DE4303039C2 - Semi-active suspension control device based on the Skyhook principle - Google Patents

Description

The present invention relates to a semi-active Suspension control device located between the wheel (or the Axis) and the body of a vehicle is arranged, and a continuous change in the damping coefficient according to the vibration condition of the vehicle evokes.

There are some suggestions in the art regarding Improvements in vibration transmission properties a suspension, which is a change of the Damping coefficients according to the state of the Vertical vibration of the vehicle can cause for example, in U.S. Patent 3,807,678, to which Pages 619-626 of the ASME Journal of Engineering for Industry, No. 96-2, published May 1974, etc. How described in these publications is a  Procedure for controlling the coefficient by judgment the sign of the product of absolute speed S a sprung mass (a body), which the Speed of vertical vibration of the body represents, and the relative speed of the sprung Mass (body) related to unsprung mass (of a wheel) known. A method of controlling the Coefficients by assessing the sign of the Product of the relative displacement of the sprung mass (the Body) with regard to the unsprung mass (the wheel) and their relative speed is known, as in the U.S. Patent No. 4,821,849.

The first-mentioned tax procedure is brief below explained.

It is known in the theory of damping that good damping properties are achieved when a Vibration damper is provided, the one Damping force with respect to the absolute speed S the sprung mass (the body) generated between the sprung mass (the body) and a point that is limited by the absolute coordinate system. However, in a vehicle it is impossible to get one Vibration damper in practice on the absolute To fix the coordinate system. Hence it is considered considered sufficient, the vibration damper between the sprung mass (the body) and unsprung Provide mass (the wheel) in parallel so that its Damping force is variable. In this Trap creates between the sprung mass and the Unsprung mass (the wheel) provided vibration damper Damping force only in the opposite direction Expansion or contraction of the vibration damper. It is common in the professional world to refer to the extension or expansion of the vibration damper as a "rebound stage" and the contraction as a "compression stage". Therefore  the vibration damper can sometimes not have a damping force in it achieve the same direction as a vibration damper that between the sprung mass and the absolute Coordinate system is present. Hence the Damping force assumed to be zero at this time.

The above concept is shown in equation form below.
If

S (S-X) <0 (1)

F = -CsS = -C (S-X) (2)

C = CsS / (S-X) (3)

if

S (S-X) <0 (4)

F = 0 (5)

C = 0 (6)

Here means:

S: absolute speed of the sprung mass (the body);
X: absolute speed of unsprung mass (wheel);
F: damping force of the vibration damper;
Cs: damping coefficient of the vibration damper provided between the sprung mass and the absolute coordinate system;
C: Damping coefficient of the vibration damper provided between the sprung mass and the unsprung mass (the wheel).

Therefore, it is possible to similar good damping properties to obtain that of a vibration damper that is between the sprung mass and the absolute coordinate system is arranged by controlling the  Damping coefficient C of the vibration damper between the sprung mass and unsprung mass (the wheel) is provided, according to equations (3) and (6), and under the conditions of equations (1) and (4).

Furthermore, from the book by J. P. Den Hartog "Mechanische Vibrations ", Verlag Springer, Berlin 1936, pages 114-121, a suspension control device is known in which the Damping is controlled according to the following scheme: with a upward vehicle acceleration and one The axis and frame of the vehicle move apart the damping is large, but with a movement of the axis and There is no damping towards the frame. At a downward acceleration of the vehicle takes place No damping when the axis and the frame differ from each other move away, but strong damping when axis and Move frames together.

However, those described above require How to measure the relative displacement between the body or the sprung mass and the Wheel or unsprung mass, or the Relative speed between these along the Vertical direction. Therefore, such a procedure for a To use the vehicle, a vehicle height sensor was required the body to be attached to the distance to measure between the body and the wheel.

If a vehicle with such a vehicle height sensor where it is snowing, snow often sticks while driving on the height sensor and brings this for freezing. The vehicle will be in the next morning Operated, destruction of the Height sensor because a lever or the like of Height sensor is operated by excessive force.

The present invention has been made in view of the preceding circumstances.

The object underlying the present invention is to provide one Suspension control device in which the Damping coefficient of the suspension simply on the Basis only of the vertical vibration of the body Vehicle can be adjusted without measuring the Relative speed or the relative displacement between the body and the wheel along the  Vertical direction, so without any Vehicle height sensor.

This task is accomplished by a suspension control device with the in claim 1 specified features solved.  

According to the present invention, the Absolute speed of the vertical vibration of the Body calculated; and based on the direction the damping coefficient becomes the absolute speed set that the damping coefficient in the rebound is big if that Damping coefficient in the compression stage of the Vibration damper has a small value, or the Damping coefficient in compression stage one has great value when the damping coefficient a small value in the rebound having. Therefore, the device according to the present invention no height sensor for determination the relative speed or the relative displacement between the body and the wheel along the Vertical direction. Therefore, it is possible to destroy the Avoid height sensor.

Here there the controller sends a control signal to the Vibration damper with variable damping coefficient from, so that the control signal during the damping coefficient the pressure level causes a small value and the damping coefficient during rebound causes it to be approximately equal to the value of the product of absolute speed and one Tax gain is when it is judged that the Body in the upward direction based on the Absolute speed moves and so that the control signal the damping coefficient causes it in the rebound of the  Has a small value, and the Damping coefficients in the compression stage approximately equal to the value of the product Absolute speed and the control gain makes if it is judged that the body is in the Downward direction based on the Absolute speed moves.

In this case, the damping coefficient is determined that it is approximately equal to the value of the product the absolute speed and a control gain. Therefore, it is possible to get the value of the required Determine damping coefficients.

Here is the control gain according to the invention variable, in such a way that as the Absolute values of the vertical acceleration of the body Control gain becomes smaller.  

Advantageous embodiments of the invention are in the Subclaims specified.

The meaning of the absolute value of the vertical acceleration of the Bodywork is shown below explained.

There is a relationship between the acceleration of the Body, the damping force of the vibration damper, and its Relative speed, namely as follows. The acceleration of those arranged in the sprung mass Body is approximately proportional to the damping force of the vibration damper. The damping force of the vibration damper is approximately proportional to the relative speed of the vibration damper.

In other words, regarding the correction of the Damping coefficient according to the reciprocal of the Absolute acceleration of the body's term "Divide by the absolute value of the acceleration of the Body "has the same meaning as part of the  Relative speed of the vibration damper. The smaller the Relative speed is the greater Damping properties can be obtained.

The invention is illustrated below with reference to drawings illustrated embodiments explained in more detail what other advantages and features emerge. It shows  

Fig. 1 is an overall view of a suspension device,

Fig. 2 is a sectional view showing the structure of a shock absorber 4 of variable damping coefficient,

Fig. 3 is a plan view of a movable plate 25 which is mounted on the vibration absorber 4 of variable damping coefficient,

Fig. 4 is a graph with a representation of the relationship between the rotation angle of the movable plate 25, the attenuation coefficient in the rebound, and the attenuation coefficient in the compression stage,

Fig. 5 is a block diagram showing the construction of the controller 6,

Fig. 6 (a) and (b) graph with a representation of the relationship between the control output signal when the vehicle runs on a stage and the displacement of the sprung mass when the vehicle is controlled by the controller 6,

Fig. 7 is a block diagram showing the construction of a further controller 40,

Fig. 8 (a) and (b) graph with a representation of the relationship between the control output signal when the vehicle runs on a stage and the displacement of the sprung mass,

Fig. 9 is a block diagram showing the construction of a further controller 45,

Fig. 10 is a block diagram showing the construction of a further controller 82,

Fig. 11 is a flow chart showing the control contents of the controller 82,

Fig. 12 is a flowchart of a subroutine in step SP5 in Figs. 11 and

Fig. 13 is a table showing the relationship of the direction of movement of the sprung mass and the stroke of the shock absorber.

A first aspect of the invention is explained below with reference to FIGS. 1 to 6.

Fig. 1 is an overall view showing a suspension apparatus for a wheel of a vehicle.

In this figure, reference numeral 1 denotes a body (the sprung mass) of a vehicle, and 2 a wheel (the unsprung mass) located on the side of an axle. Between the body 1 and the wheel 2 , a compression spring 3 and a vibration damper 4 of the type with a variable damping coefficient are provided in parallel. An acceleration sensor 5 , which represents a detection device for detecting the state of the vertical vibration of the body 1 , is attached to the body 1 , which is located on the compression spring 3 . The determined signal with respect to the acceleration is fed to a controller 6 . The controller 6 calculates in a predetermined manner based on the detected signal output from the acceleration sensor 5 , and adjusts the vibration damper 4 with a variable damping coefficient based on the calculated results according to the desired value C of a damping coefficient, as explained in detail below becomes.

The following is the theory of control according to the present invention explained.  

In the control method described in the prior art, the damping coefficient C of the vibration damper 4 , which is provided between the body 1 and the wheel 2 , is determined by the following equations.
If

S (S-X) <0 (1)

C = CsS / (S-X) (3)

if

S (S-X) <0 (4)

C = 0 (6)

However, since the present invention does not use a height sensor but the acceleration sensor 5 , it is impossible to get "SX". Therefore, equation (3) is approximated using the following two equations:

(i) Tax Rule I

The damping coefficient C is based on the Relationship of the following equations controlled under Assumption that "S-X" of equation (3) is an averaged, is constant value.

C = KvS (7)

Kv: a constant

The absolute speed S of the sprung mass (body 1 ) is obtained by integrating the value M of the acceleration sensor 5 . This control is only carried out at the speed of the sprung mass (body 1 ). Therefore, it is possible to lower the control frequency, and the controller has an advantage in the calculation time of the controller 6 .

(ii) Tax Rule II

The damping coefficient C is based on the Relationship of the following equations controlled, namely in that the value M of the sprung mass (the Body) instead of "S-X" in equation (3) is used.

C = KsS / M (8)

Ks: a constant

The acceleration M of the sprung mass is proportional to the force acting on the sprung mass. The power is determined by the sum of the damping force F Equation (2) proportional to the relative speed (S-X) is, and the spring force is expressed. However, if the Vibration is controlled is the relative speed high enough so that the change in spring force can be neglected. Hence the one on the sprung Mass acting force proportional to the damping force F, and the acceleration M of the sprung mass is proportional relative speed (S-X). Therefore, the Acceleration M of the sprung mass instead of Relative speed (F-X) can be used.

Using the above control rules I and II, it is possible to obtain the damping coefficient by using only the value M of the acceleration sensor 5 and not using the equation (3). Therefore, the damping coefficient can be determined by the following equations.

If

S (S-X) <0 (1)

C = KvS (7)

C = KsS / M (8)

or when

S (S-X) <0 (4)

C = Cmin (9)

The reason that the damping coefficient is Cmin, and not 0 if S (S-X) <0 is because of the following.

If the damping force is completely eliminated caused an unstable state before the controller this follows since the control is always related to the Vibration changes are delayed. Therefore determined that C = Cmin to a minimum To specify damping force.

The relationship described above is shown in the table in FIG. 13.

However, as shown in the table in FIG. 13, the stroke of the vibration damper can be judged by the acceleration sensor 5 . The present invention does not require such a judgment since a vibration damper is used in which the damping coefficient in the compression stage is a small constant value when the damping coefficient is variable during extension, and in which case the damping coefficient in the rebound stage is a small constant value is when the damping coefficient is variable during compression. That is, the present invention employs a vibration damper in which the damping coefficient is as variable as shown by broken lines in FIG. 4. Accordingly, the vibration damper itself selects the damping coefficient at its stroke by changing only the rotation angle Theta of the movable plate based on the direction of movement S of the sprung mass. This means that the damping coefficient can be controlled independently of the stroke of the vibration damper, corresponding to the left side (Θ <0) with respect to the coordinate axis in Fig. 4 when the direction of movement of the sprung mass is positive (S <0), and corresponding to the right side (Θ <0) with respect to the coordinate axis in Fig. 4 when the direction of movement of the sprung mass is negative (S <0).

The concrete structure of the vibration damper 4 with a variable damping coefficient, which can obtain the damping coefficient shown in FIG. 4, is described below with reference to FIG. 2.

In Fig. 2, a free piston 12 is slidably inserted into a cylinder 11 so that there is no space therebetween. The inside of the cylinder 11 is divided into a gas chamber 13 and an oil chamber 14 by the free piston 12 . The gas chamber 13 is filled with a gas under high pressure, and the oil chamber 14 is filled with an oil liquid.

A piston 15 is slidably inserted in the oil chamber 14 so that there is no space therebetween. The inside of the oil chamber 14 is divided into a lower oil chamber R1 and an upper chamber R2 by the piston 15 . A piston rod 16 , which extends through the upper chamber R2 to the outside of the cylinder 11 , is connected to the piston 15 .

The piston 15 is provided with a first connection path 17 and a second connection path 18 , each of which connects the lower chamber R1 and the upper chamber R2 to each other. A first damping valve 19 is arranged on the upper surface of the piston 15 . The first damping valve 19 , which is usually closed, is opened to connect the upper chamber R2 and the first communication path 17 when the pressure difference between the lower chamber R1 and the upper chamber R2 reaches a predetermined value by increasing the internal pressure lower chamber R1 during the retraction of the piston rod 16 . A second damping valve 20 is provided on the lower surface of the piston 15 . The second damper valve 20 , which is usually closed, is opened to connect the lower chamber R1 and the second communication path 18 when the pressure difference between the lower chamber R1 and the upper chamber R2 reaches a predetermined value by the internal pressure of the upper chamber R2 rises during the extension of the piston rod 16 .

The piston 15 is provided with a third connecting path 21 and a fourth connecting path 22 , each of which connects the lower chamber R1 to the upper chamber R2. The third connecting path R21 and the fourth connecting path 22 are arranged in the piston opposite to each other with respect to the axis of the piston rod 16 .

Check valves 23 and 24 are arranged in the third connection path 21 and the fourth connection path 22 , respectively. Check valve 23 only allows the flow of oil liquid from lower chamber R1 to upper chamber R2, and check valve 24 only allows flow of oil liquid from upper chamber R2 to lower chamber R1.

A disc-shaped, movable plate 23 is held within the piston 15 so that the movable plate 25 is rotatable about the axis of the piston rod 16 . The upper and lower surfaces of the movable plate 25 are arranged across the third connection path 21 and the fourth connection path 22 .

A pair of elongated openings 25 and 27 are provided opposite to each other in the movable plate 25 on a circle concentric therewith, as shown in FIG. 3. Each of these elongated openings 26 and 27 is elongated along the circumference of the movable plate. The area of the elongated opening 26 decreases as the elongated opening 26 increases in the clockwise direction, as indicated by the arrow P in FIG. 3. The area of the elongated opening 27 increases as the elongated opening 27 increases in the clockwise direction, as indicated by the arrow P in FIG. 3.

When the movable plate 25 is rotated about its axis, the portions of the elongated openings 26 and 27 of the movable plate 25 which face the third connection path 21 and the fourth connection path 22 change continuously. Therefore, the opening area can be changed continuously through the third or fourth connection path 21 or 22 and the elongated opening 26 or 27 . Therefore, it is possible to obtain such damping coefficient properties as shown by the broken lines in FIG. 4.

In Fig. 2, reference numeral 28 denotes an operating rod which is provided along the axis of the piston rod 16 so that it rotates relative to the piston rod. The lower end portion of the operating rod 28 is coupled to the movable plate 25 . The reference numeral 29 designates an actuator such as a stepping motor, which is connected to the upper end of the operating rod 28 to rotate the movable plate 25 in the clockwise direction P or counterclockwise direction Q through the actuating rod 28th The actuator 29 rotates the actuator rod 28 in accordance with control signals (theta) supplied from a block 32 , which will be described later.

Referring to FIGS. 3 and 4, the relationship between sections a2-c2 and c1-a1 of the elongated openings, each opposed to 26 and 27 to the connecting paths 21 and 22, and described the attenuation coefficient.

The positions of each of the communication paths 21 and 22 in the elongated openings 26 and 27 are indicated using the angle of rotation theta of the movable plate 25 . The base position (Θ = 0) of the movable plate 25 is the position in which the connecting paths 21 and 22 face the centers b2 and b1 of the elongated openings 26 and 27, respectively.

(1) When the movable plate 25 is rotated clockwise P from a reference position, that is, when the movable plate 25 is rotated in the positive direction (Θ <0), the connection path 21 faces a position a2 of the elongated opening 26 , and the connecting path 22 is opposite a position a1 of the elongated opening 27 .

Therefore, the oil liquid flows easily from the lower one Chamber R1 to the upper chamber R2, however the flows Oil fluid barely from the upper chamber R2 to the lower one Chamber R1. Therefore, the damping coefficient in the rebound of the vibration damper, and vice versa Damping coefficient in the compression stage of the Vibration damper small.

(2) When the movable plate 25 is rotated counterclockwise Q from a reference position, that is, when the movable plate 25 is rotated in the negative direction (Θ <0), the connection path 21 faces a position c2 of the elongated opening 26 , and the connection path 22 is opposite a position c1 of the elongated opening 27 . Therefore, the oil liquid hardly flows from the lower chamber R1 to the upper chamber R2, but the oil liquid easily flows from the upper chamber R2 to the lower chamber R1. Therefore, the damping coefficient in the rebound stage becomes small, and conversely, the damping coefficient in the compression stage becomes large.

In the present invention, it is theoretically desirable that the vibration damper have the characteristic indicated by the broken lines in FIG. 4. The vibration dampers in the embodiments shown in FIGS. 2 and 3 have the characteristics shown by the solid lines in FIG. 4, namely, a smooth characteristic approximately equal to the broken lines, in order to achieve a smooth change in the damping coefficient. Therefore, the damping coefficients between b1 and c1 in the rebound and between b2 and a2 in the compression are approximately constant small values, although there is some change in these values.

The structure of the controller 6 is explained below. The controller 6 calculates a desired value C of the damping coefficient used to determine the angle of rotation theta of the movable plate 25 according to the control rule I explained above, and the movable plate 25 is rotated based on the desired value C. As can be seen from FIG. 5, the control 6 comprises the blocks 30 to 32 .

The determined signal, which represents the acceleration determined by the acceleration sensor 5 , is input to block 30 . Block 30 calculates the absolute speed of the body of the vehicle by integrating the acceleration. The calculation of block 30 is output to the next block 31 . The reference character "s", which is indicated in block 30 , designates a Laplace operator, which is used for the Laplace transformation.

Block 30 is used to calculate the desired value C of the damping coefficient by multiplying the calculation of block 30 by a control gain Kv. The desired value C has a positive or negative sign. If the absolute speed of the body 1 is positive, that is to say the body has an upward speed, it is determined that the desired value C is positive. In contrast, if the absolute speed of the body 1 is negative, that is, it has a speed in the downward direction, it is determined that the desired value C is negative. The desired value C for the damping coefficient, which represents a calculation of the block 31 , is output to the next block 32 .

Block 32 determines the angle of rotation theta of movable plate 25 based on the desired value C of the damping coefficient output from block 31 . This means that the angle of rotation theta of the movable plate 25 is determined according to the magnitude of the absolute speed of the body 1 and the positive and negative directions of the absolute speed, as shown in the graph in block 32 in FIG. 5.

For example, when the absolute speed of the body 1 in the positive direction becomes larger, that is, in the upward direction of the body 1 , and then the desired value C of the damping coefficient in the positive direction becomes larger, the control signal (theta) is output to the actuator 29 which increases the angle of rotation theta of the movable plate 25 in the positive direction according to the relationship of proportionality shown in the graph in block 32 in FIG . Therefore, the damping coefficient in the rebound is made larger and the damping coefficient in the compression is made smaller, as described in paragraph (1) above.

In contrast, when the absolute speed of the body 1 becomes larger in the negative direction, that is, in the direction of the body 1 downward, and then the desired value C of the damping coefficient in the negative direction becomes larger, the control signal (theta) becomes the operation direction 29 , which makes the angle of rotation theta of the movable plate 25 larger in the negative direction, according to the relationship of proportionality shown in the graph in the block 32 in FIG. 5. Therefore, the damping coefficient is made smaller in the rebound stage and larger in the compression stage, as explained in paragraph (2) above.

In the graph in block 32 , theta is constant in the range in which the absolute value of the desired value C is very large. This is because the angle of rotation theta of the movable plate 25 has a physical limitation, since when the movable plate 25 is rotated through a larger angle than a certain angle, the connection paths 21 and 22 are closed so that they do not connect between them the upper and lower chambers R1 and R2.

The difference between the case in which the movable plate 25 of the vibration damper 4 with variable damping coefficient is controlled by the controller 6 and the case in which the movable plate 25 of the vibration damper 4 with variable damping coefficient 4 is not controlled by the controller 6 explained below with reference to FIGS. 6 (a) and (b).

The solid line in FIG. 6 (a) denotes a displacement of the body 1 controlled by the controller 6 when the vehicle hits a step, and the broken line denotes a displacement of the body 1 which is not controlled by the controller 6 . As is apparent from Fig. 6 (a), it is confirmed that when the vibration damper 4 having a variable damping coefficient is controlled by the controller 6 , the amplitude of the vibration waveform is smaller after it has reached the stage, and a good vibration waveform can be obtained , compared to the case that the vibration damper 4 with variable damping coefficient is not controlled by the controller 6 .

FIG. 6 (b) is a graph showing the type of the output signal of the control signal (theta) with the lapse of time when the vibration damper 4 having the variable damping coefficient is controlled by the controller 6 . The magnitude of the control signal (theta) corresponds to the angle of rotation theta of the movable plate 25 by which it is rotated. This means that when the angle of rotation theta of the movable plate 25 is larger, the absolute value of the control signal (theta) is larger corresponding to the angle of rotation theta.

As explained above in detail, in the above-mentioned suspension control apparatus, the absolute velocity of the body 1 on the basis of the detected signal of the fixed to the body 1 acceleration sensor 5 is calculated, then the rotation angle theta is calculated of the movable plate 25 on the basis of the absolute velocity, and the movable plate 25 is rotated in the direction indicated by the arrow P or the arrow Q in accordance with the rotation angle theta. Therefore, in this suspension device, the absolute speed of the vertical vibration of the body is calculated based on the detected signal from the acceleration sensor 5 ; and based on the absolute speed, the damping coefficient is adjusted so that the coefficient becomes larger in the rebound of the vibration damper and the coefficient in the compression stage of the vibration damper becomes smaller, or so that the damping coefficient becomes smaller in the rebound stage and larger in the compression stage becomes. Therefore, the present invention does not require any height sensor to determine the relative speed or the relative displacement between the body and the wheel along the vertical direction. Therefore, no damage due to the destruction of the height sensor can occur. Since the suspension control device according to the present invention does not require a height sensor for the vehicle at all, the manufacturing cost for the vehicle is reduced.

Here, the vibration damper 4 is controlled on the basis of the determined signal of the acceleration sensor 5 attached to the body 1 , regardless of the state in which the vibration damper 4 is moved out or compressed with a variable damping coefficient. It is therefore possible to continuously set the damping force of the vibration damper 4 to the optimum value.

A further aspect of the present invention is explained below with reference to FIGS. 7 and 8, and this concerns the damping coefficient C according to the control rule II described above, namely a different content of the block which forms the controller 40 . In the following description, the same reference numerals are assigned to the parts which have the same structure as in the first above description to simplify their explanation.

The controller 40 comprises a block 30 , which, as already described, is designed to calculate the absolute speed S of the body 1 by integrating the acceleration M of the sprung mass, and a block 41 to calculate a factor A which corresponds to the desired value C of the damping coefficient corresponds to the first aspect, by multiplying the calculation result of the block 30 by a gain Ks, a block 42 , in which a determined signal representing the acceleration is input from the acceleration sensor 5 , for calculating the absolute value B of the input acceleration; a block 43 for obtaining the desired value C by dividing the factor A obtained from the block 41 by the absolute value B of the acceleration obtained from the block 42 ; and a block 32 , which is the same as in the first aspect, for obtaining the rotation angle theta of the movable plate 25 based on the desired value C calculated in the block 43 .

In block 42 , the acceleration is regarded as a constant value in the positive and negative ranges, in which the acceleration is very small, in order to prevent the desired value C of the damping coefficient, which is calculated in block 43 , both in the positive as well as negative area becomes too large.

Next, the content of the block 43 will be explained.

In this block 43 , the factor A, which corresponds to the desired value C of the damping coefficient in the first aspect, is divided by the absolute value B of the vertical acceleration of the body 1 , which is located on the spring 3 . If the change in the spring force of the spring 3 is not taken into account, the acceleration is proportional to the damping coefficient of the vibration damper 4 . Furthermore, the damping coefficient of the vibration damper 4 is proportional to its relative speed.

The effect of dividing the factor A by the absolute value B of the acceleration of the body 1 on the spring 3 is equivalent to dividing the factor A by the relative speed of the vibration damper 4 . Therefore, the lower the relative speed of the vibration damper 4 , that is, the smaller the damping force of the vibration damper 4 , the greater the desired value C, which is calculated in block 43 . Therefore, it is possible to obtain high damping characteristics to improve the vehicle height characteristics of the vibration damper. The absolute value of the acceleration, not the acceleration itself, is used as size B. The reason for this is that the positive or negative direction of rotation of the movable plate 25 is already represented by the factor A.

Is described below with reference to FIG. 8 (a) and (b) illustrates the difference between the case in which the movable plate of the shock absorber 4 is controlled with a variable damping coefficient by the controller 40 25 and the case in which the movable plate 25 is not controlled by the controller 40 .

The solid line in FIG. 8 (a) denotes the displacement of the body 1 under control by the controller 40 when the vehicle is stepped on, and the broken line indicates the displacement of the body 1 when it is not by the controller 40 is controlled. As is apparent from Fig. 8 (a), it is confirmed that when the vibration damper 4 having the variable damping coefficient is controlled by the controller 40 , the amplitude of the vibration waveform is smaller after the step-up and that a proper vibration waveform is obtained can compared to the case where the vibration damper 4 with variable damping coefficient is not controlled by the controller 40 .

FIG. 8 (b) is a graph showing the type of the output signal of the control signal (theta) with the lapse of time when the vibration damper 4 having the variable damping coefficient is controlled by the controller 40 . The size of the control signal (theta) is equivalent to the angle of rotation theta of the movable plate 25 by which it is to be rotated. Therefore, when the angle of rotation Theta of the movable plate 25 is larger, the absolute value of the control signal (Theta) is larger corresponding to the angle of rotation Theta.

As explained in detail above, the absolute speed of the vertical vibration of the body 1 is calculated on the basis of the detected signal from the acceleration sensor 5 ; and based on the absolute speed, the damping coefficient is adjusted so that the coefficient becomes larger in the rebound of the vibration damper, and the coefficient becomes smaller in the compression, or so that the damping coefficient becomes smaller in the rebound and larger in the compression. Therefore, the suspension control device does not require a height sensor at all to measure the relative speed or the relative displacement between the body and the wheel along the vertical direction. In addition, the smaller the damping force of the vibration damper 4 , the greater the resulting damping coefficient characteristic. For example, when the vehicle hits a step, the amplitude of the vibration waveform can quickly be made smaller after reaching the step to improve the vehicle height characteristic of the damper.

In the above devices, a pair of continuous elongated openings are provided in the movable plate 25 along the circumference thereof, so that the damping coefficient can be changed continuously. However, the present invention is not limited to such embodiments. For example, a plurality of openings, such as three openings on each side, with diameters that gradually increase or decrease as the openings continue clockwise, may be provided in sections a2-c2 and a1-c1 of movable plate 25 so that the damping coefficient can be adjusted gradually. The number of openings is not limited to three with respect to each of the communication paths 21 and 22 . It is possible to provide six to seven openings in the movable plate with respect to each connection path in order to set the damping coefficient in several stages.

Another aspect of the present invention will be explained below with reference to FIG. 9.

This will now a variable control gain Kr used, while one constant control gain Kv was used. Even now are the same reference numerals Assigned to parts that have the same structure as above already described to the explanation simplify.

The controller 45 includes blocks 30 and 46 through 48 . In block 30 , the absolute speed S of the body 1 is calculated by integrating the sprung acceleration M output from the acceleration sensor 5 , and then the calculated result is output to the next block 47 , similarly to the first embodiment.

Block 46 is used to calculate a control gain Kr on the basis of the sprung acceleration M, which is determined by the acceleration sensor 5 . The calculated result is output to the next block 47 . In block 46 , the control gain Kr is set so that when the absolute value of the sprung acceleration M of the body 1 becomes larger, the control gain becomes smaller, and vice versa, to the extent that the absolute value of the sprung acceleration M of the body 1 is small, the control gain is large, as shown in the graph in block 46 of FIG. 9.

The block 47 serves to obtain the desired value of the damping coefficient by multiplying the absolute speed S of the body 1 , which represents the calculated result of the block 30 , by the control gain Kr, which represents the calculated result of the block 46 .

A block 48 determines the rotation angle Theta of the movable plate 25 based on the desired value C of the damping coefficient output from the block 47 and outputs the rotation angle. In block 48 , the angle of rotation theta of the movable plate 25 is determined based on the desired value C of the damping coefficient, using a graph indicating the relationship between the desired value C of the damping coefficient and the angle of rotation theta of the movable plate 25 , as in FIG Fig. 9 shown.

As explained in detail above, in block 46 the control gain Kr multiplied by the absolute speed S of the body 1 is set so that the control gain becomes smaller as the absolute value of the sprung acceleration M of the body 1 increases, and vice versa the control gain becomes larger when the absolute value of the sprung acceleration M of the body 1 becomes smaller. In block 47 , the desired value C of the damping coefficient of the vibration damper 4 with a variable damping coefficient is calculated by multiplying the control gain Kr by the absolute value S of the vertical vibration of the body 1 . With increasing vertical vibration of the body 1 , the damping coefficient that is to be set for the vibration damper 4 with a variable damping coefficient can therefore be limited so that it is relatively lower. Therefore, it is possible to prevent excessive control when the vertical vibration of the body 1 is large. In addition, the damping coefficient to be set for the variable damping coefficient damper 4 can be set relatively high so that the vertical vibration of the body 1 is small, thereby making it possible to prevent lack of control when the vertical movement of the body 1 is small. Therefore, if the vehicle hits a step, it is possible to mitigate the shock the driver feels when the wheel hits a protrusion.

Another aspect of the present invention will be explained below with reference to FIGS. 10 to 12.

The construction is similar to the suspension control device described first, but there is a different content of the control indicated by reference numeral 82 . The structure of the controller 82 is similar to that of the controller 40 . The difference between the controller 82 and the controller 40 is the fact that a switch 83 for changing the control gain Kv is provided to adjust the change in the control gain Kv for a block 31 '.

The type of controller 82 is explained below.

Under the condition that an engine has been started, the control gain Kv set for the block 31 'is initialized (SP1), and after a predetermined period of time has passed (SP2), the detection signal is input from the acceleration sensor 5 to the controller (SP3). Then, the rotation angle Theta of the rotating plate 25 is calculated based on the acceleration M of the sprung mass, which was determined by the acceleration sensor 5 (SP4). In the next step SP5, a control gain Kv for block 31 'is determined.

Therefore, the judgment as to whether the switch 83 for changing the control gain Kv is on or not is made in SP5A. If the switch 83 is switched off, ie logically "No", the control gain Kv is set to Kv1, for a normal mode, in step SP5B. If the switch 83 is switched on, that is to say logically “yes”, the control gain Kv is set to Kv2, for a sport mode, in step SP5C. The control gain Kv1 for a normal mode and the control gain Kv2 for a sport mode have the following relationship: Kv1 <Kv2.

After the control gain is set to Kv1 or Kv2 in step SP5C, the operation returns to step SP2, and then steps SP2 to 5 are repeated again. When the switch 83 is operated, the control gain Kv setting is quickly changed.

Here, the control gain Kv is changed by the switch 83 . However, the switch 83 can also be operated by a horizontal acceleration sensor, which is provided separately for determining the horizontal acceleration.

As described above, the control gain by the switch 83 is changed in two stages, but the present invention is not limited to this. The control gain can be changed in three or more steps or continuously.

As described above, in the suspension control device described last, it is possible to change the setting of the control gain Kv1 or Kv2 by the switch 83 , and to select the control gain according to a driver's desire or the behavior of the vehicle.

The above described suspension devices are provided on each wheel of the vehicle. However it is  possible to simplify the control without in practice to reduce the performance when vibration dampers, at which the damping coefficients both in the rebound and can also be controlled in the pressure stage, used for the front wheels which weighs more because of the engine running in one upper section provided on the side of the front wheels is, and by using vibration dampers, in which only the damping coefficient in rebound is controlled, or even vibration dampers, at which the damping coefficient is not controlled for the rear wheels with little weight.

The present invention relates to a Suspension control device for limiting the vibrations of the Body of the vehicle. However, it is possible to get one to provide comfortable driving behavior, and to improve the controllability of the vehicle, namely by combining the control according to the present Invention with a controller for limiting Rolling movements, the immersion of the front part of the Vehicle, by using a lateral acceleration or longitudinal acceleration.  

Overall, the invention is a semi-active, d. H. by actively controlling the damping of the Vibration damper effectively influencing the vehicle position Suspension control according to the Skyhook principle, at the damper adjustment in terms of driving comfort the alternately active rebound or compression stage solely from the Vertical vibration of the body is determined while body vibration stimulating wheel movements a soft damping setting in the complementary pressure or rebound have no effect.

For this purpose, the invention forms a control with which one Adaptation of the damper effect to the state of motion of the body reached and an unadjusted hard damper effect avoided becomes.

Claims (3)

1. Semi-active suspension control device based on the Skyhook principle with:
a vibration damper ( 4 ) with a variable damping coefficient (C), which is provided between the body ( 1 ) and the wheel ( 2 ) of a vehicle,
a vertical vibration detection device ( 5 ) for detecting the vertical vibration only the body ( 1 ) of the vehicle z. B. on their vertical acceleration, and
a controller ( 45 ) for determining the absolute speed (S) of the vertical vibration of the body ( 1 ) on the basis of the determined signal from the vertical vibration determination device ( 5 ), the controller ( 45 ) transmitting a control signal (Θ) to the vibration damper ( 4 ) outputs variable damping coefficient (C), so that when the absolute speed of the body ( 1 ) is directed upwards, the control signal (Θ) for the damping coefficient (C) takes a small value in the compression stage and the damping coefficient (C) for the Rebound of the vibration damper ( 4 ) becomes approximately equal to the value of the product of the absolute speed (S) and a control gain (Kr), and that when the body ( 1 ) moves downward, the control signal (Θ) for the damping coefficient (C) of the rebound takes a small value and the damping coefficient (C) for the compression is approximately equal to the value of the Prod the absolute speed (S) and the control gain (Kr)
the control gain (Kr) being variable in such a way that the control gain (Kr) becomes smaller as the absolute value of the vertical acceleration of the body ( 1 ) increases. 2. Suspension control device according to claim 1, characterized in that the control of the product (S × Kr) of the absolute speed (S) and the control gain (Kr) on the basis of the reciprocal (1 / M) of the absolute value of the vertical acceleration (M) Corrected body ( 1 ). 3. Suspension control device according to claim 1 or 2, characterized in that the controller ( 82 ) has a switch ( 83 ) through which the control gain (Kr) is gradually adjustable in at least two stages (Kv1, Kv2).