JPH02208108A - Active type suspension - Google Patents

Abstract

PURPOSE:To perform damping control on spring more accurately by rotating and driving a pump arranged in a cylinder chamber of a shock absorber or at a position connected to the cylinder chamber in response to a command signal for restricting oscillation on spring. CONSTITUTION:An oscillation control means 18 sends a command to an electric motor 14 of an axial flow pump 12 arranged in a cylinder chamber of a shock absorber 10 based on vibration of a car body 2 detected by a vertical acceleration sensor 16, or oscillation on spring of a spring 8, produces the pressure difference which correspond to the speed of the command to adjust vertical motion of a piston 10a, and restricts oscillation of the car body 2 to prevent resonance on spring. Namely, when damping force produced in the shock absorber cannot damp oscillation, and thus the car body 2 oscillates, a skyhook damper for actively restricting the oscillation is equivalently formed, and vertical vibration on the car body can thus be accurately damped due to attitude control and damping force by the shock absorber 10.

Description

[Detailed Description of the Invention] [Field of Industrial Application] The present invention relates to an active suspension for a vehicle, and in particular, a spring for supporting the vehicle body and a spring for expanding and contracting between the vehicle body and the wheels. The present invention relates to an active suspension that is equipped with a shock absorber that generates a damping force corresponding to speed, and that controls the damping force.

[Conventional technology]

As a conventional active type suspension, for example, there is one described in Japanese Patent Application Laid-Open No. 60-213512, which has already been proposed by the present applicant.

In this conventional example, a spring and a shock absorber for vibration absorption are installed between the sprung mass and the unsprung mass of the vehicle, and a detector that detects the relative displacement between the sprung mass and the unsprung mass or a change in the vehicle posture, and this detection and a controller that controls the damping characteristics of the shock absorber based on the detection signal of the shock absorber. Specifically, the shock absorber includes a cylinder filled with hydraulic fluid, a piston that is separated into an upper liquid chamber and a lower liquid chamber inside the cylinder, and a pipe line that communicates the upper liquid chamber and the lower liquid chamber. The pipe is provided with a damping force generating means such as an elastic container interposed in the middle of the pipe, a fixed orifice, or a relief valve that operates above a set differential pressure. Furthermore, a separate pipe line is provided that communicates the upper liquid chamber and the lower liquid chamber, a pump is disposed in the middle of this pipe line, and the rotation of this pump is controlled by a control signal from the controller. , which generates a pressure difference in the liquid within the pipe. This allows input from the road side,
In response to vibration input near the sprung mass resonance frequency range, which is a relatively low frequency, in order to improve ride comfort, damping force is not intentionally generated and the vibration input is applied to the relatively high frequency sprung mass resonance frequency range. The damping force generating means is designed to be effective against nearby vibration input.

[Problem to be solved by the invention]

However, in the conventional active suspension described above, the damping of vibration input near the sprung mass resonance region depends only on the pressure generated by the electric pump, and the damping force is generated according to the relative speed of the suspension. In order to attenuate the sprung mass resonance, a large amount of power is required to attenuate the sprung mass resonance, and in order to cope with this, the weight and size of the motor increases, which causes problems in vehicles. There is an unresolved problem that it is unsuitable for vehicle suspension, and on the other hand, even if it were installed in a vehicle, there is also an unresolved problem that the damping effect of sprung mass resonance is small despite active damping control. there were.

This invention was made with a focus on such unresolved problems, and aims to efficiently control damping force against vibration input in the sprung mass resonance region with less power. Make it a problem to be solved.

[Means to solve the problem]

In order to achieve the above object, the invention described in claim (1):
A spring that supports the sprung mass and a shock absorber that has a constant damping constant and generates a damping force corresponding to the speed of expansion and contraction between the sprung mass and the unsprung mass are installed between the sprung mass and each unsprung mass of the vehicle. A sprung mass rocking detection means for detecting various kinds of rocking on the spring, and a rocking control means for generating a command signal for suppressing the rocking on the spring based on the detected value of the sprung mass rocking detecting means. and a pump that is rotatably driven in response to a command signal output from the swing control means and is disposed within the cylinder chamber of the shock absorber or in a position communicating with the cylinder chamber.

The invention set forth in claim (2) is characterized in that the sprung swing detection means set forth in claim (1) is a means for detecting vertical acceleration on a spring, and the swing control means is set as a means for detecting vertical acceleration on a spring, and the swing control means is set as a means for detecting vertical acceleration on a spring. This is a means for obtaining a command signal corresponding to the absolute velocity in the vertical direction above.

In the invention described in claims (3) and (4), the sprung mass swing detection means according to claim (1) is a means for detecting longitudinal acceleration and lateral acceleration on the spring, respectively, and the rocking control means comprises: Based on the detected values, the means is used to obtain command signals for suppressing rocking in the longitudinal direction and in the lateral direction, respectively.

In addition, the invention described in claim (5) has the structure described in claim (1), and includes a cylinder in which the shock absorber is connected to the unsprung part or the unsprung part, and a communication device that moves vertically within the cylinder and communicates with the vertical part. A piston in which a hole is formed, a rod that connects the piston to the sprung upper or lower sprung side, a gas chamber that is integrally installed within the cylinder or is installed separately in communication with the cylinder, and the gas chamber and the cylinder. It has a structure in which it has a throttle provided between the chamber and the piston, and a pump is disposed between the throttle and the piston.

Furthermore, the invention described in claim (6) includes a spring supporting the sprung mass between the sprung mass and each unsprung mass of the vehicle, and a spring having a constant damping constant to adjust the expansion and contraction speed between the sprung mass and the sprung mass. A shock absorber that generates a corresponding damping force is installed, a relative displacement detection means detects the relative displacement between the sprung portion and the unsprung portion, and a negative gain is multiplied by the detected value of the relative displacement detection means. a swing control means for obtaining a command signal based on the value of the swing control means; and a pump that is rotatably driven in response to the command signal outputted from the swing control means and is disposed in a cylinder chamber of the shock absorber or in a position communicating with the cylinder chamber. are provided for each wheel.

[Effect]

In the invention described in claim (1), the rocking motion on the spring is detected by the sprung mass rocking detection means, and the rocking control means sends a command signal to the pump based on the detected value.

As a result, when the pump rotates in the commanded direction at the commanded speed, a pressure difference corresponding to the rotational drive state is generated, and this pressure difference adjusts the vertical movement of the piston, so the rocking of the car body is also suppressed, and the spring Upper resonance is prevented. In other words, when the damping force generated by the shock absorber can no longer compensate and the vehicle body swings, a sky footer damper is equivalently configured to actively suppress the swing, and this sky footer damper and shock absorber control the attitude. done jointly. On the other hand, the input vibration corresponding to the unsprung resonance region is C,=1. It is damped by the damping force generated by the vine absorber.

Further, in the invention as set forth in claim (2), in particular, vibrations in the vertical direction on the vehicle body are attenuated accurately. In the invention described in claims (3) and (4), in particular, the roll and scut/nose dive of the vehicle body are respectively suppressed accurately.

Furthermore, in the invention set forth in claim (5), since the working fluid in the cylinder chamber of the shock absorber freely flows through the upper and lower chambers of the piston through the communication hole, when the piston moves up and down, the rod A working fluid proportional to the amount of expansion and contraction flows between the gas chamber and the gas chamber through the throttle, and a damping force corresponding to the speed of expansion and contraction is generated at the throttle.

Furthermore, in the invention set forth in claim (6), the swing control means sends a command signal to the pump based on a value obtained by multiplying the detection value of the relative displacement amount detection means by a negative gain to drive the pump. The pressure difference caused by driving the pump restricts the movement of the piston so as to generate a negative spring force corresponding to the amount of relative displacement. For this reason, the spring force of the spring is equivalently weakened, the sprung mass resonance point becomes significantly lower, and the resonance point peak becomes extremely small, so the vibration transmission rate to the vehicle body hardly increases. . On the other hand, the input vibration corresponding to the unsprung resonance region is attenuated by the damping force generated by the shock absorber.

〔Example〕

(First Embodiment) Hereinafter, a first embodiment of the present invention will be described based on FIGS. 1 to 5.

FIG. 1 is a configuration diagram showing an active suspension system according to the first embodiment for arbitrary wheels that are independently controlled. In the figure, 2 is a vehicle body, 4 is a wheel, and 6 is an active suspension.

The active suspension 6 includes a coil spring 8 for load support and a shock absorber 10 for generating damping force, which are interposed in parallel between the vehicle body 2 and the wheels 4, and an axial flow pump 12 provided within the shock absorber 10. It has an electric motor 14 fixed at a predetermined position on the vehicle body 2 and whose rotating shaft is connected to the axial flow pump 12, and a vertical acceleration sensor 16 as a sprung rocking detection means fixed at a predetermined position on the vehicle body. and a swing control means 18 that controls the rotation of the electric motor 14 based on the detection signal of the sensor 16.

In this embodiment, the shock absorber 10 has a structure of a shock absorber with a gas chamber. Specifically, as shown in the figure, a cylinder 10a fixed to the vehicle body 2 side,
It is movable up and down within this cylinder 10a,
A piston 10b that separates the cylinder chamber vertically to form an upper chamber U and an upper and lower chamber, and a piston rod 10c (rod area S
), a gas chamber 10d that is installed separately from the cylinder 10a and filled with gas at a predetermined pressure, a pipe line 10e that communicates this gas chamber lOd and the cylinder upper chamber U, and this pipe line 10e.
It consists of a diaphragm 10f (damping constant C) inserted in the middle for generating damping force. The piston 10b is formed with a communication hole A that communicates the upper chamber U and the lower chamber of the cylinder lOa, so that hydraulic oil can freely flow between the upper chamber U and the upper and lower chambers. Therefore, the effective area of the piston 10b is equal to the cross-sectional area S of the piston rod 10c.

Therefore, when the wheel 4 vibrates up and down due to unevenness of the road surface, the piston wand 10C1, that is, the piston 10b is displaced up and down. As a result, eight times as much hydraulic oil as the amount of expansion and contraction of the piston rod 10c flows through the throttle 10f into the gas chamber 10d.
, a differential pressure is generated at the throttle 10f, and a damping force is generated.

This damping force characteristic is as shown in FIG. 2, and is constant regardless of the frequency of vibration input.

Further, the axial flow pump 12 is provided in the cylinder upper chamber U, as shown in the figure, and its rotation shaft is connected to the electric motor 14.
is directly connected to the rotating shaft of the The axial flow pump 12 and the electric motor 14 constitute an electric pump in this embodiment.

Furthermore, the vertical acceleration sensor 16 detects the acceleration of the vehicle body in the vertical direction, and outputs an acceleration signal S[2 corresponding to this to the swing control means 18.

The swing control means 18 includes a controller 20 and a driver 2.
2 and an alternator 24. The controller 20 receives the input vertical acceleration detection signal Mt, performs an integral calculation *z=CaSMtdt, and outputs a signal large 2 corresponding to the vertical absolute speed of the vehicle body to the driver 22.

The driver 22 outputs a motor drive signal SS (command signal) to the motor 14 according to the calculated absolute speed signal *2. The alternator 24 supplies power to the driver 22 and the like.

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

Suppose that a vehicle is traveling straight at a constant speed on a flat road. In this running state, there is no vibration input from the road surface, so no vertical displacement of the vehicle body 2 occurs. Therefore, since the detection signal S1g of the vertical acceleration sensor 16 is 0,
The vertical acceleration sensor 2 becomes 0, the motor drive signal SS also becomes 0, and the motor 14 does not rotate. Therefore, the axial flow pump 12 is also in a non-rotating state, and the pressure difference between its ends becomes 0. The vehicle body assumes a flat attitude with a constant vehicle height determined by the spring 8.

Furthermore, the vehicle travels from the above-mentioned constant speed straight ahead state to a road surface with relatively small irregularities, and as a result, high frequency vibrations (e.g. 5Hz-10Hz) near the unsprung resonance region are input to the suspension 6 from the road surface. shall be. This vibration input causes the piston 10b to move slightly up and down, and a small amount of hydraulic oil corresponding to this slight movement moves back and forth between the cylinder upper chamber, lower chambers U and L, and the gas chamber 10d via the throttle 10f. At this time, a damping force against the vibration input is generated at the aperture 10f,
Since vibrations from the road surface are accurately prevented from being transmitted to the vehicle body 2, the rough feeling transmitted from the road surface is reduced, and good riding comfort can be ensured.

Furthermore, assume that the vehicle passes through a undulating road or a very rough road where low-frequency waves with large amplitudes continue, and as a result, there is a relatively low-frequency (for example, H&) vibration input near the sprung mass resonance region. As a result, the piston 10b is largely displaced up and down at low speed, and a large flow of hydraulic fluid corresponding to this displacement is supplied to the gas chamber 10d.
The damping force is generated by the aperture 10f in the same manner as described above.

However, in this running state, depending on the amplitude, the vertical displacement of the vibration input is large, so the damping force generated by the throttle lOf alone cannot absorb the vibrations, and the vehicle body 2 may also be displaced vertically.

When the vehicle body 2 moves up and down in this way, the vertical acceleration sensor 16 detects the vertical acceleration accompanying the up and down movement, and sends a detection signal! 3 is output to the controller 20.

Then, the controller 20 calculates the vertical absolute speed large □, and the driver 22 receives this calculated value large, and the driver 22 calculates the calculated value large,
A motor drive signal SS corresponding to the motor is supplied to the motor 14.

As a result, the motor 14 rotates in the commanded direction at the command speed, so the axial flow pump 12 is also driven in the same way, and the pressure difference corresponding to the rotation speed and rotation direction is applied to both axial ends of the axial flow pump 12, that is, the gas chamber. 10d and cylinder upper and lower chambers U and L
This pressure difference actively suppresses the movement of the piston 10b.

In other words, when the vehicle body 2 is about to sink and the suspension stroke is changing in the direction of contraction, a pressure difference is generated in the direction of pushing down the piston 10b, and on the other hand,
When the vehicle body 2 is about to rise and the stroke is changing in the extension direction, a pressure difference is generated in the direction of attracting the piston tab. As a result, in addition to the damping force caused by the aperture 10f described above, a damping force corresponding to the absolute vertical speed of the vehicle body can be generated, so that the vehicle body 2
It is possible to secure a good riding comfort by reliably suppressing vertical rocking of the vehicle.

Therefore, regarding the damping control of the sprung mass resonance region described above, the present embodiment and the conventional example described above (Japanese Patent Application Laid-Open No. 213512: hereinafter,
(referred to as "conventional example") will be quantitatively analyzed and the differences will be explained.

FIG. 3(a) is an equivalent model diagram of a conventional example. In the figure, 30 is a sensor that detects the amount of relative displacement between the sprung mass and the sprung mass, and the output signal rXI of this sensor 30 is
XzJ is differentially calculated by the controller 31, and the vertical relative velocity becomes 'large', tzJ.

The controller 31 has a large vertical relative speed, and the actuator (pump) generates force F1 (=C()L) due to
l *t ). Here, xt is the vertical displacement of the vehicle body + Xl is the vertical displacement of the tire (assumed to be equal to the road surface displacement), is the spring constant of the spring, and C is the damping constant of the shock absorber. Therefore, an equivalent model diagram that further simplifies the diagram (a) is shown in the diagram (b). In other words, it actively generates a damping force that corresponds to the relative speed of the suspension, and in the vicinity of the sprung mass resonance region it produces the same function as a simple shock absorber.

Therefore, as shown by curve a in FIG. 5, the vibration characteristics near the sprung mass resonance are the same as those of the conventional passive suspension.

At this time, the vibration transmission ratio X @ / X 1 is (
m is the mass of the vehicle body, S is the Laplace operator), and the ratio of the actuator generated force F at the sprung mass resonance point (ω-ωl) to xl is F, C(*, -*ri mCs'XI
Xl ms” +Cs+, so
Substituting s −j (K/m) “”, I F−/x
+ I =K ・” (2) is obtained.

On the other hand, equivalent model diagrams corresponding to FIGS. 3(a) and 3(b) in this embodiment are shown as shown in FIGS. 4(a) and 4(b), respectively. In other words, an axial flow pump 12 has a structure (same as a conventional passive suspension) having a spring constant of 8 degrees and a shock absorber lO of a damping constant C.
This means that a new function has been added to the sky footer damper (a damper that functions in the same way as a damper with one end fixed to a fixed point in the sky), which is expressed by a variable damping constant C1 depending on the differential pressure generated.

With this configuration, the vibration transmission ratio X z / X 1 is , and the actuator (
Axial flow pump 12) The ratio of generated forces F and X is: XI XI s” + (C+C
B) s+, so 5=j(K/m)1/! ,
C1=C8 damping coefficient ζ=C/2 (mK)"" =0.
5, I F −/ x ll = K//f
... (4) is obtained. Therefore, the vibration characteristics are greatly improved compared to the conventional example of curve a, as shown in curve A in Fig. 5, and the actuator generated force F also decreases near the sprung mass resonance region. , (3) (4)
From the comparison of equations, it is reduced to about 1//r. In other words, a greater sprung mass vibration suppression effect can be obtained with less power.

(Second example) Next, a second embodiment will be explained with reference to FIG. 6.7.

Note that the same reference numerals are used for the same components as in the first embodiment, and the configuration for the front left wheel 4FL will be representatively explained (if necessary, the symbol rF corresponding to front to rear right will be used).
L-RRJ).

This second embodiment attempts to perform active control similar to the first embodiment according to the roll of the vehicle body, and as shown in FIG. A sensor 40 is mounted at a predetermined position. This lateral acceleration sensor 40 detects lateral acceleration occurring in the lateral direction of the vehicle body, and sends a lateral acceleration signal y corresponding to this to a controller 42.
This is what is output to. For the four wheels, as shown in FIG. 7, the controller 42 sets gains K, . The signal is multiplied by 42RR and output to the next-stage drivers 44FL to 44RR.

The other configurations are the same as in the first embodiment.

Therefore, in the second embodiment, the same unsprung vibration damping effect as in the first embodiment can be obtained, and the left and right reverse phase generated by each of the axial flow pumps 12FL to 12RR can also be suppressed with respect to sprung vibration damping. A force that resists the roll is generated at each axis position by the differential pressure of

(Third example) Next, a third embodiment will be explained with reference to FIG. 8.9.

The third embodiment has almost the same configuration as the second embodiment, and only the different parts will be described (the same reference numerals are used for the same configurations as the second embodiment).

The third embodiment attempts to perform active control similar to the first embodiment in response to car body scutters and dives. Therefore, as shown in FIG. 8, a longitudinal acceleration sensor 46 as a sprung rocking detection means is mounted on the vehicle body 2 at a predetermined position, and this sensor 46 detects the acceleration generated in the longitudinal direction of the vehicle body. A longitudinal acceleration signal l corresponding to this is output to the controller 48. For the four wheels, as shown in FIG. 9, the controller 48 multiplies the gains that are different and opposite in phase for the front and rear wheels by □ using the gain setters 46FL to 46RR to set the next stage drive. Output to circuits 44FL to 44RR.

The other configurations are the same as in the second embodiment.

Therefore, according to the third embodiment, the same unsprung vibration damping effect as in the first and second embodiments can be obtained, and also the sprung mass vibration can be suppressed before and after the occurrence of each axial flow pump 12FL to 12RR. Nose dive during acceleration/deceleration due to differential pressure in opposite phases,
A force that resists the scut is generated at each axis position, suppressing vehicle body fluctuations with less power consumption.

(Fourth Example) Next, a fourth example will be described with reference to FIGS. 10 to 12. Note that the same reference numerals are used for the same components as in the first embodiment,
Only the different parts will be explained.

This fourth embodiment performs so-called "negative spring control" according to the relative vibration between the vehicle body 2 and the wheels 4. Therefore, instead of the vertical acceleration sensor in the first embodiment, a stroke sensor 50 as a relative displacement detection means is provided between the sprung mass and the sprung mass, and the relative displacement detection signal h of this sensor 50 is Controller 5 in dynamic control means 18
It is configured to be output to 2. In order to generate a negative spring force, the controller 52 multiplies the input signal by a constant negative gain 'KaJ, and applies the calculated value to the driver 22.
This is what is output to.

Therefore, in the fourth embodiment, the same sprung mass damping effect as in each of the above embodiments can be obtained, and for the sprung mass,
When the amount of relative displacement between the vehicle body 2 and the wheels 4 changes, negative spring control is performed to actively generate a negative spring force.

Therefore, as shown in Fig. 11, the spring constant K of the spring 8 becomes equivalently rK-, and the suspension becomes softer, and the vibration transmission characteristic according to this equivalent model becomes curve C in Fig. 12. It will be shown. In other words, a frequency value (ω = ω, where the frequency is, for example, 0.
Although the curve has a peak at a frequency of 6 to 0.7 Hz, the curve can be suppressed to a small value, so that the sprung mass vibration damping can be almost the same as in each of the above embodiments. Also,
Using the stroke sensor 50, the controller 5
The second embodiment has the advantage that the cost of parts is lower than that of the first embodiment due to the simple structure of the second embodiment.

In addition, in each of the above embodiments, the upper and lower parts of the vehicle body 2.

We have explained the case where damping force control based on the acceleration in each direction of the lateral direction 9 and the relative displacement between the vehicle body 2 and the wheels 4 is carried out independently. It is also possible to use a suspension that has both.

Furthermore, the shock absorber and pump (electric motor and axial flow pump) of the present invention are not limited to the structures shown in each of the embodiments described above, and may have the structures shown in FIGS. 13 and 14, for example. good.

Of these, the shock absorber 54 with a gas chamber shown in FIG. 13 has a pump chamber 54g formed in the middle of a conduit 54e from the upper chamber U of the cylinder 54a to the throttle 54f. Similarly, axial flow pump 12
The rest is the same as described above. According to this structure, the vehicle body 2 can be made relatively low.

On the other hand, the shock absorber 56 with a gas chamber in FIG.
, motor 14. The axial flow pump 12° piston 56b is integrally arranged as shown in the figure, and has a structure that has the same function as each of the above-mentioned shock absorbers. This simplifies the overall structure of the suspension, and reduces the chances of oil leaking due to holes caused by rocks thrown from the road surface.

Furthermore, in the shock absorber of the present invention, the conduit or passage leading to the gas chamber may be communicated with the lower cylinder side, thereby increasing the degree of freedom in design.

〔Effect of the invention〕

As explained above, in the invention according to claims (1) to (4), basically the same damping force is generated by the shock absorber against vibration input in the sprung and unsprung resonance regions, When the sprung mass swings in each direction, up and down and horizontally, it equivalently forms a sky footer tamper, and this damper actively suppresses and controls vibrations such as bounce, roll, and pitch of the vehicle body. Although it has a relatively simple configuration compared to the previously proposed method that performs damping control for sprung resonance using only the pressure generated by the electric pump, by making the shock absorber and sky footer damper work together, The generated force related to the damper is small, and the power consumption is reduced by that amount, which makes it possible to promote energy saving, weight reduction, and miniaturization, as well as more accurate sprung mass vibration damping control. Combined with vibration damping, this has the effect of ensuring good riding comfort.

In addition, in the invention set forth in claim (5), there is no sliding resistance for the piston of the shock absorber, and vertical movement is smooth, resulting in a highly durable shock absorber. This has the advantage that a damping constant can be secured.

Furthermore, in the invention described in claim (6), in particular, the spring constant of the suspension is lowered in accordance with the relative displacement between the vehicle body and the wheels, so-called negative spring force is generated, so that the sprung mass resonance point is lowered. As a result, it is possible to obtain effects substantially equivalent to those of the above-mentioned inventions, and there is an advantage that parts costs are low.

[Brief explanation of the drawing]

FIG. 1 is a schematic configuration diagram showing a first embodiment of the present invention, and FIG.
The figure is a graph showing the attenuation constant characteristics of the first embodiment, and Figure 3 (
a) (b) are equivalent model diagrams of the conventional example for comparison with the first embodiment, Figures 4 (a) and (b) are equivalent model diagrams of the first embodiment, and Figure 5 is a comparison with the conventional example. 6 is a schematic configuration diagram showing a second embodiment of the present invention, FIG. 7 is a block diagram regarding sprung mass damping control of the second embodiment, and FIG. 8 is a graph of the vibration characteristics of the first embodiment. 9 is a schematic block diagram showing a third embodiment of the present invention, FIG. 9 is a block diagram regarding sprung mass damping control of the third embodiment, and FIG. 10 is a schematic diagram showing a fourth embodiment of the present invention. FIG. 11 is an equivalent model diagram of the fourth embodiment,
FIG. 12 is a graph of the vibration characteristics of the fourth embodiment shown in comparison with the conventional example and the first embodiment, and FIGS. 13 and 14 are schematic configuration diagrams showing modified examples of the shock absorber. In the figure, 2 is the vehicle body, 4 is the wheel, 6 is the active suspension, 8 is the spring, 10, 54, 56 is the shock absorber, 10a is the cylinder, lob is the piston, 10c is the rod, 10d is the gas chamber, and 10e is the pipe. 10f is a throttle, 12 is an axial flow pump, 14 is an electric motor, 16 is a vertical acceleration sensor, 18 is a swing control means, 4o is a lateral acceleration sensor, 46 is a longitudinal acceleration sensor, 5o is a stroke sensor, A is a communication The holes, U and L are the upper chambers of the cylinder, the royal chambers.

Claims (6)

[Claims] (1) A spring that supports the sprung mass is installed between the sprung mass and each unsprung mass of the vehicle, and the sprung mass has a certain damping constant.
A shock absorber that generates a damping force corresponding to the speed of expansion and contraction between the unsprung parts is installed, and a sprung mass rocking detection means detects various kinds of rocking on the spring, and the detection value of this sprung mass rocking detecting means is a rocking control means for obtaining a command signal for suppressing rocking on the spring based on the vibration control means; and a rocking control means that rotates in response to the command signal outputted from the rocking control means and communicates with the cylinder chamber of the shock absorber or the cylinder chamber. An active type suspension characterized by being equipped with a pump located at a certain position. (2) The sprung mass rocking detection means is a means for detecting vertical acceleration on the spring, and the rocking control means commands a command according to the vertical absolute velocity of the spring based on the detected vertical acceleration value. The active suspension according to claim 1, which is means for obtaining a signal. (3) The sprung mass rocking detection means is a means for detecting longitudinal acceleration on the spring, and the rocking control means generates a command signal for suppressing the rocking in the longitudinal direction based on the detected longitudinal acceleration value. The active suspension according to claim 1, which is a means for determining. (4) The sprung mass rocking detection means is a means for detecting lateral acceleration on the spring, and the rocking control means generates a command signal for suppressing lateral rocking based on the detected lateral acceleration value. The active suspension according to claim 1, wherein the active suspension is means for determining. (5) The shock absorber includes a cylinder connected to the unsprung or unsprung side, a piston that moves up and down within the cylinder and is formed with a communication hole that communicates with the up and down, and the piston and the unsprung or unsprung side. The structure includes a connecting rod, a gas chamber integrally installed in the cylinder or separately installed in communication with the cylinder, and a throttle provided between the gas chamber and the cylinder chamber, and the throttle and the pump is disposed between the pistons (
1) The active suspension described above. (6) A spring that supports the sprung mass is installed between the sprung mass and each unsprung mass of the vehicle, and the sprung mass has a certain damping constant.
Relative displacement detection means for detecting the relative displacement between the sprung mass and the unsprung mass, each of which is equipped with a shock absorber that generates a damping force corresponding to the speed of expansion and contraction between the unsprung mass and the relative displacement detection means. a swing control means for obtaining a command signal based on a value multiplied by a negative gain; and a swing control means that is rotatably driven in accordance with the command signal outputted from the swing control means and communicated with the cylinder chamber of the shock absorber or the cylinder chamber. An active type suspension characterized by having a pump located at each wheel for each wheel.