Suspension control device

Abstract

To solve a problem that, although there are many types of commercialized active suspensions that improve comfortableness of a vehicle, even though conventional active suspensions can work at a frequency range of 1 to 10 Hz, they are difficult to reduce sprung vibration at an unsprung resonance frequency (around 12 Hz), and as a result, improvement of comfortableness is limited.SOLUTION: In the present invention, a shock absorber and a power control type actuator are arranged in series. A shock absorber housing and an actuator housing are made integral to be a movable mass. The movable mass and a sprung mass are connected by an actuator rod. A shock absorber rod and a spring are provided between the movable mass and an unsprung mass. As a result, an active mass damper is configured and the actuator is controlled appropriately, so that sprung vibration at the entire frequency range including an unsprung resonance frequency is reduced.SELECTED DRAWING: Figure 6

Description

Technical Field

[0001] The present invention relates to a suspension control device aimed at improving the riding comfort performance.

Background Art

[0002] The present invention relates to a suspension control device aimed at improving the riding comfort performance of a vehicle, and is an invention related to the improvement of Patent Document 1. Here, Example 3 of Patent Document 1 will be described as the background art. FIG. 3 of Patent Document 1 is reproduced as FIG. 1 of the present application, and FIGS. 9 and 10 of Patent Document 1 are reproduced as FIGS. 2 and 3 of the present application. The reference numerals are also in accordance with Patent Document 1.

[0003] In FIG. 1, 25 is a motor coil. 24 is a rotating magnet and rotates integrally with the ball screw nut side 22. In response to the rotation of the ball screw nut side 22, the ball screw male side 26 strokes up and down in the figure integrally with the bolted shock absorber piston rod 30 to constitute an actuator.

[0004] FIG. 2 shows a model and a control block diagram of this device. Here, M1 is the mass under the spring, M2 is the mass above the spring, M r is the total mass of the ball screw male side and the shock absorber rod, C2 is the shock absorber damping constant, C s is the equivalent damping constant of the actuator friction, C q is the equivalent damping constant of the suspension friction, K t is the tire longitudinal spring constant, K2 is the suspension spring constant, K b is the stopper spring constant, X0 is the road surface displacement input, x1 is the displacement under the spring, x2 is the displacement above the spring, x r is the shock absorber rod displacement.

[0005] The control system consists of a sprung acceleration sensor 90, an unsprung acceleration sensor 91, a shock absorber rod acceleration sensor 92, a controller 93, and a ball screw type actuator body 94. The controller 93 integrates each acceleration sensor signal and calculates and outputs a control command value U according to the arithmetic expression shown in the block diagram of the figure. The actuator body 94 receives the control command value U and generates a force F a equal to it.

[0006] Figure 3 shows the result of calculating the PSD (Power Spectral Density) of the sprung acceleration by simulation using the model of Figure 2. The road surface input assumes good road driving. Vehicle specifications of a small passenger car are used, and the parameters of the control command value are shown in the figure. The results of the suspension of Patent Document 1 are shown in comparison with a non-controlled suspension and a conventional active suspension. Looking at the results, it can be seen that the vibration level in the range of 1 to 10 Hz is reduced compared to the conventional active suspension. The characteristics of the conventional active suspension mentioned here are inferred from Non-Patent Document 1.

[0007] Looking at the results, although the conventional active suspension and the active suspension of Patent Document 1 reduce vibration in the frequency range of 10 Hz or less, almost no effect is seen at the unsprung resonance frequency (12 Hz), so the difficulty of reducing the sprung vibration around 12 Hz is recognized.

Summary of the Invention

Problems to be Solved by the Invention

[0008] Conventional active suspensions including Patent Document 1 can produce effects in the frequency range of 10 Hz or less, but it is difficult to reduce the sprung vibration around the unsprung resonance frequency (12 Hz), and there are also limitations in improving the ride comfort.

Means for Solving the Problems

[0009] In the present invention, a shock absorber and an actuator are arranged in series, and a structure having a third movable mass capable of independent movement with respect to the movement of the sprung mass and the unsprung mass is adopted. This movable mass is utilized as an active mass damper to solve the problems.

[0010] By applying and controlling an appropriate command signal to the actuator in this configuration, it becomes possible to reduce the sprung vibration in a wide frequency range including the unsprung resonance frequency compared to the conventional active suspension, and it has become an effective means for solving the above-mentioned difficult problems.

Effects of the Invention

[0011] By applying and controlling an appropriate command signal to the mass damper, which is a movable mass with respect to the sprung and unsprung vibrations, it becomes possible to reduce the sprung vibration in a wide frequency range including the unsprung resonance frequency compared to the conventional active suspension, and a significant improvement in the riding comfort performance is obtained.

Brief Description of the Drawings

[0012]

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[0013] In a suspension control device for a vehicle in which an actuator and a shock absorber are arranged in series between the upper part of the spring (vehicle body side) and the lower part of the spring (axle side), a third movable mass capable of independent movement with respect to the movements of the unsprung mass and the sprung mass is provided to constitute an active mass damper, and an appropriate force command signal is applied to this actuator for control. Embodiments will be described below in Examples 1 to 10. Note that Examples 1 to 7 correspond to claim 5, Example 2 also corresponds to claim 7, Example 5 also corresponds to claim 6, Example 8 corresponds to claim 2, Example 9 corresponds to claim 3, and Example 10 corresponds to claim 4. Claims 8 and 9 regarding the control law are explained in paragraph numbers 44 to 53.

Example

[0014] In Example 1, the shock absorber and the actuator are arranged in series, and the shock absorber housing and the actuator housing are integrated to serve as a movable mass with respect to the movements of the upper part of the spring and the lower part of the spring. The actuator rod connects between this movable mass and the sprung mass, and a shock absorber rod and a spring are provided between this movable mass and the unsprung mass. A specific configuration is shown in FIG. 4. Example 1 in FIG. 4 shows an example using a ball screw type actuator and a shock absorber constituted by a gas chamber 12 with a bladder 11. In FIG. 4, 4 is a motor coil fixed to the housing 7. 3 is a rotating magnet that rotates integrally with the ball screw nut side 1. The upper part of the ball screw male side 2 is attached to the upper part of the spring via an insulator 18. In response to the rotation of the ball screw nut side 1, the housing 7 strokes up and down in the figure. The ball screw that converts the rotational motion of the motor into a linear motion can also be replaced with a sliding screw.

[0015] When the female ball screw side 1 rotates, a rotational reaction force is generated between it and the male ball screw side 2. This rotational reaction force is received by a disk-shaped member 9 provided at the lower part of the ball screw, which has a notch, and a guide 8 fixed to the inside of the housing. A slight gap is set between the notch and the guide 8, allowing the housing 7 to move smoothly up and down. The actuator is composed of the above parts 1 to 9.

[0016] The housing 7 is an integrated and fixed unit of the actuator housing and the shock absorber housing. The rod 17 of the shock absorber is set downward from the shock absorber housing, and its lower end is fixed to a bush 19 for attachment under the spring.

[0017] Stoppers 5 are set so that both the actuator and the shock absorber can smoothly fit within a predetermined stroke without impact on both the extension side and the contraction side. Also, a coil spring 6 is provided in the shock absorber so that the piston is always returned to the neutral position. The coil spring 6 is provided inside the shock absorber, and its function is to return the displacement between the mass under the spring and the housing 7 to the neutral position.

[0018] A port 13 is provided in the piston part of the shock absorber, and the oil chambers 15 above and below the piston always have the same pressure. Therefore, as the shock absorber expands and contracts, the oil in the volume that the rod 17 enters and exits flows through the damping valve 14, generating a damping force. The oil flowing in and out of the damping valve 14 enters and exits a separate oil chamber 10. Above the oil chamber 10, a bladder 11 that separates the gas chamber 12 and the gas chamber 12 from the oil chamber 10 is provided. Although not shown, the separation between the gas chamber 12 and the oil chamber 10 can be achieved by a free piston or a metal bellows. Also, the gas chamber 12 and the oil chamber 10 may be provided in series above the oil chamber 6. The rod guide & oil seal 16 prevents the smooth sliding of the rod 17 and oil leakage.

[0019] When the unit shown in the embodiment of FIG. 4 is incorporated into the suspension and displayed as a model, it will be as shown in FIG. 5. Here, M1 is the mass under the spring, M2 is the mass above the spring, M r is the moving mass consisting of the actuator housing and the shock absorber housing, C2 is the shock absorber damping constant, C s is the equivalent damping constant of the actuator, C q is the equivalent damping constant of the suspension friction, K t is the tire vertical spring constant, K2 is the suspension spring constant, K b is the stopper spring constant of the actuator, K d is the total spring constant of the shock absorber stopper and the coil spring, x0 is the road surface displacement input, x1 is the displacement under the spring, x2 is the displacement above the spring, x r is the displacement of the actuator housing part.

[0020] In the prior art model of FIG. 1, the moving mass M r is the total mass of the actuator rod and the shock absorber rod and is relatively small, and no contribution to performance can be expected. However, in this embodiment, the moving mass M r can be set to be relatively large. This mass M r reduces the control force of the actuator that suppresses the vibration above the spring from acting on the spring below, and reduces the acting force of the damping C2 that suppresses the vibration below the spring from reaching above the spring.

[0021] The control system consists of an acceleration sensor 90 under the spring, an acceleration sensor 91 above the spring, an acceleration sensor 96 of the actuator housing part, a controller 93, and an actuator body 94. The controller 93 integrates each acceleration sensor signal and calculates and outputs a control command value U according to the arithmetic formula shown in the block diagram of the figure. The actuator body 94 receives the control command value U and generates a force F a equal to this.

[0022] Next, the equation of motion of the model in FIG. 5 will be described. The equation of motion of the mass above the spring is as follows.

Equation

Equation

Equation

Equation

[0023] Figure 6 shows the simulation results of the vibration characteristics above the spring. The vehicle equations of motion of Equations 1, 2, and 3 and the control command value of Equation 4 are used. Vehicle specifications of a small passenger car are used. The power spectral density (PSD) of the acceleration vibration above the spring is calculated by applying a road surface input equivalent to a good road.

[0024] The suspension without control in the figure is the same as in Figure 3. Also, the suspension with a passive mass damper in the figure is the case where the actuator command value U = 0 of the suspension control device of the present invention in Figure 5, and a mass damper of mass M r is set between above and below the spring as shown in the figure. The dashed line in the figure republishes the active suspension of Patent Document 1 shown in Figure 3 for reference. The results of the suspension of the present invention are shown in comparison with these. Here, the control law U is Equation (4).

[0025] Looking at the results, the suspension with a passive mass damper shows somewhat of a vibration reduction effect at the sub-spring resonance, but the effect is small below 10 Hz and not observable around 1 Hz. In contrast, in the system of the present invention, the vibration reduction effect in the sub-spring resonance frequency range is remarkable, and a significant effect is also seen in the entire frequency range below 10 Hz. It can be seen that the vibration reduction effect is particularly large at the sub-spring resonance frequency even for the active suspension of Patent Document 1 shown for reference. Thus, in the prior art, it was difficult to reduce the vibration above the spring at around the sub-spring resonance frequency (12 Hz), and there were limitations in the improvement of ride comfort. However, the system of the present invention has been able to solve such problems of the prior art.

[0026] Fig. 7 shows the calculation of the power spectral density (PSD) of the acceleration vibration below the spring by inputting a road surface input equivalent to a good road. It can be seen that the influence of the system of the present invention on the vibration below the spring is small, and no major problems such as rattling during driving and ground contact performance occur.

Example

[0027] Fig. 8 shows a model when a road surface unevenness sensor 95 and control using its signal are added to the system of Fig. 5. The unevenness detection position of the road surface unevenness sensor 95 is near the tire tread. The control command value U is given by the following equation.

Equation

[0028] Fig. 9 shows the calculation of the power spectral density (PSD) of the acceleration vibration above the spring using the model of Fig. 8. Looking at the results, it can be seen that by using the road surface unevenness sensor, the system of the present invention has an even more remarkable vibration reduction effect in the sub-spring resonance frequency range compared to the results of Fig. 6, and also has an even more significant vibration reduction effect in the frequency range below 10 Hz.

[0029] Figure 10 shows the power spectral density (PSD) of the acceleration vibration under the spring calculated using the model of Figure 8. Similarly in this case, it can be seen that the influence on the vibration under the spring is small and no major problems such as rattling during driving or ground contact performance occur.

[0030] Figure 11 shows the PSD of the acceleration on the spring when the value of mass M is changed in the system with a road surface unevenness sensor of the present invention shown in Figure 9. To aim for a sufficient vibration reduction effect on the acceleration on the spring around the resonance frequency under the spring, it can be seen that about M = 10 Kg is necessary, and one measure to achieve this without additional mass is to integrate the actuator housing and the shock absorber housing. r around the resonance frequency under the spring, it can be seen that about M = 10 Kg is necessary, and one measure to achieve this without additional mass is to integrate the actuator housing and the shock absorber housing. r = 10 Kg is necessary, and one measure to achieve this without additional mass is to integrate the actuator housing and the shock absorber housing.

[0031] Figure 12 shows the power spectral density (PSD) of the acceleration vibration under the spring calculated in the same way. It can be seen that increasing M suppresses the resonance under the spring due to the mass - damper effect. r It can be seen that increasing M suppresses the resonance under the spring due to the mass - damper effect.

[0032] Figure 13 shows the transfer function characteristics of the total generated force F of the actuator with respect to the velocity on the spring using the model of Figure 8, and an equivalent model in the frequency range around 10 Hz. Here, F aa is the total generated force of the actuator obtained by adding the generated force by the motor caused by the control command value U, the generated force by the stopper spring K aa of the actuator, and the generated force by the equivalent damping constant C a of the actuator. From this, in the frequency range around 10 Hz, since F b is almost in phase with the velocity on the spring, F s acts as a skyhook damper, and from the gain characteristics, it can be seen that its value is C aa = 100000 N / (m / sec), generating a large amount of damping. As a result, significant vibration reduction is achieved in the resonance frequency range under the spring as shown in Figure 9, and extremely excellent ride comfort performance can be obtained. aa acts as a skyhook damper, and from the gain characteristics, it can be seen that its value is C s2 = 100000 N / (m / sec), generating a large amount of damping. As a result, significant vibration reduction is achieved in the resonance frequency range under the spring as shown in Figure 9, and extremely excellent ride comfort performance can be obtained.

[0033] FIG. 14 similarly shows the transfer function characteristics of the shock absorber generated force F (including the built-in spring force) with respect to the wheel speed using the model of FIG. 8 d and the equivalent model in the frequency range around 10 Hz. From this, around the wheel resonance frequency, F d advances approximately 60 degrees with respect to the wheel speed, so it can be seen that F d exerts the combined action of the skyhook inerter and the skyhook damper. From the gain characteristics, its value is such that the skyhook inerter is M s = 14 N / (m / sec 2 ), and the skyhook damper is C s1 = 423 N / (m / sec).

[0034] Generally, as the upper spring skyhook damper C s2 increases, the lower spring skyhook damper C s1 approaches zero, but in the present invention, it does not, and maintains a value of about 423 N / (m / sec). As a result, as shown in FIG. 10, the wheel resonance level slightly increases with respect to the conventional suspension, but it is within a range that does not affect the running performance such as the grounding property.

Embodiment

[0035] FIG. 15 shows Embodiment 3. Embodiment 3 is a modification of the structure of the shock absorber of Embodiment 1, and the functions and performance of the shock absorber are the same as those of Embodiment 1. The housing 50 of the shock absorber is composed of a triple tube. In the extension stroke of the rod 17, a damping force is generated by the damping valve 51, and in the compression stroke of the rod 17, a damping force is mainly generated by the damping valve 52 (in the compression stroke, the volume component due to the rod entry flows through the damping valve 51, but this flow rate is small and the damping force is also small, so it can be omitted as a damping force).

[0036] The check valve 53 allows the flow from the outside to the inside of the oil chamber partitioned by the tube and acts so as not to allow the reverse flow. Also, the check valve 54 provided in series with the damping valve 52 does not allow oil to flow during the extension stroke and all the oil flows through the damping valve 51. The air chamber 56 absorbs the volume change associated with the entry and exit of the rod 17. During the extension stroke of the rod 17, the oil absorbs the volume change through the port 55 and the check valve 53. During the contraction stroke of the rod 17, the oil absorbs the volume change through the damping valve 51. With the configuration of Example 3 as described above, the gas chamber and the bladder become unnecessary, gas leakage is eliminated, and durability is improved.

Example

[0037] Fig. 16 shows a cross-sectional view of the suspension control device of Example 4. The oil chamber 58 and the gas chamber 59 are isolated via a free piston 57. The spring and the stopper on the compression side of the shock absorber are installed outside the housing. With this configuration, the oil chamber and the gas chamber can be arranged on the same straight line via the free piston, and the compatibility with the conventional shock absorber is high, making it easy to mount on a vehicle. The gas chamber 59 and the oil chamber 58 may be isolated by a metal bellows.

Example

[0038] Fig. 17 shows a cross-sectional view of the suspension control device of Example 5. In Example 4, an outer cylinder 60 is provided to protect the shock absorber rod, the compression side spring, and the stopper in a harsh environment near the spring bottom, facilitate the installation of the suspension spring, and easily accommodate the strut structure with the wheel spindle fixed.

Example

[0039] Fig. 18 shows a cross-sectional view of the suspension control device of Example 6. This is an example when a coil spring type suspension spring 61 is additionally provided in Example 5. Since this suspension spring and the shock absorber can be integrated into one unit, the vehicle mountability is further enhanced.

Example

[0040] Figure 19 shows a cross-sectional view of the suspension control device of Example 7. This is an example in the case where an air spring type suspension spring is provided in addition to Example 5. The air spring consists of an air chamber 62, a diaphragm 63, and a diaphragm guide 64. By guiding the control pressure from an external air pressure source (not shown) through a solenoid valve (not shown) and adjusting the air pressure in the air chamber, vehicle height adjustment is made possible.

Example

[0041] Figure 20 shows a cross-sectional view of the suspension control device of Example 8. The difference from Example 6 is that the actuator housing 66 is mounted on the spring and the actuator rod 67 is fixed to the shock absorber housing 68. By doing so, the vibration input to the actuator can be reduced and the electrical wiring to the actuator can be concentrated on the spring, thus improving reliability. In this case, since the actuator mass is removed from the movable mass and the movable mass is reduced, there is a possibility that the performance may not be sufficient. Therefore, additional mass can be added by increasing the thickness of the upper disk member 9 or the shock absorber housing 68.

Example

[0042] Figure 21 shows a cross-sectional view of the suspension control device of Example 9. The difference from Example 6 is that the actuator housing 66 is mounted on the spring and the actuator rod 67 is fixed to the shock absorber rod 17. In this case, since the movable mass is small, an additional mass 69 is added. By doing so, the vibration input to the actuator can be reduced and the electrical wiring to the actuator can be concentrated on the spring, thus improving reliability as in Example 8. In this configuration, the gas chamber and the free piston are replaced with the air chamber 56, eliminating the concern of gas leakage and increasing durability.

Example

[0043] Figure 22 shows a cross-sectional view of the suspension control device according to Embodiment 10. The difference from Embodiment 6 is that the actuator housing 66 is fixed to the shock absorber rod 17. By doing so, similar to Embodiment 9, the gas chamber and the free piston can be replaced with the air chamber 56, eliminating the concern of gas leakage and increasing durability. Also in this case, since the mass of the shock absorber housing is removed from the moving mass and the moving mass may be reduced and the performance may not be sufficient, additional mass can be added by increasing the thickness of the actuator housing 66 or the like.

[0044] Finally, the derivation methods of the control laws in Equations 4 and 5 will be described. The control law of the present invention is derived from the energy optimal control theory shown in Patent Document 2 and Non-Patent Document 2. The outline of the energy optimal control theory will be described. First, change the idea of obtaining the control law and change the form of the control problem so as to obtain the equation of motion of an ideal system that minimizes the value of the evaluation function. Next, describe the energy flow by expressing the evaluation function in terms of power, and derive the equation of motion of an ideal system that minimizes the evaluation function. This derivation process only applies the conventional variational method and there is no process of solving the differential equation, so it is derived almost automatically. Finally, utilize the fact that "the closed-loop system solves in real time the solution of the simultaneous differential equations of the motion of the controlled object, which is hardware, and the motion that the controlled object should have, which is described by software called the control law".

[0045] Considering that many modern control systems constitute a closed-loop system by computer control, it is a natural measure to assume a closed-loop system in constructing the optimal control theory. Specifically, the characteristics obtained by reversing the input and output of the equation of motion of the aforementioned ideal system are incorporated into the closed-loop system as the control law. By doing so, optimal control is realized by the action of generating the solution of the simultaneous differential equations. Moreover, since the control law is an arithmetic expression based on state variables, real-time control becomes possible.

[0046] Figure 23 shows a comparison between the conventional optimal control problem A and the new optimal control problem B based on the energy optimal control theory. Problem A: When the motion equation of the controlled object and the evaluation function J are given, find the control law U that minimizes the evaluation function. Problem B: For any control law U, find the motion equation of the ideal system that minimizes the given evaluation function J.

[0047] In Problem A, U is to be found, while in Problem B, U is given and the problem is transformed into a system design problem of finding the ideal motion equation that minimizes the evaluation function. In Problem A, L is often unified in a quadratic form, but in Problem B, it is unified by the product of the displacement difference and the flow rate, that is, power. In this way, the motion equation of the ideal system for the input U can be almost automatically derived from the condition of minimizing this evaluation function regardless of the form of the function L.

[0048] Next, the energy optimal control theory utilizes the characteristic that the closed-loop system solves a system of simultaneous differential equations in real time. If the inverse characteristics of the input and output of this ideal system are used as the control law of the closed-loop system, the controlled object will behave as if it is an ideal system within the allowable range of its own dynamics.

[0049] Figure 24 is an explanatory diagram when this theory is applied to the system of the present invention. Define the function L by the following equation.

Equation

[0050] TIFF0007697165000009.tif11133

Mathematics

[0051]

Mathematics

Mathematics

Mathematics

[0052] Furthermore, replace r0, r1, r2, r a , r d with appropriate numerical values and substitute the determined transfer functions shown in Mathematical Formulas 8 to 10 into Mathematical Formula 7 to obtain Mathematical Formula 5. Here, the fourth term on the right side of Mathematical Formula 5 performs a 1 Hz high-pass filter process on the road surface unevenness sensor signal. This is due to specific reasons in road surface unevenness signal processing, such as removing the change caused by the steady pitch angle of the vehicle body and excluding the drift caused by the integration operation when x0 is an estimated value.

[0053] Next, perform a simulation using the optimal control law of Mathematical Formula 5, conduct a transfer function analysis, and the results of confirming the validity of the determined transfer functions in Mathematical Formulas 8 to 10 are shown in Figure 25. From top to bottom in the figure, the determined transfer functions of Mathematical Formulas 8, 9, and 10 are shown by dashed lines, and the true transfer function obtained from the simulation is shown by a solid line. Since the dashed lines approximately approximate the solid lines, it can be inductively confirmed that Mathematical Formula 5 is approximately the optimal control law.

Prior Art Documents

Patent Document

[0054]

Patent Document 1

[0055]

Patent Document 2

Non-Patent Document

[0056]

Non-Patent Document 1

Non-Patent Document 2

Industrial Applicability

[0057] The present invention is corner modularized, and it is easy to be compatible with the springs and shock absorbers of conventional suspensions, has good vehicle mountability, and can be expected to be applied to general vehicles, mainly luxury cars. In addition, the vehicle equipped with the present invention has extremely small shaking, so it is easy for passengers to operate a smartphone or the like. This is considered to be a highly important performance with high needs for future autonomous vehicles.

Explanation of Signs

[0058] 1 Female side of ball screw 2 Male side of ball screw 3 Rotating magnet 4 Motor coil 5 Stopper 6 Coil Spring 7 Housing 8 Guide 9 Disk-shaped Member 10 Separate Oil Chamber 11 Bladder 12 Gas Chamber 13 Port 14 Damper Valve 15 Oil Chamber 16 Rod Guide & Oil Seal 17 Rod 18 Insulator 19 Bush 22 Ball Screw Female Side 24 Rotating Magnet 25 Motor Coil 26 Ball Screw Male Side 30 Shock Absorber Piston Rod 50 Housing 51 Damper Valve 52 Damper Valve 53 Check Valve 54 Check Valve 55 Port 56 Air Chamber 57 Free Piston 58 Oil Chamber 59 Gas Chamber 60 Outer Cylinder 61 Coil Spring 62 Air Chamber 63 Diaphragm 64 Diaphragm Guide 65 Cover 66 Actuator Housing 67 Actuator Rod 68 Shock Absorber Housing 69 Added Mass 90 Spring Bottom Acceleration Sensor 91 Spring Top Acceleration Sensor 92 Actuator Movable Part Acceleration Sensor 93 Actuator Controller 94 Actuator Body 95 Road Surface Irregularity Sensor 96 Actuator Housing Acceleration Sensor

Claims

1. In a suspension control device having an actuator and a shock absorber for controlling the transmission force between above the spring (vehicle body side) and below the spring (axle side) of a vehicle, the actuator housing and the shock absorber housing are integrated, the rod of the actuator is mounted above the spring, and the rod of the shock absorber is mounted below the spring, so that the integrated housing constitutes a moving mass with respect to vibrations on both above and below the spring. A suspension control device for a vehicle, characterized by this.

2. In the configuration of Claim 1, an outer cylinder for sliding the shock absorber housing part is installed, and the lower end of this outer cylinder is fixed to a member for attaching the tip of the shock absorber rod below the spring. A suspension control device for a vehicle, characterized by this.

3. A suspension control device for a vehicle, characterized in that the moving mass is 10 Kg or more in Claim 1.

4. In Claim 1, it has means for measuring or estimating at least the values of the vibration above the spring, the vibration below the spring, the vibration of the moving mass, and the road surface unevenness, and has a control device that outputs a command value to the actuator by calculation using these measured values or estimated values. A suspension control device, characterized by this.

5. In Claim 4, it has a control device that outputs an actuator generated force command value according to the control law shown in the following formula. A suspension control device, characterized by this. 【Number 1】 However, Ra, Rb, Rc, and R0 are positive constants, and Hc and H0 are low-pass filter characteristics. 【Number】 are each of the measured values or their estimated values. When the road surface unevenness signal is not used, the right side of Equation 1 【Number】

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