JP2001001733A - Damping coefficient control system for vehicle - Google Patents

Abstract

PROBLEM TO BE SOLVED: To improve stability and controllability of vehicle by minimizing effect of changes in steering characteristics in a transient steering of vehicle. SOLUTION: In this damping coefficient control system, deviation Δγ between a vehicle actual yaw velocity γf and a reference vehicle yaw velocity γt is calculated (S30, S40) and vehicle steering condition is judged (S50 to S70). When the vehicle is in over-steering condition virtual shock absorber damping coefficients Cgf, Cf at a front wheel side are corrected and increased according to the deviation Δγ, and virtual shock absorber damping coefficients Cgr, Cr at a rear wheel side are corrected and decreased according to the deviation Δγ(S80, S100) When the vehicle is in under-steering condition, virtual shock absorber damping coefficients Cgf, Cf at a front wheel side are corrected and decreased according to the deviation Δγ, and virtual shock absorber damping coefficients Cgr, Cr at a rear wheel side are corrected and increased according to the deviation Δγ (S90, S100).

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a damping coefficient control device for a vehicle such as an automobile, and more particularly to a damping coefficient control device improved so as to improve the dynamic performance of the vehicle during a transient turn.

[0002]

2. Description of the Related Art One example of a damping coefficient control device for a vehicle such as an automobile provided with a shock absorber having a variable damping coefficient corresponding to each wheel is disclosed in, for example, Japanese Patent Application No. Hei 10-210, filed by the present applicant. The specification and the drawings of JP-A-10-92675 include means for detecting turning information of a vehicle, means for determining a change in a roll amount of a vehicle, and a damping coefficient of a shock absorber inside a turn in a process of increasing the roll amount of a vehicle. Means for controlling the damping coefficient relatively higher than the damping coefficient of the shock absorber on the outside of the turn.

According to the damping coefficient control device according to the prior proposal, in the process of increasing the roll amount of the vehicle body, the damping coefficient of the shock absorber inside the turning is relatively higher than the damping coefficient of the shock absorber outside the turning. As a result, the damping force of the shock absorber on the inside of the turning acting downward is controlled to be relatively higher than the damping force of the shock absorber on the outside of the turning acting upward, so that the downward force acting on the vehicle body as a whole is Is increased, thereby reducing the vehicle height and improving the kinetic performance of the vehicle during a transient turn.

[0004]

In the damping coefficient control device according to the above-mentioned proposal, the vehicle laterally moves a predetermined distance from the vehicle body to a side estimated to lift with respect to the center of gravity of the vehicle body based on the turning state of the vehicle. A first virtual shock absorber that has a virtual swing center of the vehicle body at a virtual position spaced apart from the virtual swing center and acts around the virtual swing center, and a second virtual shock absorber that acts vertically at the virtual position. Based on a vehicle model having a shock absorber, an actual damping coefficient of a variable shock coefficient provided for each wheel is controlled.

In general, when the vehicle is turning, the steering characteristics of the vehicle may change to the understeer side or the oversteer side due to the vehicle speed, acceleration / deceleration, steering angle, road surface conditions, and the like. As a result, the steering characteristic changes due to the load movement in the vehicle longitudinal direction due to acceleration and deceleration. However, in the damping coefficient control device according to the above-mentioned proposal, since the parameters of the vehicle model, such as the virtual damping coefficients of the first and second virtual shock absorbers, are constant, the steering when the vehicle makes a turn is determined. There is a problem that the actual damping coefficient of the shock absorber cannot be appropriately controlled so as to reduce the influence of the change in the characteristics.

The present invention is based on a vehicle model having a first virtual shock absorber acting around a virtual swing center and a second virtual shock absorber acting vertically at a virtual position. SUMMARY OF THE INVENTION The present invention has been made in view of the above-described problem in the damping coefficient control device according to the above-described proposal, which is configured to control the damping coefficient of an absorber. Variably set the front and rear wheel distribution ratio of the virtual damping coefficient of the first or second virtual shock absorber according to the conditions, to reduce the change in the steering characteristics during transient turning of the vehicle and improve the steering stability of the vehicle It is to make it.

[0007]

SUMMARY OF THE INVENTION According to the present invention, there is provided, in accordance with the present invention, a damper for a vehicle provided with an actual shock absorber having a variable damping coefficient corresponding to each wheel. A coefficient control device, a means for detecting a turning state of the vehicle, a means for detecting a change in the steering characteristic of the vehicle, and a means for estimating a lift relative to the center of gravity on the spring based on the turning state of the vehicle. A first virtual shock absorber having a virtual swing center on the sprung at a virtual position separated from the sprung by a predetermined distance in the vehicle lateral direction and acting around the virtual swing center and the virtual shock absorber; A front and rear wheel model having a second virtual shock absorber acting in the vertical direction at the position, and the first or second virtual shock absorber according to a change in steering characteristics of the vehicle. Virtual damping coefficient setting means for variably setting a distribution ratio between the front and rear wheels of the virtual damping coefficient of the sober; means for calculating a target damping coefficient of the actual shock absorber based on at least the virtual damping coefficient; Means for controlling the actual damping coefficient of the shock absorber based on the above.

According to the first aspect of the present invention, the spring is located at a virtual position spaced a predetermined distance laterally from the sprung to the side where it is estimated to lift with respect to the center of gravity of the sprung based on the turning state of the vehicle. Front and rear wheel sides having a first virtual shock center and having a first virtual shock absorber acting around the virtual swing center and a second virtual shock absorber acting vertically at a virtual position Since the vehicle model has the wheel side, the second virtual shock absorber suppresses the lift of the outer wheel side on the sprung by the second virtual shock absorber during the transient turning of the vehicle, thereby lowering the center of gravity on the sprung and the steering characteristic of the vehicle. The distribution ratio between the front and rear wheels of the virtual damping coefficient of the first or second virtual shock absorber is variably set according to the change of the actual shock absorber. Appropriate control is performed according to the change in the steering characteristic of the vehicle, whereby the distribution ratio between the front and rear wheels of the virtual damping coefficients of the first and second virtual shock absorbers is constant regardless of the change in the steering characteristic of the vehicle. The change in the steering characteristic at the time of the transient turning is reduced as compared with the case.

According to the present invention, in order to effectively attain the above-mentioned main object, the means for detecting a change in the steering characteristic of the vehicle may be configured to detect the turning behavior of the vehicle. The virtual damping coefficient setting means is configured to variably set a distribution ratio of the virtual damping coefficient between the front and rear wheels according to a turning behavior of the vehicle.

According to the configuration of the second aspect, the turning behavior of the vehicle is detected as a change in the steering characteristic of the vehicle, and the distribution ratio of the virtual damping coefficient between the front and rear wheels is variably set according to the turning behavior of the vehicle. Therefore, irrespective of the turning behavior of the vehicle, the steering characteristic during the transient turning is less than that in the case where the distribution ratio of the virtual damping coefficients of the first and second virtual shock absorbers between the front and rear wheels is constant. Deterioration of the turning behavior of the vehicle due to this is reliably reduced.

According to the present invention, in order to effectively attain the above-mentioned main object, the means for detecting a change in the steering characteristic of the vehicle may be arranged such that the means for detecting a change in the steering characteristic of the vehicle is provided with a pitch on the spring. The virtual damping coefficient setting means is configured to detect a motion state quantity and variably set a distribution ratio of the virtual damping coefficient between the front and rear wheels according to the pitch motion state quantity on the spring. ).

According to the configuration of the third aspect, the amount of pitch motion on the spring is detected as a change in the steering characteristic of the vehicle, and the virtual damping coefficient is distributed between the front and rear wheels according to the amount of pitch motion on the spring. Since the ratio is variably set, the change in the steering characteristic due to the pitch movement on the spring is reliably reduced.

According to the present invention, in order to effectively achieve the above-mentioned main object, in the structure of the third aspect, the pitch motion state amount on the spring includes acceleration / deceleration on the spring. The virtual damping coefficient setting means is configured to variably set the distribution ratio of the virtual damping coefficient between the front and rear wheels in accordance with the acceleration / deceleration on the spring.

According to the configuration of the fourth aspect, the distribution ratio of the virtual damping coefficient between the front and rear wheels is variably set in accordance with the acceleration / deceleration on the spring, so that the actual damping coefficient of the shock absorber is adjusted on the spring. Appropriate control is performed in accordance with the speed, whereby the change in the steering characteristic due to the load movement in the longitudinal direction of the vehicle is reliably reduced.

According to the present invention, in order to effectively achieve the above-described main object, in the configuration of the fourth aspect, the acceleration / deceleration on the spring is estimated based on the amount of acceleration / deceleration operation by the driver. (According to claim 5).

According to the fifth aspect of the present invention, the acceleration / deceleration on the spring is an estimated acceleration / deceleration estimated based on the amount of acceleration / deceleration operation by the driver. Compared to when the distribution ratio between the front and rear wheels is controlled according to the detected acceleration / deceleration, it is possible to control appropriately according to the acceleration / deceleration above the damping coefficient of the actual shock absorber without response delay compared to when the distribution is controlled according to the detected acceleration / deceleration. Become.

[0017]

As shown in FIG. 6, a two-wheel model of an actual vehicle has a vehicle body 110 supported by left and right wheels 112L and 112R, and a suspension between the vehicle body 110 and the wheels 112L and 112R. Springs 114L and 114R and shock absorber 11
6L and 116R are shown as being disposed.

In the actual vehicle model shown in FIG. 6, for example, when the vehicle turns left and an inertial force acts on the vehicle body 110 to the right, a roll moment Mroll acts outward on the vehicle body. Then, the roll moment becomes the left and right suspension springs 114L and 11L.
4R, and the damping forces Fal and Far of the left and right shock absorbers 116L and 116R are carried by the spring forces Fsl and Fsr. In the process of increasing the roll amount of the vehicle body, the moment in the roll restraining direction and the roll moment Mroll due to these forces. The vehicle body 110 rolls outward until the vehicle turns 110.

In this case, the suspension spring 114
L spring force Fsl increase and suspension spring 1
The amount of reduction of the spring force Fsr of 14R is substantially equal to each other,
Also, in a conventional vehicle, the damping coefficients of the left and right shock absorbers during turning are controlled to be equal to each other,
The damping forces Fal and Far of the left and right shock absorbers are also substantially equal to each other, so that the height of the center of gravity 118 of the vehicle does not substantially change.

On the other hand, as shown in FIG. 7, only the suspension springs 114L and 114R are disposed between the vehicle body 110 and the left and right wheels 112L and 112R, and the suspension springs 114L and 114R are disposed inside the turning with respect to the vehicle. Considering a virtual model in which one shock absorber 122 that generates a vertical damping force between a virtual wheel 120 and one shock absorber 124 that suppresses roll displacement of a vehicle body is considered, a roll moment Mroll is supported by the damping force Fas of the shock absorber 122 and the spring forces Fsl and Fsr of the left and right suspension springs 114L and 114R, and the increase in the vehicle height on the turning inner wheel side is reduced as compared with the conventional case. , Center of gravity 118
Height decreases.

Therefore, if control of the virtual model as shown in FIG. 7 can be achieved in the two-wheel model of the actual vehicle shown in FIG. 6, the height of the center of gravity 118 of the vehicle can be increased in the process of increasing the roll amount of the vehicle. Therefore, the kinetic performance of the vehicle at the beginning of turning can be improved.

As shown in FIG. 7, the spring constants of the left and right suspension springs 114L and 114R are K, the damping coefficient of the shock absorber on the turning outer wheel is Cout, the stroke of the turning outer wheel is Xout, and the turning inner wheel is Xout. The damping coefficient of the shock absorber 114L on the
In, the stroke of the turning inner wheel is Xin, the tread of the vehicle is W, the distance between the center of gravity 118 of the vehicle and the shock absorber 122 is L, and the shock absorber 12 is
Let Cg and C be the damping coefficients of 2 and 124, respectively.

Assuming that the mass of the vehicle body 110 is M, the vertical acceleration and the roll angular velocity of the vehicle body are Xbdd and θdd, respectively, and the stroke speeds of the outer and inner turning wheels are Xoutd and Xind, respectively, the virtual model shown in FIG. Equations 1 and 2 below are respectively established from the balance of the vertical force and the balance of the force around the center of gravity 118 in the above.

[0024]

(Equation 1)

Assuming that Cn = WC / 2 as a parameter for attenuating the roll motion of the vehicle body, the above equation 2 becomes the following equation 3
It is represented as

[0026]

(Equation 2)

The following formulas 4 and 5 are established from the balance of the vertical force and the balance of the force around the center of gravity 118 in the actual two-wheel model of the vehicle shown in FIG.

[0028]

(Equation 3)

From the above equations 1 and 4, the following equation 6 is established.

[0030]

(Equation 4)

If Cm = Cn / L, the following equation (7) is established from the above equations (3) and (5).

[0032]

(Equation 5)

Here, when the vertical force generated by the shock absorber 122 in the virtual model shown in FIG. 7 is set to T according to the following equation 8, the following equations 9 to 11 are established from the above equations 6 to 8. I do.

[0034]

(Equation 6)

[0035]

(Equation 7)

From Equations 9 and 11, the damping coefficient Cin of the shock absorber for the turning inner wheel can be obtained as follows.

[0037]

(Equation 8)

By substituting equation (12) into equation (9), the damping coefficient Cout of the shock absorber for the turning outer wheel can be obtained as follows.

[0039]

(Equation 9)

Further, by rearranging the above equations (12) and (13), the damping coefficient Cin of the shock absorber of the turning inner wheel and the turning outer wheel is obtained.
And Cout are represented by the following equations 14 and 15, respectively.

[0041]

(Equation 10)

The damping force generated by the shock absorbers on the turning inner wheel side and the turning outer wheel side is given by the following equation (1).
6 and Equation 17.

[0043]

[Equation 11]

Also, based on the same concept, in the process of decreasing the roll amount of the vehicle body, based on a virtual model in which virtual shock absorbers 122 and 124 are arranged outside the vehicle turning, the turning inner wheel side and the turning outer wheel side are used. The damping coefficients Cin and Cout of the shock absorber of
And the control as in Equation 19, the center of gravity 11 of the vehicle is obtained.
8, the height of the vehicle at the end of turning can be improved.

[0045]

(Equation 12)

As shown in FIG. 8, L, W, T, Cg, and C for the vehicle models on the front wheel side and the rear wheel side are respectively represented by Lf and Lr, Wf and Wr, Tf and Tr,
Let Cgf and Cgr, Cf and Cr be the stroke speeds of the inside turning front wheel and the outside turning front wheel Xfind and Xfo, respectively.
utd, the stroke speeds of the inside rear wheel and the outside rear wheel are Xrind and Xroutd, respectively, and Tf and Tr
In the process of increasing the roll amount of the vehicle body, the damping coefficient Cfin of the shock absorber of the front inside wheel and the damping coefficient Cfout of the shock absorber of the front outside wheel are respectively defined as It is preferable that the damping coefficient Crin of the shock absorber for the turning inside rear wheel and the damping coefficient Crout of the shock absorber for the turning outside rear wheel be calculated according to the following equations 24 and 25, respectively.

[0047]

(Equation 13)

[0048]

[Equation 14]

Therefore, according to another preferred embodiment of the present invention, in the structure of the first aspect, the damping coefficient Cfin of the shock absorber of the front wheel on the inside of the turning and the outside of the turning on the outside during the process of increasing the roll amount of the vehicle body. The damping coefficient Cfout of the front wheel shock absorber is calculated according to the above formulas 22 and 23, respectively. The damping coefficient Crin of the shock absorber of the turning inside rear wheel and the damping coefficient Crout of the shock absorbing rear wheel of the turning outside are respectively calculated by the formulas 24 and 24 above. 25 (preferred mode 1).

According to another preferred embodiment of the present invention, in the construction of the first aspect, the damping coefficient Cfin of the shock absorber of the front wheel on the inside of the turning and the outside of the turning on the outside during the process of reducing the roll amount of the vehicle body. The damping coefficient Cfout of the front wheel shock absorber is calculated according to the following equations (26) and (27), respectively. The damping coefficient Crin of the turning inside rear wheel shock absorber and the damping coefficient Crout of the turning outside rear wheel are respectively expressed by the following equation (28). And Equation 29 (preferred mode 2).

[0051]

(Equation 15)

According to another preferred embodiment of the present invention, in the configuration of the first aspect, the virtual damping coefficient setting means includes a roll stiffness on the front wheel side when the change in the steering characteristic of the vehicle is on the oversteer side. Is increased or the distribution ratio between the front and rear wheels of the virtual damping coefficient of the first or second virtual shock absorber is variably set so as to reduce the roll stiffness on the rear wheel side (preferred embodiment 3).

According to another preferred embodiment of the present invention, in the configuration of the first aspect, the virtual damping coefficient setting means includes a roll stiffness on the front wheel side when the change in the steer characteristic of the vehicle is on the understeer side. Is configured to variably set the distribution ratio between the front and rear wheels of the virtual damping coefficient of the first or second virtual shock absorber so as to reduce the stiffness or increase the roll stiffness on the rear wheel side (preferred mode 4).

According to another preferred embodiment of the present invention, in the configuration of the first aspect, the virtual damping coefficient setting means includes a front and rear wheel of the virtual damping coefficient of the first or second virtual shock absorber. The distribution ratio between the vehicle models is variably set, and the ratio between a predetermined distance of the vehicle model on the front wheel side and a predetermined distance of the vehicle model on the rear wheel side is variably set (preferred mode 5).

According to another preferred aspect of the present invention, in the configuration of the second aspect, the means for detecting a change in the steer characteristic of the vehicle includes a deviation between a reference yaw rate of the vehicle and an actual yaw rate. It is configured to determine the turning behavior of the vehicle based on this (preferred mode 6).

According to another preferred embodiment of the present invention, in the configuration of the above-mentioned preferred embodiment 6, the virtual damping coefficient setting means is configured to reduce the deviation between the reference yaw rate and the actual yaw rate by reducing the first or the second yaw rate. The configuration is such that the distribution ratio of the virtual damping coefficient of the second virtual shock absorber between the front and rear wheels is variably set (preferred mode 7).

According to another preferred embodiment of the present invention, in the configuration of claim 5, the acceleration / deceleration operation amount by the driver is configured to be a stroke of a brake pedal (preferred embodiment 8). .

According to another preferred embodiment of the present invention, in the configuration of claim 5, the acceleration / deceleration operation amount by the driver is configured to be a throttle opening speed (preferred embodiment 9). .

[0059]

BRIEF DESCRIPTION OF THE DRAWINGS The invention will now be described in more detail with reference to the accompanying drawings, in which several preferred embodiments are shown.

[0060] First Embodiment FIG. 1 is a schematic configuration diagram showing a first preferred embodiment of the damping coefficient control apparatus according to the present invention.

In FIG. 1, 10FL and 10FR denote left and right front wheels of the vehicle 12, respectively, and 10RL and 10RR denote left and right rear wheels, respectively. The left and right front wheels 10FL and 10FR, which are steered wheels, are steered via tie rods 18L and 18R by a rack and pinion type power steering device 16 driven in response to turning of the steering wheel 14 by the driver.

[0062] Shock absorbers 22FL to 22RR of variable damping coefficient are provided between the wheels 10FL to 10RR as unsprung parts and the vehicle body 20 as sprung parts, and the damping coefficients Ci (Ci ( i = fl, f
r, rl, rr) are controlled by the electric control device 24 when the vehicle turns, as described later.

The electric control device 24 includes a vehicle height sensor 26F.
Signals indicating strokes Xi (i = fl, fr, rl, rr) of wheels 10FL to 10RR from L, 26FR, 26RL, 26RR, signals indicating lateral acceleration Gy of the vehicle body from lateral acceleration sensor 28,
A signal indicating the yaw rate γ of the vehicle is input from the yaw rate sensor 30, a signal indicating the vehicle speed V from the vehicle speed sensor 32, and a signal indicating the steering angle δ from the steering angle sensor 34.

Although not shown in detail in the drawing, the electric control device 24 has, for example, a CPU, a ROM, a RAM, and an input / output port device, which are generally connected to each other by a bidirectional common bus. Microcomputer with a typical configuration. The vehicle height sensors 26FL to 26RR detect the stroke Xi of the wheel with the bounding direction of the wheel being positive, and the lateral acceleration sensor 28 and the steering angle sensor 34 detect the lateral acceleration and steering angle respectively with the positive left turning direction of the vehicle. .

Each of the electric control devices 24 corresponds to FIG.
The damping coefficients of the shock absorbers 22FL to 22RR are controlled based on the vehicle models on the front wheel side and the rear wheel side shown in (A) and (B). In particular, the electric control device 2 of this embodiment
4 determines whether or not the vehicle is in a transient turning state based on the lateral acceleration Gy according to the flowcharts shown in FIGS. 2 and 3 as described later. When the vehicle is in a steady turning state, the shock absorber of each wheel is determined. The damping coefficient Ci is controlled to a predetermined hard damping coefficient Chigh, so that even when the vehicle is in a transient turning state, the damping coefficient of the shock absorber inside the turning turns to the outside of the turning when the roll amount of the vehicle body increases. The damping coefficient of each shock absorber is controlled so that the damping coefficient of the shock absorber on the outside of the turn becomes higher than the damping coefficient on the inside of the turn in the process of decreasing the roll amount of the vehicle body. Control the
As a result, the vehicle height during a transient turn is reduced, and the center of gravity of the vehicle body is reduced.

The electric control unit 24 calculates the reference yaw rate γt of the vehicle based on the vehicle speed V and the steering angle δ, calculates the deviation Δγ between the actual yaw rate γ of the vehicle and the reference yaw rate γt, and calculates the deviation Δγ based on the deviation Δγ. It is determined whether the vehicle is in the oversteer state or the understeer state, and when the vehicle is in the oversteer state or the understeer state, the correction amount of the attenuation coefficient of each virtual shock absorber on the front wheel side and the rear wheel side so that the deviation Δγ becomes small. ΔCgf, ΔCg
r, ΔCf, ΔCr, and correction amounts ΔLf, ΔLr for a predetermined distance
Is calculated, and the actual damping coefficient of each shock absorber is controlled based on the calculation result.

Next, the control of the attenuation coefficient in the illustrated first embodiment will be described with reference to the flowcharts shown in FIGS. The control according to the flowcharts shown in FIGS. 2 and 3 is started by closing an ignition switch (not shown) and is repeatedly executed at predetermined time intervals.

First, in step 10, a signal indicating the stroke Xi of each wheel is read, and in step 20, a filtering process for removing a noise component from the detected yaw rate γ is performed. Calculates the yaw rate γf after the filter processing.

In step 30, the actual steering angle δf of the front wheels is calculated based on the steering angle δ, the target yaw rate γe is calculated according to the following equation 30 using H as a wheel base and Kh as a stability factor, and T Is used as a time constant, and s is a Laplace operator, and the reference yaw rate γt of the vehicle based on the vehicle speed V and the steering angle δ is calculated according to the following equation 31. The target yaw rate γe may be calculated in consideration of the dynamic yaw rate by taking into account the lateral acceleration Gy of the vehicle. γe = Vδ / (1 + KhV ² ) H (30) γt = γe / (1 + Ts) (31)

In step 40, the yaw rate deviation Δγ, that is, the deviation between the filtered yaw rate γf and the reference yaw rate γt is calculated according to the following equation (32). Δγ = γf−γt (32)

In step 50, it is determined whether or not the yaw rate γf after the filter processing is a positive value, that is, whether or not the vehicle is turning to the left, and a negative determination is made. In some cases, the process proceeds to step 70, and when an affirmative determination is made, the process proceeds to step 60.

In step 60, the yaw rate deviation Δ
It is determined whether or not γ is a positive value, that is, whether or not the vehicle is in an oversteer state. If the determination is affirmative, the process proceeds to step 80; if the determination is negative, the process proceeds to step 90. Proceed to.

Similarly, in step 70, it is determined whether or not the yaw rate deviation Δγ is a negative value, that is, whether or not the vehicle is in an oversteer state. The routine proceeds to 80, and when a negative determination is made, the routine proceeds to step 90.

In step 80, the yaw rate deviation Δ
From the maps corresponding to the graphs shown in the first and fourth quadrants of FIG. 4 based on γ, the correction amount ΔCgf of the damping coefficient of the virtual shock absorber 122F on the front wheel side and the damping coefficient of the virtual shock absorber 124F on the front wheel side are calculated. The correction amount ΔCf, the correction amount ΔCgr of the damping coefficient of the virtual shock absorber 122R on the rear wheel side, the virtual shock absorber 12 on the rear wheel side
The correction amount ΔCr of the 4R damping coefficient, the correction amount ΔLf of the predetermined distance of the vehicle model on the front wheel side, and the correction amount ΔLr of the predetermined distance of the vehicle model on the rear wheel side are calculated.

Similarly, in step 90, based on the yaw rate deviation Δγ, the correction amounts ΔCgf, ΔCgf of the attenuation coefficient are obtained from the maps corresponding to the graphs shown in the second and third quadrants of FIG.
Cf, ΔCgr, ΔCr and correction amounts ΔLf, Δ for a predetermined distance
Lr is calculated.

In step 100, Cgfo, Cgro,
Cfo and Cro are respectively assumed to be a virtual shock absorber 122F on the front wheel side and a virtual shock absorber 122 on the rear wheel side.
R, a basic damping coefficient preset for the virtual shock absorber 124F on the front wheel side and the virtual shock absorber 124R on the rear wheel side, and Lfo and Lro as basic predetermined predetermined values for the vehicle model on the front wheel side and the rear wheel side, respectively. As the distances, the damping coefficients Cgf, Cgr, Cf and Cr of the respective virtual shock absorbers and predetermined distances Lf and Lr are calculated according to the following equations 33 to 38. Cgf = Cgfo + ΔCgf (33) Cgr = Cgro + ΔCgr (34) Cf = Cfo + ΔCf (35) Cr = Cro + ΔCr (36) Lf = Lfo + ΔLf (37) Lr = Lro + ΔLr (38)

In step 110, for example, the deviations ΔCgfa and ΔCgfa between the previous value and the current value of the damping coefficients Cgf and Cgr, respectively.
Cgra is calculated, and when the absolute values of the deviations ΔCgfa and ΔCgra exceed the reference value Cgo (positive constant), the current value is corrected so that the absolute value of the deviation becomes Cgo, so that the attenuation coefficient Cgf Processing for limiting the rate of change of Cgr is performed. Further, for example, deviations ΔCfa and ΔCra between the previous values and the present values of the damping coefficients Cf and Cr are calculated, and when the absolute values of the deviations ΔCfa and ΔCra exceed the reference value Co (positive constant), the absolute values of the deviations Is corrected to become Co, whereby a process of limiting the rate of change of the damping coefficients Cf and Cr is performed.

In step 120, for example, the deviations ΔLfa and ΔL between the previous value and the current value of the predetermined distances Lf and Lr, for example,
is calculated, and when the absolute values of the deviations ΔLfa and ΔLra exceed a reference value Lo (positive constant), the current value is corrected so that the absolute value of the deviation becomes Lo. The process of limiting the rate of change of Lr is performed, and the process then proceeds to step 420.

In step 420, it is determined whether or not the absolute value of the lateral acceleration Gy exceeds a reference value Gyo (positive constant) as a control threshold value, that is, whether or not the shock absorber at the time of turning the wheel. It is determined whether the control of the attenuation coefficient is necessary. If the determination is affirmative, the process proceeds to step 440; if the determination is negative, the process proceeds to step 430.

In step 430, the damping coefficient of the shock absorber of each wheel is set in accordance with a normal control routine when the vehicle is not turning, and thereafter, in step 5
Go to 80. The control of the damping coefficient in this case may be performed in any manner known in the art.

In step 440, the time differential value ΔGy of the lateral acceleration Gy is calculated, and the time differential value ΔG
It is determined whether or not the absolute value of y exceeds the reference value ΔGyo (positive constant), that is, whether or not the vehicle is in a transient turning state. If an affirmative determination is made, the process proceeds to step 460. Proceeds to step 45 if a negative determination is made
At 0, the damping coefficient Ci of the shock absorber of each wheel
Is set to the preset hard damping coefficient Ch, and the routine proceeds to step 580.

In step 460, the time differential value (stroke speed) Xid (i = f) of the stroke Xi of each wheel is obtained.
l, fr, rl, rr) are calculated, and in step 470, it is determined whether or not the lateral acceleration Gy is positive, that is, whether or not the vehicle is turning to the left. Is performed, the process proceeds to step 480. If a negative determination is made, the process proceeds to step 490.

In step 480, the stroke speed Xfind of the front wheel on the inside of the turn is changed to the stroke speed Xfld of the left front wheel.
The stroke speed Xfoutd of the turning outer front wheel is set to the stroke speed Xfrd of the right front wheel, and the stroke speed Xrind of the turning inner rear wheel is set to the stroke speed X of the left rear wheel.
rld and the stroke speed Xrout
d is set to the stroke speed Xrrd of the right rear wheel.

Similarly, in step 490, the stroke speed Xfind of the front inside wheel is set to the stroke speed Xfrd of the right front wheel, and the stroke speed Xfo of the front outside wheel is set to Xfod.
utd is set to the stroke speed Xfld of the left front wheel, the stroke speed Xrind of the turning inside rear wheel is set to the stroke speed Xrrd of the right rear wheel, and the stroke speed Xroutd of the turning outside rear wheel is set to the stroke speed Xrld of the left rear wheel. Is done.

In step 500, the sign differential Gy is used as the sign of the lateral acceleration Gy, and the time derivative ΔGy
It is determined whether or not the product with gnGy is positive, that is, whether or not the magnitude of the lateral acceleration due to the turning of the vehicle is in the process of increasing and the roll amount of the vehicle body is increasing. If an affirmative determination is made, in step 510, the damping coefficients Cj (j = fin, fout, j) of the shock absorbers of the inside front wheel, the outside front wheel, the inside rear wheel, and the outside rear wheel are turned.
(rin, rout) are calculated in accordance with the above equations 22 to 25, and when a negative determination is made, in step 520, the damping coefficient Cj of each shock absorber is calculated in accordance with the above equations 26 to 29.

In step 530, the damping coefficient Cfl of the left front wheel shock absorber is set to the damping coefficient Cfin of the turning inside front wheel, and the damping coefficient Cfr of the right front wheel shock absorber is set to the damping coefficient Cfout of the turning outside front wheel. ,
The damping coefficient Crl of the left rear wheel shock absorber is set to the damping coefficient Crin of the turning inner rear wheel, and the damping coefficient Crr of the right rear wheel shock absorber is set to the damping coefficient Cro of the turning outer rear wheel.
Set to ut.

Similarly, in step 540, it is determined whether or not the product of the time differential value ΔGy of the lateral acceleration and the sign Gy is positive.
At 50, the damping coefficient Cj of the shock absorber of the inside turning front wheel, the outside turning front wheel, the inside turning rear wheel, and the outside turning rear wheel is shown.
Is calculated in accordance with the above equations 22 to 25, and when a negative determination is made, in step 560, the damping coefficient Cj of each shock absorber is calculated in accordance with the above equations 26 to 29.

At step 570, the damping coefficient Cfl of the left front wheel shock absorber is set to the damping coefficient Cfout of the turning front outer wheel, and the damping coefficient Cfr of the right front wheel shock absorber is set to the damping coefficient Cfin of the turning inside front wheel. ,
The damping coefficient Crl of the left rear wheel shock absorber is set to the damping coefficient Crout of the turning outer rear wheel, and the damping coefficient Crr of the right rear wheel shock absorber is set to the damping coefficient Cri of the turning inner rear wheel.
Set to n.

In step 580, control is performed so that the damping coefficient of each shock absorber becomes the damping coefficient set in step 430, 450, 530 or 570. Thereafter, the flow returns to step 10.

Thus, according to the illustrated first embodiment,
At step 420, it is determined whether or not it is necessary to control the damping coefficient of the shock absorber during turning of the wheels. At step 440, it is determined whether or not the vehicle is in a transient turning state. In step 470, the turning direction of the vehicle is determined, and in steps 460, 480, and 490, the stroke speed of each wheel is obtained.
In steps 500 and 540, it is determined whether or not the roll amount of the vehicle body is in the process of increasing. When the roll amount of the vehicle body is in the process of increasing, steps 510 and 5
In step 50, the damping coefficient Cj of each shock absorber is calculated according to equations 22 to 25. When the roll amount of the vehicle body is in the process of decreasing, in steps 520 and 560, the damping coefficient Cj of each shock absorber is calculated in equations 26 to 29.
Is calculated according to

Therefore, according to the first embodiment, when the vehicle is in a transient turning state in which the roll amount of the vehicle body increases, the damping coefficient of the shock absorber inside the turning is higher than the damping coefficient outside the turning. When the damping coefficient of each shock absorber is controlled and, conversely, the vehicle is in a transient turning state in which the roll amount of the vehicle body is reduced, each shock absorber is set so that the damping coefficient of the shock absorber on the outside of the turn is higher than the damping coefficient on the inside of the turn. Is controlled, the vehicle height can be reduced, the center of gravity of the vehicle body can be reduced, and the kinetic performance of the vehicle during transient turning can be improved.

According to the illustrated first embodiment, the damping coefficients of the left and right front wheel shock absorbers and the right and left rear wheel shock absorbers are controlled independently of each other. Wr and Wr
Is appropriately set, and a map (FIG. 4) for calculating the correction amounts ΔCgf and ΔCgr, ΔCf and ΔCr, ΔLf and ΔLr is appropriately set, whereby the posture of the vehicle body in the front-rear direction during the transient turning of the vehicle is obtained. For example, the nose dive of the vehicle body at the beginning of a turn can be reduced, and the nose lift of the vehicle body at the end of a turn can be reduced.

Further, according to the first embodiment shown in the drawings, whether the roll amount of the vehicle body is in the process of increasing or decreasing is determined based on the lateral acceleration Gy of the vehicle body. The actual roll amount of the vehicle body is calculated based on the stroke Xi of each wheel detected by the above, and it is determined whether or not the vehicle body roll amount is in the process of increasing or decreasing based on the actual roll amount. Thus, the damping coefficient of each shock absorber can be controlled with good responsiveness.

Steps 420 to 580 are common to the first and second embodiments, so that the above-described respective effects can be obtained in a second embodiment to be described later.

In particular, according to the first embodiment shown in the figure, in step 30, the reference yaw rate γt of the vehicle is calculated, and in step 40, the deviation Δγ between the actual yaw rate γf and the reference yaw rate γt is calculated. 50-7
At 0, it is determined whether the vehicle is in the oversteer state or the understeer state. When the vehicle is in the oversteer state, in steps 80 and 100, the damping coefficients Cgf, Cf and Cff of the virtual shock absorber on the front wheel side are determined in steps 80 and 100. The predetermined distance Lf of the vehicle model on the front wheel side is increased and corrected in accordance with the deviation Δγ, and the damping coefficients Cgr and Cr of the virtual shock absorber on the rear wheel side and the predetermined distance Lr of the vehicle model on the rear wheel side are deviated. Reduction correction is performed according to Δγ.

Conversely, when the vehicle is in the understeer state, in steps 90 and 100, the damping coefficients Cgf and Cf of the virtual shock absorber on the front wheel side and the predetermined distance Lf of the vehicle model on the front wheel side decrease according to the deviation Δγ. While being corrected, the damping coefficients Cgr and Cr of the virtual shock absorber on the rear wheel side and the predetermined distance Lr of the vehicle model on the rear wheel side
Is increased in accordance with the deviation Δγ.

Therefore, according to the first embodiment, when the vehicle is in the oversteer state, the roll stiffness on the front wheel side is increased and the roll stiffness on the rear wheel side is reduced according to the degree of the oversteer state. In this state, the roll stiffness on the front wheel side is reduced and the roll stiffness on the rear wheel side is increased in accordance with the degree of the understeer state. Driving stability can be improved.

According to the illustrated first embodiment, the damping coefficients Cgf, Cf, Cgr, and Cr of the virtual shock absorber are set.
In addition, the predetermined distances Lf and Lr of the vehicle model also have a deviation Δγ.
, The actual damping coefficient of the actual shock absorber can be more accurately controlled as compared with the case where only the virtual damping coefficient of the virtual shock absorber is increased / decreased in accordance with the deviation Δγ.

Further, according to the first embodiment shown in the figure, the rate of change of the damping coefficients Cgf, Cf, Cgr, Cr is limited in step 110, and the change of the predetermined distances Lf, Lr is determined in step 120. Since the rate is limited, it is possible to reliably prevent a sudden change in the damping force of the shock absorber and a deterioration in ride comfort of the vehicle due to the sudden change in the damping force of the shock absorber as compared with a case where the processing for limiting the change rate is not performed. it can.

In the illustrated first embodiment, the yaw rate γ of the vehicle is detected by the yaw rate sensor 30, but the wheel speeds Vwfl and Vwfr of the left and right front wheels, which are the steered wheels, are detected. , Tr as the tread of the vehicle, and may be calculated according to the following equation 39 based on the wheel speed. γ = (Vwfr−Vwfl) / Tr (39)

In the illustrated first embodiment, the damping coefficients Cgf, Cf, Cgr,
The predetermined distances Lf and Lr of the vehicle model in addition to Cr are also increased or decreased according to the deviation Δγ, but the increase or decrease of the predetermined distances Lf and Lr may be omitted. In the illustrated embodiment, the attenuation coefficient Cgf and the amount of increase / decrease correction of Cgr are determined based on the same deviation Δγ.
And Cr is set to be larger than the increase / decrease correction amount.
The amount of correction for the increase or decrease of the damping coefficients Cf and Cr is equal to the damping coefficients Cgf and Cgf.
It may be set to be larger than the increase / decrease correction amount of gr, and furthermore, the increase / decrease correction of one of the attenuation coefficients Cgf and Cgr and the attenuation coefficients Cf and Cr may be omitted.

Further, in the illustrated first embodiment, the target yaw rate γe of the vehicle is calculated according to the above equation 30, but is shown in FIG. 5 based on the vehicle speed V and the steering angle δ. It may be calculated from a map corresponding to the graph.

Second Embodiment FIG. 9 is a flowchart showing the first half of a damping coefficient control routine in a damping coefficient control device according to a second preferred embodiment of the present invention.

Although not shown, a signal indicating the yaw rate γ of the vehicle, a signal indicating the vehicle speed V, and a signal indicating the steering angle δ are not input to the electric control device 24 of the second embodiment. , Throttle opening Th from the throttle opening sensor
And a signal indicating the depressing stroke Sb of the brake pedal from the brake stroke sensor are also input.

In step 210 of the damping coefficient control routine of the second embodiment, a signal indicating the stroke Xi of each wheel is read, and step 220 is executed.
For example, the throttle opening speed Vt is calculated as a time differential value of the throttle opening Th.

In step 230, based on the throttle opening speed Vt and the vehicle speed V, a virtual front wheel side of a map corresponding to the graph shown in FIG. 10 (front wheel drive vehicle) or FIG. 11 (rear wheel drive vehicle). The distribution ratio Kag to the damping coefficient Cgf of the shock absorber 122F is calculated.

Similarly, in step 240, the front wheel side of the map corresponding to the graph shown in FIG. 12 (front wheel drive vehicle) or FIG. 13 (rear wheel drive vehicle) based on the throttle opening speed Vt and the vehicle speed V. Virtual shock absorber 12
The distribution ratio Ka for the 4F damping coefficient Cf is calculated.

In step 250, the distribution ratio Kbg for the damping coefficient Cgf of the virtual shock absorber 122F on the front wheel side is calculated from the map corresponding to the graph shown in FIG. 14 based on the brake stroke Sb.

Similarly, in step 260, the virtual shock absorber 12 on the front wheel side from the map corresponding to the graph shown in FIG. 15 based on the brake stroke Sb.
The distribution ratio Kb for the 4F damping coefficient Cf is calculated.

When the vehicle is a vehicle such as a rally car or a sports car, in which it is preferable to quickly saturate the cornering power of the rear wheels to make it easy to turn, the distribution ratios Kbg and Kb
Are calculated from maps corresponding to the graphs shown in FIGS. 16 and 17, respectively.

In step 270, the virtual shock absorber 122F for the front wheel and the virtual shock absorber 122 for the rear wheel are calculated according to the following equations 40 to 43, respectively.
R, the damping coefficient Cg of the virtual shock absorber 124F on the front wheel side and the virtual shock absorber 124R on the rear wheel side
f, Cgr, Cf and Cr are calculated. Cgf = KagKbgCgfo (40) Cgr = (1-Kag) (1-Kbg) Cgro ... (41) Cf = KaKbCfo ... (42) Cr = (1-Ka) (1-Kb) Cro ... (( 43)

In step 280, the process of limiting the rate of change of the damping coefficients Cgf, Cgr, Cf, and Cr is performed in the same manner as in step 110 of the above-described first embodiment. Proceed to.

Thus, according to the illustrated second embodiment,
In step 220, the throttle opening speed Vt is calculated, and in step 230, the throttle opening speed Vt is calculated.
A distribution ratio Kag to the damping coefficient Cgf of the front-wheel virtual shock absorber 122F is calculated based on the vehicle speed V and the front wheel-side virtual shock absorber 1 based on the throttle opening speed Vt and the vehicle speed V in step 240.
The distribution ratio Ka for the damping coefficient Cf of the 24F is calculated, and in step 270, the damping coefficients Cgf, Cgr, Cf, and Cr of each virtual shock absorber 124R are calculated based on the distribution ratio of the front and rear wheels based on the distribution ratio Kag and Ka. Is done.

In general, in the case of a front wheel drive vehicle, the steer characteristic of the vehicle changes to the understeer side due to the driving force of the front wheel during the turning acceleration of the vehicle. Although the steering characteristic of the vehicle changes to the oversteer side due to the force, according to the second embodiment shown in the figure, when the vehicle is a front wheel drive vehicle, the virtual shock absorber of the front wheel at the time of turning acceleration of the vehicle. Since the ratio of the damping coefficient of the virtual shock absorber of the rear wheels to the damping coefficient is increased, it is possible to reduce the change in the steer characteristics toward the understeer side of the vehicle, and when the vehicle is a rear wheel drive vehicle. During the turning acceleration of the vehicle, the ratio of the damping coefficient of the front wheel virtual shock absorber to the damping coefficient of the front wheel virtual shock absorber is increased, so that the change in the steer characteristics toward the oversteer side of the vehicle. It can be reduced, thus it is possible to improve the steering stability of the vehicle.

According to the illustrated second embodiment, in step 250, the damping coefficient Cg of the virtual shock absorber 122F on the front wheel side based on the brake stroke Sb.
The distribution ratio Kbg for f is calculated, and in step 260, the distribution ratio Kb for the damping coefficient Cf of the virtual shock absorber 124F on the front wheel side is calculated based on the brake stroke Sb, and in step 270, the distribution ratios Kbg and Kb are calculated. The damping coefficients Cgf, Cgr, Cf, and Cr of the respective virtual shock absorbers 124R are calculated based on the distribution ratio of the front and rear wheels.

Therefore, according to the second embodiment, the change in the steer characteristic to the oversteer side due to the load movement to the front of the vehicle during the turning braking of the vehicle is reduced, thereby improving the steering stability of the vehicle. Can be done.

According to the second embodiment, the acceleration / deceleration of the vehicle is estimated based on the brake stroke Sb, which is the driver's braking operation amount, and the throttle opening speed, which is the driver's acceleration operation amount. In addition, the damping coefficient of the shock absorber of each wheel can be controlled with higher responsiveness than when the acceleration / deceleration of the vehicle is detected by, for example, a longitudinal acceleration sensor.

Also in the illustrated second embodiment, the rate of change of the damping coefficients Cgf, Cf, Cgr, and Cr is limited in step 280, so that the processing of limiting the rate of change is not performed. As a result, it is possible to reliably prevent a sudden change in the damping force of the shock absorber and a deterioration in ride comfort of the vehicle caused by the sudden change.

In the illustrated second embodiment, the acceleration of the vehicle as the pitching state quantity of the vehicle is estimated based on the throttle opening speed Vt.
The acceleration of the vehicle may be estimated based on, for example, the output torque of the torque converter of the automatic transmission. Further, the deceleration of the vehicle as the pitching state quantity of the vehicle is estimated based on the brake stroke Sb. However, the deceleration of the vehicle is, for example, the depressing force of a brake pedal not shown in the drawing or the brake master cylinder. May be estimated on the basis of the pressure.

Although the present invention has been described in detail with reference to specific embodiments, the present invention is not limited to the above-described embodiments, and various other embodiments may be included within the scope of the present invention. It will be clear to those skilled in the art that is possible.

For example, in each of the above-described embodiments, in the process of increasing the roll amount of the vehicle body, the damping coefficient of the inside shock absorber is controlled to be relatively higher than the damping coefficient of the outside shock absorber. Conversely, in the process of decreasing the body roll amount, the damping coefficient of the shock absorber on the outside of the turn is controlled to be higher than the damping coefficient of the shock absorber on the inside of the turn. At the end of turning, the behavior of the vehicle may become unstable.
Therefore, in the process of decreasing the roll amount of the vehicle body, the control for making the damping coefficient of the shock absorber on the outside of the turning higher than the damping coefficient of the shock absorber on the inside of the turning may be omitted.

Specifically, the calculation of the damping coefficient Cj in steps 520 and 560 is omitted, and instead, the damping coefficient Ci of each shock absorber is replaced with, for example, step 45.
As in the case of 0, it may be set to a hard damping coefficient Chigh, and then may be modified to proceed to step 580.

In each of the above embodiments, it is determined whether or not the roll amount of the vehicle body is in the process of increasing or decreasing based on the sign of the time differential value ΔGy of the lateral acceleration Gy of the vehicle body. This determination is made, for example, using Kh as the stability factor, Rg as the steering gear ratio, H as the wheel base, and the vehicle speed V and the steering angle sensor 34 detected by the vehicle speed sensor 32 shown in FIG.
The lateral acceleration Gys of the vehicle may be estimated based on the steering angle δ detected by the following equation (44), and may be performed based on the estimated lateral acceleration. Gys = V ² δ / [(1 + Kh V ² ) Rg H] (44)

Similarly, whether the body roll amount is in the process of increasing or decreasing is determined on the basis of the sign of the vehicle body roll rate calculated based on the stroke Xi detected by the vehicle height sensors 26FL-26RR. You may. In this case, the roll rate may be detected by a roll rate sensor not shown in FIG.

In each of the above-described embodiments, the stroke speed Xid of each wheel is calculated based on the detection results of the vehicle height sensors 26FL to 26RR. The vertical acceleration Gbi of the vehicle detected by the provided vertical acceleration sensors 28FL-28RR.
May be estimated by the observer based on the vehicle height, and the vehicle height sensor may be omitted.

Further, in each of the above-described embodiments, the rate of change of the damping coefficients Cgf, Cgr, Cf, Cr and the predetermined distances Lf, Lr of the virtual shock absorber is limited. The process of limiting the damping coefficient of the shock absorber and the rate of change of the predetermined distance may be omitted.

Further, in the above-described second embodiment, the front and rear wheel distribution ratio of the damping coefficient of the virtual shock absorber is changed based on both the throttle opening speed Vt and the brake stroke Sb. However, the front and rear wheel distribution ratio of the damping coefficient of the virtual shock absorber may be modified so as to be changed based on only one of the throttle opening speed Vt and the brake stroke Sb.

[0128]

As is apparent from the above description, according to the structure of the first aspect of the present invention, the lift on the revolving outer wheel side on the spring is suppressed by the second virtual shock absorber during the transient turning of the vehicle. Therefore, the center of gravity on the sprung can be lowered to improve the kinetic performance during turning of the vehicle, and the damping coefficient of the actual shock absorber is appropriately controlled in accordance with the change in the steering characteristics of the vehicle. Despite the change in the steering characteristic of the vehicle, the change in the steering characteristic during transient turning is reduced as compared with the case where the distribution ratio of the virtual damping coefficients of the first and second virtual shock absorbers between the front and rear wheels is constant. As a result, the steering stability of the vehicle can be improved.

Further, according to the configuration of claim 2, the distribution ratio of the virtual damping coefficient between the front and rear wheels is variably set according to the turning behavior of the vehicle, so that the first and second irrespective of the turning behavior of the vehicle. As compared with the case where the distribution ratio of the virtual damping coefficient of the second virtual shock absorber between the front and rear wheels is constant, it is possible to surely reduce the deterioration of the turning behavior of the vehicle due to the change in the steering characteristic during the transient turning. it can.

According to the third aspect of the present invention, the distribution ratio of the virtual damping coefficient between the front and rear wheels is variably set in accordance with the amount of state of the pitch movement on the spring. Changes can be reliably reduced.

According to the structure of the fourth aspect, the control is appropriately performed in accordance with the actual acceleration / deceleration of the shock absorber over the damping coefficient, whereby the change in the steering characteristic caused by the load movement in the longitudinal direction of the vehicle. Can be reliably reduced.

According to the fifth aspect of the present invention, the acceleration / deceleration on the spring is an estimated acceleration / deceleration estimated based on the amount of acceleration / deceleration operation by the driver. Compared to the case where the distribution ratio between the front and rear wheels is controlled in accordance with the detected acceleration / deceleration, the control can be appropriately performed in accordance with the acceleration / deceleration of the actual shock absorber over the damping coefficient without a response delay.

[Brief description of the drawings]

FIG. 1 is a schematic configuration diagram showing a first preferred embodiment of a damping coefficient control device according to the present invention.

FIG. 2 is a flowchart illustrating a first half of a damping coefficient control routine according to the first embodiment.

FIG. 3 is a flowchart showing the second half of a damping coefficient control routine according to the first embodiment.

FIG. 4 shows the correction amount ΔCg of the yaw rate deviation Δγ and the damping coefficient.
f, ΔCgr, ΔCf, ΔCr and the correction amount ΔL for a predetermined distance
It is a graph which shows the relationship with f and (DELTA) Lr.

FIG. 5 shows a steering angle δ, a vehicle speed V, and a target yaw rate γ of the vehicle.
It is a graph which shows the relationship with e.

FIG. 6 is an explanatory diagram showing a two-wheel model of an actual vehicle.

FIG. 7 is an explanatory diagram showing a virtual model in which a virtual shock absorber is provided inside the turning of the vehicle.

FIG. 8 is an explanatory diagram showing virtual models on the front wheel side and the rear wheel side in which virtual shock absorbers are provided inside the turning of the vehicle.

FIG. 9 is a flowchart illustrating the first half of a damping coefficient control routine according to the second embodiment.

FIG. 10 shows a throttle opening speed Vt for a front wheel drive vehicle.
9 is a graph showing a relationship between the distribution ratio Kag and the distribution coefficient Kag with respect to the damping coefficient Cgf.

FIG. 11 shows a throttle opening speed Vt for a rear wheel drive vehicle.
9 is a graph showing a relationship between the distribution ratio Kag and the distribution coefficient Kag with respect to the damping coefficient Cgf.

FIG. 12 shows a throttle opening speed Vt for a front wheel drive vehicle.
4 is a graph showing the relationship between the distribution ratio Ka and the damping coefficient Cf.

FIG. 13 shows a throttle opening speed Vt for a rear wheel drive vehicle.
4 is a graph showing the relationship between the distribution ratio Ka and the damping coefficient Cf.

FIG. 14 shows a brake stroke Sb for a normal vehicle
6 is a graph showing the relationship between the distribution coefficient Kbg and the damping coefficient Cgf.

FIG. 15 shows a brake stroke Sb for a normal vehicle
6 is a graph showing the relationship between the distribution ratio Kb and the damping coefficient Cf.

FIG. 16 shows a brake stroke Sb for a special vehicle.
6 is a graph showing the relationship between the distribution coefficient Kbg and the damping coefficient Cgf.

FIG. 17 shows a brake stroke Sb for a special vehicle.
6 is a graph showing the relationship between the distribution ratio Kb and the damping coefficient Cf.

[Explanation of symbols]

 DESCRIPTION OF SYMBOLS 14 ... Steering wheel 16 ... Power steering device 20 ... Body 24 ... Electric control device 26FL-26RR ... Vehicle height sensor 28 ... Lateral acceleration sensor 30 ... Yaw rate sensor 32 ... Vehicle speed sensor 34 ... Steering angle sensor 110 ... Body 112L, 112R ... Wheels 114L, 114R: Suspension springs 116L, 116R: Shock absorbers 122, 124: Shock absorbers

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[Procedure amendment]

[Submission date] June 24, 1999 (1999.6.2
4)

[Procedure amendment 1]

[Document name to be amended] Statement

[Correction target item name] 0016

[Correction method] Change

[Correction contents]

According to the fifth aspect of the present invention, the acceleration / deceleration on the spring is an estimated acceleration / deceleration estimated based on the amount of acceleration / deceleration operation by the driver. It can be appropriately controlled according to the acceleration of the upper ne attenuation coefficient Woba actual shock absorber without a delay in response as compared to the case where the distribution ratio between the front and rear wheels is controlled in accordance with the acceleration speed detected become.

[Procedure amendment 2]

[Document name to be amended] Statement

[Correction target item name]

[Correction method] Change

[Correction contents]

[0131] According to the fourth aspect, suitably controlled according to the actual acceleration or deceleration of the upper ne damping coefficient Woba of the shock absorber, thereby the steering characteristic due to load movement in the longitudinal direction of the vehicle Changes can be reliably reduced.

[Procedure amendment 3]

[Document name to be amended] Statement

[Correction target item name]

[Correction method] Change

[Correction contents]

According to the fifth aspect of the present invention, the acceleration / deceleration on the spring is an estimated acceleration / deceleration estimated based on the amount of acceleration / deceleration operation by the driver. it can be appropriately controlled in accordance with the acceleration of the upper ne attenuation coefficient Woba actual shock absorber without a delay in response as compared to the case where the distribution ratio between the front and rear wheels is controlled in accordance with the acceleration speed detected .

Continued on the front page (72) Inventor Toshio Onuma 1 Toyota Town, Toyota City, Aichi Prefecture Toyota Motor Co., Ltd. F-term (reference) 3D001 AA02 AA12 BA01 CA01 DA17 EA08 EA22 EA36 EA42 EB32 ED02

Claims (5)

[Claims] 1. A vehicle damping coefficient control device provided with an actual shock absorber having a variable damping coefficient corresponding to each wheel, means for detecting a turning state of the vehicle, and detecting a change in steering characteristics of the vehicle. Means for performing a virtual swing on the spring at a virtual position spaced a predetermined distance laterally from the sprung vehicle to a side estimated to be lifted with respect to the center of gravity on the spring based on the turning state of the vehicle. A first virtual shock absorber having a moving center and acting around the virtual swing center and a second virtual shock absorber acting in the vertical direction at the virtual position; A virtual damping coefficient setting means for variably setting a distribution ratio of a virtual damping coefficient of the first or second virtual shock absorber between the front and rear wheels according to a change in a steering characteristic of the vehicle model; Means for calculating a target damping coefficient of the actual shock absorber based on at least the virtual damping coefficient; and means for controlling the damping coefficient of the actual shock absorber based on the target damping coefficient. A vehicle damping coefficient control device. A means for detecting a change in the steering characteristic of the vehicle; detecting a turning behavior of the vehicle; and a virtual damping coefficient setting means for distributing the virtual damping coefficient between the front and rear wheels according to the turning behavior of the vehicle. The vehicle damping coefficient control device according to claim 1, wherein the ratio is variably set. 3. A means for detecting a change in the steering characteristic of the vehicle detects a pitch movement state quantity on the spring, and the virtual damping coefficient setting means sets the virtual damping coefficient in accordance with the pitch movement state quantity on the spring. 2. The damping coefficient control device for a vehicle according to claim 1, wherein a distribution ratio of the coefficient between the front and rear wheels is variably set. 4. The amount of pitch motion on the spring includes the acceleration / deceleration on the spring, and the virtual damping coefficient setting means determines a distribution ratio of the virtual damping coefficient between the front and rear wheels according to the acceleration / deceleration on the spring. The damping coefficient control device for a vehicle according to claim 3, wherein is variably set. 5. The vehicle damping coefficient control device according to claim 4, wherein the acceleration / deceleration on the sprung is an estimated acceleration / deceleration estimated based on an acceleration / deceleration operation amount by a driver.