JP2010221864A - Rolling behavior control device for vehicle - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To prevent a rolling behavior in steering from exceeding a tolerance through a reaction of lateral displacement when restraining the rolling behaviors in steering due to a lateral displacement of a structure. <P>SOLUTION: When a lateral displacement amount control of a movable structure for a rolling control is started, a steering angle θ, a wheel speed ω, a yaw rate Φ, forward and backward acceleration α<SB>xT</SB>, lateral acceleration α<SB>yT</SB>, and a motor rotation angle ϕ are first detected by a vehicle condition detecting part 21. Then, a variable-structure-displacement-amount-target-value calculation part 22 calculates a target-lateral-displacement-amount for restraining a rolling behavior change of the variable structure which gets smaller as a steering frequency is higher using the steering angle θ and the wheel speed ω. Thereafter, a variable structure driving part 23 calculates a motor driving torque command value required for realizing the target lateral displacement amount of the variable structure from the target lateral displacement amount of the variable structure, the yaw rate Φ, the forward and backward acceleration α<SB>xT</SB>, the lateral acceleration α<SB>yT</SB>, and the motor rotation angle ϕ, so as to output it to a servo driver for a motor. <P>COPYRIGHT: (C)2011,JPO&INPIT

Description

  The present invention relates to a rolling behavior control device for a vehicle that performs behavior control so as to suppress a change in behavior of a vehicle body in a rolling direction due to a steering wheel or a disturbance such as a cross wind.

  As such a rolling behavior control device for a vehicle, a device as described in Patent Document 1, for example, has been proposed.

This proposed technology can set a movable component that can be displaced left and right (vehicle width direction) with respect to the vehicle body, and the wheel suspension means can be displaced in the left and right direction (vehicle width direction) with respect to the vehicle body.
By displacing these movable components and suspension means in the left-right direction relative to the vehicle body so that the contact load acting on the left wheel of the vehicle and the contact load acting on the right wheel have the same value, The vehicle behavior is controlled so as to suppress the behavior change in the rolling direction.

Japanese Patent Laid-Open No. 2007-030566

However, the conventional rolling behavior control device described above has the following problems.
That is, for example, when the steering frequency is high (rapid steering is performed), this causes a sudden change in behavior in the rolling direction in which the vehicle height rapidly rises with respect to the vehicle turning direction outside. Therefore, the center of gravity of the vehicle needs to be quickly moved in the direction toward the inside of the turn, and the above-described movable components and suspension means are rapidly displaced in the direction toward the inside of the turn.

By the way, when the movable component or suspension means is displaced in the left-right direction so as to move the vehicle center of gravity in the left-right direction, the reaction acts on the vehicle body, and the vehicle body receives a roll moment in the opposite direction. The faster the displacement speed of an object or suspension means, the greater.
Therefore, in order to suppress sudden changes in the rolling direction behavior of the vehicle due to the above-described sudden steering, the movable components and the suspension means are rapidly displaced in the turning inner direction. Roll moment increases.

This large roll moment in the direction of the turn is a direction that further promotes the change in the rolling direction behavior of the vehicle due to the sudden steering, and the change in the rolling direction behavior of the vehicle becomes large, resulting in further behavior instability.
However, the above-described conventional rolling behavior control device has not proposed any means for solving such a problem, and causes a problem that the vehicle is greatly rolled toward the outer side of the turn during sudden steering and the behavior becomes unstable.

Although the present invention follows the rolling behavior control technology in which the behavior change in the rolling direction of the vehicle body is suppressed by the lateral displacement of the movable component as in the past,
During sudden steering, the amount of lateral displacement of the movable component is reduced, so that the roll moment generated by the reaction of the movable component in the lateral direction on the vehicle body is reduced, and large rolling outward The purpose of the present invention is to propose a rolling behavior control device for a vehicle that can solve the problem of behavioral instability caused by the vehicle.

For this purpose, the rolling behavior control device for a vehicle according to the present invention is:
The vehicle rolling behavior control device having a movable component that can be displaced in a direction that suppresses the behavior change in the rolling direction of the vehicle body is assumed as a basic premise of the gist configuration,
Steering information detection means for detecting vehicle steering information by the driver;
Based on the steering amount and the steering frequency among the steering information detected by the means, the displacement amount of the movable component corresponding to the steering amount, which is necessary for suppressing the rolling behavior change, becomes smaller as the steering frequency is higher. The movable component displacement amount setting means for setting as described above is provided.

According to the rolling behavior control device of the present invention,
Because the displacement amount of the movable component according to the steering amount necessary to suppress the rolling behavior change at the time of steering is made smaller as the steering frequency is higher,
During sudden steering, the amount of displacement of the movable component is reduced, and the roll moment generated by the reaction caused by the displacement of the movable component on the vehicle body can be reduced. The problem of behavioral instability due to can be solved.

1 is a schematic plan view of a vehicle equipped with a rolling behavior control device according to a first embodiment of the present invention. FIG. 2 is a schematic rear view showing the vehicle of FIG. 1 as viewed from the rear. FIG. 2 is a control system diagram of a behavior control motor by a controller in FIG. 1 and 2 show the rolling behavior control operation of the vehicle, (a) is a rear view similar to FIG. 2 showing the vehicle state before starting rolling behavior control, and (b) is the rolling behavior control performed. FIG. 3 is a rear view similar to FIG. 2 showing a vehicle state at the time. FIG. 2 is a flowchart showing a motor drive torque control processing process executed by the controller in FIG. 1 when controlling rolling behavior. FIG. FIG. 6 is a characteristic diagram illustrating a target wheel load movement rate characteristic set in the first example of FIGS. FIG. 6 is a characteristic diagram illustrating a change characteristic of a steady gain set in the first example of FIGS. FIG. 6 is a characteristic diagram showing the frequency response of the wheel load movement rate at low vehicle speeds for each cutoff frequency in the first embodiment of FIGS. FIG. 6 is a characteristic diagram showing the frequency response of the wheel load movement rate at high vehicle speeds for each cut-off frequency in the first example of FIGS. FIG. 6 is a characteristic diagram illustrating a change characteristic of a cutoff frequency set in the first example of FIGS. The operation when the rolling behavior control device of the first embodiment in FIGS. 1 to 5 is applied to a vehicle having a relatively narrow tread, when the rolling behavior control is not performed at all, and the conventional rolling behavior control is performed. It is an operation time chart shown in comparison with the case. In the rolling behavior control apparatus according to the second embodiment of the present invention, the difference between the phase delay amount in the frequency response of the movable component swing displacement amount and the phase delay amount in the frequency response of the vehicle body lateral acceleration is the smallest It is explanatory drawing for demonstrating the point at the time of calculating | requiring the combination of a gain and a cutoff frequency, (a) is a steering frequency when a stationary gain is temporarily placed at a certain value and a cutoff frequency is set to 0.5 Hz. (B) is a characteristic diagram showing the frequency response of the swinging displacement of the movable component and the lateral acceleration of the vehicle body, (b) is the movable relative to the steering frequency when the steady-state gain is temporarily placed at the same value and the cutoff frequency is 1 Hz. (C) is a characteristic diagram showing the frequency response of the swing displacement of the component and the lateral acceleration of the vehicle body. (C) shows the possible response to the steering frequency when the steady-state gain is temporarily placed at the same value and the cutoff frequency is 2 Hz. Construct swing displacement and is a characteristic diagram showing the frequency response of the vehicle lateral acceleration. 10 is an operation time chart showing a case where steering equivalent to emergency avoidance is performed by comparing the rolling behavior control of the second embodiment with the rolling behavior control of the first embodiment. FIG. 6 is a flowchart similar to FIG. 5, showing a motor driving torque control process executed by the controller of the rolling behavior control apparatus according to the third embodiment of the present invention in the rolling behavior control. FIG. 15 is a flowchart showing a control program related to a movable component displacement amount target value limiting process during the processing process in FIG. 14. FIG. FIG. 15 is a flowchart showing a control program related to a motor drive torque limiting process in the process in FIG.

Hereinafter, embodiments of the present invention will be described in detail based on first to third embodiments shown in the drawings.
<Configuration of the first embodiment>
1 is a schematic plan view of a vehicle equipped with a rolling behavior control apparatus according to a first embodiment of the present invention, FIG. 2 is a schematic rear view showing the vehicle viewed from the rear, and FIG. 3 is a controller in FIG. It is a control system figure of the motor for behavior control by.

1 and 2, reference numeral 1 denotes a vehicle body, and 2 denotes four wheels. Each wheel 2 is suspended from the vehicle body 1 by a suspension device 3 so as to be bound and rebound.
Each suspension device 3 is composed of an upper arm 3a, a lower arm 3b, and a strut 3c. The lower end, which is a wheel side attachment point of the strut 3c, is generally connected to the lower arm 3b. The upper end, which is the attachment point, is connected to the rolling member 4.

One rolling member 4 is provided for the left and right front wheels and one for the left and right rear wheels. Each of the rolling members 4 can swing around an axis extending in the front-rear direction at the center position in the vehicle width direction of the vehicle body 1. And pivot on the car body 1.
The upper ends of the struts 3c constituting the left and right front wheel suspension devices 3 are respectively connected to both ends of the left and right front wheel rolling members 4 at the front, and the left and right rear wheel suspension devices 3 are respectively connected to both ends of the left and right rear wheel rolling members 4 at the rear. Connect the upper ends of the struts 3c.

The front and rear rolling members 4 are coupled to an output shaft 5a (see FIG. 1) of the motor 5 so that the swing position can be controlled by individual motors 5 provided on the vehicle body 1.
According to the swing position control of the rolling member 4 via the motor 5, the case where the rolling member 4 is swung in the arrow direction from the horizontal position in FIG. As shown in b), the attachment point on the vehicle body side of the right wheel strut 3c is raised in the direction of the arrow, and the attachment point on the vehicle body side of the left wheel strut 3c is lowered in the direction of the arrow.

Thus, the left and right wheel struts 3c move up and down in opposite directions, so that the left and right wheel suspension device 3 moves the vehicle body 1 in the swing direction of the rolling member 4 as shown by arrows in FIG. The position of the center of gravity in the left-right direction of the vehicle can be moved in the corresponding direction.
Due to such movement of the center of gravity of the vehicle in the left-right direction, the wheel load is increased or decreased between the left and right wheels, and by controlling the swing position of the rolling member 4 via the motor 5, changes in the rolling behavior of the vehicle can be suppressed.

Therefore, the left and right wheel suspension device 3 and the vehicle body 1 act as movable components that can be displaced in a direction that suppresses changes in the rolling behavior of the vehicle,
The motor 5 and the rolling member 4 constitute a movable component driving means that controls the displacement.

However, the movable structure that can be displaced in a direction that suppresses changes in the rolling behavior of the vehicle is not limited to the left and right wheel suspension device 3 and the vehicle body 1 as in this embodiment,
Only the cabin can be swung in the rolling direction with respect to the vehicle body, and its swing position can be controlled,
Only the cabin can be moved horizontally to the left and right with respect to the vehicle body, and its movement position can be controlled,
It is possible to move heavy objects such as wheels and batteries horizontally in the left-right direction, and to control the movement position,
It goes without saying that a movable component that can be swung or horizontally moved in a direction that suppresses a change in rolling behavior of the vehicle may be provided.

It is assumed that the swing control of the front and rear yawing members 4 is performed by the controller 11 via the individual motors 5 respectively.
Therefore, the controller 11 has a signal from the steering angle sensor 12 that detects the steering angle θ of the steering wheel 6 that steers the left and right front wheels of the wheels 2, and
A signal from the wheel speed sensor 13 that individually detects the wheel speed ω of each wheel 2 (for convenience, the wheel speeds of all the wheels are indicated by the same symbol);
A signal from the vehicle behavior sensor group 14 provided at the center of gravity of the vehicle,
A signal from the motor rotation sensor 15 that detects the rotation angle φ from the neutral position of the motor 5 is input.
The sensors 12 to 15, the controller 11, and the motor 5 are connected as shown in FIG.

As shown in FIG. 3, the vehicle behavior sensor group 14 detects the yaw rate sensor 14a for detecting the yaw rate Φ of the vehicle, the front / rear G sensor 14b for detecting the longitudinal acceleration α _xT of the vehicle, and the lateral acceleration α _{yT of the} vehicle. It is assumed that it comprises a lateral G sensor 14c.

  As will be described in detail later, the controller 11 calculates the target displacement amount (target swing angle of the rolling member 4) of the movable component (suspension device 3 and vehicle body 1) based on the input information described above, and calculates this target value. The drive torque command value of the motor 5 necessary to achieve is determined.

The controller 11 supplies the motor drive torque command value to the motor servo driver 16 shown in FIGS. 1 and 3, and the servo driver 16 converts the electric power of the battery 17 in response to the motor drive torque command value. Then, electric power corresponding to the motor drive torque command value is supplied to the motor 5.
Thus, the motor 5 is controlled so that the drive torque becomes the same as the above command value, and the movable component (suspension device 3 and the vehicle body 1) can be oscillated and displaced by the target displacement amount via the rolling member 4. .

<Calculation of motor drive torque command value>
The controller 11 calculates the motor drive torque command value as follows.
As will be described in detail with reference to FIG. 5, when the displacement control of the movable components (suspension device 3 and vehicle body 1) is started, first the vehicle state detection unit 21 performs steering angle θ, wheel speed ω, yaw rate Φ, longitudinal acceleration. alpha _xT, lateral acceleration alpha _yT, and detects the motor rotation angle phi.
Therefore, the vehicle state detection unit 21 corresponds to the steering information detection means in the present invention.

Next, the movable component displacement amount target value calculation unit 22 constituting the movable component displacement amount setting means in the present invention uses the detected steering angle θ and the wheel speed ω to detect the movable component (suspension device). 3 and the target displacement amount of the vehicle body 1) (the target swing angle of the rolling member 4) is calculated.
Finally, in the movable component drive unit 23, the target displacement amount (target swing angle of the rolling member 4) of the movable component (suspension device 3 and vehicle body 1), the yaw rate Φ, the longitudinal acceleration α _xT , the lateral acceleration described above. Based on the acceleration α _yT and the motor rotation angle φ, the drive torque command for the motor 5 necessary to realize the target displacement (target swing angle of the rolling member 4) of the movable component (suspension device 3 and vehicle body 1). A value is calculated, and this motor drive torque command value is output to the motor servo driver 16.

In calculating the target displacement amount (target swing angle of the rolling member 4) of the movable component (suspension device 3 and vehicle body 1) by the movable component displacement amount target value calculation unit 22, first, "(A) vehicle body speed" is calculated. In the “calculation section”, the vehicle body speed V is calculated as follows.
Therefore, the “(A) vehicle body speed calculation unit” corresponds to the vehicle body speed calculation means in the present invention.
(A) Vehicle speed calculation unit The vehicle speed V is calculated by using the wheel speed ω. At this time, the average value of the wheel speeds ω of the four wheels may be used as the vehicle speed V, or the wheel speed ω of the non-driven wheels may be calculated. The average value may be the vehicle speed V.
In the present embodiment adopts the method of calculating the vehicle speed V from only the wheel speed omega, subjected to correction by addition to the wheel speed omega example vehicle longitudinal acceleration alpha _xT, it may be calculated vehicle speed V .

  When calculating the target displacement amount (target swing angle of the rolling member 4) of the movable component (suspension device 3 and vehicle body 1) by the movable component displacement amount target value calculation unit 22, next, "(B) steady state In the “gain / cutoff frequency calculation section”, the steady gain and the cutoff frequency are calculated as follows.

(B) Steady Gain / Cutoff Frequency Calculation Unit In this embodiment, the former steady gain value is first calculated, and then the cut-off frequency is determined in accordance with the steady gain value.
(B-1) Calculation of steady gain The steady gain here refers to how much the movable component swings and displaces per unit steering angle during steady steering (during steady steering at a steady steering determination frequency of 0.2 Hz or less). It is a value indicating
Therefore, changing the value of the steady gain changes the response of the wheel load change to the steering and the response of the lateral motion / yaw motion to the steering.

In the present embodiment, from the viewpoint of preventing the vehicle from performing rolling behavior exceeding the allowable limit in any vehicle state, the right and left wheel load movement amount of either the front wheel or the rear wheel is It is assumed that a steady gain that is a desired value is calculated at any vehicle speed.
In order to ensure the stability of the vehicle, it is necessary to manage the rear wheel load movement amount so that it does not become too large. Here, the rear wheel load movement amount is set to a desired value. A method for calculating a steady gain will be described in detail.

(B-1-1) Calculation of wheel load movement rate First, parameters determined from vehicle specifications are defined as follows.
However, the unit is SI unit system.
m _s : Mass of movable component m _T : Mass of vehicle body (excluding movable component)
hs: Length of the roll arm of the movable component (distance from the height of the center of gravity of the movable component to the center of the roll rotation of the vehicle)
h _T : Body roll arm length (distance from the center of gravity height of the mass other than movable components to the center height of the roll rotation motion of the body)
G: Average roll stiffness before and after l: Foil base
l _f : Distance from front wheel axle to vehicle center of gravity
l _r : Distance from vehicle center of gravity to rear wheel axle b: Tread
K _f : Front wheel cornering stiffness
_Kr : Rear wheel cornering stiffness
K _cf : Front wheel canvas stiffness
K _cr : Rear wheel canvas stiffness
N: Steering gear ratio (here, treated as a constant)
F _z0 : Wheel load per wheel at rest (rear wheel load is used here)
g: Gravity acceleration

The following are set as variables.
V: Body speed
K _φQ δ _f0 : Steady gain

Here, a value obtained by dividing the steady wheel load movement amount per unit steering angle by the wheel load is defined as a steady wheel load movement rate ΔW ₀ F _zo θ ₀ per unit steering angle.
By solving the equation of motion of the vehicle, the steady wheel load movement rate per unit steering angle is obtained as a function of the vehicle body speed V and the steady gain K _φQ δ _f0 as follows.

(B-1-2) Setting the target wheel load mobility characteristics As for the vehicle speed dependence of the steady wheel load mobility per unit steering angle in the previous section, by determining the target characteristics, the steady gain depends on the vehicle speed. How sex should be set is determined.
In this embodiment, an ideal normative vehicle model is assumed, and the vehicle speed dependency of the steady wheel load movement rate of the vehicle model is set as a target.

For example, the characteristic shown in FIG. 6 is defined as a target wheel load movement rate characteristic ΔW ₀ F _zo θ ₀ (V).
In a vehicle with this desired characteristic, for example, when the wheel load per wheel is 2500 N, when traveling at a vehicle speed of 100 km / h, and slowly steering 30 deg,
ΔW = 0.01 (wheel load movement rate per unit steering angle) x 30 x 2500 = 750 [N]
The left and right wheel load movement will occur.

のように求まる。
上式に各パラメータの値をそれぞれ代入して得た、車速に対する定常ゲインの変化特性を図7に示す。
ここでの符号の向きは、負の値のときに転舵と逆方向に可動構成物が揺動変位し、転舵と同方向に車体1が傾くことを意味する。 (B-1-3) Calculation of steady-state gain in which wheel load movement amount has desired characteristics From the previous and previous studies, steady-state gain K _φQ δ _f0 (V) that realizes desired characteristics is

It is obtained like this.
FIG. 7 shows the steady-state gain variation characteristics with respect to the vehicle speed, obtained by substituting the values of the parameters in the above equation.
The direction of the sign here means that when the value is a negative value, the movable component swings and displaces in the direction opposite to that of the steering, and the vehicle body 1 tilts in the same direction as the steering.

In the vehicle of the present embodiment, the value of the steady gain is always a negative value. Therefore, it can be understood that the vehicle body must always be tilted inward in order to achieve a desired wheel load movement rate.
In FIG. 7, the steady-state gain of the wheel load is set to a desired characteristic, as can be seen from the fact that the steady-state gain value at the vehicle speed V = 100 km / h is four times or more that at the vehicle speed V = 30 km / h. Therefore, it is necessary to greatly change the steady gain value depending on the vehicle speed V.

(B-2) Calculation of cut-off frequency Here, the cut-off frequency means that when the gain tends to decrease according to the signal frequency in the system, the gain is only a set value (for example, 3 dB) rather than the steady value. This is the steering frequency when it decreases.
For example, when the cut-off frequency is lowered, the gain starts to decrease from a lower steering frequency, and when the cut-off frequency is increased, the gain is not easily lowered even at a high steering frequency.

Therefore, in this embodiment, by changing this cutoff frequency, the swing displacement amount target value of the movable component in the region where the steering frequency is high is adjusted.
While the setting was made so that the wheel load movement amount at the time of steady steering becomes a desired value by the steady gain, in the setting of the cut-off frequency, the range of about 0.5 Hz to 2 Hz which is a practical steering frequency band In this case, it is necessary to set the wheel load movement amount to a desired value.

で表される一次のローパスフィルタの形で想定し、定常ゲインK_φQδ_f０(V)に対応するカットオフ周波数F₀(V)を求めることとする。 When a high-order filter is used, the phase delay becomes large, and there is a problem that the responsiveness of the movable component is deteriorated with respect to steering. Therefore, in this embodiment, the swing displacement target of the movable component with respect to the steering angle θ is generated. The characteristic of the value φ * is

The cut-off frequency F ₀ (V) corresponding to the steady-state gain K _φQ δ _f0 (V) is obtained.

FIG. 8 shows the frequency response of the wheel load transfer rate when the vehicle speed is low (for example, 30 km / h to 40 km / h) by comparing each cut-off frequency.
FIG. 9 shows the frequency response of the wheel load movement rate for each cutoff frequency when the vehicle speed is high (for example, 120 km / h to 160 km / h).

  In the present embodiment, the following criteria are provided as criteria for selecting a suitable cutoff frequency for each vehicle speed.

(Criteria for determining the lower limit of the cutoff frequency)
If the gain increases between the steering frequencies of 0.5 Hz to 1 Hz required during emergency avoidance steering, excessive wheel load movement occurs during emergency avoidance steering, and vehicle rolling behavior exceeding the limit occurs.
Therefore, in this embodiment, a lower threshold is provided for the steering frequency of 1 Hz, and in the steering frequency region below the lower threshold, the cutoff is performed so that the gain of the wheel load movement amount with respect to the steering does not exceed the gain during steady steering. The frequency is selected.
8 and 9, it corresponds to selecting a cut-off frequency that does not have a gain characteristic curve in the hatched area in these gain diagrams.

(Criteria for determining the upper limit of the cut-off frequency)
In the frequency characteristic of the wheel load movement amount with respect to the steering, if the phase shift is large, a sudden wheel load fluctuation may occur during the steering, and the vehicle behavior may become unstable.
Therefore, in this embodiment, threshold values are respectively set for the phase delay of −45 deg and the phase advance of 45 deg, and the cutoff frequency is selected so that the phase shift of the wheel load movement amount with respect to the steering is within these threshold ranges. To do.
8 and 9, this corresponds to selecting a cut-off frequency that does not have a phase characteristic curve in the hatched region in these phase diagrams.

  In light of the above criteria, it can be seen that the cutoff frequency that satisfies the condition at low vehicle speed in FIG. 8 is approximately 2 Hz, and that the cutoff frequency that satisfies the condition at high vehicle speed in FIG. 9 is approximately 1 Hz.

FIG. 10 illustrates the vehicle speed dependency of the cutoff frequency f ₀ (V).
Compared to the characteristics of the steady gain K _φQ δ _f0 (V) in Fig. 7, in order to ensure the stability of the vehicle while avoiding the rolling behavior of the vehicle exceeding the allowable limit, the steady gain is set high as a result. It can be seen that it is effective to set the cut-off frequency low in the high vehicle speed range, and to set the cut-off frequency high in the low vehicle speed range where the steady gain is set low.

  When calculating the target displacement amount (target swing angle of the rolling member 4) of the movable component (suspension device 3 and vehicle body 1) by the movable component displacement amount target value calculation unit 22 in FIG. ) The displacement target value calculation unit using the low-pass filter ”calculates the swing displacement target value of the movable component as follows.

で表される伝達関数を持った一次のローパスフィルタを構成することができる。
このローパスフィルタを用いることで、操舵角入力θに対する可動構成物の揺動変位量目標値φ*を決めることができる。 (C) Displacement target value calculation section using low-pass filter The steady gain K _φQ δ _f0 (V) and the cut-off frequency f ₀ (V) are obtained by the above calculation, so the characteristics change according to the vehicle speed V To

A first-order low-pass filter having a transfer function represented by
By using this low-pass filter, it is possible to determine the swing displacement amount target value φ * of the movable component with respect to the steering angle input θ.

5, the target swing displacement φ * of the movable components (suspension device 3 and vehicle body 1), the yaw rate Φ detected by the vehicle state detection unit 21, the longitudinal acceleration α _xT , the lateral When calculating the drive torque command value of the motor 5 necessary for realizing the target displacement φ * of the movable component (suspension device 3 and the vehicle body 1) from the acceleration α _yT and the motor rotation angle φ, D) The motor drive torque command value calculation unit performs the calculation as follows.

(D) Motor drive torque command value calculation unit Calculates the motor drive torque command value according to the deviation between the target swing displacement φ * of the movable component calculated in the previous section and the actual swing displacement detection value φ. To do.
The value of the motor driving torque required to move the movable component to the target swing displacement amount changes according to the vehicle state amount such as the yaw rate Φ, the longitudinal acceleration α _xT , and the lateral acceleration α _yT .
Therefore, vehicles such as motor drive torque, yaw rate Φ, longitudinal acceleration α _xT , and lateral acceleration α _yT according to the deviation between the target swing displacement amount φ * of the movable component and the actual swing displacement detection value φ A motor drive torque command value is determined together with a correction torque that takes the state quantity into consideration.

(G) Motor drive torque command value output unit In the movable component drive unit 23 of FIG. 5, the motor drive torque command value obtained as described above is further converted into the motor drive torque command value output unit "(G) Motor drive torque command value output unit". Output to servo driver 16.

<Effect of the first embodiment>
By controlling the motor 5 as described above, the movable components (the suspension device 3 and the vehicle body 1) are controlled such that the swing displacement amount with respect to the steering angle input θ becomes the target swing displacement amount φ * with respect to the steering angle input θ. Is done.
By the way, the target swing displacement amount φ * with respect to the steering angle input θ is obtained by using the first-order low-pass filter capable of changing the steady-state gain K _φQ δ _f0 (V) and the cut-off frequency F ₀ (V) according to the vehicle state. Because it is the one that has been specifically requested,
The swing displacement amount of the movable component with respect to the steering angle input θ necessary for suppressing the rolling behavior change during steering can be reduced as the steering frequency increases.

For this reason, during sudden steering with a high steering frequency, the swing displacement amount of the movable component with respect to the steering angle input θ is reduced.
It is possible to reduce the roll moment that the reaction caused by the displacement of the movable component reaches the vehicle body, and to prevent the rolling behavior of the vehicle from exceeding the allowable limit even under a high steering frequency. It is possible to avoid instability of behavior due to large rolling exceeding the allowable limit in the turning outer direction.

In this embodiment, the primary low-pass filter is applied as described above to set the swing displacement amount of the movable component with respect to the steering angle input θ to be smaller as the steering frequency is higher, Because it is a configuration that can change the filter characteristics accordingly,
Even if the state of the vehicle changes, the amount of swing displacement of the movable component is appropriately adjusted each time, so that the above-mentioned effects can always be achieved, and the rolling behavior performance of the vehicle can be improved. Regardless of the situation, it will always be possible.

Further, in the present embodiment, the steady gain K _φQ δ _f0 (V) described above depends on the vehicle body speed V, or on the difference between the vehicle body speed V and the vehicle body speed V (time change rate of the vehicle body speed V). To change
Even if the steering amount is the same, the steady-state gain K _φQ δ _f0 (V) is changed in consideration of the property that the lateral acceleration and yaw rate of the vehicle differ when the vehicle body speed V and the time change rate differ. Become.

By the way, regardless of the vehicle speed V, if the movable component is always oscillated and displaced with respect to the steering angle,
For example, the swing displacement amount of the movable component tends to be too large at a low vehicle speed range where a large steering angle is frequently used, and the swing displacement amount of the movable component is small at a high vehicle speed region where only a small steering angle is used. There is a tendency to become too much.

By the way, as in this embodiment, if the steady gain K _φQ δ _f0 (V) is changed in accordance with the vehicle body speed V and its time change rate, it can be moved based on any vehicle speed V and its time change rate. The swing displacement amount of the object can be made appropriate without excess and deficiency, and the rolling behavior performance of the vehicle can always be improved regardless of the vehicle body speed V and its time change rate.

As described above with reference to FIGS. 7 and 10, as the absolute value of the steady gain K _φQ δ _f0 (V) is set larger, the cutoff frequency f ₀ (V) is set lower and the steady gain K _φQ δ _f0 (V) If the cut-off frequency f ₀ (V) is set higher as the absolute value is set smaller,
For each steady gain K _φQ δ _f0 (V), the _highest possible cut-off frequency f ₀ (V) can be set, and if the cut-off frequency f ₀ (V) can be set high, the vehicle for steering It is possible to reduce the driver's uncomfortable feeling at the time of steering and improve the steering feeling while maintaining the improvement of the rolling behavior performance of the vehicle while the movement response delay of the center of gravity is reduced.

Furthermore, as described above with reference to FIGS. 8 and 9, when setting the cutoff frequency f ₀ (V),
(1) _Restriction that the gain of the wheel load movement amount with respect to the steering amount is equal to or less than the gain K _φQ δ _f0 (V) during steady steering in a steering frequency region of a relatively low predetermined steering frequency (1 Hz) or less. To define the lower limit of the cut-off frequency f ₀ (V),
(2) In a steering frequency region of a relatively high predetermined steering frequency (3 Hz) or less, the phase shift of the wheel load movement amount with respect to the steering amount is changed from a predetermined phase delay amount (−45 deg) to a predetermined phase advance amount (45 deg.). ) To set the upper limit of the cut-off frequency f ₀ (V)
When the cutoff frequency f ₀ (V) is set between the upper limit and the lower limit of these cutoff frequencies, the following effects can be obtained.

In other words, according to the condition (1), the amount of wheel load movement in the range of 0.5 Hz to 1 Hz required for emergency avoidance can be suppressed to less than that in steady state, and rolling behavior exceeding the allowable limit of vehicles occurs during emergency steering. This can be prevented more reliably.
In addition, the phase shift of the wheel load movement amount with respect to the steering is within a range of about ± 45 deg. Can fit in.
Incidentally, if the phase shift is large, sudden wheel load fluctuations may occur in the middle of the steering operation, and there is a possibility that the vehicle behavior may become unstable. The fluctuation becomes gentle and the vehicle behavior can be stabilized.

  The above-described idea of the present embodiment can be applied to any specification vehicle. However, the above-mentioned various effects are maximized with respect to the height of the center of gravity. This is the case for narrow vehicles.

FIG. 11 is an operation time chart when the rolling behavior control device of the present embodiment described above is applied to a vehicle having a relatively narrow tread.
This figure shows the case of no control for the steering input equivalent to emergency avoidance, the case of applying the control of the prior art that controls the left and right wheel loads to the same value, and the control of this embodiment described above. This is a comparison of the wheel load transfer rate.

Without control, the gain of the displacement amount is of course 0, but the tread is narrow. Therefore, the absolute value of the wheel load movement rate exceeds 1 and the wheel load on either the left or right side becomes 0, and the rolling behavior is not effective. It becomes stable.
When the control of the prior art is applied, the left and right wheel loads are controlled to have the same value. Therefore, the gain of the displacement amount is too high, and the reaction of the sudden movement of the center of gravity makes the remarkably and earlier than the case without control. Cause instability in rolling behavior.
On the other hand, when the control of the present embodiment is applied, the gain of the displacement amount is not too high and is kept at an appropriate gain even under a high steering frequency, and the center of gravity moves slowly, and its reaction The rolling behavior instability caused by can be avoided.

<Second embodiment>
The present embodiment has basically the same configuration as the first embodiment described above, but “(B) steady gain / cutoff frequency calculation section in the movable component displacement amount calculation section 22 shown in FIG. ”In particular calculates the steady-state gain and the cut-off frequency based on the following control law.

(B) Steady Gain / Cutoff Frequency Calculation Unit In the present embodiment, the cut-off frequency F ₀ (in the same manner as that described in the section “(B-2) Calculation of cut-off frequency” in the first embodiment, above. Set upper and lower limits for V), but in addition, add the following phase characteristic conditions.

The “phase characteristic condition” in this example is
Of the combination of the steady gain value and the cutoff frequency value, in the region of the steering frequency below the threshold value, the phase lag amount in the frequency response of the swing displacement amount of the movable component with respect to the steering angle, and the vehicle body lateral acceleration with respect to the steering angle. Means the combination of steady gain and cutoff frequency so that the difference from the phase lag in the frequency response is as small as possible.
In this embodiment, a combination of a steady gain and a cutoff frequency that satisfies the phase characteristic condition is selected.
In this case, since the two variables of the steady gain and the cut-off frequency are determined simultaneously, the steady gain and the cut-off frequency are determined using the method described below (so-called qualifying final method).

(B-2-1) Temporary placement of steady gain First, the steady gain value is placed temporarily.
The steady gain is obtained by the same method as that described in the section “(B-1-3) Calculation of steady gain so that wheel load movement amount has desired characteristics” in the first embodiment. It is obtained by determining the mobility characteristics.

(B-2-2) Selection of cut-off frequency satisfying phase characteristic condition Next, when setting the cut-off frequency, for example, in order to make the characteristic of the steering frequency 0.5 Hz to 1 Hz necessary for emergency avoidance a desired characteristic, a threshold value is set. Is defined as 1 Hz.
Then, as described in the section of (B-2) Calculation of cut-off frequency in the first embodiment, the threshold 1 Hz is selected from the cut-off frequencies that satisfy the upper and lower limit restrictions of the cut-off frequency. In the following steering frequency region, a cutoff frequency that satisfies the above-mentioned phase characteristic condition is selected.

(B-2-3) Repeating qualifying The qualifying process described in “(B-2-1) Temporary placement of steady gain” and “(B-2-2) Selection of cut-off frequency satisfying phase characteristics” is performed. In this case, the qualifying process is repeated while changing the steady gain value little by little.

(B-2-4) Final Of the combination of steady gain and cut-off frequency obtained by repeating the qualifying process described above, the frequency response of the movable component swing displacement amount that best satisfies the phase characteristic condition described above Finally, the combination of the steady gain and the cutoff frequency is selected so that the difference between the phase lag amount at and the phase lag amount in the frequency response of the vehicle body lateral acceleration is minimized.

<Operational effects of the second embodiment>
12 (a), 12 (b), and 12 (c) show the movable component swing displacement amount and the vehicle body with respect to the steering frequency when the steady-state gain is temporarily placed at a certain value and the cutoff frequency is sequentially changed. The frequency response of a lateral acceleration is illustrated.
The three types of cutoff frequencies 0.5 Hz, 1 Hz, and 2 Hz listed as candidates in FIGS. 12 (a), (b), and (c) are the upper and lower limits of the cutoff frequency described above for the first embodiment, respectively. It satisfies the related constraints.

  As is clear from the comparison of FIGS. 12 (a), (b), and (c), when the cutoff frequency is 1 Hz, the phase characteristic condition is best satisfied, and the frequency response of the swing displacement amount of the movable component It can be seen that the difference between the phase delay amount and the phase delay amount in the frequency response of the vehicle body lateral acceleration is the smallest combination of the steady gain and the cut-off frequency.

According to the rolling behavior control of the second embodiment using the combination of the steady gain and the cut-off frequency, in addition to the same operational effects as the first embodiment, the following operational effects can also be achieved. .
Referring to FIG. 13 which shows an operation time chart when steering equivalent to emergency avoidance is performed, the frequency response of the movable component swing displacement amount becomes the frequency response of the vehicle body lateral acceleration by satisfying the phase characteristic condition described above. Close to.

For this reason, as apparent from the smoothness of the wheel load movement rate indicated by the solid line, the operation of the vehicle body becomes smoother than the case of the first embodiment indicated by the broken line, and the rolling behavior of the vehicle does not exceed the allowable limit. In this manner, the driver can feel an improvement in steering feeling while maintaining stable rolling behavior performance.
Further, as is apparent from the time series change of the motor output indicated by the solid line, it is also possible to reduce the burden on the motor 5 that swings and displaces the movable component than that of the first embodiment indicated by the broken line. .

Incidentally, the advantage of the first embodiment is that the “phase characteristic condition” as in the second embodiment is not set, so the degree of freedom in selecting a steady gain is large,
Therefore, a relatively large steady-state gain can be set, and as is clear from the time series change in the left and right wheel load movement rate indicated by the broken line in FIG. 13, the wheel load movement is greater than in the second embodiment indicated by the solid line. The rate (wheel load movement amount) can be reduced, and improvement in vehicle motion performance can be expected.

<Third embodiment>
As shown in FIG. 14, the present embodiment has basically the same configuration as that of the first embodiment or the second embodiment described above, but in the movable component displacement amount calculation unit 22 shown in FIG. E) “displacement target value limiter” is added, and “(F) drive torque limiter” is added to the movable component drive unit 23 shown in FIG.
Hereinafter, control rules of the added “(E) displacement amount target value limiting unit” and “(F) drive torque limiting unit” will be described.

(E) Displacement target value limiter For example, when the vehicle's center of gravity is suddenly moved in the direction of turning inside, such as when the steering frequency is high, the roll moment in a direction that further facilitates left and right wheel load movement due to the reaction of quick center of gravity movement And the vehicle is rolling beyond the allowable limit.

This reaction roll moment is generated by the relative displacement acceleration of the vehicle body relative to the unsprung mass (relative lateral acceleration of the vehicle body coordinate system relative to the unsprung mass coordinate system).
Therefore, attention is paid to the value of the relative displacement acceleration of the vehicle body, and control is performed so that this value does not exceed the threshold value, thereby preventing the vehicle from rolling beyond the allowable limit.

FIG. 15 is a flowchart of a control program executed by “(E) displacement amount target value limiting unit” in FIG. 14, and the processing contents in each step will be described in detail below.
(E-1) Prediction of the relative displacement acceleration of the vehicle body against the unsprung mass Before the relative displacement acceleration of the vehicle body that causes the vehicle to roll beyond the permissible limit, the swing displacement target value of the movable component is obtained in advance. Need to be changed,
Therefore, in the first step (E-1), a predicted value of the relative displacement acceleration of the vehicle body is calculated in advance from the control target value instead of the detected value, and this predicted value is used for threshold determination.

The predicted value α _rel of the relative displacement acceleration of the vehicle body is
φ *: Target swing value of movable components φ _Roll : Roll angle predicted value based on suspension stroke h: Distance from the height of the center of rotation when the vehicle rolls to the center of gravity of the vehicle (substantially roll Arm length)
Can be calculated by the following equation.
α _rel = h {(d2 / dt) φ * + (d2 / dt) φ _Roll }

Note that the roll angle predicted value φ _Roll based on the suspension stroke is calculated using the transfer function φ _Roll (s) / θ (s) with respect to the steering angle θ.
The transfer function may be calculated from a linearly approximated equation of motion, or may be identified from measured characteristics.

なおここでは、前輪の方が後輪よりも荷重移動量が大きく設定されているものと仮定し、閾値に前輪の諸元値を用いる。
但し、
ｍ_ｓ：可動構成物の質量
ｍ_T：車体質量（可動構成物を除く）
hs：可動構成物のロールアーム長（可動構成物重心高さから、車体のロール回転運動中心高さまでの距離）
h_T：車体ロールアーム長（可動構成物以外のマスの重心高さから、車体のロール回転運動中心高さまでの距離）
b_f：前輪トレッド
ｘ：前輪のロール剛性配分比から決まる定数(0<x<1を満たす)
ΔW_fMAX：前輪荷重移動量の許容最大値（１輪あたり静止時輪荷重よりわずかに少ない値）
φ：可動構成物の揺動変位量の値（現在の値）
α_yT：車体の横加速度（計測値・操舵角からの推定値いずれを用いてもよい） (E-2) Calculation of Acceleration Threshold The acceleration threshold α _{rel_Limit} may be determined by selecting an appropriate value based on actual measurement, but can be determined, for example, as follows based on an equation of motion in the roll direction.

Here, it is assumed that the load movement amount is set larger for the front wheels than for the rear wheels, and the specification values of the front wheels are used as the threshold values.
However,
m _s : Mass of movable component m _T : Mass of vehicle body (excluding movable component)
hs: Length of the roll arm of the movable component (distance from the height of the center of gravity of the movable component to the center of the roll rotation of the vehicle)
h _T : Body roll arm length (distance from the center of gravity height of the mass other than movable components to the center height of the roll rotation motion of the body)
b _f : Front wheel tread x: Constant determined by the roll rigidity distribution ratio of the front wheel (0 <x <1 is satisfied)
ΔW _fMAX : Maximum allowable front wheel load travel (slightly less than stationary wheel load per wheel)
φ: Value of the swing displacement of the movable component (current value)
alpha _yT: vehicle lateral acceleration (may use any estimate from the measured value, the steering angle)

The meaning of the calculation formula of the acceleration threshold α _{rel_Limit} described above is as follows.
The acceleration threshold α _{rel_Limit} increases as the allowable maximum value ΔW _fMAX of the front wheel load movement amount increases and the vehicle tread b _f increases.
Further, as the lateral acceleration α _{yT of the} vehicle body increases, the wheel load movement due to the lateral acceleration occurs and the margin decreases, so the acceleration threshold α _{rel_Limit} decreases.
Further, as the swing displacement amount φ of the movable component increases, the center of gravity of the vehicle is greatly inclined toward the inside of the turn, so that the acceleration threshold α _{rel_Limit} increases.

And (E-3) threshold determination or more acceleration threshold alpha _{Rel_Limit,} by comparing the predicted value alpha _rel of the vehicle relative displacement acceleration, exceeds or predicted value alpha _rel is less than the acceleration threshold value alpha _{Rel_Limit} the (permissible limit Whether or not rolling behavior occurs) is determined.
(E-4) Suppression of swing displacement target value If the predicted value α _rel exceeds the acceleration threshold α _{rel_Limit} based on the above judgment result (when rolling behavior exceeding the allowable limit occurs) The swing displacement target value φ * of the movable component is suppressed until the predicted value α _rel becomes lower than the acceleration threshold value α _{rel_Limit} by repeated calculation.
(E-5) Determination of swing displacement target value When the predicted value α _rel falls below the acceleration threshold α _{rel_Limit} due to the above suppression, the swing displacement target value φ * of the movable component is suppressed. The processing is stopped, and the swing displacement target value φ * of the movable component at this time is determined as the final target value.

FIG. 16 is a flowchart of a control program executed by “(F) Drive torque limiter” in FIG. 14, and the processing contents in each step will be described in detail below.
In FIG. 15, before the relative displacement acceleration of the vehicle body that actually causes the rolling behavior exceeding the allowable limit to occur, the vehicle body displacement from the target value is suppressed for the purpose of suppressing the swing displacement target value of the movable component. The threshold value was determined by calculating the predicted value of the relative displacement acceleration.

However, in practice, even if the correct target value is set, the relative displacement acceleration of the vehicle body may exceed the threshold value due to unexpected factors such as disturbance.
From this viewpoint, in FIG. 16, the actual vehicle body relative displacement acceleration is calculated, and if the calculated value exceeds the threshold value, the drive torque command value of the motor 5 is suppressed regardless of the target value of the swing displacement amount. This prevents the vehicle from performing rolling behavior that exceeds the allowable limit.

(F-1) Calculation of relative displacement acceleration of body against unsprung The basic idea is the same as described in section (E-1) above, but here the relative displacement acceleration of vehicle body is obtained using measured values. .
The relative displacement acceleration α _{rel_detec} of the car body is
φ: Measured swing displacement of the movable component (motor rotation angle)
φ _Roll : Predicted roll angle based on suspension stroke (or calculated roll angle based on measured lateral acceleration)
h: Distance from the height of the center of rotation when the body rolls to the height of the center of gravity of the body (effective roll arm length)
Can be calculated by the following equation.
α _{rel_detec} = h {(d2 / dt) φ + (d2 / dt) φ _Roll }
Therefore, the “(F-1) calculation of relative displacement acceleration of the vehicle body relative to the unsprung” processing corresponds to the vehicle body ground displacement acceleration calculation means in the present invention.

(F-2) Calculation of acceleration threshold value Acceleration threshold value α _{rel_Limit} is calculated in the same manner as described in “(E-2) Calculation of acceleration threshold value”.
(F-3) Threshold judgment _{Compare the} above acceleration threshold α _{rel_Limit} with the relative displacement acceleration α _{rel_detec of} the vehicle body. If the relative displacement acceleration α _{rel_detec} of the vehicle body is below the acceleration threshold α _{rel_Limit} ( _{exceeding the} allowable limit) Whether or not rolling behavior occurs) is determined.
(F-4) Suppression of motor drive torque command value Based on the above judgment results, if the relative displacement acceleration α _{rel_detec} of the vehicle body exceeds the acceleration threshold α _{rel_Limit} (if rolling behavior exceeding the allowable limit occurs) ), _Until the relative displacement acceleration α _{rel_detec} of the vehicle body becomes _less than the acceleration threshold α _{rel_Limit} by repeated calculation, the drive torque command value of the motor 5 is suppressed regardless of the swing displacement target value φ *.

When the relative displacement acceleration α _{rel_detec} of the vehicle body falls below the acceleration threshold α _{rel_Limit} due to the above suppression, the control is terminated from the threshold determination of (F-3), and the motor drive torque command value suppression processing (F -4) is stopped, and thereafter, the motor drive torque command value is determined according to the swing displacement target value φ * so as to realize this.

<Operational effects of the third embodiment>
According to the rolling behavior control of the third embodiment described above, basically the same configuration as that of the first embodiment or the second embodiment is followed. The following effects can also be achieved.

That is, as described above with reference to FIG. 15, the predicted value α _rel of the relative displacement acceleration of the vehicle body relative to the unsprung mass calculated from the swing displacement target value of the movable component is the weight of the vehicle body and the movable component, In order to limit the amount of displacement of the movable component (its target value φ *) so as not to exceed a predetermined acceleration threshold α _{rel_Limit} determined from the height of the center of gravity of the movable component and the tread before and after the vehicle,
Even if the rolling behavior control according to the present embodiment is applied to a vehicle of any specification such as a narrow vehicle or a high center of gravity vehicle, it is possible to prevent the vehicle from performing a rolling behavior exceeding an allowable limit.

In addition, as described above with reference to FIG. 16, when the relative displacement acceleration α _{rel_detec} of the vehicle body exceeds the acceleration threshold α _{rel_Limit} , the relative displacement acceleration α _{rel_detec} of the vehicle body decreases to a value equivalent to the acceleration threshold α _{rel_Limit.} To swing and move the movable component,
Even if the relative displacement acceleration α _{rel_detec} of the vehicle body against the unsprung _force exceeds the acceleration threshold α _{rel_Limit} due to some unforeseen factors such as disturbance including cross wind, the vehicle will swing before the rolling behavior exceeds the allowable limit. The displacement acceleration α _{rel_detec} can be dealt with immediately so as to be within the acceleration threshold α _{rel_Limit} , and abnormal rolling behavior due to unexpected factors such as disturbance can be prevented in advance.

1 body
2 wheels
3 Suspension device
3a Upper arm
3b Lower arm
3c strut
4 Rolling member
5 Motor
6 Steering wheel
11 Controller
12 Steering angle sensor
13 Wheel speed sensor
14 Vehicle behavior sensors
14a Yaw rate sensor
14b Front / rear G sensor
14c Lateral G sensor
15 Motor rotation sensor
16 Servo driver
17 battery

Claims (8)

In a rolling behavior control device for a vehicle having a movable component that can be displaced in a direction that suppresses a change in behavior in the rolling direction of the vehicle body,
Steering information detection means for detecting vehicle steering information by the driver;
Based on the steering amount and the steering frequency among the steering information detected by the means, the displacement amount of the movable component corresponding to the steering amount, which is necessary for suppressing the rolling behavior change, becomes smaller as the steering frequency is higher. A vehicle rolling behavior control device, comprising: a movable component displacement amount setting unit configured to set as described above. The rolling behavior control device for a vehicle according to claim 1,
The movable component displacement amount setting means sets the steady gain defined as the displacement amount of the movable component per unit steering amount at the time of steady steering where the steering frequency is less than the steady steering determination value, and the gain is set more than the steady value. The amount of displacement of the movable component relative to the steering amount is set by using a low-pass filter that can change the cutoff frequency, which is the steering frequency when the value decreases by a value, according to the vehicle state. Vehicle rolling behavior control device. In the rolling behavior control device for a vehicle according to claim 2,
A vehicle speed calculation means for calculating the vehicle speed is provided.
A rolling behavior control device for a vehicle, characterized in that the steady gain is changed according to a vehicle speed calculated by the means or according to both a vehicle speed and a difference value of the vehicle speed. In the rolling behavior control device for a vehicle according to claim 2 or 3,
A rolling behavior control device for a vehicle, wherein the cutoff frequency is set low as the absolute value of the steady gain is set large, and the cutoff frequency is set high as the absolute value of the steady gain is set small. . In the rolling behavior control device for a vehicle according to claim 2 or 3,
The lower limit of the cutoff frequency is determined by providing a constraint that the gain of the wheel load movement amount with respect to the steering amount is equal to or less than the gain at the time of steady steering in a steering frequency region of a relatively low predetermined steering frequency or less. ,
By providing a constraint that the phase shift of the wheel load movement amount with respect to the steering amount falls within the range of the predetermined phase advance amount from the predetermined phase delay amount in a steering frequency region of a relatively high predetermined steering frequency or less, Defining an upper limit of the cutoff frequency;
A rolling behavior control device for a vehicle, wherein the cutoff frequency is set between an upper limit and a lower limit of the cutoff frequency. The rolling behavior control device for a vehicle according to claim 1,
When setting the displacement amount of the movable component,
Phase lag in the frequency response of the displacement amount with respect to the steering amount, and the phase lag in the frequency response of one of the lateral acceleration of the vehicle body with respect to the steering amount and the lateral acceleration of the component with respect to the steering amount, in the steering frequency region below the predetermined steering frequency A rolling behavior control device for a vehicle, wherein a displacement amount of the movable component is set so that a difference from the amount is minimized. The rolling behavior control device for a vehicle according to claim 1,
When setting the displacement amount of the movable component,
The predicted relative displacement acceleration value of the vehicle body with respect to the unsprung amount calculated from the target value of the displacement amount of the movable component includes the weight of the vehicle body and the movable component, the height of the center of gravity of the vehicle body and the movable component, and the tread before and after the vehicle. A rolling behavior control device for a vehicle, wherein a limit is set to a displacement amount of the movable component so as not to exceed a predetermined acceleration threshold value determined from the following. The rolling behavior control device for a vehicle according to claim 1 or 7,
A vehicle body ground displacement acceleration calculating means for calculating the relative displacement acceleration of the vehicle body with respect to the unsprung portion is provided,
When the ground displacement acceleration of the vehicle body calculated by the means exceeds a predetermined acceleration threshold value determined from the height of the center of gravity of the vehicle body and the movable component and the tread before and after the vehicle, the vehicle body ground acceleration is equal to the acceleration threshold value. A rolling behavior control device for a vehicle, characterized in that the movable component is controlled to be displaced in such a manner as to decrease to a value of.

Images