JP3060803B2 - Vehicle characteristic control device for four-wheel steering vehicle - Google Patents

Abstract

PURPOSE:To perform optimum damping force control and steering control to correspond with the vertical movement of the body of a four-wheel steering vehicle. CONSTITUTION:A suspension is provided with variable damping force shock absorbers and a rear-wheel auxiliary steering device, and the frequency fFL-fRR of normal acceleration detected by a normal acceleration sensor provided in each of the wheel positions of the vehicle body is calculated (step S32.) When the frequency is in the range of sprung resonance frequencies, a damping force control flag F is set to '0' and a damping coefficient is set in accordance with the normal speed and relative speed of the vehicle body, and a control gain kP for calculating the amount of steering of the rear wheels is set to a high control gain kPH (step S340). When the frequency is in the range of unsprung resonance frequencies, the flag F is set to '1' and the damping coefficient is set to its maximum, and the control gain kP is set to the high control gain kPH (step S36); when the frequency is in the range of intermediate frequencies between the sprung and unsprung resonance frequencies, the flag is set to '0' and the control gain kP is lowered to the standard control gain kPN.

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 capable of assisting at least the rear wheels in accordance with a steering angle or the like to improve ride comfort and steering stability. is there.

[0002]

2. Description of the Related Art A conventional vehicle characteristic control device for a four-wheel steering vehicle is disclosed in, for example, Japanese Patent Application Laid-Open No.
There is one described in No. 5969. This conventional example includes a suspension capable of switching and controlling suspension characteristics such as a spring constant, a damping force, and a roll rigidity, an auxiliary steering device for auxiliary steering of at least one of a front wheel and a rear wheel, and an auxiliary steering device for adjusting a steering angle and the like. In a four-wheel steering vehicle provided with a steering control means for controlling the suspension characteristic change, the suspension characteristic change detection means for detecting a change in the suspension characteristic, and the steering control means according to a characteristic detection value of the suspension characteristic change detection means. By providing an 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 the change in the suspension characteristics.

[0003]

However, in the above-mentioned conventional vehicle characteristic control apparatus for a four-wheel steering vehicle, a change in steering stability caused by a change in suspension characteristics is simply corrected to maintain the initial steering characteristics. The suspension control characteristic and the steering characteristic are not controlled in conjunction with each other according to the traveling state of the vehicle, but only the suspension characteristic and the steering characteristic are independently controlled. Can only correct the amount of auxiliary steering according to changes in suspension characteristics, but cannot properly control the ride comfort control by suspension characteristics control and the steering stability control by auxiliary steering characteristics control according to the vertical movement of the vehicle body There is an unsolved problem.

Accordingly, the present invention has been made in view of the above-mentioned unsolved problems of the prior art, and is capable of appropriately controlling ride comfort and steering stability in accordance with the vertical movement of a vehicle body.
It is an object of the present invention to provide a vehicle characteristic control device for a wheel steering vehicle.

[0005]

In order to achieve the above object, a vehicle characteristic control apparatus for a four-wheel steering vehicle according to the first aspect of the present invention includes a suspension having a variable damping force shock absorber and an auxiliary steering of at least a rear wheel. An auxiliary steering device; damping force control means for controlling the damping force of the variable damping force shock absorber in accordance with vertical movement of the vehicle body; and steering control means for controlling the auxiliary steering device in accordance with a steering angle or the like. In a vehicle characteristic control device for a four-wheel steering vehicle, a vehicle vertical acceleration detecting means for detecting vehicle vertical acceleration, a vertical acceleration frequency detecting means for detecting a frequency of vehicle vertical acceleration of the vehicle vertical acceleration detecting means, and a vertical acceleration frequency A control gain change that changes a control gain in the damping force control means and the steering control means according to the vehicle body vertical acceleration frequency detected by the detection means. It is characterized in that a means.

According to a second aspect of the present invention, in the vehicle characteristic control apparatus for a four-wheel steering vehicle, the control gain changing means controls the control gain in the damping force control means when the vertical acceleration frequency is near the sprung resonance frequency. The control gain of the steering control means is set to a high control gain, and the control gain of the damping force control means is set in an intermediate frequency range between the sprung resonance frequency and the unsprung resonance frequency. Is changed according to the vehicle vertical speed and the relative displacement speed, and the control gain of the steering control means is set to a control gain lower than the high control gain.

[0007]

In the vehicle characteristic control device for a four-wheel steering vehicle according to the first aspect, the vertical acceleration of the vehicle body, that is, the vertical acceleration of the sprung body is detected, and the frequency thereof is detected by the vertical acceleration frequency detecting means, so that the sprung frequency variation is detected. Optimum damping force control and steering characteristic control according to the sprung frequency are performed by changing the control gains in the damping force control means and the steering control means in accordance with the above.

According to a second aspect of the present invention, when the frequency of the vehicle body vertical acceleration is near the sprung resonance frequency, the damping force control means uses the vehicle body vertical speed and the relative speed. By setting the control gain, the damping performance is improved and the ride comfort is emphasized, and the steering stability is also emphasized by selecting the high control gain with the steering control means, but the sprung resonance frequency and the unsprung resonance frequency are emphasized. In the intermediate frequency range intermediate between the above, the control gain of the steering control means is reduced, the response is reduced, and unnecessary behavior of the vehicle body is suppressed, so that riding comfort is emphasized.

[0009]

Embodiments of the present invention will be described below with reference to the drawings. FIG. 2 is a schematic configuration diagram showing one embodiment in which the present invention is applied to a four-wheel steering vehicle.
The variable damping force shock absorbers 3FL to 3RR constituting a suspension device are disposed between the vehicle body 2 and the vehicle body 2, respectively. Is controlled by a control signal from

The front wheels 1FL, 1RR are connected to a knuckle (not shown) at one end of tie rods 73L, 73R, and the other ends of the tie rods 73L, 73R are connected to a rack shaft 74a of a rack and pinion type steering device 74.
A 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, 1RR are connected to a knuckle (not shown) through tie rods 78L, 78R, and a piston rod 79a of a rear wheel auxiliary steering cylinder 79. The rear wheels 1RL and 1RR are connected to the output side of a differential device 81 via axles 80L and 80R, and the input side of the differential device 81 is connected to a transmission 84 to which the rotational force of an engine 83 is input via a propeller shaft 82. And is rotationally driven.

Further, the rear wheel assist steering cylinder 79 has pressure chambers 89L, 89R defined by a piston 79b.
Are connected to a protruding side of a hydraulic pump 88 which is rotationally driven by an engine 83 via an unload valve 87, and drain ports are connected to each other and connected to an oil tank 89 via an unload valve 87. I have. Reference numeral 90 denotes an accumulator for accumulating the line pressure. Here, a rear wheel steering device is constituted by the rear wheel assist steering cylinder 79, the servo valve 85, the unload valve 87, the hydraulic pump 88, the oil tank 89, and the accumulator 90.

Variable damping force shock absorber 3FL-3RR
Is a twin-tube gas-filled strut type having a cylinder tube 7 composed of an outer cylinder 5 and an inner cylinder 6 as shown in FIGS. 8 define upper and lower pressure chambers 9U and 9L. 4 to 7, the piston 8 has a cylindrical lower half 11 having a sealing member 9 molded on the outer peripheral surface thereof in sliding contact with the inner cylinder 6 and having a center opening 10 on the inner peripheral surface. The lower half 11 has an upper half 12 fitted therein.

The lower half body 11 has an extension oil passage 13 penetrating vertically and a sealing member 9 extending downward from the upper surface side.
The hole 14 a having a diameter larger than that of the extension-side oil flow path 13 and extending from the outer peripheral surface of the cylindrical body 11 to the hole 14.
a, a pressure-side oil flow path 14 formed of a hole 14b drilled in communication with the bottom of the hole a, annular grooves 15U, 15L formed at the upper and lower open ends of the central hole 10, and formed on the upper surface side. A long groove 16 communicating with the annular groove 15U and the expansion-side oil flow path 13 and a long groove 17 formed on the lower surface side and communicating with the annular groove 15L are formed. The long groove 17 is closed by the extension-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 inserted into the center opening 10 of the lower half body 11 and the shaft portion 2.
The lower end face of the small-diameter shaft portion 21 is formed at the center of the small-diameter shaft portion 21 and the large-diameter shaft portion 22. 23a extending from the side to the middle of the large-diameter shaft portion 22, and a hole 23b communicating with the upper end of the hole 23a and having a smaller diameter than the hole 23a.
And a hole 23c having a larger diameter than the hole 23c communicating with the upper end side of the hole 23b. The through hole 23 is formed at a position facing the annular grooves 15U and 15L of the small diameter shaft 21 respectively. Pair of through holes 24 penetrating the inner peripheral surface side in the direction
a, 24b and 25a, 25b are drilled, and the upper end side of the hole 23a of the large-diameter shaft portion 22 is connected to the arc-shaped groove 2 communicating therewith.
6 is formed, and an L-shaped pressure-side oil flow path 27 communicating with the arc-shaped groove 26 and the lower end face is formed.
Is blocked by

The lower half 11 and the upper half 12 are positioned below the lower half 11 of the small-diameter shaft 21 with the small-diameter shaft 21 inserted into the central opening 10 of the lower half 11. The nut 29 is screwed into the lower end protruding from the nut, and the nut 29 is tightened to be integrally connected. Further, a cylindrical valve body 31 whose upper end is closed in a hole 23a of the upper half body 12 and constitutes a variable throttle is rotatably disposed. As shown in FIG. 4, the upper half 1
2, a through-hole 32 is formed at a position facing the arc-shaped groove 26 of the large-diameter shaft portion 22 so as to reach the inner peripheral surface in the radial direction, and the small-diameter shaft of the upper half body 12 as shown in FIGS. A communication groove 33 is formed in the outer peripheral surface corresponding to the space between the through holes 24a and 25a of the portion 21, and 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.
An elongated hole 34 extending in the axial direction is formed in the outer peripheral surface corresponding to the area between the inner peripheral surface and the inner peripheral surface side.
Then, the positional relationship between the through hole 32, the communication groove 33, and the long hole 34 is selected so as to obtain the damping force characteristic with respect to the rotation angle of the valve body 31 shown in FIG. 8, that is, the step angles of the 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 arc-shaped groove 26, so that the piston 8 descends. For pressure side movement, lower pressure chamber 9L
A pressure-side flow path C1 (shown by a dashed line) passing through the orifice formed by the open end of the pressure-side oil flow path 14 and the pressure-side disc valve 19 toward the upper pressure chamber 9U, and a lower pressure chamber 9L.
Through the inner peripheral surface of the valve body 31, through hole 32, arc-shaped groove 2
6. A pressure-side flow path C2, shown by a broken line, which passes through the pressure-side oil flow path 27, passes through an orifice formed by the opening end thereof and the pressure-side disc valve 28, and goes to the upper pressure chamber 9U, and the piston 8 rises. For the extension side movement, the upper pressure chamber 9U passes through the long groove 16 and the extension side flow path 13 and passes through the orifice formed by the opening end and the extension side disc valve 18 to the broken line toward the lower pressure chamber 9L. Only the expansion side flow path T1 shown in the figure is formed, and a high damping force is generated on the expansion side, which rapidly increases in accordance with an increase in the piston speed, and a low damping force on the compression side is slightly increased in accordance with the increase in the piston speed. Generates damping force.

By rotating the valve body 31 in the counterclockwise direction 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. The opening area between the communication groove 33 and the through holes 24a and 25a gradually increases as the rotation angle increases. For this reason, as shown in FIG. 5 (a), when the piston 8 moves on the extension side, 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 disc valve 18.
A flow path T2 toward L is formed, and 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. 5B, for the movement on the extension side, the state where the flow paths C1 and C2 are formed is maintained, so that the minimum damping force state is maintained.

Further, when the valve element 31 is rotated counterclockwise to a position near the position B, as shown in FIG. 6, a state is established in which the through hole 24b, 25b of the valve element 31 is communicated by the elongated hole 34. Become. For this reason, as shown in FIG. 6A, the elongated groove 16, the annular groove 15U, the through hole 24a, the elongated hole 3 are arranged in parallel with the flow paths T1 and T2 with respect to the extension side movement of the piston 8.
4. Flow path T3 passing through hole 23a toward lower pressure chamber 9L
Is formed, the extension-side damping force becomes the minimum damping force state, and the piston 23 moves against the compression side in addition to the flow passages C1 and C2 as well as the hole 23a, the long hole 34, and the through hole 2
4b, the flow path C reaching the long groove 16 through the annular groove 15U
3, a flow path C4 is formed that reaches the long groove 16 through 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. As shown in FIG. 8, the minimum damping force state is maintained.

Further, when the valve element 31 is rotated in the counterclockwise direction, the opening area between the elongated hole 34 and the through holes 24b and 25b is reduced, and the elongated hole 34 and the through hole 24 are rotated at the rotation angle θ _B2.
7, the opening area between the through hole 32 and the arc-shaped groove 26 gradually decreases from the rotation angle θ _B2 . For this reason, between the rotation angle θ _B2 and the maximum rotation angle θ _{C in the} counterclockwise direction, the flow path T1 and T2 coexist for the movement of the piston 8 on the extension side, so that the minimum damping force state is set. On the contrary, when the piston 8 is moved on the pressure side, the opening area between the through hole 32 and the arc-shaped groove 26 is gradually reduced, so that the maximum damping force is gradually increased, and the valve body 31 is moved to the position. As shown in FIG. 7, when the piston reaches pressure C, the space between the through-hole 32 and the arc-shaped groove 26 is cut off.
The flow path from L to the upper pressure chamber 9U is only the flow path C1, and the pressure side is in a high damping force state.

On the other hand, a cylindrical piston rod 35 is fitted into the hole 23c of the upper half body 12, and the upper end of the piston rod 35 projects upward from the cylinder tube 7, as shown in FIG. The upper end side is the vehicle body side member 36.
Is fixed by a nut 39 via rubber bushes 38U and 38L to a bracket 37 attached to the piston rod 35, and the step motors 41FL to 41RR are mounted on the upper end of the piston rod 35 via a bracket 40 by the rotation shaft 41a.
The rotating shaft 41a and the above-described valve element 31 are connected by a connecting rod 42 loosely inserted into the piston rod 35. 43 is a bumper rubber. The lower end of the cylinder tube 7 is connected to a wheel-side member (not shown).

The controller 4 has, on its input side,
As shown in the figure, according to the vertical acceleration provided on the vehicle body side corresponding to each wheel position, the vertical acceleration detection values X _2FL ″ to X composed of analog voltages that are upwardly positive and downwardly negative are provided.
Vertical acceleration sensors 51FL to 51RR as vertical acceleration detecting means for outputting _2RR ", and an inductance which is built in, for example, the cover of each of the damping force variable shock absorbers 3FL to 3RR and which corresponds to the relative displacement between the vehicle body side member and the wheel side member. Relative displacement detection value X consisting of analog voltage due to change
_DFL (= X _2FL −X _1FL ) to X _DRR (= X _2RR −
X _1RR ), stroke sensors 52FL to 52RR as relative displacement detecting means, a vehicle speed sensor 53 for detecting a vehicle speed, a steering angle sensor 54S for detecting a steering angle of the steering wheel 76, and a rear wheel assist steering cylinder 79. A rear wheel steering angle sensor 54R for detecting a rear wheel steering angle by detecting a moving amount of the vehicle and a yaw rate sensor 55 for detecting a yaw rate generated by the vehicle body are connected, and each of the damping force variable shock absorbers 3FL to 3RR is provided on the output side. The step motors 41FL to 41RR for controlling the damping force of the motor and 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 device 56c, and a storage device 56d;
The vertical displacement detection values X _2FL ″ to X _2RR ″ of the _FL to _51RR are converted into digital values and supplied to the input interface circuit 56a, and the relative displacement detection values X of the A / D converters 57FL to 57RR and the stroke sensors 52FL to 52RR. input interface circuit 56a converts the _DFL to X _DRR to a digital value
A / D converters 58FL to 58RR supplied to the A / D converter 6 and an A / D converter 6 which converts the detected steering angle θ _S of the steering angle sensor 54 into a digital value and supplies the digital value to the input interface circuit 56a.
0R and the rear wheel steering angle detection value δ _rd of the rear wheel steering angle sensor 54R are converted into digital values and input interface circuit 56a
A / D converter 60R that supplies the yaw rate sensor 5
5 of the yaw rate detection value Y _D converted to a digital value input interface circuit 56a to supply the A / D converter 6
0Y, a step control signal for each of the step motors 41FL to 41RR output from the output interface circuit 56b is input, and the motor drive circuits 59FL to 59RR for converting this into step pulses to drive each of the step motors 41FL to 41RR. Drive circuits 61a and 61 for driving the servo valve 85 of the rear wheel steering device with the drive control signals CS _ra and CS _rb output from the output interface circuit 56b.
1b.

Here, the arithmetic processing unit 56c of the microcomputer 56 executes the processing of FIGS.
A vehicle body vertical velocity X _2FL '~X _2RR' obtained by integrating the vehicle body vertical acceleration detection value X _2FL "~X _2RR" input from the vertical acceleration sensor 51FL～51RR, stroke sensor 52FL
To 52RR, the relative displacement detection values X _DFL (= X _2FL -X _1FL ) to X _DRR (= X _2RR -X)
_1RR ), the damping force coefficient C for performing the skyhook control is determined based on the relative speeds X _DFL ′ to X _DRR ′, and the step motor 4 corresponding to the determined damping coefficient C is determined.
Calculates a target step angle theta _T of 1FL～41RR, this calculates the difference value between the target step angle theta _T and the current step angle theta _P, the motor driving circuit a step control amount corresponding to 5
And outputs the 9FL～59RR, calculates a front wheel steering angle [delta] _f based on the steering angle detected value theta _S of the steering angle sensor 54S, then the steering angle ratio k of the front and rear wheels based on the vehicle speed detection value V of the vehicle speed sensor 53 Is calculated based on the steering angle ratio k.
calculating a _r, and outputs a switching control signal CS _ra and CS _rb as difference value between the rear wheel steering angle [delta] _r and the rear wheel steering angle detected value [delta] _rd becomes zero, further vertical acceleration detection value X _2FL " It detects the frequency f _FL ~f _RR of to X _2RR ", when the frequency f _FL ~f _RR is in sprung resonance frequency range, the yaw rate feedback control in and the steering control process performed skyhook control in the damping force control process the gain k _P a high control gain, when in the unsprung resonance frequency range, the damping force control processing is held and steering control processing high damping force state control gain k _P a high control gain, sprung resonance frequency range and when in the middle frequency range intermediate the unsprung resonance frequency is set to a low standard control gain control gain k _P than the height control gain in and steering control processing performed skyhook control in the damping force control processing.

The storage device 56d stores the arithmetic processing device 5
A program necessary for the arithmetic processing of FIG. 6c is stored in advance, necessary values in the arithmetic processing process and arithmetic results are sequentially stored, and a target yaw rate map for calculating a target yaw rate is stored in advance. Here, as shown in FIG. 13, the target yaw rate map has a steering angle detection value θ _S of the steering angle sensor 54S on the X axis and a vehicle speed sensor 53 on the Y axis.
And a target yaw rate Y _O on the Z-axis, respectively, and a three-dimensional map that peaks when the steering angle detection value θ _S is 90 degrees and the vehicle speed detection value V is 60 km / h, for example. I have.

Next, the operation of the above embodiment will be described with reference to a micro computer.
Of the damping force control processing of the arithmetic processing unit 56c of the computer 56
FIG. 10 showing an example, FIG. 11 showing an example of a steering control process, and
A description will be given with reference to FIG. 12 showing an example of the attenuation coefficient setting process.
You. That is, the damping force control processing of FIG.
(For example, every 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), and then proceeds to step S2, where each phase is read.
Displacement detection value X_DiAnd then proceed 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′ Are calculated and these are stored in the storage device 56 d
Is temporarily stored in a predetermined storage area, and then the process proceeds to step S4.
And the relative displacement detection value X read in step S2._DiThe example
For example, by performing high-pass filter processing,
Speed X_Di′ Are calculated, and these are stored in a predetermined
After the temporary storage in the storage area, the process proceeds to step S5.

In step S5, it is determined whether or not the damping force control flag F has been set to "1" in a control gain setting process of FIG. 12 described later, and the damping force control flag F is reset to "0". If so, the process proceeds to step S6 to calculate the following formula (1) based on the vehicle body vertical speed X _2i ′ and the relative speed X _Di ′ and the control gain C _S calculated in steps S3 and S4, and calculate the sky. After calculating the damping coefficient C for performing the hook control, the process proceeds to step S8, and when the damping force control flag F is set to "1", the process proceeds to step S7, where the damping coefficient C is set in advance. transition from set to C _MAX to step S8.

C = C _{S ×} X _2i ′ / X _Di ′ (1) In step S8, the damping coefficient C calculated in step S6 or S7 is set to a predetermined value by the damping force variable shock absorber 3i. It is determined whether or not the minimum damping force is less than or equal to the minimum damping force _CMIN . If C> _CMIN , the process proceeds to step S9 to determine whether or not the vehicle body vertical speed _X2i 'is positive.
_{When 2i} ′> 0, the process proceeds to step S10,
In order to set the damping coefficient C calculated in step S6 or S7 on the extension side, θ _{A of} the control map corresponding to FIG.
After calculating the target step angle θ _T with reference to the area of θ _B1, the process proceeds to step S11.

In step S11, the storage device 56d
Current step angle θ stored in _PAnd target step angle
θ_TIs calculated as a step control amount S.
While updating and storing in a predetermined storage area of the storage device 56d,
The target step angle θ_TIs the current step angle θ_PUpdate as
New storage is performed, and then the process proceeds to step S12,
Step control stored in a predetermined storage area of the storage 56d
Outputs the amount S to the motor drive circuit 59i and then interrupts the timer
The process ends and returns to the predetermined main program.

The result of the determination in step S9 is X _2i ′ <
When the value is 0, the process proceeds to step S13, and refers to the range of θ _B2 to θ _{C in} the control map corresponding to FIG. 8 so that the damping coefficient C calculated in step S6 or S7 is set on the pressure side. After calculating the target step angle θ _T , the process proceeds to step S11. Further, when the result of the determination in step S8 is C ≦ C _MIN , the process _proceeds to step S14.
The process proceeds to, migrate with reference to the region of the theta _B1 through? _B2 of the corresponding control map in FIG. 8 calculates the target step angle theta _T in the step S11.

The processing in FIG. 10 corresponds to the damping force control means. The steering control process of FIG. 11 is executed as a timer interrupt process every predetermined time (for example, 20 msec), similarly to the above-described damping force control process. First, in step S21, the vehicle speed detection value V of the vehicle speed sensor 53 and the steering angle sensor 54S the steering angle detection value theta _S, the yaw rate detected value Y of the yaw rate sensor 55
_D 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 S22, in which the steering angle detection value θ _S is divided by the steering gear ratio N to obtain the front wheel steering angle δ _F (=
θ _S / N) is calculated.

Next, the routine proceeds to step S23, where the following equation (2) is calculated based on the detected vehicle speed V to calculate the steering angle ratio k of the front and rear wheels. k = {bL−mV ² (a / C _r )} / {aL−mV ² (a / C _f )} (2) Next, the process proceeds to step S24, where the vehicle speed detection value V and the steering angle detection value θ Based on _S , the target yaw rate Y _O is calculated with reference to the target yaw rate map of FIG. 13, and then the process proceeds to step S25, where the deviation between the target yaw rate Y _O and the yaw rate detection value Y _D read in step S21. ε (=
Y _O −Y _D ), and then proceeds to step S26 to calculate the yaw rate deviation differential value ε ′ by differentiating the yaw rate deviation ε by, for example, a high-pass filter process, and then proceeds to step S27 to obtain the following (3) ) is calculated rear wheel steering angle [delta] _r by performing the calculation of the equation.

Δ _r = k · δ _f + k _P · ε + k _D · ε ′ (3) Here, k _P is a yaw rate feedback control gain, and is a storage device in a control gain setting process of FIG. 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. Then, the process proceeds to step S28, the rear wheel steering angle [delta] _r
Deviation between the rear wheel steering angle detected value [delta] _rd and _{Δδ r (= δ r -δ rd} )
Calculates, when the difference value .DELTA..delta _r is zero, the servo valve 8
It is both a logic value "0" to the control signal CS _ra and CS _rb for 5, when a difference value .DELTA..delta _r is positive (Δδ _r> 0), the control signal CS _ra to the logical value "1", the control signal CS _rb
To logic value "0", when the difference value .DELTA..delta _r is negative (Δδ _r <0), the control signal CS _ra to the logical value "0", respectively set the control signal CS _rb to the logical value "1" Drive circuit 60
a, 60b, then terminates the timer interrupt process and returns to the predetermined main program.

The processing in FIG. 11 corresponds to the steering control means. Further, the control gain setting process of FIG. 12 is executed as a timer interrupt process every predetermined time (for example, 20 msec), similarly to the above-described damping force control process and steering control process. First, in step S31, the vertical acceleration detection value X _{2FL is} detected. ″ ~
X _2RR ″ is read, and then the process proceeds to step S 32.
The frequency f _{FL to} f _RR is calculated by measuring the number of zero crossings per unit time of the vertical acceleration detection values X _2FL ″ to X _2RR ″ or the time from zero cross to zero cross.

Then, the process proceeds to step S33, in which the calculated frequencies f _{FL to} f _RR are adjusted to the sprung resonance frequency range (1 to 2 Hz).
It is determined whether or not the damping force is within the sprung resonance frequency range, and the process proceeds to step S34 to reset the damping force control flag F to “0” indicating that the damping force is skyhook controlled. Then, the yaw rate feedback control gain k _P in the steering control processing is set to the high control gain k _PH , 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 processing returns to the predetermined main program. I do.

If the result of the determination in step S33 is not in the sprung resonance frequency range, the flow shifts to step S35, where the frequencies f _{FL to} f _RR are changed to the unsprung resonance frequency range (10 to 10).
It is determined whether or not the damping force is in the unsprung resonance frequency range. If it is in the unsprung resonance frequency range, the process proceeds to step S36 to set the damping force control flag F to “1” indicating that the damping force is maintained in the high damping force state. At the same time, the yaw rate feedback control gain k _P is set to the high control gain k _PH , which is updated and stored in a predetermined storage area of the storage device 56d, and then the timer interrupt processing is ended to return to the predetermined main program. .

Further, the result of the determination in step S35 is not the unsprung resonance frequency range but the intermediate frequency range (about 4 to 8 Hz) between the sprung resonance frequency range and the unsprung resonance frequency range, or is equal to or higher than the unsprung resonance frequency range. when the frequency, the process proceeds to step S37, the damping force control flag F is reset to "0", sets the yaw rate feedback control gain k _P at a high control gain k lower than _PH standard control gain k _PN, This is updated and stored in a predetermined storage area of the storage device 56d, and then the timer interrupt processing is terminated to return to a predetermined main program.

Therefore, if it is assumed that the vehicle is traveling at a constant speed on a flat, good road, the vertical acceleration of the vertical acceleration sensors 51FL to 51RR is small since there is almost no vertical movement of the vehicle in this state.
The vertical acceleration detection values X _2FL ″ to X _2RR ″ output from are substantially zero. Therefore, when the control gain setting process of FIG. 12 is executed, the vehicle body vertical speed frequencies f _{FL to} f _RR calculated in step S32 also become zero. Therefore, the process proceeds to step S37 via steps S33 and S35, the yaw rate feedback control gain k _P is set to the standard control gain k _PN , and the damping force control flag F is reset to “0”.

For this reason, when the damping force control process of FIG. 10 is executed, the vehicle body vertical speeds X _2FL ′ to X _2RR ′ calculated in step S3 become substantially zero, and the damping force control flag F
Has been reset to "0", the damping coefficient C calculated from step S5 to step S6 also becomes substantially zero, so the step moves from step S7 to step S13, where the extension side and compression side minimum damping coefficients are calculated. the step angle in the range of C _Nmin and C _Amin a step angle theta _B1 through? _B2 is set as the target step angle theta _T, the step motor 41FL~
Step angle 41RR are driven so as to match the target step angle theta _T. Therefore, the valve element 31 of variable damping force shock absorber 3FL~3RR is set to the position B shown in FIG. 6, whereby the minimum s damping coefficient C of the extension side and the compression side husband damping piston 8 coefficients C _Nmin and C _Amin Is set to Therefore, in this state, even if vibrations due to fine irregularities on the road surface are input to the wheels, the vibrations are absorbed by the damping force variable shock absorbers 3FL to 3RR and are not transmitted to the vehicle body.
Good ride comfort can be ensured.

On the other hand, when the steering control process shown in FIG. 11 is executed, the vehicle is in a straight running state, so that the detected steering angle θ
_{Since S} is zero and the target yaw rate Y _O calculated in step S25 is also zero, the rear wheel steering angle δ _r calculated in step S27 is also zero, so that the control signals CS _ra and CS _ra
rb both assume the logical value “0” and maintain the straight traveling state.

In this good road running state, for example,
When passing through a temporary step such as a difference,
When the vehicle does not move up and down due to
_2FL'~ X_2RR′ Remain zero, so that the minimum damping coefficient C
_aMINAnd C_nMINWheels ride on steps to maintain condition
Although it can absorb the pushing force when raised,
Riding on a relatively large step, absorbing the thrust
Otherwise, the vehicle will be displaced upward.
Therefore, the vehicle vertical speed X_2FL'~ X_2RR′ Is positive
Will increase. Thus, the vehicle vertical speed X
_2FL'~ X_2RR′ Increases in the positive direction, step S7
And the step angle θ in FIG._A~ Θ_B1Attenuation in the area
Target step angle θ according to coefficient C_TIs calculated, so
The damping force variable shock absorber 3FL ~ 3RR valve body 31
Switching control is performed as shown in FIG. As a result, step riding
Relative speed X by raising_DFL'~ X_DRR′ Is negative, ie the body
Side displacement speed X_2i′ To the wheel side displacement speed X_1i'But
When the piston 8 moves to the compression side at high speed, the compression
Small damping coefficient C_aMINVibration to the wheel side
Input can be absorbed, and from this state the step
As a result, the rising speed on the wheel side is
Relative speed X_DFL'~ X_DRR′ Is positive
As a result, the piston 8 moves to the extension side. At this time
Means that the damping coefficient C has a large value,
Demonstrate the damping effect, then stop the lift of the body
And body vertical speed X_2FL'~ X_2RR′ Becomes zero
As described above, the step motors 41FL to 41RR are
It is rotated counterclockwise to return to position B, thereby
The minimum damping coefficient C on both the compression side and the extension side_aMINAnd C_nMINTo
Control, and then when the vehicle begins to descend,
Body speed X_2FL'~ X _2RR′ Increases in the negative direction
As a result, the process moves from step S9 to step S13.
And the step angle θ with reference to the control map of FIG._B2~ Θ_C
Target step angle θ according to the damping coefficient C in the range of_TCalculate
As a result, the valve element 31 is further rotated counterclockwise.
Then, it is turned to the turning position shown in FIG. Because of this, the car
Body descends and relative speed X_DFL'~ X_DRR′ Is negative
When the piston 8 moves to the compression side, the damping force is large.
Greater vibration exerts a great damping effect.

Conversely, when the wheels pass through the step of descending forward, the relative speeds X _DFL ′ to X _DRR ′ increase in the forward direction due to the rebound of the wheels, but the vehicle does not move up and down at this time. Body vertical speed X _2FL '~ X
_{Since 2RR} 'is zero, the damping coefficients of the damping force variable shock absorbers 3FL to 3RR are the minimum damping coefficients C _aMIN and C _{nM IN}
Maintaining, allowing the lowering of the wheels, then the vehicle starts to descend, the vehicle body vertical velocity X _2FL '~X _2RR' is increased in the negative direction, is the attenuation coefficient C is a large value, the step angle θ will _{be B2} through? _C target step angle theta _T ranging is calculated, since the valve body 31 is rotated to the position shown in FIG. 7, a large damping force against movement of the compression side of the piston 8 As a result, the valve body 31 is rotated clockwise as the vehicle vertical speeds X _2FL ′ to X _2RR ′ become smaller and the damping coefficient C becomes smaller. Return to the B side, and the vehicle vertical speed _X2FL '~
When X _2RR ′ becomes zero, the valve element 31 is at the position B, and has the minimum damping coefficients C _aMIN and C _nMIN . Thereafter, when the vehicle body starts rising by _swinging back, the vehicle body vertical speeds X _2FL ′ to X _2RR ′ increase in the positive direction, and the relative speeds X _DFL ′ to X _DRR ′ become the positive direction. is the target step angle theta _T as a step angle theta _a side with an increase in the calculation, by the valve body 31 is a position shown in FIG. 5 is rotated in the clockwise direction, the movement of the extension side of the piston 8 , A large damping force can be given to exert a vibration damping effect. In this way, when passing through a temporary step while traveling on a good road, a good vibration suppression effect can be exerted by the skyhook control, and even when traveling on a bad road, Body vertical speed X _2FL '~
'By target step angle theta _T positive (or negative) by the step angle theta _A side (or step angle theta _C side) is calculated, the relative speed X _DFL body rises' X _2RR to X
_DRR 'is negative and the body descends and the relative speed X _DFL ' ~ X
_{When DRR} 'is in the excitation direction where it is positive, the damping coefficient C is controlled to the minimum damping coefficients C _aMIN and C _nMIN , and conversely, the vehicle body rises and the relative speeds X _DFL ' to X _DRR 'become positive and the vehicle body becomes When the relative velocity X _DFL 'to X _DRR ' is lowered and the relative velocity X _DFL 'to X _DRR ' is negative, the damping coefficient C is _{adjusted according} to the vertical velocity X _2FL 'to X _2RR ' and the relative velocity X _DFL 'to X _DRR '. By controlling to the optimum damping coefficient, a good ride comfort can be secured.

Even when the vehicle is traveling on a rough road, the vehicle body rises and the relative speeds X _DFL 'to X
_DRR 'is negative and the body descends and the relative speed X _DFL ' ~ X
_{When DRR} 'is in the excitation direction where it is positive, the damping coefficient C is controlled to the minimum damping coefficients C _aMIN and C _nMIN , and conversely, the vehicle body rises and the relative speeds X _DFL ' to X _DRR 'become positive and the vehicle body becomes When the relative velocity X _DFL 'to X _DRR ' is lowered and the relative velocity X _DFL 'to X _DRR ' is negative, the damping coefficient C is _{adjusted according} to the vertical velocity X _2FL 'to X _2RR ' and the relative velocity X _DFL 'to X _DRR '. By controlling to an optimal damping coefficient, a good ride comfort can be secured.

When the steering wheel 76 is turned right (or left) from a straight running state on a good road, and then left (or right) turned to a slalom running state, the turning outer wheel is first attached to the vehicle body. The left wheel side sinks, and the right wheel side, which is the turning inner wheel side, rises and rolls downward, as viewed from the rear wheel side. As a result, a downward-sloping roll is produced. Thus, when the roll by slalom occurs,
In response to this, the vertical acceleration sensors 51FL and 51
In RL, the vehicle vertical acceleration detection values X _2FL ″ and X _2RL ″ increase from zero in the negative direction and then increase in the positive direction.
Would be increased in the negative direction after the right wheel side of the vertical acceleration sensor 51FR and 51 _RR in the vehicle body vertical acceleration detection value X _2FR "and X _2RR" is increased from zero in the positive direction, the process of FIG. 12 is executed At this time, when the vertical acceleration frequencies f _{FL to} f _RR increase from zero and are in the sprung resonance frequency range (about 1 to 2 Hz), steps S33 to S3 are performed.
4, the yaw rate feedback control gain k
_P is set to the high control gain k _PH and the damping force control flag F is reset to “0”.

For this reason, when the slalom running is started, the control state in which the ride comfort during the straight running immediately before the running is emphasized is continued, but when the vertical acceleration frequency reaches the sprung resonance frequency range, the yaw rate feedback control is performed. By setting the gain k _P to the high control gain k _PH , when the processing in FIG. 11 is started, the correction term of the yaw rate deviation ε in step S27 becomes a large value, so that the responsiveness of the rear wheel steering is improved. While the control state shifts to a control state that emphasizes steering stability, while the damping force control processing in FIG.
Since the damping force control flag F is set to "0" in the same manner as in the straight running state, the skyhook control is continued, and the control emphasizing the riding comfort is continued. For this reason,
Even if the wheel load changes due to the damping force control, the steering stability can be secured by the yaw rate feedback control by the rear wheel steering, so that the slalom traveling can be performed stably.

In addition, the vehicle passes over rough roads during slalom traveling.
The vertical acceleration frequency f_FL~ F_RRIs sprung
Beyond the resonance frequency and become the intermediate frequency range (about 4 to 8 Hz)
Then, when the processing of FIG. 12 is executed, step S3
After going through step S35 and step S37, the
Feedback control gain k_PIs the standard control gain k _PN
And the damping force control flag F is reset to "0".
The responsiveness of maneuvering stability is reduced because the set state is maintained.
The driver's sense of discomfort by suppressing the steering stability control feeling
And change to a control state that emphasizes ride comfort.
It is.

Further, when the vehicle travels on a road surface where fine irregularities are continuous during the slalom traveling, the vertical acceleration frequencies f _{FL to} f _RR become lower than the unsprung resonance frequency range (about 10 to 12 Hz).
When the process of FIG. 12 is executed, the process proceeds from step S35 to step S36, the yaw rate feedback control gain k _P is increased to the high control gain k _PH , and the damping force control flag F Is "1"
When the damping force control process of FIG. 10 is executed, the process proceeds from step S5 to step S7, where the damping coefficient is set to the maximum damping coefficient C _MAX , so that the unsprung flutter is greatly reduced. In addition, when the steering control process of FIG. 11 is executed, the correction term of the yaw rate deviation ε in step S27 becomes a large value. Stability can be improved, and control with emphasis on steering stability is performed.

In the above embodiment, the case where the valve element 31 for controlling the damping force is formed in a rotary type is described. However, the present invention is not limited to this. Different flow paths may be formed on the side and in this case, in this case, step motors 41FL to 4FL
A pinion may be connected to the rotation shaft 41a of the 1RR, and a rack that meshes with the pinion may be attached to the connection rod 42 or an electromagnetic solenoid may be used to control the sliding position of the valve body 31. Instead of changing the damping force, a damping force variable shock absorber capable of switching the damping force in a plurality of stages can be applied.

In the above embodiment, the rear wheel auxiliary steering cylinder 79 is provided with a closed sensor type servo valve 85.
However, the present invention is not limited to this, but is applicable to an open center type servo valve, and the piston rod 79a of the four-wheel drive 79 is returned to the neutral position accordingly. The control may be performed by interposing a spring.

Further, in the above-described embodiment, the case where the yaw rate feedback control of the above-described formula (3) is performed in the rear wheel steering control is described. However, the present invention is not limited to this. May be changed according to the conditions. Furthermore, in the above embodiment,
In the damping force control, the case of switching between the skyhook control and the high damping force fixed control has been described. However, the present invention is not limited to this, and the damping force may be finely changed according to the vertical acceleration frequency, and similarly in the steering control. Alternatively, the control gain may be finely changed according to the vertical acceleration frequency.

Furthermore, in the above-described embodiment, the case where the control is performed by using the microcomputer 56 has been described. However, the present invention is not limited to this, and the integration for integrating the outputs of the zero-cross detector and the vertical acceleration sensor 51i is performed. An electronic circuit such as a device, a differentiating circuit for differentiating the output of the stroke sensor 52i, and a function generator may be combined.

In the above embodiment, the case where the potentiometer is used as the stroke sensor has been described. However, the present invention is not limited to this. The ultrasonic distance sensor for detecting the relative distance between the vehicle body and the road surface, the detection coil Any relative displacement detection means such as a displacement sensor that detects displacement by impedance change or inductance change using the above method can be applied.

Furthermore, in the above embodiment, the vertical acceleration sensors 51FL to 51FL are located at the respective wheels 1FL to 1RR of the vehicle body 2.
Although the case where the RR is provided has been described, any one of the vertical acceleration sensors may be omitted, and the vertical acceleration at the omitted position may be estimated from the values of the other vertical acceleration sensors. Furthermore, in the above embodiment, the case where the step motors 41FL to 41RR are controlled by open loop has been described. However, the present invention is not limited to this, and the rotation angle of the step motors is detected by an encoder or the like, and this is fed back to perform closed loop control. You may make it.

[0054]

As explained above, according to the first aspect of the present invention,
According to the vehicle characteristic control device for a wheel-steered vehicle, the vertical acceleration of each wheel position of the vehicle body is detected, the frequency of the detected vertical acceleration is detected, and the damping force control unit and the steering control unit are detected according to the detected frequency. Is changed, the damping force characteristic and the steering characteristic can be controlled to an optimum state according to the vertical movement of the vehicle body.

According to the vehicle characteristic control apparatus for a four-wheel steering vehicle, when the frequency of the vehicle body vertical acceleration is in the sprung resonance frequency range, the steering control gain is increased and the damping force control is performed. By controlling the gain based on the vehicle vertical speed and the relative speed, control that places importance on both steering stability and ride comfort is performed, and the frequency of the vehicle vertical acceleration is between the sprung resonance frequency range and the sprung resonance frequency range. When the vehicle is in the intermediate frequency range, the steering control gain is reduced and the damping force control gain is controlled based on the vehicle body vertical speed and the relative speed, so that control is performed with emphasis on ride comfort, and the control according to the vehicle body vertical movement is performed. The effect that optimal characteristic control can be performed is obtained.

[Brief description of the drawings]

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

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

FIG. 3 is a partially sectional front view showing an example of a variable damping force shock absorber.

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

FIGS. 5A and 5B are enlarged cross-sectional views showing a damping force adjusting mechanism in an intermediate damping force state when the vehicle body is lifted, wherein FIG. 5A shows a hydraulic oil path on the extension side and FIG.

FIGS. 6A and 6B are enlarged cross-sectional views showing a damping force adjustment mechanism when the vehicle body does not fluctuate, wherein FIG. 6A shows a hydraulic oil path on the extension side and FIG.

FIGS. 7A and 7B are enlarged cross-sectional views showing a damping force adjusting mechanism in a state of maximum damping force when the vehicle body descends, wherein FIG. 7A shows a hydraulic oil path on the extension side and FIG.

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

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

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

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

FIG. 12 is a flowchart illustrating 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 detected steering angle, a detected vehicle speed, and a target yaw rate.

[Explanation of symbols]

 1FL to 1RR Wheel 2 Body 3FL to 3RR Damping force variable shock absorber 4 Controller T1 to T3 Expansion 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-59RR Motor drive circuit 76 Steering wheel 79 Rear wheel auxiliary steering cylinder 85 Servo valve

Claims (2)

(57) [Claims] 1. A suspension having a variable damping force shock absorber, an auxiliary steering device for auxiliary steering of at least a rear wheel, and damping force control means for controlling a damping force of the variable damping force shock absorber in accordance with a vertical movement of a vehicle body. When,
A vehicle characteristic control apparatus for a four-wheel steering vehicle, comprising: a steering control means for controlling the auxiliary steering device according to a steering angle and the like. A vehicle vertical acceleration detection means for detecting a vehicle vertical acceleration, and a vehicle vertical acceleration detection means Vertical acceleration frequency detecting means for detecting the frequency of the vehicle vertical acceleration, and control gain changing for changing control gains in the damping force control means and the steering control means according to the vehicle vertical acceleration frequency detected by the vertical acceleration frequency detecting means. And a vehicle characteristic control device for a four-wheel steering vehicle. 2. The control gain changing means changes a control gain in the damping force control means in accordance with a vehicle vertical speed and a relative displacement speed when the vertical acceleration frequency is near a sprung resonance frequency, and further comprises a steering control means. In the intermediate frequency region between the sprung resonance frequency and the unsprung resonance frequency, the control gain of the damping force control means is changed according to the vehicle vertical speed and the relative displacement speed, and the steering The vehicle characteristic control device for a four-wheel steering vehicle according to claim 1, wherein the control gain of the control means is set to a control gain lower than the high control gain.