions .
Calculate HV _RL .

The sprung absolute velocity thus obtained ZV _FR ~
Based on ZV _RL and the sprung and unsprung relative velocities HV _{FR to} HV _RL , in step S6, target damping coefficients C _{SFR to} C _SRL as ideal target values by skyhook damper control are obtained.
Is calculated from the following formula.

C _Si = −Cs × ZV _i / HV _i (5) where Cs is a coefficient of skyhook damper control, i = F
R, FL, RR, and RL are shown, and the wheels of interest are shown (the same applies hereinafter).

Next, in step S7, the magnitudes of these target damping coefficients C _Si are compared with the maximum damping coefficient C _max and the minimum damping coefficient C _min which can be set by the shock absorbers 3FR to 3RL, and the skyhook is determined according to the comparison result. The damper control damping coefficients C _{1FR to} C _1RL are set as in the following equations (6) to (8).

When C _Si ≦ C _min C _1i = C _min (6) When C _min <C _Si <C _max C _1i = C _Si (7) When C _Si > C _max C _1i = C _max (8) On the other hand, a step S8 differentiates the relative speeds HV _{FR to} HV _RL of the wheels 2FR to 2RL obtained in the step S5, and the relative accelerations HG of the sprung and unsprung positions at the positions of the wheels 2FR to 2RL. Calculate _{FR to} HG _RL .

In step S9, the relative acceleration HG _FR
~ HG _RL and sprung acceleration ZG _FR ~ ZG _RL each wheel 2F
The unsprung acceleration ZG _1FR ~ZG _1RL in R~2RL be calculated by the following equation.

ZG _1i = ZG _i + HG _i (9) Next, in step S10, the wheels 2FR to 2RL are each.
Unsprung acceleration ZG _1FR ~ZG _1RL and sprung acceleration ZG _FR of
From ZG _RL , relative displacements H _{0FR to} H _0RL between the road surface and the unsprung _part are calculated by the following equations.

HV _0i = (M _1i × ZG _1i + M _2i × Z _Gi ) / K ₁ (10) where M _1i ; unsprung mass of each wheel M _2i ; unsprung mass of each wheel K ₁ ; Vertical Spring Constant Here, each wheel 2FR to 2R based on the above equation (10).
The relative displacements H _{0FR to} H _0RL between the road surface of L and the unsprung _part will be described based on the ¼ vehicle model shown in FIG. 20, and the sprung and unsprung motion equations are expressed as the following equations, respectively. .

[0063]

[Equation 1]

However, X ₀ : road surface displacement X ₁ ; unsprung displacement X ₂ ; sprung displacement From these equations (11) and (12),

[0065]

[Equation 2]

Therefore, the relative displacements H _{0FR to} H _0RL between the road surface and the unsprung _parts can be obtained based on the equation (13).

Next, at step S11, the relative displacements H _{0FR to} H _0RL obtained as described above are differentiated to differentiate the relative velocity HV between the road surface and the unsprung wheels 2FR to 2RL.
After calculating the _{0FR ~HV 0RL,} an ideal target damping coefficient Cu _FR to CU _RL for unsprung attenuation control of each wheel 2FR~2RL in step S12, and the relative velocity HV _0i under the road surface and the spring , Is calculated from the unsprung and unsprung relative velocities HV _i by the following equation.

Cu _i = −Cu × HV _0i ÷ HV _i (14) However, Cu; unsprung damping coefficient (= control gain) Next, in step S13, the magnitude of these target damping coefficients Cu _i is set to the shock absorber 3FR to. The maximum damping coefficient C _max and the minimum damping coefficient C _min that can be set in 3RL are compared, and the target damping coefficients C _{2FR to} C _{2RL of the} respective wheels are determined according to the comparison result.
Is set as in the following equations (15) to (17).

When Cu _i ≦ C _min C _2i = C _min (15) When C _min <Cu _i <C _max C _2i = Cu _i (16) When Cu _i > C _max C _2i = C _max ... (17) in step S14, the skyhook damper control damping coefficients C _1FR -C _1RL obtained above step S7, the target damping coefficient C _2FR -C _2RL obtained in step S13 respectively added as follows Then, the damping coefficients C _{FR to} C _RL of each wheel are calculated.

[0070] difference between _{C i = C 1i + C 2i} ... (18) Next, at step S15, the damping coefficient C _i obtained in the current processing and the attenuation coefficient C _i obtained in the previous processing (n-1) ΔC _i is calculated respectively, and in steps S16, the actuators 7FR to 7RL corresponding to the difference ΔC _{i of the} damping coefficients are calculated.
Rotation angle [delta] _i by calculating the outputs a command signal corresponding to the rotational angle [delta] _i in step S17 to the actuator 7FR~7RL of each wheel 2FR～2RL.

By repeating the above steps S1 to S17 every predetermined time, the damping of the sprung and unsprung portions of the vehicle is performed by the damping coefficients C _{FR to} C _RL obtained according to the relative velocity HV _0i of the road surface and the unsprung portion. Can be done at the same time.

Considering a case where a running vehicle passes through unevenness on the road surface, as shown in FIG. 13, the damping force F generated by the shock absorbers 3FR to 3RL is a relative speed HV _0i between the road surface and the unsprung portion. Since the damping coefficient C _{2FR to} C _2RL is set in consideration, the relative displacement between the road surface and the unsprung part (X ₀ −X ₁ ) is greater than that with the fixed damping coefficient shown by the solid line in the figure. The ground contact of the wheel can be improved as compared with the conventional example shown by the broken line in FIG. 21, and the displacement on the spring is also reduced as compared with the conventional example, so that the riding comfort of the vehicle is improved. It is possible to improve.

In the conventional example shown by the broken line in FIG. 21, the damping force of the shock absorber is controlled so as to be a force proportional to the vertical absolute velocity of the unsprung part.
The unsprung displacement becomes small regardless of the vertical displacement of the road surface, which deteriorates the ground contact property. On the other hand, in the present invention, the damping coefficient C _2FR ~ according to the relative velocity HV _0i between the road surface and the unsprung surface is reduced.
Since the damping force F taking C 2 _RL into consideration is set, it is possible to secure the grounding property.

As shown in FIG. 14, in the frequency response of the suspension according to the present invention, the unsprung resonance at the unsprung resonance frequency is reduced, while the unsprung resonance frequency to the unsprung resonance frequency range (about 1 in the figure). .5-10H
The unsprung vibration in z) is also reduced, and it is possible to both suppress unsprung vibration at the unsprung resonance frequency and reduce sprung vibration in the frequency range from the unsprung resonance frequency to the unsprung resonance frequency, and to ground the wheel. It is possible to improve the ride quality and the riding comfort at the same time.

On the other hand, in the frequency response of the conventional example, as shown by the broken line in FIG. 22, unsprung vibration at the unsprung resonance frequency is reduced, but from the sprung resonance frequency to the unsprung resonance frequency. The sprung vibration in the frequency range of is worse than the shock absorber with a fixed damping coefficient, and it is not possible to suppress both sprung and unsprung vibrations at the same time.

Regarding the vibration on the spring at the resonance frequency on the spring, as compared with the shock absorber having the fixed damping coefficient, as shown in FIG. 15, by the skyhook damper control according to the above equations (5) to (8). Therefore, it is possible to significantly reduce the vibration, and to further improve the riding comfort by further promoting the vibration suppression on the sprung part and the unsprung part.

It should be noted that, in the above embodiment, the vertical acceleration sensor 6FL having the vertical acceleration under the spring arranged on the spring.
6R and the outputs from the vehicle height sensors 5FR to 5RL were obtained from the second-order differentiated value, but although not shown, the output of the vertical acceleration sensor provided under the spring can be used.

Further, the relative displacement H _0i between the road surface and the unsprung portion is obtained from the vertical acceleration on the spring and the vertical acceleration under the spring, and the relative velocity HV _0i between the road surface and the unsprung portion is obtained by differentiating this relative displacement H _0i. However, the differential value of the vertical acceleration on the spring,
The relative velocity HV _i between the road surface and the unsprung portion may be obtained from the differential value of the unsprung vertical acceleration.

Further, in the above embodiment, the skyhook damper control is combined to suppress the vibration on the spring, but it may be combined with other damping force control or may be controlled independently. Good.

FIG. 16 is a flow chart showing the second embodiment, and step S12 shown in FIG. 12 of the first embodiment.
The control gain Cu of the equation (14) performed in step S14 is made variable according to the vehicle speed V, and the control gain C is set after the step S11.
A step S20 for calculating u is added, and the other points are the same as those in the first embodiment.

In FIG. 16, in step S20, the relationship between the vehicle speed V and the control gain Cu corresponding to the detected value of the vehicle speed sensor 9 is set in advance as a map or a function.
The control gain Cu corresponding to is calculated.

This control gain Cu is, for example, an input from the road surface is not so large at low speed running, and vibration under the spring is unlikely to cause a problem. Therefore, in a predetermined low speed range, the control gain Cu is set to a small value and unnecessary damping force is reduced. While it is necessary to secure the stability of the vehicle during high-speed traveling while suppressing the occurrence of the above, the control gain Cu is set to a large value as compared with during low-speed traveling so as to improve the ground contact of the wheels. .

FIG. 17 is a flowchart showing the third embodiment, and step S20 shown in FIG. 16 of the second embodiment.
The calculation of the control gain Cu performed in step S30 is replaced with step S30 of calculating the control gain Cu according to the turning state of the vehicle, and the other points are the same as those in the first embodiment.

In FIG. 17, a step S30 estimates the turning radius R of the vehicle from the steering angle θ detected by the steering angle sensor 8 and increases the control gain Cu in accordance with the increase of the turning radius R, while The control gain Cu is set to a small value.

Therefore, it is possible to improve the stability of the vehicle during turning by setting so that the grounding property is improved when the vehicle is turning than when the vehicle is straight ahead. Occurrence is suppressed and riding comfort can be improved.

The turning state of the vehicle may be estimated from the steering angular velocity of the steering wheel, or may be estimated from the steering steering angle θ and the vehicle speed V or the steering angular velocity and the vehicle speed V. Although not shown, a yaw rate sensor or a sensor for detecting lateral acceleration (acceleration in the vehicle width direction) may be provided to estimate the turning state of the vehicle from the yaw rate, yaw angular acceleration or lateral acceleration, and lateral speed.

FIG. 18 is a flow chart showing the fourth embodiment, and step S20 shown in FIG. 16 of the second embodiment.
The calculation of the control gain Cu performed in step S4 is replaced with step S40 in which the control gain Cu is calculated according to the unsprung vertical acceleration, and the others are the same as in the first embodiment.

[0088] In FIG. 18, step S40 is unsprung calculated at the step S9 acceleration ZG _1FR ~ZG _1RL
One which increase the control gain Cu with an increase, when the unsprung acceleration ZG _1FR ~ZG _1RL is small, a small input from the road surface, since the vibration of the unsprung is less likely a problem, a small value of control gain Cu set to be to suppress the occurrence of unnecessary damping force, while measuring the ride comfort improvements, because when a large unsprung acceleration ZG _1FR ~ZG _1RL large input from the road surface, to suppress the vibration of the unsprung Therefore, the control gain Cu is set to a large value.

FIG. 19 is a flow chart showing the fifth embodiment, and step S20 shown in FIG. 16 of the second embodiment.
The calculation of the control gain Cu that is performed by the control gain Cu is performed according to the stroke of the shock absorbers 3FR to 3RL.
It is replaced with step S50 for calculating, and the other points are the same as in the first embodiment.

In FIG. 18, a step S50 is the shock absorber 3FR-read in the above step S2.
Stroke of 3RL (= relative displacement H _FR above and below the spring H _FR
.About.H _RL ), the control gain Cu is increased. When the stroke of the shock absorbers 3FR to 3RL is small, the input from the road surface is small and vibration under the spring is less likely to cause a problem. By setting a small value to suppress the generation of unnecessary damping force,
While the ride comfort is improved, the input from the road surface is large when the stroke is large, so the control gain Cu is set to a large value in order to suppress unsprung vibration.

In the above embodiment, the damping coefficients C _FR to _FL of the shock absorbers 3FR to 3RL were changed by changing the oil passages according to the driving of the actuators 7FR to 7RL, which is not shown. , A shock absorber whose damping coefficient can be changed by an electrorheological fluid, etc., can be set to a predetermined damping coefficient in accordance with a command signal corresponding to the damping coefficient C _i .

[0092]

As described above, according to the first aspect of the invention, the target corresponding to the force proportional to the relative velocity HV ₀ between the road surface and the unsprung surface estimated based on the sprung acceleration and the unsprung acceleration of each wheel. The damping coefficient Cu _i is calculated, and by suppressing the excessive fluctuation of the relative displacement between the road surface and the unsprung portion, the vibration on the spring from the sprung resonance frequency to the unsprung resonance frequency is suppressed while improving the grounding property of the wheel. Therefore, it is possible to improve the ground contact property and the riding comfort at the same time.

Further, the second aspect of the invention is that the target damping coefficient C corresponding to the force proportional to the relative velocity HV ₀ between the road surface and the unsprung surface estimated based on the sprung mass acceleration and the unsprung mass acceleration of each wheel.
It is possible to calculate u _{i based} on the detected control gain Cu corresponding to the traveling state and to improve the ground contact property of the wheels and the riding comfort while generating the damping force according to the traveling state. .

A third aspect of the present invention is the control gain Cu.
It is possible to make the damping coefficient small at low speeds and increase the damping coefficient at high speeds to secure riding comfort at low speeds and improve the grounding of the wheels at high speeds. Become.

In the fourth aspect of the invention, the control gain Cu is changed in accordance with the detected turning state, and the damping coefficient is small when the vehicle is traveling straight, and the damping coefficient is increased when the vehicle is traveling to ensure a comfortable ride when traveling straight. It is possible to achieve both improvement of the grounding property of the wheel during turning.

Further, in the fifth aspect of the invention, the control gain Cu is changed according to the detected road surface condition, and the damping coefficient is changed according to the quality of the road surface condition to secure the riding comfort and improve the grounding property of the wheels. It becomes possible to be compatible.

According to the sixth aspect of the invention, the control gain Cu is changed according to the detected unsprung acceleration to suppress the generation of unnecessary damping force when the unsprung acceleration is small and the input from the road surface is small. When the downward acceleration is large and the input from the road surface is large, the unsprung vibration is suppressed to ensure both the riding comfort and the improved grounding of the wheels.

According to the seventh aspect of the invention, the control gain Cu is changed according to the stroke of the detected shock absorber to suppress the generation of unnecessary damping force when the stroke is small and the input from the road surface is small, while the stroke is reduced. When the input from the road surface is large, the unsprung vibration can be suppressed to ensure both riding comfort and improvement of the ground contact of the wheels.

In the eighth aspect of the invention, the relative velocity HV ₀ between the road surface and the unsprung portion can be obtained by differentiating the relative displacement H ₀ between the road surface and the unsprung portion obtained from the sprung mass acceleration and the unsprung mass acceleration.

The ninth aspect of the present invention is the target damping coefficient Cu _i.
Is calculated so as to correspond to the resultant force of the force proportional to the relative velocity HV ₀ between the road surface and the unsprung portion and the force proportional to the sprung portion velocity ZV. It is possible to suppress both the sprung mass vibration at the frequency and the unsprung mass vibration at the unsprung resonance frequency to improve both the grounding property and the riding comfort.

[Brief description of drawings]

FIG. 1 is a perspective view of a vehicle showing an embodiment of the present invention.

FIG. 2 is a block diagram of the same.

FIG. 3 is a schematic diagram of an acceleration sensor.

FIG. 4 is a characteristic diagram of the acceleration sensor.

FIG. 5 is a schematic plan view of a vehicle body showing an arrangement of acceleration sensors.

FIG. 6 is a schematic view of a vehicle height sensor.

FIG. 7 is a characteristic diagram of a vehicle height sensor.

FIG. 8 is a cross-sectional view of a shock absorber.

9 is an enlarged view of part A in FIG.

FIG. 10 shows the relationship between the spool position and the oil passage, (A)
FIG. 4B is a sectional view of the spool, and FIG. 6B is an explanatory diagram showing the relationship between the pressure side oil passage and the circular oil passage.

FIG. 11 is an explanatory diagram showing how the damping force is adjusted.
Shows the state of expansion side; hard, compression side; soft, (B) shows the state of both expansion side and compression side soft, and (C) shows the state of expansion side; soft, compression side; hard.

FIG. 12 is a flowchart showing an example of control.

FIG. 13 shows the relationship between the state of the suspension and the time when passing through unevenness of the road surface. The horizontal axis represents time, the vertical axis represents road surface displacement X ₀ , sprung displacement X ₂ , and road surface and unsprung relative displacement (X ₀ −X ₁ ) And damping force F, the broken line shows the present embodiment and the solid line shows the case where the damping coefficient is fixed.

14A and 14B are graphs showing the frequency response of the suspension, where FIG. 14A shows the sprung displacement / road surface displacement, and FIG. 14B shows the unsprung / road surface displacement. The broken line shows the present embodiment and the solid line shows the damping coefficient. The case where they are fixed is shown.

FIG. 15 is a graph showing the frequency response of the suspension to which the skyhook damper control is added, (A) shows the sprung displacement / road surface displacement, (B) shows the unsprung / road surface displacement, and the broken line shows the present embodiment. The solid line shows the case where the skyhook damper control is added, and the solid line shows the case where the damping coefficient is fixed.

FIG. 16 is a flowchart showing an example of control according to the second embodiment.

FIG. 17 is a flowchart showing an example of control according to the third embodiment.

FIG. 18 is a flowchart showing an example of control according to the fourth embodiment.

FIG. 19 is a flowchart showing an example of control according to the fifth embodiment.

FIG. 20 is a ¼ vehicle model of the suspension.

FIG. 21 shows the relationship between the state of the suspension and the time when passing through uneven road surface according to the conventional example, where the horizontal axis represents time, the vertical axis represents road surface displacement X ₀ , sprung displacement X ₂ , and road surface and unsprung relative displacement (X ₀ −
X ₁ ) and the damping force F, the broken line shows the case where the damping coefficient is controlled, and the solid line shows the case where the damping coefficient is fixed.

FIG. 22 is a graph showing a frequency response of a suspension according to a conventional example, (A) shows sprung displacement / road displacement,
(B) shows the unsprung / road surface displacement, the broken line shows the case where the damping coefficient is controlled, and the solid line shows the case where the damping coefficient is fixed.

FIG. 23 is a claim correspondence diagram corresponding to any one of the first to ninth inventions.

[Explanation of symbols]

 2FR to 2RL Wheels 3FR to 3RL Shock absorber 5FR to 5RL Vehicle height sensor 6FR to 6R Acceleration sensor 7FR to 7FR Actuator 8 Steering angle sensor 9 Vehicle speed sensor 10 Controller 30 Control rod 33A, 33B, 33C Oil passage 34 Spool 35 Pressure side oil passage 36 Pressure side oil passage 37 Extension side oil passage 38 Extension side oil passage 50 Damping coefficient changing means 51 Sprung acceleration detection means 52 Unsprung acceleration detection means 53 Control means 54 Road surface-Unsprung relative speed estimation means 55 Target damping coefficient calculation means 56 Relative Displacement calculating means 57 Differentiating means 58 Sprung speed calculating means 60 Running state detecting means 61 Vehicle speed detecting means 62 Turning detecting means 63 Road surface state detecting means 65 Stroke detecting means

Claims (9)

[Claims] 1. A shock absorber having a damping coefficient changing means for changing a damping coefficient, which is interposed between an unsprung portion and an unsprung portion of each wheel, and a vertical acceleration of a vehicle body on the spring of each wheel is detected. Sprung acceleration detection means, unsprung acceleration detection means for detecting the unsprung acceleration of each wheel in the vertical direction of the vehicle body, and target damping of the shock absorber based on the detection values of these sprung and unsprung acceleration detection means In a suspension control device for a vehicle, which calculates a coefficient and which controls the damping coefficient changing means so that the damping coefficient of the shock absorber matches the target damping coefficient, the control means controls the sprung acceleration and road surface estimating the relative velocity HV ₀ under the road surface and the spring from the unsprung acceleration - a lower relative velocity estimating means spring is proportional to the road surface and the relative velocity of the unsprung Target damping coefficient Cu _i corresponding to the force
A suspension control device for a vehicle, comprising: a target damping coefficient calculation means for calculating 2. A shock absorber having a damping coefficient changing means for changing a damping coefficient, which is interposed between an unsprung portion and an unsprung portion of each wheel, and a vertical acceleration of the vehicle body on the spring of each wheel is detected. Sprung acceleration detection means, unsprung acceleration detection means for detecting the unsprung acceleration of each wheel in the vertical direction of the vehicle body, and target damping of the shock absorber based on the detection values of these sprung and unsprung acceleration detection means In a suspension control device for a vehicle, which calculates a coefficient and has a control means for driving the damping coefficient changing means so that the damping coefficient of the shock absorber matches the target damping coefficient, a running state detection for detecting a running state of the vehicle means and said control means is road estimates the relative velocity HV ₀ under the road surface and the spring from the sprung acceleration and the unsprung acceleration - unsprung relative speed estimation Means and the target damping coefficient Cu _i corresponding to the road surface and the force proportional to the relative velocity of the unsprung and a target damping coefficient calculation means for calculating, based on the control gain Cu corresponding to the running state of the detected vehicle A suspension control device for a vehicle, comprising: 3. The running state detecting means is constituted by a vehicle speed detecting means for detecting a vehicle speed, and the target damping coefficient calculating means changes the control gain Cu according to the vehicle speed. The vehicle suspension control device according to. 4. The traveling state detecting means is constituted by a turning detecting means for detecting a turning state of the vehicle, and the target damping coefficient calculating means is adapted to control the gain C according to the turning state.
The suspension control device for a vehicle according to claim 2, wherein u is changed. 5. The running condition detecting means is constituted by a road condition detecting means for detecting a condition of a road surface, and the target damping coefficient calculating means calculates the control gain C according to the road condition.
The suspension control device for a vehicle according to claim 2, wherein u is changed. 6. The road surface condition detecting means is constituted by means for detecting an unsprung acceleration in the vertical direction of the vehicle body, and the target damping coefficient calculating means changes the control gain Cu according to the unsprung acceleration. 6. The method according to claim 5, wherein
The vehicle suspension control device according to. 7. The road surface condition detecting means is constituted by means for detecting a stroke of a shock absorber, and the target damping coefficient calculating means changes the control gain Cu according to the stroke. Item 5. The vehicle suspension control device according to item 5. 8. The road surface-unsprung relative speed estimation means comprises:
2. A relative displacement calculating means for calculating a relative displacement H ₀ between the road surface and the unsprung portion from the sprung acceleration and the unsprung acceleration, and a differentiating means for differentiating the relative displacement H _0. The suspension control device for a vehicle according to claim 7. 9. The target damping coefficient calculation means includes sprung speed calculation means for calculating a speed ZV in the vehicle body vertical direction on the spring based on the sprung acceleration, and a relative speed HV ₀ between the road surface and the unsprung surface. 9. The vehicle suspension according to claim 1, wherein a target damping coefficient Cu _i corresponding to a resultant force of a force proportional to the sprung speed ZV and a force proportional to the sprung speed ZV is calculated. Control device.