WO2024142840A1 - Control device - Google Patents

Abstract

This controller comprises: a weighting coefficient calculation unit that, when an actual target damping force is calculated on the basis of a calculated value from a vehicle behavior calculation unit, calculates a weighting coefficient for making the actual target damping force approach a value close to a predetermined target damping force; a vertical movement calculation unit that determines the vertical-direction movement state of a vehicle; a pitch movement calculation unit that determines the pitch-direction movement state of the vehicle; an actual target damping force calculation unit that calculates the actual target damping force on the basis of the weighting coefficient from the weighting coefficient calculation unit and the calculation results of the vertical movement calculation unit and the pitch movement calculation unit; a target damping coefficient calculation unit that calculates a target damping coefficient on the basis of the value calculated by the actual target damping force calculation unit; a correction unit that calculates a corrected damping coefficient produced by lowering an upper limit of the target damping coefficient in a low-speed range in which the relative speed between sprung and unsprung parts of the vehicle is below a prescribed value; and a control signal output unit that outputs a control signal corresponding to the corrected damping coefficient to a damping force regulating-type shock absorber.

Description

Control device

This disclosure relates to a control device that controls a damping force adjustable shock absorber, for example, in a four-wheeled vehicle.

Control devices that control damping force adjustable shock absorbers that are interposed between the vehicle body and wheels are known (Patent Documents 1 and 2). Conventional control devices control the damping force of the damping force adjustable shock absorbers so as to suppress vibrations in the vertical direction of the vehicle body.

JP 2014-69759 A JP 2011-131876 A

Incidentally, conventional control devices perform, for example, bilinear optimal control (BLQ) to suppress vertical vibrations of the vehicle body. In this case, it is preferable to be able to suppress pitch vibrations in addition to vertical (heave) vibrations. However, in order to use BLQ to suppress pitch vibrations in addition to vertical vibrations, optimization must be performed using a full vehicle model, which poses the problem of poor tunability. Meanwhile, control using gain scheduling parameters (GSP) is known for skyhook control. However, when GSP is applied to BLQ, there is an issue that two controllers are required, resulting in poor controllability.

One of the objectives of the present invention is to provide a control device for a damping force adjustable shock absorber that can suppress vertical and pitch vibrations while ensuring tunability.

One embodiment of the present invention is a control device for a damping force adjustable shock absorber that is provided in a vehicle and capable of adjusting the damping force generated, comprising: a weighting coefficient calculation unit that calculates a weighting coefficient for bringing the actual target damping force closer to a pre-determined target damping force when calculating an actual target damping force based on a calculated value from a vehicle behavior calculation unit; a vertical motion calculation unit that calculates the vertical motion state of the vehicle; a pitch motion calculation unit that calculates the pitch motion state of the vehicle; an actual target damping force calculation unit that calculates the actual target damping force based on the weighting coefficient calculated by the weighting coefficient calculation unit and the calculation results of the vertical motion calculation unit and the pitch motion calculation unit; a target damping coefficient calculation unit that calculates a target damping coefficient based on the calculated value from the actual target damping force calculation unit; and a weighting coefficient calculation unit that calculates a weighting coefficient between the sprung and unsprung portions of the vehicle. The damping force control system has a correction unit that calculates a corrected damping coefficient that lowers the upper limit of the target damping coefficient in a region where the relative speed is slower than a predetermined value, and a control signal output unit that outputs a control signal corresponding to the corrected damping coefficient to the damping force adjustable shock absorber, the correction unit sets the corrected damping coefficient so that the damping force increases as the relative speed increases, and the slope of the damping force with respect to the relative speed is small when the relative speed is slower than the predetermined value and the slope of the damping force with respect to the relative speed is large when the relative speed is faster than the predetermined value, the correction unit has a maximum damping coefficient according to the relative speed, and when the target damping coefficient exceeds the maximum damping coefficient, corrects the target damping coefficient to the maximum damping coefficient.

According to one embodiment of the present invention, it is possible to suppress vertical and pitch vibrations while ensuring tunability.

1 is an overall configuration diagram showing a four-wheeled automobile to which a controller according to an embodiment of the present invention is applied; 1 is a diagram illustrating a vehicle behavior control device according to an embodiment of the present invention. FIG. 2 is a control block diagram showing a controller according to the first embodiment. FIG. 4 is a block diagram showing a GSP calculation unit in FIG. 3 . FIG. 4 is a block diagram showing a BLQ controller in FIG. 3 . FIG. 6 is an explanatory diagram illustrating a damping coefficient upper limit calculation unit in FIG. 5 . FIG. 6 is an explanatory diagram illustrating a command value calculation unit in FIG. 5 . 6 is a block diagram showing a target damping force calculation unit and a corrected damping coefficient calculation unit in FIG. 5 . FIG. 6 is a block diagram showing the sprung vertical vibration damping BLQ in FIG. 5 . FIG. 6 is a block diagram showing the roll vibration suppression BLQ in FIG. 5 . FIG. 6 is a block diagram showing the pitch vibration damping BLQ in FIG. 5 . FIG. 1 is an explanatory diagram showing a vehicle model used in designing a sprung vertical vibration damping BLQ. FIG. 2 is an explanatory diagram showing a vehicle model used in designing a roll vibration damping system BLQ. FIG. 2 is an explanatory diagram showing a vehicle model used in designing a pitch vibration damping BLQ. 5A to 5C are characteristic diagrams showing an example of changes over time in a sprung displacement, a sprung acceleration, a pitch angle, a right front wheel command current, and a right rear wheel command current for the first embodiment and a comparative example. FIG. 11 is a control block diagram showing a controller according to a second embodiment. FIG. 17 is a block diagram showing a GSP calculation unit in FIG. 16 . FIG. 11 is a control block diagram showing a controller according to a third embodiment. FIG. 20 is a block diagram showing a BLQ controller in FIG. 18 . FIG. 20 is a block diagram showing the sprung vertical vibration damping BLQ in FIG. 19 . FIG. 20 is a block diagram showing the roll vibration suppression BLQ in FIG. 19 . FIG. 20 is a block diagram showing the pitch vibration damping BLQ in FIG. 19 .

Below, we will explain in detail an example of a control device according to an embodiment of the present invention applied to a four-wheeled vehicle, with reference to the attached drawings.

FIG. 1 shows a vehicle to which a controller 11, which is a control device, is applied. A vehicle behavior control device 1 including the controller 11 is applied to this vehicle. As shown in FIG. 2, the vehicle behavior control device 1 is composed of a suspension device 5 constituting a damping force generating device, and a controller 11 constituting a vehicle control device. Here, in FIG. 2, for example, left and right front wheels and left and right rear wheels (hereinafter collectively referred to as wheels 3) are provided on the underside of a vehicle body 2 constituting the body of the vehicle. The wheels 3 are composed of tires 4, which act as springs that absorb small irregularities in the road surface.

The suspension device 5 is installed between the vehicle body 2 and the vehicle wheels 3. This suspension device 5 is composed of a suspension spring 6 (hereinafter referred to as the spring 6) and a damping force adjustable shock absorber (hereinafter referred to as the variable damper 7) that is installed between the vehicle body 2 and the wheels 3 in a parallel relationship with the spring 6.

Note that FIG. 2 shows a case where one set of suspension devices 5 is provided between the vehicle body 2 and the wheels 3. However, a total of four sets of suspension devices 5 are provided, for example, individually and independently between the four wheels 3 and the vehicle body 2, and FIG. 2 shows only one of these sets diagrammatically.

Here, the variable damper 7 of the suspension device 5 is configured using a damping force adjustable hydraulic shock absorber interposed between the vehicle body 2 and the wheel 3. This variable damper 7 is provided with a variable damping force actuator 8 consisting of a damping force adjustment valve or the like in order to continuously adjust the characteristics of the generated damping force (i.e., the damping force characteristics) from hard characteristics to soft characteristics. The variable damper 7 constitutes a force generating mechanism that generates an adjustable force between the vehicle body 2 and the wheel 3. Note that the variable damping force actuator 8 does not necessarily have to be configured to continuously adjust the damping force characteristics, and may be capable of adjusting the damping force in multiple stages, for example, two or more stages. The variable damper 7 may also be a pressure control type or a flow rate control type. The variable damper 7 may also be a type that controls viscosity, such as a magnetorheological fluid or an electrorheological fluid.

The controller 11 constitutes a control device for the variable damper 7 and controls the damping characteristics of the variable damper 7. The controller 11 is composed of, for example, a microcomputer. The controller 11 is connected to, for example, a CAN9 (Controller Area Network), which is a line network required for data communication. The controller 11 acquires vehicle driving specifications through the CAN9. At this time, the vehicle driving specifications include, for example, wheel speed, slip ratio, longitudinal acceleration, engine torque, brake fluid pressure, road surface friction coefficient, and independent braking/driving force control flags for each of the four wheels. The output side of the controller 11 is connected to the damping force variable actuator 8 of the variable damper 7.

The controller 11 also has a storage unit 11A consisting of a ROM, a RAM, a non-volatile memory, etc. The storage unit 11A of the controller 11 stores various programs, information (vehicle information), data, etc. for controlling the variable damper 7.

As shown in FIG. 3, the controller 11 includes a state estimation unit 12, a GSP calculation unit 13, and a BLQ controller 20.

The state estimation unit 12 estimates the vehicle state (vehicle motion, vehicle behavior) using the CAN signal. That is, the state estimation unit 12 estimates the sprung-unsprung relative velocity, vehicle body absolute vertical velocity (sprung velocity), unsprung-road relative velocity, unsprung absolute vertical velocity (unsprung velocity), roll rate, roll angle, pitch rate, and pitch angle based on the signal (CAN signal) flowing through CAN 9. For example, the state estimation unit 12 estimates the sprung-unsprung relative velocity, vehicle body absolute vertical velocity, unsprung-road relative velocity, unsprung absolute vertical velocity, roll rate, roll angle, pitch rate, and pitch angle based on the vehicle steering angle, longitudinal acceleration, lateral acceleration, yaw rate, wheel speed of each wheel 3, presence or absence of brake operation, and master cylinder hydraulic pressure.

Note that technology for estimating the sprung-unsprung relative velocity, vehicle body absolute vertical velocity, unsprung-road relative velocity, unsprung absolute vertical velocity, roll rate, roll angle, pitch rate, pitch angle, etc. using a vehicle model (equation of motion) that models the vehicle from longitudinal acceleration, lateral acceleration, yaw rate, wheel speed, etc., and a Kalman filter, etc., is described in various documents (documents related to estimating vehicle states), including JP2012-47553A, and therefore will not be described further here.

The state estimation unit 12 constitutes a vehicle behavior calculation unit that calculates vehicle behavior. Specifically, the state estimation unit 12 includes a vertical movement calculation unit that determines the state related to the vertical movement of the vehicle. The state estimation unit 12 includes a roll movement calculation unit that determines the state related to the roll movement of the vehicle. The state estimation unit 12 includes a pitch movement calculation unit that determines the state related to the pitch movement of the vehicle. The state estimation unit 12 includes an attitude change detection unit that detects changes in the attitude of the vehicle body.

The state estimation unit 12 outputs signals corresponding to the estimated state values (motion values), i.e., the sprung-unsprung relative velocity, the vehicle body absolute vertical velocity, the unsprung-road relative velocity, the unsprung absolute vertical velocity, the roll rate, the roll angle, the pitch rate, the pitch angle, etc., to the BLQ controller 20.

The vehicle behavior calculation unit may be various sensors that detect the sprung speed, roll rate, pitch rate, etc. The vehicle behavior calculation unit may also estimate the sprung acceleration, roll rate, pitch rate, etc. based on the wheel speed and road surface information (preview) in the vehicle's traveling direction.

The GSP calculation unit 13 (gain scheduling parameter calculation unit) constitutes a weighting coefficient calculation unit that calculates a weighting coefficient that brings the actual target damping force based on the calculation value of the state estimation unit 12 (vehicle behavior calculation unit) closer to a value close to a pre-determined target damping force. The GSP calculation unit 13 calculates the heave GSP, roll GSP _r and pitch GSP _p based on the vehicle speed acquired from the CAN 9, for example. The heave GSP is a weighting coefficient that adjusts the heave target damping force for suppressing the vertical movement of the vehicle. The roll GSP _r is a weighting coefficient that adjusts the roll target damping force for suppressing the movement of the vehicle in the roll direction. The pitch GSP _p is a weighting coefficient that adjusts the pitch target damping force for suppressing the movement of the vehicle in the pitch direction.

That is, when the BLQ controller 20 calculates an actual target damping force based on the vehicle state (e.g., sprung-unsprung relative velocity, vehicle body absolute vertical velocity, unsprung-road surface relative velocity, unsprung absolute vertical velocity, roll rate, roll angle, pitch rate, pitch angle, etc.) which is a calculated value of the state estimation unit 12 (vehicle behavior calculation unit), the GSP calculation unit 13 calculates weighting coefficients (heave GSP, roll GSP _r , pitch GSP _p ) that bring this actual target damping force closer to a value close to a pre-calculated target damping force. The GSP calculation unit 13 outputs the heave GSP, roll GSP _r , and pitch GSP _p to the BLQ controller 20.

As shown in Fig. 4, the GSP calculation unit 13 includes a lookup table that records the relationship between the vehicle speed and the heave GSP. At this time, the heave GSP is determined by tuning in an actual vehicle test. Specifically, the heave GSP is finely adjusted by referring to vibration data and the like after the rough characteristics are determined by a sensory evaluation in an actual vehicle test. The GSP calculation unit 13 also includes a lookup table that records the relationship between the vehicle speed and the roll GSP, and a lookup table that records the relationship between the vehicle speed and the pitch GSP. The roll _GSPr and the pitch _GSPp are also determined by tuning in an actual vehicle test, like the heave GSP.

The heave GSP, roll GSP _r and pitch GSP _p are normalized values between 0 and 1. The heave GSP, roll GSP _r and pitch GSP _p have small values at low speeds and large values at high speeds, as shown in Fig. 4 for example. Note that the heave GSP, roll GSP _r and pitch GSP _p are not limited to this and may be constant values that are the same regardless of the vehicle speed.

The BLQ controller 20 calculates a command current (control signal) that becomes a command value based on bilinear optimal control theory. As shown in FIG. 5, the BLQ controller 20 includes a sprung vertical vibration damping BLQ21, a roll vibration damping BLQ22, a pitch vibration damping BLQ23, a target damping force calculation unit 24, a damping coefficient upper limit calculation unit 25, a corrected damping coefficient calculation unit 26, and a command value calculation unit 29.

The vehicle state output from the state estimation unit 12 and the heave GSP output from the GSP calculation unit 13 are input to the sprung vertical vibration damping BLQ21. The sprung vertical vibration damping BLQ21 calculates the heave target damping force of the variable damper 7 to reduce vertical vibration based on the sprung-unsprung relative velocity, vehicle body absolute vertical velocity, unsprung-road surface relative velocity, and unsprung absolute vertical velocity output from the state estimation unit 12, and the heave GSP output from the GSP calculation unit 13. The sprung vertical vibration damping BLQ21 is designed based on the vehicle model shown in FIG. 12, for example. FIG. 12 illustrates an example in which one set of suspension devices 5 is provided between the vehicle body 2 and the wheels 3. However, the suspension devices 5 are provided individually and independently in total, for example, between the four wheels 3 and the vehicle body 2, and only one of these is illustrated diagrammatically in FIG. 12. The sprung vertical vibration damping BLQ21 is designed based on the state equation of the vehicle model in FIG. 12, as shown in the design method described below.

As shown in FIG. 9, the sprung vertical vibration damping BLQ21 includes a Bp calculation unit 21A, a multiplier 21B, a Bp calculation unit 21C, a multiplier 21D, a multiplier 21E, and a heave target damping force calculation unit 21F. The Bp calculation unit 21A calculates the maximum value of Bp according to the relative speed based on the previously calculated Bp and the heave GSP. The multiplier 21B multiplies the maximum value of Bp output from the Bp calculation unit 21A by the relative speed. The Bp calculation unit 21C calculates the maximum value of Bp according to the sprung speed based on the previously calculated Bp and the heave GSP. The multiplier 21D multiplies the maximum value of Bp output from the Bp calculation unit 21C by the sprung speed. The multiplier 21E multiplies the value calculated by the multiplier 21B by the value calculated by the multiplier 21D. The heave target damping force calculation unit 21F multiplies the calculated value by the multiplier 21E by the inverse matrix (R ^-1 ) of R obtained in advance to calculate the heave target damping force. In this way, the sprung vertical damping BLQ21 calculates a value obtained by multiplying the right side of the equation (12) described later by the relative speed from the left side. As a result, the sprung vertical damping BLQ21 calculates the heave target damping force by multiplying the left side of the equation (12) described later by the relative speed from the left side. Note that, although the sprung vertical damping BLQ21 has exemplified a case in which the heave target damping force is calculated based on Bp according to the relative speed and Bp according to the sprung speed, the present invention is not limited to this. The sprung vertical damping BLQ21 may calculate the heave target damping force based on Bp according to other vehicle conditions in addition to Bp according to the relative speed and sprung speed.

The roll vibration suppression BLQ22 receives the vehicle state output from the state estimation unit 12 and the roll GSP _r output from the GSP calculation unit 13. The roll vibration suppression BLQ22 calculates the roll target damping force of the variable damper 7 for reducing roll vibration based on the roll rate and roll angle output from the state estimation unit 12 and the roll GSP _r output from the GSP calculation unit 13. The roll vibration suppression BLQ22 is designed based on the vehicle model shown in FIG. 13, for example. In FIG. 13, the simplest one-degree-of-freedom rotational motion model is illustrated as a vehicle model taking roll into consideration. In an actual vehicle, two sets of this motion model are provided on the front wheel side and the rear wheel side, and only one of them is illustrated in FIG. 13. The roll vibration suppression BLQ22 is designed based on the state equation of the vehicle model in FIG. 13, as shown in the design method described later.

As shown in FIG. 10, the roll vibration suppression BLQ22 includes a Bp calculation unit 22A, a multiplier 22B, and a roll target damping force calculation unit 22C. The Bp calculation unit 22A calculates the maximum value of B _roll p _roll according to the roll rate based on the previously calculated B _roll p _roll and roll GSP _r . The multiplier 22B multiplies the maximum value of B _roll p _roll output from the Bp calculation unit 22A by the roll rate. The roll target damping force calculation unit 22C multiplies the calculated value by the multiplier 22B by the previously calculated inverse matrix (R _roll ^-1 ) of R _roll to calculate the roll target damping force. In this way, the roll vibration suppression BLQ22 calculates a value obtained by multiplying the right side of the formula 20 described later by the relative velocity from the left side. As a result, the roll vibration suppression BLQ22 calculates a roll target damping force obtained by multiplying the left side of the formula 20 described later by the roll rate from the left side.

The pitch vibration damping BLQ23 receives the vehicle state output from the state estimation unit 12 and the pitch GSP _p output from the GSP calculation unit 13. The pitch vibration damping BLQ23 calculates the pitch target damping force of the variable damper 7 for reducing pitch vibration based on the pitch rate and pitch angle output from the state estimation unit 12 and the pitch GSP _p output from the GSP calculation unit 13. The pitch vibration damping BLQ23 is designed based on the vehicle model shown in FIG. 14, for example. FIG. 14 illustrates a simplest one-degree-of-freedom rotational motion model as a vehicle model taking pitch into consideration. In an actual vehicle, two sets of this motion model are provided on the left and right sides, and only one of them is illustrated in FIG. 14. The pitch vibration damping BLQ23 is designed based on the state equation of the vehicle model in FIG. 14, as shown in the design method described later.

As shown in FIG. 11, the pitch vibration damping BLQ 23 includes a Bp calculation unit 23A, a multiplier 23B, and a pitch target damping force calculation unit 23C. The Bp calculation unit 23A calculates the maximum value of B _pitch p _pitch according to the pitch rate based on the previously calculated B _pitch p _pitch and pitch GSP _p . The multiplier 23B multiplies the maximum value of B _pitch p _pitch output from the Bp calculation unit 23A by the pitch rate. The pitch target damping force calculation unit 23C multiplies the calculated value by the multiplier 23B by the previously calculated inverse matrix (R _pitch ^-1 ) of R _pitch to calculate the pitch target damping force. In this way, the pitch vibration damping BLQ 23 calculates a value obtained by multiplying the right side of Equation 27 described later by the relative speed from the left side. As a result, the pitch vibration damping BLQ 23 calculates a pitch target damping force obtained by multiplying the left side of Equation 27 described later by the pitch rate from the left side.

The target damping force calculation unit 24 calculates an actual target damping force of each variable damper 7 based on three target damping forces (heave target damping force, roll target damping force, pitch target damping force) output from the sprung vertical vibration damping BLQ21, the roll vibration damping BLQ22, and the pitch vibration damping BLQ23 (see FIG. 5). The sprung vertical vibration damping BLQ21, the pitch vibration damping BLQ23, and the target damping force calculation unit 24 constitute an actual target damping force calculation unit that calculates an actual target damping force based on the heave GSP, roll GSP _r , and pitch GSP _p (weighting coefficients) by the GSP calculation unit 13 (weighting coefficient calculation unit) and the calculation results by the state estimation unit 12 (vertical movement calculation unit, pitch movement calculation unit).

As shown in FIG. 8, the target damping force calculation unit 24 includes an adder 24A and a damping force dead zone calculation unit 24B. The adder 24A adds the heave target damping force from the sprung vertical vibration damping BLQ21, the roll target damping force from the roll vibration damping BLQ22, and the pitch target damping force from the pitch vibration damping BLQ23, and calculates the sum of these to form a BLQ target damping force.

The damping force deadband calculation unit 24B sets a deadband for the BLQ target damping force output from the adder 24A. The damping force deadband calculation unit 24B outputs the actual target damping force that is set to the deadband for the BLQ target damping force.

Because bilinear optimal control (BLQ) operates even with very small road inputs, there is a risk that ride comfort will deteriorate if a high damping force is generated in response to a very small road input. Therefore, the damping force dead zone calculation unit 24B sets a dead zone for the BLQ target damping force. This prevents the damping force dead zone calculation unit 24B from operating with control for very small target damping forces.

If the control gain of the BLQ varies depending on the vehicle speed, the dead band selection process also sets the dead band according to the vehicle speed. If the control gain of the BLQ varies depending on the car mode, the dead band selection process also sets the dead band according to the car mode.

When driving on undulating roads, the target damping force becomes smaller when the sprung speed or relative displacement crosses zero, and if a damping force dead band is set at this time, the control amount is not output, resulting in a deterioration in vibration control. Therefore, in order to prioritize vibration control when driving on undulating roads, a mitigation process is performed for the provisional BLQ damping force dead band, which reduces the dead band when the undulating road index is high, so that a control amount is output even with a small target damping force. Also, in order to improve tuning, the dead band mitigation process may be made switchable between enabled and disabled.

Note that the target damping force calculation unit 24 adds the heave target damping force, roll target damping force, and pitch target damping force, but the present invention is not limited to this. The target damping force calculation unit 24 may select, for example, the largest value among the heave target damping force, roll target damping force, and pitch target damping force.

The damping coefficient upper limit calculation unit 25 calculates a maximum damping coefficient Cmax, which is an upper limit value of the damping coefficient C. As shown in FIG. 6, the damping coefficient upper limit calculation unit 25 is a maximum damping coefficient map that calculates the maximum damping coefficient Cmax based on the vehicle state. The damping coefficient upper limit calculation unit 25 includes, for example, a characteristic line 25A that shows the relationship between the relative speed x ^* between the sprung and unsprung parts and the maximum damping coefficient Cmax. The damping coefficient upper limit calculation unit 25 outputs the maximum damping coefficient Cmax based on the relative speed x ^* . At this time, the maximum damping coefficient Cmax is set to a value within a range that does not exceed the maximum value of the damping coefficient that can be generated by the variable damper 7. The maximum damping coefficient Cmax is set to a small value when the relative speed x ^* is slower than a predetermined threshold value Vt, and is set to a large value when the relative speed x ^* is faster than the threshold value Vt.

Specifically, when the relative speed x ^* is lower than the threshold value Vt (-Vt<x ^* <Vt), the maximum damping coefficient Cmax is set to a small low-speed setting value C1. On the other hand, when the relative speed x ^* on the extension side (positive side) is higher than the threshold value Vt (x ^* >Vt), the maximum damping coefficient Cmax is set to a high-speed setting value C2 that is higher than the low-speed setting value C1. Similarly, when the relative speed x ^* on the contraction side (negative side) is higher than the threshold value Vt (x ^* <-Vt), the maximum damping coefficient Cmax is set to a high-speed setting value C3 that is higher than the low-speed setting value C1.

When the relative velocity x ^* is close to the threshold value Vt, the maximum damping coefficient Cmax may be set to a value between the low speed setting value C1 and the high speed setting value C2. Similarly, when the relative velocity x ^* is close to the threshold value (-Vt), the maximum damping coefficient Cmax may be set to a value between the low speed setting value C1 and the high speed setting value C3.

The high speed setting values C2, C3 are appropriately set in consideration of the structure, specifications, damping force characteristics, etc. of the variable damper 7. In addition, although the low speed setting value C1 and the high speed setting values C2, C3 are all constant values in the above example, they may be configured to change according to the relative speed x ^* .

The threshold value Vt is obtained experimentally, for example, taking into consideration the occurrence state of jerk, and is set appropriately depending on the structure and damping force characteristics of the variable damper 7. The threshold value Vt may be set to the same value on the positive side and the negative side of the relative velocity x ^* , or may be set to different values on the positive side and the negative side of the relative velocity x ^* .

The corrected damping coefficient calculation unit 26 calculates a target damping coefficient C corresponding to the actual target damping force, and outputs a corrected damping coefficient Ca based on the target damping coefficient C and the maximum damping coefficient Cmax (see FIG. 8). The corrected damping coefficient calculation unit 26 includes a target damping coefficient calculation unit 27 and a minimum value selection unit 28.

The target damping coefficient calculation unit 27 constitutes a target damping coefficient calculation unit that calculates the target damping coefficient C based on the calculated value of the actual target damping force calculation unit 24. The target damping coefficient calculation unit 27 calculates the target damping coefficient C corresponding to the actual target damping force output from the target damping force calculation unit 24. Specifically, the target damping coefficient calculation unit 27 is constituted by a divider, and calculates the target damping coefficient C by dividing the actual target damping force by the relative speed x ^* .

The minimum value selection unit 28 compares the target damping coefficient C output from the target damping coefficient calculation unit 27 with the maximum damping coefficient Cmax output from the damping coefficient upper limit calculation unit 25, selects the smaller of these coefficients C and Cmax, and outputs it as the corrected damping coefficient Ca. Therefore, the minimum value selection unit 28 and the damping coefficient upper limit calculation unit 25 constitute a correction unit that calculates the corrected damping coefficient Ca by lowering the upper limit of the target damping coefficient C in the region where the relative speed x ^* between the sprung and unsprung parts of the vehicle is low.

At this time, the correction unit sets the corrected damping coefficient Ca so that the damping force increases with an increase in the relative speed x ^* , the slope of the damping force with respect to the relative speed x ^* is small when the relative speed x ^* is low, and the slope of the damping force with respect to the relative speed x ^* is large when the relative speed x ^* is high. In addition, the correction unit has a maximum damping coefficient Cmax according to the relative speed x ^* , and corrects the target damping coefficient C to the maximum damping coefficient Cmax when the target damping coefficient C exceeds the maximum damping coefficient Cmax.

The command value calculation unit 29 constitutes a control signal output unit that outputs a control signal corresponding to the corrected damping coefficient Ca to the variable damper 7 (damping force adjustable shock absorber). The command value calculation unit 29 outputs a command current value I as a control signal corresponding to the corrected damping coefficient Ca. As shown in FIG. 7, the command value calculation unit 29 is a damping coefficient map that variably sets the relationship between the corrected damping coefficient Ca and the command current value I according to the relative speed x ^* . This damping coefficient map was created based on test data by the inventors. Then, the command value calculation unit 29 specifies a command current value I for adjusting the damping force characteristics of the variable damper 7 based on the corrected damping coefficient Ca and the relative speed x ^* , and outputs this command current value I to the damping force variable actuator 8 of the variable damper 7.

The command value calculation unit 29 also outputs a control signal (command current value I) for controlling the variable damper 7 so as to conform the damping force adjustable shock absorber to the skyhook theory, for example. This command value calculation unit 29 has a hard-side characteristic line 29A shown by a solid line in FIG. 7, and a soft-side characteristic line 29B shown by a dashed line in FIG. 7. At this time, the hard-side characteristic line 29A is positioned in a range where the corrected damping coefficient Ca is larger than that of the soft-side characteristic line 29B.

When the relative velocity x ^* and the corrected damping coefficient Ca are input, an intersection between the corrected damping coefficient Ca and the relative velocity x ^* is found in the damping coefficient map of the command value calculation unit 29. When this intersection is located in a range where the corrected damping coefficient Ca is larger than the hard side characteristic line 29A, the command current value I is increased to set the damping force characteristics to hard characteristics. On the other hand, when the intersection is located in a range where the corrected damping coefficient Ca is smaller than the soft side characteristic line 29B, the command current value I is decreased to set the damping force characteristics to soft characteristics. Furthermore, when the intersection is located in a range between the hard side characteristic line 29A and the soft side characteristic line 29B, the command current value I is adjusted according to the corrected damping coefficient Ca, and the damping force characteristics are set to intermediate characteristics between hard and soft.

As a result, the damping force generated by the variable damper 7 is variably adjusted continuously or in multiple stages between hard and soft according to the command current value I supplied to the damping force variable actuator 8.

Next, we will explain how to design sprung vertical vibration damping BLQ21, roll vibration damping BLQ22, and pitch vibration damping BLQ23, and how to tune them using gain scheduling parameters (GSP).

(1) Design of the vertical movement control system (sprung vertical vibration suppression BLQ21) A control design model (1/4 vehicle model) targeting vertical movement is shown in Figure 12. Here, the absolute vertical displacement of the vehicle body 2 (sprung mass) is _zb , the absolute vertical displacement of the unsprung mass is _zt , the absolute vertical displacement of the road surface is _z0 , the sprung mass (vehicle mass) is _mb , the unsprung mass _mt , the spring constant between the vehicle body 2 and the unsprung mass (sprung-unsprung) is _ks , the tire spring constant is _kt , the damper damping coefficient is c, and the control force acting between the sprung mass and unsprung mass is f. Also, considering that the control force f acts in addition to the soft damping force characteristics, the damper damping coefficient c is set to the damping coefficient of the soft damping force characteristics.

In this case, if the relative displacement between the sprung and unsprung parts is z _bt and the relative displacement between the unsprung part and the road surface is z _t0 , the relative displacements z _bt , z _t0 and the state equations are as shown in the following formulas 1 to 3. In the state equations, the state variable x and the output y are as shown in formulas 4 and 5. A dot in the equations means a first-order differential with respect to time t (d/dt). Two dots mean a second-order differential (d2/dt2).

The external force u acting between the vehicle body 2 and the unsprung mass is the control force f (u=f), and the disturbance w is the detected road surface displacement _z0 (w=z0). The elements of the state equation are as shown in Equation 6.

Based on the above-mentioned state equation, a method for designing a vertical motion control system (sprung vertical vibration suppression BLQ21) based on bilinear optimal control theory will be described. In the above-mentioned state equation, a control force f was used, but in the case of a bilinear system, the input is treated as a damping coefficient u _c . Therefore, the formula 3 is described by the following state equation.

Here, x ^* is a matrix including the state variable x, so when the control force f is expressed in terms of the damping coefficient input u _c and the relative velocity, it becomes as in Equation 8. At this time, x ^* is as shown in Equation 9.

Generally, when the input is irregular, the state variables become random variables, so the evaluation function uses the formula (10) expressed as an expected value. However, the state quantities (sprung-unsprung relative displacement, sprung velocity, unsprung-road relative displacement, and unsprung velocity) are used in the evaluation function. Also, C ^T QC is a weighting matrix for the output, and R is a weighting matrix for the input.

Here, an optimal control input u _c ⁰ that satisfies Expression 11 is obtained. If the above problem is solved by dynamic programming and the optimal control input u _c ⁰ is considered to be a stationary problem, the optimal control input u _c ⁰ can be obtained as shown in Expression 12.

However, p in equation (12) is the unique positive definite solution to equation (13).

Here, the values of Q and R of the sprung upper and lower vibration damping BLQ21 were set to the values shown in formula 14, and p was calculated.

(2) Design of the roll motion control system (roll vibration suppression BLQ22) In the same manner as the above-mentioned vertical motion control system, the roll vibration suppression BLQ22 is designed based on the bilinear control theory. The motion model is a one-degree-of-freedom rotational motion model, which is the simplest vehicle model considering roll, as shown in FIG.

Here, the roll angle of the vehicle body is θ, the absolute vertical displacements of the left and right wheels on the road surface are x _0l and x _0r , respectively, the vehicle body roll inertia is I, the spring constant between the sprung and unsprung masses is k _s , the spring constant of the stabilizer is k _st , the damper damping coefficient is c, and the external forces acting between the sprung and unsprung masses at the left and right wheels are F _r and F _l . From Fig. 13, the equation of motion in the roll direction is derived as shown in equation 15.

Because this is a bilinear system, the state equation is as shown in Equation 16.

Here, because it is a bilinear system, the external force can be expressed by the damping coefficient inputs c _r and c _l of the left and right wheels and the relative speed as in Eq. (17).

Moreover, the state variables x _roll , x ^* _roll , the control force u _croll , and the disturbance w _roll are expressed by the following equation (18).

At this time, each element of the state equation is as shown in Equation 19.

From this state equation, the optimum control input u ⁰ _croll is obtained as in the case of the vertical vibration control system, as in Equation 20.

Here, p _roll in equation (20) is a unique positive definite solution to equation (21).

Here, the values of Q _roll and R _roll of the roll vibration suppression BLQ22 are set to those shown in the formula 22, and p _roll is calculated.

(3) Design of pitch motion control system (pitch vibration suppression BLQ23) In the same manner as the above-mentioned vertical motion and roll motion control systems, the pitch vibration suppression BLQ23 is designed based on bilinear control theory. The motion model is the simplest one-degree-of-freedom rotational motion model as a vehicle model considering pitch, as shown in Figure 14.

Here, the vertical forces acting on the vehicle body from the front and rear suspensions are F _Fr and F _Rr , the front and rear spring constants are k _Fr and k _Rr , the front and rear reduction coefficients are c _Fr and c _Rr , the horizontal distances from the center of gravity to the front and rear suspensions are L _Fr and L _Rr , the pitch angle is Φ, the front and rear vehicle height displacements are x _{2 Fr} and x _{2 Rr} , the front and rear road surface heights are x _{0 Fr} and x _{0 Rr} , the vertical forces generated by the front and rear actuators are F _FrAct and F _RrAct , and the pitch direction sprung inertia moment around the 1/2 center of gravity is I _yy . From FIG. 14 , the equation of motion in the roll direction is derived as shown in Equation 23.

Here, the state equation for the bilinear system is as shown in Equation 24.

Moreover, the state variables x _pitch , x ^* _pitch , the control force u _cpitch , and the disturbance w _pitch are expressed by the following equation 25.

At this time, each element of the state equation is as shown in Equation 26.

From this state equation, the optimum control input u ⁰ _cpitch can be calculated as in the case of the vertical vibration control system, as given by equation (27).

However, p _pitch in equation (27) is a unique positive definite solution to equation (28).

Here, the Q _pitch and R _pitch of the pitch vibration damping BLQ23 were set to the values shown in Equation 29, and p _pitch was calculated.

(4) Tuning Using Gain Scheduling Parameters The optimal control inputs to the control systems for vertical vibration, roll vibration, and pitch vibration are expressed by equations 11, 20, and 27. At this time, the theoretical maximum values (maximum control amount) and minimum values (minimum control amount) of B ^T p, B _roll ^T p _roll , and B _pitch ^T p _pitch are found in advance by calculation. For example, an example of the maximum and minimum values of B _roll ^T p _roll and B _pitch ^T p _pitch is shown below. B ^T p can also be found in advance by calculation in the same way.

A value between the maximum value and the longest value is linearly interpolated using the parameters (heave GSP, roll GSP _r , pitch GSP _p ) to select an arbitrary control amount. Based on the control amount calculated using the heave GSP, roll GSP _r , and pitch GSP _p and equations 11, 20, and 27, the target damping forces for vertical vibration suppression, roll vibration suppression, and pitch vibration suppression can be calculated.

When tuning the BLQ without using gain scheduling parameters, it is necessary to change Q, R, Q _roll , R _roll , Q _pitch , and R _pitch and re-solve the Riccati equation each time. In contrast, when tuning the BLQ using gain scheduling parameters (heave GSP, roll GSP _r , pitch GSP _p ), it is not necessary to re-solve the Riccati equation, and tuning can be performed with a low man-hour.

The vehicle behavior control device 1 according to the first embodiment has the configuration described above. Next, we will explain the process of variably controlling the damping force characteristics of the variable damper 7 using the controller 11.

When the vehicle is traveling, vehicle travel-related specifications are input to the controller 11 via the CAN 9. Based on the CAN signals, the state estimation unit 12 of the controller 11 estimates the vehicle state, including the sprung-unsprung relative velocity, vehicle body absolute vertical velocity (sprung velocity), unsprung-road relative velocity, unsprung absolute vertical velocity (unsprung velocity), roll rate, roll angle, pitch rate, and pitch angle.

At this time, the vehicle state calculated by the state estimation unit 12 and the heave GSP, roll GSP _r and pitch GSP _p calculated by the GSP calculation unit 13 are input to the BLQ controller 20 of the controller 11. The sprung vertical vibration damping BLQ 21 of the BLQ controller 20 calculates a heave target damping force based on the vehicle state and the heave GSP. The roll vibration damping BLQ 22 of the BLQ controller 20 calculates a roll target damping force based on the vehicle state and the roll GSP _r . The pitch vibration damping BLQ 23 of the BLQ controller 20 calculates a heave target damping force based on the vehicle state and the pitch GSP _p .

A target damping force calculation unit 24 of the BLQ controller 20 calculates an actual target damping force based on the heave target damping force, the roll target damping force, and the pitch target damping force. A target damping coefficient calculation unit 27 of the BLQ controller 20 calculates a target damping coefficient C by dividing the actual target damping force by the relative speed x ^* . A minimum value selection unit 28 of the BLQ controller 20 selects the smaller of the target damping coefficient C output from the target damping coefficient calculation unit 27 and the maximum damping coefficient Cmax output from the damping coefficient upper limit calculation unit 25, and outputs it as a corrected damping coefficient Ca. A command value calculation unit 29 of the BLQ controller 20 calculates a command current value I according to the corrected damping coefficient Ca and the relative speed x ^* .

The command current value I is then input to the variable damping force actuator 8 of the variable damper 7, controlling the drive of the variable damping force actuator 8. As a result, the damping force characteristics of the variable damper 7 are variable between hard characteristics and soft characteristics and are continuously controlled.

In particular, the controller 11 of the first embodiment is equipped with pitch vibration damping BLQ23, and calculates the command current value I based on the pitch target damping force calculated by the pitch vibration damping BLQ23. Therefore, pitch vibration can be suppressed more effectively than when the pitch vibration damping BLQ23 is omitted.

In order to confirm the effect of improving pitch vibration by this embodiment, an actual vehicle test was conducted on the first embodiment and the comparative example when the vehicle was traveling on an undulating road. The results are shown in Fig. 15. In the actual vehicle test, the vehicle speed was 100 km/h, the heave GSP was 0.6, and the pitch GSP _p was 0.3.

In FIG. 15, the solid line indicates the characteristics according to the first embodiment. However, when the controller 11 calculates the roll target damping force, the gain of this roll target damping force is set to 0. That is, the solid line in FIG. 15 indicates the characteristics when the roll vibration suppression BLQ22 is omitted from the controller 11 of the first embodiment. On the other hand, in FIG. 15, the dashed line indicates the characteristics according to the comparative example (prior art). At this time, the comparative example corresponds to the case where the controller according to Patent Document 1 is used. That is, in the comparative example, the variable damper 7 is controlled based on the target damping force that suppresses only the sprung vertical vibration. As shown in FIG. 15, in the first embodiment, the amplitude of the pitch angle is reduced compared to the comparative example, and it can be seen that, for example, the pitch vibration is improved (reduced) by about 7.3%.

Thus, the controller 11 as a control device according to this embodiment includes a weighting coefficient calculation unit (GSP calculation unit 13) that calculates weighting coefficients (heave GSP, pitch GSP _p ) for making an actual target damping force based on a calculated value by a vehicle behavior calculation unit (state estimation unit 12) approach a value close to a previously determined target damping force, a vertical motion calculation unit (state estimation unit 12) that calculates a state related to the vertical motion of the vehicle, a pitch motion calculation unit (state estimation unit 12) that calculates a state related to the pitch motion of the vehicle, weighting coefficients by the weighting coefficient calculation unit, the vertical motion calculation unit, and an actual target damping force calculation unit (sprung vertical vibration damping BLQ21, pitch vibration damping BLQ23, target damping force calculation unit 24) that calculates the actual target damping force based on the calculation results of the pitch motion calculation unit, a target damping coefficient calculation unit (target damping coefficient calculation unit 27) that calculates a target damping coefficient based on a calculated value by the actual target damping force calculation unit, and a relative velocity x between the sprung and unsprung parts of the vehicle. ^* has a correction unit (minimum value selection unit 28) that calculates a corrected damping coefficient Ca that lowers the upper limit of the target damping coefficient C in a low-speed region, and a control signal output unit (command value calculation unit 29) that outputs a control signal corresponding to the corrected damping coefficient Ca to a damping force adjustable shock absorber (variable damper 7).

At this time, the weighting coefficient calculation unit (GSP calculation unit 13) calculates, as weighting coefficients, a heave GSP related to the vertical movement of the vehicle and a pitch GSP _p related to the movement of the vehicle in the pitch direction, and the actual target damping force calculation unit includes a heave BLQ (sprung vertical vibration damping BLQ21) that calculates a target damping force in the vertical direction of the vehicle based on _the heave GSP and the calculation result of the vertical movement calculation unit, and a pitch BLQ (pitch vibration damping BLQ23) that calculates a target damping force in the pitch direction of the vehicle based on the pitch GSP p and the calculation result of the pitch movement calculation unit, and calculates the actual target damping force based on the target damping force calculated by the heave BLQ (heave target damping force) and the target damping force calculated by the pitch BLQ (pitch target damping force).

Therefore, the controller 11 can control the damping force of the variable damper 7 based on the heave target damping force for suppressing the vertical vibration of the sprung mass and the pitch target damping force for suppressing the pitch vibration. As a result, both the vertical vibration of the sprung mass and the pitch vibration can be reduced. In addition, the controller 11 calculates the actual target damping force based on the weighting coefficients (heave GSP, pitch GSP _p ) and the vehicle state (sprung-unsprung relative velocity, vehicle body absolute vertical velocity (sprung velocity), unsprung-road relative velocity, unsprung absolute vertical velocity (unsprung velocity), pitch rate, pitch angle). Therefore, it is not necessary to solve the Riccati equation again, and by adjusting the weighting coefficients (heave GSP, pitch GSP _p ), the tuning parameters can be reduced and the controller 11 (sprung vertical vibration damping BLQ21, pitch vibration damping BLQ23) can be easily tuned.

Further, the correction section (minimum value selection section 28) of the controller 11 sets the corrected damping coefficient Ca so that the damping force increases as the relative velocity x ^* increases, the slope of the damping force with respect to the relative velocity x ^* is small when the relative velocity x ^* is low, and the slope of the damping force with respect to ^the relative velocity is large when the relative velocity x* is high, and the correction section has a maximum damping coefficient Cmax according to the relative velocity x ^* , and when the target damping coefficient C exceeds the maximum damping coefficient Cmax, outputs the corrected damping coefficient Ca obtained by correcting the target damping coefficient C to the maximum damping coefficient Cmax.

For this reason, when the relative speed x ^* is low, such as when the variable damper 7 reverses between the extension stroke and the retraction stroke, the controller 11 outputs a corrected damping coefficient Ca that lowers the upper limit of the target damping coefficient C, and outputs a command current value I corresponding to this corrected damping coefficient Ca to the variable damper 7. This makes it possible to reduce the occurrence of abnormal noise and jerks caused by a sudden change in the damping force.

On the other hand, when the relative velocity x ^* is high, the controller 11 calculates a corrected damping coefficient Ca with a higher upper limit of the target damping coefficient C compared to when the relative velocity x* is low. In this case, when the relative velocity x ^* is high, such as during the extension stroke or retraction stroke of the variable damper 7, it is possible to calculate a large value of the corrected damping coefficient Ca without restricting the target damping coefficient C as much as possible. As a result, when the relative velocity x ^* is high, a large damping force can be generated by the variable damper 7, vibration control performance can be ensured, and ride comfort can be improved.

The controller 11 further includes a posture change detection unit (state estimation unit 12) that detects posture changes of the vehicle body 2, and the correction unit (minimum value selection unit 28) reduces the amount of correction when it determines that a posture change will occur based on the detection result of the posture change detection unit.

At this time, when there is no change in the posture of the vehicle body 2, the correction unit increases the correction amount and lowers the upper limit of the target damping coefficient when the relative speed is low. This makes it possible to suppress sudden changes in the damping force. On the other hand, when there is a change in the posture of the vehicle body, the correction unit decreases the correction amount and relaxes the upper limit of the target damping coefficient. This makes it possible to generate a damping force that resists the change in the posture of the vehicle body, and ensures vibration control performance.

The controller 11 further includes a roll motion calculation unit (state estimation unit 12) that determines a state related to the motion of the vehicle in the roll direction. In this case, the weighting coefficient calculation unit (GSP calculation unit 13) calculates, as the weighting coefficients, a heave GSP related to the motion of the vehicle in the vertical direction, a roll GSP _r related to the motion of the vehicle in the roll direction, and a pitch GSP _p related to the motion of the vehicle in the pitch direction, and the actual target damping force calculation unit calculates a heave BLQ (sprung vertical damping BLQ21) that calculates a target damping force in the vertical direction of the vehicle based on the heave GSP and the calculation result of the vertical motion calculation unit, a roll BLQ (roll damping BLQ22) that calculates a target damping force in the roll direction of the vehicle based on the roll GSP _r and the calculation result of the roll motion calculation unit, and a pitch GSP p related to the motion of the vehicle in the pitch direction. and a pitch BLQ (pitch vibration suppression BLQ23) that calculates a target damping force in the pitch direction of the vehicle based on _p and the calculation result of the pitch motion calculation unit, and calculates the actual target damping force based on the target damping force (heave target damping force) calculated by the heave BLQ, the target damping force (roll target damping force) calculated by the roll BLQ, and the target damping force (pitch target damping force) calculated by the pitch BLQ.

As a result, the actual target damping force calculation unit (sprung vertical vibration damping BLQ21, roll vibration damping BLQ22, pitch vibration damping BLQ23, target damping force calculation unit 24) can calculate the actual target damping force based on the calculation results of the roll motion calculation unit in addition to the calculation results of the vertical motion calculation unit and pitch motion calculation unit. In other words, the controller 11 can control the damping force of the variable damper 7 based on the roll target damping force for suppressing roll vibration in addition to the heave target damping force for suppressing vertical vibration in the sprung mass and the pitch target damping force for suppressing pitch vibration. As a result, in addition to the vertical vibration and pitch vibration in the sprung mass, roll vibration can also be reduced.

In the first embodiment, the controller 11 calculates the actual target damping force based on the weighting coefficients (heave GSP, roll GSP _r , pitch GSP _p ) and the vehicle state (sprung-unsprung relative velocity, vehicle body absolute vertical velocity (sprung velocity), unsprung-road relative velocity, unsprung absolute vertical velocity (unsprung velocity), roll rate, roll angle, pitch rate, pitch angle). Therefore, it is not necessary to solve the Riccati equation again, and by adjusting the weighting coefficients (heave GSP, roll GSP _r , pitch GSP _p ), the tuning parameters can be reduced and the controller 11 (sprung vertical vibration damping BLQ21, roll vibration damping BLQ22, pitch vibration damping BLQ23) can be easily tuned.

Next, Figs. 1, 2, 16 and 17 show a second embodiment. The second embodiment is characterized in that the weighting coefficient calculation unit changes the coefficients according to the road surface condition, the vehicle weight and the selected damping force mode. In the second embodiment, the same components as those in the first embodiment described above are given the same reference numerals and their explanations will be omitted.

The controller 31 according to the second embodiment constitutes a control device for the variable damper 7 and controls the damping characteristics of the variable damper 7 (see Figures 1 and 2). The controller 31 is configured in a manner similar to the controller 11 according to the first embodiment. The controller 31 is configured, for example, by a microcomputer and is connected to the CAN 9. The controller 31 acquires specifications related to the running of the vehicle through the CAN 9. The output side of the controller 31 is connected to the damping force variable actuator 8 of the variable damper 7.

The controller 31 also has a storage unit 31A consisting of a ROM, a RAM, a non-volatile memory, etc. The storage unit 31A of the controller 31 stores various programs, information (vehicle information), data, etc. for controlling the variable damper 7.

As shown in FIG. 16, the controller 31 includes a state estimation unit 12, a GSP calculation unit 34, and a BLQ controller 20. In addition, the controller 31 includes a road surface index calculation unit 32 and a weight estimation unit 33.

The road surface index calculation unit 32 obtains, for example, the sprung velocity etc. from the state estimation unit 12, and uses these to output a road surface index. The road surface index calculation unit 32 estimates the vertical acceleration of the vehicle body 2 (sprung mass) from the sprung velocity etc. The road surface index calculation unit 32 is equipped with, for example, a band pass filter, and extracts undulating road components of a predetermined frequency band (for example, 0.5 to 2 Hz) from the sprung mass vertical acceleration. The road surface index calculation unit 32 outputs a road surface index (undulating road index) corresponding to the magnitude of the undulating road component (undulating road level) to the GSP calculation unit 34. At this time, if the undulating road component is not extracted and is at its minimum value, the road surface index corresponding to the undulating road is 0. On the other hand, if the undulating road component is at its maximum value, the road surface index corresponding to the undulating road is 1. In this way, the road surface index is a normalized value between 0 and 1. That is, the road surface index is an index of the up-down, pitch, and roll movements (0.5 to 2 Hz) of the vehicle body 2 caused by road surface input, expressed as a value between 0 and 1.

The road surface index calculation unit 32 is not limited to acquiring the sprung velocity and the like from the state estimation unit 12, but may acquire detection values from a vertical acceleration sensor to calculate the road surface index. The road surface index calculation unit 32 may also acquire road surface information in the traveling direction of the vehicle from a camera or the like to calculate the road surface index.

The weight estimation unit 33 acquires the absolute vertical displacement of the vehicle body 2 based on, for example, a CAN signal. The weight estimation unit 33 estimates the weight of the vehicle body 2 based on, for example, the absolute vertical displacement of the vehicle body 2, and outputs the estimated weight to the GPS calculation unit 34. Note that the weight estimation unit 33 is not limited to estimating the weight of the vehicle body 2 from, for example, the absolute vertical displacement of the vehicle body 2, and may estimate the weight of the vehicle body 2 based on, for example, a signal from a stroke sensor that detects the stroke of the variable damper 7. The weight estimation unit 33 may obtain the weight of the vehicle body 2 by acquiring a detection value from a weight sensor that detects the weight of the vehicle body.

The GSP calculation unit 34 (gain scheduling parameter calculation unit) constitutes a weighting coefficient calculation unit that calculates a weighting coefficient that brings the actual target damping force based on the calculation value of the state estimation unit 12 (vehicle behavior calculation unit) closer to a value close to a target damping force obtained in advance. For example, the vehicle speed acquired from the CAN 9, the road surface index from the road surface index calculation unit 32, and the estimated weight from the weight estimation unit 33 are input to the GSP calculation unit 34. In addition, a car mode as a damping force mode is input to the GSP calculation unit 34. In this case, the car mode includes three types of modes, for example, a normal mode (Normal), a sports mode (Sport), and a comfort mode (Comfort). The car mode is selected, for example, by a mode selection switch (not shown) provided on the vehicle. The GSP calculation unit 34 calculates the heave GSP, roll GSP _r , and pitch GSP _p based on the vehicle speed, the road surface index, the weight estimation unit, and the car mode. The car mode is not limited to three types of modes, but may have two types of modes, or may have four or more types of modes.

That is, when the BLQ controller 20 calculates an actual target damping force based on the vehicle state (e.g., sprung-unsprung relative velocity, vehicle body absolute vertical velocity, unsprung-road surface relative velocity, unsprung absolute vertical velocity, roll rate, roll angle, pitch rate, pitch angle, etc.) which is a calculated value of the state estimation unit 12 (vehicle behavior calculation unit), the GSP calculation unit 34 calculates weighting coefficients (heave GSP, roll GSP _r , pitch GSP _p ) that bring this actual target damping force closer to a value close to a pre-calculated target damping force. The GSP calculation unit 34 outputs the heave GSP, roll GSP _r , and pitch GSP _p to the BLQ controller 20.

As shown in FIG. 17, the GSP calculation unit 34 includes a lookup table that records, for example, the relationship between the vehicle speed, road surface index, estimated weight, car mode, and the heave GSP, roll GSP _r , and pitch GSP _p . At this time, the heave GSP, roll GSP _r , and pitch GSP _p are determined by tuning in an actual vehicle test. Specifically, the heave GSP is finely adjusted by referring to vibration data and the like after the rough characteristics are determined by sensory evaluation in an actual vehicle test. In addition, the GSP calculation unit 34 includes a lookup table that records the relationship between the vehicle speed, road surface index, estimated weight, car mode, and roll GSP, and a lookup table that records the relationship between the vehicle speed, road surface index, estimated weight, car mode, and pitch GSP. The roll GSP _r and pitch GSP _p are also determined by tuning in an actual vehicle test, like the heave GSP.

As shown in FIG. 17, for example, the GSP calculation unit 34 calculates a small value of the heave GSP at low speeds and a large value of the heave GSP at high speeds. When the GSP calculation unit 34 judges the road to be undulating based on the road surface index, the GSP calculation unit 34 increases the value of the heave GSP. When the estimated weight is large, the GSP calculation unit 34 increases the value of the heave GSP compared to when the estimated weight is small. When the car mode is the sports mode, the GSP calculation unit 34 increases the value of the heave GSP compared to when the car mode is the normal mode. When the car mode is the comfort mode, the GSP calculation unit 34 decreases the value of the heave GSP compared to when the car mode is the normal mode. The GSP calculation unit 34 calculates the roll GSP _r and pitch GSP _p according to the vehicle speed, road surface index, estimated weight, and car mode, similar to the heave GSP.

17 shows an example in which the GSP calculation unit 34 is provided with three types of lookup tables corresponding to the road surface index, the estimated weight, and the car mode. The GSP calculation unit 34 calculates the heave GSP, the roll GSP _r , and the pitch GSP _p by taking these three types of lookup tables into consideration. The present invention is not limited to this, and the GSP calculation unit 34 may calculate the heave GSP, the roll GSP _r , and the pitch GSP _p based on a single lookup table that combines the three types of lookup tables in FIG. 17.

Thus, in the second embodiment, it is possible to obtain substantially the same effect as in the first embodiment. In the second embodiment, the weighting coefficient calculation unit (GSP calculation unit 34) changes the coefficient according to the road surface condition. Therefore, the actual target damping force can be changed according to the road surface condition, thereby improving ride comfort and vehicle stability. The weighting coefficient calculation unit (GSP calculation unit 34) changes the coefficient according to the weight of the vehicle body 2. Therefore, the actual target damping force can be changed according to the weight of the vehicle body 2, thereby improving ride comfort and vehicle stability. The weighting coefficient calculation unit (GSP calculation unit 34) changes the coefficient according to the selected damping force mode (car mode). Therefore, the actual target damping force can be changed according to the selected car mode, thereby improving ride comfort and vehicle stability.

In the second embodiment, the GSP calculation unit 34 calculates the weighting coefficients (heave GSP, roll GSP _r , pitch GSP _p ) based on the road surface index, the estimated weight, and the car mode, but the present invention is not limited to this. The GSP calculation unit 34 may calculate the weighting coefficients (heave GSP, roll GSP _r , pitch GSP _p ) based on any one or two of the road surface index, the estimated weight, and the car mode.

In the second embodiment, the road surface index is an undulating road index, but the present invention is not limited to this. The road surface index may be an index corresponding to the condition of a road surface other than an undulating road, such as a rough road.

Next, Figs. 1, 2, 18 to 22 show a third embodiment. The third embodiment is characterized in that the weighting coefficient calculation unit changes the coefficients according to the pre-determined front-rear weight balance of the vehicle body and the weight of the vehicle body. In the third embodiment, the same components as those in the first and second embodiments described above are given the same reference numerals, and their explanation will be omitted.

The controller 41 according to the third embodiment constitutes a control device for the variable damper 7 and controls the damping characteristics of the variable damper 7 (see Figures 1 and 2). The controller 41 is configured in a manner similar to the controller 11 according to the first embodiment. The controller 41 is configured, for example, by a microcomputer and is connected to the CAN 9. The controller 41 acquires specifications related to the running of the vehicle through the CAN 9. The output side of the controller 41 is connected to the variable damping force actuator 8 of the variable damper 7.

The controller 41 also has a storage unit 41A consisting of a ROM, a RAM, a non-volatile memory, etc. The storage unit 41A of the controller 41 stores various programs, information (vehicle information), data, etc. for controlling the variable damper 7.

As shown in FIG. 18, the controller 41 includes a state estimation unit 12, a GSP calculation unit 34, and a BLQ controller 42. In addition, the controller 31 includes a road surface index calculation unit 32 and a weight estimation unit 33.

The BLQ controller 42 calculates a command current (control signal) that becomes a command value based on bilinear optimal control theory. The BLQ controller 42 is configured in the same manner as the BLQ controller 20 according to the first embodiment. In this case, as shown in FIG. 19, the BLQ controller 42 includes a sprung vertical vibration damping BLQ43, a roll vibration damping BLQ44, a pitch vibration damping BLQ45, a target damping force calculation unit 24, a damping coefficient upper limit calculation unit 25, a corrected damping coefficient calculation unit 26, and a command value calculation unit 29.

The sprung vertical vibration damping BLQ43 is configured in the same manner as the sprung vertical vibration damping BLQ21 according to the first embodiment. The vehicle state output from the state estimation unit 12, the heave GSP output from the GSP calculation unit 34, and the estimated weight output from the weight estimation unit 33 are input to the sprung vertical vibration damping BLQ43. The sprung vertical vibration damping BLQ43 calculates the heave target damping force of the variable damper 7 for reducing vertical vibration, separately for the front and rear wheels, based on the sprung-unsprung relative velocity, vehicle body absolute vertical velocity, unsprung-road surface relative velocity, and unsprung absolute vertical velocity output from the state estimation unit 12, the heave GSP output from the GSP calculation unit 34, and the estimated weight output from the weight estimation unit 33.

As shown in FIG. 20, the sprung vertical vibration damping BLQ 43 includes a correction gain calculation unit 43A, a Bp calculation unit 43B, a multiplier 43C, a Bp calculation unit 43D, a multiplier 43E, a multiplier 43F, and a heave target damping force calculation unit 43G. The correction gain calculation unit 43A calculates a correction gain for the front wheels and a correction gain for the rear wheels based on the estimated weight. The Bp calculation unit 43B calculates the maximum value of Bp according to the relative speed based on the previously obtained Bp and heave GSP. The multiplier 43C multiplies the correction gain output from the correction gain calculation unit 43A, the maximum value of Bp output from the Bp calculation unit 43B, and the relative speed. The Bp calculation unit 43D calculates the maximum value of Bp according to the sprung speed based on the previously obtained Bp and heave GSP. The multiplier 43E multiplies the correction gain output from the correction gain calculation unit 43A, the maximum value of Bp output from the Bp calculation unit 43D, and the sprung speed. The multiplier 43F multiplies the value calculated by the multiplier 43C and the value calculated by the multiplier 43E. The heave target damping force calculation unit 43G multiplies the value calculated by the multiplier 43F and the inverse matrix (R ^-1 ) of R obtained in advance to calculate the heave target damping force for the front wheels and the heave target damping force for the rear wheels. Note that, although the sprung vertical damping BLQ 43 has exemplified a case in which the heave target damping force is calculated based on the Bp according to the relative speed and the Bp according to the sprung speed, the present invention is not limited to this. The sprung vertical damping BLQ 43 may calculate the heave target damping force based on the Bp according to other vehicle conditions in addition to the Bp according to the relative speed and the sprung speed.

The roll vibration suppression BLQ 44 receives as input the vehicle state output from the state estimation unit 12, the roll GSP _r output from the GSP calculation unit 34, and the estimated weight output from the weight estimation unit 33. The roll vibration suppression BLQ 22 calculates the roll target damping force of the variable damper 7 for reducing roll vibration separately for the front wheels and the rear wheels, based on the roll rate and roll angle output from the state estimation unit 12, the roll GSP output from the GSP calculation unit 34, and the estimated weight output from the weight estimation unit 33.

As shown in Fig. 21, the roll vibration suppression BLQ 44 includes a correction gain calculation unit 44A, a Bp calculation unit 44B, multipliers 44C and 44D, and a roll target damping force calculation unit 44E. The correction gain calculation unit 44A calculates a correction gain for the front wheels and a correction gain for the rear wheels based on the estimated weight. The Bp calculation unit 44B calculates a maximum value of B _roll p _roll according to the roll rate based on the previously calculated B _roll p _roll and roll GSP _r . The multiplier 44C multiplies the correction gain output from the correction gain calculation unit 44A by the maximum value of B _roll p _roll output from the Bp calculation unit 44B. The multiplier 44D multiplies the value calculated by the multiplier 44C by the roll rate. The target roll damping force calculation section 44E multiplies the value calculated by the multiplier 44D by the inverse matrix (R _roll ^-1 ) of R _roll determined in advance, to calculate a target roll damping force for the front wheels and a target roll damping force for the rear wheels.

The pitch vibration suppression BLQ 45 receives as input the vehicle state output from the state estimation unit 12, the pitch GSP _p output from the GSP calculation unit 34, and the estimated weight output from the weight estimation unit 33. The pitch vibration suppression BLQ 45 calculates the pitch target damping force of the variable damper 7 for reducing pitch vibration separately for the front wheels and the rear wheels, based on the pitch rate and pitch angle output from the state estimation unit 12, and the pitch GSP _p output from the GSP calculation unit 34.

As shown in Fig. 22, the pitch vibration suppression BLQ 45 includes a correction gain calculation unit 45A, a Bp calculation unit 45B, multipliers 45C and 45D, and a pitch target damping force calculation unit 45E. The correction gain calculation unit 45A calculates a correction gain for the front wheels and a correction gain for the rear wheels based on the estimated weight. The Bp calculation unit 45B calculates a maximum value of B _pitch p _pitch according to the pitch rate based on the previously calculated B _pitch p _pitch and pitch GSP _p . The multiplier 45C multiplies the correction gain output from the correction gain calculation unit 45A by the maximum value of B _pitch p _pitch output from the Bp calculation unit 45B. The multiplier 45D multiplies the value calculated by the multiplier 45C by the pitch rate. The pitch target damping force calculation unit 45E multiplies the value calculated by the multiplier 45D by the inverse matrix (R _pitch ^-1 ) of R _pitch obtained in advance to calculate a pitch target damping force for the front wheels and a pitch target damping force for the rear wheels.

At this time, the GSP calculation unit 34 and the correction gain calculation units 43A, 44A, and 45A constitute a weighting coefficient calculation unit.

Thus, the third embodiment can achieve substantially the same effects as the first and second embodiments. In the third embodiment, the weighting coefficient calculation unit (GSP calculation unit 34, correction gain calculation unit 43A, 44A, 45A) changes the coefficients according to the weight balance between the front and rear of the vehicle body 2 and the weight of the vehicle body 2, which are determined in advance. Specifically, the heave BLQ (spring vertical damping BLQ21), roll BLQ (roll damping BLQ22) and pitch BLQ (pitch damping BLQ23) calculate the heave target damping force, roll target damping force and pitch target damping force taking into account the front and rear weight balance of the vehicle body. Therefore, the ratio between the target damping force on the front wheel side and the target damping force on the rear wheel side can be adjusted at a ratio according to the weight of the vehicle. As a result, the ride comfort can be improved even when the ride comfort and vehicle stability of the vehicle change depending on the number of passengers in the vehicle, the load weight, etc.

In the third embodiment, the controller 41 includes the GSP calculation unit 34 of the second embodiment, but may include the GSP calculation unit 13 of the first embodiment. Also, in the third embodiment, the weighting coefficient calculation unit is configured by the GSP calculation unit 34 and the correction gain calculation units 43A, 44A, and 45A. The present invention is not limited to this, and the GSP calculation unit 34 and the correction gain calculation units 43A, 44A, and 45A may be integrated to form a lookup table, and the weighting coefficient calculation unit may be configured by this lookup table.

Note that Figs. 4 and 17 show examples of lookup tables constituting the GSP calculation units 13 and 34, Fig. 6 shows an example of the damping coefficient upper limit calculation unit 25, and Fig. 7 shows an example of the command value calculation unit 29, but the present invention is not limited to these. The GSP calculation unit, the damping coefficient upper limit calculation unit, and the command value calculation unit are set appropriately taking into account the vehicle specifications, etc.

In the above embodiments, the controllers 11, 31, 41 have been described as acquiring vehicle driving parameters via the CAN 9, but the present invention is not limited to this. The controllers 11, 31, 41 may be connected to, for example, an acceleration sensor that detects the vertical acceleration of the sprung and unsprung parts, or a wheel speed sensor that detects the wheel speed, and the detected values of these sensors may be acquired as vehicle driving parameters. The controllers 11, 31, 41 may also acquire vehicle driving parameters from other controllers, etc.

In each of the above embodiments, the controllers 11, 31, 41 are provided with a state estimation unit 12 including a vertical motion calculation unit, a roll motion calculation unit, and a pitch motion calculation unit. In this case, the state estimation unit 12 determines the state related to the vertical motion of the vehicle from the CAN signal, determines the state related to the roll motion of the vehicle from the CAN signal, and determines the state related to the pitch motion of the vehicle from the CAN signal. The present invention is not limited to this, and for example, the controllers 11, 31, 41 may be connected to a vertical acceleration sensor, a roll sensor, a pitch sensor, etc., and calculate the state related to the vertical motion of the vehicle, the state related to the roll motion of the vehicle, and the state related to the pitch motion of the vehicle based on the detection values of these sensors. In this case, the parts of the controllers 11, 31, 41 that calculate the respective states constitute the vertical motion calculation unit, the roll motion calculation unit, and the pitch motion calculation unit.

In the above embodiments, the damping force adjustable shock absorber is configured with a variable damper 7 made of a semi-active damper. The present invention is not limited to this, and an active damper (either an electric actuator or a hydraulic actuator) may be used as the damping force adjustable shock absorber. In the above embodiments, the damping force adjustable shock absorber that generates an adjustable force between the vehicle body 2 side and the wheel 3 side is configured with a variable damper 7 made of a hydraulic shock absorber. The present invention is not limited to this, and the damping force adjustable shock absorber may be configured with an air suspension, an electromagnetic suspension, or the like, in addition to a hydraulic shock absorber.

In each of the above embodiments, a vehicle behavior device for use in a four-wheeled vehicle has been described as an example. However, the present invention is not limited to this, and can also be applied to, for example, two-wheeled and three-wheeled vehicles, or work vehicles and transport vehicles such as trucks and buses.

The specific numerical values described in each of the above embodiments are merely examples, and are not limited to the values shown. It goes without saying that each of the above embodiments is merely an example, and partial substitution or combination of the configurations shown in different embodiments is possible.

The present invention is not limited to the above-described embodiments, but includes various modified examples. For example, the above-described embodiments have been described in detail to clearly explain the present invention, and are not necessarily limited to those having all of the configurations described. It is also possible to replace part of the configuration of one embodiment with the configuration of another embodiment, and it is also possible to add the configuration of another embodiment to the configuration of one embodiment. It is also possible to add, delete, or replace part of the configuration of each embodiment with other configurations.

This application claims priority to Japanese Patent Application No. 2022-207940, filed December 26, 2022. The entire disclosure of Japanese Patent Application No. 2022-207940, filed December 26, 2022, including the specification, claims, drawings, and abstract, is hereby incorporated by reference in its entirety into this application.

1: Vehicle behavior control device, 2: Vehicle body, 3: Wheels, 5: Suspension device, 7: Adjustable damping shock absorber (variable damper), 8: Variable damping actuator, 9: CAN, 11, 31, 41: Controller (control device), 12: State estimation unit (vehicle behavior calculation unit, vertical movement calculation unit, roll movement calculation unit, pitch movement calculation unit, attitude change detection unit), 13, 34: GPS calculation unit (weighting coefficient calculation unit), 20, 42: BLQ controller, 21, 43 : sprung upper and lower vibration damping BLQ (heave BLQ), 22, 44: roll vibration damping BLQ (roll BLQ), 23, 45: pitch vibration damping BLQ (pitch BLQ), 24: target damping force calculation unit (actual target damping force calculation unit), 25: damping coefficient upper limit calculation unit, 26: corrected damping coefficient calculation unit, 27: target damping coefficient calculation unit (target damping coefficient calculation unit), 28: minimum value selection unit (correction unit), 29: command value calculation unit (control signal output unit), 32: road surface index calculation unit, 33: weight estimation unit

Claims (10)

A control device for a damping force adjustable shock absorber that is provided in a vehicle and is capable of adjusting a generated damping force,
a weighting coefficient calculation unit that calculates a weighting coefficient for making the actual target damping force closer to a predetermined target damping force when calculating an actual target damping force based on a calculated value of the vehicle behavior calculation unit;
A vertical motion calculation unit that calculates a vertical motion state of the vehicle;
a pitch motion calculation unit that calculates a motion state in a pitch direction of the vehicle;
an actual target damping force calculation unit that calculates the actual target damping force based on the weighting coefficient calculated by the weighting coefficient calculation unit and the calculation results of the vertical movement calculation unit and the pitch movement calculation unit;
a target damping coefficient calculation unit that calculates a target damping coefficient based on the calculated value of the actual target damping force calculation unit;
a correction unit that calculates a corrected damping coefficient by lowering an upper limit of the target damping coefficient in a region where a relative speed between a sprung portion and an unsprung portion of the vehicle is lower than a predetermined value;
a control signal output unit that outputs a control signal corresponding to the corrected damping coefficient to the damping force adjustable shock absorber,
the correction unit sets the corrected damping coefficient so that the damping force increases as the relative velocity increases, and so that a slope of the damping force with respect to the relative velocity is small when the relative velocity is slower than the predetermined value and a slope of the damping force with respect to the relative velocity is large when the relative velocity is faster than the predetermined value,
The correction unit has a maximum damping coefficient according to the relative speed, and when the target damping coefficient exceeds the maximum damping coefficient, corrects the target damping coefficient to the maximum damping coefficient. The control device according to claim 1 ,
Further comprising a posture change detection unit for detecting a posture change of the vehicle body,
The correction unit reduces an amount of correction when it determines that a posture change will occur based on a detection result from the posture change detection unit. The control device according to claim 1 ,
The control device further includes a roll motion calculation unit that determines a motion state in a roll direction of the vehicle. The control device according to claim 1 ,
The weighting coefficient calculation unit changes the weighting coefficient depending on a road surface condition. The control device according to claim 1 ,
The weighting coefficient calculation unit changes the weighting coefficient according to a weight of a vehicle body. The control device according to claim 1 ,
The weighting coefficient calculation unit changes the weighting coefficient in accordance with a selected damping force mode. The control device according to claim 1 ,
The weighting coefficient calculation unit changes the weighting coefficient according to a predetermined front-rear weight balance of the vehicle body and the weight of the vehicle body. The control device according to claim 1 ,
The weighting coefficient calculation unit calculates, as the weighting coefficient, a heave GSP related to a vertical movement of the vehicle and a pitch GSP related to a pitch movement of the vehicle,
The actual target damping force calculation unit includes a heave BLQ that calculates a target damping force in the vertical direction of the vehicle based on the heave GSP and the calculation result of the vertical movement calculation unit, and a pitch BLQ that calculates a target damping force in the pitch direction of the vehicle based on the pitch GSP and the calculation result of the pitch movement calculation unit,
A control device wherein the actual target damping force calculation unit calculates the actual target damping force based on the target damping force calculated by the heave BLQ and the target damping force calculated by the pitch BLQ. The control device according to claim 3,
the weighting coefficient calculation unit calculates, as the weighting coefficients, a heave GSP related to a vertical motion of the vehicle, a roll GSP related to a roll motion of the vehicle, and a pitch GSP related to a pitch motion of the vehicle;
The actual target damping force calculation unit includes a heave BLQ that calculates a target damping force in the vertical direction of the vehicle based on the heave GSP and the calculation result of the vertical movement calculation unit, a roll BLQ that calculates a target damping force in the roll direction of the vehicle based on the roll GSP and the calculation result of the roll movement calculation unit, and a pitch BLQ that calculates a target damping force in the pitch direction of the vehicle based on the pitch GSP and the calculation result of the pitch movement calculation unit,
A control device wherein the actual target damping force calculation unit calculates the actual target damping force based on the target damping force calculated by the heave BLQ, the target damping force calculated by the roll BLQ, and the target damping force calculated by the pitch BLQ. The control device according to claim 9,
A control device that calculates target damping forces for the heave BLQ, the roll BLQ and the pitch BLQ by taking into account the front and rear weight balance of a vehicle body.