JPH01195109A - Active suspension - Google Patents

Abstract

PURPOSE:To reduce sprung vibration by interposing a damping mechanism between a wheel side member and a car body side member parallel to an active suspension for controlling a hydraulic cylinder by vertical acceleration and vertical speed. CONSTITUTION:Active suspensions 1FL-1RR respectively comprising hydraulic cylinders 5FL-5RR, coil springs 6FL-6RR, pressure control valves 8FL-8RR and shock absorbers 7FL-7RR as damping mechanisms are interposed between a wheel side member 4 and a car body side member. In this arrangement, according to detection signals of vertical acceleration detectors 29FL-29RR of the respective wheels, a control device 30 obtains vertical acceleration and vertical speed of each wheel and controls pressure control valves 8FL-8RR to control working pressure. Simultaneously, the interposed shock absorbers 7FL-7RR function to reduce sprung vibration without increasing a damping ratio of a hydraulic cylinder.

Description

[Detailed Description of the Invention] C. Industrial Field of Application] This invention provides a system in which a fluid pressure cylinder is interposed between each wheel and the vehicle body of a vehicle, and the operating pressure of this fluid pressure cylinder is changed only according to a command value. The present invention relates to an improvement of an active suspension that controls changes in the attitude of a vehicle body by controlling a pressure control valve.

[Conventional technology]

As a conventional active type suspension, there is one described in Japanese Patent Laid-Open No. 62-289420, which was previously filed by the present applicant.

In this prior art, a fluid pressure cylinder is interposed between each wheel and the vehicle body, and the operating pressure of this fluid pressure cylinder is controlled according to the outputs of a vertical acceleration sensor and a vertical speed sensor. This provides good control over changes in posture.

If one form of such a control system is represented in a block diagram, the eighth
As shown in the figure, the transfer function of this control system is as shown in the following equation.

Note that ξ is the damping ratio (active damping ratio) of the hydraulic cylinder, and ω. is the natural circular frequency at the resonance point, and if m is the equivalent mass of the sprung mass at each wheel position and k is the spring constant of the coil spring, it becomes (m / k) l / l, and S
is the Laplace operator.

Therefore, an equivalent model of the above equation (1) is shown in Figure 9, where the spring constant k of the coil spring, the equivalent mass m at each wheel position, and the damping constant 03 of the fluid pressure cylinder are connected in series. It becomes a relationship.

Then, the amplitude ratio (gain) of this transfer function is IG(i
ω)l=1/2ξ1.

[Problem to be solved by the invention] However, in the conventional active suspension described above, the control method was to generate a force proportional to the speed on the spring, so in order to reduce the sprung mass resonance, the consumption There was an unresolved issue that the horsepower would increase.

Furthermore, as a means of detecting the degree of spring improvement, a method is used in which the spring maxillary velocity is integrated using an integrator.

Therefore, in order to reduce sprung mass resonance, the gain of the integrator must be increased, but since the DC offset of the vertical acceleration sensor due to drift etc. may be detected, it is necessary to significantly reduce sprung mass resonance. There were also unresolved issues that it was difficult to do.

Therefore, in addition to the conventional active suspension, this invention significantly reduces sprung resonance by interposing a damping mechanism between each wheel side member and the vehicle body side member,
The purpose of this invention is to solve the problems of the above-mentioned prior art.

[Means to solve the problem]

In order to achieve the above object, the active suspension of the present invention includes a fluid pressure cylinder interposed between each wheel side member and a vehicle body side member, and an operating pressure of the fluid pressure cylinder that is controlled only according to a command value. a pressure control valve that can be varied; a vertical acceleration detection means that detects the vertical acceleration approximately directly above each wheel of the vehicle body; a vertical speed detection means that detects the vertical speed at each of the positions; and the vertical acceleration. a control means for calculating a command value for the pressure control valve based on the vertical acceleration detection value of the detection means and the detection value of the vertical speed detection means, and a control means parallel to the fluid pressure cylinder between each wheel side member and the vehicle body side member. A damping mechanism is provided.

[Effect]

The vertical acceleration and vertical speed at positions corresponding to each wheel of the vehicle body are detected by the vertical acceleration detection means and the vertical speed detection means, respectively, and the control device controls the operating pressure of the fluid pressure cylinder based on these detected values. For, roll, pitch,
Good suppression control can be performed for each wheel against posture changes such as bounce.

In addition, a damping mechanism with an optimal damping constant is installed between each wheel side member and the vehicle body side member, so compared to conventional active suspensions, the vibration transmission rate is lowered to the resonance point. can be significantly reduced.

〔Example〕

Embodiments of the present invention will be described below based on the drawings.

1 to 7 show an embodiment of the present invention.

In Figure 1, IFL, IFR, IRL, I
RR is the vehicle body side member 2 and each wheel 3FL, 3, respectively.
FR, 3RL.

It is an active suspension that is interposed between the wheel side member 4 that individually supports 3RR, and includes hydraulic cylinders 5FL to 5RR as actuators, coil springs 6FL to 6RR, and shock absorbers 7FL to 7FL as damping mechanisms. 7 RR, and hydraulic cylinder 5FL~
Pressure control valves 8FL to 8 that control the hydraulic pressure for 5RH only in response to a command value from a control device 30, which will be described later.
It is equipped with RR.

Here, each of the hydraulic cylinders 5FL to 5RH has its cylinder tube 5a attached to the wheel side member 4,
The piston rod 5b is attached to the vehicle body side member 2, and the working oil pressure in the pressure chamber 19 closed by the piston 5c is controlled by the pressure control valve? Controlled by FL~7RR. Also,
Between the vehicle body side member 2 and the wheel side member 4, coil springs 6FL to 6RR supporting the static load of the vehicle body and shock absorbers 7PL to 7RR as damping mechanisms are provided in parallel with the hydraulic cylinders 5FL to 5RR. is installed.

Moreover, each of the pressure control valves 8FL to 8RR has a cylindrical valve housing 11 and a proportional solenoid 12 integrally provided therein, as shown in FIG.

In the center of the valve housing 11, an upper insertion hole 11U and a lower insertion hole 11L in FIG. 2, which are defined by a partition wall 11A having a valve seat 11a of a predetermined diameter, are coaxially formed. .

Further, a fixed throttle 13 is provided at a lower position above the insertion hole 11L and spaced a predetermined distance from the partition wall 11A, thereby forming a pilot chamber C between the fixed throttle 13 and the partition wall 11A. . Further, a main spool 14 is disposed below the fixed throttle 13 in the insertion hole ILL so as to be slidable in the axial direction thereof, and feedback chambers F and Ft are formed above and below the main spool 14, respectively. At the same time, the upper and lower ends of the main spool 14 are regulated by offset springs 15A and 15B disposed in the feedback chambers Fu and Ft, respectively. and,
Input port 111° control baud) 1 in through hole 11L
and a drain boat 11o are connected to each other in this order, and the input port 11i is connected to the hydraulic oil supply side of the hydraulic power source 24 via the hydraulic piping 25, and the drain port 11o is connected to the hydraulic fluid supply side of the hydraulic power source 24 via the hydraulic piping 26. It is connected to the drain side, and a control bow (1 inch) is further connected to the pressure chambers 19 of the hydraulic cylinders 5FL to 5RR via a hydraulic pipe 27.

The main spool 14 has a land 14a facing the input port lli and a land 1 facing the drain boat llo.
4b, and both of these lands 14a.

It includes a pressure chamber 14c formed of an annular groove formed between the pressure chambers 14b and a pilot passage 14d that communicates the pressure chamber 14c with the lower feedback chamber FL.

In addition, a poppet 16 is disposed in the upper insertion hole 11U so as to be slidable in the axial direction with the valve portion facing the valve seat 11a. The flow rate of the hydraulic oil flowing through the valve seat 11a, that is, the pressure in the pilot chamber C can be adjusted.

Furthermore, the input port lli is a pilot passage 11s.
The drain boat 110 is communicated with the insertion hole 11U via a drain passage lit.

On the other hand, the proportional solenoid 12 includes a plunger 17 that is slidable in the axial direction, and a poppet 16 of the plunger 17.
It has an actuator 17A fixed to the side and an excitation coil 18 that drives the plunger 17 in its axial direction, and this excitation coil 18 is appropriately controlled by a command value I made of a direct current from a control device 30. be done. Thereby, movement of the plunger 17 controls the position of the poppet 16 via the actuator 17A, thereby controlling the flow rate passing through the communication hole 11A. Then, when the pressure in both the feedback chambers FL and FD is balanced with the pressing force being applied to the poppet 16 by the proportional solenoid 12, the spool 14 is in the neutral position and the control boat 1 in and the input port lli and the drain boat llo are cut off.

Here, the relationship between the command value ■ and the control oil pressure P output from the control boat I In is P when the command value I is around zero, as shown in FIG. 1H, and when the command value I increases in the positive direction from this state, the control oil pressure P increases with a predetermined proportional gain, and is saturated with the pressure PL of the oil pressure source 24.

In addition, in FIG. 1, 28H indicates pressure control valves 8FL to 8.
The high pressure side accumulator 28L is connected to the middle of the hydraulic piping 25 between the RR and the hydraulic power source 24, and the pressure control valve 8FL~
This is a low-pressure side accumulator that communicates with the hydraulic pipe 27 between the hydraulic cylinder 8RR and the hydraulic cylinders 5FL to 5RR via a throttle 28V.

On the other hand, each wheel 3FL, 3FR, 3RL,
There are four vertical acceleration detectors 29FL, 29FR, 29RL as vertical acceleration detection means directly above 3RR.
, 29RR are arranged, and these vertical acceleration detectors 2
Vertical acceleration detection signals from 9FL to 29RR are sent to the control device 30.
is input.

As shown in FIG. 1, the control device 30 has control units 31FL to 31RR to which the vertical acceleration detection signals Se and L-2 are individually input. Each control unit 31FL~3
1RR, as shown in FIG. 4, includes a gain adjustment unit 32 that multiplies the vertical acceleration detection signal by a predetermined gain Km, and a vertical speed detection unit that multiplies the integral value of the vertical acceleration detection signal by a predetermined gain Kn. and an addition section 34 as a control means that adds the outputs of the gain adjustment section 32 and the calculation section 33, and the addition output of the addition section 34 corresponds to the command value of the pressure control valves 8FL to 8RH. I 4FL~I
4RII is supplied to each pressure control valve 8.

Next, the operation of the above embodiment will be explained.

Assuming that the vehicle is currently traveling straight at a constant speed on a flat road with no irregularities, and the vehicle height is within the appropriate range, in this condition, the vehicle body will not experience any vibrations such as pitch, roll, or bounce. Therefore, the vertical acceleration detection signal M! of each vertical acceleration detector 29FL to 29RR is FL~'i2□ is zero. This detection signal 5ezy, ~'it□ is the control device 3
0, so the gain adjustment section 32 and calculation section 3
3 and 3 are respectively zero, and the addition output of the adder 34, that is, the command value I4FL~■4□, is also zero.

In this state, if you turn the steering wheel (not shown) clockwise to make a right turn, lateral acceleration will occur in the vehicle body depending on the vehicle speed and steering angle at that time, causing the vehicle body to tilt downward to the left. A roll occurs. When the vehicle enters a roll state in this way, at the start of the roll, the positions corresponding to the wheels 3FR and 3RR on the right side of the vehicle body are upwardly moved.
The positions corresponding to the left wheels 3FL and 3RL are respectively displaced downward, and positive vertical acceleration detection signals are received from the vertical acceleration detectors 29FR and 29RR disposed at these positions! ,,I,! □ is output, and negative vertical acceleration detection signals M and 2 are output from the vertical acceleration detectors 29FL and 29RL, and these are supplied to the control device 3.0.

For this reason, each control unit 30FL to 30RR of the control device 30
In the gain adjustment section 32 and calculation section 33, respectively, the vertical acceleration detection value is L-5! □ is multiplied by the gain Km, and the vertical acceleration detected value'! Vertical speed detection value 5'i, which is an integral value of rt-father □, and Ldt-5 father 11dt are multiplied by a predetermined gain Kn, and both are added in addition section 34 to obtain oil pressure corresponding to right wheels 3FR and 3RR. cylinder 5FR,
Relatively small command value for 5RH■4□+IJRR
A relatively large command value r4rtl14RL is outputted to the hydraulic cylinders 5FL and 5RL corresponding to the left wheels 3FL and 3RL.

Therefore, in the right hydraulic cylinders 5FR, 5RR, the excitation current of the proportional solenoid 12 decreases, so the actuator 17a rises, while the left hydraulic cylinder 5F
As for L and 5RL, the excitation current of the proportional solenoid 12 increases, so the actuator 17a lowers. For this reason, the right hydraulic cylinder 5FR.

5RR is the contraction direction, and the left hydraulic cylinder 5FL
.

Since 5RL is the stretching direction, an anti-roll effect can be exhibited.

By the way, the control mode for exerting the above-mentioned anti-roll effect is such that the spring mass at each wheel position is changed to the spring constant of the M1 coil spring 6 by 1 to the pressure chamber 1 of the hydraulic cylinder 5.
9 pressure is P, the pressure receiving area of the piston 5C is A, the command value of the control device 30 is I4, the vehicle height target value is I, and this vehicle height target value 1. The deviation amount between the command value ■4 is 13, the gain of the pressure control valve 8 is KI, and the damping constant of the shock absorber 7 is CC.
+ Let the amount of displacement under the spring be X, and the amount of displacement on the spring be I2, then the feedback system including the control device 30 corresponding to each wheel is expressed isometrically as shown in FIG.

Here, the inertial resistance MS (2) of the vehicle body can be expressed by the following equation (1), which is the sum of the resistance of the coil spring 6, the resistance of the shock absorber 7, and the resistance of the hydraulic cylinder 5.

MS! , =k (x, -x, ) 1CC・fc
, -*, ) + P-A --(1) Also, the deviation amount ■ can be expressed by the following equation (2).

13F1+Is...(2) Furthermore, the pressure P in the upper hydraulic chamber A is expressed by the following equation (3) by multiplying the output ■3 of the control device 30 by the gain. .

P=, ・■, ......(3) Furthermore, the control device 300 command value ■4 is as described below (4)
It can be expressed by the formula.

14=Km Mz +Kn sentence, ...-(4) Therefore, by substituting formulas (2) to (4) into formula 11, M Se z=k(x+ xz) +Cc(sentence 1- Large, )+KI(Vl-Kmalsoz-Knta2)^=-(K
+KmA X 2+, KnA sentence z+Kxz)+Cc*
, -Ccfatz + kxt + KIAVI...
・・・・・・・・・(5)

If we replace this equation (5) with the Laplace operator S, we get MS”
xz = (K+KmAS”Xz++KnAS
Xz + CC5 When we rearrange the equation and calculate the transfer characteristic formula Xz/X+, we get:
A + Cc) S + K (7). Here, (K, to, A) is the damping constant C1 of the hydraulic cylinder 5, and the sprung equivalent mass at the wheel position (
If M+KI K,A) is set to m, the isometric model of the above equation (7) will be as shown in FIG.

Therefore, the above equation (7) is expressed as follows:
damping ratio (conventional damping ratio) ξo=Cc/ (
2(mk) I′t ), the damping ratio of the hydraulic cylinder 5 (
active damping ratio) ξ, = C, / (2(mk)""
), the above equation (8) is X2 2ξゎω. S+ω. '・・・・・・・・・
...(9) becomes.

Note that the damping ratio Cc of the shock absorber 7 is 0 below.
0) is set so that the formula holds true.

Now, when the amplitude ratio (gain) of the above equation (9) is determined, the following evaluation function is then calculated based on the above equation 01 in order to calculate the relationship between both damping ratios that minimizes the above equation 09. and perform analysis on this evaluation function.

That is, the competitive damping ratio ξ. as a variable, and write the following (
(conversion) formula.

Then, when we analyze the above equation (2) and find the extreme value (convex downward), we get this ξ. It can be seen that when , the above equation (2), that is, the above aυ takes the minimum value.

From the relationship between these two damping ratios, the optimal relationship between the damping constant of the shock absorber 7 and the damping constant of the fluid pressure cylinder 5 is found to be C (= m k / 4 Ca, which is The above 0 represents the relationship between both damping constants.
Matches formula 1.

Further, from the relationship of the α ring, the minimum value of the above equation 01 can be transformed as shown in the following equation 04.

2 (ξ, ′+ξ1 ξC) ・・・・・・・・・・・・αa Here, since (ξ, ”+1/4)≧ξ,
It can be seen that the value of the above equation 041 is always a smaller value than the amplitude ratio (1/2ξ1) calculated in the conventional example. Therefore, as shown in FIG. 7, in this embodiment, the amplitude ratio at the resonance point can be reduced without increasing the damping ratio of the hydraulic cylinder so much. That is, in this embodiment, in an active suspension equipped with a fluid pressure cylinder, a damping mechanism is interposed between each wheel side member and the vehicle body side member, and the damping constant of the damping mechanism and the damping constant of the fluid pressure cylinder are The relationship with the above 0
Since the relationship is as shown in Equation 0, the main spring vibration can be further reduced than in the conventional active suspension.

In the above embodiment, suspension control when the vehicle rolls has been described, but when the vehicle pitches in the longitudinal direction or bounces in the vertical direction, the vertical acceleration detector 29FL detects changes in the vehicle body posture at that time.
~29RR can be detected, so changes in the posture of the vehicle body due to these can be suppressed.

Further, in the above embodiment, the vertical velocity generated almost directly above each wheel of the vehicle body is obtained by integrating the detected value of the vertical acceleration sensor, but the present invention is not limited to this, and for example, a potentiometer can be used. A stroke sensor may be installed approximately directly above each wheel of the vehicle body, and the vertical speed may be calculated by differentiating the detected value of this stroke sensor.

Furthermore, in the above embodiment, a case has been described in which vertical acceleration detectors are provided corresponding to all wheels, but instead of this, for example, three vertical acceleration detectors are provided corresponding to three wheels. Alternatively, the vertical acceleration corresponding to the other wheel may be estimated based on the detected values of these detectors, or the vertical acceleration of the two wheels corresponding to the left (or right) wheel may be estimated. At the same time, a roll rate detector is installed at the intersection of the diagonal lines of each wheel, and based on the detected value of the vertical acceleration detector and the detected value of the roll rate detector, the right side (or left side) The vertical acceleration corresponding to the two wheels may be estimated.

〔Effect of the invention〕

As explained above, in the active suspension of the present invention, a fluid pressure cylinder is provided between each wheel side member and the vehicle body side member, and the fluid pressure cylinder is operated according to the vertical acceleration and vertical speed of the vehicle body. In addition to controlling the pressure, a damping mechanism was interposed between each wheel side member and the vehicle body side member, so
It is not necessary to increase the damping ratio of the fluid pressure cylinder so much, and the main vibration of the spring can be significantly reduced compared to a conventional active suspension, thereby achieving the effect that the horsepower consumption of the suspension can be reduced.

[Brief explanation of the drawing]

Fig. 1 is a configuration diagram of an embodiment of the present invention, Fig. 2 is a sectional view showing an example of a pressure control valve, Fig. 3 is a graph showing the characteristics of the pressure control valve shown in Fig. 2, and Fig. 4 is a control FIG. 5 is a block diagram showing the control section of the device, FIG. 5 is a block diagram of the control system, FIG. 6 is a block diagram showing an equivalent model of this embodiment, and FIG. 7 compares the characteristics of this embodiment and a conventional example. 8 is a block diagram of a conventional control system, and FIG. 9 is a block diagram showing an equivalent model of the conventional example. IFL, IFR, IRL, IRR...active suspension, 3FL, 3FR, 3RL, 3
R...Wheels, 5FL, 5FR. 5RL, 5RR...Hydraulic cylinder, 7FL,
7FR, 7RL. 7RR...Shock absorber (damping mechanism), 8FL
. 8FR, 8RL, 8RR...Pressure control valve, 29F
L, 29PR. 29RL, 291? R...Vertical acceleration detector, 3
0...control device.

Claims (2)

[Claims] (1) A fluid pressure cylinder interposed between each wheel side member and the vehicle body side member, a pressure control valve that can change the operating pressure of the fluid pressure cylinder only according to a command value, and a Vertical acceleration detection means for detecting the vertical acceleration at substantially directly above each wheel; Vertical speed detection means for detecting the vertical speed at each position; and the vertical acceleration detection value of the vertical acceleration detection means and the vertical speed detection means. The vehicle is characterized by comprising: a control means for calculating a command value for the pressure control valve based on a detected value; and a damping mechanism interposed in parallel with the fluid pressure cylinder between each wheel side member and the vehicle body side member. Active suspension. (2) When the damping constant of the damping mechanism is C_c, the damping constant generated in the fluid pressure cylinder is C_a, the sprung equivalent mass at each wheel position is m, and the spring constant of the coil spring is k, the damping of the damping mechanism Let the constant C_c be C_c=m
The active suspension according to claim 1, wherein the active suspension is set to xk/C_a.