JP2573193B2 - Vehicle suspension device - Google Patents

Description

The present invention relates to a suspension device for a vehicle.

(Prior art and its problems) Some of the suspension devices for vehicles include Europe (EPC)
As seen in the specification identified in Application Publication No. 0 114 757, a liquid cylinder is erected between the vehicle body and each wheel,
A so-called active suspension is known in which the characteristics of the suspension are variably controlled by supplying and discharging a working liquid to and from the liquid cylinder.

By the way, one of the characteristics required for a vehicle suspension device is roll rigidity, and factors determining the roll rigidity include a suspension geometry and a spring characteristic.

For example, increasing the spring constant of the suspension and decreasing the roll angle is effective in improving the controllability and stability of the vehicle. However, on the other hand, there is a problem that ride comfort is impaired.

For this reason, the spring constant of each of the front and rear wheels is preferably given in terms of ride comfort, and a stabilizer is provided to set the relative magnitude of the suspension roll rigidity of the front and rear wheels. That is, the stabilizer has a function of increasing the spring constant of the suspension only during rolling as an auxiliary spring for the suspension device.

Incidentally, the relative magnitude of the front-wheel-side suspension roll stiffness and the rear-wheel-side suspension roll stiffness, that is, the roll stiffness ratio between the front wheel and the rear wheel, has a great effect on the traveling performance of the vehicle. Taking the steering characteristic as an example, when the suspension roll stiffness on the front wheel side is increased, the tendency of understeer increases and the roll angle tends to decrease. On the other hand, when the suspension roll rigidity on the rear wheel side is increased, the understeer tends to be weakened. For this reason, in the current suspension device, the roll rigidity ratio suitable for the vehicle type is set at present.

On the other hand, the steering characteristics of a vehicle tend to change in response to a load change in the vehicle front-rear direction. For example, when the load on the rear wheel side increases, the steering characteristics change in the direction of oversteer, and the turning performance of the vehicle tends to be sensitive.

By the way, recently, various active suspension devices capable of arbitrarily changing suspension characteristics have been proposed, and have already been put to practical use in some vehicles. In this active suspension device, roll control for suppressing roll of the vehicle body, pitching control for suppressing pitching of the vehicle body, and the like are performed. JP 58
Japanese Patent Application Laid-Open No. -206409 proposes a technique in which the roll stiffness ratio between the front wheel side and the rear wheel side is changed according to the load to optimize the turning characteristics. Japanese Patent Laying-Open No. 61-193908 proposes an apparatus that optimizes the amount of load movement between the front wheels and the rear wheels, that is, the roll rigidity, in accordance with the lateral acceleration acting on the vehicle body. Further, Japanese Unexamined Patent Publication No. 61-181713 discloses that the turning characteristic at the time of turning is changed by changing the distribution amount of the hydraulic fluid to the front wheel-side vehicle height adjusting actuator and the rear wheel-side vehicle height adjusting actuator. There has been proposed a device in which any one of a neutral characteristic, an over-steering characteristic, and an under-steering characteristic can be selected.

In the active suspension device described above,
When performing roll control and pitching control, a load change occurs between the front wheel side and the rear wheel side due to pitching,
This change in load in the front-rear direction results in a change in suspension characteristics as described above.
This is undesirable.

The present invention has been made in view of the above circumstances, and an object of the present invention is to provide an active suspension apparatus that performs roll control for suppressing rolls and pitching control for suppressing pitching. It is an object of the present invention to provide a suspension device capable of preventing the deterioration of the suspension.

(Means for Solving the Problems) In order to achieve the above object, the present invention has the following configuration. That is, in a vehicle suspension apparatus in which suspension characteristics are controlled by supplying and discharging a working liquid to and from a liquid cylinder provided between a vehicle body and wheels through a control valve, Roll change detecting means for outputting a detected signal, pitching change detecting means for outputting a signal relating to the pitching change of the vehicle body, and a signal from the roll change detecting means for inputting a signal from the roll change detecting means so as to prevent a roll change of the vehicle body. A roll control unit for calculating a control amount of liquid to be supplied to and discharged from the cylinder; and a signal from the pitching change detecting unit being input, and calculating a control amount of liquid to be supplied to and discharged from the liquid cylinder so as to prevent a pitching change of the vehicle body. A pitching control unit, and a predetermined coefficient value is added to the control amount calculated by the roll control unit. A control amount to change the roll rigidity ratio of the front and rear wheels into a predetermined relationship, and a signal from the roll rigidity control unit and the pitching control unit, and control the two control units. A supply / discharge control means for controlling the supply / discharge of liquid to each liquid cylinder by adding an amount to the coefficient value of the roll stiffness control unit, as the control amount of the pitching control unit increases, the roll stiffness of the front wheel increases. And a correction unit that corrects the roll rigidity to be smaller than the roll rigidity of the rear wheel.

(Effects of the Invention) According to the present invention, while controlling the posture of the roll suppression control and the pitching suppression control to optimize the vehicle body posture, the roll of the front wheel and the rear wheel is controlled according to the control amount for suppressing the pitching. A suitable stiffness ratio can also provide suitable maneuverability.

(Example) Hereinafter, an example of the present invention will be described with reference to the drawings.

First Embodiment In FIG. 1, reference numeral 1 denotes a suspension device. Hereinafter, in the description of the elements included in the suspension device 1, when the elements are generically identified by a numeral, and when distinguished for each wheel, "FR" is used. (For front right wheel), "F
Symbols “L” (for left rear wheel), “RR” (for right rear wheel), and “RL” (for left rear wheel) are added for identification.

The suspension device 1 includes a vehicle body and wheels (not shown).
Each of the cylinders 2 is slidably fitted into the cylinder 2 and integrated with a piston rod 3 as is known. The cylinder liquid chamber 6 is defined by 4. Each cylinder fluid chamber 6 has gas springs 8FR, 8FL, 8RR, 8RL and oil passage 10F.
The orifices 12FR, 12FL, 12RR, and 12RL are provided in each oil passage 10 through R, 10FL, 10RR, and 10RL.
Each of the gas springs 8 has the same configuration, and has a gas chamber 16 and a liquid chamber 18 defined by a diaphragm 14 as a movable partition, and the liquid chamber 18 communicates with the oil passage 10. I have. The unit 20 including such a combination of the cylinder 2, the gas spring 8 and the orifice 12 has a basic function as a suspension by the damping action of the gas spring 8 and the damping action of the orifice 12. The characteristics of the suspension unit 20 are uniformly determined by the elastic modulus (spring coefficient) of the gas spring 8 and the throttle resistance of the orifice 12.

On the other hand, an external pipe 22 is connected to the cylinder 2,
The supply and discharge of the oil liquid to and from the cylinder 2, that is, the cylinder liquid chamber 6, is performed through a supply and discharge passage formed by the external pipe 22.

The hydraulic circuit for the cylinder 2 will be described. In FIG. 1, reference numeral 30 denotes a pump driven by an engine, and hydraulic fluid pumped up from a reservoir tank 32 by the pump 30 passes through a supply passage 33 for each wheel. It is supplied to the cylinder 2. That is, the supply passage
33 is a common passage 34 on the upstream side. The common passage 34 is branched into a front wheel passage 35 and a rear wheel passage 36,
35 is branched into a right front wheel passage 38FR and a left front wheel passage 38FL, and the rear wheel passage 36 is branched into a right rear wheel passage 40RR and a left rear wheel passage 40RL. , 38FL,
40RR and 40RL are the supply / discharge passages 22F leading to the cylinder 2 for each wheel.
R, 22FL, 22RR, and 22RL, respectively. The common passage 34 is provided with a switching valve 42, a check valve 44, and an accumulator 46 in this order from the upstream side. The accumulator 46 has the same configuration as the gas spring 8 and has a pressure accumulating function. Have been. On the other hand, each wheel passage 3
8, 40 and the supply / discharge passage 22, a flow control valve 4
8 is provided to adjust the amount of hydraulic fluid passing per unit time, that is, the flow rate of the hydraulic fluid.

On the other hand, the recirculation passage 50 extends from each flow control valve 48 to the reservoir tank 32 via the recirculation passage 52 for each wheel and the common recirculation passage 54, and the common recirculation passage 54 is switched from the switching valve 42. A valve return passage 56 is connected.

Next, the operation of the hydraulic circuit will be described. First, when the flow control valve 48 is closed, the suspension unit 20 exhibits characteristics based on the throttle resistance of the orifice 12 and the elastic modulus of the gas spring 8. That is, if the amount of change in load applied to the cylinder 2 is represented by ΔF and the amount of displacement of the piston 4 is represented by Δx, the dynamic spring constant K defined by ΔF / Δx is
The suspension unit 20 closed as a system is defined by the restricting resistance of the orifice 12 and the elastic modulus of the gas spring 8, so that a so-called passive (passiv)
e) to form a system.

On the other hand, when the flow control valve 48 is opened, for example, when the hydraulic fluid is supplied into the cylinder 2 when the piston rod 3 is displaced in the shortening direction, the supplied hydraulic fluid causes As a result of suppressing the shortening movement of the rod 3, the dynamic spring constant K changes in a direction to increase. In other words, by supplying and discharging the hydraulic fluid in the cylinder 2, the same effect as changing the throttling resistance of the orifice 12 and the elastic modulus of the gas spring 8 can be obtained. The unit 20 forms a so-called active system.

The flow control valve 48 is actuated by a control signal from a control unit 60 composed of a microcomputer, and the control unit 48 generates the control signal.
A signal from a pressure sensor 62 that picks up the pressure in each cylinder 2 is input to 60, and the pressure signal from the pressure sensor 62 is filtered by a high-pass filter 64 (a type of differential filter) in the control unit 60. After the processing, it is inputted to the control circuit 66. The control unit 60 has a common passage 34.
When the pressure signal from the pressure sensor 68 provided is input and the pressure of the hydraulic circuit becomes equal to or higher than a predetermined pressure, the switching valve 42 is switched, and the hydraulic fluid pumped up by the pump 30 is returned to the return passage 56, The liquid is returned to the reservoir tank 32 through 54. On the other hand, when the pressure of the hydraulic circuit becomes lower than the predetermined pressure, the switching valve 42 is switched to switch the pump
The hydraulic fluid pumped by the pump 30 is caused to flow through the supply passage 33, whereby the pressure in the hydraulic circuit is maintained at a predetermined pressure.

Next, an outline of the active control will be described. First, the basic model of control determines a target flow rate in accordance with a change in oil pressure in the cylinder 2 for each wheel, and supplies and discharges hydraulic fluid to and from each cylinder 2 based on the target flow rate. ing. In addition to this basic model, by combining the signals from the pressure sensors 62 for each wheel, four body input modes of bounce, pitch, roll, and warp are detected,
Control for suppressing the vehicle body input mode is performed. In addition, adding a warp moment to the vehicle body while the vehicle body is rolling imposes a roll stiffness anyway. Therefore, based on the roll moment detection value of the vehicle input mote, The target warp moment is determined, and a warp moment according to the roll is given, and the above-mentioned pitch moment is added to the calculation of the target warp moment so as to be in accordance with the load ratio between the front wheel and the rear wheel. By determining the target warp moment, the roll rigidity ratio is changed according to the load ratio. FIG. 2 is a block diagram showing such a control system. In this figure, the circuit for calculating the mode target flow rate is represented by transfer functions GB (S), GP (S), GR (S), and KW, where GB (S) is bound and GP (S) is pitch,
GR (S) is for roll and KW is for warp. RW is the transfer characteristic of the warp moment target value calculation circuit,
KP indicates the transfer characteristic of the filter with respect to the pitch moment.

The transfer function GB (S) and the like are obtained as follows. In the following description, in order to facilitate understanding, description will be made based on basic control of only one wheel, which is a basic unit of this control. Therefore, in the following description, the transfer functions GB (S), GP (S), etc. are collectively referred to as G (S), and the transfer function G (S) is derived based on the basic model from which the mode analysis is omitted. It shall be.

First, the transfer characteristic of each element in the control system is represented by the following relational expression.

ΔP = ΔF / A (1) Here, ΔF: Load change amount to cylinder 2 A: Pressure receiving area of piston 4 ΔP: Fluid pressure change amount in cylinder 2 ΔPN = ΔP−ΔPC (2) where ΔPC : Pressure change amount of liquid spring 8 ΔPN: Change amount of throttle pressure difference at orifice 12 QN = ΔPN / KN (3) where, KN: throttle resistance of orifice 12 QN: flow rate of oil liquid passing through orifice 12 ΔVC = QN / S (4) where, ΔVC: volume change amount of the fluid spring 8 ΔPC = KC · ΔVC (5) where KC: elastic modulus of the fluid spring 8 Δe = Ke · ΔF (6) Here, Ke: sensor characteristic of the pressure sensor 62 Δe: output of the pressure sensor 62 Δi = G (S) · Δe (7) where, Δi: flow control valve 48 output from the control circuit 66
Control current corresponding to the target flow rate of QV: flow rate of the oil liquid flowing through the flow control valve 48 ΔVL: QT / S (9) where ΔVL: change amount of the oil liquid in the cylinder 2 ΔV = ΔVC−ΔVL (10) where ΔV: cylinder 2 (cylinder fluid chamber 6) volume change amount Δx = ΔV / A (11) where, Δx: displacement amount of piston 4 Next, the target characteristic in the control system, that is, the frequency characteristic of dynamic spring constant When the target characteristic is set as shown in FIG. 3, the target characteristic is expressed by the following equation.

Here, S: Laplace operator T: Time constant By replacing the above equation (12), By the way, from the above equations (1) to (5), the volume change amount ΔVC of the fluid spring 8 is given by ΔVC = QN / S = ΔPN / (KN · S) = (ΔP−ΔPC) / (KN · S) = ( ΔP−KCΔVC) / (KN · S) = (ΔF / A−KC · ΔVC) / (KN · S) It is represented by

Further, the change amount ΔVL of the oil liquid in the cylinder 2 is calculated from the above equations (6) to (9). Is represented by

Further, the change amount Δx of the piston 4 is (10) to (1)
5) From the formula, Therefore, if this equation (16) is replaced, Becomes In comparison between the expression (17) and the expression (13) indicating the control target, in the expression (17), K ₁ = A ² · KC (18) K ₂ = A ² · KN (19) T = Putting N · TV… (20) and substituting these equations (18) to (20) into equation (13), Becomes

Therefore, from the above equations (17) and (21), Δx / Δ
When F is eliminated and the transfer function G (S) is obtained, And the characteristics shown in FIG. That is, (2)
By giving the transfer function G (S) shown in the equation 2) or FIG. 4, the dynamic spring characteristic ΔF / Δx shown in FIG. 3 is obtained. Such a transfer function G (S) is equivalent to a high-pass filter. In other words, the suspension device 1 of each wheel has its dynamic spring constant K variable according to the frequency, and the suspension device 1 responds to the frequency only by picking up the load acting on the suspension device 1. Also,
As shown in FIG. 3, the suspension device 1 is an active type suspension device in a low frequency range, so that a large dynamic spring constant K (hard) can be realized in a low frequency range. The change in the posture of the vehicle body such as the roll and the pitch, which is a problem in the region, can be suppressed to a small value. For comparison, the characteristic of only the passive control is shown by a broken line in FIG. In other words, in the high frequency range, the flow control valve 48 is closed to form a passive system, so that the dynamic spring constant of the passive system serving as a base is suppressed low (for example, the spring constant of the gas spring 8 is reduced). It is possible to improve riding comfort in a high frequency range under a soft suspension. In addition, since the flow control valve 48 does not require responsiveness in a high frequency range, there is an advantage that the flow control valve 48 can be simple. Further, when there is a failure in the hydraulic circuit, by closing the flow control valve 48,
Since the basic functions of the suspension are maintained in the active system, safety against failure is not impaired.

For the above basic model, detection of the body input mode
This is performed as follows.

(1) Rebound Bounce is a motion mode in the vertical direction of the vehicle body, and therefore, the motion directions of all four wheels are the same. From this,
Bound detection is based on the following equation.

ΔeB = ΔeFR + ΔeFL + ΔeRR + ΔeRL (23) Where, Δe: Bound detection value corresponding to the pressure change amount in the bound mode ΔeFR: Output of right front wheel pressure sensor 62FR ΔeFL: Output of left front wheel pressure sensor 62FL ΔeRR: Right rear wheel Output of the pressure sensor 62RR ΔeRL: Output of the left rear wheel pressure sensor 62RL (2) Pitch The pitch is a motion mode in which the motion direction of the front of the vehicle body and the motion direction of the rear of the vehicle are in opposite phases (upward or downward motion). From this, the detection of the pitch is based on the following equation.

Δep = (ΔeFR + ΔeFL) − (ΔeRR + ΔeRL) (22) where, Δep: pitch detection value corresponding to the pressure change amount in the pitch mode. Is a motion mode having an opposite phase (rotational motion about an axis extending in the front-rear direction of the vehicle body), and the roll detection is based on the following equation.

ΔeR = (ΔeFR−ΔeFL) + (ΔeRR−ΔeRL) (25) where ΔeR is a roll detection value corresponding to the amount of pressure change in the roll mode. (4) Warp The torsion moment acting on the vehicle body, and the right front wheel (FR ) And the left rear wheel (RL) have the same components, and other combinations (FL,
RR). From this, the detection of warp is based on the following equation.

ΔeW = (ΔeFR−ΔeFL) − (eRR−ΔeRL) (26) where, ΔeW is a warp detection value corresponding to the pressure change amount in the warp mode. The target flow rates ΔiB, ΔiP in each mode obtained in this manner. Are distributed in the same manner as in the above-described mode analysis, and the target flow rates ΔiF of the respective flow control valves 48FR, FL, RR, and RL are distributed.
R, ΔiFL, ΔiRR, and ΔiRL.

That is, the bound target flow rate ΔiB is distributed to each flow control valve 48 with the same sign, the pitch target flow rate ΔiP is distributed under the opposite sign for the front and rear wheels, and the roll target flow rate ΔiR is reversed for the right and left wheels. The warp target flow rate ΔiW is distributed under the opposite sign in a combination of the wheels located on the vehicle body diagonal line in units of each combination. If this is shown from the side of the target flow rates ΔiER, ΔiFL, ΔiRR, ΔiRL of each wheel, it is expressed by the following equation.

ΔiFR = ΔiB + ΔiP + ΔiR + ΔiW (27) ΔiFL = (ΔiB + ΔiP) − (ΔiR + ΔiW) (28) ΔiRR = (ΔiB−ΔiP) + (ΔiR−ΔiW)… (29) ΔiB = (ΔiB−ΔiP) − (Δi) (30) That is, control is added to the basic model in the direction of suppressing the moment in each mode in accordance with the vehicle body input mode. When the vehicle body is rolling, the target warp moment is determined, and as a result, a predetermined roll rigidity ratio is given. And, for example, the number of occupants, changes in the load ratio between the front wheels and the rear wheels according to the size of the load capacity of the load, or acceleration / deceleration, lift, road gradient, changes in the load ratio due to changes in the amount of fuel, the said As a result of being reflected in the target warp moment, the roll rigidity ratio between the front wheel and the rear wheel is changed according to the change in the load ratio.

That is, the transfer characteristic of the target warp moment calculation circuit
Assuming that the magnitude of RW is KRW, this KRW is represented by KRW = A-BX with respect to the pitch moment (X).

Therefore, when the pitch moment (X) is large, that is, the load ratio on the front wheel side is increased, and the tendency of understeer is increased, the KRW is reduced accordingly, and the target warp moment is changed to a small value. Accordingly, the rigidity of the suspension roll on the front wheel side is changed to a relatively small value, and the tendency of understeer is reduced. Conversely, if the load ratio on the rear wheel increases and the tendency of oversteer becomes stronger, the target warp moment is changed to a larger value, and the suspension roll stiffness on the rear wheel is changed to a relatively smaller value, and the tendency of oversteer is increased. Will be weakened.

Here, in the embodiment, the roll rigidity ratio between the front wheel and the rear wheel is changed by using the warp control indicated by KW in FIG. 2, and the coefficient value for determining the roll rigidity ratio is: Second
It becomes RW of the transfer characteristic in the figure (corresponding to the roll rigidity control unit).
Then, the transfer characteristic RW as this coefficient value is corrected by the KP in FIG. 2 corresponding to the control amount of the pitching suppression (corresponding to the correction unit).

[Brief description of the drawings]

 1 is an overall system diagram in the embodiment, FIG. 2 is a block diagram in the embodiment, FIG. 3 is a target dynamic spring characteristic diagram in the embodiment, and FIG. 4 is a characteristic diagram of a transfer function. 1: Suspension device 2: Cylinder 8: Gas spring 30: Pump 46: Accumulator 48: Flow control valve 60: Control unit 62: Load sensor 64: High pass filter (differential filter) 66: Control circuit 68: Pressure sensor

Continuation of the front page (72) Inventor Shoichi Uemura 3-1, Shinchi, Fuchu-cho, Aki-gun, Hiroshima Prefecture Mazda Co., Ltd. (56) References JP-A-60-64014 (JP, A) JP-A-60-199181 ( JP, A) JP-A-58-206409 (JP, A) JP-A-61-193908 (JP, A) JP-A-61-181713 (JP, A) JP-A-60-163104 (JP, U)

Claims (1)

(57) [Claims] A suspension apparatus for a vehicle in which a suspension characteristic is controlled by supplying and discharging a working liquid to and from a liquid cylinder provided between a vehicle body and wheels via a control valve. Roll change detecting means for outputting a signal related to the change; pitching change detecting means for outputting a signal related to the pitching change of the vehicle body; and inputting a signal from the roll change detecting means to prevent a roll change of the vehicle body. A roll control unit for calculating a control amount of the liquid to be supplied to and discharged from the liquid cylinder; and a signal from the pitching change detecting means being input to control the liquid to be supplied to and discharged from the liquid cylinder so as to prevent a pitching change of the vehicle body. A pitching control unit for calculating an amount, and adding a predetermined coefficient value to the control amount calculated by the roll control unit. By changing the control amount to the wheels, a roll stiffness control unit having a predetermined relationship between the roll stiffness ratios of the front and rear wheels, and inputting signals from the roll stiffness control unit and the pitching control unit, Supply / discharge control means for controlling the supply / discharge of liquid to / from each liquid cylinder by adding a control amount; and setting the coefficient value of the roll rigidity control unit to a roll of the front wheel as the control amount of the pitching control unit increases. And a correction unit that corrects the rigidity to be smaller than the roll rigidity of the rear wheel.