JPH07117439A - Vehicle characteristic control device of four-wheel steering vehicle - Google Patents

Abstract

PURPOSE:To perform optimum characteristic control to correspond with the behavior of a vehicle body by detecting a yaw rate of the vehicle, and varying the control gains of variable damping force shock absorbers and an auxiliary steering device according to the peak value of the yaw rate. CONSTITUTION:Variable damping force shock absorbers 3FL-3RR are disposed between each wheel 1FL-1RR and a vehicle body, with their damping forces switched by step motors 41FL-41RR. An auxiliary rear-wheel steering cylinder 79 is connected to the rear wheels 1RL, 1RR and is driven via a servo valve 85, etc. The step motors 41FL-41RR and the servo valve 85, etc., are controlled by a controller 4 in accordance with detection signals from sensors 51FL-51RR, 52FL-52RR, 53, 54S, 54R, 55 for detecting various operating conditions. In this case, the control gains of the shock absorbers and the cylinder 79 are varied according to the peak value of yaw rate which is detected by the sensor 55.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a vehicle characteristic control device for a four-wheel steering vehicle in which at least rear wheels can be assist-steered according to a steering angle or the like, and is intended to improve riding comfort and steering stability. is there.

[0002]

2. Description of the Related Art As a conventional vehicle characteristic control device for a four-wheel steering vehicle, for example, Japanese Patent Application Laid-Open No. 1-9 previously proposed by the applicant of the present application.
Some are described in Japanese Patent No. 5969. In this conventional example, a suspension that can switch and control suspension characteristics such as spring constant, damping force, and roll rigidity, an auxiliary steering device that assists steering of at least one of the front wheels and the rear wheels, and the auxiliary steering device that adjusts the steering angle and the like. In a four-wheel steering vehicle including steering control means for controlling the suspension characteristics, a suspension characteristic change detecting means for detecting a change in the suspension characteristics, and a steering control means for the steering control means according to the characteristic detection value of the suspension characteristic change detecting means. With the configuration including the auxiliary steering amount correcting means for correcting the auxiliary steering amount, the steering characteristics of the four-wheel steering vehicle are maintained in an appropriate state regardless of changes in the suspension characteristics.

[0003]

However, in the above-described conventional vehicle characteristic control device for a four-wheel steering vehicle, the change in the steering stability due to the change in the suspension characteristic is simply corrected to maintain the initial steering characteristic. However, the suspension characteristics and the steering characteristics are not linked to each other according to the running state of the vehicle, but both the suspension characteristics and the steering characteristics are independently controlled. Since it only corrects the amount of auxiliary steering according to changes in suspension characteristics, it may cause discomfort due to steering stability control, or may cause vehicle vibration when a side wind is received. There is an unsolved problem that the steering stability control cannot be properly controlled according to the behavior of the vehicle.

Therefore, the present invention has been made by paying attention to the unsolved problem of the above-mentioned conventional example, and performs the optimal characteristic control according to the behavior of the vehicle while suppressing the occurrence of discomfort due to the steering stability control. An object of the present invention is to provide a vehicle characteristic control device for a four-wheel steering vehicle that can be used.

[0005]

In order to achieve the above-mentioned object, a vehicle characteristic control device for a four-wheel steering vehicle according to a first aspect of the present invention assists steering at least a rear wheel and a suspension having a variable damping force shock absorber. An auxiliary steering device, damping force control means for controlling the damping force of the variable damping force shock absorber according to vertical movement of the vehicle body, and steering control means for controlling the auxiliary steering device according to the steering angle and the like are provided. In a vehicle characteristic control device for a four-wheel steering vehicle, a yaw rate detecting means for detecting a yaw rate of the vehicle, a peak detecting means for detecting a peak value of a yaw rate detection value of the yaw rate detecting means, and a peak value detected by the peak detecting means. And a control gain changing means for changing the control gain of the damping force variable shock absorber and the auxiliary steering device according to It is.

Further, in the vehicle characteristic control device for a four-wheel steering vehicle according to claim 2, the control gain changing means is configured to change the yaw rate feedback control gain of the auxiliary steering device according to the yaw rate peak value. It is characterized by being.

[0007]

In the vehicle characteristic control device for a four-wheel steering vehicle according to the first aspect, it is possible to detect the fluctuation of the vehicle by detecting the yaw rate of the vehicle and detecting the peak value thereof, and from the peak value. By controlling the damping force characteristic and the steering characteristic in accordance with the magnitude of, the optimum vehicle characteristic control is performed without causing a feeling of strangeness.

Further, in the vehicle characteristic control device for a four-wheel steering vehicle according to the second aspect, the yaw rate feedback control gain in the steering control is changed according to the yaw rate peak value to reduce the yaw rate peak and to improve the steering stability. Improve.

[0009]

Embodiments of the present invention will be described below with reference to the drawings. FIG. 2 is a schematic configuration diagram showing an embodiment in which the present invention is applied to a four-wheel steering vehicle, and each wheel 1FL to 1RR.
Damping force variable shock absorbers 3FL to 3RR constituting a suspension device are respectively disposed between the vehicle body 2 and the vehicle body 2, and step motors 41FL to 41RR for switching the damping force of these damping force variable shock absorbers 3FL to 3RR are provided as a controller 4 which will be described later. Controlled by the control signal from.

In the front wheels 1FL and 1RR, one end of tie rods 73L and 73R is connected to a knuckle (not shown), and the other ends of the tie rods 73L and 73R are connected to a rack shaft 74a of a rack and pinion type steering device 74.
The steering shaft 75 of the rack and pinion type steering device 74 is connected to a steering wheel 76, and by steering the steering wheel 76, the front wheels 1FL, 1RR are steered in the same direction as the steering direction.

On the other hand, the rear wheels 1RL and 1RR are connected to a knuckle (not shown) with piston rods 79a of a rear wheel auxiliary steering cylinder 79 via tie rods 78L and 78R. The rear wheels 1RL, 1RR are connected to the output side of the differential device 81 via the axles 80L, 80R, and the input side of the differential device 81 receives the rotational force of the engine 83 via the propeller shaft 82. It is connected to the output side of and is driven to rotate.

The rear wheel auxiliary steering cylinder 79 has pressure chambers 89L and 89R defined by a piston 79b.
Is connected to a closed center type servo valve 85 and is connected to a protruding side of a hydraulic pump 88 which is rotationally driven by the engine 83 via an unload valve 87, drain ports are connected to each other and an oil tank 89 is connected. There is. Incidentally, 90 is an accumulator for accumulating the line pressure. Here, the rear wheel auxiliary steering cylinder 79, the servo valve 85, the unload valve 87, the hydraulic pump 88, the oil tank 89, and the accumulator 90 constitute a rear wheel steering device.

Variable damping force shock absorber 3FL to 3RR
As shown in FIGS. 3 to 7, the piston is configured as a twin-tube type gas-filled strut type having a cylinder tube 7 composed of an outer cylinder 5 and an inner cylinder 6, and the inside of the inner cylinder 6 is in sliding contact with the piston. 8 define upper and lower pressure chambers 9U and 9L. 4 to 7, the piston 8 includes a cylindrical lower half body 11 having a central opening 10 formed in the inner peripheral surface of a seal member 9 that is slidably contacted with the inner cylinder 6 on the outer peripheral surface. It is composed of an upper half body 12 fitted in the lower half body 11.

In the lower half body 11, an expansion-side oil passage 13 is formed so as to vertically penetrate therethrough, and a sealing member 9 is provided downward from the upper surface side.
Hole 14a having a diameter larger than that of the expansion-side oil passage 13 and extending from the outer peripheral surface of the cylindrical body 11 to the hole 14a.
The pressure side oil flow passage 14 formed by a hole portion 14b that is formed by communicating with the bottom portion of a, the annular grooves 15U and 15L formed at the upper and lower open ends of the central opening 10, and formed on the upper surface side. A long groove 16 that communicates with the annular groove 15U and the expansion-side oil passage 13 is formed, and a long groove 17 that is formed on the lower surface side and that communicates with the annular groove 15L is formed. The long groove 17 is closed by the expansion side disk valve 18, and the upper end side of the compression side oil flow path 14 is closed by the compression side disk valve 19.

The upper half 12 has a small-diameter shaft portion 21 fitted in the central opening 10 of the lower half body 11 and the shaft portion 2.
1 and a large diameter shaft portion 22 having a diameter smaller than the inner diameter of the inner cylinder 6 integrally formed at the upper end of the small diameter shaft portion 21 and the lower end surface of the small diameter shaft portion 21 at the center position. Hole 23a reaching from the side to the middle portion of the large-diameter shaft portion 22 and a hole portion 23b having a smaller diameter than the hole portion 23a communicating with the upper end side of the hole portion 23a.
And a through hole 23 composed of a hole portion 23c having a larger diameter than this and communicating with the upper end side of the hole portion 23b is formed. Pair of through holes 24 penetrating to the inner peripheral surface side in the direction
a, 24b and 25a, 25b, and an arc-shaped groove 2 communicating with the upper end side of the hole portion 23a of the large-diameter shaft portion 22.
6 is formed, and an L-shaped pressure-side oil passage 27 that connects the arc-shaped groove 26 and the lower end surface is formed.
Is blocked by.

The lower half body 11 and the upper half body 12 are located below the lower half body 11 of the small diameter shaft portion 21 with the small diameter shaft portion 21 fitted in the central opening 10 of the lower half body 11. The nut 29 is screwed into the lower end portion projecting to the end and tightened with the nut to be integrally connected. Further, a cylindrical valve body 31 having a closed upper end which constitutes a variable throttle is rotatably disposed in the hole 23a of the upper half body 12. As shown in FIG. 4, the valve body 31 has an upper half body 1
2 has a through hole 32 that reaches the inner peripheral surface in the radial direction at a position facing the arcuate groove 26 of the large-diameter shaft portion 22, and the small-diameter shaft of the upper half body 12 is formed as shown in FIGS. A communication groove 33 is formed on the outer peripheral surface of the portion 21 corresponding to the space between the through holes 24a and 25a. Further, as shown in FIG. 6, the through hole 24b of the small diameter shaft portion 21 of the upper half body 12 is formed.
And 25b, an elongated hole 34 extending in the axial direction is formed on the outer peripheral surface corresponding to between the inner peripheral surface 25b and the inner peripheral surface 25b.
The positional relationship among the through hole 32, the communication groove 33, and the elongated hole 34 is selected so as to obtain a damping force characteristic with respect to the rotation angle of the valve body 31 shown in FIG. 8, that is, the step angle of step motors 41FL to 41RR described later. ing.

That is, for example, at the position A in FIG. 8 which is the maximum rotation angle position in the clockwise direction, as shown in FIG. 4, only the through hole 32 communicates with the arcuate groove 26, and therefore the piston 8 descends. For pressure side movement, lower pressure chamber 9L
From the pressure side oil flow path 14 to the upper pressure chamber 9U through an orifice formed by the opening end and the pressure side disk valve 19, and the pressure side flow path C1 shown by a broken line and the lower pressure chamber 9L.
Through the inner peripheral surface of the valve body 31, the through hole 32, the arc-shaped groove 2
6. A pressure side flow path C2, which is shown by a broken line, is formed toward the upper pressure chamber 9U through the orifice formed by the opening end of the pressure side oil flow path 27 and the pressure side disk valve 28, and the piston 8 rises. With respect to the extension side movement, the broken line extending from the upper pressure chamber 9U to the lower pressure chamber 9L through the elongated groove 16 and the extension side flow path 13 and the orifice formed by the open end and the extension side disk valve 18. Only the extension side flow path T1 shown in the figure is formed, and a high damping force that rapidly increases as the piston speed increases is generated on the extension side, and a low damping force slightly increases on the compression side as the piston speed increases. Generates damping force.

By rotating the valve body 31 counterclockwise from the position A, the communication groove 33 of the valve body 31 and the through holes 24a, 25a of the small-diameter shaft portion 21 communicate with each other, as shown in FIG. Therefore, the opening areas of the communication groove 33 and the through holes 24a and 25a gradually increase as the turning angle increases. Therefore, for the extension side movement of the piston 8, as shown in FIG. 5A, the long groove 16 and the annular groove 15 are arranged in parallel with the flow path T1.
U, the through hole 24a, the communication groove 33, the through hole 25a, the annular groove 15L, the long groove 17, and the lower pressure chamber 9 through the orifice formed by the long groove 17 and the pressure side disk valve 18.
Since the flow path T2 toward L is formed, the maximum value of the damping force gradually increases as the opening area between the communication groove 33 and the through holes 24a and 25a of the small diameter shaft portion 21 increases, as shown in FIG. As shown in FIG. 5 (b), the minimum damping force state is maintained in order to maintain the state in which the flow paths C 1 and C 2 are formed, as shown in FIG. 5B.

Further, when the valve body 31 is rotated counterclockwise to the vicinity of the position B, as shown in FIG. 6, the through holes 24b and 25b of the valve body 31 are communicated by the elongated hole 34. Become. Therefore, for the extension side movement of the piston 8, as shown in FIG. 6A, the elongated groove 16, the annular groove 15U, the through hole 24a, and the elongated hole 3 are arranged in parallel with the flow paths T1 and T2.
4, the flow path T3 that goes toward the lower pressure chamber 9L through the hole 23a
Is formed, and the extension side damping force becomes the minimum damping force state, and with respect to the pressure side movement of the piston 8, in addition to the flow paths C1 and C2, the hole portion 23a, the long hole 34, and the through hole 2 are formed.
4b, a flow path C reaching the long groove 16 through the annular groove 15U
3 and the hole 23a, the long hole 34, the through hole 25b, the annular groove 15L, the through hole 25a, the communication groove 33, the through hole 24a, and the annular groove 15U to form the flow path C4 that reaches the long groove 16. , The minimum damping force state is maintained, as shown in FIG.

Further, when the valve body 31 is rotated in the counterclockwise direction, the opening area between the elongated hole 34 and the through holes 24b and 25b becomes smaller, and the elongated hole 34 and the through hole 24 at the rotation angle θ _B2.
As shown in FIG. 7, the area between b and 25b is blocked, but the opening area between the through hole 32 and the arcuate groove 26 gradually decreases from the rotation angle θ _B2 . Therefore, between the rotation angle θ _B2 and the maximum counter-clockwise rotation angle θ _C , the flow paths T1 and T2 coexist with respect to the extension side movement of the piston 8, so that the minimum damping force state is set. On the contrary, with respect to the pressure side movement of the piston 8, the maximum damping force is gradually increased by gradually decreasing the opening area between the through hole 32 and the arcuate groove 26, and the valve body 31 is positioned. When reaching C, as shown in FIG. 7, the through-hole 32 and the arcuate groove 26 are cut off from each other, so that the lower pressure chamber 9 is prevented from moving toward the pressure side of the piston.
The flow path from L to the upper pressure chamber 9U is only the flow path C1 and is in the compression side high damping force state.

On the other hand, a cylindrical piston rod 35 is fitted in the hole portion 23c of the upper half body 12, and the upper end of the piston rod 35 is projected above the cylinder tube 7 as shown in FIG. , Its upper end side is the vehicle body side member 36
Is fixed to the bracket 37 attached to the bracket 37 by a nut 39 via rubber bushes 38U and 38L, and the step motors 41FL to 41RR are mounted on the upper end of the piston rod 35 via a bracket 40 so that the rotary shafts 41a to 41RR of the step motors 41FL to 41RR are rotated.
Is fixed in a downwardly projecting relationship, and the rotary shaft 41a and the valve element 31 described above are connected by a connecting rod 42 that is loosely inserted in the piston rod 35. In addition, 43 is a bumper rubber. The lower end of the cylinder tube 7 is connected to a wheel side member (not shown).

The input side of the controller 4 is shown in FIG.
As shown in, the vertical acceleration detection values X _2FL ″ to X 2 which are analog voltages that are positive in the upward direction and negative in the downward direction according to the vertical acceleration provided on the vehicle body side corresponding to each wheel position.
Vertical acceleration sensors 51FL to 51RR as vertical acceleration detecting means for outputting _2RR ″, and an inductance corresponding to relative displacement between a vehicle body side member and a wheel side member, for example, built in a cover of each damping force variable shock absorber 3FL to 3RR. Relative displacement detection value X which is analog voltage due to change
_DFL (= X _2FL -X _1FL ) ~ X _DRR (= X _2RR-
Stroke sensors 52FL to 52RR as relative displacement detecting means for outputting X _1RR ), a vehicle speed sensor 53 for detecting a vehicle speed, a steering angle sensor 54S for detecting a steering angle of a steering wheel 76, and a rear wheel auxiliary steering cylinder 79. The rear wheel steering angle sensor 54R for detecting the rear wheel steering angle by detecting the movement amount of the vehicle and the yaw rate sensor 55 for detecting the yaw rate generated by the vehicle body are connected, and the damping force variable shock absorbers 3FL to 3RR are provided on the output side. The step motors 41FL to 41RR for controlling the damping force of the servo valve 85 are connected.

The controller 4 includes a microcomputer 56 having at least an input interface circuit 56a, an output interface circuit 56b, an arithmetic processing unit 56c and a storage unit 56d, and a vertical acceleration sensor 51.
FL-51RR vertical acceleration detection values X _2FL ″ _{-X 2RR} ″ are converted into digital values and supplied to the input interface circuit 56 a. A / D converters 57FL-57RR and stroke sensors 52FL-52RR relative displacement detection values X Input interface circuit 56a by converting _{DFL to} X _DRR into digital values
And the A / D converters 58FL to 58RR supplied to the input interface circuit 56a and the steering angle detection value θ _S of the steering angle sensor 54 converted to a digital value.
0R and the rear wheel rudder angle sensor 54R rear wheel rudder angle detection value δ _rd are converted into digital values to input interface circuit 56a.
To the A / D converter 60R and the yaw rate sensor 5
A / D converter 6 that converts the yaw rate detection value Y _D of 5 into a digital value and supplies it to the input interface circuit 56a
0Y and step control signals for the step motors 41FL to 41RR output from the output interface circuit 56b are input, and the step drive signals are converted to step pulses to drive the step motors 41FL to 41RR, and motor drive circuits 59FL to 59RR, Drive circuits 61a, 6 for driving the servo valve 85 of the rear wheel steering system by the drive control signals CS _ra and CS _rb output from the output interface circuit 56b.
1b and.

Here, the arithmetic processing unit 56c of the microcomputer 56 executes the processing of FIGS.
A vehicle body vertical velocity X _2FL ′ to X _2RR ′ _obtained by integrating the vehicle body vertical acceleration detection values X _2FL ″ to X _2RR ″ input from the vertical acceleration sensors 51FL to _51RR, and a stroke sensor 52FL.
.About.52RR relative displacement detection value between the wheel and the vehicle body X _DFL (= X _2FL -X _1FL ) to X _DRR (= X _2RR -X
_1RR ) is differentiated to determine a damping force coefficient C for performing skyhook control based on relative speeds X _DFL ′ to X _DRR ′, and the step motor 4 corresponding to the determined damping coefficient C is determined.
The target step angle θ _T of 1 FL to 41 RR is calculated, the difference value between the target step angle θ _T and the current step angle θ _P is calculated, and the step control amount corresponding to this is calculated.
9FL to 59RR, the front wheel steering angle δ _f is calculated based on the steering angle detection value θ _S of the steering angle sensor 54S, and then the front and rear wheel steering angle ratio k is calculated based on the vehicle speed detection value V of the vehicle speed sensor 53. And the rear wheel steering angle δ based on the steering angle ratio k.
_r is calculated, the opening / closing control signals CS _ra and CS _rb are output so that the difference between the rear wheel steering angle δ _r and the rear wheel steering angle detection value δ _rd becomes zero, and based on the yaw rate detection value Y _D. The peak value is detected, and the yaw rate feedback control gain k _P in steering control is changed according to the magnitude of this peak value, and the damping coefficient for the damping force variable shock absorber on the turning outer wheel side is changed.

Further, the storage device 56d is the processing unit 5
The program necessary for the arithmetic processing of 6c is stored in advance, the values required in the arithmetic processing and the arithmetic results are sequentially stored, and the target yaw rate map for calculating the target yaw rate is stored in advance. Here, in the target yaw rate map, as shown in FIG. 13, the steering angle detection value θ _S of the steering angle sensor 54S is on the X axis, and the vehicle speed sensor 53 is on the Y axis.
Of the vehicle speed detection value V and the target yaw rate Y _O on the Z axis, respectively. For example, when the steering angle detection value θ _S is 90 degrees and the vehicle speed detection value V is 60 km / h, a three-dimensional map is formed. There is. Note that this target yaw rate map is determined only by the tire characteristics. For example, in the case of a tire having a high grip, the peak point extends over a large range of the steering angle detection value θ _s, which is different from FIG. If it passes, the characteristics will decrease sharply.

Next, the operation of the above-described embodiment will be performed by a microcomputer.
Of the damping force control processing of the arithmetic processing unit 56c of the computer 56
10 showing an example, FIG. 11 showing an example of steering control processing, and
An example of the attenuation coefficient setting process will be described with reference to FIG.
It That is, the damping force control process of FIG.
It is executed as a timer interrupt process every (for example, 20 msec).
First, in step S1, the vehicle speed detection value V and the steering angle detection value θ
_SAnd each vertical acceleration detection value X_2i″ (I = FL, FR, RL, R
R), then move to step S2,
Displacement detection value X_DiRead, and then move to step S3
Then, the vertical acceleration detection value X read in step S1_2i″
Is integrated by, for example, low-pass filtering
Body vertical speed X_2i′ Is calculated and stored in the storage device 56d.
Temporarily store in a predetermined storage area, and then move to step S4.
Relative displacement detection value X read in step S2_DiExample
For example, a high-pass filter is applied to differentiate and
Speed X_Di′ Is calculated, and these are stored in a predetermined memory of the storage device 56d.
After temporarily storing in the memory area, the process proceeds to step S5.

In this step S5, the above-mentioned step S3
And the vertical speed X of the vehicle body calculated in S4 _2i′ And relative velocity X
_Di′ And control gain C_SBased on and, the calculation of the following formula (1)
And calculate the damping coefficient C for skyhook control.
After outputting, the process proceeds to step S6. C = C_S・ X_2i′ / X_Di'(1) In step S6, the control gain setting process of FIG.
Therefore, the damping force control flags FM and FH are both reset to "0".
It is determined whether or not the damping force control flag F is
When it is reset to "0", the step S
16, and one of the control flags FM and FH
If it is set to "1", move to step S7.
To go.

In step S7, it is determined whether or not the control flag FM is set to "1", and when the control flag FM is set to "1", step S8.
And the damping coefficient C of each damping force variable shock absorber 3i (i = FL, FR, RL, RR) calculated in step S5.
It is determined whether or not _i is equal to or greater than a preset moderate damping coefficient C _M , and if C _i ≧ C _M , the direct step S
16, and if C _i <C _M , then step S
After shifting to 9, the damping coefficient C _FO on the turning outer wheel side of the front wheel is set to the medium damping coefficient C _M , and then the procedure shifts to step S16.

On the other hand, if the result of determination in step S7 is that the control flag FM has been reset to "0", it is determined that the control flag FH has been set to "1", and step S12 follows. Whether the damping coefficient C _i for each damping force variable shock absorber 3i calculated in step S5 is equal to or higher than the preset high damping coefficient C _H between the middle damping coefficient C _M and the maximum damping coefficient C _MAX set in step S5. If C _i ≧ C _H , step S1 is directly executed.
6, and when C _i <C _H , step S1
After shifting to step 3 and setting the damping coefficient C _i to the high damping coefficient C _H , the processing shifts to step S16.

In step S16, the above steps S5 and
The damping coefficient C calculated in S9 or S13 is the preset minimum damping force C in the damping force variable shock absorber 3i.
_If C> C _MIN , it is determined whether or not the vehicle body vertical speed X _2i ′ is positive, and if C 2> C _MIN, it is determined whether or not X _2i ′> 0. ,
Going to step S18, the target step is referred to by referring to the regions of θ _A to θ _B1 of the control map corresponding to FIG. 8 so that the damping coefficient C calculated in step S5, S9 or S13 is set on the extension side. Step S1 after calculating the angle θ _T
Move to 9.

In step S19, the storage device 56d
Current step angle θ stored in _PAnd target step angle
θ_TAnd the deviation is calculated as step control amount S
While updating and storing in a predetermined storage area of the storage device 56d,
The target step angle θ_TThe current step angle θ_PAs
Newly memorize, then move to step S20
Step control stored in a predetermined storage area of the unit 56d
Output the quantity S to the motor drive circuit 59i and then interrupt the timer
The process is terminated and the process returns to the predetermined main program.

Further, the determination result of step S17 is X _2i ′.
When <0, the process proceeds to step S21, and the region of θ _B2 to θ _C of the control map corresponding to FIG. 8 is set so that the damping coefficient C calculated in step S5, S9 or S13 is set on the pressure side. Target step angle θ _T
After calculating, the process proceeds to step S19. Furthermore, when the determination result of step S16 is C ≦ C _MIN , the process proceeds to step S22, and the target step angle θ _T is calculated with reference to the regions of θ _B1 to θ _B2 of the control map corresponding to FIG. Then, the process proceeds to step S19.

The processing of FIG. 10 corresponds to the damping force control means, and the processing of steps S7 to S15 and the damping coefficient setting processing of FIG. 12 described later correspond to the damping coefficient changing means.
The steering control process of FIG. 11 is executed as a timer interrupt process for each predetermined time (for example, 20 msec), similarly to the damping force control process. First, in step S31, the vehicle speed detection value V of the vehicle speed sensor 53 and the steering angle sensor 54S are processed. Of the steering angle detection value θ _S , the yaw rate detection value Y _D of the yaw rate sensor 55, and the rear wheel steering angle detection value δ _rd of the rear wheel steering angle sensor 54R are read, and then the process proceeds to step S32 to detect the steering angle detection value θ
_The front wheel steering angle δ _F (= θ) is obtained by dividing _S by the steering gear ratio N.
Calculate _S / N).

Next, in step S33, the steering angle ratio k of the front and rear wheels is calculated by performing the following equation (2) based on the detected vehicle speed V. k = {bL-mV ² (a / C _r )} / {aL-mV ² (a / C _f )} ... (2) Next, the process proceeds to step S34 and the vehicle speed detection value V and the steering angle detection value θ are obtained. Based on _S , the target yaw rate Y _O is calculated with reference to the target yaw rate map in FIG. 13, then the process proceeds to step S35, and the deviation between the target yaw rate Y _O and the yaw rate detection value Y _D read in step S31. ε (=
Y _O -Y _D) is calculated, then the process proceeds to step S36, calculates a yaw rate deviation differential value epsilon 'by differentiating the yaw rate deviation epsilon example by high-pass filtering, then following the processing proceeds to step S37 (3 ) Is calculated to calculate the rear wheel steering angle δ _r .

[0035] _{δ r = k · δ f +} k P · ε + k D · ε '............ (3) where, k _P is the yaw rate feedback control gain storage device in the control gain setting process of FIG. 12 described later The value updated and stored in the predetermined storage area of 56d is read and used, and k _D is a preset fixed value control gain. Next, in step S38, the rear wheel steering angle δ _r
Between the rear wheel steering angle detection value δ _rd and Δδ _r (= δ _r −δ _rd )
When the difference value Δδ _r is zero, the servo valve 8
When both the control signals CS _ra and CS _rb for 5 have a logical value “0” and the difference Δδ _r is positive (Δδ _r > 0), the control signal CS _ra has a logical value “1” and the control signal CS _rb.
Is set to a logical value “0”, and the difference value Δδ _r is negative (Δδ _r <0), the control signal CS _ra is set to a logical value “0” and the control signal CS _rb is set to a logical value “1”. Drive circuit 61
After outputting to a and 61b, the timer interrupt processing is terminated and the predetermined main program is restored.

The processing of FIG. 11 corresponds to the steering control means. Further, the control gain setting process of FIG. 12 is executed as a timer interrupt process at every predetermined time (for example, 20 msec) similarly to the damping force control process and the steering control process, and first, at step S41, the yaw rate detection value Y _D ( n) is read, and then the process proceeds to step S42 to determine whether the absolute value of the yaw rate detection value Y _D (n) read this time is smaller than the absolute value of the yaw rate detection value Y _D (n-1) read last time. By determining whether or not, the yaw rate peak value Y _P is detected.

Then, the routine proceeds to step S43, where it is judged whether or not the yaw rate peak value Y _P is equal to or more than a preset relatively small first set value Y _PS1 , and when Y _P <Y _PS1 , step S43 In S44, the medium damping force control flag FM and the high damping force control flag FH in the damping force control process are both reset to "0", and the yaw rate feedback control gain k _P in the steering control process is set.
Is set to a low control gain k _PL , which is updated and stored in a predetermined storage area of the storage device 56d, and then the timer interrupt processing is terminated and the predetermined main program is restored.

The determination result of step S43 is Y _P ≧
When it is Y _PS1 , the process proceeds to step S45, and the yaw rate peak value Y _P is the first set value Y set in advance.
It is determined whether or not the second set value Y _PS2 larger than _{PS1 is} greater than or equal to the second set value. If Y _P <Y _PS2 , step S46.
, The medium damping force control flag FM is set to "1", and the yaw rate feedback control gain k _P
Is set to a standard control gain k _PN that is higher than the low control gain k _PL, and this is updated and stored in a predetermined storage area of the storage device 56d, then the timer interrupt process is terminated and the process returns to the predetermined main program.

Further, the determination result of step S45 is Y _P
When ≧ Y _PS2 , the process proceeds to step S47,
While setting the high damping force control flag FH to "1",
The yaw rate feedback control gain k _P is set to a high control gain k _PH higher than the standard control gain k _PN , and this is updated and stored in a predetermined storage area of the storage device 56d, and then the timer interrupt processing is terminated and a predetermined main program is executed. Return to.

Therefore, if it is assumed that the vehicle is traveling straight at a constant speed on a flat good road, there is almost no vertical movement of the vehicle body in this state.
Vertical acceleration detection value X _2FL ″ output from 1RR
X _2RR ″ is substantially zero, and the steering angle detection value θ _S is also zero.
As a result, when the damping coefficient setting process of FIG. 12 is executed, the target yaw rate Y _O calculated in step S42 also becomes substantially zero. Therefore, after step S43, the process proceeds to step S44 where both the medium damping force control flag FM and the high damping force control flag FH are reset to "0", and
Yaw rate feedback control gain k _P in the steering control is set to a low control gain k _PL.

Therefore, when the damping force control process of FIG. 10 is executed, the vehicle body vertical velocities X _2FL ′ to X _2RR ′ calculated in step S3 also become substantially zero, and the middle and high damping force control flags FM and Since both FH are reset to "0", the damping coefficients C _{FL to} C _RR for the damping force variable shock absorbers 3 _FL to 3 _RR also become substantially zero in step S5. Therefore, the flow proceeds from step S6 to step S16 to step S22. After the transition, the step angles within the range of step angles θ _{B1 to} θ _B2 that become the _expansion side and _compression side minimum damping coefficients C _nMIN and C _aMIN are set as the target step angle θ _T , and the step angles of the step motors 41 FL to 41 RR are set. Is the target step angle θ
Driven to match _T. Therefore, the valve bodies 31 of the damping force variable shock absorbers 3FL to 3RR are set to the position B shown in FIG. 6, whereby the damping coefficients C on the extension side and the _compression side of the piston 8 are the minimum damping coefficients C _nMIN and C _{n, respectively.}
Set to _aMIN . Therefore, in this state, even if vibrations due to fine unevenness of the road surface are input to the wheels, the vibrations are absorbed by the variable damping force shock absorbers 3FL to 3RR and are not transmitted to the vehicle body, and a good riding comfort can be secured. It is possible to perform damping force control with emphasis on riding comfort.

On the other hand, when the steering control process of FIG. 11 is executed, the vehicle is in a straight traveling state, so the steering angle detection value θ
_{Since S} is zero and the target yaw rate Y _O calculated in step S35 is also zero, the rear wheel steering angle δ _r calculated in step S37 is also zero, so the control signals CS _ra and CS
Both _rb have the logical value "0", and the straight traveling state is maintained. At this time, the yaw rate feedback control gain k
_{Since P} is controlled to the low control gain k _PL , the feedback term of the second term on the right side of the above equation (3) will be a small value even if the vehicle body slightly sways on a gravel road or the like. Since the responsiveness is low, it is possible to prevent the occupant from feeling uncomfortable due to excessive steering stability control.

In this running condition on a good road, for example, the step up to the front
When passing a temporary step such as a difference, pass this step
If the vehicle body does not move up and down due to
_2FL’～ X_2RRSince ′ maintains zero, the minimum damping coefficient C
_aMINAnd C_nMINTo maintain the condition, the wheels should ride on the step
It can absorb the pushing force when raised, but
Riding on a relatively large step and absorbing the thrust force
If you can not move it, the vehicle body will also be displaced upward.
Therefore, the vertical speed X of the vehicle body_2FL’～ X_2RR′ Is the positive direction
Will increase. Thus, the vehicle body vertical velocity X
_2FL’～ X_2RR′ Increases in the positive direction, step S1
7 to step S18, the step angle θ in FIG.
_A~ Θ_B1Target step angle θ according to the damping coefficient C in the region of
_TIs calculated, the damping force variable shock absorber 3
The valve bodies 31 of FL to 3RR are switched and controlled as shown in FIG.
It As a result, the relative velocity X_DFL′
~ X_DRR′ Is negative, that is, the displacement speed X on the vehicle body side_2i′ Against car
Wheel side displacement speed X_1i′ Is fast and the piston 8 moves to the pressure side.
The minimum damping coefficient C on the compression side_aMINHave maintained
Since it is possible to absorb the vibration input to the wheel side,
From the state of
When the degree becomes smaller than the ascending speed on the vehicle body side, the relative speed
X _DFL’～ X_DRR′ Becomes positive and the piston 8 moves to the extension side.
Will move. At this time, the damping coefficient C is large.
Since it is a value, it exhibits a damping effect that suppresses the rise of the vehicle body
Then, when the ascent of the vehicle body stops after that, the vehicle body vertical speed X
_2FL’～ X_2RRSince ′ becomes zero, as described above
Step motors 41FL to 41RR are rotated counterclockwise.
Is returned to the position B, which causes both the compression side and the extension side to
The minimum damping coefficient C _aMINAnd C_nMINControlled by then car
When the body starts to descend, the vehicle body vertical speed X
_2FL’～ X_2RRAs ′ increases in the negative direction,
The control of FIG. 8 is performed by shifting from step S17 to step S21.
Step angle θ with reference to the map_B2~ Θ _CAttenuator in the range of
Target step angle θ according to number C_TBy calculating
7, the valve body 31 is further rotated counterclockwise, and as shown in FIG.
It is rotated to the rotation position shown. For this reason, the vehicle body descends,
And relative speed X_DFL’～ X_DRR′ Becomes negative and piston
When 8 moves to the pressure side, the damping force should be large.
As a result, a large damping effect is exerted.

On the contrary, when the wheels pass through the step of falling forward, the relative speeds X _DFL ′ to X _DRR ′ increase in the positive direction due to the rebounding of the wheels, but at this time, the vehicle body does not move up and down. Vertical speed of the vehicle X _2FL '~ X
_{Since 2RR} 'is zero, the damping coefficients of the variable damping force shock absorbers 3FL to 3RR are the minimum damping coefficients C _aMIN and C _{nM IN.}
Then, when the vehicle body starts to descend after that, when the vehicle body vertical speeds X _2FL ′ to X _2RR ′ increase in the negative direction, the damping coefficient C becomes a large value and the step angle θ _Since the target step angle θ _{T in} the range of _{B 2 to} θ _C is calculated, and the valve body 31 is rotated to the position shown in FIG. 7, a large damping force is applied to the pressure side movement of the piston 8. A large damping effect can be exerted by giving the same, and thereafter, the valve body 31 is rotated clockwise in accordance with the decrease in the vehicle body vertical speeds X _2FL ′ to X _2RR ′ and the decrease in the damping coefficient C. Return to the B side, and the vehicle body vertical speed X _2FL ′ ~
When X _2RR ′ becomes zero, the valve body 31 becomes the position B, and the minimum damping coefficients C _aMIN and C _nMIN are obtained. After that, when the vehicle body starts to rise by _swinging back, the vehicle body vertical velocities X _2FL ′ to X _2RR ′ increase in the positive direction and the relative velocities X _DFL ′ to X _DRR ′ become the positive direction, so that the damping coefficient C As the target step angle θ _T on the side of the step angle θ _A is calculated with an increase in the valve angle, the valve body 31 is rotated clockwise to the position shown in FIG. It is possible to exert a large damping force on the vibration damping effect.

As described above, when passing a temporary step while traveling on a good road, a good damping effect can be exerted by skyhook control, and when traveling on a bad road. Also, the target step angle θ _T on the step angle θ _A side (or step angle θ _C side) is calculated by the positive (or negative) of the vehicle body vertical velocities X _2FL ′ to X _2RR ′, so that the vehicle body rises. When the relative speeds X _DFL ′ to X _DRR ′ are negative and the vehicle body is descending so that the relative speeds X _DFL ′ to X _DRR ′ are positive, the damping coefficient C is the minimum damping coefficient C.
_By controlling to _aMIN and _CnMIN , conversely, the vehicle body rises and the relative velocity X _DFL ′ to X _DRR ′ is positive, and the vehicle body descends and the relative velocity X.
When the damping direction is such that _DFL ′ to X _DRR ′ are negative, the damping coefficient C is set to the vertical velocity X _2FL ′ to X _2RR ′ and the relative velocity X.
Controlling to the optimum damping coefficient according to _DFL '~ X _DRR ',
A good ride comfort can be secured.

Even when the vehicle is traveling on a rough road, the vehicle body is lifted and the relative speeds X _DFL ′ to X X are reached as in the case of passing the step.
_{When DRR} 'is negative and the car body is descending, the relative speed X _DFL ' ~ X
_The damping coefficient C is controlled to the minimum damping coefficients C _aMIN and C _nMIN when _DRR ′ is in the positive excitation direction, and conversely, the vehicle body rises and the relative speeds X _DFL ′ to X _DRR ′ are positive and the vehicle body is The damping coefficient C is _{adjusted according} to the vertical speeds X _2FL ′ to X _2RR ′ and the relative speeds X _DFL ′ to X _DRR ′ in the damping direction in which the relative speeds X _DFL ′ to X _DRR ′ become negative and become negative. It is possible to secure a good riding comfort by controlling the damping coefficient to the optimum value.

Therefore, when the vehicle is traveling straight ahead in a stable manner, the damping force variable shock absorbers 3FL to 3FL ...
The control width of the damping force of 3RR is wide as shown by the solid line in FIG. 14, and good skyhook control can be performed. In this stable traveling state, for example, when a cross wind suddenly acts through a tunnel and a yaw occurs in the vehicle body, the yaw rate detection value Y _D of the yaw rate sensor 55 starts from time t ₁ as shown in FIG. The absolute value of yaw rate increases rapidly.

In this state, the peak value cannot be detected even if the process of FIG. 12 is executed until the peak value at the time point t ₂ is reached, and the process proceeds from step S43 to step S44. The middle damping coefficient control flag FM and the high damping coefficient control flag FH both remain in the reset state of "0", and the yaw rate feedback control gain k _P in the steering control is maintained in the low control gain k _PL. Therefore, the yaw rate cannot be suppressed.

However, when the yaw rate peak value Y _P is detected at time t ₂ , this yaw rate peak value Y _P
Is greater than or equal to the second set value Y _PS2 , the process proceeds from step S45 to step S47 when the process of FIG. 12 is executed, the high damping coefficient control flag FH is set to "1", and the steering The yaw rate feedback control gain k _P in control is set to the high control gain k _PH .

Therefore, the damping force control process of FIG. 10 is executed.
When it is done, it goes through steps S6 to S7
Move to step S12 to change each damping force shock absorber.
Damping coefficient C for server 3i_iIs a high damping force C_HAnd above
Whether or not_i≧ C _HWhen it is
The extinction coefficient is set as it is and C_i<C_HWhen
Shifts to step S13 and the damping coefficient C_iHas high damping
Number C_HRegulated by. Therefore, each damping force variable shock
The damping force of the absorber 3i is as shown by the one-dot chain line in FIG.
And high damping force C_HWill not fall anymore, each wheel
Improves ground contact from 1FL to 1RR and emphasizes steering stability
Will be done.

On the other hand, in the steering control process of FIG. 11, the yaw rate feedback control gain k _P is set to the high control gain k _PH , so in step S37 the aforesaid deviation based on the deviation between the target yaw rate Y _O and the yaw rate detection value Y _D is calculated. (3)
Since the feedback control amount of the second term on the right side of the equation has a large value, it is possible to effectively suppress the fluctuation of the vehicle body and significantly improve the steering stability.

Therefore, as shown in FIG. 14, time t ₂
After that, the yaw rate detection value Y _D gradually decreases, and when it reaches a peak smaller than the second set value Y _PS2 at the time point t ₃ , thereafter, when the processing of FIG. After shifting from S45 to step S46, the medium damping force control flag FM is set to "1" and the yaw rate feedback control gain k _P in steering control is set to the standard control gain k _PN .

Therefore, when the damping force control process of FIG. 10 is executed, the process proceeds from step S6 to step S7 to step S8, and the damping coefficient C _i for each damping force variable shock absorber 3i is set to the medium damping force. It is determined whether or not C _M or more. When C _i ≧ C _M , the damping coefficient is set as it is, and when C _i <C _M ,
In step S9, the damping coefficient C _i is the medium damping coefficient C _M.
Regulated by. Therefore, the damping coefficient C is the medium damping coefficient C _M.
Due to the above limitation, as shown by the broken line in FIG. 14, the control range of the damping force is expanded to the middle damping force, and it is possible to improve the riding comfort while ensuring a proper grounding property and steering stability. it can.

On the other hand, also in the steering control process of FIG. 11, the yaw rate feedback control gain k _P is equal to the standard control gain k.
Since it is set to _PN , the feedback control amount based on the deviation between the target yaw rate Y _O and the yaw rate detection value Y _D becomes smaller than that at the high yaw rate in step S37, which also causes an uncomfortable feeling in steering stability. It can be done to the extent that there is nothing to do.

Further, in this way, as shown in FIG. 15, when the yaw rate gradually converges and the yaw rate peak value Y _P becomes less than the first set value Y _PS1 , the processing of FIG. 12 is executed. Sometimes from step S43 to step S
After shifting to 44, the control state is returned to the same control state as the straight traveling stable traveling state described above, and the control state in which the riding comfort is emphasized is set. Incidentally, when the steering control system and the damping force control system are independent as in the conventional example, as shown in FIG. 16, the amount of fluctuation in the high yaw rate state is large and the convergence of the yaw rate decreases, Steering stability cannot be ensured.

Even when the lane is changed or the slalom traveling is performed at a high speed, the damping coefficient C and the yaw rate feedback control gain k _P can be set to the optimum values based on the yaw rate peak value in the same manner as described above. It is possible to perform the vehicle characteristic control while maintaining the riding comfort and the steering stability in the optimum state while improving the convergence of the vehicle.

In the above embodiment, when the yaw rate peak value Y _P is greater than or equal to the second set value Y _PS2 ,
Although the case where the damping coefficient C _i is controlled to be equal to or higher than the high damping coefficient C _H has been described, the present invention is not limited to this, and the maximum damping coefficient C _MAX may be fixed. Further, in the above-described embodiment, the case where the valve body 31 for controlling the damping force is configured to be a rotary type has been described, but the present invention is not limited to this, and it is configured to be a spool type, and the compression side and the extension side are configured. Different flow paths may be formed. In this case, a pinion is connected to the rotary shaft 41a of the step motors 41FL to 41RR, and a rack that meshes with the pinion is attached to the connecting rod 42 or an electromagnetic solenoid is applied. It suffices to control the sliding position of the valve body 31, and further, instead of changing the damping force continuously, it is also possible to apply a damping force variable shock absorber capable of switching the damping force in a plurality of stages.

Further, in the above embodiment, the rear wheel auxiliary steering cylinder 79 is provided with the closed sensor type servo valve 8.
Although the case where the feedback control is performed using 5 is described above, the present invention is not limited to this, and an open center type servo valve is applied, and accordingly, the piston rod 79a of the four-wheel trust 79 is returned to the neutral position. You may make it control by interposing a return spring.

Furthermore, in the above embodiment, the case where the microcomputer 56 is applied for control has been described, but the present invention is not limited to this, and a comparator,
It is also possible to combine electronic circuits such as an integrator that integrates the output of the vertical acceleration sensor 51i, a differentiation circuit that differentiates the output of the stroke sensor 52i, and a function generator.

Further, in the above embodiments, the case where the potentiometer is applied as the stroke sensor has been described, but the present invention is not limited to this, and an ultrasonic distance sensor and a detection coil for detecting the relative distance between the vehicle body and the road surface. Can be used to apply any relative displacement detection means such as a displacement sensor that detects displacement by impedance change or inductance change.

Further, in the above embodiment, the vertical acceleration sensors 51FL to 51FL are arranged at the positions of the wheels 1FL to 1RR of the vehicle body 2.
Although the case where RR is provided has been described, any one vertical acceleration sensor may be omitted and the vertical acceleration at the omitted position may be estimated from the values of other vertical acceleration sensors. Furthermore, in the above embodiment, the case where the step motors 41FL to 41RR are subjected to open loop control has been described, but the present invention is not limited to this, and the closed angle control is performed by detecting the rotation angle of the step motor with an encoder or the like and feeding it back. You may do it.

[0062]

As described above, according to claim 1,
According to the vehicle characteristic control device for a wheel steering vehicle, the yaw rate is detected by the yaw rate detecting means, the yaw rate peak value is detected by the peak value detecting means, and the damping force control or steering is performed by the control gain changing means based on the peak value. Since the control gain of the control is changed, when the fluctuation of the vehicle is small, the control gain due to the rear wheel steering is reduced to weaken the steering stability control and prevent the occurrence of discomfort, and the riding comfort is emphasized. When the vehicle characteristics control is performed and the vehicle has a lot of fluctuations, the control gains of the damping force control and the steering control can be increased to smoothly shift from the riding comfort-oriented control state to the steering stability-oriented control state. It is possible to obtain the effect of being able to perform optimal vehicle characteristic control according to the above.

Further, according to the vehicle characteristic control device for a four-wheel steering vehicle according to claim 2, the control gain changing means changes the yaw rate feedback control gain in the steering control according to the yaw rate peak value. The effect that the fluctuation of the vehicle can be effectively suppressed and the convergence of the yaw rate can be improved is obtained.

[Brief description of drawings]

FIG. 1 is a schematic configuration diagram showing a basic configuration of the present invention.

FIG. 2 is a schematic configuration diagram showing an embodiment of the present invention.

FIG. 3 is a front view with a part in section showing an example of a damping force variable shock absorber.

FIG. 4 is an enlarged cross-sectional view showing a damping force adjusting mechanism in a maximum damping force state when the vehicle body is raised.

FIG. 5 is an enlarged cross-sectional view showing a damping force adjusting mechanism in an intermediate damping force state when the vehicle body is raised, (a) showing an extension side and (b) showing a pressure side hydraulic fluid path, respectively.

FIG. 6 is an enlarged cross-sectional view showing a damping force adjusting mechanism when there is no change in the vehicle body.

FIG. 7 is an enlarged sectional view showing a damping force adjusting mechanism in a maximum damping force state when the vehicle body is descending, (a) showing an extension side and (b) showing a pressure side hydraulic fluid path, respectively.

FIG. 8 is an explanatory diagram showing damping force characteristics with respect to a step angle of a damping force variable shock absorber.

FIG. 9 is a block diagram showing an example of a controller.

FIG. 10 is a flowchart showing an example of a damping force control processing procedure of the controller.

FIG. 11 is a flowchart showing an example of a steering control processing procedure of the controller.

FIG. 12 is a flowchart showing an example of a control gain setting processing procedure of the controller.

FIG. 13 is a characteristic diagram showing a target yaw rate map showing a relationship between a steering angle detection value, a vehicle speed detection value, and a target yaw rate.

FIG. 14 is a characteristic diagram showing the relationship between piston speed and damping force.

FIG. 15 is a time chart for explaining the operation of the present invention.

FIG. 16 is a time chart for explaining the operation of the conventional example.

[Explanation of symbols]

 1FL to 1RR Wheels 2 Vehicles 3FL to 3RR Damping force variable shock absorber 4 Controller T1 to T3 Extension side flow path C1 to C4 Pressure side flow path 41FL to 41RR Step motor 51FL to 51RR Vertical acceleration sensor 52FL to 52RR Stroke sensor 53 Vehicle speed sensor 54S Steering Angle sensor 54R Rear wheel steering angle sensor 55 Yaw rate sensor 56 Microcomputer 59FL to 59RR Motor drive circuit 76 Steering wheel 79 Rear wheel auxiliary steering cylinder 85 Servo valve

Claims (2)

[Claims] 1. A suspension having a variable damping force shock absorber, an auxiliary steering device for auxiliary steering at least the rear wheels, and a damping force control means for controlling the damping force of the variable damping force shock absorber according to the vertical movement of the vehicle body. When,
A vehicle characteristic control device for a four-wheel steering vehicle, comprising: a steering control device that controls the auxiliary steering device according to a steering angle or the like; and a yaw rate detection device that detects a yaw rate of the vehicle, and a yaw rate detection value of the yaw rate detection device. And a control gain changing means for changing the control gains of the damping force variable shock absorber and the auxiliary steering device according to the peak value detected by the peak detecting means. A vehicle characteristic control device for a four-wheel steering vehicle. 2. The four-wheel steering vehicle according to claim 1, wherein the control gain changing means is configured to change a yaw rate feedback control gain of the auxiliary steering device according to a yaw rate peak value. Vehicle characteristic control device.