JP2022145369A - Suspension control device - Google Patents

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 The present invention relates to a suspension control device intended to improve riding comfort.

The present invention relates to a suspension control device for the purpose of improving ride comfort performance of a vehicle, and is an improvement of Patent Document 1.
Here, Example 3 of Patent Document 1 will be described as a background art. FIG. 3 of Patent Document 1 is reprinted as FIG. 1 of the present application, and FIGS. 9 and 10 of Patent Document 1 are reprinted as FIGS. 2 and 3 of the present application. The code is also in accordance with Patent Document 1.

In FIG. 1, 25 is a motor coil. A rotating magnet 24 rotates integrally with the ball screw female side 22 .
As the ball screw female side 22 rotates, the ball screw male side 26 moves up and down together with the screwed shock absorber piston rod 30 to form an actuator.

FIG. 2 shows a model and control block diagram of this device.
where M1 is the _unsprung mass, _M2 is the sprung mass, _Mr is the total mass of the ball screw male side and the shock absorber rod, _C2 is the damping constant of the shock absorber, _Cs is the equivalent damping constant of the actuator friction, C _q is equivalent damping constant of suspension friction, _Kt is tire _longitudinal spring constant, K2 is suspension spring constant, _Kb is stopper spring constant, _X0 is road displacement input, _x1 is unsprung displacement, _x2 is sprung The displacement, _xr , is the shock absorber rod displacement.

The control system consists of an unsprung acceleration sensor 90 , a sprung 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, calculates and outputs a control command value U according to the arithmetic expression shown in the block diagram of FIG. Actuator body 94 receives control command value U and generates force _Fa equal thereto.

FIG. 3 shows the result of calculating the PSD (Power Spectral Density) of the sprung mass acceleration by simulation using the model of FIG. The road surface input assumes driving on a good road.
The vehicle specifications used are those of a small passenger car, and the parameters of the control command values are shown in the figure. The results of the suspension of US Pat. No. 5,700,000 are shown in comparison with the uncontrolled suspension and the conventional active suspension. Looking at the results, it can be seen that the vibration level of 1 to 10 Hz is reduced compared to the conventional active suspension. The characteristics of the conventional active suspension referred to here are assumed from Non-Patent Document 1.

Looking at the results, the conventional active suspension and the active suspension of Patent Document 1 reduce vibration in the frequency range of 10 Hz or less, but there is almost no effect at the unsprung resonance frequency (12 Hz). Difficulty in reducing upper vibration is recognized.

Conventional active suspensions, including Patent Document 1, are effective in the frequency range of 10 Hz or less, but it is difficult to reduce sprung mass vibration around the unsprung resonance frequency (12 Hz), and there is a limit to improving ride comfort. rice field.

In the present invention, the shock absorber and the actuator are arranged in series, and the third movable mass capable of independent movement with respect to the movement of the sprung mass and the unsprung mass is provided, and this movable mass is used as an active mass damper. We are using it to solve problems.

By applying an appropriate command signal to the actuator in this configuration and controlling it, it is possible to reduce sprung mass vibration over a wider frequency range, including the unsprung resonance frequency, compared to conventional active suspensions, thus solving the above difficult problem. It has become an effective means to

Appropriate command signals are applied to control the mass dampers, which are movable masses against sprung and unsprung vibrations, to reduce sprung mass vibrations over a wider frequency range, including unsprung resonance frequencies, compared to conventional active suspensions. This has made it possible to achieve a significant improvement in ride comfort performance.

Conventional suspension control device sectional view Conventional suspension model and control block diagram Sprung acceleration PSD diagram of conventional suspension control system Sectional view of suspension control device of embodiment 1 Suspension model of Example 1 Sprung acceleration PSD diagram of the suspension control device of the first embodiment Unsprung acceleration PSD diagram of the suspension control device of the first embodiment Suspension model of Example 2 Sprung acceleration PSD diagram of the suspension control device of the second embodiment Unsprung acceleration PSD diagram of the suspension control device of the second embodiment PSD diagram of the effect of Mr on the _sprung mass acceleration of Example 2 PSD diagram of the influence of Mr on the _unsprung acceleration of Example 2 Sprung Transfer Function of Suspension Control Device of Embodiment 2 Unsprung transfer function of suspension control device of embodiment 2 Suspension control device sectional view of embodiment 3 Suspension control device sectional view of embodiment 4 Suspension control device sectional view of embodiment 5 Suspension control device sectional view of embodiment 6 Cross-sectional view of the suspension control device of Embodiment 7 Suspension control device sectional view of Embodiment 8 Suspension control device sectional view of embodiment 9 Suspension control device sectional view of embodiment 10 Schematic diagram of energy optimal control theory Application diagram of the energy optimum control theory to the present invention Transfer function representation of the partial derivative term in the optimal control law

A vehicle suspension control system in which an actuator and a shock absorber are placed in series between a sprung mass (body side) and an unsprung mass (axle side). We have devised a suspension control system that constructs an active mass damper by providing a possible third movable mass and controls this actuator by applying an appropriate force command signal. Embodiments will be described below in Examples 1 to 10. FIG.
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, The ninth embodiment corresponds to claim 3, and the tenth embodiment corresponds to claim 4. Claims 8 and 9 relating to control laws were discussed in paragraphs 44-53.

In Embodiment 1, the shock absorber and the actuator are arranged in series, and the shock absorber housing and the actuator housing are integrated to form a movable mass with respect to the sprung and unsprung movements. An actuator rod connects between them, and a shock absorber rod and a spring are provided between the movable mass and the unsprung mass.
A specific configuration is shown in FIG.
Embodiment 1 of FIG. 4 shows an embodiment using a ball screw type actuator and a shock absorber in which a gas chamber 12 is formed by a bladder 11 . In FIG. 4, a motor coil 4 is fixed to the housing 7 . A rotating magnet 3 rotates integrally with the ball screw female side 1 . The upper portion of the ball screw male side 2 is mounted on the spring via an insulator 18 . As the ball screw female side 1 rotates, the housing 7 strokes up and down in the figure. The ball screw that converts the rotary motion of the motor into linear motion can also be replaced with a slide screw.

When the ball screw female side 1 is rotated, a rotational reaction force is generated between it and the ball screw male side 2 . This rotational reaction force is received by a disc-shaped member 9 provided below the ball screw and provided with 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 to allow the housing 7 to move up and down smoothly. The above parts 1 to 9 constitute the actuator.

The housing 7 is formed by integrally fixing the actuator housing and the shock absorber housing, and the rod 17 of the shock absorber is set downward from the shock absorber housing, and its lower end is fixed to the bush 19 for mounting under the spring. It is

A stopper 5 is set so that the actuator and the shock absorber can be smoothly accommodated within a predetermined stroke without impact on both the extension side and the compression side. Also, the shock absorber is provided with a coil spring 6 so that the piston is always returned to the neutral position. A coil spring 6 is provided in the shock absorber and its function is to return the displacement between the unsprung mass and the housing 7 to the neutral position.

A port 13 is provided in the piston portion of the shock absorber, and oil chambers 15 above and below the piston are always at the same pressure. Accordingly, as the shock absorber expands and contracts, the volume of oil that the rod 17 moves in and out flows through the damping valve 14, thereby generating a damping force. The oil flowing in and out of the damping valve 14 flows in and out of the separate oil chamber 10 . A gas chamber 12 and a bladder 11 separating the gas chamber 12 and the oil chamber 10 are provided above the oil chamber 10 . Although not shown, the gas chamber 12 and the oil chamber 10 may be separated from each other 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 and oil seal 16 prevent smooth sliding of the rod 17 and oil leakage.

When the unit shown in the embodiment of FIG. 4 is incorporated into the suspension and displayed as a model, it becomes as shown in FIG.
where M1 is the _unsprung mass, _M2 is the sprung mass, _Mr is the moving mass consisting of the actuator housing and the shock absorber housing, _C2 is the damping constant of the shock absorber, _Cs is the equivalent damping constant of the actuator, C _q is the equivalent damping constant of the suspension friction, _Kt is the tire longitudinal spring constant, K2 is the suspension spring constant, _Kb is the actuator stopper spring constant, _Kd is the _total spring constant of the shock absorber stopper and coil spring, _x0 is the road surface displacement input, x1 is the unsprung displacement, _x2 is the _{sprung displacement, and xr is} the displacement of the actuator housing.

In the prior art model shown in FIG. 1, the moving mass _Mr , which is the total mass of the actuator rod and the shock absorber _rod , is relatively small and cannot be expected to contribute to performance. . This mass Mr reduces the control force of the actuator that suppresses the _sprung mass from acting on the unsprung part, and reduces the acting force of the damping _C2 that suppresses the unsprung mass from acting on the sprung part.

The control system consists of an unsprung acceleration sensor 90 , a sprung acceleration sensor 91 , an actuator housing acceleration sensor 96 , a controller 93 and an actuator body 94 .
The controller 93 integrates each acceleration sensor signal, calculates and outputs a control command value U according to the arithmetic expression shown in the block diagram of FIG. Actuator body 94 receives control command value U and generates force _Fa equal thereto.

バネ下質量の運動方程式は以下の通りである。

質量Ｍ_ｒの運動方程式は以下の通りである。

制御指令値Ｕは次式である。この指令値を受けてアクチュエーターはＦ_ａを発生させる。

Next, the equation of motion of the model in FIG. 5 is described.
The equation of motion for the sprung mass is as follows.

The equation of motion for the unsprung mass is as follows.

The equation of motion for mass M _r is:

The control command value U is given by the following equation. Upon receiving this command value, the actuator generates _Fa .

FIG. 6 shows the simulation results of sprung mass vibration characteristics. Equations 1, 2 and 3 of vehicle motion equations and Equation 4 of control command values are used. Vehicle specifications used are those of a small passenger car. The power spectral density (PSD) of the sprung acceleration vibration is calculated by adding a road surface input equivalent to a smooth road.

The uncontrolled suspension in the figure is the same as in FIG. The suspension with a _passive mass damper in the figure is a case where the actuator command value U of the suspension control system of the present invention in FIG. Equivalent to setting The dashed line in the figure is for reference the active suspension of Patent Document 1 shown in FIG. 3 . The results of the suspension of the present invention are shown in comparison with these. Here, the control law U is Equation (4).

Looking at the results, the suspension with a passive mass damper has some vibration reduction effect in unsprung resonance, but the effect is small at 10 Hz or less, and no effect is seen at around 1 Hz. On the other hand, in the system of the present invention, the effect of reducing vibration in the unsprung resonance frequency range is remarkable, and a significant effect can be seen in the entire frequency range of 10 Hz or less. It can be seen that the active suspension of Patent Document 1 shown as a reference also has a large vibration reduction effect particularly at the unsprung resonance frequency.
As described above, with the conventional technology, it is difficult to reduce the sprung mass vibration around the unsprung resonance frequency (12 Hz), and there is a limit to improving the ride comfort. I was able to solve it.

FIG. 7 shows the power spectrum density (PSD) of unsprung acceleration vibration calculated by inputting a road surface input equivalent to a smooth road. It can be seen that the influence of the system of the present invention on unsprung vibration is small, and there are no major problems such as fluttering feeling during running or road contact.

図５の実施例１の制御則に対して右辺第４項を追加したものである。FIG. 8 shows a model in which a road surface unevenness sensor 95 and control using its signal are added to the system of FIG. The unevenness detection position of the road surface unevenness sensor 95 is in the vicinity of the tire tread surface. A control command value U is given by the following equation.

The fourth term on the right side is added to the control law of the first embodiment in FIG.

FIG. 9 shows the power spectral density (PSD) of the sprung mass acceleration vibration calculated using the model of FIG.
Looking at the results, by using the road surface unevenness sensor, the system of the present invention has a more pronounced vibration reduction effect in the unsprung resonance frequency range than the results in FIG. It can be seen that there is a vibration reduction effect.

FIG. 10 similarly calculates the power spectral density (PSD) of the unsprung acceleration vibration using the model of FIG. In this case as well, the influence on the unsprung vibration is small, and no major problems such as fluttering feeling and grounding during running do not occur.

FIG. 11 shows the sprung mass acceleration PSD when the value of the mass _Mr is changed in the system with the road surface unevenness sensor of the present invention shown in FIG. In order to aim for a sufficient effect of reducing sprung mass vibration around the unsprung resonance frequency, M _r =10 Kg is required. To achieve this without additional mass, the actuator housing and shock absorber housing must be integrated. is one measure.

FIG. 12 similarly calculates the power spectrum density (PSD) of the unsprung acceleration vibration. It can be seen that the unsprung resonance can be suppressed by increasing _Mr due to the mass damper effect.

FIG. 13 shows transfer function characteristics of the total actuator force F _aa with respect to the sprung speed and an equivalent model in a frequency range around 10 Hz using the model of FIG. 8 . Here, _Faa is the force generated by the entire actuator, which is the force Fa generated by the motor according to the control command value U, the force generated by the _{stopper spring Kb of} the actuator, and the force generated by the equivalent damping constant _Cs of the actuator. From this, since F _aa is almost in phase with the sprung velocity in the frequency range around 10 Hz, F _aa acts as a skyhook damper and its value is C _s2 = 100000 N/(m/sec) from the gain characteristic. It can be seen that a large attenuation occurs. As a result, vibrations are greatly reduced in the unsprung resonance frequency region as shown in FIG. 9, and extremely excellent riding comfort performance is obtained.

FIG. 14 similarly shows the transfer function characteristics of the shock absorber generated force (including the built-in spring force) _Fd with respect to the unsprung speed and an equivalent model in the frequency range around 10 Hz using the model of FIG.
From this, Fd around the unsprung resonance frequency advances the unsprung speed by approximately 60 degrees, so it can be seen that Fd exerts a combined action of the skyhook _inerter and the skyhook damper _. . From the gain characteristics, it can be seen that the skyhook inerter has a value of about M _s =14 N/(m/sec ² ) and the skyhook damper has a value of about C _s1 =423 N/(m/sec).

In general, when the sprung skyhook damper _Cs2 is increased, the unsprung skyhook damper _Cs1 approaches zero, but in the present invention, the value is maintained at about 423 N/(m/sec). As a result, as shown in FIG. 10, the unsprung resonance level tends to increase slightly compared to the conventional suspension, but it is within a level that does not affect the running performance such as road contact.

Example 3 is shown in FIG. Example 3 is a modification of the structure of the shock absorber of Example 1, and the function and performance of the shock absorber are the same as those of Example 1. The housing 50 of the shock absorber is composed of a triple tube, and the damping force is generated by the damping valve 51 during the extension stroke of the rod 17, and the damping force is generated mainly by the damping valve 52 during the contraction stroke of the rod 17 (in the contraction stroke, the rod enters Although the volume due to this flows through the damping valve 51, this flow rate is small and the damping force is also small, so it can be omitted as a damping force).

The check valve 53 allows the flow from the outside to the inside of the oil chamber partitioned by the tube, and prevents the reverse flow. Also, the check valve 54 provided in series with the damping valve 52 allows all the oil to flow through the damping valve 51 without allowing it to flow during the extension stroke. The air chamber 56 absorbs changes in volume that accompany the movement of the rod 17 in and out. During the extension stroke of the rod 17, the oil passes through the port 55 and the check valve 53 and absorbs the volume change. During the contraction stroke of the rod 17, the oil passes through the damping valve 51 and absorbs the volume change.
With such a configuration of the third embodiment, the gas chamber and the bladder are unnecessary, and the gas leakage is eliminated, thereby improving the durability.

FIG. 16 shows a cross-sectional view of the suspension control device of the fourth embodiment.
The oil chamber 58 and the gas chamber 59 are separated from each other via the free piston 57 . The spring and stopper on the shock absorber compression side 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, which makes it highly compatible with conventional shock absorbers and easy to mount on a vehicle. A metal bellows may be used to separate the gas chamber 59 and the oil chamber 58 .

FIG. 17 shows a cross-sectional view of the suspension control device of the fifth embodiment.
A strut structure in which an outer cylinder 60 is provided in the fourth embodiment to protect the shock absorber rod, the spring on the compression side, and the stopper in a harsh environment close to the spring, while facilitating the parallel installation of the suspension spring and fixing the wheel spindle. This makes it easier to deal with

FIG. 18 shows a cross-sectional view of the suspension control device of the sixth embodiment.
This is an embodiment in which a coil spring type suspension spring 61 is added to the fifth embodiment. The suspension springs and shock absorbers can be integrated into one unit, further increasing the on-vehicle compatibility.

FIG. 19 shows a cross-sectional view of the suspension control device of the seventh embodiment.
This is an embodiment in which an air spring type suspension spring is added to the fifth embodiment.
The air spring consists of an air chamber 62 , a diaphragm 63 and a diaphragm guide 64 .
A control pressure is introduced from an external air pressure source (not shown) through a solenoid valve (not shown) to adjust the air pressure in the air chamber, thereby making it possible to adjust the vehicle height.

FIG. 20 shows a cross-sectional view of the suspension control device of the eighth embodiment.
The difference from the sixth embodiment is that the actuator housing 66 is mounted on a 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, so reliability can be improved. In this case, since the actuator mass is removed from the movable mass, the movable mass may be reduced and the performance may not be sufficient. .

FIG. 21 shows a cross-sectional view of the suspension control device of the ninth embodiment.
The difference from the sixth embodiment is that the actuator housing 66 is mounted on a spring and the actuator rod 67 is fixed to the shock absorber rod 17 . In this case, the additional mass 69 is added because the movable mass is small. 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 as in the eighth embodiment, so that the reliability can be improved. In this configuration, the gas chamber and the free piston are replaced with the air chamber 56, eliminating the fear of gas leakage and increasing the durability.

FIG. 22 shows a sectional view of the suspension control device of the tenth embodiment.
The difference from the sixth embodiment is that the actuator housing 66 is fixed to the shock absorber rod 17 . By doing so, the gas chamber and the free piston are replaced with the air chamber 56 in the same manner as in the ninth embodiment. In this case as well, the mass of the shock absorber housing is removed from the movable mass, and the movable mass is reduced, which may lead to insufficient performance.

Finally, a method of deriving the control rules of Equations 4 and 5 will be described.
The control law of the present invention is derived from the energy optimum control theory shown in Patent Document 2 and Non-Patent Document 2.
An outline of the energy optimum control theory is explained.
First, we change the form of the control problem by changing the concept of finding the control law and finding the equation of motion of the ideal system that minimizes the value of the evaluation function.
Next, the energy flow is described by the power expression of the evaluation function, and the equation of motion of the ideal system that minimizes the evaluation function is derived. This derivation process only applies the conventional variational method, and since there is no process of solving differential equations, the derivation is almost automatic.
Finally, the fact that "the closed-loop system solves simultaneous differential equations in real time between the motion of the controlled object, which is hardware, and the desired motion of the controlled object, which is described in software called the control law." take advantage of

Considering that many modern control systems are computer-controlled closed-loop systems, it is natural to assume closed-loop systems in constructing optimal control theory. Specifically, the characteristics of the equation of motion of the above-described ideal system with the inputs and outputs reversed are incorporated into the closed loop system as the control law. In this way, optimum control is realized by the action of generating solutions to the simultaneous differential equations. Moreover, since the control law is an arithmetic expression based on state variables, real-time control is possible.

FIG. 23 shows the conventional optimum control problem A and the new optimum control problem B based on the energy optimum control theory in comparison.
Problem A: Given the equation of motion of the controlled object and the evaluation function J, find a control law U that minimizes the evaluation function.
Problem B: For an arbitrary control law U, find the equation of motion of the ideal system that minimizes the given evaluation function J.

In Problem A, U is obtained, while in Problem B, U is given and converted into a system design problem of obtaining an ideal equation of motion that minimizes the evaluation function.
L in Problem A is often unified in a quadratic form, but in Problem B, it is unified in the product of the position difference amount and the circulation amount, that is, the power. By doing so, the equation of motion of the ideal system for the input U can be almost automatically derived from the condition that minimizes this evaluation function regardless of the form of the function L.

Next, the energy optimum control theory utilizes the property that the closed-loop system solves the 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 were an ideal system within the allowable range of its own dynamics.

ここで、ｒ_０、ｒ_１、ｒ_２、ｒ_ａ、ｒ_ｄは重み係数、上式右辺第１項は路面からの入力パワー項に負符号をつけたもの、第２項はアクチュエーターの入力パワー項、第３項はダンパーの発熱項、第４項はバネ下スカイフックダンパーの発熱項、第５項はバネ上スカイフックダンパ

FIG. 24 is an explanatory diagram when this theory is applied to the system of the present invention.
A function L is defined by the following equation.

Here, r ₀ , r ₁ , r ₂ , r _a , and r _d are weighting factors, the first term on the right side of the above equation is the input power term from the road surface with a negative sign, and the second term is the input power of the actuator. term, the third term is the heat generation term of the damper, the fourth term is the heat generation term of the unsprung skyhook damper, and the fifth term is the sprung skyhook damper

項が未知のまま残っている。これを理論的に求めることは難しいので、まずこの偏微分項をラプラス変換した関数を定数や１次の伝達関数として仮置きして試行錯誤的シミュレーションを繰り返しある程度適切な結果が得られるような伝達関数を確定する。

より確定した伝達関数は各数式の矢印ようになる。

terms remain unknown. Since it is difficult to obtain this theoretically, first, by temporarily setting the function obtained by Laplace transform of this partial differential term as a constant or a first-order transfer function, trial-and-error simulations are repeated, and transfer is performed so that appropriate results can be obtained to some extent. Confirm the function.

A more definite transfer function is indicated by the arrows in each equation.

Furthermore, by replacing r ₀ , r ₁ , r ₂ , r _a , and r _d with appropriate numerical values and substituting the fixed transfer functions shown in Equations 8 to 10 into Equation 7, Equation 5 is obtained. Here, the fourth term on the right side of Equation 5 applies high-pass filter processing of 1 Hz to the road surface unevenness sensor signal. This is due to reasons peculiar to road unevenness signal processing, such as to remove the amount of change due to the stationary pitch angle of the vehicle body, and to remove the amount of drift due to integration when _x0 is an estimated value.

Next, a simulation using the optimum control law of Equation 5 was performed, and transfer function analysis was performed to confirm the validity of the determined transfer functions of Equations 8 to 10. The results are shown in FIG. From the top of the drawing, the determined transfer functions of Equations 8, 9, and 10 are indicated by dashed lines, and the true transfer functions obtained from simulation are indicated by solid lines. Since the dashed line approximates the solid line, it can be confirmed by induction that Equation 5 is approximately the optimum control law.

Patent application 2020-154981 Suspension control device

Patent No. 4609767 System optimal control method

"JSAE SYMPOSIUM Vehicle Dynamics Control Learning from the Past" No. 1-19 text p. 58 Figure 30, Date: Friday, July 5, 2019, Venue: Urban Tech Hall, Kogakuin University (1-24-2 Nishi-Shinjuku, Shinjuku-ku), Organizer: Society of Automotive Engineers of Japan, Planning: Vehicle Movement performance department committee Naoto Fukushima, Ichiro Hagiwara, Energy Optimal Control Theory -New Framework of Optimal Control Theory and Its Expandability-, Applied Mathematics, Vol. 21, No. 4, 2011, p. 259-275

The present invention is a corner module, and the conventional suspension springs and shock absorbers are easily compatible with each other. It is highly mountable and can be expected to be applied to general cars, mainly luxury cars.
In addition, since the vehicle equipped with this invention has extremely small shaking, it is easy for passengers to operate smartphones. This is considered to be an important performance that is in high demand for future self-driving vehicles.

1 Ball screw female side 2 Ball screw male side 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 Damping valve 15 Oil chamber 16 Rod Guide & oil seal 17 Rod 18 Insulator 19 Bushing 22 Ball screw female side 24 Rotating magnet 25 Motor coil 26 Ball screw male side 30 Shock absorber piston rod 50 Housing 51 Damping valve 52 Damping 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 Additional mass 90 Unsprung acceleration sensor 91 Sprung acceleration sensor 92 Actuator moving part acceleration sensor 93 Actuator controller 94 Actuator body 95 Road unevenness sensor 96 Actuator housing acceleration sensor

Claims (9)

A third movable mass that can move independently of the vibrations of both the sprung mass and the unsprung mass in a vehicle suspension control device that is installed on the sprung mass (body side) and unsprung mass (axle side). and at least one actuator and one shock absorber as elements connecting the movable mass to the sprung mass and the unsprung mass. In claim 1, the actuator housing is mounted on the spring, the rod of the actuator is fixed to the shock absorber housing, and the rod of the shock absorber is mounted under the spring, so that the shock absorber housing has a spring and an unsprung structure. A vehicle suspension control device comprising a movable mass capable of independent motion with respect to both vibrations. In claim 1, the actuator housing is mounted on the spring, the shock absorber housing is mounted under the spring, the actuator rod and the shock absorber rod are fixed, and an additional mass that moves integrally with these rods is added. A vehicle suspension control device characterized by: In claim 1, the actuator housing is fixed to the shock absorber rod, the actuator rod is mounted on the spring, and the shock absorber housing is mounted under the spring, so that the actuator housing can withstand both on-sprung and unsprung vibrations. A suspension control device for a vehicle, characterized in that a movable mass is configured for the suspension. In claim 1, the actuator housing and the shock absorber housing are integrally structured, and the rod of the actuator is mounted on the spring, and the rod of the shock absorber is mounted under the spring, so that the housing of the integrated structure is formed on the spring and the unsprung. A suspension control device for a vehicle, characterized in that a movable mass is configured for both vibrations. In the constructions of claims 2 and 5, an outer cylinder for sliding the shock absorber housing is installed, and the lower end of this outer cylinder is fixed to a member that attaches the tip of the shock absorber rod to the bottom of the spring. vehicle suspension control device. 2. A suspension control system for a vehicle according to claim 1, wherein the movable mass is 10 kg or more. In any one of claims 1 to 7, it has means for measuring or estimating at least values of sprung mass vibration, unsprung mass vibration, movable mass vibration, and road surface unevenness, and a command to the actuator is given by calculation using these measured values or estimated values. A suspension control device comprising a control device for outputting a value. 9. The suspension control device according to claim 8, further comprising a control device for outputting an actuator generated force command value according to a control rule expressed by the following equation.

However, R _a , R _b , R _c , and R ₀ are positive constants, and H _c and H ₀ are low-pass filter characteristics.

It is a measured value or its estimated value. If the road surface unevenness signal is not used, the fourth term on the right side of Equation 1 is

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