JPH09249015A - Damping force control device for suspension system - Google Patents

Abstract

PROBLEM TO BE SOLVED: To perform proper damping on a spring by providing damping force characteristics to attach importance to either vertical vibration characteristics or rolling characteristis according to the operation state of a suspension system. SOLUTION: A damping constant target value inputted to a damping force variation type shock absorber 4 from a damping force target value computing means 3 is selectively switched to a damping constant target value, suitable to suppression of mainly vertical vibration on a spring and a damping constant target value suitable to suppression of mainly rolling on a spring according to a magnitude relation between a steering speed detected by a steering speed detecting means 5 or lateral acceleration detected by a lateral acceleration detecting means and a given reference value.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a suspension system in which a shock absorber capable of variably controlling a damping force and a spring are combined to accurately control the damping force in accordance with a vibration state which changes from moment to moment. The present invention relates to a vehicle capable of accurately suppressing rolling of a vehicle in accordance with driving conditions.

[0002]

2. Description of the Related Art In a suspension system using a shock absorber having a variable damping force, a damping coefficient is changed according to a vibration state to increase, for example, a damping rate of vibration from above a spring and isolation of vibration from below a spring. Proposals for improving the effect are made by the following references and the like.

Reference 1 ... D. Karnopp et
al. : Vibration Control Usi
ng Semi-Active Force Gene
rats, J. et al. E. FIG. I. ASME, May, 619
/ 626 (1974), Reference 2 ... Takehiko Fujioka, Koji Kido; Study on control method of variable damper (Vehicle vibration control viewed from VSS theory), Preprint 862 197/200 (1989) of the Society of Automotive Engineers of Japan , Reference 3 ... A
n Introduction to Sliding
mode Variable Structure
and Lyapunov Control, 1/2
0, Springer (1994).

Furthermore, the applicant of the present invention filed Japanese Patent Application No. 7-32
In 6169, by adjusting the damping force of the damping force variable shock absorber using signals that are relatively easy to detect, such as the relative speed and relative displacement between the sprung and unsprung, the damping force characteristics as desired can be obtained. There has been proposed a damping force control device that can be used.

[0005]

By the way, since the damping force control device of Japanese Patent Application No. 7-326169 uses a one-wheel model, it is only necessary to consider the vertical movement of the sprung and unsprung portions. In this case, a high damping effect can be obtained, but when this is applied to a four-wheel model and pitching and rolling on the spring must also be considered, this damping force control device can be used for pitching and rolling. Appropriate damping force characteristics are not always obtained by applying them individually.

That is, when this damping force control device is used, the damping force characteristic is determined by performing tuning with emphasis on any one of the vertical vibration, pitching and rolling vibration characteristics on the spring of the vehicle. However, there is a trade-off relationship among these vertical movement characteristics, pitching characteristics, and rolling characteristics, and as a result, it is difficult to obtain good vibration characteristics in total.

Specifically, if tuning is performed with emphasis on vertical vibration characteristics so as to be suitable for straight running, the rolling characteristics required when changing lanes are deteriorated. On the contrary, if the tuning that emphasizes the rolling characteristic is performed, the vertical vibration characteristic deteriorates.

The present invention pays attention to such a problem, and exerts damping force characteristics that emphasize either vertical vibration characteristics or rolling characteristics according to the operating condition of the suspension system, thereby suppressing vibration on the spring. It is an object of the present invention to provide a suspension system damping force control device capable of accurately performing.

[0009]

A first aspect of the present invention is directed to a sprung and unsprung relative velocity detecting means, a sprung and unsprung relative displacement detecting means, and a relative velocity and relative displacement detected from these. A damping force target value calculating means for setting a damping constant target value of the damping force variable shock absorber as an input, and a damping force variable shock bubbersaver for inputting the damping constant target value to adjust the damping constant are provided. In the suspension-type damping force control device, the damping force target value calculating means includes a damping constant target value mainly suitable for suppressing vertical vibration on a spring and a damping constant target value mainly suitable for suppressing rolling on a spring. Values can be output selectively.

According to a second aspect of the invention, the damping force target value calculating means is a target value setting means for setting a basic damping constant target value which is a predetermined constant, and an ideal damping calculated based on the relative speed and relative displacement. And a correction value calculation means for calculating a damping constant correction value in which the force is corrected, and outputs a damping constant target value including these basic damping constant target value and damping constant correction value, while the damping force target value calculation means is , A control gain calculating means for outputting a control gain for calculating the ideal damping force, and by this control gain, a damping constant target value suitable for suppressing vertical vibration on the spring and a rolling on the spring Selectively switches between the damping constant target value suitable for suppression.

A third aspect of the present invention comprises a steering speed detecting means for detecting a steering speed of the suspension system, and is suitable for suppressing a damping constant target value suitable for suppressing vertical vibration on the spring and suppressing rolling on the spring. The damping constant target value is switched according to the magnitude relationship between the detected steering speed and a predetermined reference value.

A fourth aspect of the present invention comprises a lateral acceleration detecting means for detecting a lateral acceleration of a suspension system, and is suitable for a damping constant target value suitable for suppressing vertical vibration on the spring and a suppressing value for suppressing rolling on the spring. The damping constant target value is switched according to the magnitude relationship between the detected lateral acceleration and a predetermined reference value.

[0013]

Therefore, in the present invention, the damping constant target value input to the damping force variable shock absorber has appropriate values when mainly suppressing vertical vibration on the spring and when mainly suppressing rolling on the spring. Since the setting is made, the damping force characteristics are controlled so that an appropriate damping force is generated according to the driving situation.

Further, by providing the steering speed detecting means, whether the vertical vibration on the spring should be mainly suppressed by comparing the steering speed detected from this with a predetermined reference value,
It is easily judged whether or not the rolling on the spring should be mainly suppressed. Based on this judgment, the damping constant target value suitable for suppressing the vertical vibration on the spring and the damping constant target value suitable for suppressing the rolling on the spring are set. Can be switched accurately.

Further, by providing the lateral acceleration detecting means, whether the vertical vibration mainly on the spring should be suppressed by comparing the lateral acceleration detected from this time with a predetermined reference value,
It is easily judged whether or not the rolling on the spring should be mainly suppressed. Based on this judgment, the damping constant target value suitable for suppressing the vertical vibration on the spring and the damping constant target value suitable for suppressing the rolling on the spring are set. Can be switched accurately.

[0016]

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

In FIG. 1, 1 is a relative displacement detecting means for detecting the relative displacement of the variable damping shock absorber 4 between the sprung and unsprung portions, and 2 is the relative velocity for similarly detecting the relative velocity of the sprung and unsprung portions. The detection means 5 is the steering speed λ of the vehicle.
The steering speed detecting means 3 for detecting the target value calculates the target value of the damping force based on the relative displacement, the relative speed and the steering speed, and inputs the result to the damping force variable shock absorber 4 to obtain the target damping force. It is a damping constant target value calculation means for controlling to generate.

FIG. 2 shows a concrete configuration of the damping constant target value calculating means 3, and as shown in the figure, the relative displacement x ₁ between the sprung part and the unsprung part (hereinafter simply referred to as relative displacement), Unsprung relative velocity x ₂ (hereinafter simply referred to as relative velocity),
Switching input calculation means 11 for calculating a switching input based on a state quantity x _{c of} a filter W _c (s) 13 described later and control gains M and N from a control gain calculation means 18 described later.
Similarly, relative velocity, relative displacement, filter state quantity x _c ,
The linear input calculation means 12 calculates a linear input based on the control gain L from the control gain calculation means 18, an adder 16 for adding these switching inputs and the linear input, and an output of this adder 16. A filter W _c (s) 13 that outputs an ideal damping force based on an ideal input;
_The correction value calculating means 14 for calculating the correction value of the damping force constant from the ideal damping force output from the _c (s) 13, the target value setting means 15 for setting the basic damping constant target value, and these outputs are added. Adder 17 which outputs as a damping constant target value
And a control gain calculating means 18 for switching the control gains M and N of the switching input calculating means 11 and the control gain L of the linear input calculating means 12 from the steering speed λ of the vehicle.

This will be described in more detail below. In the following description, (d / dt) means the first derivative with respect to time, and (d ² / dt ² ) means the second derivative.

The switching input computing means 11 calculates the switching input u _sw from the relative velocity, the relative displacement, the state quantity x _{c of the} filter, the control gains N and M, the predetermined gain ρ, and the constant δ as follows.

U _sw = ρMx _e / (‖Nx _e ‖ + δ) (1) However, x _e = [x ^T x _C ^T ] ^T , x = [x ₁ x ₂ x ₃
x is a _{4 x 5 x 6 x 7 x} 8] T.

The linear input calculation means 12 calculates the linear input u _ln from the control gain L as follows.

U _ln = Lx _e (2) The filter Wc (s) 13 has an ideal input u _ideal = u _sw +
The ideal damping force f _ideal is calculated from u _ln as follows.

(D / dt) x _c = A _c x _c + B _c u _ideal (3) f _ideal = C _c (d / dt) x _c (4) where A _C , B _C , and C _C are It is a constant matrix of an appropriate dimension that represents the attenuation characteristic of the filter W _c (s) 13. For example, the filter W _c (s) is a low-pass filter having a cutoff frequency of 10 [Hz]; W _c (s) = 2π × 10 / (s + 2π × 10), and s is a Laplace operator,

[0025]

[Equation 1]

It may be set as follows.

The damping force target value calculation means 3 is f _ideal =
As [f ₁ f ₂ f ₃ f ₄ ] ^T , c ₁ ′ = −f ₁ / x ₂ + c _0f c ₂ ′ = −f ₂ / x ₄ + c _0f c ₃ ′ = −f ₃ / x ₆ + c _0r c ₄ ′ = − f ₄ / x ₈ + c _0r (5), damping force target values c ₁ ′, c ₂ ′, c ₃ ′, c _{4 of the} right front wheel, left front wheel, right rear wheel, and left rear wheel, respectively ′ Is output. Here, c _0f and c _0r are predetermined constants determined from the vibration characteristics in the high frequency band. In addition, in relation to the description of the claim, c
_0f, c _0r the damping constant target _{value, -f 1 / x 2, -f} 2 /
x ₄ , -f ₃ / x ₆ , and -f ₄ / x ₈ correspond to damping constant correction values, respectively.

The damping force variable shock absorber 4 has a c '=
It is assumed that [c ₁ ′ c ₂ ′ c ₃ ′ c ₄ ′] ^T is input and that the generated damping force c = [c ₁ c ₂ c ₃ c ₄ ] ^T is c = c ′ (6) To do.

The method of deriving each control gain will be described below.

Consider the suspension system shown in FIG. The equation of motion on the spring of this system can be written as

M ₂ (d ² / dt ² ) z ₂ = F ₁ + F ₂ + F ₃ + F ₄ i _θ ² (d ² / dt ² ) θ ₂ = −L _f (F ₁ + F ₂ ) + L _r (F
_{3 + F 4) i ψ2 (} d 2 / dt 2) ψ 2 = W f (-F 1 + F 2) + W r (-
F ₃ + F ₄ ) where F ₁ = −k _2f (z ₂₁ −z ₁₁ ) −c _2f [(d / dt) z ₂₁ − (d / dt) z ₁₁ ] F ₂ = −k _2f (z ₂₂ − _{z 12) -c 2f [(d} / dt) z 22 - (d / dt) z 12] F 3 = -k 2r (z 23 -z 13) -c 2r [(d / dt) z 23 - (d _{/ dt) z 13] F 4} = -k 2r (z 24 -z 14) -c 2r [(d / dt) z 24 - (d / dt) z 14] ... (7) where, m ₂ is the spring Upper (overall) mass, i _θ2 is the moment of inertia on the spring for pitching (rotation in the θ ₂ direction in the figure), and i _ψ2 is the moment of inertia on the spring for rolling (rotation in the ψ ₂ direction in the figure) , K _2f is the spring rigidity of the front wheel side, k _2r is the spring rigidity of the rear wheel side, c ₂₁ is the shock absorber damping constant of the right front wheel, c ₂₂ is the shock absorber damping constant of the left front wheel, and c ₂₃ is the shock of the right rear wheel. Absorber reduction Decay constant, c
₂₄ is the shock absorber damping constant of the left rear wheel, (d ² / d
t ² ) z ₂ is the vertical acceleration relative to the absolute coordinates on the spring,
(D ² / dt ² ) θ ₂ is the pitch angular acceleration on the spring, and (d ² /
dt ² ) ψ ₂ is the roll angular acceleration on the spring, z ₂₁ is the spring on the right front wheel, z ₂₂ is the spring on the left front wheel, z ₂₃ is the spring on the right rear wheel, z ₂₄ is the left rear wheel Sprung displacement, z ₁₁ is unsprung displacement of the right front wheel, z ₁₂ is unsprung displacement of the left front wheel, z ₁₃ is unsprung displacement of the right rear wheel, z ₁₄ is unsprung displacement of the left rear wheel, and L _f is The horizontal length from the sprung mass center to the front wheel, L _r is the horizontal length from the sprung mass center to the rear wheel, W _f is the distance between the left and right front wheels, and W _r is the distance between the left and right rear wheels. The purpose of the suspension system is disturbances z ₀₁ , z ₀₂ , z ₀₃ and z ₀₄ (unsprung displacements z ₁₁ and z respectively).
(Which affects ₁₂ , z ₁₃ , and z ₁₄ ).

The relative displacement between the sprung and unsprung parts of the right front wheel is x ₁
= Z ₂₁ -z _11, the right front wheel of the spring and the relative velocity of the unsprung x
₂ = (d / dt) z ₂₁ − (d / dt) z ₁₁ , the relative displacement of the left front wheel on and off the spring is x ₃ = z ₂₂ −z ₁₂ , the relative speed of the left front wheel on and off the spring. X ₄ = (d / dt) z ₂₂ −
(D / dt) z ₁₂ , the relative displacement between the sprung and unsprung portions of the right rear wheel is x ₅ = z ₂₃ −z ₁₃ , and the relative velocity between the sprung and unsprung portions of the right rear wheel is x ₆ = (d / dt _{) z 23 - (d / dt} ) z 13, x 7 = z 24 -z 14 a relative displacement on the unsprung spring of the left rear wheel, a sprung relative speed of the unsprung of the left rear wheel x ₈ = (D / dt)
z _24- (d / dt) z ₁₄

[0033]

[Equation 2]

Is obtained. However, α ₁ = 1 / m ₂ + L _f ² / i _θ ² + W _f ² / i _{ψ 2} , α ₂ = 1 / m ₂ + L _f ² / i _θ ² −W _f ² / i _ψ ² , α ₃ = 1 / m ₂ + L _f L _r / i _θ 2 + W _f W _r / i _{ψ 2} , α ₄ = 1 / m ₂ −L _f L _r / i _{θ 2} −W _f W _r / i _{ψ 2} , β ₁ = 1 / m ₂ −L _f L _r / i _θ2 + W _f W _r / i _ψ2 , β ₂ = 1 / m ₂ −L _f L _r / i _{θ 2} −W _f W _r / i _{ψ 2} , β ₃ = 1 / m ₂ + L _r ² / i _θ2 + W ^{_{r 2 / i ψ2, β 4}} = 1 / m 2 + L r 2 / it θ2 -W r 2 / i ψ2, is.

(D ² / dt ² ) z ₁ = [(d ² / dt ² ) z
₁₁ (d ² / dt ² ) z ₁₂ (d ² / dt ² ) z ₁₃ (d ²
/ Dt ² ) z ₁₄ ] ^T , the equation (8) is rewritten as follows.

[0036]

(Equation 3)

However,

[0038]

(Equation 4)

Is as follows. c _u1 , c _u2 , c _u3 , and c _u4 are damping coefficient variable margins of the damping force variable shock absorber 4 for the right front wheel, the left front wheel, the right rear wheel, and the left rear wheel, respectively. c _0f , c _0r
Is determined considering the vibration characteristics in the high frequency band, as will be described later, regardless of the actual variable amount of the shock absorber.

U is generated through the filter W _c (s) as follows: u (s) = W _c (s) u _ideal (10) Thus, for example, W if _c (s) is Oke as a low-pass filter is attenuated high frequency components of u (s), the high frequency _{band, u ≒ 0 ⇒c u1 ≒ 0} , c u2 ≒ 0, c u3 ≒ 0, _{c u4 ≒ 0 ⇒c 21 ≒ c} 0f, c 22 ≒ c 0f, c 23 ≒ c 0r, the _{c 24 ≒ c 0r ... (11} ). This effect will be described later.

When u is generated as in the equation (10), the suspension system and the filter expansion system can be expressed as follows.

[0042]

(Equation 5)

Further, the acceleration [(d^Two/ Dt^Two) Z_Two, (D^Two
/ Dt^Two) Θ_Two, (D^Two/ Dt^Two) Ψ_Two]^TIs [(d^Two/ Dt^Two) Z_Two, (D^Two/ Dt^Two) Θ_Two, (D^Two/ Dt^Two) Ψ_Two]^T= [C_P0 D _P0 C_c] X_e … (13) Where x_e= [X^T x_c ^T]^TAnd

[0044]

(Equation 6)

Is as follows.

Below, the control gain is derived based on the sliding mode control theory.

Equation (12) is given by T ₁ B _c = [0 B ₂ ^T ] ^T , B ₂
Is a regular matrix, and a similarity transformation using T _{1 such} that

[0048]

(Equation 7)

By

[0050]

(Equation 8)

Can be rewritten as

When u _ideal is generated as in equation (9),
Within a sufficiently short time for practical use after starting the control, the Sli
The ding Mode can be generated on the switching surface; σ = 0, σ = Fy ₁ + y ₂ (15). That is, the state quantity of the expansion system is constrained to the plane of σ = 0 [see Reference 3]. At this time, the motion of the expanding system is (d / dt) y ₁ = A ₁₁ y ₁ + A ₁₂ y ₂ + B ₁₁ (d ² / dt ² ) z ₁ y ₂ = −Fy ₁ (16) That is, (d / dt) y ₁ = (A ₁₁ −A ₁₂ F) y ₁ + B ₁₁ (d ² / dt ² ) z ₁ (17)

To meet the control objective, the state may be constrained to the desired switching surface. In other words, the expression (17) may be expressed as a preferable vibration characteristic. This is equivalent to the problem of damping the system represented by (dy ₁ / dt) = A ₁₁ y ₁ + A ₁₂ v with state feedback v = −Fy ₁ . In order to design such a feedback gain F (switching surface), the following known method may be used.

When the state is restricted to the switching surface σ = 0, the following evaluation function

[0055]

[Equation 9]

The switching surface that minimizes σ _∞ = 0, σ _∞ = F _∞
y ₁ + y ₂ can be designed from the equation (16) by H _∞ optimization. Here, Q is a positive definite symmetric weight matrix, y = [y ₁
y ₂ ] ^T and γ are constants determined by the characteristics of the suspension system and the characteristics of the filter. Alternatively, if (d ² / dt ² ) z ₁ is assumed to be an impulse function or white noise, Slidein
The following evaluation function when gMode is generated;

[0057]

(Equation 10)

The switching surface that minimizes σ ₂ = 0, σ ₂ = F ₂ y ₁
+ Y ₂ can be designed by H ₂ optimization [see Reference 3].

Here, for example, the sprung vertical acceleration (d^Two/
dt^Two) Z_Two, Sprung pitch angular acceleration (d^Two/ Dt^Two) Θ_Two
 Sprung roll angular acceleration (d^Two/ Dt^Two) Ψ_TwoWant to suppress
Sometimes, [(d^Two/ Dt^Two) Z_Two (D^Two/ Dt^Two) Θ_Two 
(D^Two/ Dt^Two) Ψ_Two]^TShould be put in the evaluation function, [(d^Two/ Dt^Two) Z_Two (D^Two/ Dt^Two) Θ_Two (D^Two/ Dt^Two) Ψ_Two]^T= [C_P0 D _P0 C_c] X_e, T_Twox_e= Y (18), Q = Q₀+ (Q_W[C_P0 D_P0C_c] T_Two ^-1)^T(Q_W[C_P0 D_P0C_c] T_Two ^-1) (18a) However, x_e= [X^T x_c ^T]^T, Q₀Is Q
Is a matrix for satisfying the positive definite symmetry of. here,

[0060]

[Equation 11]

For vertical movement, pitching, and rolling, vertical weight q _z , pitching weight q _p , and
The rolling weight q _r allows individual weighting.

The ideal input u _ideal for realizing the sliding mode may be generated as u _ideal = Lx _e + ρMx _e / (‖Nx _e ‖ + δ) (19). Here, the control gains L, M, and N are the appropriate matrices shown in the above-mentioned reference document 3, and are obtained as follows. First, Σ, Θ, and Φ are defined as follows.

Σ = A ₁₁ -A ₁₂ F Θ = F Σ-A ₂₂ F + A ₂₁ Φ = FA ₁₂ + A _{22 From} these, L, M and N are L = −B ₂ ⁻¹ (Θ Φ−Φ ′) T ^{_{3 T 2 M = -B 2 -1}} (0 P 2) T 3 T 2 N = (0 P 2) becomes T ₃ T _2. Here, P ₂ is a positive definite symmetric solution of the following Lyapunov equation.

[0064] In the ^{_{P 2 Ф '+ Ф' T}} P 2 = -I Ф ' and ρ is the design parameters, respectively Sliding
A stability matrix that determines the speed of convergence to the mode, and a positive constant that determines the size of u _sw . Also,

[0065]

(Equation 12)

Is as follows.

Δ ≧ 0 is a design parameter, which is an established number for preventing chattering. The response of the damping constant is fast enough,
In addition, when the relative displacement and the relative velocity can be detected sufficiently fast, if δ = 0, u _sw becomes a discontinuous switching input, and the state can be reliably restrained to the switching surface.

From u _ideal , the ideal damping force is generated as previously shown: f _ideal (s) = W _c (s) u _ideal (s). With ideal damping force, f _sabs that can be generated by the shock absorber is -f _ideal / x ₂ = c ″ ≧ 0, f _sabs = −c ″ x ₂ (20) and when −f _ideal / x ₂ ≧ 0. Limited Here, it is not always guaranteed that −f _ideal / x ₂ ≧ 0, but the filter W
_{If c} (s) and the switching surface σ = 0 are properly set, a sufficient control effect can be expected in the practical area. From this fact, the target value of the damping constant of the damping force variable shock absorber is as shown in equation (5) in consideration of the fixed component c ₀ of the damping force, as shown in equation (5): c ′ = c _u + c ₀ , c ′ = [c ₁ ′ C ₂ ′ c ₃ ′ c ₄ ′] ^T _cu = [− f _ideal / x ₂ −f _ideal / x ₄ −f _ideal /
x ₆ −f _ideal / x ₈ ] ^T c ₀ = [c _0f c _0f c _0r c _0r ] ^T.

Next, the effects of c _0f and c _0r will be described. As can be seen from the equation of motion (7) of the suspension system, when damping the vibration on the spring, z ₁₁ , z ₁₂ , z ₁₃ , z ₁₄ ,
(D / dt) z ₁₁ , (d / dt) z ₁₂ , (d / dt) z
₁₃ , (d / dt) z ₁₄ is a disturbance, and k _2f , k _2r ,
c ₂₁ , c ₂₂ , c ₂₃ , and c ₂₄ have respective input gains. Therefore, if the values of c ₂₁ , c ₂₂ , c ₂₃ , and c ₂₄ are reduced, the disturbances (d / dt) z ₁₁ , (d / dt) z ₁₂ ,
Although the effects of (d / dt) z ₁₃ and (d / dt) z ₁₄ are small, damping is insufficient and vibration near the sprung resonance frequency becomes large.

However, in the control method described above, if, for example, the filter W _c (s) is a low-pass filter and c _0f and c _0r are sufficiently small values, equation (1)
As described in 1), at high frequencies, c _{21 ≈c 0f} , c _{22 ≈c
0f} , c _{23 ≈c 0r} , c _{24 ≈c 0r} , the disturbance blocking property is improved, and the sliding mode is used in the low frequency band.
The effect of the control can be realized and the damping of the suspension system can be improved.

In the present invention, as described above, the weighting matrix Q for determining the control gains L, M and N is given by the equation (18).
Since it is given as a), by the magnitude relation between q _z and q _r ,
It is possible to select which of the sprung vertical acceleration (d ² / dt ² ) z ₂ and the sprung roll angular acceleration (d ² / dt ² ) ψ ₂ is more suppressed.

That is, in Equation (18), the sprung vertical acceleration (d ² / dt ² ) z ₂ and the sprung roll angular acceleration (d ² / dt) when the vertical weight q _z is set to be larger than the rolling weight q _r ² ) ψ ₂ is, for example, as shown in FIGS.
The characteristic is as shown by the line (A). On the other hand, the sprung vertical acceleration (d ² / dt ² ) z ₂ and the sprung roll angular acceleration (d ² / dt ² ) ψ ₂ when the rolling weight q _r is set larger than the vertical weight q _z are, for example, The characteristics shown by the line (B) in FIGS. 4 and 5 are taken.

As can be seen by comparing these lines (A) and (B), if q _z > q _r , the sprung vertical acceleration is suppressed more than the sprung roll angular acceleration. On the other hand, q
_{If r} > q _z , the sprung roll angular acceleration is suppressed more than the sprung vertical acceleration is higher than the sprung roll angular acceleration.

Therefore, when the steering speed λ is smaller than a predetermined set value (reference value), the vertical weight q _z and the rolling weight q _r are set to q _z > q _r, and the control gain L that is switched according to these weights is set. , M, N are used to mainly suppress the sprung vertical acceleration. As a result, in a driving situation where the steering speed λ is small and straight rolling is not required, and sprung vertical acceleration is suppressed, priority is given to suppression of sprung vertical acceleration.

On the other hand, when the steering speed λ is larger than the set value, q _z <q _r is set, and by using the control gains L, M and N which are switched according to these weights, Try to suppress mainly.
As a result, when the steering speed λ is increased at the time of changing lanes, when turning is started, and when turning is ended, and especially when rolling is required to be suppressed, the sprung roll angular acceleration is suppressed more than the sprung vertical acceleration is suppressed. It will be given priority.

6 and 7 show another embodiment of the present invention.

As shown in FIG. 6, in this embodiment,
A lateral acceleration detecting means 6 for detecting a lateral acceleration G is provided in place of the steering speed detecting means of FIG. 1, and the damping constant target value determining means 3 includes a relative displacement between the sprung portion and the unsprung portion from the relative displacement detecting means 1. The sprung and unsprung relative velocities from the relative velocity detecting means 2 and the lateral acceleration G of the vehicle from the lateral acceleration detecting means 6.
The target value of the damping force is calculated based on the above, and the result is input to the variable damping force shock absorber 4 so that the target damping force is generated.

That is, as shown in the concrete configuration of the damping constant target value calculating means 3 in FIG. 7, the lateral acceleration G of the vehicle is input to the control gain calculating means 18, and the lateral acceleration G is used to switch input calculating means. Control gain M of 11,
N and the control gain L of the linear input calculation means 12 can be switched to appropriate ones.

Specifically, when the lateral acceleration G is smaller than a set value, for example, when traveling straight ahead, the sprung vertical acceleration is changed by switching to the control gains L, M and N when q _z > q _r. Mainly suppress. On the other hand, when the lateral acceleration G is larger than the set value, the sprung roll angular acceleration is mainly suppressed by using the control gains L, M, and N when q _z <q _r to start turning when changing lanes. Also, when the lateral acceleration G becomes large, such as at the end of rolling, and it is particularly necessary to suppress rolling, it is possible to cope with it.

As described above, according to the present invention, an appropriate damping force is generated in accordance with the change in the driving situation judged by the magnitude of the steering speed λ or the lateral acceleration G, and the rolling of the vehicle is appropriately suppressed. You can

[0081]

As described above, according to the present invention, the damping constant target value input to the variable damping force shock absorber is
Since appropriate values are set mainly when vertical vibration on the spring should be suppressed and when rolling on the spring should be mainly suppressed, damping force characteristics are set so that an appropriate damping force is generated according to the driving situation. Is controlled.

Whether or not the vertical vibration on the spring should be mainly suppressed by comparing the steering speed detected from now on with a predetermined reference value by providing the steering speed detecting means,
It is easily judged whether or not the rolling on the spring should be mainly suppressed. Based on this judgment, the damping constant target value suitable for suppressing the vertical vibration on the spring and the damping constant target value suitable for suppressing the rolling on the spring are set. Can be switched accurately.

Whether the vertical acceleration on the spring should be suppressed mainly by comparing the lateral acceleration detected from the lateral acceleration with a predetermined reference value by providing the lateral acceleration detecting means.
It is easily judged whether or not the rolling on the spring should be mainly suppressed. Based on this judgment, the damping constant target value suitable for suppressing the vertical vibration on the spring and the damping constant target value suitable for suppressing the rolling on the spring are set. Can be switched accurately.

[Brief description of drawings]

FIG. 1 is a block diagram showing an embodiment of the present invention.

FIG. 2 is a block diagram showing a damping force calculation means.

FIG. 3 is a model diagram showing a suspension system of the same.

FIG. 4 is a characteristic diagram similarly showing a vibration damping effect.

FIG. 5 is a characteristic diagram similarly.

FIG. 6 is a block diagram showing another embodiment of the present invention.

FIG. 7 is a block diagram showing a damping force calculation means according to another embodiment of the present invention.

[Explanation of symbols]

 1 Relative Displacement Detecting Means 2 Relative Velocity Detecting Means 3 Damping Constant Target Value Calculating Means 4 Shock Absorber 5 Steering Speed Detecting Means 6 Lateral Acceleration Detecting Means 11 Switching Input Calculating Means 12 Linear Input Calculating Means 13 Filters 14 Correction Value Calculating Means 15 Target Values Setting means 18 Control gain calculation means

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[Procedure amendment]

[Submission date] March 26, 1996

[Procedure amendment 1]

[Document name to be amended] Statement

[Name of item to be corrected] 0024

[Correction method] Change

[Correction contents]

[0024] Here, A _c , B _c , and C _c are constant matrices of appropriate dimensions that represent the attenuation characteristics of the filter W _c (s) 13. For example, the filter W _c (s) is a low-pass filter having a cutoff frequency of 10 [Hz]; W _c (s) = 2π × 10 / (s + 2π × 10), and s is a Laplace operator,

[Procedure amendment 2]

[Document name to be amended] Statement

[Correction target item name] 0031

[Correction method] Change

[Correction contents]

[0031] Here, m ₂ is the mass on the spring (entire), i _θ2 is the moment of inertia on the spring with respect to pitching (rotation in the _θ2 direction in the figure), and i _φ2 is rolling (rotation in the ψ ₂ direction in the figure) Moment of inertia on the spring, k _2f is spring rigidity on the front wheel side, k _2r is spring rigidity on the rear wheel side, c ₂₁ is the shock absorber damping constant of the right front wheel, c ₂₂ is the shock absorber damping constant of the left front wheel, c ₂₃ Is the shock absorber damping constant of the right rear wheel, c ₂₄ is the shock absorber damping constant of the left rear wheel, (d ² / dt ² ) z ₂ is the vertical acceleration relative to the absolute coordinates on the spring, and (d ² / dt ² ) θ. ₂ is the pitch angular acceleration on the spring, (d ² / dt ² ) ψ ₂ is the roll angular acceleration on the spring, z ₂₁ is the sprung displacement of the right front wheel, z ₂₂ is the sprung displacement of the left front wheel, and z ₂₃ is the right Rear sprung displacement of the rear wheel, z ₂₄
Is the sprung displacement of the left rear wheel, z ₁₁ is the unsprung displacement of the right front wheel,
z ₁₂ is the unsprung displacement of the left front wheel, z ₁₃ is the unsprung displacement of the right rear wheel, z ₁₄ is the unsprung displacement of the left rear wheel, L _f is the horizontal length from the sprung mass center to the front wheel, and L _r is The horizontal length from the sprung center of gravity to the rear wheel, W _f is 1 / the distance between the left and right front wheels
2, W _r is 1/2 of the distance between the left and right rear wheels. The purpose of the suspension system is disturbances z ₀₁ , z ₀₂ , z ₀₃ , z ₀₄ (influencing unsprung displacements z ₁₁ , z ₁₂ , z ₁₃ , z ₁₄ respectively)
In the presence of, it is to suppress the movement on the spring.

[Procedure 3]

[Document name to be amended] Statement

[Correction item name] 0053

[Correction method] Change

[Correction contents]

To meet the control objective, the state may be constrained to the desired switching surface. In other words, the expression (17) may be expressed as a preferable vibration characteristic. This is equivalent to the problem of damping the system represented by (d / dt) y ₁ = A ₁₁ y ₁ + A ₁₂ v with state feedback v = −Fy ₁ . In order to design such a feedback gain F (switching surface), the following known method may be used.

[Procedure amendment 4]

[Document name to be amended] Statement

[Correction target item name] 0067

[Correction method] Change

[Correction contents]

Δ ≧ 0 is a design parameter, which is a positive constant for preventing chattering. The response of the damping constant is fast enough,
In addition, when the relative displacement and the relative velocity can be detected sufficiently fast, if δ = 0, u _sw becomes a discontinuous switching input,
It is possible to surely restrain the state of the switching surface.

[Procedure Amendment 5]

[Document name to be amended] Statement

[Correction target item name] 0073

[Correction method] Change

[Correction contents]

As can be seen by comparing these lines (A) and (B), if q ₂ > q _r , the sprung vertical acceleration is suppressed more than the sprung roll angular acceleration. on the other hand,
If q _r > q ₂ , the sprung roll angular acceleration is suppressed more than the sprung vertical acceleration.

Claims (4)

[Claims] 1. A sprung and unsprung relative velocity detecting means, a sprung and unsprung relative displacement detecting means, and a damping constant of a damping force variable shock absorber using the relative velocity and relative displacement detected from these as input. A damping force control device for a suspension system, comprising: damping force target value calculation means for setting a target value; and a damping force variable shock babsobber for inputting the damping constant target value to adjust the damping constant. The damping force target value calculation means can selectively output a damping constant target value mainly suitable for suppressing vertical vibration on the spring and a damping constant target value mainly suitable for suppressing rolling on the spring. A damping force control device for a suspension system characterized by the above. 2. The damping force target value calculating means corrects the ideal damping force calculated based on the relative speed and the relative displacement, and a target value setting means for setting a basic damping constant target value which is a predetermined constant. The damping force target value calculating means includes a correction value calculating means for calculating a damping constant correction value, and outputs a damping constant target value including the basic damping constant target value and the damping constant correction value. A control gain calculation means for outputting a control gain for calculating a force is provided, and by this control gain, a damping constant target value suitable for suppressing vertical vibration on the spring and a control for suppressing rolling on the spring are suitable. The damping force control device for a suspension system according to claim 1, wherein the damping constant target value is selectively switched. 3. A steering speed detecting means for detecting a steering speed of a suspension system, a damping constant target value suitable for suppressing vertical vibration on the spring, and a damping constant target suitable for suppressing rolling on the spring. 3. The suspension system damping force control device according to claim 1, wherein the value is switched according to the magnitude relationship between the detected steering speed and a predetermined reference value. 4. A lateral acceleration detecting means for detecting a lateral acceleration of a suspension system, comprising a damping constant target value suitable for suppressing vertical vibration on the spring and a damping constant target suitable for suppressing rolling on the spring. The damping force control device for a suspension system according to claim 1 or 2, wherein the value is switched according to the magnitude relation between the detected lateral acceleration and a predetermined reference value.