JPH11151923A - Damping force control device for vehicle - Google Patents

Abstract

PROBLEM TO BE SOLVED: To enhance the controllability of a vehicle by controlling the damping force of a hydraulic shock absorbing means on the basis of the grounding load detected by a grounding load detecting means. SOLUTION: In the damping force control of a shock absorber by means of an ECU 230, first, the detection results of a steering angle sensor 203, a vehicle speed sensor 204, a vertical acceleration sensor 205, and load sensors 206 are read. While the vehicle is turned, and when the grounding load is increased as compared with the preceding routine, the damping force is reduced by subtracting 1 from the control step number for setting the stopping position of a stepping motor being the driving source of an actuator 2 in a shock absorber. When the grounding load is reduced as compared with the preceding routine, the damping force is increased by adding 1 to the control step number. Thus, a force acts in the direction the movement of a load, the reduction in the grounding load in the turning inner wheel is prevented, and the increase in the grounding load in the turning outer wheel is prevented to improve the controllability.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a damping force control device for a vehicle for controlling a damping force of a hydraulic shock absorber (shock absorber) disposed between a sprung member and a unsprung member of a vehicle.

[0002]

2. Description of the Related Art An example of a conventional damping force control device is disclosed in, for example, Japanese Patent Laid-Open No. 5-294122. This damping force control device employs the so-called skyhook control theory to control the damping force of the hydraulic pressure damping device. The relative speed between the sprung member and the unsprung member of the vehicle and the vertical direction of the sprung member are controlled. The damping force is controlled based on the speed, thereby increasing the effect of suppressing the vertical vibration of the sprung member based on the vertical input from the road surface.

[0003] In addition, in order to ensure the riding comfort of the vehicle,
It is desirable that the damping force be small in a region where the displacement speed of the piston rod that constitutes the hydraulic shock absorber is low. On the other hand, in order to ensure good steering stability of the vehicle, it is desirable that the damping force be large in a region where the displacement speed of the piston rod is high.
Generally, the displacement speed of the piston rod is 0.1 to 0.3.
It is known that at a point exceeding m / s, the riding comfort and the steering stability can be made compatible by reducing the gradient of the damping force characteristic (characteristic indicating the relationship between the displacement speed of the piston rod and the damping force). ing.

[0004]

As a result of research, it has been found that the riding comfort and the steering stability of a vehicle are limited to a region where the displacement speed of the piston rod is extremely low (for example, a region of 0.02 m / s or less; It is also found that it greatly depends on the damping force characteristic in the above.

However, such a large change in the riding comfort and the driving stability of the vehicle depending on the damping force characteristics in the low speed range has not been recognized so far, and the riding comfort and the steering stability of the vehicle have not been recognized. It is important to control the damping force characteristic in the low-speed region in order to improve the damping force.

Accordingly, an object of the present invention is to provide a damping force control device for a vehicle which can further improve the controllability of the vehicle by controlling the damping force characteristic in the low speed range.

[0007]

Accordingly, a vehicle damping force control device according to a first aspect of the present invention damps vibration generated between a sprung member and a unsprung member of a vehicle with a predetermined damping force, Hydraulic buffer means having a variable damping force function in a region where the relative speed between the upper member and the unsprung member is low, a contact load detecting means for detecting a contact load on the wheel, and a contact load detected by the contact load detecting means And control means for controlling the damping force of the hydraulic pressure relief means.

[0008] During turning, the load on the vehicle moves and the contact load on the turning inner wheel decreases. However, when the contact load decreases, the cornering force generated on the turning inner wheel decreases and the turning performance deteriorates. Becomes Further, the ground load of each wheel can be increased or decreased by changing the damping force of the corresponding hydraulic pressure buffer. Therefore, the control unit controls the damping force of each hydraulic pressure buffer unit based on the detection result of the ground load detection unit, and controls the increase and decrease of the ground load on each wheel to improve the maneuverability of the vehicle. Let it. According to a second aspect of the present invention, in the control device according to the first aspect, the damping force of the hydraulic buffer is reduced when the contact load increases, and the hydraulic buffer is reduced when the contact load decreases. Is controlled to increase the damping force.

[0009] By reducing the damping force of the hydraulic buffer, an increase in the ground contact load is suppressed in the wheel provided with the hydraulic buffer. In addition, by increasing the damping force of the hydraulic buffer, a decrease in the ground contact load of the wheel provided with the hydraulic buffer is suppressed.

According to a third aspect of the present invention, the grounding load detecting means estimates a grounding load based on an internal pressure of a tire mounted on a corresponding wheel.

When the contact load of the wheel increases, the internal pressure of the tire mounted on the wheel increases, and conversely, when the contact load decreases, the internal pressure of the tire decreases. Therefore, the contact load detecting means estimates the contact load based on the internal pressure of the tire mounted on the corresponding wheel.

According to a fourth aspect of the present invention, there is provided a vehicle damping force control device, wherein the ground contact load detecting means estimates a ground contact load based on a turning radius of a tire mounted on a corresponding wheel.

When the ground contact load of a wheel increases during running, the tire mounted on the corresponding wheel is flattened and the turning radius of the tire decreases, and conversely, when the contact load decreases, the turning radius of the tire increases. Therefore, the contact load detecting means estimates the contact load based on the turning radius of the tire mounted on the corresponding wheel.

According to a fifth aspect of the present invention, in the vehicle damping force control device, the ground contact load detecting means according to the third or fourth aspect includes a wheel speed detecting means for detecting a wheel speed of a corresponding wheel.

The tire internal pressure and the tire turning radius can be estimated by detecting the rotation speed of the wheel on which the tire is mounted. Therefore, the contact load detecting means includes a wheel speed detecting means, and estimates the internal pressure of the tire or the turning radius of the tire based on the detection result.

According to a sixth aspect of the present invention, there is provided a vehicle damping force control apparatus for damping a vibration generated between a sprung member and a unsprung member of a vehicle with a predetermined damping force, and further comprising a damping force for the sprung member and the unsprung member. A hydraulic buffer having a damping force variable function in a region where the relative speed is low, and a displacement state detecting means disposed corresponding to the hydraulic pressure buffer and detecting a relative displacement between the sprung member and the unsprung member. On the basis of the detection result of the displacement state detecting means, the damping force of the hydraulic buffer means is increased when the hydraulic buffer means is displaced in the contracting direction, and the hydraulic pressure is increased in the situation where the hydraulic buffer means is displaced in the extending direction. Control means for reducing the damping force of the buffer means.

Due to the load movement of the vehicle, the hydraulic pressure buffer located on the turning inner wheel side is displaced in the elongating direction and the ground contact load of the wheel is reduced. And the ground contact load of the wheel increases. Therefore, the damping force of the hydraulic buffer is increased by the control unit when the hydraulic buffer is displaced in the contracting direction, and the damping force of the hydraulic buffer is decreased when the hydraulic buffer is displaced in the extending direction. Control is performed so that the As a result, the load movement toward the turning outer wheel is suppressed, and the turning inner wheel easily follows the road surface. As a result, the increase in the contact load on the turning outer wheel is suppressed, and the decrease in the contact load on the turning inner wheel is suppressed. .

[0018]

BEST MODE FOR CARRYING OUT THE INVENTION Hereinafter, embodiments of the present invention will be described.
This will be described with reference to the accompanying drawings.

FIG. 1 schematically shows a vehicle equipped with the damping force control device according to the first embodiment. Wheels L and R are connected to the left and right sides of the vehicle body 200 via suspension arms 210, respectively. A shock absorber 1 is provided between the vehicle body 200 and the left and right suspension arms 210 to attenuate vertical vibration generated in the vehicle body 200.
0 is arranged. The shock absorber 10 includes an actuator 2 as described later, and has a mechanism capable of adjusting a generated damping force by controlling the drive of the actuator 2.

The steering shaft 2 connected to the steering handle 201
In 02, a steering angle sensor 203 is provided, and the operation amount of the steering wheel 201 is detected based on the detection result of the steering angle sensor 203. A load sensor 206 is provided between each of the shock absorbers 10 and the coil spring 10c and the vehicle body 200, and detects a ground contact load of each wheel based on a detection result of the load sensor 206 (see FIG. 3). . A wheel speed sensor 208 for detecting the rotation state of the wheel is provided corresponding to each of the wheels L and R.
Then, various calculations such as the wheel speed are performed based on the detection result of the wheel speed sensor 208.

A stroke sensor 207 is provided corresponding to each wheel. The sensor main body is fixed to the vehicle body 200, and the tip of a detection rod 207a extending from the sensor main body is connected to the suspension arm 210. The detection rod 207a moves with the swing of the suspension arm 210.
Swings, and the movement of the detection rod 207a is detected by a potentiometer built in the sensor body. Therefore, the stroke sensor 207 detects the relative stroke amounts of the wheels L and R with respect to the vehicle body 200. Since the connection portion of the detection rod 207a to the suspension arm 210 is located near the center of the vehicle body 200, the actual relative stroke amount of the wheel is a constant multiple of the detection result of the stroke sensor 207. Therefore, a value obtained by multiplying the detection result of the stroke sensor 207 by a predetermined constant is obtained as the relative stroke amount Y of the wheel.

FIG. 1 schematically shows only the configuration of the front wheels, but the wheels L, R, the shock absorber 10, the suspension arm 210, the stroke sensor 207, etc. have the same configuration on the rear wheels. . Then, as shown in FIG. 2, an electronic control unit (hereinafter, referred to as “ECU”).
Reference numeral 230 denotes a steering angle sensor 203, a stroke sensor 207 corresponding to each wheel, a wheel speed sensor 208, a vehicle speed sensor 204 for detecting a vehicle speed, and a vertical acceleration acting on the vehicle body 200 as a sprung member. A detection result from each sensor such as the vertical acceleration sensor 205 is given. The ECU 230 executes a calculation process described later based on these detection results, and based on the calculation results,
The drive control of the actuator 2 in the shock absorber 10 corresponding to each wheel is performed.

Here, the shock absorber 10 will be described. As shown in FIG. 3, the shock absorber 10
Has a piston rod 16 and an outer cylinder 18. A guide 10a is fixed to the outer periphery of the outer cylinder 18, and a bracket 10b is hung on the upper end of the piston rod 16. The coil spring 10c described above is disposed between the guide 10a and the bracket 10b.
The body 200 is elastically supported by the coil spring 10c.

Inside the outer cylinder 18, the inner cylinder 20 is
And are arranged coaxially. An annular chamber 21 is formed between the outer cylinder 18 and the inner cylinder 20. A rod guide 22 is fitted into the upper end of the outer cylinder 18. The rod guide 22 is a columnar rigid member having a large diameter portion 22a and a small diameter portion 22b. The outer peripheral surface of the small diameter portion 22b is engaged with the inner peripheral surface of the inner cylinder 20, and the outer peripheral surface of the large diameter portion 22a is engaged with the inner peripheral surface of the outer cylinder 18. The rod guide 22 has a through hole at the center thereof. The piston rod 16 is slidably inserted through the through hole in a liquid-tight manner. A cap 24 is fixed to the upper end of the outer cylinder 18 so that the piston rod 16 passes through the center of the cap 24.

The piston rod 16 is a columnar member whose lower end has a small diameter. The piston rod 16 is arranged so that its small diameter portion is housed inside the inner cylinder 20. A rebound stopper 26 and a rebound stopper plate 28 are mounted on the piston rod 16 at positions accommodated inside the inner cylinder 20.

The rebound stopper plate 28 is an annular rigid member, and is fixed to the outer periphery of the piston rod 16. The rebound stopper 26 is an annular member having elasticity, and is mounted on an upper portion of the rebound stopper plate 28. When the piston rod 16 is displaced upward by a predetermined distance, the rebound stopper 26 comes into contact with the rod guide 22, and further displacement of the piston rod 16 is restricted.

A sub-piston 30 and a main piston 32 are fixed to a lower end portion of the piston rod 16, and the sub-piston 30 and the main piston 32 are attached in this order from above. The inner space of the inner cylinder 20 is formed by the sub piston 30 and the main piston 32 by the sub piston 30.
The upper chamber 34 is partitioned into an upper chamber 34, a middle chamber 36 between the sub piston 30 and the main piston 32, and a lower chamber 38 below the main piston 32.

Sub piston 30 and main piston 32
Are provided with orifices and valve mechanisms that allow fluid to flow between the upper chamber 34 and the middle chamber 36 and between the middle chamber 36 and the lower chamber 38, respectively, in accordance with the advance and retreat of the piston rod 16. Generates damping force. Details of the configuration of the sub piston 30 and the main piston 32 will be described later.

A passage 40 penetrating in the axial direction is provided inside the piston rod 16. Passage 40
A large-diameter portion 40a and a small-diameter portion 4 extending below the large-diameter portion 40a
0bb. A step 40c is formed at the boundary between the large diameter portion 40a and the small diameter portion 40b of the passage 40. The large diameter portion 40a of the passage 40 has an adjustment rod 42
Is inserted.

A base valve 41 is fixed to a lower end of the outer cylinder 18. The base valve 41 is configured to allow a fluid to flow between the lower chamber 38 and the annular chamber 21. The working fluid is contained inside the outer cylinder 18 so as to fill the inner space of the inner cylinder 20 and fill the annular chamber 21 to a predetermined height.

The upper end of the adjusting rod 42 reaches the upper part of the piston rod 16 and is engaged with the actuator 2 mounted on the vehicle body 200. Actuator 2
Is provided with, for example, a stepping motor as a drive source, and adjusts the adjustment rod 42 according to a signal from the ECU 230.
To rotate.

Next, referring to FIG.
The configuration of the main piston 32, the main piston 32, and its peripheral parts will be described. FIG. 4 is an enlarged view of the sub-piston 30, the main piston 32, and a peripheral portion thereof. The left half of FIG. 4 shows a component that allows fluid to flow from the upper chamber 34 to the lower chamber 38, and the right half of FIG. Fig. 2 shows components that allow fluid to flow through.

As shown in FIG. 4, the adjusting rod 42 is one of the damping force variable mechanisms, and includes a small-diameter portion 42a having an outer diameter smaller than the inner diameter of the large-diameter portion 40a of the passage 40, and a small-diameter portion 42a. And a conical portion 42b formed at a lower end portion. The adjusting rod 42 is configured such that the tip of the conical portion 42 b is
It is arranged so as to enter the small-diameter portion 40b of zero. A clearance C is formed between the outer peripheral surface of the conical portion 42b and the step 40c of the passage 40.

An O-ring 43 is mounted on the outer periphery of the adjusting rod 42 above the small diameter portion 42a. The O-ring 43 defines an annular communication space 44 between the outer periphery of the small diameter portion 42 a of the adjustment rod 42 and the inner periphery of the large diameter portion 40 a of the passage 40. The communication space 44 communicates with the internal space of the small diameter portion 40b of the passage 40 via the clearance C.

A communication passage 46 extending in the radial direction of the piston rod 16 and communicating the upper chamber 34 with the communication space 44.
Is provided. Further, the piston rod 16 is provided with a communication passage 47 extending in the radial direction and communicating the internal space of the small diameter portion 40 b of the passage 40 with the middle chamber 36.

The adjusting rod 42 is screwed into a large diameter portion 40a of the passage 40 at a screw portion (not shown), and the upper end thereof is engaged with the actuator 2. Therefore, the clearance C can be adjusted by rotating the adjustment rod 42 by the actuator 2 to change the opening position (up and down position) of the adjustment rod 42.

On the outer periphery of the small diameter portion of the piston rod 16, a stopper plate 48, a leaf seat 49, and a leaf valve 5 are arranged in this order from the large diameter portion 16a side (the upper side in FIG. 4).
0, the sub piston 30, the leaf valve 54, and the leaf seat 56 are fitted.

The leaf valves 50 and 54 are members having a low bending rigidity made of a thin plate. Annular grooves 58 and 60 are provided on the upper end surface and the lower end surface of the sub piston 30, respectively. Leaf valves 50 and 54
Are arranged to close the annular grooves 58 and 60, respectively. The sub-piston 30 has an annular groove 58.
A through passage 62 that communicates the inner space of the inner chamber with the middle chamber 36 and a through passage 6 that communicates the inner space of the annular groove 60 with the upper chamber 34
4 are provided.

The leaf valve 50 opens by flexing and deforming when the hydraulic pressure of the middle chamber 36 becomes higher than the hydraulic pressure of the upper chamber 34 by a predetermined valve opening pressure P1. To the upper chamber 34 from the working fluid. The leaf valve 54 opens by flexing and deforming when the hydraulic pressure of the upper chamber 34 becomes higher than the hydraulic pressure of the middle chamber 36 by a predetermined valve opening pressure P2. The flow of the working fluid toward the middle chamber 36 is allowed.

A piston ring 66 is mounted on the outer periphery of the sub piston 30. The seal between the sub-piston 30 and the inner cylinder 20 is ensured by the piston ring 66. Leaf sheet 5 around piston rod 16
6, a communication member 68, a leaf seat 70, a spacer 72, a spring seat 74,
And a spacer 76 are fitted.

The communication member 68 has a communication passage 77 which penetrates in the radial direction and communicates the communication passage 47 of the piston rod 16 with the intermediate chamber 36. A spring seat 78 is fitted on the outer periphery of the spacer 76 so as to be slidable in the axial direction. Spring seat 74 and spring seat 7
8, a spring 80 is provided.

The spacer 76 on the outer periphery of the piston rod 16
A leaf valve 82, a main piston 32, and a leaf valve 86 are fitted in order from the upper side further below. A plurality of seat surfaces 92 are provided on the upper end surface of the main piston 32. Also, the main piston 32
A plurality of sheet surfaces 94 are provided at positions not corresponding to the sheet surface 92 at the lower end surface of the. The leaf valves 82 and 86 are members formed by laminating a plurality of thin plate members, and are disposed so as to contact the top surfaces of the seat surfaces 92 and 94, respectively. A piston ring 95 is mounted on the outer periphery of the main piston 32. Piston ring 95
Thereby, the sealing property between the main piston 32 and the inner cylinder 20 is ensured.

The main piston 32 is provided with through passages 96 and 98 penetrating in the axial direction.
The through-passage 96 is configured to open at a concave portion between the seat surfaces 92 at an upper end thereof, and to open at a top surface of the seat surface 94 at a lower end thereof. In addition, the through passage 9
Numeral 8 is configured to open at the top of the seat surface 92 at its upper end and to open into the recess between the seat surfaces 94 at its lower end.

In a state where the leaf valve 82 is in contact with the seat surface 92, a first communication between the through passage 98 and the middle chamber 36 is established.
An orifice (not shown) is formed and leaf valve 86
Is in contact with the seat surface 94 and the through passage 96 and the lower chamber 3
A second orifice (not shown) for communicating with the second orifice 8 is formed.

A spacer 99 is fitted below the leaf valve 86 on the outer periphery of the piston rod 16.
A thread 16c is formed at the lower end of the piston rod 16, and a spring seat 100 is screwed to the thread 16c. A spring seat 102 is fitted on the outer periphery of the spacer 99 so as to be slidable in the axial direction. Spring seat 102 and spring seat 10
A spring 104 is provided between the spring 104 and the spring.

At the lower end of the small diameter portion of the piston rod 16, a screw 105 for closing the passage 40 is mounted. Therefore, communication between the passage 40 and the lower chamber 38 is interrupted, and the passage 40 communicates only with the upper chamber 34 and the middle chamber 36.

The member disposed on the outer periphery of the small diameter portion at the lower portion of the piston rod 16 is pressed by the spring seat 100 toward the step surface at the boundary between the large diameter portion 16a and the small diameter portion, whereby the piston rod 16 is pressed. 16 are integrally fixed.

The leaf valves 82 and 86 are pressed toward the top surfaces of the seat surfaces 92 and 94 of the main piston 32 by the urging forces of springs 80 and 104, respectively. The leaf valve 82 adjusts the hydraulic pressure of the lower chamber 38 to the middle chamber 3.
When the pressure becomes higher than the predetermined valve opening pressure P3 as compared with the hydraulic pressure of No. 6, the valve is opened by flexing and deforming upward against the urging force of the spring 80, and the operation is performed from the lower chamber 38 to the middle chamber 36. Allow fluid flow. The leaf valve 86 has a predetermined valve opening pressure P that is higher than the hydraulic pressure of the middle chamber 36 compared to the hydraulic pressure of the lower chamber 38.
When the pressure becomes 4 or more, the valve opens by flexing and deforming downward against the urging force of the spring 104, allowing the flow of the working fluid from the middle chamber 36 to the lower chamber 38.

In this embodiment, since the leaf valves 50 and 54 are formed of low rigidity thin plate members,
These valve opening pressures P1 and P2 are set to very small values. On the other hand, since the leaf valves 82 and 86 are pressed by the springs 80 and 104, respectively, the valve opening pressures P3 and P4 are set to relatively large values.

Next, the operation of the shock absorber 10 will be described. FIG. 5 shows a damping force characteristic realized by the shock absorber 10. In FIG. 5, the horizontal axis shows the displacement speed V of the piston rod 16, and the vertical axis shows the damping force F generated by the shock absorber 10. In addition,
In FIG. 5, the damping force F when the piston rod 16 is displaced in the direction in which the piston rod 16 retreats from the inner cylinder 20, that is, in the extension direction, is shown as positive.

When the piston rod 16 is displaced in the extending direction, the volume of the upper chamber 34 decreases and the volume of the lower chamber 38 increases. In order to compensate for these volume changes, the working fluid flows from the upper chamber 34 through the middle chamber 36 to the lower chamber 38. Further, when the piston rod 16 retreats from the inner cylinder 20, the volume of the inner cylinder 20 increases. In order to compensate for the increase in the volume of the inner cylinder 20, the working fluid flows from the annular chamber 21 into the lower chamber 38 via the base valve 41.

When the displacement speed V of the piston rod 16 is sufficiently low, the differential pressure between the upper chamber 34 and the middle chamber 36 and the differential pressure between the middle chamber 36 and the lower chamber 38 are small, and the leaf valve Both the valve 54 and the leaf valve 86 are kept closed. Therefore, the working fluid in the upper chamber 34 flows through the communication passage 46 of the piston rod 16, the communication space 44, the clearance C, the small diameter portion 40 b of the passage 40, the communication passage 47, and the communication passage 77 of the communication member 68. The air flows into the intermediate chamber 36 through a road (hereinafter, referred to as a bypass passage). The working fluid in the middle chamber 36 is supplied to the through passage 9 of the main piston 32.
6 and into the lower chamber 38 through the second orifice.

In this case, when the working fluid flows through the bypass passage and the second orifice, a damping force is generated due to the flow resistance. The damping force F exerted by the shock absorber 10 includes a damping force Fa generated according to the flow resistance R1 when the working fluid flows from the upper chamber 34 to the middle chamber 36, and a damping force F generated from the middle chamber 36 to the lower chamber 38. This is the sum with the damping force Fb generated according to the flow resistance R2 when flowing. For this reason. As shown by reference numeral A1 in FIG. 5, the damping force F rises with a large gradient as the displacement speed V increases.

When the flow resistance R1 when the working fluid flows from the upper chamber 34 to the middle chamber 36 increases, the upper chamber 34 and the middle chamber 36
And the pressure difference between them rises. Further, when the flow resistance R2 when the working fluid flows from the middle chamber 36 to the lower chamber 38 increases,
The pressure difference between the middle chamber 36 and the lower chamber 38 increases. And
When the displacement speed V increases until the pressure difference between the upper chamber 34 and the middle chamber 36 reaches the valve opening pressure P2 of the leaf valve 54, the leaf valve 54 opens. Hereinafter, the displacement speed V of the piston rod 16 when the leaf valve 54 is opened and the damping force F generated by the shock absorber 10 will be referred to as a first force, respectively.
These are referred to as a valve opening speed V1 and a first valve opening damping force F1. As described above, the first valve opening damping force F1 has a very small value, for example,
The valve opening pressure P2 of the leaf valve 54 is set sufficiently small so as to be 3 to 5 kgf. When the valve opening pressure P2 of the leaf valve 54 is set as described above, the first valve opening speed V1 becomes a very low speed of 0.05 m / s or less.

When the leaf valve 54 opens, the upper chamber 34
The movement of the working fluid from to the middle chamber 36 is performed through the through passage 64 together with the bypass passage. Therefore, the flow resistance R1 when the working fluid flows from the upper chamber 34 to the middle chamber 36 decreases. Then, as the distribution resistance R1 decreases, as indicated by reference numeral A2 in FIG.
In a region where the displacement speed V exceeds the first valve opening speed V1, the increasing gradient of the damping force F decreases.

When the displacement speed V further increases and the pressure difference between the middle chamber 36 and the lower chamber 38 reaches the valve opening pressure P4 of the leaf valve 86, the leaf valve 86 opens. Hereinafter, the displacement speed V and the damping force F when the leaf valve 86 opens are referred to as a second valve opening speed V2 and a second valve opening damping force F2, respectively. In the present embodiment, the second valve opening damping force F2 is, for example, 50 k.
gf, the valve opening pressure P4 of the leaf valve 86
Is set. In this case, the second valve opening speed V2 is 0.2
The value is about m / s.

When the leaf valve 86 is opened, the middle chamber 36
By increasing the flow passage area of the flow passage from the lower chamber 38 to the lower chamber 38,
The flow resistance R2 when the working fluid flows from the middle chamber 36 to the lower chamber 38 is reduced. For this reason, FIG.
In the region where the displacement speed V exceeds the second valve opening speed V2, the increasing gradient of the damping force F further decreases as shown by.

On the other hand, when the piston rod 16 is displaced in the direction of entering the inner cylinder 20, that is, in the contracting direction, the capacity of the upper chamber 34 increases and the capacity of the lower chamber 38 decreases. In order to compensate for these volume changes, the working fluid flows from the lower chamber 38 to the upper chamber 34 via the middle chamber 36. Further, the volume of the inner cylinder 20 is reduced by the piston rod 16 entering the inner cylinder 20. In order to compensate for the decrease in the volume of the inner cylinder 20, the working fluid is supplied from the lower chamber 38 to the base valve 4.
It flows out into the annular chamber 21 through 1.

In this embodiment, the valve opening pressure P1 of the leaf valve 50 is provided so as to substantially coincide with the valve opening pressure P2 of the leaf valve 54. Therefore, when the displacement speed V reaches v1 substantially equal to the first valve opening speed V1, and the damping force F becomes f1 substantially equal to the first valve opening damping force F1,
The leaf valve 50 opens. Also, the leaf valve 82
The valve opening pressure P3 is set to be slightly smaller than the valve opening pressure P4 of the leaf valve 86. Therefore, the displacement speed V reaches v2 (for example, about 0.15 m / s) smaller than the second valve opening speed V2, and the damping force F becomes f2 (for example, about 30 kgf) smaller than the second valve opening damping force F2. At this point, the leaf valve 82 opens. Note that
The displacement speeds v1 and v2 of the piston rod 16 when the leaf valves 50 and 82 open are also the first speeds, respectively.
F1 and f, which are referred to as a valve opening speed and a second valve opening speed, and are damping forces F when the leaf valves 50 and 82 open.
2 is also referred to as a first valve opening damping force and a second valve opening damping force, respectively.

Accordingly, even when the piston rod 16 is displaced in the contracting direction, similarly to the case where the piston rod 16 is displaced in the extending direction, the displacement speed V of the piston rod 16 reaches the first valve opening speed v1. 5, the damping force F rises with a relatively large gradient, as indicated by reference numeral B1 in FIG. Then, the displacement speed V becomes equal to the first valve opening speed v1.
, The leaf valve 50 is opened, and FIG.
, The increasing gradient of the damping force F decreases. Further, when the displacement speed V reaches the second valve opening speed v2, the leaf valve 82 is opened, so that the increasing gradient of the damping force F is further reduced as indicated by reference numeral B3 in FIG.

As described above, according to the shock absorber 10, the displacement speed V of the piston rod 16 is limited to the low speed range (first range).
Damping such that the increasing gradient of the damping force F gradually decreases in accordance with the transition from the valve opening speed V1, v1 or lower region to the high speed region (region exceeding the first valve opening speed V1, v1). Force characteristics are realized.

Incidentally, the degree of opening of the bypass passage changes according to the size of the clearance C. As the opening degree of the bypass passage increases, the flow resistance of the working fluid flowing through the bypass passage decreases. When the flow resistance when flowing through the bypass passage decreases, the differential pressure between the upper chamber 34 and the middle chamber 36 generated for a constant displacement speed V decreases, and the damping force F decreases. That is, the gradient of the damping force characteristic is small as shown by the broken lines with the reference numerals a1 and b1 in FIG.

Therefore, by adjusting the clearance C, the displacement speed V of the piston rod 16 becomes equal to the first valve opening speed V.
1, while the damping force characteristic in a region larger than v1, that is, a high-speed region is maintained substantially constant, the first valve opening speed V
It is possible to change the damping force characteristic at 1, v1 or less. As described above, the first valve opening speed V1, v1 is 0.05
It is set to a low value of m / s or less. Therefore, according to the shock absorber 10 according to the present embodiment, by changing the clearance C, the damping of the shock absorber 10 in the low-speed region of 0.05 m / s or less without affecting the damping force characteristics in the high-speed region. Only the force characteristics can be adjusted. Further, by controlling the driving of the actuator 2 to change the clearance C stepwise, it is possible to change the gradient of the damping force characteristic of the shock absorber 10 stepwise in the low speed range of the piston rod 16.

The shock absorber 10 according to the present embodiment
According to the experiment performed using the vehicle, it is known that the riding comfort and the steering stability of the vehicle greatly change depending on the damping force characteristics in a low speed range. For example, if the clearance C is reduced to increase the gradient of the damping force characteristic in the low speed range, the steering holding force at the time of cornering increases, and the feeling of steering response increases. In addition, the responsiveness of the rolling speed of the vehicle during turning and the responsiveness of the yawing change of the vehicle during steering are sensitive to changes in the damping force characteristics in the low speed range. Therefore, according to the shock absorber 10 according to the present embodiment, by adjusting the clearance C and changing the damping force characteristic in the low speed range, more optimal riding comfort and steering stability can be obtained. In each of the embodiments described below, EC
In U230, the damping force characteristic in this low speed range is mainly controlled.

Here, a first embodiment of the damping force change control (damping force change routine) of the shock absorber 10 executed by the ECU 230 will be described with reference to the flowchart of FIG. Note that the flowchart of FIG. 6 representatively shows the damping force change control of the shock absorber 10 corresponding to the wheel L in FIG. 1, and the same damping force control is performed for the other three wheels. .

The routine shown in FIG. 6 is repeatedly executed from when an ignition switch (not shown) is turned on until it is turned off.

First, in step 102 (hereinafter, the step is referred to as “S”), the steering angle sensor 203 and the vehicle speed sensor 20 are used as detection results of the sensors indicating the running state of the vehicle.
4. Read the detection results of the vertical acceleration sensor 205 and the load sensor 206.

At S104, the control step number step set in the previous routine is read. The control step number will be described later at S112.

In S106, the value of the ground contact load of the wheel, which is the detection result of the load sensor 206, is set as fn.
In subsequent S108, the contact load detected in the previous routine is set as fn-1.

At S110, it is determined whether the vehicle is traveling straight. This determination is performed, for example, based on the detection result of the steering angle sensor 203, calculates the steering angle θ from the detection result of the steering angle sensor 203 read in S102, and calculates the value of the calculated steering angle θ. A predetermined steering determination constant is compared. When the steering angle θ is equal to or smaller than the steering determination constant, it is determined that the vehicle is traveling straight. In step S110, the determination may be made based on the yaw rate applied to the vehicle.

If "Yes" is determined in S110, the process proceeds to S112 and normal skyhook control is performed.

In step S112, a relative stroke speed (hereinafter referred to as a stroke speed) Vi of the wheel L is calculated based on the relative stroke amount Y of the wheel L detected by the stroke sensor 207, and the vertical acceleration sensor 205 The sprung speed Zd, which is the speed in the vertical direction of the vehicle body 200, is calculated from the detection result. Then, based on the speed ratio Zd / Vi between the stroke speed Vi and the sprung speed Zd and the vehicle speed read in S102, the control step number step is set based on the skyhook control theory. The control step number step is used to set a stop position of the stepping motor serving as a drive source of the actuator 2 in the shock absorber 10. The control step number step is set to the adjustment rod 42 in the shock absorber 10. Corresponds to the up and down position. The larger the control step number is, the smaller the clearance C of the oil passage formed by the adjusting rod 42 is, and the damping force is increased (the damping force is shifted to the hard side).
The smaller the ep, the larger the clearance C of the oil passage and the lower the damping force (the damping force shifts to the soft side).

On the other hand, if "No" in S110, that is, if it is determined that the vehicle is turning, the flow advances to S114 to determine whether fn-fn-1> 0. If "Yes" in S114, the contact load has increased compared to the previous routine. In this case, the process proceeds to S116, where 1 step is subtracted from the currently set latest control step number step, and a new Set as the control step number step. If “No” is determined in S114, S118
To determine whether fn-fn-1 <0. S118
If “Yes”, the ground contact load is smaller than in the previous routine. In this case, the process proceeds to S120, where 1 step is added to the currently set latest control step number step, and new control is performed. Set as step number step.

If the determination is "No" in S118, it means that fn = fn-1. In this case, S11
Proceeding to 2, the normal skyhook control is performed.

The load movement during turning reduces the contact load of the turning inner wheel and increases the contact load of the turning outer wheel. By adopting such a damping force changing routine, the turning inner wheel side where the contact load decreases. The damping force of the shock absorber 10 is increased, and at the same time, the damping force of the shock absorber 10 on the turning outer wheel side where the ground contact load increases is reduced. As a result, a force acts in a direction that inhibits the movement of the load. As a result, a decrease in the ground contact load is suppressed in the turning inner wheel, and an increase in the ground contact load is suppressed in the turning outer wheel.

Next, a second embodiment will be described.

FIG. 7 shows a flowchart of the damping force change control (damping force change routine) of the shock absorber 10 executed by the ECU 230. Note that the flowchart of FIG. 7 shows representatively the damping force change control of the shock absorber 10 corresponding to the wheel L in FIG. 1, and the same damping force control is performed for the other three wheels. .

The routine shown in FIG. 7 is repeatedly executed between the time when an ignition switch (not shown) is turned on and the time when it is turned off.

First, in S202, the detection results of the steering angle sensor 203, the vehicle speed sensor 204, the vertical acceleration sensor 205, and the wheel speed sensor 208 are read as the detection results of the sensors indicating the running state of the vehicle.

At S204, the control step number step set in the previous routine is read.

In S206, based on the wheel speed signal detected by the wheel speed sensor 208, the tire pressure at which the wheel speed has been detected is estimated. Specifically, as shown in FIG. 8, the wheel is treated as a linear system in which a rim side a and a belt side b are connected by a torsion spring c having a spring constant K. When the tire air pressure changes, the spring constant K of the torsion spring c changes, which is reflected as a disturbance in this linear system. Therefore, the tire pressure is estimated by estimating the disturbance using an observer.

At S207, the ground contact load is estimated based on the tire pressure. In this case, a map indicating the relationship between the tire pressure and the contact load is stored in advance, and S
A map search is performed for the corresponding contact load from the tire pressure estimated in 206.

At S208, the contact load estimated at S207 is set as fn. At S210, the contact load set at the previous routine is set as fn-1.

The subsequent steps from S212 are the same as those of the first embodiment.
The same contents as those described from 110 to S120 are performed.

By employing such a damping force changing routine, similarly to the first embodiment, a decrease in the ground contact load is suppressed in the turning inner wheel, and an increase in the ground contact load is suppressed in the turning outer wheel.

Next, a third embodiment will be described.

FIG. 9 shows a flowchart of the damping force change control (damping force change routine) of the shock absorber 10 executed by the ECU 230. Note that the flowchart of FIG. 9 representatively shows the damping force change control of the shock absorber 10 corresponding to the wheel L in FIG. 1, and the same damping force control is performed for each of the other three wheels. .

The routine shown in FIG. 9 is repeatedly executed from when an ignition switch (not shown) is turned on until it is turned off.

First, in S302, the detection results of the steering angle sensor 203, the vehicle speed sensor 204, the vertical acceleration sensor 205, and the wheel speed sensor 208 are read as the detection results of the sensors indicating the running state of the vehicle.

At S304, the control step number step set in the previous routine is read.

In the following S306, the turning radius of the tire is calculated based on the wheel speed signal detected by the wheel speed sensor 208, and in the following S307, the ground contact load is estimated based on the turning radius of the tire. In this case, a map indicating the relationship between the turning radius of the tire and the ground contact load is stored in advance, and S3
A map search is performed for a corresponding ground contact load from the tire turning radius calculated in step 06.

In the following S308, the contact load estimated in S307 is set as fn, and in the next S310, the contact load set in the previous routine is set as fn-1.

[0093] Subsequent steps from S312 are the same as those in the first embodiment.
The same contents as those described from 110 to S120 are performed.

By adopting such a damping force changing routine, similarly to the first embodiment, a decrease in the contact load is suppressed on the inner turning wheel, and an increase in the contact load is suppressed on the outer turning wheel.

Next, a fourth embodiment will be described.

FIG. 10 shows a flowchart of the damping force change control (damping force change routine) of the shock absorber 10 executed by the ECU 230. FIG.
Flowchart 0 representatively shows the damping force change control of the shock absorber 10 corresponding to the wheel L in FIG. 1, and the same damping force control is performed for the other three wheels.

The routine shown in FIG. 10 is repeatedly executed between the time when an ignition switch (not shown) is turned on and the time when it is turned off.

First, in S402, the detection results of the steering angle sensor 203, the vehicle speed sensor 204, the vertical acceleration sensor 205, and the stroke sensor 207 are read as the detection results of the sensors indicating the running state of the vehicle.

At S404, the control step number step set in the previous routine is read.

At S406, it is determined whether the vehicle is traveling straight. This determination is performed based on the detection results of the steering angle sensor 203, the yaw rate sensor, and the like, as in the first embodiment.

If "Yes" is determined in S406, the process proceeds to S408, and the normal skyhook control described in S112 is performed.

On the other hand, if "No" in S406, that is, if it is determined that the vehicle is turning, S410
Then, the stroke speed Vi is calculated based on the relative stroke amount Y obtained as the detection result of the stroke sensor 207. This calculation is performed by differentiating the relative stroke amount Y with time.

In the following S412, the stroke speed Vi>
It is determined whether it is 0. If "Yes" in S412, the piston rod 16 of the shock absorber 10 is displaced in the extension direction. In this case, the flow proceeds to S414, and the current set control step number step is changed by one step from the currently set control step number step.
The value is subtracted and set as a new control step number step. If “No” in S412, the process proceeds to S416,
It is determined whether or not the stroke speed Vi <0. S416
In the case of "Yes", the piston rod 16 of the shock absorber 10 is displaced in the contracting direction. In this case, the process proceeds to S418, where 1 step is added to the currently set latest control step number step, and Number of control steps st
Set as ep.

[0104] If "No" is determined in S416, it means that the stroke speed is Vi = 0, and in this case, the process proceeds to S408, and normal skyhook control is performed.

Although the grounding load of the turning inner wheel decreases and the grounding load of the turning outer wheel increases due to the load movement during the turning, the turning outer wheel side displaced in the contraction direction by adopting such a damping force changing routine. The damping force of the shock absorber 10 is increased, and at the same time, the damping force of the shock absorber 10 on the turning inner wheel side displaced in the extension direction is reduced.

Therefore, in the shock absorber 10 on the turning outer wheel side, the damping force is increased, so that the shock absorber 10 works against the road surface and acts to push back the load moving to the turning outer wheel side, and the turning outer wheel side In this case, an increase in the contact load is suppressed. On the other hand, in the shock absorber 10 on the turning inner wheel side, the damping force is reduced, so that the movement of the wheels is eased and the vehicle easily follows the road surface. Will join. By both of these actions, a decrease in the ground contact load is suppressed on the turning inner wheel side.

Further, by this operation, the vehicle body 200 can be maintained flat against the load movement generated during the turn, so that the riding comfort of the vehicle is improved.

In each of the embodiments described above, the number of steps to be increased or decreased is 1 step in the flowchart, but is not limited to this example.
For example, a predetermined number of steps may be increased or decreased. Further, in the first to third embodiments, the number of steps to be increased or decreased may be set according to the amount of change in the contact load. In the fourth embodiment, the number of steps to be increased or decreased according to the magnitude of the stroke speed Vi. It is also possible to set the number.

[0109]

As described above, according to the damping force control device for a vehicle according to the first to fifth aspects, based on the grounding load detected by the grounding load detecting means, the hydraulic pressure relief is performed by the control means. Since the damping force of the means is controlled, it is possible to carry out the increase / decrease control of the contact load on each wheel, and in particular, it is possible to suppress the decrease of the contact load on the turning inner wheel,
Thereby, the maneuverability of the vehicle at the time of turning can be improved.

Further, according to the damping force control device for a vehicle according to claim 6, based on the detection result of the displacement state detecting means,
By the control means, for example, the damping force of the hydraulic buffer means is increased when the hydraulic buffer means is displaced in the contraction direction, and the damping force of the hydraulic buffer means is decreased when the hydraulic buffer means is displaced in the extension direction. Thus, it is possible to suppress a decrease in the ground contact load on the turning inner wheel, and thereby it is possible to improve the controllability and the riding comfort of the vehicle at the time of turning.

[Brief description of the drawings]

FIG. 1 is a front view schematically showing a vehicle equipped with a damping force control device according to a first embodiment.

FIG. 2 is a block diagram showing an input / output target in damping force change control and an ECU that performs control calculation.

FIG. 3 is a longitudinal sectional view of a shock absorber used in each embodiment.

FIG. 4 is an enlarged sectional view showing a main part of the shock absorber.

FIG. 5 is a graph showing a relationship between a displacement speed V of a piston rod and a generated damping force F.

FIG. 6 is a flowchart illustrating damping force change control according to the first embodiment.

FIG. 7 is a flowchart illustrating damping force change control according to a second embodiment.

FIG. 8 is a diagram showing a dynamic model of a wheel.

FIG. 9 is a flowchart illustrating damping force change control according to a third embodiment.

FIG. 10 is a flowchart illustrating damping force change control according to a fourth embodiment.

[Explanation of symbols]

2 ... actuator, 10 ... shock absorber, 20
0: body, 201: steering wheel, 203: steering angle sensor, 206: load sensor, 207: stroke sensor,
210: suspension arm, 230: electronic control unit.

 ──────────────────────────────────────────────────続 き Continued on the front page (72) Inventor Kazuya Sasaki 1 Toyota Town, Toyota City, Aichi Prefecture Inside Toyota Motor Corporation (72) Inventor Shigeki Ikeda 1 Toyota Town, Toyota City, Aichi Prefecture Inside Toyota Motor Corporation (72) Inventor Satoshi Suzuki 1 Toyota Town, Toyota City, Aichi Prefecture Inside Toyota Motor Corporation (72) Inventor Yoshiyuki Hashimoto 1 Toyota Town Toyota City, Aichi Prefecture Inside Toyota Motor Corporation (72) Inventor Takeshi Takeshi Aichi 1-1 1-1 Showa-cho, Kariya-shi, Japan Inside DENSO Corporation

Claims (6)

[Claims] 1. A vibration generated between a sprung member and a unsprung member of a vehicle is damped by a predetermined damping force, and the damping force is variable in a region where the relative speed between the sprung member and the unsprung member is low. A hydraulic pressure buffer having a function, a ground load detecting means for detecting a ground load on a wheel, and a ground load detected by the ground load detecting means,
Control means for controlling the damping force of the hydraulic pressure damping means. 2. The control device according to claim 1, wherein the damping force of the hydraulic buffer is reduced when the contact load increases, and the damping force of the hydraulic buffer is increased when the contact load decreases. The damping force control device for a vehicle according to any one of the preceding claims. 3. The damping force control device for a vehicle according to claim 1, wherein the ground contact load detecting means estimates a ground contact load based on an internal pressure of a tire mounted on a corresponding wheel. 4. The damping force control device for a vehicle according to claim 1, wherein the ground contact load detecting means estimates a ground contact load based on a turning radius of a tire mounted on a corresponding wheel. 5. The damping force control device for a vehicle according to claim 3, wherein said ground contact load detecting means includes a wheel speed detecting means for detecting a wheel speed of a corresponding wheel. 6. A vibration generated between a sprung member and a unsprung member of a vehicle is attenuated by a predetermined damping force, and the damping force is variable in a region where the relative speed between the sprung member and the unsprung member is low. A hydraulic pressure buffer having a function, a displacement state detecting means disposed corresponding to the hydraulic pressure buffer and detecting a relative displacement between the sprung member and the unsprung member, and detection of the displacement state detecting means Based on the result, the damping force of the hydraulic buffer is increased when the hydraulic buffer is displaced in the contracting direction, and the damping force of the hydraulic buffer is increased when the hydraulic buffer is displaced in the extending direction. And a control means for reducing the damping force.