JP4968006B2 - Suspension control device - Google Patents

Description

  The present invention relates to a so-called preview control of a suspension.

Examples of a suspension control device that performs preview control are described in Patent Documents 1 to 4.
In the suspension control device described in Patent Document 1, a road surface sensor is provided in front of the front wheels of the vehicle, and the damping characteristics of the shock absorbers for the front wheels and the rear wheels are controlled based on the detection values of the road surface sensors. The control command is output after a delay time determined by (a) the distance between the position where the road surface sensor is provided and the wheel to be controlled, (b) the traveling speed of the vehicle, and (c) the control delay time. .
In the suspension control device described in Patent Document 2, a front wheel up / down behavior sensor for detecting up / down behavior is provided for the front wheel, and the damping characteristic of the shock absorber of the rear wheel is controlled based on a detection value by the front wheel up / down behavior sensor. The The control signal for controlling the damping characteristics of the rear wheel shock absorber is both a control signal based on the value detected by the front wheel vertical motion sensor and a control signal (preview control signal) obtained by delaying the phase of the signal. The ratio of the control signal delayed in phase is high when the vehicle traveling speed is smaller than a predetermined value. When the vehicle traveling speed is larger than the predetermined value, the ratio of the control signal delayed in phase is the traveling speed. It is reduced with the increase of. As a result, it is possible to create a control signal that changes in the same phase as the actual vertical behavior of the rear wheel, and it is possible to satisfactorily suppress the vertical vibration of the rear wheel. It is known that the vertical behavior of the rear wheels does not occur in the same phase as when the actual vertical behavior of the front wheels is delayed by a time determined by the vehicle speed etc., but occurs by a phase advanced from that delayed by a time determined by the vehicle speed etc. It has been. This is because the vehicle body is a rigid body, so that the influence of the vertical behavior on the front wheel side affects the vertical behavior of the rear wheel. On the other hand, it is known that the phase advance amount is smaller when the vehicle speed is large than when the vehicle speed is small. As a result, when the vehicle speed is high, if the ratio of the control signal delayed in phase is smaller than when the vehicle speed is low, a control signal that changes in phase close to the actual vertical behavior of the rear wheels can be obtained.
In the suspension control device described in Patent Document 3, a rear wheel control signal is created based on a detection value by a sprung acceleration sensor provided on the front wheel side. In this case, the detection value by the sprung acceleration sensor is filtered to generate a control signal, but filters having different phase characteristics are used based on the traveling speed of the vehicle. As a result, regardless of the vehicle speed, the phase of the control signal can be brought close to the phase of the actual vertical behavior of the rear wheels.
In the suspension control device described in Patent Document 4, the preview total gain is determined by the longitudinal acceleration, lateral acceleration, and traveling speed of the vehicle. When a control response delay occurs due to high speed traveling, or when the traveling trajectory of the front and rear wheels differs due to turning or the like, the preview gain is made variable and the control output is lowered.
JP-A-5-262118 JP-A-7-237419 JP 7-186660 A Japanese Patent Laid-Open No. 7-205629

  An object of the present invention is to ensure that preview control is performed well even while the vehicle is turning.

Means and effects for solving the problem

The suspension device according to claim 1 is provided on the front wheel side of the vehicle, includes at least one sensor for detecting a vertical behavior of a front wheel side portion, and the vehicle is based on a detection value by the at least one sensor. A suspension control device for controlling a suspension of a rear wheel that is a control target wheel of the vehicle, wherein when the vehicle turns, an overlap between a road surface on which the tire of the front wheel passes and a road surface on which the tire of the rear wheel is predicted to pass A trajectory corresponding overlap acquisition unit that predicts a value smaller than a small value when the difference between the front wheel trajectory and the rear wheel trajectory is large, and when the value representing the overlap is small than, viewed contains a gain determination section that determines the gain to be used for control of the suspension to a small value, and the track corresponding overlapping acquisition unit, the front wheel A lap ratio, which is a value obtained by dividing a dimension in the width direction of the vehicle by a width of the tire of the rear wheel, where a road surface through which the tire passes and a road surface in which the tire of the rear wheel is predicted to overlap is overlapped. A lap rate acquisition unit to be acquired is included.
In the suspension device described in this section, so-called preview control (also referred to as preview control) is performed, and the suspension of the rear wheel that is the control target wheel is controlled based on the vertical behavior of the front wheel side portion. . However, when the vehicle is turning, if the road surface through which the front wheels pass does not overlap the road surface where the rear wheels are expected to pass, that is, if the rear wheel passes through a different road surface than the road surface through which the front wheels have passed, When the control is performed, the vibration in the vertical direction on the rear wheel side cannot be satisfactorily suppressed, or on the contrary, the ride comfort may be deteriorated (hereinafter sometimes referred to as an adverse effect). is there. On the other hand, it is possible to prevent the preview control from being performed during a turn, but even in that case, the vibration in the vertical direction on the rear wheel side may not be satisfactorily suppressed.
Therefore, in the suspension device described in this section, the overlap between the road surface on which the front wheel tire passes and the road surface on which the rear wheel tire is predicted to pass is predicted. Decrease to a small value. When the overlap is small, the effect of the preview control is less than when the overlap is large, and on the contrary, the possibility of an adverse effect is increased, so the gain is reduced. As a result, it is possible to suppress adverse effects when the same preview control is performed when the overlap is small or the overlap is large. Also, since preview control is performed even when the vehicle is turning, it is possible to suppress the vertical vibration of the rear wheels compared to the case where the preview control is not performed.
When the vehicle is almost straight, there is usually much overlap between the road surface through which the front wheels pass and the road surface where the rear wheels are expected to pass, so there is no need to acquire the gain by acquiring the overlap. It is considered low.
On the other hand, since there is not always much overlap during turning, it is appropriate to determine the gain based on the amount of overlap. For example, when the absolute value of the steering angle of the steering wheel of the vehicle is greater than or equal to a set value, it can be considered that the vehicle is turning, and the gain is determined according to the overlap. Whether the vehicle is turning or going straight is not limited to the absolute value of the steering wheel steering angle, but the absolute value of the steering amount of the steering member (for example, the absolute value of the steering angle of the steering wheel), The determination can also be made based on physical quantities representing the turning state of the vehicle such as the turning radius, the lateral acceleration, and the absolute value of the yaw rate.
The sensor detects the vertical behavior of the front wheel side portion. For example, the sensor detects the vertical behavior of the unsprung member of the front wheel or the vertical behavior of the sprung member. Or detecting the distance between the sprung and unsprung springs. Also, the sensor for detecting the behavior in the vertical direction may be a vertical acceleration sensor for detecting the vertical acceleration of the sprung member or the unsprung member, or a stroke sensor for detecting the stroke between the sprung springs. Can do. Further, the number of sensors for detecting the vertical behavior of the front wheel side portion may be one or two or more.
The road surface on which the front tire passes and the road surface on which the rear tire is predicted to pass are acquired based on the wheel trajectory and the wheel tire width. The trajectory of a wheel is represented by a line in the present specification, and a trajectory is a set of contact points between the center surface of the wheel (the surface passing through the center line in the width direction of the tire) and the road surface, A set of arbitrary points on the rotation center line can be traced. For example, a contact point with the road surface of the center plane of the wheel (tire) or a turning radius of an arbitrary point on the rotation center line of the wheel can be set as a trace. it can. In addition, the front wheel trajectory and the rear wheel trajectory may be represented by intermediate trajectories between the left and right wheels on the front wheel side and the rear wheel side, respectively. This intermediate trajectory is a trajectory of the center point on the front wheel side in the width direction of the vehicle (a surface extending in a vertical direction including a front-rear direction line passing through the center of gravity of the vehicle when the vehicle is traveling straight on a horizontal road surface, and The trajectory of the intersection with the line passing through the center line of the left and right front wheel axle), the trajectory of the center point of the rear wheel side (the trajectory of the intersection of the surface extending in the vertical direction and the line passing through the center line of the axle of the left and right rear wheels) It can also be expressed.
In addition, the overlap between the road surface on which the front wheel tire passes and the road surface on which the rear wheel tire is predicted to pass is expressed by the area of the road surface of the overlap portion when the vehicle has traveled a predetermined travel distance, It can be represented by the length of the width direction of the vehicle of the part. In addition, the road surface area of the above-mentioned overlapping portion is represented by a value divided by the area of the road surface through which the rear wheel tire passes, or the length in the width direction of the overlapping portion is represented by a value divided by the tire width. Can do.
Patent Document 4 describes that the preview gain is determined according to the lateral force, and that the preview gain is decreased when the traveling trajectories of the front and rear wheels are different. There is no description of making the gain smaller than the case where there are many cases. There is a positive correlation between the degree of overlap and the expected effect of the preview control. If the gain is determined according to the degree of overlap as in the suspension device according to the present invention, Thus, preview control can be effectively performed while avoiding the occurrence of adverse effects .

Patentable invention

  In the following, the invention that is claimed to be claimable in the present application (hereinafter referred to as “claimable invention”. The claimable invention is at least the “present invention” to the invention described in the claims. Some aspects of the present invention, including subordinate concept inventions of the present invention, superordinate concepts of the present invention, or inventions of different concepts) will be illustrated and described. As with the claims, each aspect is divided into sections, each section is numbered, and is described in a form that cites the numbers of other sections as necessary. This is merely for the purpose of facilitating understanding of the claimable invention, and is not intended to limit the set of components constituting the claimable invention to those described in the following sections. In other words, the claimable invention should be construed in consideration of the description accompanying each section, the description of the embodiments, etc., and as long as the interpretation is followed, another aspect is added to the form of each section. In addition, an aspect in which constituent elements are deleted from the aspect of each item can be an aspect of the claimable invention.

(1) A rear wheel that is provided on the front wheel side of the vehicle and includes at least one sensor that detects the behavior in the vertical direction of the front wheel side portion, and is a wheel to be controlled of the vehicle based on a detection value by the at least one sensor. A suspension control device for controlling a suspension of a wheel,
Predicts the overlap between the road surface through which the front wheel tire passes and the road surface through which the rear wheel tire is predicted to pass, and if the overlap is small, the gain used for controlling the suspension is smaller than when there are many. suspension control device characterized by comprising a gain determining unit for determining the.
(2) The gain determination unit includes an overlap acquisition unit that acquires a value representing an overlap between a road surface on which the front wheel tire passes and a road surface on which the rear wheel tire is predicted to pass. Suspension control device.
The value representing the overlap is a value corresponding to the overlap one-to-one, and is a value representing the degree of overlap. For example, it can be acquired based on the difference between the trajectories of the front wheels and the rear wheels and the widths of the tires of the front wheels and the rear wheels.
For example, as described above, when the travel distance of the vehicle is a predetermined set distance, the area of the overlapping portion with respect to the area of the road surface through which the tire of the rear wheel passes, or the length of the overlapping portion in the width direction of the vehicle Express the value by dividing the road surface area of the overlapping part by the area of the road surface through which the rear wheel tire passes, or by dividing the width of the overlapping part in the width direction by the tire width. Etc.
(3) The gain determination unit includes a trajectory corresponding overlap acquisition unit that determines that the value representing the overlap is a smaller value than when the difference between the trajectory of the front wheel and the trajectory of the rear wheel is small. The suspension control device according to item (1) or (2) .
When the difference between the trajectory of the front wheel and the trajectory of the rear wheel is large, it can be assumed that there is less overlap than when the trajectory is small. In this case, it is necessary to consider the widths of the front and rear tires, but if the front and rear tire widths are known in advance, it is necessary to directly consider the tire width. Is low, and the degree of overlap can be obtained from the difference in trajectory.
The difference between the trajectory of the front wheel and the trajectory of the rear wheel is the difference between the trajectories of the wheels on the same side of the vehicle (for example, the difference between the trajectory of the right front wheel and the trajectory of the right rear wheel, the trajectory of the left front wheel and the trajectory of the left rear wheel The difference between the locus on the front wheel side and the locus on the rear wheel side may also be used.
(4) The suspension control device according to (3), wherein the trajectory corresponding overlap acquisition unit includes a trajectory difference acquisition unit that acquires a difference between a trajectory of a front wheel and a trajectory of a rear wheel on the same side in the vehicle width direction.
(5) When the difference between the locus of the front wheel and the locus of the rear wheel is equal to or less than the width of the tire, the locus corresponding overlap acquisition unit has the overlap, but the difference in the locus is larger than the width of the tire. In some cases, the suspension control device according to (3) or (4), including an overlap presence / absence determination unit that determines that there is no overlap.
If the difference in trajectory (the difference between the turning radii is the value obtained by subtracting the smaller turning radius from the larger turning radius) is less than or equal to the tire width, it can be estimated that there is an overlap, and is greater than the tire width. It can be estimated that there is no overlap.
The width of the tire can be the width of the tire when the front wheel and the rear wheel are the same, and can be the average of these when the front wheel and the rear wheel are different.
(6) The trajectory-corresponding overlap acquisition unit calculates the length in the width direction of the vehicle of the portion predicted to overlap the road surface on which the front wheel tire passes and the road surface on which the rear wheel tire is predicted to pass. to (3) no claim including overlapping ratio acquisition unit for acquiring overlap ratio is a value obtained by dividing the width of the tire of the rear wheel (5) suspension controller according to any one of claims (claim 1).
When the lap rate is high, the ratio of the overlapping portion with respect to the road surface on which the rear tires pass is higher than when the lap rate is low, and it can be considered that the preview control is more effective. Therefore, when the wrap rate is high, it is appropriate to set the gain to a larger value than when the lap rate is low.
The lap rate can be obtained as a value obtained by dividing a value obtained by subtracting a track difference from the width of the tire and dividing the tire width by the tire width when the tire width is the same between the front wheel and the rear wheel.
(7) The trajectory corresponding overlap acquisition unit includes a turning radius acquisition unit that acquires the trajectory as a turning radius of each of the front and rear wheels when the vehicle is turning. The suspension control device according to any one of claims 2 (Claim 2) .
The turning radius can be adopted as the trajectory, and when the difference in turning radius (the value obtained by subtracting the smaller one from the larger turning radius) is large, it can be considered that the difference in the trajectory is larger than when the turning radius is small.
Further, when the slip of the wheel is large, it is desirable not to obtain the turning radius. This is because when the slip is large, it is difficult to accurately obtain the turning radius. When the slip is large, the gain can be set to 0 without acquiring the overlap or acquiring the gain according to the overlap.
(8) The control command value for the suspension of the rear wheel is determined from the time when the detected value by the at least one sensor is acquired, from the margin time determined by the wheel base of the vehicle and the traveling speed of the vehicle. (1) to claim includes a preview control unit time obtained by subtracting the output after a lapse of (7) the suspension controller according to any one of claims (claim 3).
If the control command value is output after the time obtained by subtracting the control delay time from the allowance time, the vertical vibration of the wheel to be controlled can be satisfactorily suppressed.
(9) Gain gradual reduction in which the gain determining unit determines the gain to be a smaller value as the value representing the overlap becomes smaller when the value representing the overlap is smaller than a predetermined first set value The suspension control device according to any one of (1) to (8), including a section (claim 4) .
(10) In the item (9), the gain determining unit includes means for determining the gain to be a value equal to or greater than a predetermined set gain when a value representing the overlap is equal to or greater than the first set value. The suspension control device described.
While the vehicle is turning, the lap ratio is 1, that is, the road surface through which the front wheel tire passes and the road surface through which the rear wheel tire do not become the same, so that preview control may not be performed. it can.
On the other hand, for example, when the lap rate is 0.8 (first set value) or more, the gain can be set to a value greater than or equal to the set gain, assuming that preview control is effective. The gain may be a fixed value or a variable value, but in any case, the gain is not made smaller than the set gain.
Further, when the wrap rate is smaller than 0.8, the gain becomes a small value as the wrap rate decreases. The gain may be a continuously decreasing value or a gradually decreasing value as the lap rate decreases, and if it is a continuously decreasing value, it is a linearly decreasing value. Or a value that decreases curvilinearly.
(11) The gain determination unit includes a 0 determination unit that sets the gain to 0 when a value representing the overlap is equal to or smaller than a predetermined second set value smaller than the first set value. Item to (10)
The suspension control device according to any one of the items.
For example, the second set value may be a value indicating that there is no overlap, and the gain is set to 0 when there is no overlap. Also, the second set value can be a value that is considered to have little overlap and the effect of preview control is very small. In this case, the gain is set to 0 even if there is an overlap.
(12) The gain determination unit includes any one of the items (1) to (11), wherein the gain determination unit includes a vehicle speed corresponding gain determination unit that determines the gain to a smaller value than when the vehicle traveling speed is large.
Suspension control device described in one.
During turning, the rear wheel suspension can be more appropriately controlled by setting the gain to a smaller value when the vehicle traveling speed is large than when the vehicle is traveling at a low speed. The gain may be a value that is decreased when the vehicle speed is higher than the vehicle speed determined by the control delay of the suspension to be controlled.
In the suspension device described in this section, the gain is determined by both the turning state and the vehicle speed.
(13) The suspension of the wheel to be controlled includes a vertical force generator provided between an unsprung member and a sprung member that holds the wheel to be controlled, and generates a force in a substantially vertical direction. Any one of the items (1) to (12), wherein the control device includes a vertical force control device that controls the vertical force generation device based on a value detected by the at least one sensor and the gain. A suspension control device according to claim 5 (Claim 5) .
The vertical force generator is provided between the unsprung member and the sprung member, and generates a force in the vertical direction. The vertical force is a force having a vertical component, and may be a vertical direction of the vehicle or a direction slightly inclined with respect to the vertical direction of the vehicle.
The direction of the vertical force generated by the vertical force generator is determined by the body of the unsprung member, the structure of the connecting portion with respect to the wheel, the state of connection between the vertical force generator and the unsprung member, and the like. If the unsprung member is connected so as to be rotatable in the vertical direction, and the movement (or rotation) in the front-rear direction and the width direction is not permitted, the generated force is considered to be a vertical force. be able to.
If the vertical force generator is controlled based on the detection value and gain by the sensor, the vertical vibration of the wheel to be controlled can be satisfactorily suppressed. The vertical force can be a damping force or an elastic force, as will be described later.
(14) The vertical force generating device includes a damping force generating device that generates a damping force, and the vertical force control device corresponds to the wheel to be controlled based on a detection value by the at least one sensor. Of the absolute speed in the vertical direction of the sprung member, the absolute speed in the vertical direction of the unsprung member holding the wheel to be controlled, and the relative speed in the vertical direction of the sprung member and the unsprung member A target damping force determining unit that estimates one or more speeds and at least a value determined based on the estimated one or more speeds and the gain, and controls the damping force generator. And a damping force control unit that outputs the target damping force determined by the target damping force determination unit.
A damping force is generated by the control of the vertical force generator, thereby suppressing vertical vibrations. The damping force may be a magnitude corresponding to the absolute speed of the unsprung member, whether it is a magnitude corresponding to the absolute speed of the sprung member in the vertical direction or a magnitude corresponding to the relative speed between the sprung member and the unsprung member. It may be good. When determining the magnitude of the damping force or determining the damping coefficient, two or more of these speeds may be considered.
Further, the vertical force generated in the vertical force generator can be a magnitude including these two or more damping forces. For example, the magnitude can be controlled to include a damping force corresponding to the absolute speed of the sprung member and a damping force corresponding to the absolute speed of the unsprung member.
When the absolute speed of the sprung member is acquired based on the detection value by the sensor, when the absolute speed of the unsprung member is acquired, the relative speed of the unsprung member may be acquired, etc. The value detected by the sensor and the acquired speed are not necessarily the same.
(15) The vertical force generator includes an elastic force generator that generates an elastic force, and the vertical force controller corresponds to the wheel to be controlled based on a detection value by the at least one sensor. One of a displacement in the vertical direction of the sprung member, a displacement in the vertical direction of the unsprung member that holds the control target wheel, and a relative displacement in the vertical direction between the sprung member and the unsprung member. A target elastic force determination unit that sets a target elastic force to a value determined by at least one estimated displacement and the gain, and the elastic force generator is controlled and determined by the target elastic force determination unit The suspension control device according to item (13) or (14), further including an elastic force control unit that outputs the target elastic force.
By controlling the vertical force generating device, an elastic force is generated, thereby suppressing the vertical vibration of the wheel to be controlled.
Further, the vertical force generated in the vertical force generator can have a magnitude including these two or more elastic forces. For example, the magnitude can be controlled to include an elastic force corresponding to the displacement of the sprung member and an elastic force corresponding to the displacement of the unsprung member. Further, the vertical force can be controlled so as to be the sum of the damping force and the elastic force.
(16) The vertical force generator includes an elastic member having one end connected to one of the unsprung member and the sprung member and the other end connected to the other, and the vertical force The control device includes a drive source that elastically deforms the elastic member against a restoring force, and includes an elastic deformation amount control unit that controls the vertical force by changing an elastic deformation amount of the elastic member (13 The suspension control device according to any one of claims 1 to 15 (Claim 6 ).
(17) The elastic member is a generally L-shaped bar including a first bar portion extending in the front-rear direction of the vehicle and a second bar portion extending in the width direction, and the drive source is the L The suspension control device according to item (16), including an electric motor that rotates one of the first bar portion and the second bar portion of the character-shaped bar about its axis.
(18) The elastic member is a rod composed of one of a first bar portion extending in the front-rear direction of the vehicle and a second bar portion extending in the width direction, and the drive source is bent to the rod. The suspension control device according to (16) or (17), including an electric motor that applies a moment.
The elastic member may be an L-shaped member or a member extending in a straight line when viewed from above and below. In other words, the shape may be curved in the vertical direction. (19) A suspension spring is provided between the unsprung member and the sprung member in parallel with the vertical force generator as an elastic member different from the elastic member included in the vertical force generator. The suspension control device according to any one of (13) to (18).
Between the unsprung member and the sprung member, an elastic member of the vertical force generator and a suspension spring which is an elastic member different from the elastic member are provided. As described above, the elastic member included in the vertical force generator is elastically deformed by the drive source to generate the vertical force, but the suspension spring is not elastically deformed by the drive source. It is elastically deformed by a load applied to the.
The load applied to the wheel is received by the suspension spring and the elastic member. However, when the elastic member is not elastically deformed when the drive source is not in operation, no force is generated in the elastic member, and the load is received by the suspension spring. This state is the reference state of the drive source of the vertical force generator. The reference state is determined by the load applied to the wheel. When the load is large, the distance between the sprung springs is shorter than when the load is small.
For example, when the electric motor of the drive source is rotated in one direction from the reference state, the distance between the sprung member and the unsprung member increases. The direction of the elastic force of the suspension spring and the direction of the elastic force of the elastic member are the same. When the distance between the unsprung springs increases and the elastic force of the suspension spring decreases, the elastic force of the elastic member increases, and the sum of these is maintained at a magnitude corresponding to the load.
When the electric motor is rotated in the other direction from the reference state, the distance between the sprung member and the unsprung member is reduced. The direction of the elastic force of the suspension spring is opposite to the direction of the elastic force of the elastic member. When the distance between the unsprung springs becomes smaller and the elastic force of the suspension spring becomes larger, the elastic force in the opposite direction of the elastic member becomes larger.
When the elastic member is an L-shaped bar, when one of the first bar portion and the second bar portion (hereinafter referred to as the shaft portion) is rotated around the axis, the other (hereinafter referred to as the arm portion). ) Is rotated, thereby changing the distance between the sprung sprung. Also, when the shaft portion is twisted, the torsional moment applied to the shaft portion (same as the torque applied by the electric motor) and the bending moment applied to the arm portion become equal, so the vertical force of the magnitude determined by it is Added to the unsprung member.
When the elastic member is a linear rod, the torque applied to the rod by the electric motor is equal to the bending moment, so that a vertical force of a magnitude determined by the torque is applied to the unsprung member.
Comparing the case where the elastic member is an L-shaped bar and the case where it is a linear rod, in either case, the vertical direction according to the magnitude at which the torque and bending moment applied to the elastic member by the electric motor are equal. The same is true for the point at which the force is generated (assuming that the torsional stress and the bending stress reach the allowable stress at the same time).
Further, when the elastic member is an L-shaped bar, the arm portion is rotated by rotation around the axis of the shaft portion, whereas when the elastic member is a linear rod, the linear rod is It is rotated directly by an electric motor. Therefore, when the length of the linear rod is the same as the length of the arm portion of the L-shaped bar, the L-shaped bar is the driving source compared to the linear rod. Can be provided in a portion away from the wheel of the vehicle body (sprung member).
(20) A suspension control device for controlling a suspension of a wheel to be controlled behind a detection target portion of at least one sensor based on a detection value by at least one sensor provided in the vehicle,
When the overlap between the road surface through which the front wheel tire passes and the road surface through which the rear wheel tire is predicted to pass is predicted, and the value representing the overlap is equal to or greater than a predetermined first set value, A gain determination for determining a gain used for controlling the suspension to be a value equal to or greater than a predetermined set gain and determining the gain to be 0 when the gain is equal to or less than a second set value smaller than the first set value. A suspension control device comprising a portion.
The suspension control device described in this section can employ the technical features described in any one of the items (1) to (19).
When the value representing the overlap is smaller than the first set value and greater than the second set value, the gain can be a value that decreases as the value representing the overlap decreases.
(21) A suspension control device for controlling a suspension of a wheel to be controlled behind a detection target portion of the at least one sensor based on a detection value by at least one sensor provided in the vehicle,
When the vehicle is traveling straight ahead, the gain used for controlling the suspension is determined to be a smaller value when the vehicle traveling speed is high, and when the vehicle is turning. Includes a gain determining unit that determines the gain to be a smaller value when the absolute value of the steering angle of the steering wheel of the vehicle is large.
The suspension control device described in this section can employ the technical features described in any one of the items (1) to (20).
(22) The vehicle includes at least one road surface sensor that detects an uneven state of the same detection target portion on the road surface, and is behind the detection target portion of the road surface based on a detection value by the at least one road surface sensor. A suspension control device for controlling the suspension of the wheel to be controlled in
The overlap between the detection target portion of the road surface and the road surface on which the tire of the wheel to be controlled is predicted to pass is predicted, and the gain used for controlling the suspension is set to a smaller value than when the overlap is small. A suspension control device including a gain determining unit for determining .
The suspension control device described in this section can employ the technical features described in any one of the items (1) to (21).
(23) The vehicle includes at least one road surface sensor that detects the uneven state of the same detection target portion on the road surface, and is located behind the detection target portion of the road surface based on a detection value by the at least one road surface sensor. A suspension control device for controlling the suspension of the wheel to be controlled in
Control of the suspension based on the relative position of the detection target portion with respect to the road surface predicted to pass the tire is predicted by predicting the detection target portion of the road surface and the road surface predicted to pass the tire of the wheel to be controlled. A suspension control device including a gain determining unit that determines a gain used in the vehicle.
The suspension control device described in this section can employ the technical features described in any one of the items (1) to (22).
For example, when the detection target portion is located at the center portion in the width direction of the road surface through which the tire of the control subject wheel passes, the gain can be set to a larger value than when the detection target portion is located at a portion away from the center.

Hereinafter, a suspension system including a suspension control apparatus according to an embodiment of the present invention will be described.
2 and 3, a suspension 16 is provided between the front and rear wheels 12FR, FL, RL and RR of the vehicle and the vehicle body 14 as a sprung member.
The suspension 16 includes coil springs 20FR, FL, RL, and RR as suspension springs, shock absorbers (hereinafter sometimes abbreviated as "absorbers") 22FR, FL, RL, RR, vertical force generators 24FR, FL, RL, Includes RR etc. Hereinafter, the subscripts FR, FL, RL, and RR indicating the wheel position are described when it is necessary to distinguish them, but are not described when it is not necessary to distinguish them. Further, when it is necessary to distinguish the front wheel side and the rear wheel side, suffixes F and R may be added. Furthermore, in the case of representing any one of the front, rear, left, and right wheels, a suffix ij (i = F, R j = R, L) may be attached to indicate that the wheels are the same.
As shown in FIG. 3, the suspension 16 includes a plurality of suspension arms such as a first upper arm 40, a second upper arm 42, a first lower arm 44, a second lower arm 46, and a toe control arm 48. In this embodiment, the suspension 16 is a multi-link type. These five suspension arms 40, 42, 44, 46, 48 are rotatably connected to the vehicle body 14 at one end, and an axle carrier 50 that holds the wheel 12 so as to be relatively rotatable at the other end. It is connected to the pivotable. The axle carrier 50 is movable relative to the vehicle body 14 in the vertical direction along a predetermined trajectory by the five arms 40, 42, 44, 46, and 48.

As shown in FIG. 4, the shock absorber 22 is basically incapable of relative movement in the vertical direction and swings between the vehicle body 14 as the sprung member and the second lower arm 46 as the unsprung member. Connected as possible. The shock absorber 22 includes a damping characteristic control device 56, and the damping characteristic can be continuously controlled.
The shock absorber 22 includes a housing 60 and a piston 62. The shock absorber 22 is connected to the second lower arm 46 in the housing 60, and is connected to the vehicle body 14 via a mount portion 54 at a piston rod 64 of the piston 62. The piston rod 64 is brought into sliding contact with the lid portion 66 of the housing 60 via a seal 68 at the intermediate portion.
As shown in FIG. 5, the housing 60 includes an outer cylinder 71 and an inner cylinder 72, and a space between them is a buffer chamber 74. The above-described piston 62 is fitted inside the inner cylinder 72 in a liquid-tight and slidable manner, and the inner side of the inner cylinder 72 is partitioned into an upper chamber 75 and a lower chamber 76.
The piston 62 is provided with a plurality of concentric connection passages 77 and 78 that connect the upper chamber 75 and the lower chamber 76 (two of them are shown in FIG. 5). The valve plate 79 disposed on the lower surface of the piston 62 and the valve plates 80 and 81 disposed on the upper surface are respectively supported by stepped portions and nuts provided on the piston rod 64.
The valve plate 79 does not cover the opening of the connection passage 78 on the outer peripheral side, but is large enough to cover the opening of the connection passage 77 on the inner peripheral side, and the hydraulic pressure difference between the upper chamber 75 and the lower chamber 76 is a set value. Thus, when a force greater than the valve opening pressure is applied, the fluid is deflected, and the flow of hydraulic fluid from the upper chamber 75 to the lower chamber 76 is allowed. A leaf valve 84 is configured by the valve plate 79, the opening of the connection passage 77, and the like.
The valve plates 80 and 81 are provided so as to overlap in the vertical direction. The valve plate 80 covers the opening of the connection passage 78, and an opening is formed in a portion corresponding to the opening of the connection passage 77 of the valve plates 80 and 81. When the hydraulic pressure difference between the lower chamber 76 and the upper chamber 75 exceeds a set value and a force greater than the valve opening pressure is applied, the fluid is deflected, and the flow of hydraulic fluid from the lower chamber 76 toward the upper chamber 75 is allowed. A leaf valve 86 is configured by the valve plates 80 and 81, the opening of the connection passage 78, and the like.
A base valve body 88 having a leaf valve is provided between the lower chamber 76 and the buffer chamber 74 described above.

As shown in FIG. 4, the damping characteristic control device 56 includes an electric motor 90, a motion conversion mechanism 91 that converts the rotation of the electric motor 90 into a linear motion, an adjustment rod 92, and the like. A through hole 94 extending in the axial direction of the piston rod 64 is formed in the piston rod 64, and an adjustment rod 92 is disposed in the through hole 94. The adjustment rod 92 is connected to the output member of the motion conversion mechanism 91 at the upper end, and is linearly moved relative to the piston rod 64 as the electric motor 90 rotates. The rotation angle of the electric motor 90 is detected by a motor rotation angle sensor 96.
As shown in FIG. 5, the through hole 94 has a stepped shape, and the upper part is a large diameter part 98 and the lower part is a small diameter part 100. The small diameter portion 100 opens to the lower chamber 76, and the large diameter portion 98 communicates with the upper chamber 75 via the connection passage 102. The upper chamber 75 and the lower chamber 76 are communicated with each other by a through hole 94 and a connection passage 102.
On the other hand, the outer diameter of the intermediate portion of the adjustment rod 92 is smaller than the inner diameter of the large diameter portion 98 and larger than the inner diameter of the small diameter portion 100, and the outer diameter of the lower end portion 106 is gradually reduced as it goes downward (for example, The shape of the lower end portion 106 can be a conical shape). The adjustment rod 92 is disposed in a state where the intermediate portion is located at the large diameter portion 98 and the lower end portion 106 is located near the step portion between the large diameter portion 98 and the small diameter portion 100.
A clearance (opening area) between the outer peripheral surface of the lower end portion 106 of the adjustment rod 92 and the peripheral edge 107 of the step portion of the large diameter portion 98 and the small diameter portion 100 is a piston rod 64 of the adjustment rod 92 (lower end portion 106). As the relative position changes with respect to, it is continuously changed. The relative position of the adjustment rod 92 is known from the rotation angle of the electric motor 90. The opening area of the variable throttle 108 is controlled by the control of the electric motor 90, and the flow control valve (variable throttle) 108 is controlled by the inner peripheral surface of the through hole 94 including the lower end 106 of the adjustment rod 76 and the peripheral edge 107. Composed.
A seal member 109 is provided above the portion of the through hole 94 where the connection passage 102 is connected, and the space between the inner peripheral surface of the through hole 94 and the outer peripheral surface of the adjustment rod 92 is kept liquid-tight.

For example, when a force in a direction to bring the sprung member 14 and the second lower arm 46 (wheel 12) closer, that is, a force in a direction to move the piston 62 downward relative to the housing 60 is applied. The hydraulic pressure in the lower chamber 76 is increased.
Part of the hydraulic fluid in the lower chamber 76 flows to the upper chamber 75 through the variable throttle 108 in the through hole 94. Further, when the force corresponding to the hydraulic pressure difference applied to the valve plate 80 (81) becomes equal to or higher than the valve opening pressure and the leaf valve 86 is switched to the open state, the working fluid flows into the upper chamber 75 via the connection passage 78. Further, it flows out to the buffer chamber 74 through the leaf valve of the base valve body 88. The damping characteristic of the shock absorber 22 is mainly determined by the opening area of the variable diaphragm 108. The resistance force applied when the hydraulic fluid flows through the variable throttle 108 increases as the opening area of the variable throttle 108 decreases when the flow velocity is the same. In the present embodiment, the opening area of the variable diaphragm 108 is controlled by controlling the electric motor 90 so that the damping coefficient becomes a desired magnitude in the entire shock absorber.

On the contrary, when a force in a direction to separate the sprung member 14 and the second lower arm 46 (wheel 12), that is, a force in a direction to move the piston 62 relatively upward with respect to the housing 60 is applied. The hydraulic pressure in the upper chamber 75 increases.
A part of the hydraulic fluid in the upper chamber 75 flows to the lower chamber 76 through the variable throttle 108 of the through hole 94. Further, when the force applied to the valve plate 79 becomes equal to or higher than the valve opening pressure, the leaf valve 84 is switched to the open state, and the working fluid flows into the lower chamber 76 through the connection passage 77. A part of the working fluid in the buffer chamber 74 flows in through the leaf valve of the base valve body 88. The attenuation characteristic is controlled by controlling the opening area of the variable stop 108.
If the moving speed of the piston 62 (the flow rate of the hydraulic fluid) is the same, the damping force is controlled by controlling the damping characteristic (damping coefficient), so the damping characteristic control is considered to be the damping force control. You can also.

The coil spring 20 is disposed between a lower retainer 110 provided at an intermediate portion of the housing 60 of the shock absorber 22 and an upper retainer 114 provided on the mount portion 54 via an anti-vibration rubber 112. Since the housing 60 is supported by the second lower arm 46 and attached to the vehicle body 14 at the mount portion 54, the coil spring 20 is provided in parallel with the shock absorber 22 between the second lower arm 46 and the vehicle body 14. .
An elastic member 116 is provided in a portion of the piston rod 64 located inside the housing 60, and the upper surface of the elastic member 116 abuts on the lower surface of the lid 66, so that the rebound direction (the separation between the wheel and the vehicle body) is achieved. Direction) movement limits are defined. Further, when the upper surface of the lid portion 66 of the housing 60 abuts against the anti-vibration rubber 112, a movement limit in the bound direction (the approach direction between the wheel and the vehicle body) is defined. The elastic member 116 or the elastic member 116 and the lower surface of the lid 66 constitute a rebound stopper, and the anti-vibration rubber 112 or the anti-vibration rubber 112 and the upper surface of the lid 66 constitute a rebound stopper. .

As shown in FIGS. 3 and 6, the vertical force generator 24 is a bar 122 that is generally L-shaped when viewed from the vertical direction (which is an embodiment of an elastic member, and hereinafter abbreviated as an L-shaped bar). ) And an actuator 124 (which is one mode of the drive source) that rotates the L-shaped bar 122 about the axis Ls. The L-shaped bar 122 includes a shaft portion 130 that extends substantially in the vehicle width direction and an arm portion 132 that intersects the shaft portion 130 and extends generally rearward of the vehicle so that force can be transmitted integrally (for example, one bar). (Bending the bar). The actuator 124 is fixed to the vehicle body 14 at the attached portion 134.
The L-shaped bar 122 is supported by the vehicle body 14 by being connected to the actuator 124 at the end opposite to the side where the arm portion 132 of the shaft portion 130 is provided, and at the end on the arm portion side. The mounting tool 136 is supported on the vehicle body 14 in a state that allows rotation around the axis Ls of the shaft portion 130. Further, the end portion of the arm portion 132 opposite to the shaft portion 130 is connected to the second lower arm 46 via a link rod 137. The link rod 137 is swingably connected at one end to a link rod connecting portion 138 provided on the second lower arm 46, and swingable to the end of the arm portion 132 of the L-shaped bar 122 at the other end. It is connected to.

As shown in FIG. 7, the actuator 124 includes an electric motor 140 and a speed reducer 142, and the shaft portion 130 of the L-shaped bar 122 is connected to the output shaft of the electric motor 140 via the speed reducer 142. The rotation of the electric motor 140 is decelerated and transmitted to the shaft portion 130.
In the housing 144, an electric motor 140 and a speed reducer 142 are provided in series. The output shaft 146 of the electric motor 140 and the output shaft 148 of the speed reducer 142 are respectively held in the housing 144 via bearings 150 and 152 so as to be relatively rotatable. The output shafts 146 and 148 are hollow, and the shaft portion 130 is inserted inside them. The shaft portion 130 is held by the housing 144 via the bush type bearing 153 so as to be relatively rotatable.
The electric motor 140 includes a plurality of coils 154 provided on the inner peripheral surface of the housing 144, an output shaft 146, and a plurality of permanent magnets 155 provided on the output shaft 146 (the permanent magnet 155 is provided on the outer peripheral surface of the output shaft 146. Provided or embedded). The electric motor 140 is a three-phase DC brushless motor, and the rotation angle of the electric motor 140 is detected by a motor rotation angle sensor 156.
The reduction gear 142 is configured as a harmonic gear mechanism, and includes a wave generator (wave generator) 157, a flexible gear (flex spline) 158, and a ring gear (circular spline) 160. The wave generator 157 includes an elliptical cam and a ball bearing fitted on the outer periphery thereof, and is fixed to one end of the motor output shaft 146. The flexible gear 158 has a cup shape in which the cylindrical portion can be elastically deformed. A plurality of teeth (400 teeth in the present speed reducer 142) are formed on the outer peripheral surface of the cylindrical portion, and the flexible gear 158 is formed on the bottom portion. The shaft portion 130 of the L-shaped bar 122 is fitted into the hole, and can rotate integrally. The ring gear 160 is generally ring-shaped and has a plurality of teeth (402 teeth in the present speed reducer 142) formed on the inner periphery, and is fixed to the housing 144. The flexible gear 158 has a cylindrical portion located on the outer peripheral side of the wave generator 157, elastically deformed in an elliptical shape, and meshes with the ring gear 160 at two locations located in the major axis direction of the ellipse, and meshes at other locations. It is in a state not to do.

When the wave generator 157 is rotated once (360 degrees), that is, when the output shaft 146 of the electric motor 140 is rotated once, the flexible gear 158 and the ring gear 160 are relatively rotated by two teeth, and the shaft portion 130 is rotated. Is rotated. In the present embodiment, a portion that can rotate integrally with the shaft portion 130 of the flexible gear 158 serves as the output shaft 148 of the speed reducer 142.
The reduction ratio of the reduction gear 142 is 1/200, which is relatively large. The rotational speed of the output shaft 148 of the speed reducer 142 is smaller than the rotational speed of the electric motor 140, and as a result, the control delay of the actuator 124 increases (after the control command value is output to the electric motor 140, the shaft portion 130 The time until torque is applied is long).

On the other hand, the positive efficiency η _P of the actuator 124 is defined as the ratio of the external input to the minimum motor force required to rotate the shaft portion 130 of the L-shaped bar 122 against a certain external input, and the reverse efficiency η _{When N} is defined as the ratio of the minimum motor force at which the actuator 124 cannot be rotated by some external input to the external input, the actuator force (the force applied from the outside to the actuator 124 and considered as the actuator torque) If the motor force (which can be considered as the motor torque) generated by the electric motor 140 is Fm, then the positive efficiency η _P and the reverse efficiency η _N can be expressed as Can be expressed.
η _P = Fa / Fm
η _N = Fm / Fa
The normal efficiency η _P and the reverse efficiency η _P are necessary for the motor force Fm _P of the electric motor 140 required under the normal efficiency characteristics and the reverse efficiency characteristics even when the actuator force Fa having the same magnitude is generated. The motor power Fm _N is different (Fm _P > Fm _N ). Further, the product of the positive efficiency η _P and the reverse efficiency η _N of the electric motor 140 (referred to as the positive / reverse efficiency product η _P · η _N ) is necessary to operate the actuator against a certain amount of external input. It can be considered as a ratio (Fm _N / Fm _P ) between the motor force required and the motor force necessary for the actuator not to be operated by its external input.
The smaller the forward / reverse efficiency product η _P · η _N is, the smaller the motor force Fm _N required under the reverse efficiency characteristic is relative to the motor force Fm _P required under the normal efficiency characteristic. It can be considered that the actuator is hard to move.
In this embodiment, since the forward / reverse efficiency product η _P · η _N of the actuator is small, there is an advantage that a small current is sufficient to hold the force applied to the L-shaped bar 122.

  As described above, between the unsprung member 46 and the sprung member 14, the coil spring 20, the shock absorber 22, and the L-shaped bar 122 as an elastic member are provided in parallel to each other. Therefore, the load applied to the wheel 12 is received by the coil spring 20, the shock absorber 22, and the L-shaped bar 122 in cooperation. When no electric current is supplied to the electric motor 140, no force is generated in the L-shaped bar 122, so that a load is received by the coil spring 20 and the shock absorber 22 (mainly because the coil spring 20 receives the load). Hereinafter, it is described that the coil spring 20 receives). In this embodiment, this state is a reference state where the rotation angle of the electric motor 140 is 0 (a reference state of the actuator 124).

When the electric motor 140 is driven from this reference state, the torque of the electric motor 140 is applied to the shaft portion 130. The arm part 132 is rotated and the shaft part 130 is twisted. The rotation angle of the electric motor 140 and the rotation angle of the actuator 124 have a one-to-one correspondence. The control command value is a target value of the rotation angle of the electric motor 140 as will be described later.
As shown in FIG. 8 (a) (in FIG. 8, the relationship between the rotation of the electric motor 140, the rotation of the arm portion 132, and the rotation of the unsprung member 46 is described in an easily understandable state. This is different from the actual posture of the bar 122). When the actuator 124 is rotated by the rotation angle θ _{MA in the} P direction and the arm 132 portion is rotated by θ _A , the sprung sprung The stroke in between increases. When the arm portion 132 is rotated in the P direction by the rotation angle θ _A , the stroke between the sprung springs increases by a distance corresponding to θ _A (sin θ _A ), and the elastic force of the suspension spring 20 is Minutes get smaller.
The shaft portion 130 is twisted at an angle obtained by subtracting the rotation angle θ _A of the arm portion 132 from the rotation angle θ _MA of the actuator 124. Since the torsional moment T _M (torque applied by the actuator 124) applied to the shaft portion 130 is equal to the bending moment generated in the arm portion 132, the expression T _M = F _B · L (1)
Is established. Here, L is the length of the arm portion 132, F _B is the force applied to the arm portion 132 (a reaction force of the force applied to the second lower arm 46), F _B · L is, the arm portion 132 The resulting bending moment. A downward force (a force having a downward component) is applied to the second lower arm 46.
On the other hand, the torsional moment T _M of the shaft portion 130, the shear modulus G _S, when the polar moment of inertia of area I _P, wherein _{T M = G S · I P} · (θ MA -θ A) ··· (2)
Is established.
From formulas (1) and (2),
F _B = G _S · I _P · (θ _MA −θ _A ) / L (3)
And the force applied to the second lower arm 46 (corresponding to the vertical force and corresponding to the force applied to the arm portion 132) is proportional to the torsion angle (θ _MA −θ _A ). I understand.
Further, there is a predetermined relationship between the rotation angle θ _MA of the actuator 124 and the rotation angle θ _A of the arm portion 132 (amount of change in vehicle height).
From the above circumstances, once the rotation angle theta _MA of the actuator 124 (electric motor 140), and than the force F _B exerted on the change amount and the second lower arm 46 of the distance between the sprung and unsprung determined, this embodiment , The rotational angle θ _M of the electric motor 140 is controlled so that the vertical force applied to the second lower arm 46 (vertical force applied by the L-shaped bar 122) has a desired magnitude.
As described above, since the shaft portion 130 is held by the vehicle body 14 in the vicinity of the arm portion 132, it is not necessary to consider the bending of the shaft portion 130.
In addition, since the elastic member is the L-shaped bar 122, the actuator 124 can be provided in a portion away from the wheel 12 of the vehicle body 14 as compared with the case where the rod is a linear rod. Can be improved.

As shown in FIG. 8B, when the electric motor 140 (actuator 124) is rotated in the Q direction, the arm portion 132 is rotated by θ _{A in the} arrow Q direction. The stroke between the sprung spring and the unsprung spring is reduced, and the force generated by the coil spring 20 is increased. The shaft portion 130 is twisted in the Q direction by (θ _MA −θ _A ), and a vertical force is applied to the second lower arm 46 in such a direction that the stroke between the sprung springs becomes smaller. The direction of the force applied to the second lower arm 46 by the L-shaped arm 122 is opposite to the direction applied by the coil spring 20.
Also in this case, the vertical force applied to the second lower arm 46 is controlled by controlling the rotation angle θ _M of the electric motor 140.
8 (a) and 8 (b), the direction of the vertical force is determined by the rotation direction of the electric motor 140, and depends on the magnitude of the rotation angle θ _M of the electric motor 140 (sometimes referred to as the absolute value of the rotation angle). The magnitude of the vertical force and the distance (or the amount of change in distance) between the sprung sprung are determined.

In this embodiment, the shock absorber 22, the vertical force generator 24, etc. are controlled by a suspension control unit 168 shown in FIG. Suspension control unit 168 includes a vertical force generator control unit (ECU) 170 and an absorber control unit (ECU) 172. The vertical force generating device control unit 170 controls the vertical force applied to the second lower arm 46 by the L-shaped bar 122, and the absorber control unit 172 controls the damping force generated in the shock absorber 22.
The vertical force generator control unit 170 includes a controller 176 mainly composed of a computer having an execution unit 173, an input / output unit 174, and a storage unit 175, and an inverter 178 as a drive circuit. 174 includes the motor rotation angle sensor 156, the sprung acceleration sensor 196, the vehicle height sensor 198, the left and right front wheels (steering wheels) 12FL, the steering angle sensor 200 for detecting the steering angle of the FR, and the steering amount of the steering member (this embodiment) In the example, a steering angle sensor 204 that detects a steering angle of a steering wheel (not shown) is connected, and the above-described inverter 178 is connected. The sprung acceleration sensor 196 is provided on the mount portion 54 of the vehicle body 14 and detects the vertical acceleration of the sprung member 14. The vehicle height sensor 198 detects the vertical displacement of the sprung member 14 relative to the unsprung member 46 (distance between the sprung sprung). The sprung acceleration sensor 196 and the vehicle height sensor 198 are provided corresponding to the front, rear, left and right wheels 12FL, FR, RL, RR. The storage unit 175 stores a plurality of tables, programs, and the like.

Similarly, the absorber control unit 172 includes a controller 220 mainly composed of a computer including an execution unit 210, an input / output unit 211, a storage unit 212, and the like, and an inverter 222 as a drive circuit. A sprung acceleration sensor 196, a vehicle height sensor 198, a steering angle sensor 200, a steering angle sensor 204, a motor rotation angle sensor 96, and the like are connected, and the above-described inverter 222 is connected.
Similarly, the brake control unit 224 includes a controller mainly composed of a computer. The brake control unit 224 is connected to a wheel speed sensor 226 that detects the rotational speed of each wheel 12FR, FL, RR, RL, and the vehicle traveling speed and slip state are acquired based on the detected values by the wheel speed sensor 226. Is done.
The vertical force generator control unit 170, the absorber control unit 172, the brake control unit 224 and the like are connected to each other via a CAN (Car Area Network), and information indicating the traveling speed of the vehicle acquired in the brake control unit 224, Information indicating the slip states of the front, rear, left and right wheels 12FL, FR, RL, RR and the like are supplied to the vertical force generator control unit 170, the absorber control unit 172, and the like.

  It should be noted that the controller 176 of the vertical force generator ECU 170 and the controller 220 of the absorber ECU 172 can be provided individually for each wheel (for each inverter).

As shown in FIG. 9, the electric motor 140 is a Δ-connected three-phase DC brushless motor, and corresponding to each phase (U, V, W), the energization terminals 230u, 230v, 230w (hereinafter collectively referred to as “general names”). In some cases, it may be referred to as “energization terminal 230”.
The inverter 178 has two switching elements (UHC, ULC) provided corresponding to the high (positive) side and the low (negative) side for each energization terminal, that is, each phase (U, V, W). , VHC, VLC, WHC, WLC) and a switching element switching circuit for switching these switching elements. Switching element switching circuit, three Hall elements H _A provided in the electric motor 140, H _B, (in the figure, is indicated as H) H _C determines the rotation angle (electrical angle) by the detection signal of, Based on the rotation angle, each of the six switching elements is switched ON / OFF. Inverter 178 is connected to battery 236 via converter 232.
Thus, since a constant voltage controlled by the converter 232 is applied to the electric motor 140, the amount of power supplied to the electric motor 140 is changed by changing the amount of supplied current. The supply current amount is changed by changing the ratio (duty ratio) between the pulse on time and the pulse off time by PWM (Pulse Width Modulation) in the inverter 178.
The operation mode of the electric motor 140 is changed by controlling the operation state of the inverter 178 configured as described above. The operation modes include a control energization mode in which power is supplied from the battery 236 to the electric motor 140, a standby mode in which power is not supplied, a brake mode, and a free mode.

In the control energization mode, as shown in FIGS. 9 and 10, the ON / OFF of each switching element UHC, ULC, VHC, VLC, WHC, WLC is turned on by the electric motor 140 by a so-called 120 ° energization rectangular wave drive method. Is switched according to the rotation angle. In this case, duty control is not executed for the high-side switching elements UHC, VHC, WHC, but duty control is executed for the low-side switching elements ULC, VLC, WLC. By changing the duty ratio, the amount of current supplied to the electric motor 140 is changed. “1 *” in FIG. 10 indicates that duty control is performed. The switching mode of each switching element differs depending on the rotation direction of the electric motor 140, and is referred to as a clockwise direction (CW direction) and a counterclockwise direction (CCW direction) for convenience.
In the control energization mode, the power supplied to the electric motor 140 is controlled, thereby controlling the direction and magnitude of the torque.

In the standby mode, even if switching of each switching element is executed, the power from the battery 236 to the electric motor 140 is not actually supplied.
As shown in FIG. 10, compared with the case in the control energization mode, the switch elements WHC, VHC, UHC existing on the high side are switched in the same manner as in the control energization motor, but each switching element existing on the low side is switched. In any of the elements ULC, VLC, WLC, the duty ratio is 0, and the low-side switching elements ULC, VLC, WLC are always in the OFF state (open state). As a result, power is not supplied to the electric motor 140 in this mode. “0 *” in FIG. 10 indicates this.
Thus, in this mode, only one switching element among the switching elements UHC, VHC, WHC, ULC, VLC, and WLC is turned on (closed state). Conduction between one and the high-potential side terminal 234H of the converter 232 is ensured. For this reason, this operation mode can be considered as a kind of specific terminal energization mode. Even in the standby mode, there are two forms of switching elements, the CW direction and the CCW direction.
In the brake mode, the energization terminals of the electric motor 140 are electrically connected to each other by ON / OFF of the switching element. This mode can be considered as a kind of all-terminal conduction mode. Regardless of the rotation angle of the electric motor 140, all of the switching elements arranged on one of the high side and the low side (in this embodiment, the low side) are maintained in the closed state, and the high side and the low side are maintained. Everything arranged on the other side (high side in this embodiment) is kept open. Each phase of the electric motor 140 is short-circuited by the high-side switching elements UHC, VHC, and WHC that are turned on. In this mode, the effect of short circuit braking is obtained.
In the free mode, all of the switching elements UHC, ULC, VHC, VLC, WHC, and WLC are in the OFF state (open state). In this mode, the electric motor 140 is in a free state.

As described above, the operation of the electric motor 140 (actuator 124) is controlled by the switching control of the switching element in the inverter 178, whereby the vertical force F _B applied to the unsprung member 46 by the L-shaped bar 122 is increased. Be controlled. In the present embodiment, the direction of the vertical force F _B is the vertical movement direction opposite to the direction of the unsprung member 46, is controlled to be a magnitude corresponding to the absolute velocity of the unsprung member 46 size . The vertical force is controlled so as to act as a damping force.
The direction of the vertical force F _B is determined by the rotation direction of the electric motor 140 from the reference position, and the magnitude of the vertical force F _B is controlled by the rotation angle of the electric motor 140. As described above, since between the rotation angle theta _M of the electric motor 140 and the magnitude of the vertical force F _B is a predetermined relationship, based on the relationship between these, the vertical force F _B The target rotation angle θ _M * (rotation direction and size) can be determined so as to obtain a desired direction and size.

The amount of current supplied to the electric motor 140 is set according to the target rotation angle θ _M * of the electric motor 140 and the motor rotation angle deviation Δθ (= θ _M * −θ) between the actual rotation angle θ. . In this embodiment, PI control is performed, and the amount of supplied current is expressed by the following equation: i = K _P · Δθ + K _I · Int (Δθ)
It depends on. K _P and K _I are a proportional gain and an integral gain, respectively, and Int (Δθ) corresponds to an integral value of the motor rotation angle deviation Δθ. When the absolute value of the rotation angle deviation Δθ is large, the supply current i is supplied such that the amount of supply current is increased and the actual rotation angle θ quickly approaches the target rotation angle θ _M *.
In the present embodiment, when the absolute value of the target value F _B * of the vertical force described later is increased, the electric motor 140 is driven based on the magnitude (absolute value) of the supply current i. The duty ratio is determined. The sign of the supply current i corresponds to the desired torque (rotation direction) direction of the electric motor 140. When control commands for the duty ratio and the rotation direction are output to inverter 178, switching elements are controlled in inverter 178 in accordance with the control command value. In this case, the supply current i * (having a sign) corresponds to the control command value. On the other hand, when the absolute value of the target value F _B * is held or decreased, the control command value for the duty ratio and the rotation direction is not output, but as described later, the brake mode, the free mode A control command value indicating switching to is output.

In the present embodiment, the supply current i is determined according to the PI control law. However, the supply current i may be determined according to the PID control law. In this case, for example, the equation i = K _P · Δθ + K _I · Int (Δθ) + K _D · Δθ
Accordingly, the supply current i is determined. K _D is a differential gain, and the third term means a differential term component.

In this embodiment, normal control is performed for the front wheel side vertical force generator 24F, and in principle, preview control is performed for the rear wheel side vertical force generator 24R. In the case where it is not possible to effectively suppress the vibration, normal control may be performed.
Normal control refers to control in which the detection target wheel whose vertical behavior is detected by the sensors 196 and 198 is the same as the control target wheel whose vertical force is controlled. The normal control is sometimes referred to as normal suspension control.
In the normal control, an absolute speed (hereinafter referred to as an unsprung absolute speed) V _L of the unsprung member 46 of the wheel to be controlled (also a detection target wheel) 12ij is acquired, and a damping force corresponding to the unsprung absolute speed V _L is obtained. The vertical force generator 24ij is controlled so that
The absolute speed (hereinafter referred to as the sprung absolute speed) V _U of the sprung member 14 is acquired by integrating the detection value G _U by the sprung acceleration sensor 196 provided corresponding to the controlled wheel 12ij with time, By differentiating the detection value by the vehicle height sensor 198 with respect to time, the relative speed between the sprung springs (the changing speed of the stroke between the sprung springs) V _S is obtained. Then, by subtracting the relative speed V _S from the sprung absolute speed V _U , the unsprung absolute speed V _L is acquired.
V _L = V _U −V _S = V _U − (V _U −V _L )
The vertical force target value (target damping force) F _B *, the unsprung absolute speed when the V _L, wherein _{F B * = - G 0 ·} C · V L
The target rotation angle θ _M * of the electric motor 140 is calculated according to the equation θ _M * = f (F _B *)
As required. C is an attenuation coefficient, and is a predetermined fixed value. G ₀ is a gain of normal control and is a predetermined fixed value. f is a predetermined function, and the target rotation angle θ _M * is acquired according to the function formula. The sign (−) indicates that the direction of the target damping force F _B * is opposite to the direction of the unsprung absolute speed. When the unsprung absolute speed is upward, the target damping force F _B * is downward.
Then, as described above, the rotation angle deviation Δθ is acquired from the target rotation angle θ _M * and the actual rotation angle θ, the supply current i corresponding to the rotation angle deviation Δθ is obtained, and based on the supply current i and the like As described above, the control command value is created and output. In normal control, the control command value is output as soon as it is created.

On the other hand, as described above, the actuator 124 has a large control delay (poor response). Therefore, when the actuator 124ij of the same wheel 12ij is controlled based on the vertical behavior of the own wheel 12ij, the vibration may not be suppressed satisfactorily, or the ride comfort may be deteriorated.
On the other hand, it has been confirmed by actual vehicle tests, simulations, and the like that even if control is started with a phase delay of １ ／, the vibration can be suppressed. Further, the control delay time (the time from when the control command value is output until the torque of the electric motor 140 is actually applied to the L-shaped bar 122) is determined in advance by the structure of the actuator 124, the performance of the inverter 178, and the like.
Therefore, in this embodiment, even if the suspension control is started with a delay with respect to the actual vibration, if the delay time is equal to or less than the time corresponding to 1/8 phase of the vibration, Control (normal control) of the actuator 124ij is performed.
When the control delay time of the actuator 124 is T _D , the frequency f _D corresponding to the 1/8 phase of the control delay time T _D is expressed by the formula f _D = 1 / (8 · T _D ).
Is required. Therefore, actual frequency f of the vertical vibration of the wheel 12ij is, the frequency (hereinafter, referred to as lag corresponding frequency) higher than f _D (large), more if high-frequency vibration occurs (f> f _D) since the response delay time T _D is longer than the 1/8 phase component, control of damping force by the vertical force generator 24ij is not performed. On the other hand, when the actual frequency f is equal to or less than the delay corresponding frequency f _D and the low frequency vibration occurs (f ≦ f _D ), the control delay time T _D is equal to or less than 1/8 phase. The vibration control effect can be obtained. Therefore, in this case, the damping force is controlled by the vertical force generator 24ij.

In the preview control, the control target wheel is a rear wheel and the detection target wheel is a front wheel (the detection target portion by the sensor is a front wheel side portion). The vertical behavior of each of the front wheels 12FL, FR is detected, and based on the detected vertical behavior, the vertical force generators 24RL, RR of the rear wheels 12RL, RR on the same side in the vehicle width direction are detected. Each is controlled.
As shown in FIG. 1, when it is assumed that the rear wheels 12RL and RR pass through the same road surface through which the front wheels 12FL and FR pass, the same input as the road surface input applied to the front wheels 12FL and FR is applied to the rear wheels 12RL. , RR are also received, and the same behavior as that of the unsprung members 46FL, FR of the front wheels 12FL, FR occurs in the unsprung members 46RL, RR of the rear wheels 12RL, RR after a predetermined time has elapsed. If this is used to control the vertical force generator 24R of the rear wheel 12R based on the vertical behavior of the unsprung member 46F of the front wheel 12F, the control delay of the vertical force generator 24R is controlled. Can be made small or 0, and it is possible to satisfactorily suppress the vertical vibration of the unsprung member 46R of the rear wheel 12R.
In this embodiment, the unsprung absolute speed V _L of the front wheel 12F is acquired, the target damping force F _B * corresponding to the unsprung absolute speed V _L is obtained, and a control command value is created. Then, after a lapse of a predetermined time, that is, in accordance with the vertical behavior of the unsprung member 46R of the rear wheel 12R, the damping force corresponding to the control command value is generated in the rear wheel 12R. The directional force generator 24R is controlled.

The target damping force F _B * is determined based on the unsprung absolute velocity V _L , and the formula F _B * = − G · C · V _L
Get according to. G is a preview gain (a gain used for preview control).
As described above, the target rotation angle θ _M is expressed by the equation θ _M * = f (F _B *)
As described above, the rotation angle deviation Δθ is obtained from the target rotation angle θ _M * and the actual rotation angle θ, and the supply current i corresponding to the rotation angle deviation Δθ is obtained. A value is created.
Control command value is, in principle, from the detection of the vertical behavior of the front-wheel side portion, the wheel base L _W time minus the time T _D delayed previewable time T _P is obtained by dividing the vehicle speed V and T _Q (hereinafter, This is output after the waiting time).
T _P = L _W / V _S
T _Q = T _P -T _D

Afford time T _P, the unevenness of the same road, a time until the rear wheel 12R passes from through the front wheels 12F, as shown in FIG. 12 (a), for the same vehicle (wheel base L _W is When the vehicle speed V is high, the vehicle speed V is shorter than that when the vehicle speed V is low.
When the margin time T _P is equal to or longer than the control delay time T _D , that is, the time obtained by subtracting the control delay time T _D (in this embodiment, the control delay time corresponds to the first set time) from the margin time T _P. When the waiting time _TQ is 0 or more, the preview control can be performed effectively. Therefore, as shown in FIG. 12B, when the waiting time _TQ is 0 or more, the preview gain is set to 1. If latency is greater than or equal to 0, corresponding to when the previewable time T _P is the response delay time T _D above, the vehicle speed V is, wheelbase L _W the response delay time T _D divided by the magnitude V _D (V _D = L _W / T _D ) or less (V _S ≦ V _D ).
In contrast, the vehicle speed is increased, the margin time T _P becomes shorter than the response delay time T _D, even if the control command value without waiting T _Q is output, the vertical force generator 24R for the rear wheel 12R Since the control with respect to is delayed with respect to the actual vertical behavior of the rear wheel 12R, the vertical vibration of the rear wheel 12R cannot be satisfactorily suppressed by the preview control, or the ride comfort is deteriorated. Sometimes. Therefore, in this embodiment, when the margin time T _P becomes shorter than the control delay time T _D , as the vehicle speed V increases (as the margin time decreases, as indicated by the solid line in FIG. 12B). The preview gain G is gradually decreased linearly.

Also, when the margin time T _P becomes shorter than the effective time T _L (T _P <T _L ), the preview gain G is set to 0, and preview control is not performed.
As described above, it is known that even if the control is delayed with respect to the actual vibration, the control effect can be obtained when the delay is equal to or shorter than the 1/8 phase equivalent time. Based on this fact, as shown in FIGS. 12 (a) and 12 (c), even if the control command value is immediately output without waiting time, the actual vertical vibration of the rear wheel 12R is reduced to 1 / The margin time in the case of delaying 8 phases is defined as effective time T _L.
The effective time T _L is a time T _X {1 / (8 × N) sec} that is 1/8 phase delayed from the control delay time T _D when the vibration frequency of the unsprung member 46 is N (Hz). In this embodiment, the effective time T _L corresponds to the second set time.
In the case where the control delay time T _D is the same (when the actuator 124 is the same), the time T _X corresponding to the 1/8 phase becomes longer in the vibration with the small frequency N, so the effective time T _L becomes shorter and the frequency In vibrations with a large N, the effective time _TL becomes longer. In this embodiment, N is set to 3. The frequency of 3 Hz is a relatively high frequency in the vibration of the vehicle that normally occurs, and is the maximum frequency that can be controlled by the actuator 124. As a result, the longest effective time T _L is set as a threshold value within the controllable range of the actuator 124.
The vehicle speed V when the margin time T _P is the effective time T _L is
V _SMAx = L _W / T _L
It becomes. Therefore, when the vehicle speed V becomes _higher than _VSMAx , the preview control is not performed.
Relationship between the preview gain G and the previewable time T _P, as shown in FIG. 12 (b), are stored previously been tabulated.

Incidentally, when the previewable time T _P is shorter than the response delay time T _D, the preview gain G is also possible to curvedly gradually decreases as shown by the broken line in FIG. 12 (b), as indicated by the two-dot chain line It may be 0.
Further, the table may represent the relationship between the preview gain G and the vehicle speed V.
Furthermore, the first set time is a time longer than the control delay time T _D (a time when the set value is added), or a time shorter than the control delay time T _D (the set value is subtracted). Time multiplied by a ratio less than 1).
Further, the vibration frequency is acquired each time, and the second set time (effective time T _L ) or the vehicle speed upper limit value V _{SMAX at} which preview control is performed is determined each time, and the gain can be determined each time. . For example, when the preview gain G is an actual vehicle speed V, the equation G = V / (V _SMAX −V _D )
Can be asked according to.

When the vehicle is turning, the rear wheels 12RL and RR do not always pass along the same road surface through which the front wheels 12FL and FR pass. If the rear wheels 12RL and RR do not pass through the same road surface as the road through which the front wheels 12FL and FR pass, preview control may not be performed effectively. Therefore, in this embodiment, as shown in FIGS. 13A and 13B, the lap rate Lap is acquired based on the track difference ΔR and the tire width W _T, and the case where the lap rate Lap is low is high. Accordingly, the preview gain G is set to a small value. As shown in FIG. 16B, when the lap rate Lap is 0 or less, the preview gain is set to 0 so that preview control is not performed.
The lap rate Lap is the length ΔW in the vehicle width direction (turning radius direction) of the overlapping portion of the road surface on which the tires WF of the front wheels 12FL and FR pass and the road surface on which the tires WR of the rear wheels 12RL and RR are expected to pass. _T is a value obtained by dividing the width dimension W _T of the tire ([Delta] W _T / W _T). The width dimension of the tire (hereinafter, width abbreviated tire) W _T is the front wheel 12F and rear wheel 12R, when the width of the tire is the same, its dimensions.

The trajectory (trajectory) of each wheel 12FL, FR, RL, RR is represented by a single continuous line. In this embodiment, the wheel 12 (or tire) is represented by a set of contacts with the road surface on the center plane in the width direction, and is represented by a turning radius R of the set of contacts.
Further, a trajectory in the middle of the trajectory of the left front wheel 12FL and the trajectory of the right front wheel 12FR can be used as a trajectory on the front wheel side. The intermediate trajectory is represented by an average value of the turning radius of the left front wheel 12FL and the turning radius of the right front wheel 12FR, or as shown in FIG. 14, the trajectory (turning radius) of the center point PF in the width direction on the front wheel side of the vehicle. Or can be represented by Center point PF is strictly speaking, when the vehicle is traveling straight horizontal road surface, a vertical plane including the line L _V extending in the longitudinal direction through the center of gravity G of the vehicle, the left and right front wheels 12FL, FR axle It is the intersection with the line that passes through the axis. The trajectory of the intersection point PF can be considered to be composed of a set of points obtained by projecting the intersection point PF onto the road surface.
Similarly, the trajectory difference (trajectory difference) ΔR is a value obtained by subtracting the turning radius Rr of the rear wheel 12R from the turning radius Rf of the front wheel 12F. However, the difference between the turning radii of the front wheel 12F and the rear wheel 12R on the same side. The difference between the turning radius on the front wheel side and the turning radius on the rear wheel side may be used. Locus of the rear wheel side, similar to the front wheel side of the track, passing through a vertical plane including the center line L _V, left and right rear wheels 12RL, a line passing through the axis of the RR of the axle (left and right rear wheels 12RL, the centers of the RR Line) and the turning radius Rr of the intersection (center point) PR.

As shown in FIG. 14, when the slips of the wheels 12FL, FR, RL, and RR are small, it is known that the turning center is on an extension of the line passing through the center line of the axles of the left and right rear wheels 12RL and RR. ing. Therefore, if the absolute value of the steering angle of the steered wheel (front wheel) 12F is δ _W and the wheel base is L _W , the turning radius of the center point PF on the front wheel side and the center point PR on the rear wheel side is given by the equation Rf = L _W / sinδ _W · 10 ^-3
Rr = L _W / tan δ _W · 10 ⁻³
Can be asked according to. Since 10 ^-3 is the unit of the wheel base L _W (mm) and the unit of the turning radii Rf and Rr is (m), it is a value for conversion of these units. In this embodiment, since the turning direction is not related, the absolute value of the steering angle is used.
On the other hand, the turning radii of the front wheel 12F and the rear wheel 12R inside the turn are respectively expressed by the formula Rfin≈Rf−Tf / 2.
Rrin = Rr-Tr / 2
The turning radii of the front wheel 12F and the rear wheel 12R outside the turn are respectively
Rfout≈Rf + Tf / 2
Rrout = Rr + Tr / 2
Can be expressed as Tf and Tr are wheel treads of the left and right front wheels 12FL and FR, and the left and right rear wheels 12RL and RR, respectively.

As a result, the difference (trajectory difference) ΔRin between the turning radii of the front wheel 12F and the rear wheel 12R inside the turn, and the difference ΔRout between the turning radii between the front wheel 12F and the rear wheel 12R outside the turn are respectively expressed by the equation ΔRin≈Rf− {Rr + ( Tf−Tr) / 2}
= (Rf-Rr)-(Tf-Tr) / 2 (4)
ΔRout≈Rf− {Rr− (Tf−Tr) / 2}
= (Rf-Rr) + (Tf-Tr) / 2 (5)
It becomes the size expressed.
On the other hand, the difference in turning radius between the front wheel side and the rear wheel side is expressed by the equation ΔR = Rf−Rr.
Therefore, from equation (4), the turning radius difference ΔRin between the front wheel 12F and the rear wheel 12R of the turning inner wheel is the tread difference (Tf−) from the turning radius difference (Rf−Rr) between the front wheel side and the rear wheel side. Tr) is smaller by 1/2 of Tr), and the turning radius difference ΔRout between the front wheel 12F and the rear wheel 12F of the outer turning wheel is tread from the difference (Rf−Rr) of the turning radius between the front wheel side and the rear wheel side. It can be seen that the difference is increased by ½ of the difference (Tf−Tr).
As shown in FIG. 15, from the above formula, the turning radius on the front wheel side (turning radius of the center point PF) Rf (the turning radius of the front wheel 12 is Rfin, Rfout) is the turning radius on the rear wheel side (turning of the center point PR). Radius) Rr (rear wheel turning radii Rrin, Rrout) is relatively larger, and it turns out that the turning radius decreases as the absolute value δ _W of the front wheel rudder angle increases. It can also be seen that the difference in turning radius, that is, the trajectory difference increases as the absolute value δ _W of the front wheel steering angle increases and the turning radius R decreases.

As shown in FIG. 24, the turning radius Rf on the front wheel side and the turning radius Rr on the rear wheel side are the distance L between the turning radius Rg of the center of gravity G of the vehicle and the center point PF on the front wheel side. It can also be obtained according to the following equation from _W f and the distance L _W r between the center of gravity G and the center point PR on the rear wheel side.
Rf = √ (Rg ² + L _W f ² )
Rr = √ (Rg ² + L _W r ² )
L _W = L _W f + L _W r
In this case, the turning radius Rg of the center of gravity G of the vehicle is expressed by the equation Rg = V / γ from the yaw rate γ and the traveling speed V of the vehicle.
Can be asked according to.
Further, the turning radius Rg is given by the formula Rg = L _W · (1 + K · V ² ) / δ _W
You can also ask for it. Here, K is a stability factor, and when the equivalent cornering power of the front and rear wheels is Kf, Kr and the mass of the vehicle is m, the equation K = m (L _W r · Kr−L _W f・ Kf) / (2 ・ L _W ²・ Kf ・ Kr)
Can be asked according to.
Furthermore, the turning radius Rg of the center of gravity G of the vehicle can also be acquired based on road information from the navigation system. Based on information representing the radius of curvature of the corner of the road, the turning radius of the vehicle can be obtained.

As shown in FIG. 13 (b), the length ΔW _T in the width direction of the portion where the road surface through which the front wheel 12F passes and the road surface predicted to pass through the rear wheel 12R overlaps with the turning radius Rf on the front wheel side and the rear wheel side. When the turning radius Rr is considered, the equation ΔW _T = (Rr + W _T / 2) − (Rf−W _T / 2)
= W _T- (Rf-Rr)
= W _T -ΔR (6)
It becomes the length represented by. This can be obtained by subtracting the turning radius of the inner peripheral side edge of the tire of the front wheel 12F from the turning radius of the outer peripheral side edge of the tire of the rear wheel 12R.
From (6), the width [Delta] W _T of the overlapping portion, it can be seen that the front wheel side, the value obtained by subtracting the turning radius difference for the rear wheels (track difference) [Delta] R from the tire width W _T. From this equation, it can be seen that when the track difference is smaller than the width of the tire, there is an overlapping portion, but when the track difference is equal to or greater than the width of the tire, there is no overlapping portion.
When the turning radii are obtained for each of the inside of the turning and the outside of the turning, respectively, ΔRin and ΔRout may be substituted for ΔR in the equation (6), respectively.
Accordingly, the lap rate Lap is expressed by the formula Lap = (W _T −ΔR) / W _T = 1−ΔR / W _T
Can be asked according to.

As shown in FIG. 16A, the lap rate Lap is a value that decreases as the trajectory difference ΔR increases. When the track difference ΔR becomes large and the overlapping portion [Delta] W _T of the road surface through the tire WR road and the rear wheel 12 through the tire WF of the front wheel 12 is reduced. Further, when the absolute value δ _W of the front wheel steering angle reaches the set value δ _W0 , the lap rate Lap becomes 0, and then the lap rate becomes a value smaller than 0 as the absolute value δ _W of the front wheel steering angle increases. Become. That the lap rate Lap is 0 or less means that there is no overlapping portion between the road surface through which the front wheel 12F passes and the road surface through which the rear wheel 12R passes. As described above, when the track difference ΔR of the front wheel side and rear wheel side is equal to or greater than the width W _T of the tire, there is no overlapping portion, the overlapping ratio is 0 or less.
The preview gain G is set to 1 when the lap rate Lap is equal to or larger than the set value Lapth as indicated by the solid line in FIG. Since the setting value Lapth has many overlaps, it is considered that the preview control can be effectively performed even during turning, and can be set to a value in the vicinity of 0.8, for example. When the lap rate Lap is smaller than the set value Lapth, the preview gain is also reduced as the lap rate Lap decreases. As the overlap is reduced, the preview gain is reduced. When the lap rate Lap is 0, the preview gain is also 0. When the lap rate Lap is 0 or less, preview control is not performed. The relationship between the lap rate Lap and the preview gain is previously tabulated and stored.

Note that the preview gain can also be a value that gradually decreases as the lap rate Lap decreases, as indicated by the broken line in FIG.
Further, the table can represent the relationship between the absolute value δ _W of the front wheel steering angle and the preview gain.

The normal control is performed according to the execution of the normal control program represented by the flowchart of FIG. This program is executed at predetermined time intervals for each of the left and right front wheels 12FL, FR.
Step 101 (hereinafter, simply, the same applies for. Other steps abbreviated as S101) in, sprung acceleration G _U of the control object wheels (e.g., will be described the front left wheel 12FL is controlled wheel) is In step S102, a vehicle height (unsprung unsprung stroke) H is detected. In steps S103 and 104, a control command value corresponding to the vertical force generator 24FL of the left front wheel 12FL is created. Specifically, the sprung absolute speed V _U is obtained by integrating the sprung acceleration G _U, and the relative height (stroke change) ΔH (V _S ) of the sprung spring is obtained by differentiating the vehicle height H. From these, the unsprung absolute speed V _L is obtained. Then, the target damping force F _B * is obtained from the gain G ₀ , the damping coefficient C, and the unsprung absolute speed V _L , the target rotation angle θ _M * is obtained, and the actual rotation angle θ and the target rotation angle θ _M * are obtained. The supply current i is acquired from the rotation angle deviation Δθ that is the difference between the two.
In S105, the frequency f of the vertical vibration of the unsprung member 46 is acquired based on the unsprung absolute speed V _L. When the absolute speed is 0, the unsprung member 46 is in a position where the absolute value of the amplitude is maximized. Based on this, the frequency can be acquired. In S106, the frequency f is, or less than the predetermined delay corresponding frequency f _D is determined. If the actual vertical vibration frequency f of the left front wheel 12FL is small and is equal to or less than the delay-corresponding frequency f _D , the control is effective, and thus a control command value is output in S107.
In contrast, when a high frequency than the corresponding frequency f _D lag the actual vertical vibration of the frequency f of the front left wheel 12FL, since control is not considered valid, S106 the determination is NO, control No command value is output (the vertical force generator 24FL is not controlled).

The output of the control command value in S107 is performed according to the flowchart of FIG. In S121, it is determined whether or not the absolute value of the target damping force F _B * has increased. If it has increased, a control command value representing the supply current i is output to the inverter 176FL in S122.
On the other hand, if the absolute value of the target damping force F _B * is not less than the threshold value Fth, it is determined in S123 if it has not increased, that is, if it has decreased or is almost constant. Is done. If it is equal to or greater than the threshold value Fth, the brake mode is selected in S124, and a control command value indicating that is output. If it is smaller than the threshold value Fth, the free mode is selected in S125, and a control command value indicating that is output.
As shown in FIG. 23, when the absolute value of the damping force is increased, the electric motor 140 is energized. However, when the absolute value of the damping force is decreased, no current is supplied. The force applied to the wheel 12 to return the sprung portion to the reference state is applied to the actuator 124 via the second lower arm 46 and the L-shaped bar 120, so that the supply current is controlled. Even if not, it is returned to the reference position. The actuator 124 has a small forward / reverse efficiency product and is not easily affected by the external input. However, when the free mode is set, the actuator 124 is operated by the external input and returned to the reference position. Thus, if current is not supplied when the absolute value of the damping force is reduced, power consumption can be reduced. Also, when the absolute value of the target damping force F _B * is large, the brake is Since the mode is set, it is possible to avoid that the absolute value of the damping force is suddenly reduced by an external force.
Furthermore, when the absolute value of the target damping force F _B * is reduced, it is possible to perform energy regeneration. In this way, energy efficiency can be further improved.
Further, since the energized state is not set when the absolute value of the target damping force F _B * is reduced, the rotation direction of the actuator 124 can be quickly reversed as compared with the case where current control is performed. , A decrease in responsiveness can be suppressed.

  Note that the frequency of the vertical vibration of the wheel to be controlled can be acquired based on a change in the sprung absolute speed, or can be acquired based on the displacement of the sprung or unsprung. . It can also be acquired using Fourier transform or the like.

The preview control is performed according to the execution of the preview control program represented by the flowchart of FIG. Each of the left and right rear wheels 12RL and RR is executed at predetermined time intervals. The vertical force generator 24RL of the left rear wheel 12RL is controlled based on the vertical behavior of the left front wheel 12FL, and the vertical force generator 24RR of the right rear wheel 12RR is based on the vertical behavior of the right front wheel 12FR. Controlled.
In S1, the control object wheel (e.g., described in the case the rear left wheel 12RL is controlled wheel) is detected sprung acceleration G _U and the same side front wheel (left front wheel) 12FL, at S2, the vehicle height H Is detected. In S3, as described above, the unsprung absolute speed V _L is acquired based on these. In step S4, the preview gain G is determined. In step S5, it is determined whether the determined preview gain G is zero.
If the preview gain is greater than 0, preview control is performed in S6-10. In S6, the target damping force F _B * is acquired from the preview gain G, the damping coefficient C, and the unsprung absolute speed V _{L, the} target rotation angle θ _M * is acquired from the target damping force F _B *, and the rotation angle deviation Δθ is acquired. Supply current i is acquired. In S7, the waiting time T _Q is determined from the margin time T _P determined in S4. In S8, by comparing the margin time T _P and the response delay time T _D, when the previewable time T _P is the response delay time T _D above, in S9, the supply current amount i is stored. The control command value is output after the waiting time _TQ has elapsed.
In contrast, when the previewable time T _P is shorter than the response delay time T _D, in S10, the control command value is immediately outputted.
If the preview gain is 0, the same control as the normal control represented by the flowchart of FIG. 21 is performed in S11. When the control target wheel is the left rear wheel 12RL, the vertical force generator 24RL is controlled based on the vertical behavior of the left rear wheel 12RL.
Further, the output of the control command values in S9 and S10 is performed in the same manner as in the normal control according to the flowchart of FIG. When the target damping force F _B * tends to decrease, the electric motor 140 is not energized, so that power consumption can be reduced as compared with the case where the electric motor 140 is energized.

In the S9, after a waiting time T _Q, in accordance with the flowchart of FIG. 22, even as the normal control similar run is performed, to create the pre-control command value, and stores the control command value itself, The control command value may be output after the waiting time _TQ has elapsed.

The preview gain in S4 is determined according to the flowchart of FIG.
In the present embodiment, when the vehicle is in a straight traveling state, a vehicle speed corresponding gain (which can also be referred to as a margin time corresponding gain) G _V determined based on the vehicle speed (margin time) is used as it is as a preview gain. Is done. When the vehicle is in a turning state, the geometric mean of the turning-time gain G _R and the vehicle speed corresponding gain G _V determined by the turning state, that is, a value obtained by multiplying the product G _R · G _V by a power of {{(G _R G _V )} is the preview gain G. When the absolute value δ _W of the steering angle of the steered wheel (front wheel 12F) is equal to or less than the set value, the vehicle is in a straight traveling state, and when it is greater than the set value, the vehicle is in a turning state. The set value is a size that can be considered that the vehicle is traveling straight ahead.
In S21, the determined vehicle speed-basis gain G _V is, the steering angle of the front wheels 12F step S22 is detected, in S23, whether the absolute value [delta] _W of the steering angle is equal to or smaller than the set value [delta] _MIN is determined. If it is equal to or _smaller than the set value δMIN, it is not necessary to consider the turning state, and therefore the vehicle speed corresponding gain G _V is set as the preview gain in S24 (G ← G _V ).
On the other hand, if the absolute value δ _W of the front wheel rudder angle is larger than the set value δ _MIN , the slip ratio of at least one of the front, rear, left and right wheels 12FL, FR, RL, RR is determined in advance in S25. It is determined whether or not the set slip ratio is greater than or equal to. The slip ratio (braking slip or drive slip) in the front-rear direction of at least one wheel is greater than or equal to a predetermined first set slip ratio, and the second set slip ratio in which the lateral slip ratio is predetermined. If at least one of the above is satisfied, the determination is YES. The turning gain is not determined, and the vehicle speed corresponding gain G _V is set as the preview gain G in S24. The first and second set slip ratios are magnitudes that reduce the estimation accuracy of the turning radius, and are preset fixed values. If the determination is NO in S26, the determined turning-state gain G _R are in S27, the vehicle speed-basis gain G _V, geometric mean of turning-state gain G _R is determined as a preview gain G.

The determination of the vehicle speed corresponding gain in S21 is executed according to the flowchart of FIG.
In S51, the vehicle speed V is acquired in S52, is retrieved from the vehicle speed V and the wheel base L _W previewable time T _P is, in S53, it determines whether the previewable time T _P is the response delay time T _D or Is done. If it is the response delay time T _D above, in S54, the vehicle speed-basis gain G _V is one.
On the other hand, if the margin time T _P is shorter than the control delay time T _D , it is further determined whether or not the effect time T _L is shorter. If it is shorter than the effective time T _L , the vehicle speed corresponding gain G _V is set to 0 in S56. Although the previewable time T _P is shorter than the response delay time T _D, if it is effective limit time T _L or more is determined to be the table value at S57. As the margin time T _P decreases, the vehicle speed corresponding gain G _V is determined to be a smaller value. The vehicle speed-corresponding gain G _V is the same for both the left rear wheel 12RL and the right rear wheel 12RR.

The turning gain determination in S25 is executed according to the flowchart of FIG.
In S71, the turning radius on the front wheel side (point PF) and the turning radius on the rear wheel side (point PR) are obtained, and in S72, the turning radius difference (trajectory difference) is acquired. ΔRout is acquired when the left rear wheel 12RL that is the control target wheel is a turning outer wheel, and ΔRin is acquired when the left rear wheel 12RL is a turning inner wheel. In S73, the lap rate Lap is obtained.
In S74-78, the turning gain is determined based on the lap rate Lap. In S74, it is determined whether or not the lap rate Lap is equal to or greater than a set value Lapth. If the set value Lapth above, in S75, the turning time of the gain G _R are considered to be one. When the overlap ratio Lap is set value Lapth smaller is determined whether or not greater than zero, when greater than 0, in S77, a table value is a turning-state gain G _R, if it is 0 or less Is set to 0 in S78.

Incidentally, the overlap ratio Lap is the rear left wheel 12RL, obtained for each of the right rear wheel 12RR (turning inner acquired for each of the turning outer wheel), the turning-state gain G _R using the acquired overlap ratio Lap it is obtained, may be acquired pivot gain G _R using the average value of the overlapping ratio Lapout turning inner wrap ratio Lapin a turning outer wheel. In the former case, the preview gain values may be different for the left and right rear wheels.

As described above, in this embodiment, the preview control is performed on the vertical force generating device 24R of the rear wheel 12R. Therefore, even if the control delay of the actuator 124R is large, the delay can be reduced to zero or small. The vibration in the vertical direction of 12RL and RR can be satisfactorily suppressed.
Further, and when the previewable time T _P is shorter than the response delay time T _D, the preview control is performed when the overlap between the road surface through which the rear wheel 12R and the road surface which the front wheel 12F is through a small, rather, ride comfort is poor However, since the preview gain G is determined to a value smaller than 1 based on the vehicle speed and the turning state, it is possible to avoid the deterioration of the riding comfort by performing the preview control, and the rear wheels 12RL, RR. Can be suppressed satisfactorily.
Further, the normal control, the response delay time T _D of the actuator 124 is performed when the low-frequency vibration than the vibration corresponding to 1/8 phase occurs, when the high-frequency vibration occurs is not performed, The preview control is performed if the vibration is below a controllable frequency determined by the response of the actuator 124. As a result, if the preview control is performed on the actuator 124, vibrations at higher frequencies can be suppressed.
Further, in the operating state of the vertical force generator 24, vibrations having a large frequency can be absorbed by elastic deformation of the L-shaped bar 122.

As described above, in this embodiment, the preview control program represented by the flowchart of FIG. 17 of the vertical force generator ECU 170 included in the sprung acceleration sensor 196, the vehicle height sensor 198, and the suspension ECU 168, the flowchart of FIG. The vertical force control device is constituted by a part for storing a normal control program represented by the following, a part for executing the normal control program, and the like. In the present embodiment, the vertical force generator 24 functions as a damping force generator, and the vertical force controller functions as a damping force controller. The damping force control unit is also an unsprung motion dependent vertical force control unit.
Also, in the vertical force control device, the part for storing S4 of the preview control program of FIG. 17, the part to be executed, the table represented by the map of FIG. 12B, and the map of FIG. 16B. The gain determination unit is configured by a part that stores the table. Among them, the gain reduction unit is configured by a part for storing the table represented by the solid line or the broken line in FIG. 12B, a part for storing S56 and 57 in the flowchart in FIG. The part for storing the table in FIG. 12B, the part for storing S56 in the flowchart of FIG. Further, a turning gain determining unit is configured by a part storing the table of FIG. 16B, a part storing S26 of the flowchart of FIG. Note that the 0 determining unit described in claim 4 includes a part for storing a table represented by a one-dot chain line in FIG. 12B, a part for storing S57, a part for executing, and the like.

  Furthermore, a locus-corresponding overlap acquisition unit is configured by a portion that stores S71 to S73 in the flowchart of FIG. The trajectory corresponding overlap acquisition unit is also a lap rate acquisition unit. Of these, the turning radius acquisition unit is configured by the part that stores S71, the part that executes, etc., and the part that stores the table of FIG. 6B, the part that stores S77 of the flowchart of FIG. A decreasing part is constructed.

In addition, when acquiring the vehicle speed corresponding gain G _V , it is not essential to obtain the allowance time T _P, and it can be acquired based on the vehicle speed V. As described above, a table representing the relationship between vehicle speed and gain can be created.
Similarly, when retrieving a turn when the gain G _R, rather it is essential to obtain the overlapping ratio can also be obtained based on the width [Delta] W _T and orbital difference of the overlapping portion (the turning radius difference).
Further, in the above embodiment, during turning, the geometric mean of the running-speed-basis gain G _V and the turning time gain G _R are, have been preview gain G, it is not essential to do so. During turning, the turning gain G _R is set as a preview gain (G _R → G), and during straight traveling, the vehicle speed corresponding gain G _V is set as a preview gain G (G _V → G). it can. In this case, step S26 is not necessary.
Furthermore, the damping force F _B * can be controlled based on the skyhook theory in at least one of normal control and preview control.
Further, the target damping force F _B * can be set to a value corresponding to the sprung absolute speed V _U or a value corresponding to the sprung unsprung relative speed V _S. In these cases, the control based on the unsprung absolute speed _VL and the rule for obtaining the target deceleration F _B * can be changed.
F _B * = − G ・ C ・ V _U
F _B * =-G ・ C ・ V _S

Further, in the above embodiment, the case has been described where the damping force is applied by the control of the vertical force generator 24, so that the elastic force corresponding to the displacement X _L of the unsprung member 46 (vertical force) is applied It can also be.
The target value of the vertical force (target elastic force) F _B * is expressed by the formula F _B * = G · K · X _L
As required.
Displacement of the unsprung member 46 (hereinafter, referred to as unsprung displacement) when X _L is lower than the reference position, the target elastic force F _B * orientation is also the lower. When the distance between the sprung springs increases, the elastic force of the coil spring 20 decreases. By generating a decrease in the elastic force of the coil spring 20 by the vertical force generator 24, the displacement of the sprung member 14 accompanying the displacement of the unsprung member 46 is suppressed. Incidentally, the rotation of the arm portion 132 due to the rotation of the electric motor 140, the distance between the sprung and unsprung a distance corresponding to the unsprung displacement X _L.
When the unsprung displacement _XL is above the reference position, the direction of the target elastic force F _B * is upward. When the distance between the unsprung springs decreases, the elastic force of the coil spring 20 increases. Therefore, a reverse force corresponding to the increased amount is generated by the vertical force generator 24 and the unsprung member 46 This suppresses the displacement of the sprung member due to the vertical displacement.

Unsprung displacement X _L is or can be obtained from or calculated by integrating the absolute velocity V _L unsprung, the sprung acceleration G _U and stroke H between the integrated value and sprung under the two. K is the spring constant of the L-shaped bar 122, and is a fixed value determined based on the shear coefficient of the shaft portion 130, the secondary moment of section, the bending rigidity of the arm portion 132, and the like.
An example of control in this case will be described based on the flowcharts of FIGS. Steps in which the same execution as in the above embodiment (the same execution as the program shown in the flowcharts of FIGS. 17 and 21) is performed are denoted by the same step numbers and description thereof is omitted.
When the left front wheel 12FL is a wheel to be controlled, the normal control is performed according to the execution of the normal control program represented by the flowchart of FIG. In S103b, a sprung acceleration G _U and vehicle height H of the left rear wheel 12RL, as described above, unsprung displacement X _L of the left rear wheel 12RL is acquired, at S104b, * is obtained target elastic force F _B Accordingly, the target rotation angle θ _M * is obtained, and the supply current i is obtained. Then, in S105 and S106, the actual vibration frequency f of the front left wheel 12FL unsprung member 46 is obtained, or less than the response delay corresponding frequency f _D is determined. The frequency may be acquired based on the unsprung absolute speed or may be acquired based on the unsprung displacement. When the actual vibration frequency f is equal to or less than the control delay corresponding frequency f _D , the control command value determined based on the supply current i and the target elastic force F _B * in S107 is the same as in the above embodiment. Is output.

When the control target wheel is the left rear wheel 12RL, the preview control is performed in accordance with the execution of the preview control program represented by the flowchart of FIG. In S3b, the detection value by the sprung acceleration sensor 196 of the front left wheel 12FL, the vehicle height H and unsprung displacement from X _L is obtained. If the preview gain is greater than 0, the target elastic force (target value of the vertical force) F _B * is acquired from the unsprung displacement X _L , the elastic coefficient K, and the preview gain G in S6b, and the target elastic force is obtained. The target rotation angle θ _M * is determined so that the force F _B * is obtained, and the supply current i is determined. When the previewable time T _P is the response delay time T _D or more, after a waiting time T _Q, the control command value is outputted to the inverter 178 of the vertical force generator 24 of the rear wheels 12FR, margin time If T _P is shorter than the response delay time T _D, the control command value is immediately outputted.
If the preview gain is 0, normal control is performed in S11b.
Thus, the preview control can be applied not only to the damping force control but also to the elastic force control.
In the present embodiment, the vertical force generator functions as an elastic force generator, and the vertical force controller functions as an elastic force controller.

The target elastic force F _B * may be a value corresponding to the displacement X _U of the sprung member 14 or may be a value corresponding to the relative displacement (vehicle height) X _S under the sprung spring. Preview control similar to that in the embodiment can be performed.
F _B * = G ・ K ・ X _U
F _B * = G ・ K ・ X _S

In the above embodiment, the vertical force is controlled by controlling the vertical force generator 24. However, the damping force is controlled by controlling the shock absorber 22. You can also. In the present embodiment, the case where the damping force is controlled based on the skyhook theory will be described.
A control example in that case will be described with reference to the flowcharts of FIGS. In the present embodiment, steps that are executed in the same manner as in the above-described embodiment (flowcharts of FIGS. 17 and 21) are denoted by the same step numbers and description thereof is omitted. In the present embodiment, it is determined whether or not the frequency of vibration of the sprung member 14 is equal to or lower than the control delay corresponding frequency.
In the normal control represented by the flowchart of FIG. 28, when the wheel to be controlled is the left front wheel 12FL, the sprung acceleration G _U of the left front wheel 12FL and the distance H between the sprung unsprung are acquired, and in S105b, frequency f is obtained in S 106 b, the obtained frequency f is equal to or less than the response delay corresponding frequency f _D is determined. The frequency f can be acquired based on the sprung acceleration or can be acquired based on the sprung absolute speed. That is, as in the case of the above embodiment, when the absolute velocity is 0, the absolute value of the amplitude is obtained using the maximum position, or obtained using the Fourier transform or the like. Can do it. When control delay corresponding at frequency f _D or less, the determination is YES in S 106 b, in S103d, sprung absolute speed V _U, the relative velocity V _S of the sprung and unsprung portions is obtained. In S104d to 104f, the target attenuation coefficient C * is obtained. In S104d, it is determined whether or not the product value of the sprung absolute speed V _U and the relative speed V _S between the sprung springs is positive. If it is positive (V _U · V _S > 0), the target damping coefficient C * is set to (G ₀ · C · V _U / V _S) in S104e, and the product value is negative (V _{For U} · V _S <0), the target damping coefficient C * is set to a small value C _MIN in S104f. G ₀ is a normal control gain, and C is a constant. In S104g, a supply current i that achieves the target damping coefficient C * is obtained, and in S107, a control command value is output.
In this embodiment, a current is supplied to the electric motor 90 regardless of the increase / decrease of the damping coefficient. Therefore, supply current i corresponds to the control command value, and control command value i is output to inverter 222. This is because the electric power consumption in the electric motor 90 is small.

In the preview control shown in the flowchart of FIG. 27, when the wheel to be controlled is the left rear wheel 12RL, the sprung acceleration G _U and the sprung unsprung stroke H of the left front wheel 12FL are acquired. The upper absolute speed V _U and the relative speed V _S are acquired. In S4, the preview gain is determined in the same manner as in the above embodiment. When the determined preview gain G is not 0, the target attenuation coefficient is determined in S6d to 6h as in the above case. When the sign of the product value of the sprung absolute velocity V _U and the relative velocity V _S is positive, the target damping coefficient is C * = G · C · V _U / V _S, and is negative. Is C _MIN . In step S6g, the supply current i is obtained based on the target attenuation coefficient C *. Hereinafter, as in the above embodiment, when the margin time T _P is equal to or longer than the control delay time T _D , the control command value i is output after the elapse of the waiting time T _Q and is shorter than the control delay time T _D. In the case, it is output immediately.
On the other hand, when the preview gain G is 0, in S11d, the damping characteristic control device 56 for the left rear wheel 12RL is controlled based on the vertical behavior of the left rear wheel 12RL in accordance with the flowchart of FIG. The target damping characteristic C * is determined to have a magnitude determined by the product value of the sprung absolute speed V _U and the sprung unsprung relative speed V _S of the left rear wheel 12RL.

When this embodiment is compared with the above-described embodiment (when the control target is the damping characteristic control device 56 and the vertical force generating device 24), the control target is the damping characteristic control device 56. The control delay time is shorter and the controllable frequency is higher. Therefore, the control delay corresponding frequency f _D becomes larger (a higher frequency side value), the control delay time T _D becomes shorter, and the control effect effective time T _L becomes shorter.
Therefore, normal control is performed even if vibrations with a higher frequency than in the control of the vertical force generator 24 occur. In normal control, the frequency range in which suspension control is effectively performed becomes wider (the determination in S106b in the flowchart of FIG. 28 is more likely to be YES). Further, the preview gain is set to 1 even if the vehicle speed is higher than in the case of the control of the vertical force generator 24. The vehicle speed range in which the preview control is effectively performed is widened. Furthermore, preview control is performed even if the vehicle speed is higher than in the above embodiment (the case where the determination of S5 in the flowchart of FIG. 27 is YES is less).

When the vehicle speed V is high, the preview control is performed on the shock absorber 22 even if the vertical force generator 24 is not preview-controlled. It is possible to satisfactorily suppress the vertical vibrations of
Similarly, the shock absorber 22 is controlled even if the high-frequency vibration is generated and the vertical force generator 24 is not controlled. Even when the high-frequency vibration is generated, the vertical vibration is generated. Vibration can be suppressed satisfactorily.

  As shown in FIG. 29, the shock absorber is controlled in a manner that a coil spring 284 and a shock absorber 286 are provided in parallel between an unsprung member 280 and a sprung member 282 for holding the wheel 12. Therefore, the present invention can be applied to a suspension in which the vertical force generator 24 is not provided. The damping characteristic control device 288 of the shock absorber 286 is controlled in the same manner as in the above embodiment based on the command of the suspension control unit 290.

  The present invention can also be applied to the suspension shown in FIG. In the present embodiment, a coil spring 310 and a hydraulic cylinder device 312 are provided in parallel between the unsprung member 300 and the sprung member 302. The hydraulic cylinder device 312 includes a hydraulic cylinder 314, a pump 316, an electric motor 318, and the like. The hydraulic cylinder 314 includes a housing 320 and a piston 322 that is fluid-tightly and slidably fitted to the housing 320. A piston rod 324 of the piston 322 is swingably connected to the unsprung member 300. Is pivotably coupled to the sprung member 302. A pump 316 is connected between the two liquid chambers 330 and 332 partitioned by the piston 322 inside the housing 320. The pump 316 pumps hydraulic fluid from one of the two liquid chambers 330 and 332 to the other. It can be supplied, or the working fluid can be pumped from the other and supplied to the other. Thereby, the fluid pressure in the fluid chambers 330 and 332 and the stroke of the piston 322 are controlled. A hydraulic fluid compensator 340 is provided in parallel with the hydraulic cylinder 314.

The electric motor 318 is controlled based on a command from the suspension ECU 350 mainly including a computer. The suspension ECU 350 includes a controller mainly composed of a computer having an input / output unit 352, a storage unit 354, an execution unit 356, and the like. A provided vehicle height sensor (vertical stroke sensor) 360, sprung acceleration sensor 362, and the like are connected, and a pump motor 318 provided for each wheel is connected via a drive circuit (not shown).
The storage unit 354 stores a plurality of programs, tables, and the like.

In this embodiment, the vertical force that is the sum of the elastic force corresponding to the unsprung absolute speed and the damping force based on the Skyhook theory is generated by the control of the electric motor 318 of the hydraulic cylinder device 312. The vertical force corresponds to the hydraulic pressure of the hydraulic cylinder device 312.
Similarly to the case of the above embodiment, the load applied to the wheel is received by the spring 310 and the hydraulic cylinder device 312, so that the stroke between the reference position and the hydraulic pressure of the hydraulic cylinder device 312 is between Have a certain relationship. Therefore, when the target value of the vertical force is determined, the pump motor 318 is operated so that a stroke corresponding to the target value is realized. In this embodiment, the target value F _B * of the vertical force is the sum of the elastic force according to the displacement of the unsprung member 300 and the damping force according to the absolute speed of the sprung member 302.

A control example in that case will be described based on the flowcharts of FIGS. In the present embodiment, steps that are executed in the same manner as in the above-described embodiments (the flowcharts of FIGS. 17 and 21 and the flowcharts of FIGS. 27 and 28) are denoted by the same step numbers and description thereof is omitted.
When the control target wheel is the left rear wheel 12RL, the preview control program represented by the flowchart of FIG. 31 is executed. The sprung acceleration G _U and the sprung unsprung stroke H of the left front wheel 12FL are acquired, and in S3e, the sprung absolute speed V _U , the relative speed V _S , and the unsprung absolute speed V _L are acquired. Then, the preview gain G is determined in the same manner as in the above embodiment, and when the preview gain G is not 0, the sign of the product value of the sprung absolute velocity V _U and the relative velocity V _S is positive in S6d. If negative, the attenuation coefficient C is set to C _MID (predetermined value) in S6h. If negative, the attenuation coefficient C is determined in S6f. Is C _MIN (predetermined value). In S6i, the formula F _B * = (G · K · X _L ) + (− G · C · V _U )
Accordingly, the target value F _B * of the vertical force is determined, and the supply current i to the electric motor 318RL is determined according to the target value F _B * of the vertical force. In the above equation, K is the spring constant K of the spring 310.
Hereinafter, as in the above embodiment, when the margin time T _P is equal to or longer than the control delay time T _D , the control command value i is output after the elapse of the waiting time T _Q and is shorter than the control delay time T _D. In the case, it is output immediately.
In this embodiment, since the electric power consumed by the electric motor 318 is large, the absolute value of the target value F _B * of the vertical force is maintained or reduced, as in the case of the control of the vertical force generator 24. In this case, it is desirable that no electric current be supplied to the electric motor 318.

On the other hand, when the preview gain G is 0, the electric motor 318RL of the left rear wheel 12RL is controlled based on the vertical behavior of the left rear wheel 12RL in S11e. Normal control is performed.
In S103e, the unsprung displacement X _L and the unsprung absolute speed V _U are acquired based on the sprung acceleration and the vehicle height. Similarly, in S104d, 104f, and 104h, the damping coefficient is determined. In S104i, the vertical force The target value F _B * is the formula
F _B * = (− G ₀ · K · X _L ) + (− G ₀ · C · V _U )
And the supply current i determined by the target value F _B * is determined. G ₀ is a gain of normal control and is a fixed value.
In S105c and 106c, the greater one of the frequency acquired based on the unsprung absolute speed V _L of the left rear wheel 12RL and the frequency acquired based on the sprung absolute speed V _U is the control delay corresponding frequency. or less than f _D is determined. When the frequency is equal to or less than the control delay corresponding frequency, the control command value is immediately output.
In S105c and 106c, the frequency may be acquired based on the unsprung absolute speed or may be acquired based on the sprung absolute speed.

Thus, in the present embodiment, a vertical force having a magnitude that combines both the elastic force according to the unsprung displacement and the damping force according to the sprung absolute velocity is applied. Both suppression control of unsprung vibration and skyhook control can be realized, vibration in the vertical direction of the wheel 12 can be suppressed satisfactorily, and riding comfort can be improved.
Further, in the hydraulic cylinder device 312R for the rear wheel 12R, since preview control is performed, the control delay can be reduced or reduced to 0, and vibrations in the vertical direction of the rear wheel 12R can be suppressed well. it can.

Note that the target value F _B * for the vertical force is not necessarily the sum of the elastic force and the damping force described above, and can be a magnitude determined by one of the following equations.
F _B * = G ・ K ・_XL
F _B * = − G ・ C ・ V _U
In addition, the target value F _B * is expressed by the formula F _B * = − G · C · V _L
F _B * = G ・ K ・ X _U
F _B * =-G ・ C ・ V _S
It is also possible to use a value determined according to the above, or a value obtained by summing two or more of these values may be used as a target value.

Furthermore, the present invention can also be applied to the suspension control shown in FIG.
In this embodiment, the vertical force generator 370 includes a linear rod 372 instead of an L-shaped bar. One end of the linear rod 372 is connected to the actuator 374, and the other end is connected to the unsprung member 380 via the link member 378. The actuator 374 is attached to a vehicle body 382 as a sprung member, and a linear rod 372 is disposed between the sprung member 382 and the unsprung member 380. A coil spring 384 is provided between the sprung member 382 and the unsprung member 380, and the coil spring 384 and a linear rod 372 as an elastic member are provided in parallel.
The actuator 374 includes an electric motor and a speed reducer. A linear rod 372 is connected to the output shaft of the electric motor via the speed reducer, and a motor torque _TM is applied by driving the electric motor. Further, in the linear rod 372, since the motor torque T _M and the bending moment L · F _B * are equal, the reaction force F _B * is expressed by the formula F _B * = T _M / L
Is the size required by Reaction force F _B * is the vertical force generator 370, the force F _B * reaction force exerted on the unsprung member 380.
The actuator 374 is connected to a controller 392 mainly including a computer via an inverter 390. The controller 392 is connected to a sprung acceleration sensor, a vehicle height sensor, a rudder angle sensor, a steering angle sensor, a brake ECU, and the like, as in the above-described embodiment. And the output torque of the electric motor 374 is controlled. In this embodiment, the controller 392 and the inverter 390 constitute a suspension control unit.
Target value of the vertical force F _B * may be similarly appropriately determined as in the above embodiments, by controlling the torque T _M of the electric motor 374, it is possible to control the vertical force.

In each of the above-described embodiments, the case where the preview control is performed on the suspension of the rear wheel 12R based on the detection values of the sensors 196F and 198F that detect the vertical behavior of the front wheel side portion has been described. As shown in FIG. 34, the invention is similarly applied to the case where preview control is performed on the suspension of the front wheel 12F and the suspension of the rear wheel 12R based on the detection value by the road surface sensor 402 provided in the front bumper 400. can do.
The road surface sensor 402 can detect road surface irregularities by, for example, ultrasonic waves. The distance to the road surface is acquired based on the time until the transmitted ultrasonic wave is received and the ultrasonic wave is received (the time until it is reflected by the road surface and returned), and the unevenness of the road surface is acquired. As shown in FIG. 35, the road surface sensor 402 is provided on both the front of the right front wheel 12FR and the front of the left front wheel 12FL of the bumper 400 (402R, 402L). The portion of the road surface detected by the road surface sensors 402R, L (the road surface detection target portion) corresponds to a portion almost directly below the position where the road surface sensors 402R, L are provided when the vehicle is stopped.
Therefore, when acquiring the preview gain, the distance from the position where the vehicle road surface sensor 402 is provided to the wheel to be controlled {if the wheel to be controlled is the front wheel, the distance L _P , and the wheel to be controlled is the rear wheel. In such a case, a margin time determined by the distance (L _P + L _W )} and the traveling speed of the vehicle, and an overlap between the detection target portion of the road surface and the road surface through which the control target wheel passes are problematic.

When the wheel to be controlled is the front wheel 12F, the distance in the front-rear direction between the road surface sensors 402R, L and the left and right front wheels 12FL, FR is a line passing through the pair of road surface sensors 402 (the width of the vehicle) A distance L _{P in} the front-rear direction between the left and right front wheels 12FL and the line through which the center line of the axle of the FR passes. Therefore, a value obtained by dividing the distance L _P by the vehicle speed is the margin time T _P.
T _P = L _P / V
When the wheel to be controlled is a rear wheel, the sum of the distance L _P and the wheel bale L _W corresponds to the distance in the front-rear direction between the road surface sensors 402R, L and the left and right rear wheels 12RL, RR. . As a result, the margin time T _P is given by the formula T _P = (L _P + L _W ) / V
The time represented by
Hereinafter, the vehicle speed corresponding gain G _V can be acquired in the same manner as in the above embodiment.

The turning radius on the front wheel side, the turning radius on the rear wheel side, the turning inner side, the turning wheel front and rear turning radii can be obtained in the same manner as in the above embodiment.
Further, the trajectory of the detection target portion can be considered to be the same as the trajectories of the road surface sensors 402R and L. The trajectory of the road surface sensor 402 is acquired as a trajectory of a predetermined point in each of the road surface sensors 402R and L, is acquired as an intermediate trajectory of the trajectories of the road surface sensors 402R and L, or the center point P of the front portion of the vehicle It can also be acquired as a _fv trajectory. The center point P _fv is an intersection point P _fv between a vertical plane including a line extending in the front-rear direction passing through the center of gravity G and a line passing through the pair of road surface sensors 402R and L when the vehicle is stopped on a horizontal road surface. . When the locus is considered as a set of points on the road surface, a set of projection points on the road surface of the center point P _fv becomes the locus. The locus (turning radius) of the center point P _{fv is referred to as} a locus (turning radius) on the sensor side.
As shown in FIG. 35, the turning radius on the sensor side is expressed by the equation R _fv = (L _P + L _W ) / sin δ _W · 10 ^−3.
Can be asked according to.
Since the distance L _P is smaller than the turning radius R _fv , the center angle (P _fv −0−PR) can be regarded as the same as the absolute value δ _W of the steering angle of the steered wheels.

When the wheel to be controlled is the front wheel 12F, the difference in the turning radius (trajectory difference) between the road surface sensor 402 and the front wheel inside the turning is expressed by the equation ΔR _f in = (R _fv −T _s / 2) − (R _f − T _f / 2) = R _fv −R _f
Therefore, the difference in turning radius between the road surface sensor 402 and the front wheels on the outside of the turn is expressed by the equation ΔR _f out = (R _fv + T _s / 2) − (R _f + T _f / 2) = R _fv − R _f
Can be asked according to. Ts is the distance between the pair of road surface sensors 402, and in this embodiment, it is equal to the wheel tread Tf of the front wheel.
When the detection target portion by the road surface sensor 402 is a portion surrounded by a circle having a diameter D, as shown in FIG. 36, the length ΔW _T of the overlapping portion in the width direction is expressed by the equation ΔW _T = (R _f + W _T / 2) − (R _fv −D / 2) = (W _T / 2 + D / 2) −ΔR (7)
The lap rate Lap can be calculated according to the formula Lap = ΔW _T / W _T
Can be asked according to.

When the wheel to be controlled is a rear wheel, the difference in turning radius between the road surface sensor 402 and the rear wheel on the inside of the turn and the difference in turning radius between the road surface sensor 402 and the rear wheel on the outside of the turn are respectively _expressed by the equation ΔR _r in _{= (R fv -T s / 2} ) - (R r -T r / 2) = R fv - {R r + (T s -T r) / 2}
ΔR _r out = (R _fv + T _s / 2) − (R _r + T _r / 2) = R _fv − {R _r − (T _s −T _r ) / 2}
Can be expressed as
The length ΔW _T in the width direction of the overlapping portion is expressed by the equation ΔW _T = (R _r + W _T / 2) − (R _fv −D / 2) = (W _T / 2 + D / 2) −ΔR (8)
The lap rate Lap can be calculated according to the formula Lap = ΔW _T / W _T
Can be asked according to.
Hereinafter, as in the embodiments described above, and a is the turning-state gain G _R are obtained based on the overlap ratio Lap, preview gain G is obtained.

On the other hand, when the area of the detection target portion is very small and can be regarded as almost “point”, the diameter D may be set to 0 in the equations (7) and (8).
In this case, from the equations (7) and (8), when the track difference ΔR is larger than 1/2 of the tire width (W _T / 2), there is no overlap and 1/2 of the tire width. It can be seen that there is an overlap if: The length [Delta] W _T represent the relative position with respect to the road surface through the tire to be detected portion (length from the outer peripheral edge of the road surface of the tire through the tire). When the length ΔW _T is ½ (W _T / 2) of the tire width (ΔW _T = W _T / 2), the turning radius difference ΔR = 0, and the center of the tire is the detection target Go through the department. When the length ΔW _T is close to 0 (ΔW _T ≈0), the turning radius difference ΔR = W _T / 2, and the outer peripheral edge of the tire passes through the detection target portion. When the length ΔW _T is close to the tire width W _T (ΔW _T ≈W _T ), the turning radius difference ΔR = −W _T / 2, and the inner peripheral edge of the tire defines the detection target portion. I can see that it passes. In other words, when the center portion of the tire passes the detection target portion, it can be considered that there are more overlapping portions with the road surface irregularities around the detection target portion detected by the road surface sensor 402 than when the edge portion passes. . Given that way, if the overlapping ratio obtained by dividing the length [Delta] W _T above the tire width W _T, it is possible to employ as in the above embodiment, if the overlap ratio is large is small It is reasonable to increase the gain.

  Based on these values, the preview gain is obtained in the same manner as in the above embodiment. Further, the unevenness of the road surface detected by the road surface sensor 402 can be considered as a displacement of the unsprung member of the wheel to be controlled. Based on these, the vertical force generator 24F for the front wheel 12F and the vertical force generator 24R for the rear wheel 12R are controlled. As a result, preview control can be performed not only on the rear wheel 12R but also on the front wheel 12F, so that the control delay can be reduced and vibrations in the vertical direction of the front wheel 12F can be satisfactorily suppressed.

In the above-described embodiment, the road surface sensor 402 detects the unevenness of the road surface substantially vertically below. In that case, the margin time is acquired by the distance between the portion of the road surface to be controlled and the center line of the axle of the wheel to be controlled and the vehicle speed.
Also, considering the model, the sprung displacement X _U , the sprung absolute speed V _U , the sprung unsprung relative speed V _S, etc. are acquired based on the road surface irregularities (unsprung displacement), and the preview control is performed based on these. It can also be done.
Furthermore, the present embodiment can also be applied in combination with the suspension described in FIGS. 29, 30 and 33.

  Although a plurality of aspects have been described above, the present invention is not limited to the above-described embodiments, such as the structure of the suspension and the content of suspension control. It can be implemented in an improved manner.

1 is a diagram conceptually illustrating an entire vehicle including a suspension control apparatus according to an embodiment of the present invention. It is a figure which shows notionally the whole suspension system containing the said suspension control apparatus. It is a side view of the up-down direction force generator included in the suspension system. It is sectional drawing of the shock absorber contained in the said suspension system. It is detailed sectional drawing of the said shock absorber. It is a top view of the up-and-down direction force generator included in the above-mentioned suspension. It is sectional drawing of the actuator of the said up-down direction force generator. It is a figure showing the action | operation of the said up-down direction force generator. It is a circuit diagram of the inverter which controls the electric motor of the said actuator. It is a figure which shows the operating state of the said inverter. 2 is a circuit diagram conceptually showing the periphery of a suspension control unit included in the suspension system. FIG. (a) It is a figure which shows the relationship between a vehicle speed and margin time. (b) A map conceptually showing a vehicle speed corresponding preview gain determination table stored in the storage unit of the control unit. (c) It is a figure which shows notionally the relationship between margin time, control delay time, and effective time. (a) It is a figure which shows the track | orbit of a wheel during the turning of a vehicle. (b) It is a figure showing the overlap of the road surface where the front wheel passed, and the road surface where the rear wheel is expected to pass. It is a figure which shows the relationship between the turning radius of each wheel and the front-wheel steering angle during the turning of a vehicle. It is a figure which shows the relationship between the turning radius of a front wheel and a rear wheel, and a track difference. (a) It is a figure which shows the relationship between the absolute value of a front-wheel steering angle, and a lap | lap rate. (b) A map conceptually showing a turning preview gain determination table stored in the storage unit of the control unit. It is a flowchart showing the preview control program memorize | stored in the memory | storage part of the up-down direction force generator control unit contained in the said suspension system. It is a flowchart which shows a part of said preview control program (preview gain determination). It is a flowchart which shows another part of the said preview control program (vehicle speed corresponding | compatible gain determination). It is a flowchart which shows another part of the said preview control program (gain determination at the time of turning). It is a flowchart showing the normal control program memorize | stored in the memory | storage part of the said up-down direction force generator. It is also a part of the preview control program. It is a flowchart which shows a part of said normal control program (output of a control command value). It is a figure which shows an example of the control in the said suspension system. It is a figure which shows the relationship between another turning radius and a front-wheel steering angle. It is a flowchart showing another preview control program memorize | stored in the memory | storage part of the said up-down direction force generator control unit. It is a flowchart showing another normal control program memorize | stored in the memory | storage part of the said up-down direction force generator. It is a flowchart showing the preview control program memorize | stored in the memory | storage part of the absorber control unit contained in the said suspension system. It is a flowchart showing the normal control program memorize | stored in the memory | storage part of the said absorber control unit. It is a figure which shows notionally the structure of another suspension included in the said suspension system. It is a figure which shows notionally the structure of another suspension included in the said suspension system. It is a flowchart showing another preview control program memorize | stored in the memory | storage part of the suspension control unit contained in the said suspension system. It is a flowchart showing another normal control program memorize | stored in the said memory | storage part. It is a figure which shows notionally the structure of another suspension included in the said suspension system. It is a figure which shows notionally the whole vehicle containing the suspension control apparatus which is another one Example of this invention. It is a figure which shows the relationship between the turning radius of the said vehicle, a front wheel steering angle, and a wheel base. (a) It is a figure which shows the locus | trajectory of the detection target part in the turning of a vehicle, and the locus | trajectory of a wheel. (b) It is a figure showing the overlap of the locus | trajectory of a detection target part (road surface part), and the road surface where it is estimated that a control object wheel passes.

Explanation of symbols

  12: Wheel 14: Car body 20: Coil spring 22: Shock absorber 24: Vertical force generator 46: Second lower arm 56: Damping characteristic controller 90: Electric motor 122: L-shaped bar 124: Actuator 130: Shaft part 132: Arm unit 140: Electric motor 168: Suspension ECU 170: Vertical force generator ECU 172 Shock absorber ECU 178: Inverter 222: Inverter 196: On-spring acceleration sensor 198: Vehicle height sensor 200: Rudder angle sensor 284: Coil spring 286: Shock absorber 288: damping characteristic control device 312: hydraulic cylinder device 316: pump 318; electric motor 350: suspension ECU 360: stroke sensor 362: sprung acceleration Degree sensor 370: Vertical force generator 372: Linear rod 374: Actuator 390: Inverter 392: Suspension ECU 402: Road surface sensor

Claims (6)

A suspension of a rear wheel, which is provided on the front wheel side of the vehicle, includes at least one sensor for detecting the vertical behavior of the front wheel side portion, and is a wheel to be controlled of the vehicle based on a detection value by the at least one sensor. Suspension control device for controlling
When the vehicle is turning , the difference between the trajectory of the front wheel and the trajectory of the rear wheel is a value representing the overlap between the road surface through which the tire of the front wheel passes and the road surface predicted to pass through the tire of the rear wheel. In some cases, a trajectory-corresponding overlap acquisition unit that predicts a smaller value than the small case is provided, and when the value representing the overlap is small, the gain used for determining the gain used for controlling the suspension is smaller than that when the value is large. part saw including a and the track corresponding overlapping acquisition unit, in the width direction of the vehicle portion and the road surface is expected to overlap predicted to tire passes of the rear wheel and the road surface on which the front wheel tires passes A suspension control apparatus, comprising: a lap ratio acquisition unit that acquires a lap ratio that is a value obtained by dividing a dimension by a width of a tire of the rear wheel . The track corresponding overlapping acquisition unit, the trajectory, during turning of the vehicle, serial mounting of the suspension control apparatus in claim 1 including a turning radius acquiring unit that acquires a turning radius of the front and rear wheels, respectively. The control command value for the suspension of the rear wheels is obtained by subtracting the control delay time from the margin time determined by the wheel base of the vehicle and the traveling speed of the vehicle from the time when the detection value by the at least one sensor is acquired. The suspension control device according to claim 1 , further comprising a preview control unit that outputs after a lapse of time. Wherein the gain determination unit, when the lap ratio is smaller than a predetermined set value, as the said overlap ratio decreases, claims 1 comprises a gain decreasing unit for determining the gain to a small value 3 The suspension control apparatus according to any one of the above. The suspension of the wheel to be controlled includes a vertical force generator that is provided between an unsprung member and a sprung member that holds the wheel to be controlled, and generates a force in a substantially vertical direction. The suspension control device according to any one of claims 1 to 4, further comprising a vertical force control device that controls the vertical force generation device based on a detection value obtained by the at least one sensor and the gain. The vertical force generation device includes an elastic member having one end connected to one of the unsprung member and the sprung member and the other end connected to the other, and the vertical force control device includes: 6. The apparatus according to claim 5 , further comprising a drive source that elastically deforms the elastic member against a restoring force, and an elastic deformation amount control unit that controls the vertical force by changing an elastic deformation amount of the elastic member. Suspension control device.

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