JP6136136B2 - Steering device for vehicle - Google Patents

Description

  The present invention relates to a vehicle steering apparatus that includes a vehicle suspension apparatus that supports steered wheels and a steer-by-wire system that performs steering control of the steered wheels.

Conventionally, in a suspension device for a vehicle, a desired suspension performance is achieved by setting a kingpin shaft.
For example, in the technique described in Patent Document 1, the upper and lower pivots that constitute the king pin by supporting the lower arm and the upper arm constituted by two I-type arms at the same point on the lower side and the upper side of the axle of the axle carrier. Steering performance and stability are improved by adopting a link arrangement that suppresses movement in the longitudinal direction of the vehicle when turning points.

  Further, for example, in the technique described in Patent Document 2, a lower transverse link composed of two arms crossing each other on the lower side and the upper side of the axle of the axle carrier and two arms with a common link bearing. A kingpin shaft connecting a virtual lower pivot point represented by the intersection of the two arms of the lower transverse link and an upper pivot point represented by the center of the link bearing of the upper transverse link. Is extended with a negative tilt angle so that the contact point with the road surface is on the inner side in the vehicle width direction of the steered wheels.

JP 2010-126041 A International Publication No. 2009/062823 Pamphlet

  However, when steering is performed while the vehicle is traveling, the lateral force corresponding to the traveling speed is input to the tire ground contact point. However, the technique described in Patent Document 1 does not consider the influence of the lateral force. Furthermore, when braking is performed at the time of turning, a force in the vehicle front-rear direction (rearward force) is applied to the wheels in addition to the lateral force, and it is necessary to consider changes in the link exerted by the front-rear force.

  By the way, as in the technique described in Patent Document 2, when a lower transverse link is configured with two arms, a cross link structure that intersects the two arms is used so that the intersection between the two is a virtual pivot point. can do. At this time, generally, since the scrub radius is increased and the kingpin inclination angle is decreased, the rack axial force at the time of turning can be reduced.

  However, since the virtual pivot point of the lower link is sequentially changed by turning, it is difficult to obtain a desired scrub radius and kingpin inclination angle at the time of turning. In particular, in the technique described in Patent Document 2, when the kingpin shaft is viewed from the front side of the vehicle, the kingpin tilt angle passes through the inner end in the vehicle width direction of the steered wheel at the upper side and the inner side in the vehicle width direction of the steered wheel at the lower side. Therefore, when the steered wheels are steered, the tire ground contact area is small and the stability limit of the vehicle is low.

On the other hand, in a vehicle steering apparatus equipped with a steer-by-wire system that controls the steering of a steered wheel separated from the steering wheel according to the steering state of the steering wheel, the steer-by-wire system controls the steered wheels. In this case, in the vehicle suspension device, the steering response is emphasized as compared with the straight traveling performance of the vehicle, and when the abnormality occurs in the steer-by-wire system, the straight traveling performance of the vehicle suspension device remains lowered. Become.
SUMMARY OF THE INVENTION An object of the present invention is to provide a fail-safe function in a vehicle suspension device by ensuring straight travel when steering control of a steered wheel by a steering control device becomes insufficient.

  In order to solve the above-described problems, a vehicle suspension device according to the present invention includes a steering control device that controls the turning of a steered wheel by operating an actuator according to the steering state of the steering wheel, and the steered wheel is rotatable. And a vehicle suspension device supported by the vehicle. The vehicle suspension device includes an axle carrier having an axle that rotatably supports the steered wheels, a lower link structure that connects a vehicle body side support portion and the axle carrier below the axle, and An upper link structure for connecting the vehicle body side support portion and the axle carrier on the upper side of the axle. Furthermore, at least when the steering control of the steering control device is insufficient, the pivot point of at least one kingpin shaft of the lower link structure and the upper link structure is tilted backward in the vehicle side view. It has a pivot point moving part to move

  According to the present invention, the pivot point of the kingpin shaft of at least one of the lower link structure and the upper link structure is moved in a state where the steered wheel steering control by the steered control device is insufficient and the straight traveling performance is deteriorated. Thus, since the kingpin shaft is tilted backward when the steered wheels are steered, straightness can be ensured and a fail-safe function can be exhibited.

1 is a schematic diagram illustrating a configuration of an automobile 1 according to a first embodiment. It is a perspective view which shows typically the structure of the suspension apparatus 1B. It is a top view which shows typically the structure of the suspension apparatus 1B. It is a front view which shows typically the structure of the suspension apparatus 1B. It is a side view which shows typically the structure of the suspension apparatus 1B. It is a figure showing (a) partial top view (left front wheel) and (b) tire grounding surface (right front wheel) which show typically the lower link structure of suspension system 1B. It is a schematic diagram which shows the structure of the suspension apparatus 1B, Comprising: (a) is a schematic diagram which shows an upper link structure, (b) is a schematic diagram which shows a lower link structure. It is a schematic diagram which shows a lower link structure, (a) shows the moving direction of the virtual lower pivot point of a parallel link structure, (b) shows the moving direction of the virtual lower pivot point of a cross link structure. It is a figure which shows the relationship between the rack stroke at the time of steering, and a rack axial force. It is a figure which shows the locus | trajectory of the tire ground-contact surface center at the time of steering. FIG. 5 is an isoline diagram showing an example of rack axial force distribution at coordinates with a kingpin tilt angle and a scrub radius as axes. It is a figure which shows the relationship between the toe angle and the scrub radius in the case of the tension type suspension device and the present invention in which the lower link members do not intersect. It is a graph which shows the relationship between the road landing point of a kingpin axis | shaft, and lateral force. It is a figure which shows (a) lateral force compliance steer and (b) lateral rigidity in suspension apparatus 1B and a comparative example. It is a figure which shows the longitudinal force compliance steer in the suspension apparatus 1B and a comparative example. It is a block diagram which shows the specific structure of the steering control apparatus of FIG. It is a figure which shows the generation | occurrence | production torque control map for estimating a self-aligning torque. It is a figure which shows the characteristic of a suspension apparatus, Comprising: (a) is a figure which shows the relationship between a caster angle, responsiveness, and stability, (b) is a figure which shows the relationship between a caster trail, lateral force reduction allowance, and straightness It is. It is a figure which shows a steering response characteristic, Comprising: (a) is a characteristic diagram which shows the change of the response characteristic of a vehicle, (b) is a figure which shows the switching timing of a control characteristic. It is a flowchart which shows an example of a steering control processing procedure. It is a flowchart which shows an example of an abnormality determination processing procedure. It is a perspective view which shows the specific structure of the 2nd Embodiment of this invention. It is a top view which shows typically the structure of the suspension apparatus 1B. It is a front view which shows typically the structure of the suspension apparatus 1B. It is a side view which shows typically the structure of the suspension apparatus 1B. It is a schematic diagram which shows the upper link structure and lower link structure of 2nd Embodiment of this invention. It is a schematic diagram which shows the structure of the suspension apparatus 1B, Comprising: (a) is a schematic diagram which shows an upper link structure, (b) is a schematic diagram which shows a lower link structure. It is an enlarged view of Fig.27 (a). It is a schematic diagram which shows 3rd Embodiment of this invention. It is a figure which shows the vehicle body side support structure of a compression link. It is a flowchart which shows an example of the abnormality determination processing procedure which can be applied to 3rd Embodiment. It is a schematic diagram with which description of operation | movement of 3rd Embodiment is provided, Comprising: (a) shows an upper link structure, (b) shows a lower link structure. It is an enlarged view of Fig.32 (a). It is a schematic block diagram which shows 4th Embodiment of this invention. It is a block diagram which shows the specific structure of steering control apparatus CT. FIG. 5 is a diagram illustrating the characteristics of a suspension device, where (a) illustrates the relationship between caster angle, responsiveness, and stability, and (b) illustrates the relationship between caster trail, lateral force reduction allowance, and straightness. It is.

Embodiments of an automobile to which the present invention is applied will be described below with reference to the drawings.
(First embodiment)
(Constitution)
FIG. 1 is a schematic diagram showing a configuration of an automobile 1 according to the first embodiment of the present invention.
In FIG. 1, an automobile 1 includes a vehicle body 1A, a steering wheel 2, an input side steering shaft 3, a steering angle sensor 4, a steering torque sensor 5, a steering reaction force actuator 6, and a steering reaction force actuator angle sensor 7. A steering actuator 8, a steering actuator angle sensor 9, an output side steering shaft 10, a steering torque sensor 11, a pinion gear 12, a pinion angle sensor 13, a steering rack member 14, a tie rod 15, A tie rod axial force sensor 16, wheels 17FR, 17FL, 17RR, 17RL, a vehicle state parameter acquisition unit 21, wheel speed sensors 24FR, 24FL, 24RR, 24RL, a control / drive circuit unit 26, and a mechanical backup 27 are provided. I have.

The steering wheel 2 is configured to rotate integrally with the input side steering shaft 3, and transmits a steering input by the driver to the input side steering shaft 3.
The input-side steering shaft 3 includes a steering reaction force actuator 6, and applies a steering reaction force by the steering reaction force actuator 6 to the steering input input from the steering wheel 2.

The steering angle sensor 4 is provided on the input side steering shaft 3 and detects the rotation angle of the input side steering shaft 3 (that is, the steering input angle to the steering wheel 2 by the driver). Then, the steering angle sensor 4 outputs the detected rotation angle of the input side steering shaft 3 to the control / drive circuit unit 26.
The steering torque sensor 5 is installed on the input side steering shaft 3 and detects the rotational torque of the input side steering shaft 3 (that is, the steering input torque to the steering wheel 2). Then, the steering torque sensor 5 outputs the detected rotational torque of the input side steering shaft 3 to the control / drive circuit unit 26.

In the steering reaction force actuator 6, a gear that rotates integrally with the motor shaft meshes with a gear formed in a part of the input-side steering shaft 3, and input by the steering wheel 2 in accordance with an instruction from the control / drive circuit unit 26. A reaction force is applied to the rotation of the side steering shaft 3.
The steering reaction force actuator angle sensor 7 detects the rotation angle of the steering reaction force actuator 6 (that is, the rotation angle by the steering input transmitted to the steering reaction force actuator 6), and sends the detected rotation angle to the control / drive circuit unit 26. Output.

The steered actuator 8 has a gear that rotates integrally with the motor shaft meshes with a gear formed on a part of the output side steering shaft 10, and the output side steering shaft 10 is moved according to an instruction from the control / drive circuit unit 26. Rotate.
The steering actuator angle sensor 9 detects the rotation angle of the steering actuator 8 (that is, the rotation angle output by the steering actuator 8) and outputs the detected rotation angle to the control / drive circuit unit 26. To do.

The output side steering shaft 10 includes a steering actuator 8, and transmits the rotation input by the steering actuator 8 to the pinion gear 12.
The steering torque sensor 11 is installed on the output side steering shaft 10 and detects the rotational torque of the output side steering shaft 10 (that is, the steering torque of the wheels 17FR and 17FL via the steering rack member 14). Then, the steering torque sensor 11 outputs the detected rotational torque of the output side steering shaft 10 to the control / drive circuit unit 26.

The pinion gear 12 meshes with a spur tooth formed on the steering rack member 14 and transmits the rotation input from the output side steering shaft 10 to the steering rack member 14.
The pinion angle sensor 13 detects the rotation angle of the pinion gear 12 (that is, the turning angle of the wheels 17FR and 17FL output via the steering rack member 14), and controls / detects the detected rotation angle of the pinion gear 12 26.

The steering rack member 14 has a rack that meshes with the pinion gear 12 and converts the rotation of the pinion gear 12 into a linear motion in the vehicle width direction. In the present embodiment, the steering rack member 14 is located on the vehicle front side with respect to the front wheel axle.
The tie rod 15 connects both ends of the steering rack member 14 and the knuckle arms of the wheels 17FR and 17FL via ball joints.

The tie rod axial force sensor 16 is installed in each of the tie rods 15 installed at both ends of the steering rack member 14 and detects the axial force acting on the tie rod 15. The tie rod axial force sensor 16 outputs the detected axial force of the tie rod 15 to the control / drive circuit unit 26.
Here, a steer-by-wire system SBW is configured by the steering reaction force actuator 6, the steering actuator 8, the pinion gear 12, the steering rack member 14, the tie rod 15, and the control / drive circuit unit 26.

  The wheels 17FR, 17FL, 17RR, and 17RL are configured by attaching tires to tire wheels, and are installed on the vehicle body 1A via the suspension device 1B. Among these, for the front wheels (wheels 17FR and 17FL), the direction of the wheels 17FR and 17FL with respect to the vehicle body 1A is changed by the knuckle arm swinging by the tie rod 15 driven by the steer-by-wire system SBWU.

The vehicle state parameter acquisition unit 21 acquires the vehicle speed based on a pulse signal indicating the rotation speed of the wheels output from the wheel speed sensors 24FR, 24FL, 24RR, and 24RL. Moreover, the vehicle state parameter acquisition part 21 acquires the slip ratio of each wheel based on the vehicle speed and the rotational speed of each wheel. Then, the vehicle state parameter acquisition unit 21 outputs the acquired parameters to the control / drive circuit unit 26.
The wheel speed sensors 24FR, 24FL, 24RR, 24RL output a pulse signal indicating the rotational speed of each wheel to the vehicle state parameter acquisition unit 21 and the control / drive circuit unit 26.

  The control / driving circuit unit 26 controls the entire automobile 1, and based on signals input from sensors installed in each part, the steering reaction force of the input side steering shaft 3, the steering angle of the front wheels, or the mechanical backup 27, various control signals are output to the steering reaction force actuator 6, the steering actuator 8, the mechanical backup 27, or the like.

  Further, the control / drive circuit unit 26 converts the detection value by each sensor into a value corresponding to the purpose of use. For example, the control / drive circuit unit 26 converts the rotation angle detected by the steering reaction force actuator angle sensor 7 into a steering input angle, or converts the rotation angle detected by the steering actuator angle sensor 9 into the wheel turning angle. Or the rotation angle of the pinion gear 12 detected by the pinion angle sensor 13 is converted into the turning angle of the wheel.

  The control / drive circuit unit 26 includes a rotation angle of the input side steering shaft 3 detected by the steering angle sensor 4, a rotation angle of the steering reaction force actuator 6 detected by the steering reaction force actuator angle sensor 7, and a steering actuator. The rotation angle of the steering actuator 8 detected by the angle sensor 9 and the rotation angle of the pinion gear 12 detected by the pinion angle sensor 13 are monitored, and the occurrence of a failure in the steering system is detected based on these relationships. be able to. When a failure in the steering system is detected, the control / drive circuit unit 26 outputs an instruction signal for connecting the input side steering shaft 3 and the output side steering shaft 10 to the mechanical backup 27.

The mechanical backup 27 connects the input side steering shaft 3 and the output side steering shaft 10 in accordance with instructions from the control / drive circuit unit 26, and ensures transmission of force from the input side steering shaft 3 to the output side steering shaft 10. Mechanism. Here, the control / drive circuit unit 26 normally instructs the mechanical backup 27 not to connect the input side steering shaft 3 and the output side steering shaft 10. If the steering system needs to perform a steering operation without passing through the steering angle sensor 4, the steering torque sensor 5, the steering actuator 8, and the like due to the occurrence of a failure in the steering system, the input side steering shaft 3 and the output side steering shaft 10 is input.
The mechanical backup 27 can be configured by, for example, a cable type steering mechanism.

  FIG. 2 is a perspective view schematically showing the suspension device 1B according to the first embodiment. FIG. 3 is a plan view schematically showing the configuration of the suspension device 1B of FIG. 4 is a front view schematically showing the configuration of the suspension device 1B of FIG. 2, and FIG. 5 is a side view schematically showing the configuration of the suspension device 1B of FIG. FIGS. 6A and 6B are views schematically showing a lower link structure of the suspension device 1B of FIG. 2A, a partial plan view (the left front wheel) and (b) a tire contact surface (the right front wheel).

  As shown in FIGS. 2 to 6, the suspension device 1B is a suspension device that suspends the wheels 17FR and 17FL attached to the wheel hub mechanism WH, and an axle 32 that supports the wheels 17FR and 17FL rotatably. A plurality of link members arranged in the vehicle body width direction from the vehicle body side support portion and connected to the axle carrier 33, and a spring member 34 such as a coil spring.

  The plurality of link members constitute a compression link (compression link member) 37 as a first lower link member constituting a lower link structure, a tension link (tension link member) 38 as a second lower link member, and an upper link structure. A transverse link (transverse link member) 39 as a first upper link member, a compression link (compression link member) 40 as a second upper link member, a tie rod (tie rod member) 15, a strut (spring member 34 and a shock absorber) 41) and a stabilizer 42.

  The compression link 37 and the tension link 38 constituting the lower link structure connect the support portion on the vehicle body side located below the axle 32 and the lower end of the axle carrier 33. In the present embodiment, the compression link 37 and the tension link 38 are I arms made of independent members. The compression link 37 and the tension link 38 are connected to the vehicle body side by one support portion, and are connected to the axle carrier 33 side by one attachment portion. Further, the compression link 37 and the tension link 38 in the present embodiment connect the vehicle body 1A and the axle carrier 33 in a state of crossing each other (hereinafter, an intersection of virtual links formed by the compression link 37 and the tension link 38). It will be referred to as “virtual lower pivot point PL” as appropriate.)

  Of these lower link structures, the compression link 37 is inclined forward with respect to the axle 32 (arranged so that the wheel side support point CPLa is rearward of the axle 32 and the vehicle body side support point CPLb is forward of the axle 32). It is installed. Further, the tension link 38 is inclined backward with respect to the axle 32 (disposed in such a direction that the wheel side support point TSLa is rearward of the axle 32 and the vehicle body side support point TSLb is more forward) than the compression link 37. It is installed.

The reason why such link arrangement is performed will be described below with reference to FIG.
When a centrifugal force directed toward the outer corner of the vehicle body acts on the tire ground contact center point (applying force point) during turning as a turning outer wheel, a lateral force toward the turning center is generated against the centrifugal force. When this lateral force (force inward of the vehicle) is input, the wheel side support point CPLa of the compression link 37 moves outward of the vehicle, and the wheel side support point TSLa of the tension link 38 moves outward of the vehicle. Therefore, it is possible to realize the compliance steer that directs the wheels in the toe-out direction with respect to the input lateral force.

  The tie rod 15 is positioned below the axle 32 and connects the steering rack member 14 and the axle carrier 33. The steering rack member 14 transmits the rotational force (steering force) input from the steering wheel 2 to steer. Generate axial force for Accordingly, the tie rod 15 applies an axial force in the vehicle width direction to the axle carrier 33 in accordance with the rotation of the steering wheel 2, and the wheels 17FR and 17FL are steered through the axle carrier 33.

  In the lower link structure of the suspension device 1B according to the present embodiment, the support point Xa on the wheel side (axle carrier 33 side) of the tie rod 15 is the wheel side support point CPLa, TSLa of the compression link 37 and tension link 38 in the vehicle top view. It is located outside the vehicle width direction. Further, as shown in FIG. 2, the vehicle body side support point Xb of the tie rod 15 (the ball joint position serving as the connecting portion with the end portion of the steering rack member 14) is located on the rear side in the vehicle longitudinal direction from the wheel side support point Xa. positioned. As described above, the wheel side support point CPLa of the compression link 37 is located on the front side in the vehicle longitudinal direction from the wheel center, and the wheel side support point TSLa of the tension link 38 is located on the vehicle longitudinal direction rear side from the wheel center. Further, the vehicle body side support point CPLa of the compression link 37 is rearward in the vehicle longitudinal direction from the wheel side support point TSLa of the tension link 38, and the vehicle body side support point TSb of the tension link 38 is from the wheel side support point CPLa of the compression link 37. It is located on the front side in the vehicle longitudinal direction.

  Because of this link arrangement, in the situation where the vehicle longitudinal force is dominant (such as during turning braking where relatively strong braking is applied), the vehicle longitudinal force input to the tire contact point (vehicle rearward direction) Force), the wheel side support point Xa of the tie rod 15 rotates about the vehicle body side support point Xb and moves outwardly of the vehicle, and the wheel side support point TSa of the tension link 38 moves inward of the vehicle. Further, the wheel side support point CPLa of the compression link 37 moves outward of the vehicle. Therefore, it is possible to realize a compliance steer that directs the wheels in the toe-out direction. That is, the front-rear direction compliance steer of the vehicle can be ensured.

  In the present invention, as shown in FIG. 6 (b), the contact point between the kingpin shaft of the suspension device 1B and the road surface of the kingpin shaft is located within the tire ground contact surface with the steering wheel 2 in the neutral position. I am letting. Further, the caster trail is set so as to be located within the tire ground contact surface. More specifically, in the suspension device 1B in the present embodiment, the caster angle is set to a value close to zero, and the kingpin axis is set so that the caster trail approaches zero. Thereby, the tire twisting torque at the time of turning can be reduced, and the moment around the kingpin axis can be further reduced. The scrub radius is a positive scrub with zero or more. Thereby, since the caster trail corresponding to the scrub radius is generated with respect to the tire skid angle at the time of turning, straight running performance can be ensured.

  In the present invention, the compression link 37 and the tension link 38 constituting the lower link structure connect the vehicle body 1A and the lower end of the axle carrier 33 in a state of crossing each other. This makes it possible to reduce the initial kingpin tilt angle and increase the initial scrub radius toward the positive scrub as compared to a structure in which the compression link 37 and the tension link 38 do not intersect. Therefore, the tire twisting torque at the time of turning can be reduced, and the rack axial force required for turning can be reduced. Furthermore, in the present invention, since the virtual lower pivot point PL moves to the outside of the vehicle body due to the lateral force acting on the wheel at the time of turning, the turning response can be improved, and the turning response can be improved compared to the straight running performance. It is possible to make the suspension configuration important.

(Specific configuration example)
FIG. 7 shows that the upper link structure of the suspension device 1B according to the present invention is configured so that the two link members do not intersect with each other, but the extension lines on the wheel side intersect, and the lower link structure intersects with each other. It is a schematic diagram which shows the example made into the structure to do. FIG. 7A schematically shows the upper link structure, and FIG. 7B schematically shows the lower link structure.
As shown in FIG. 7B, the lower link structure of the suspension device 1B includes the compression link 37 and the tension link 38 having the above-described lower link structure.

  In a suspension device having a cross link structure, when a double pivot system is used in which two lower link members intersect each other, each lower link member rotates as a turning outer wheel by rotating forward about the vehicle body side support point. Steering is possible (the state shown in the dashed line). At this time, the virtual lower pivot point PL is a point where the lower link member intersects. However, since this virtual lower pivot point PL can be formed on the inner side in the vehicle body width direction than the suspension type where the lower link member does not intersect, the initial scrub The radius can be increased in the positive scrub direction. Specifically, the virtual lower pivot point PL is slightly outside the wheel width direction inner end of the wheel 17FL serving as a steered wheel (inside the wheel 17FL), as indicated by a circle in FIG. It is set slightly behind the vehicle from the central axis of 17FL.

  In the lower link structure shown in FIG. 7B, since the rotation angle of the compression link 37 and the rotation angle of the tension link 38 at the time of turning are substantially equal, the virtual lower pivot point PL moves forward in the vehicle front-rear direction. In this case, focusing on the distance from the tire center line WC to the virtual lower pivot point PL in the vehicle width direction in the vehicle plan view, when the upper link structure is configured by an A arm or the like, for example, the virtual upper pivot point PU is seen in the plan view. Since the virtual lower pivot point PL hardly moves outward in the vehicle width direction, a positive scrub state can be maintained. In addition, when the virtual lower pivot point PL moves to the front side in the vehicle front-rear direction from the tire center line, the caster angle increases, the caster trail increases, and straight travel performance can be ensured.

  Therefore, in the suspension device in which the lower link structure is set to a structure in which two link members intersect each other, when the present invention is applied, the scrub radius suppresses a positive scrub change by turning as a turning outer wheel. The rack axial force value can be set smaller than that of a suspension device in which the lower link members do not intersect.

  By the way, it is composed of two I links 37P and 38P in which the distance between the vehicle body side support points is larger than the distance between the wheel side support points where the lower link member does not intersect, and does not intersect on the link member. In the case of a suspension device having a parallel link structure in which the extension lines intersect, the virtual lower pivot point PL is the intersection of the wheel-side extension lines of both links as shown in FIG. Compared to a link structure in which two link members intersect each other, the virtual lower pivot point PL moves to the rear side in the vehicle front-rear direction at the time of turning as a turning outer wheel. For this reason, the caster angle becomes negative, the straight running performance is greatly reduced, the entrainment phenomenon occurs, and the running stability is lowered.

  On the other hand, the transverse link 39 and the compression link 40 that constitute the upper link structure, as shown in FIGS. 4 and 5, intersect the vehicle body side support portion located above the axle 32 and the upper end of the axle carrier 33. Do not connect. In the present embodiment, the transverse link 39 and the compression link 40 are I arms made of independent members.

  The transverse link 39 and the compression link 40 are connected to the vehicle body side by one support portion, and are connected to the axle carrier 33 side by one attachment portion. Furthermore, the transverse link 39 and the compression link 40 in the present embodiment connect the vehicle body 1A and the axle carrier 33 (hereinafter, the intersection of the transverse link 39 and the wheel-side extension line of the compression link 40 is appropriately “virtual upper”. Pivot point PU ").

  Among these upper link structures, the transverse link 39 is installed substantially parallel to the axle, and the wheel side support point TBHa of the transverse link 39 is the vehicle longitudinal direction from the wheel center (axle) in the vehicle plan view. On the rear side, it is on the inner side from the inner end in the vehicle width direction of the wheel. In addition, the compression link 40 is inclined rearward and inward with respect to the axle with respect to the transverse link 39 (arrangement in which the vehicle body side support point CPUb is on the vehicle width direction inner side and the vehicle longitudinal direction rear side with respect to the wheel side support point CPUa). It is installed.

  And the wheel side support point CPUa of the compression link 40 is on the rear side in the vehicle longitudinal direction from the wheel center. Further, the vehicle body side support point TBUb of the transverse link 39 is on the front side in the vehicle front-rear direction with respect to the wheel side support point CPUa of the compression link 40. Further, the vehicle body side support point CPUb of the compression link 40 is on the rear side in the vehicle front-rear direction with respect to the vehicle body side support point TBUb of the transverse link 39.

  Thus, by making the upper link structure a compression-type link structure in which two link members intersect with the extension line on the wheel side, the intersection of the wheel side extension line in a plan view of the transverse link 39 and the compression link 40 Becomes the virtual upper pivot point PU. The virtual upper pivot point PU moves to the rear side in the vehicle front-rear direction and turns outward in the vehicle width direction when the steered wheel 17FL is steered to become a turning outer wheel, contrary to the virtual lower pivot point PL described above. Move slightly.

  For this reason, the virtual lower pivot point PL and the virtual upper pivot point PU move in directions opposite to each other in the vehicle front-rear direction. Therefore, the rearward tilt amount of the kingpin axis KS can be increased as compared with the case where the upper link structure is constituted by an A arm or the like and the upper pivot point does not move in the vehicle front-rear direction in the vehicle plan view. For this reason, a caster angle can be enlarged, a caster trail can also be enlarged, and straight advanceability can be ensured.

Hereinafter, the suspension geometry in the suspension device 1B will be examined in detail.
(Analysis of rack axial force component)
FIG. 9 is a diagram illustrating the relationship between the rack stroke and the rack axial force during steering.
As shown in FIG. 9, the rack axial force component mainly includes a tire twisting torque and a wheel lifting torque, and of these, the tire twisting torque is dominant.
Therefore, the rack axial force can be reduced by reducing the torsional torque of the tire.

(Minimizing tire twisting torque)
FIG. 10 is a diagram illustrating a trajectory of the center of the tire ground contact surface at the time of turning.
In FIG. 10, the case where the movement amount of the tire ground contact surface center at the time of turning is large and the case where it is small are shown together.
From the analysis result of the rack axial force component, in order to reduce the rack axial force, it is effective to minimize the tire twisting torque at the time of turning.
In order to minimize the tire twisting torque at the time of turning, as shown in FIG. 10, the trajectory at the center of the tire contact surface may be made smaller.
That is, the tire torsion torque can be minimized by matching the center of the tire contact surface with the kingpin contact point.
Specifically, it is effective to set the caster trail to 0 mm and the scrub radius to 0 mm or more.

(Effect of kingpin tilt angle)
FIG. 11 is an isoline diagram showing an example of rack axial force distribution in coordinates with the kingpin tilt angle and the scrub radius as axes.
In FIG. 11, the contour lines in the case where the rack axial force is small, medium and large are shown as examples.
As the kingpin tilt angle increases with respect to tire torsion torque input, the rotational moment increases and the rack axial force increases. Therefore, it is desirable to set the kingpin tilt angle to be smaller than a certain value, but from the relationship with the scrub radius, for example, if the kingpin tilt angle is 15 degrees or less, the rack axial force can be reduced to a desired level.

  In addition, the area surrounded by the one-dot chain line (boundary line) in FIG. 11 is smaller than the kingpin inclination angle 15 degrees at which the lateral force can be estimated to exceed the friction limit in the turning limit area, and the viewpoint of the tire twisting torque. Thus, an area having a scrub radius of 0 mm or more is shown. In the present embodiment, this region (the direction in which the kingpin tilt angle decreases from 15 degrees on the horizontal axis and the direction in which the scrub radius increases from zero on the vertical axis) is a region that is more suitable for setting. However, even if the scrub radius is a negative region, a certain effect can be obtained by indicating other conditions in this embodiment.

Specifically, when determining the scrub radius and the kingpin tilt angle, for example, the isoline indicating the rack axial force distribution shown in FIG. 11 is approximated as an nth-order curve (n is an integer of 2 or more), and the one point A value determined by the position of the inflection point (or peak value) of the n-th order curve from the region surrounded by the chain line can be adopted.
In the example shown in FIG. 7, the wheel center moves to the inside of the turn at the time of turning in the vehicle top view. Therefore, the effect of reducing the rack axial force can be further enhanced by positioning the steering rack member 14 in front of the axle as in the present embodiment.

FIG. 12 is a diagram showing the relationship between the toe angle and the scrub radius in the case of the tension type suspension device in which the lower link members do not intersect and the present invention.
As shown in FIG. 12, in the case of the present invention, the scrub radius near the neutral position (toe angle is 0) can be made larger than in the case where the lower link members are not crossed. Moreover, in the direction (the minus direction in FIG. 12) in which the turning angle that becomes the turning outer wheel becomes larger, the scrub radius becomes larger and the rack axial force can be made smaller.
Also, setting the caster angle to 0 degrees can improve the suspension rigidity.

  Further, FIG. 13 is a diagram showing the relationship between the road landing point of the kingpin axis KS and the lateral force. In FIG. 13, road landing points of the kingpin axis KS are denoted by reference numerals 1 to 5 from the vehicle front side to the vehicle rear side. Setting the caster trail to 0 mm means that the road surface landing point of the kingpin axis KS is the tire ground contact point on the tire ground contact surface (the force applied), as shown in FIG. Point) This means that it coincides with O, and this can further improve the lateral force reduction effect.

  Even when the ground contact point of the kingpin axis KS within the tire ground contact surface including the tire ground contact center point (adhesion point) O is the front reference numeral 2 and the rear reference numeral 4, the ground contact point of the kingpin axis KS is the reference sign. As shown by 1 and 5, the lateral force can be reduced as compared with the case where the position is deviated from the tire ground contact surface in the front-rear direction. In particular, the lateral force when the road landing point of the kingpin axis KS is closer to the vehicle front side than the tire ground contact point (force point) O is greater than the lateral force compared to the case where the road contact point is the vehicle rear side from the tire ground contact point (force point) O. Can be suppressed small.

(Ensuring straightness by positive scrub)
FIG. 6B described above is a conceptual diagram for explaining the self-aligning torque in the case of a positive scrub.
As shown in FIG. 6B, the restoring force (self-aligning torque) acting on the tire increases in proportion to the sum of the caster trail and the pneumatic trail.
Here, in the case of positive scrubbing, the distance εc from the wheel center determined by the position of the foot of the perpendicular line drawn from the contact point of the kingpin shaft to the straight line in the direction of the side slip angle β of the tire passing through the tire contact center (FIG. 6B) Can be considered a caster trail.
Therefore, the greater the scrub radius of the positive scrub, the greater the restoring force acting on the tire during turning.

  In the present embodiment, the setting of the kingpin axis is positive scrub and the initial scrub radius can be secured larger than when the lower link member is not crossed. Is reduced. In addition, since the steer-by-wire system is adopted, it is possible to ensure the final straightness by the steered actuator 8.

(Operation of suspension device)
Next, the operation of the suspension device 1B according to this embodiment will be described.
In the suspension device 1B according to the present embodiment, the two lower link members are I-arms. Then, the compression link 37 is installed to extend obliquely rearward from the axle carrier 33 at the rear side in the vehicle front-rear direction and inside the vehicle width direction, and the tension link 38 intersects the compression link 37 from the lower end of the axle carrier 33. It is installed obliquely forward and obliquely forward on the front side in the vehicle front-rear direction and on the inner side in the vehicle width direction. Specifically, the wheel side support point TBLa of the compression link 37 is on the front side in the vehicle front-rear direction with respect to the wheel center, and the wheel side support point TSLa on the tension link 38 is on the rear side in the vehicle front-rear direction with respect to the wheel center. Further, the vehicle body side support point TBLb of the compression link 37 is the rear side in the vehicle longitudinal direction from the wheel side support point TSLa of the tension link 38, and the vehicle body side support point TSLb of the tension link 38 is the wheel side support point TBLa of the compression link 37. Rather than the vehicle front-rear direction.

  In the case of the above suspension structure, as shown in FIG. 6 (a), in the situation where the force in the vehicle front-rear direction is dominant, the force in the vehicle front-rear direction input to the tire contact point (force in the vehicle rearward direction) The wheel side support point CPLa of the compression link 37 moves outward of the vehicle. Further, the wheel side support point TSLa of the tension link 38 moves inward of the vehicle. Therefore, it is possible to realize a compliance steer that directs the wheels in the toe-out direction with respect to the force applied to the vehicle rearward direction.

  In the suspension structure, as shown in FIG. 6A, the wheel side support points TBLa, TSLa and TBUa, TBUb and the vehicle body side support points TBLb, TSLb and TBUb, TSBb for each lower link structure and each upper link structure. Imagine a straight line connecting Then, for the lower link structure, the intersection of these straight lines becomes the virtual lower pivot point PL, and for the upper link structure, the intersection of the extension lines on the wheel side becomes the virtual upper pivot point PU. A straight line connecting these virtual lower pivot points PL and PU is a kingpin axis KS.

In the present embodiment, the kingpin axis KS is set so as to pass through the tire ground contact surface. The kingpin axis KS is set so that the caster trail is located within the tire ground contact surface.
More specifically, for example, the setting of the kingpin axis is a positive scrub having a caster angle of approximately 0 degrees, a caster trail of 0 mm, and a scrub radius of 0 mm or more. The kingpin inclination angle is set within a range where the scrub radius can be a positive scrub and a smaller angle (for example, 15 degrees or less).

By setting it as such a suspension geometry, the locus | trajectory of the tire ground-contact surface center at the time of steering becomes smaller, and a tire torsion torque can be reduced.
Therefore, since the rack axial force can be made smaller, the moment around the kingpin axis can be made smaller, and the output of the steering actuator 8 can be reduced. Moreover, the direction of the wheel can be controlled with a smaller force. That is, maneuverability and stability can be improved.

Further, by installing the lower link members so as to intersect with each other, the support point of the lower link member can be located close to the center of the wheel, so that the weight of the axle carrier 33 can be reduced.
In the suspension device 1B according to the present embodiment, since the two lower link members are installed so as to intersect with each other, the virtual lower pivot point PL is easily arranged on the inner side of the vehicle body than the center of the tire contact surface.
Therefore, it becomes easy to make the kingpin inclination angle close to 0 degrees, and it becomes easy to make the scrub radius large on the positive scrub side.

In addition, since the caster angle is set to approximately 0 degrees and the caster trail is set to 0 mm, there is a possibility that the straight traveling performance on the suspension structure may be affected. By setting the positive scrub, the influence is reduced. . Further, in addition to the control by the steering actuator 8, the straightness is ensured. That is, maneuverability and stability can be improved.
Further, since the turning by the turning actuator 8 is performed for limiting the kingpin tilt angle to a certain range, it is possible to avoid the driver from feeling heavy in the steering operation. Further, the kickback due to the external force from the road surface can be countered by the steering actuator 8, so that the influence on the driver can be avoided. That is, maneuverability and stability can be improved.

As described above, in the suspension device 1B according to the present embodiment, with respect to the lower link structure, the compression link 37 is installed to extend obliquely rearward in the vehicle width direction on the front side in the vehicle front-rear direction, and in the vehicle plan view, the tension link 38 is arranged so as to intersect with the compression link 37. Therefore, the virtual lower pivot point PL can be brought closer to the inside of the vehicle body in the vehicle width direction. The kingpin axis KS defined by the virtual lower pivot point PL and the virtual upper pivot point PU is assumed to have a small kingpin tilt angle, and the kingpin axis KS passes through the tire contact surface and the caster trail is located within the tire contact surface. Because of the setting, the moment around the kingpin axis can be made smaller.
Accordingly, the steering can be performed with a smaller rack axial force, and the direction of the wheel can be controlled with a smaller force, so that the maneuverability and stability can be improved.

Further, since the moment around the kingpin axis can be further reduced, the load applied to the steering rack member 14 and the tie rod 15 can be reduced, and the member can be simplified.
Further, as the steering actuator 8 for realizing the steer-by-wire, one having a lower driving ability can be used, and the cost and weight of the vehicle can be reduced.
For example, when compared with a conventional steer-by-wire type suspension device, the configuration of the present invention is approximately 10% in weight and approximately 50% in cost mainly due to simplification of the lower link member and downsizing of the steering actuator 8. Can be reduced.

In addition, since the caster trail increases at the time of turning, it is possible to suppress the turning angle from increasing when turning at a high lateral acceleration.
Further, at the time of turning as a turning outer wheel, the virtual lower pivot point PL of the lower link structure moves forward in the vehicle front-rear direction, so that the caster angle can be increased and the caster trail can be increased. Moreover, since the virtual upper pivot point PU of the upper link moves to the rear side in the vehicle front-rear direction contrary to the virtual lower pivot point PL, the caster angle can be increased and the caster trail can be increased. Straightness is improved and there is an understeer tendency.

  Therefore, by controlling the steered actuator 8 by the steer-by-wire system SBW, the necessary steered force can be supplied to the steered wheels 17FL and the tie rods 15 by controlling the steerable wire system SBW. It is possible to obtain good steering characteristics by transmitting to 17FR.

On the other hand, in the steer-by-wire system SBW, when it is not possible to control the turning actuator 8 by detecting an abnormality by an abnormality determination process described later, it is possible to improve straightness in the turning state that becomes the turning outer wheel, Driving stability can be ensured. Therefore, it is possible to ensure traveling stability during turning and to exhibit a fail-safe function.
In addition to this, the caster angle increases at the time of turning as a turning outer wheel, but the change in the positive scrub and the change in the kingpin tilt angle can be suppressed, and the change in the straightness due to these can be suppressed.

  Further, in the present invention, the compression link 37 is installed to extend obliquely rearward of the vehicle as a lower link structure, and the wheel-side support point CPLa of the compression link 37 is set to the front side in the vehicle longitudinal direction from the wheel center in the vehicle plan view. It is said. Further, the tension link 38 is installed to extend obliquely forward of the vehicle with respect to the axle, contrary to the compression link 37. And the wheel side support point TSLa of the tension link 38 is set to the rear side in the vehicle longitudinal direction from the wheel center. Further, the vehicle body side support point CPLb of the compression link 37 is set to the front side in the vehicle front-rear direction with respect to the wheel side support point TSLa of the tension link 38. Further, the vehicle body side support point TSLb of the tension link 38 is set to the front side in the vehicle front-rear direction with respect to the wheel side support point CPLa of the compression link 37.

  In the case of such a lower link structure, when a lateral force (force inward to the applied force point on the rear side of the wheel center) is input to the wheel, the wheel side support point CPLa of the compression link 37 moves outward in the vehicle. The wheel side support point TSLa of the tension link 38 moves inward of the vehicle. Therefore, it is possible to realize the compliance steer that directs the wheels in the toe-out direction with respect to the input lateral force.

Further, in the present invention, the wheel side support point Xa of the tie rod 15 is positioned on the outer side in the vehicle width direction than the wheel side support points CPLa and TSLa of the compression link 37 and the tension link 38. Further, the vehicle body side support point Xb of the tie rod 15 is located on the rear side in the vehicle longitudinal direction with respect to the wheel side support point Xa.
In such a link arrangement, in a situation where the vehicle longitudinal force is dominant, the wheel side support point of the compression link 37 against the vehicle longitudinal force (force toward the vehicle rear) input to the tire contact point. TBLa moves outward of the vehicle. Further, the wheel side support point Xa of the tie rod 15 rotates around the vehicle body side support point Xb and moves outward, and the wheel side support point TSLa of the tension link 38 moves inward in the vehicle width direction. Therefore, it is possible to realize a compliance steer that directs the wheels in the toe-out direction.
Therefore, according to the present invention, in the vehicle suspension device, the compliance steer characteristic with respect to the force in the vehicle longitudinal direction can be made more appropriate.

FIG. 14 is a diagram showing (a) lateral force compliance steer and (b) lateral stiffness in the suspension device 1B according to the present invention and a comparative example.
In FIG. 14, as a comparative example, a suspension having a lower link structure in which the lower link members do not intersect is assumed.
As shown in FIG. 14, in the case of the configuration of the suspension device 1B according to the present invention (solid line in FIG. 14), the lateral force compliance steer is improved by 35% compared to the comparative example (broken line in FIG. 14). The stiffness is improved by 29%.

FIG. 15 is a diagram showing the longitudinal force compliance steer in the suspension device 1B according to the present invention and the comparative example.
In FIG. 15, as a comparative example, a suspension having a lower link structure in which the lower link members do not intersect is assumed.
As shown in FIG. 15, in the case of the configuration of the suspension device 1B according to the present invention (solid line in FIG. 15), the longitudinal force compliance steer is improved by 28% compared to the comparative example (broken line in FIG. 15). .

  In this embodiment, the wheels 17FR, 17FL, 17RR, and 17RL correspond to the tire wheel, the tire, and the wheel hub mechanism WH, the compression links 37 and 40 correspond to the compression link member, and the tension link 38 corresponds to the tension link member. The transverse link 39 corresponds to the transverse link member, and the link arrangement of the compression link 37 and the tension link 38 constituting the lower link structure and the link arrangement of the transverse link 39 and the compression link 40 constituting the upper link structure are pivoted. Corresponds to the point mover. The tie rod 15 corresponds to the tie rod.

(Specific configuration example of control / drive circuit)
Next, specific configuration examples for realizing the control / drive device unit 26 will be described with reference to FIGS.
As shown in FIG. 16, the control / drive device unit 26 includes a steering control device 50. The steering control device 50 includes a target turning angle calculation unit 51, a turning angle control unit 52, a straightness complementation unit 53, a disturbance compensation unit 54, a delay control unit 56, a turning angle deviation calculation unit 58, a turning motor. A control unit 59, a current deviation calculation unit 60, and a motor current control unit 62 are provided.

The target turning angle calculation unit 51 receives the vehicle speed V and the steering angle θs detected by the steering angle sensor 4, and calculates the target turning angle δ ^* based on these.
The steered angle control unit 52 calculates the change amounts Δfl and Δfr of the steered angles of the steered wheels 17FL and 17FR due to compliance steer. These change amounts Δfl and Δfr constitute a lower link structure with the driving forces TL and TR of the left and right wheels outputted from the driving force control device 64 that distributes and controls the driving forces of the steered wheels 17FL and 17FR that are the left and right driving wheels. Based on the compliance steer coefficient af corresponding to the bending of the bush of the compression link 37 and the tension link 38, the calculation is performed by performing the following expressions (1) and (2). Then, a displacement amount difference between the calculated displacement amounts Δfl and Δfr is calculated to calculate a compliance steer control value Ac (= Δfl−Δfr) as a turning angle control value.
Δfl = af · TL (1)
Δfr = af · TR (2)

The straightness complementation unit 53 receives the left and right wheel driving forces TL and TR from the driving force control device 64 that controls the distribution of the driving wheel driving force, and the steering torque detected by the steering torque sensor 5. Ts is input, and based on these, the self-aligning torque Tsa is calculated. The calculated self-aligning torque Tsa is multiplied by a predetermined steering angle correction gain Ksa, and the self-aligning torque control value Asa ( = Ksa · Tsa).
Here, the calculation of the self-aligning torque Tsa in the linearity complementing unit 53 first calculates the driving force difference ΔT (= TL−TR) between the driving forces TR and TL of the left and right wheels, and calculates the calculated driving force difference ΔT. Based on the generated torque estimation control map shown in FIG. 17, the generated torque Th generated at the time of turning by the torque steer phenomenon is estimated.

  The generated torque estimation control map is set for a vehicle whose scrub radius is positive, that is, positive scrub. As shown in FIG. 17, the generated torque estimation control map has the driving force difference ΔT on the horizontal axis and the generated torque Th on the vertical axis, and the driving force difference ΔT increases from zero to the positive direction. When the driving force TL increases above the right wheel driving force TR, the generated torque Th is set so as to increase in proportion to this in the direction of turning the vehicle to the right (positive direction).

On the other hand, when the driving force difference ΔT increases from zero to the negative direction, that is, when the right wheel driving force TR increases above the left wheel driving force TL, the generated torque Th is proportionally proportional to the direction in which the vehicle turns left from zero. It is set to increase in the (negative direction).
The straightness complementing unit 53 calculates the self-aligning torque Tsa by subtracting the generated torque Th from the steering torque Ts detected by the steering torque sensor 5.
The calculation of the self-aligning torque Tsa is not limited to the calculation based on the left and right driving force difference ΔT as described above, and can be similarly calculated based on the left and right braking force difference.

For calculating the self-aligning torque Tsa, a yaw rate sensor for detecting the yaw rate γ of the vehicle and a lateral acceleration sensor for detecting the lateral acceleration Gy of the vehicle are provided, and the differential value of the yaw rate and the lateral acceleration Gy based on the motion equation of the vehicle are provided. Is calculated by multiplying the lateral force Fy by the pneumatic trail εn.
Furthermore, the steering angle θs detected by the steering angle sensor 4 with reference to a control map obtained by actually measuring the relationship between the steering angle θs of the steering wheel 2 and the self-aligning torque Tsa using the vehicle speed V as a parameter or by simulation. The self-aligning torque Tsa can be calculated based on the vehicle speed V.

Further, the inertia J and static friction coefficient Fr of the electric motor constituting the steering actuator are obtained in advance as constants, the motor angular velocity ωm, the motor angular acceleration αm, and the steering torque Tm generated by the steering actuator 8 (proportional to the steering torque Tm). The self-aligning torque Tsa may be calculated by performing the following expression based on a target drive current im ^* described later.
Tsa (s) = Tm (s) −J · αm (s) −Fr · sin (ωm (s))
Here, s is a Laplace operator.

The disturbance compensation unit 54 receives the steering torque Ts from the steering torque sensor 5, the rotation angle θmo from the steering actuator angle sensor 9, and the motor current imr from the motor current detection unit 61. A disturbance compensation value Adis for suppressing these disturbances is calculated by separating and estimating each frequency band.
In the disturbance compensator 54, as described in, for example, Japanese Patent Laid-Open No. 2007-237840, the steering torque Ts that is a steering input by the driver and the steering input by the steering actuator 8 are used as control inputs, and actual steering is performed. In a model using a state quantity as a control quantity, a disturbance is estimated based on a difference between a value obtained by passing the control input through a low-pass filter and a value obtained by passing the control quantity through an inverse characteristic of the model and the low-pass filter. A plurality of disturbance estimation units; Each disturbance estimation part isolate | separates a disturbance for every several frequency band by varying the cutoff frequency of a low-pass filter.

Then, the disturbance compensation value Adis and the self-aligning torque control value Asa calculated by the disturbance compensation unit 54 and the straightness complementing unit 53 are added by the adder 55a. The addition output of the adder 55a and the compliance steer control value Ac calculated by the turning angle control unit 52 are added by the adder 55b to calculate the straight travel guarantee control value δa. The straightness guarantee ensuring control value δa is supplied to the delay control unit 56.
Here, as shown in FIG. 16, the turning angle control unit 52, the straightness complementing unit 53, the disturbance compensation unit 54, and the adders 55a and 55b constitute a straightness guaranteeing unit SG. A steering response setting unit SRS is configured with the delay control unit 56 described below.

As shown in FIG. 16, the delay control unit 56 includes a steering start detection unit 56a, a monostable circuit 56b, a gain adjustment unit 56c, and a multiplier 56d.
The steering start detection unit 56a detects the timing of right steering or left steering from the state where the neutral position is maintained based on the steering angle θs detected by the steering angle sensor 4, and indicates the steering start signal SS indicating the steering start from the neutral state. Is output to the monostable circuit 56b.

Further, the monostable circuit 56b outputs a control start delay signal that is turned on for a predetermined delay time, for example, 0.1 seconds, to the gain adjustment unit 56c based on the steering start signal output from the steering start detection unit 56a.
The gain adjustment unit 56c sets the control gain Ga to “0” when the control start delay signal is on, and sets the control gain Ga to “1” when the control start delay signal is off. The set control gain Ga is output to the multiplier 56d.
In the multiplier 56d, the straightness ensuring control value δa output from the straightness ensuring unit SG is input, the straightness ensuring control value δa is multiplied by the control gain Ga, and the multiplication result is output from the target turning angle calculating unit 51. Is supplied to the adder 56e to which the target turning angle δ ^* is input.

Therefore, when the delay control unit 56 detects the steering start state in which the right steering or the left steering is performed from the state where the neutral state is maintained by the steering start detection unit 56a, the straight traveling calculated by the straight travel guarantee unit SG is performed. In the gain adjusting unit 56c, the straightness ensuring control value is set so that the straightness ensuring control for adding the property ensuring control value δa to the target turning angle δ ^* is stopped for a predetermined time set by the monostable circuit 56b, for example, 0.1 second. A control gain Ga to be multiplied by δa is set to “0”. Then, when the output signal of the monostable circuit 56b is inverted to the OFF state after 0.1 seconds have elapsed, the gain adjusting unit 56c adds the straight travel guarantee ensuring control value δa to the target turning angle δ ^* . The control gain Ga is set to “1” so as to start straight travel guarantee control.

Further, when the steering wheel 2 is continuously steered, the delay control unit 56 does not detect the steering start from the neutral state by the steering start detection unit 56a, so that the output of the monostable circuit 56b maintains the off state. Thus, the control gain Ga is set to “1” by the gain adjusting unit 56c. For this reason, the straightness ensuring control value δa calculated by the straightness ensuring unit SG is supplied to the adder 56e as it is. Therefore, the straight traveling guarantee control is performed by adding the multiplication device Ga · δa of the straight traveling guarantee control value δa and the control gain Ga to the target turning angle δ ^* .

The turning angle deviation calculating unit 58 turns the actuator 8 from the added target turning angle δ ^* a obtained by adding the straight traveling guarantee control value δa to the target turning angle δ ^* output from the adder 56c. The actual turning angle δr output from the turning actuator angle sensor 9 of the motor 8 a is subtracted to calculate the turning angle deviation Δδ, and the calculated turning angle deviation Δδ is output to the turning motor control unit 59.
The steered motor control unit 59 calculates the target drive current im ^* of the steered motor 8a constituting the actuator 8 so that the input angle deviation Δδ becomes zero, and calculates the calculated target drive current im ^* as a current deviation. To the unit 60.

The current deviation calculator 60 subtracts the actual motor drive current imr output from the motor current detector 61 that detects the motor current supplied to the steered motor 8a constituting the actuator 8 from the input target drive current im ^*. The current deviation Δi is calculated and the calculated current deviation Δi is output to the motor current control unit 62.
The motor current control unit 62 performs feedback control so that the input current deviation Δi becomes zero, that is, the actual motor drive current imr follows the target drive current im ^*, and steers the actual motor drive current imr. Output to the motor 8a.

Here, the turning angle deviation calculation unit 58, the turning motor control unit 59, the current deviation calculation unit 60, the motor current detection unit 61, and the motor current control unit 62 constitute an actuator control device 63. The actuator control device 63 controls the rotation angle δr detected by the turning actuator angle sensor 9 that detects the rotation angle of the turning motor 8a constituting the turning actuator 8 so as to coincide with the target turning angle δ ^*. . Therefore, when the vehicle is traveling straight and the target turning angle δ ^* becomes “0”, the control is performed so that the rotation angle δr coincides with the target turning angle δ ^* . When the straight traveling collateral part SG is used as the main straight traveling collateral part, the secondary straight traveling collateral part is configured.

(Operation of the steering control device)
Next, the operation of the steering control device in the first embodiment will be described with reference to FIGS.
Now, it is assumed that the vehicle is traveling straight with the steering wheel 2 held in a neutral position.
In this straight traveling state, the target turning angle δ ^* calculated by the target turning angle calculation unit 51 is zero. At this time, since the steering wheel 2 holds the neutral position, the driving force or braking force of the steered wheels 17FL and 17FR that are the left and right drive wheels becomes equal. Therefore, the steering angle displacements Δfl and Δfr of the steered wheels 17FL and 17FR by the compliance steer calculated by the steered angle control unit 52 by the above formulas (1) and (2) are equal values. Therefore, since the compliance steer correction amount Ac is a value obtained by subtracting the displacement amount Δfr from the displacement amount Δfl, the compliance steer correction amount Ac is zero.

  Similarly, since the driving force TL and TR are also equal in the straight travel complementing unit 53, the generated torque Th calculated with reference to the generated torque estimation control map shown in FIG. It becomes zero. On the other hand, since the steering wheel 2 is not steered in the straight traveling state, the steering torque Ts is also zero, the self-aligning torque Tsa is also zero, and the self-aligning torque control value Asa is also zero.

On the other hand, the disturbance compensation unit 54 calculates a circulation compensation value Adis that suppresses the disturbance. Therefore, the straight travel guarantee value δa is only the circulation compensation value Adis. The straightness ensuring control value δa is supplied to the multiplier 56 d of the delay control unit 56.
In the delay control unit 56, since the steering start is not detected by the steering start detection unit 56a, the output of the monostable circuit 56b maintains the off state. For this reason, the gain adjustment unit 56c sets the control gain Ga to “1”, and this control gain Ga is supplied to the multiplier 56d. From the multiplier 56d, the straight traveling performance ensuring control value δa is supplied to the adder 56e as it is and added to the zero target turning angle δ ^* . Therefore, the added target turning angle δ ^* a corresponding to the disturbance compensation value Adis is calculated, and the turning angle of the turning motor 8a of the actuator 8 is controlled so as to coincide with the added target turning angle δ ^* a. Is done. For this reason, it is possible to perform straight traveling without the influence of disturbance.

Therefore, when the front wheels 17FR and 17FL are steered due to disturbance caused by input from the road surface due to differences in road surface steps and road surface friction coefficients of the front wheels 17FR and 17FL, the turning actuator 8 is rotated. In response to this, when the rotation angle θmo detected by the steering actuator angle sensor 9 changes, a disturbance compensation value Adis corresponding to the change in the rotation angle θmo is output.
For this reason, the turning actuator 8 is controlled in accordance with the disturbance compensation value Adis, and it is possible to generate torque against turning by the road surface input of the suspension device 1B. Accordingly, the straight travel performance of the suspension device 1B can be secured by the straight travel performance securing section SG.

In addition, when the vehicle is traveling straight and no disturbance is detected by the disturbance compensator 54, the straight travel guarantee control value δa calculated by the straight travel guarantee unit SG becomes zero, and the target turning angle calculator 51 Since the target turning angle δ ^* output from is also zero, the post-addition target turning angle δ ^* output from the adder 56e is also zero.
For this reason, when the turning angle displacement occurs in the turning motor 8a constituting the turning actuator 8 by the actuator control device 63, the actuator control device 63 outputs the motor current imr so as to eliminate the turning angle displacement. Therefore, the steered wheels 17FR and 16FL are returned to the steered angle in the straight traveling state. Therefore, straightness can be secured by the actuator control device 63.

However, when the steering wheel 2 is steered to the right (or left) from the state of maintaining the straight traveling state in which the steering wheel 2 is held at the neutral position, the transition from the straight traveling state to the turning state by steering is performed. Is detected by the steering start detector 56a.
For this reason, the control delay signal that is in the ON state for a predetermined time, for example, 0.1 second is output from the monostable circuit 56b to the gain adjusting unit 56c. Accordingly, the gain adjustment unit 56c sets the control gain Ga to “0” while the control delay signal is kept on. For this reason, the multiplication output outputted from the multiplier 56d becomes “0”, and the output of the straightness ensuring control value δa to the adder 56e is stopped.
Therefore, since the control gain Ga is set to “0” during the initial response period T1 of 0.1 seconds from the time when the steering is started from the neutral position of the steering wheel 2, the multiplication output output from the multiplier 56d is It becomes “0”, and the straight traveling guarantee control with respect to the target turning angle δ ^* is stopped as shown by the solid line in FIG.

Therefore, the steering angle θs detected by the steering angle sensor 4 is supplied to the target turning angle calculation unit 51, and the target turning angle δ ^* calculated by the target turning angle calculation unit 51 is directly calculated as the turning angle deviation. Supplied to the unit 58. For this reason, the turning motor 8a is rotationally driven so as to coincide with the target turning angle δ ^* . During this time, the straightness guaranteeing control in the straightness guaranteeing part SG is stopped.
Therefore, in the initial response period T1, the turning by the suspension device 1B in which the road contact point of the kingpin axis KS is set at the contact center position in the contact surface of the tire and the caster angle is set to zero is started.

At this time, the caster angle of the suspension device 1B is set to zero. As shown in FIG. 18A, when the caster angle is substantially zero, the relationship between the caster angle, the steering response, and the steering stability is high (the steering response is emphasized). However, steering stability cannot be ensured, that is, there is a trade-off relationship between the steering response to the caster angle and the steering stability.
For this reason, in the initial state where the steering is started from the neutral position, the straight travel guarantee control by the steer-by-wire control is not executed, and the suspension device 1B covers this initial turning.

  In the initial response period T1, as described above, the suspension device 1B has a caster angle of approximately zero and has high steering response. Therefore, as shown by the characteristic line L1 shown by the solid line in FIG. A steered response characteristic (yaw rate) higher than a steered response characteristic (yaw rate) in a vehicle having a general steer-by-wire type steering system indicated by a characteristic line L2 shown by a chain line can be obtained. At this time, since the turning angle changes corresponding to the steering angle change caused by the steering of the steering wheel 2 by the driver, the driver does not feel uncomfortable.

However, if the steering is continued beyond the initial response period T1 only by the steering response by the suspension device 1B, the middle response period T2 and the late response period T3 as shown by the characteristic line L3 in FIG. 19 (a). The steering response of the vehicle by steering becomes sensitive. In addition, a phenomenon of getting inside the vehicle from the middle response period T2 to the later response period T3 becomes large.
For this reason, in the first embodiment, as shown in FIG. 19B, for example, 0.1 seconds after the initial response period T1 elapses, the turning angle control unit 52, the straightness complementation unit 53, and the disturbance compensation unit. The straightness ensuring control with respect to the target turning angle δ ^* by the straightness ensuring part SG constituted by 54 is started stepwise. For this reason, the steering response of the vehicle by the suspension device 1B is suppressed to suppress the vehicle wobble, and the straightness of the suspension device 1B is complemented by the steer-by-wire control as shown by the dotted line in FIG. 18B. Thus, the steering stability can be ensured.

  Thereafter, after 0.3 seconds have elapsed, for example, when the medium-term response period T2 ends, the steering response characteristic is further suppressed even when compared with a general vehicle steering response characteristic by the straight traveling security control by the straight traveling security unit SG. And understeer tendency. Thereby, as shown by the characteristic line L1 shown by the solid line in FIG. 19A, the steering stability can be improved, and an ideal vehicle turning response characteristic indicated by the characteristic line L1 can be realized. .

As described above, according to the vehicle steering apparatus of the present embodiment, in the suspension apparatus 1B, the compression link 37 serving as the first lower link member and the tension serving as the second lower link member constituting the lower link structure. Since the link 38 intersects with the vehicle top view, the kingpin axis KS passes through the tire contact surface with the steering wheel in the neutral position, and the caster trail is set in the tire contact surface. The moment around KS can be further reduced.
Therefore, also in the first embodiment, the steering can be performed with a smaller rack axial force, and the direction of the wheel can be controlled with a smaller force. That is, maneuverability and stability can be improved.

  As described above, in the first embodiment, by setting at least the kingpin axis KS to pass through the tire contact surface, the suspension device 1B itself is configured to improve the steering response, in addition to this. The straightness of the suspension device 1B is secured by performing the turning angle control for controlling the turning characteristics, the straightness compensation, and the disturbance compensation by the straightness guaranteeing part SG of the steer-by-wire system SBW.

  Therefore, when right or left steering is performed from the state where the steering wheel 2 is held at the neutral position, high response is secured by using the high steering response of the suspension device 1B itself in the initial response period T1. To do. Thereafter, when the initial response period T1 has passed and the medium-term response period T2 is entered, it is necessary to focus on steering stability rather than focusing on the steering response, and the gain adjustment unit of the delay control unit 56 in the steer-by-wire system SBW When the control gain Ga is set to “1” in 56c, the straightness ensuring control using the straightness ensuring control value δa calculated by the straightness ensuring unit SG is started.

  For this reason, by starting straightness guarantee control such as turning angle control, straightness complementation, and disturbance compensation, high steering response by the suspension device 1B is suppressed to ensure steering stability. Further, in the late response period T3, ideal steering response control is established by further reducing the steering response to further suppress the vehicle wobble as an understeer tendency so as to suppress the phenomenon of inward entrainment of the vehicle. can do.

Furthermore, the turning angle control unit 52 can be provided to perform straight travel guarantee control in consideration of the displacement amount of the steered wheels 17FL and 17FR due to compliance steer. For this reason, it is possible to set the rigidity of the bush inserted between the compression link 37 and the tension link 38 constituting the lower link structure on the support portion on the vehicle body 1A side, and through the compression link 37 and the tension link 38 Riding comfort can be improved by reducing the vibration transmission rate from the road surface to the vehicle body 1A.
In addition, in the said 1st Embodiment, although the case where the steering control apparatus 50 was comprised with hardware was demonstrated, it is not limited to this, For example, the target turning angle calculating part 51, the straight travel guarantee part SG May be constituted by an arithmetic processing unit such as a microcomputer, and the steering control process shown in FIG. 20 may be executed by this arithmetic processing unit.

  As shown in FIG. 20, in this steering control process, first, in step S1, the vehicle speed V, the steering angle θs detected by the steering angle sensor 4, the rotation angle θmo detected by the steering actuator angle sensor 9, and the driving force control are performed. Data necessary for calculation processing such as the driving forces TL and TR of the left and right wheels of the device 64 and the steering torque Ts detected by the steering torque sensor 5 are read. Next, the process proceeds to step S2, and whether or not the steering wheel 2 is steered to the right or left from the state in which the steering wheel 2 holds the neutral position based on the steering angle θs detected by the steering angle sensor 4 is determined. When the steering is not started, the process proceeds to step S3.

In this step S4, it is determined whether or not the control flag F indicating the steering start control state is set to “1”. When the control flag F is reset to “0”, the process proceeds to step S4. Then, after the control gain Ga is set to “1”, the process proceeds to step S5.
In this step S5, the target turning angle δ ^* is calculated based on the vehicle speed V and the steering angle θs as in the target turning angle calculation unit 51 described above.

Subsequently, the process proceeds to step S6, and similarly to the above-described turning angle control unit 52, the left and right wheel driving forces TL and TR are multiplied by the compliance steer coefficient sf, and the displacement amount Δfl of the steered wheels 17FL and 17FR due to the compliance steer. And Δfr are calculated, and based on these, the compliance steer control value Ac is calculated.
Next, the process proceeds to step S7, and the generated torque estimation control shown in FIG. Referring to the map, the generated torque Th generated at the time of turning due to the torque steer phenomenon is estimated. Then, the generated torque Th is subtracted from the steering torque Ts to calculate the self-aligning torque Tsa, and the self-aligning torque Tsa is multiplied by a predetermined gain Ksa to calculate the self-aligning torque control value Asa.

Next, the process proceeds to step S8, and the disturbance input to the vehicle is determined for each frequency band based on the motor rotation angle θmo from the steering actuator angle sensor 9, the steering torque Ts, and the motor current imr detected by the motor current detection unit 61. The disturbance compensation value Adis for suppressing these disturbances is calculated.
Next, the process proceeds to step S9, and the following equation (3) is calculated based on the target turning angle δ ^* , the compliance steer control value Ac, the self-aligning torque control value Asa, and the disturbance compensation value Adis. To calculate the target turning angle δ ^* a after addition.
δ ^* a = δ ^* + Ga (Ac + Asa + Adis) (3)

Next, the process proceeds to step S10, where the added target turning angle δ ^* a calculated in step S9 is output to the turning angle deviation calculating unit 58 in FIG. 16, and then the process returns to step S1.
When the determination result in step S2 is the steering start state, the process proceeds to step S11, the control flag F is set to “1”, and then the process proceeds to step S12. Furthermore, when the determination result of step S3 is that the control flag F is set to “1”, the process directly proceeds to step S12.
In step S12, it is determined whether a preset delay time (for example, 0.1 second) has elapsed. At this time, when the delay time has not elapsed, the process proceeds to step S13, the control gain Ga is set to “0”, then the process proceeds to step S5, and the target turning angle δ ^* is calculated.

If the determination result in step S12 is that a predetermined delay time (for example, 0.1 second) has elapsed, the process proceeds to step S14, the control flag F is reset to “0”, and then the process proceeds to step S4. Thus, the control gain Ga is set to “1”.
Even in the steering command angle calculation process shown in FIG. 20, when the steering wheel 2 is not in the steering start state in which the steering is started from the neutral position to the right or left, the target steering angle δ ^* is complied with. Straightness ensuring control is performed in which the straightness ensuring control value δa obtained by adding the steering control value Ac, the self-aligning torque control value Asa, and the disturbance compensation value Adis is added to the target turning angle δ ^* .

On the other hand, when the steering wheel 2 is in the steering start state in which steering is started to the right or left from the state where the steering wheel 2 is held at the neutral position, the control gain Ga is set until the preset delay time elapses. Since it is set to “0”, the straight travel guarantee control is stopped. For this reason, only the target turning angle δ ^* is output to the turning angle deviation calculating unit 58, and thereby the turning motor 8 a constituting the turning actuator 8 is rotationally driven. For this reason, the high turning response of the suspension device itself is set as the initial turning response, and the high turning response can be obtained.

Thereafter, when the delay time elapses, the control gain Ga is set to “1”, so that the compliance steering control value Ac, the self-aligning torque control value Asa, and the disturbance compensation value Adis are added to the target turning angle δ ^* . The turning motor 8a constituting the turning actuator 8 is rotationally driven by a value obtained by adding the straight traveling guarantee control value δa to the target turning angle δ ^* . For this reason, the high steering response of the suspension device 1B is suppressed, and the straightness of the suspension device 1B is ensured, and an ideal steering response characteristic can be obtained.

Even in this steering control process, when the vehicle is traveling straight, the target turning angle δ ^* is zero, and when no disturbance occurs, the target turning angle δ ^* is directly converted into the turning angle deviation calculation unit of FIG. 58, the straightness is ensured by the actuator control device 63 as described above.
In the steering control process of FIG. 20, the process of step S5 corresponds to the target steering angle calculation unit 51, the process of step S6 corresponds to the steering angle control unit 52, and the process of step S7 is a straightness complementing unit. 53, the processing of steps S5 to S7 corresponds to the straight traveling guarantee unit SG: the processing of steps S2 to S4 and S11 to S14 corresponds to the delay control unit 56, and the processing of steps S2 to S14 is a turning response. This corresponds to the sex setting unit SRS.

In the steering control process of FIG. 20, the case where the control performed by the steering response setting unit SRS is realized by software is described. However, the present invention is not limited to this, and the control of the steering response setting unit SRS is performed. In addition, software processing including control of the actuator control device 63 may be performed.
Moreover, in the said 1st Embodiment, although the case where the straight travel guarantee part SG was comprised by the turning angle control part 52, the straight travel complementation part 53, and the disturbance compensation part 54 was demonstrated, it is not limited to this. Any one or two of the turning angle control unit 52, the rectilinearity complementing unit 53, and the disturbance compensation unit 54 may be omitted.

  By the way, in the state in which the steer-by-wire system SBW is operating normally as described above, the steering response characteristic setting unit SRS compensates for the reduction in straightness due to the configuration that emphasizes the steering response characteristic of the vehicle suspension apparatus 1B. However, if an abnormality occurs in the steer-by-wire system SBW, the straightness of the vehicle suspension apparatus 1B cannot be complemented to ensure the straightness.

  That is, in the control / drive circuit unit 26, the rotation angle of the input side steering shaft 3 detected by the steering angle sensor 4, the rotation angle of the steering reaction force actuator 6 detected by the steering reaction force actuator angle sensor 7, the steering actuator The rotation angle of the steering actuator 8 detected by the angle sensor 9 and the rotation angle of the pinion gear 12 detected by the pinion angle sensor 13 are monitored, and the occurrence of a failure in the steering system is detected based on these relationships. be able to.

When the occurrence of this failure is detected, the mechanical backup 27 is activated by the control / drive circuit unit 26. For this reason, the input side steering shaft 3 and the output side steering shaft 10 are connected by the mechanical backup 27 to ensure the transmission of force from the input side steering shaft 3 to the output side steering shaft 10.
When the steering wheel 2 is steered by turning the steering wheel 2 in the state where this failure occurs, for example, by turning the steered wheel 17FL toward the outer turning wheel, the steered wheel 17FL starts to steer. In the structure, the virtual lower pivot point PL represented by the intersection of the compression link 37 and the tension link 38 moves to the front side of the vehicle as shown in FIG. 7B.

On the other hand, in the upper link structure, the virtual upper pivot point PU represented by the intersection point of the extension line on the steered wheel 17FL side in the plan view of the transverse link 39 and the compression link 40 is slightly in the vehicle width direction on the rear side in the vehicle front-rear direction. It moves to the diagonally rear side that becomes the outside.
Therefore, the inclination angle of the kingpin axis KS connecting the virtual lower pivot point PL and the virtual upper pivot point PU in the vehicle side view is increased, the caster angle is increased, and the caster trail is increased accordingly. For this reason, it becomes possible to ensure straightness.

At this time, since the virtual upper pivot point PU slightly moves inward in the vehicle width direction, the kingpin inclination angle becomes small, so that the force entering the steering rack member 14 can be reduced at the same time.
As a result, even when a failure occurs in the steer-by-wire system SBW, when the toe angle is generated in the steered wheels 17FL and 17FR, it is possible to ensure straight travel performance and to exhibit a fail-safe function.

Further, the control / drive circuit unit 26 executes the abnormality determination process shown in FIG. 21 as a timer interrupt process at predetermined time intervals (for example, about 10 μsec).
In this abnormality determination process, first, in step S21, the steering angle θs detected by the steering angle sensor 4, the steering reaction force actuator rotation angle θr detected by the steering reaction force actuator angle sensor 7, and the turning actuator angle sensor 9 are detected. After reading the turning actuator rotation angle θa and the pinion gear rotation angle θp detected by the pinion angle sensor 13, the process proceeds to step S22.

  In this step S22, the steering reaction force actuator angle θr is converted into the steering input angle θi, and then the process proceeds to step S23, where the turning actuator rotation angle θa is converted into the turning angle δr1 of the steered wheels 17FR and 17FL, and then The process proceeds to step S24, the pinion gear rotation angle θp is converted into the steered angle δr2 of the steered wheels 17FR and 17FL, and then the process proceeds to step S25.

  In this step S25, it is determined whether or not the absolute value | θs−θi | of the steering angle deviation obtained by subtracting the steering input angle θi from the steering angle θs exceeds a preset threshold θth, and | θs−θi |> θth. If it is, it is determined that the steering angle sensor 4 or the steering reaction force control system is abnormal, and the process proceeds to step S26, where the mechanical backup 27 is activated and the input side steering shaft 3 and the output side steering shaft 10 are mechanically operated. Next, the process proceeds to step S27, where the driving of the steering reaction force actuator 6 and the steering actuator 8 is stopped, and then the abnormality determination process ends.

  If the determination result in step S25 is | θs−θi | ≦ θth, it is determined that the steering angle sensor 4 and the steering reaction force control system are normal, and the process proceeds to step S28. In this step S27, it is determined whether or not the absolute value | δr1-δr2 | of the deviation from the turning angle δr1 to the turning angle δr2 exceeds a preset turning angle threshold value δth, and | δr1-δr2 |> If δth, it is determined that an abnormality has occurred in the steering control system, and the process proceeds to step S26 described above. If | δr1-δr2 | ≦ δth, it is determined that the steering control system is normal. Control goes to step S29.

  In this step S29, it is determined whether or not the absolute value | δs−δr1 | of the deviation obtained by subtracting the turning angle δr1 from the steering angle θs exceeds a preset threshold value θth1, and | δs−δr1 |> θth1. Sometimes it is determined that the steering angle sensor or the steering actuator system is abnormal, and the process proceeds to step S26. When | δs−δr1 | ≦ θth1, it is determined that the steering angle sensor or the steering actuator system is normal. The process proceeds to step S30.

In this step S30, it is determined whether or not the absolute value | δs−δr2 | of the deviation obtained by subtracting the turning angle δr1 from the steering angle θs exceeds a preset threshold θth1, and | δs−δr2 |> θth1. Sometimes it is determined that the steering angle sensor or the steering actuator system is abnormal, and the process proceeds to step S30. When | δs−δr2 | ≦ θth2, it is determined that the steering angle sensor or the steering actuator system is normal. Control goes to step S31.
In step S31, the operation of the mechanical backup 27 is stopped, and then the timer interrupt process is terminated and the process returns to a predetermined main program.

(Abnormality judgment processing operation)
In the abnormality determination process shown in FIG. 21, the steering angle θs, the steering reaction force actuator rotation angle θr, the turning actuator rotation angle θa, and the pinion gear rotation angle θp are read (step S21), and the steering reaction force actuator rotation angle θr is read as the steering input angle. The angle is converted to θi (step S22), the turning actuator rotation angle θa is converted to the turning angle δr1 (step S23), and the pinion gear rotation angle θp is converted to the turning angle θr2 (step S24).

  When the steering angle θs is compared with the steering input angle θi and the absolute value | θs−θi | of the deviation is larger than the threshold θth, an abnormality has occurred in the steering angle sensor 4 or the steering reaction force control system. (Step S25), the mechanical backup 27 is activated (step S26), and the input side steering shaft 3 and the output side steering shaft 10 are mechanically connected. Further, the driving of the steering reaction force actuator 6 and the steering actuator 8 of the steer-by-wire system SBW is stopped (step S27).

Here, when it can be determined that the steering angle θs is normal, a steering assist control process for generating a steering assist force by the steering actuator 8 may be performed using the steering angle θs.
As described above, when the abnormality is detected by the abnormality determination process, the mechanical backup 27 is activated, and the driving of the steering actuator 8 by the steering control device 50 described above is stopped, the suspension device 1B of the steering control device 50 is stopped. It becomes impossible to complement straightness, and straightness cannot be secured.

  However, in the present invention, as described above, when turning the steered wheels 17FL and 17FR as turning outer wheels, the virtual lower pivot point PL moves forward in the vehicle front-rear direction in the lower link structure, and the virtual upper pivot point in the upper link structure. When the PU moves rearward in the vehicle front-rear direction, the inclination of the kingpin axis KS is increased, the caster angle is increased, and the caster trail is increased, so that straightness can be ensured. For this reason, it becomes an understeer tendency and can exhibit the fail safe function which improves the maneuverability and stability of a vehicle.

  On the other hand, when the steering angle sensor 4 and the steering reaction force control system are normal, the turning angle δr1 is compared with the turning angle δr2, and when the absolute value | δr1-δr2 | of the deviation is larger than the threshold value δth. Then, it is determined that there is an abnormality in the steering actuator control system, and the process proceeds to step S26, and similarly to the above, the mechanical backup operation is performed and the driving of the steering reaction force actuator 6 and the steering actuator 8 is stopped. End the timer interrupt process.

  Further, when it is determined that the steering actuator control system is normal, the steering angle θs is compared with the steering angle δr1, and when the absolute value | θs−δr1 | of the deviation is larger than the threshold value δth, the steering actuator When it is determined that there is an abnormality in the control system, the process proceeds to step S26, and in the same manner as described above, the mechanical backup operation is performed and the operations of the steering reaction force actuator 6 and the steering actuator 8 are stopped and then the timer allocation is performed. Finish the process.

  Further, the steering angle θs is compared with the turning angle δr2, and when the absolute value | θs−δr2 | of the deviation is larger than the threshold value δth, it is determined that there is an abnormality in the turning actuator control system, and the process proceeds to step S26. In the same manner as described above, the mechanical backup operation is performed, the driving of the steering reaction force actuator 6 and the steering actuator 8 is stopped, and the timer interruption process is ended.

Then, assuming that the steering angle sensor 4, the steering reaction force actuator control system, and the turning actuator control system are all normal, the timer backup process is terminated after the mechanical backup 27 is deactivated, and a predetermined main program is entered. Return.
As described above, when the various sensors and the control system of the steer-by-wire system SBW are normal, the mechanical backup 27 is set in a non-operating state so that the input side steering shaft 3 and the output side steering shaft 10 are separated.
In the abnormality determination process shown in FIG. 21, the processes in steps S25 to S30 correspond to the abnormality control unit.

  In the first embodiment, the virtual lower pivot point PL of the lower link structure of the steered wheel that is the outer wheel at the time of steering is moved to the front side in the vehicle front-rear direction, and the virtual upper pivot point PU of the upper link structure is moved rearward in the vehicle front-rear direction. Although the case of moving to the side has been described, the present invention is not limited to this, and any one of the lower link structure and the upper link structure may be a link structure in which the pivot point such as the A arm does not substantially move.

  In the first embodiment, the case where the kingpin axis KS is tilted backward when it is determined that the control / drive circuit unit 26 constituting the steering control device is abnormal has been described. Naturally, the kingpin axis KS can be tilted backward at the time of steering even when the steering control of the steered wheels by 26 is insufficient.

(Effect of 1st Embodiment)
(1) The vehicle suspension includes: a steering control device that controls the steering wheel by operating an actuator according to a steering state of the steering wheel; and a vehicle suspension device that rotatably supports the steering wheel. The apparatus includes an axle carrier having an axle that rotatably supports the steered wheels, a lower link structure that couples a vehicle body side support portion and the axle carrier below the axle, and a vehicle body side above the axle. An upper link structure for connecting the support portion and the axle carrier, and at least one of the king pins of the lower link structure and the upper link structure when the steering control of the steering control device is insufficient. A pivot point moving unit is provided for moving the pivot point of the shaft so that the kingpin shaft tilts backward in a vehicle side view.
As a result, when the steering control of the steering control device is insufficient, it is possible to move the pivot point of at least one of the lower link structure and the upper link structure to tilt the kingpin shaft backward, ensuring straightness It is possible to ensure maneuverability and stability.

(2) A steering control device that turns an steered wheel by operating an actuator according to a steering state of the steering wheel, and a vehicle suspension device that rotatably supports the steered wheel. The vehicle suspension device includes an axle carrier having an axle that rotatably supports the steered wheels, a lower link structure that connects a vehicle body side support portion and the axle carrier below the axle, and An upper link structure for connecting the vehicle body side support portion and the axle carrier on the upper side of the axle. Further, the kingpin shaft passing through the lower pivot point of the lower link structure and the upper pivot point of the upper link structure is set so as to pass through the tire ground contact surface at a neutral position of the steering wheel, and the steering control device includes: The steered wheels are steered so as to generate self-aligning restoring force on the steered wheels by actuating an actuator to ensure straightness of the suspension device, and at least steer control of the steered control device A pivot point moving part that moves the pivot point of at least one kingpin shaft of the lower link structure and the upper link structure so that the kingpin shaft tilts backward in a side view of the vehicle. Yes.

  As a result, when the straightness guarantee control by the steering control device is insufficient, the pivot point of the lower link structure and the upper link structure can be moved to tilt the kingpin shaft backward, thereby improving the straightness. It can be ensured to exhibit a fail-safe function, and the maneuverability and stability can be ensured.

In addition, since the kingpin shaft is set to pass through the tire ground contact surface at the neutral position of the steering wheel, the moment around the kingpin shaft can be made smaller, so steering can be performed with a smaller rack axial force. And can control the direction of the wheels with a smaller force.
Therefore, in this embodiment, it is possible to improve the maneuverability and stability of the vehicle while reducing the weight of the suspension device.

(3) A steered control device comprising a steer-by-wire system that operates and steers a steered wheel mechanically separated from the steering wheel according to the steering state of the steering wheel, and the steered wheel is rotatable. And a vehicle suspension device supported by the vehicle. The vehicle suspension device includes an axle carrier having an axle that rotatably supports the steered wheels, a lower link structure that connects a vehicle body side support portion and the axle carrier below the axle, and An upper link structure for connecting the vehicle body side support portion and the axle carrier on the upper side of the axle. Furthermore, at least when the steering control of the steering control device is insufficient, the pivot point of at least one kingpin shaft of the lower link structure and the upper link structure is tilted backward in the vehicle side view. A pivot point moving unit that moves as described above is provided.

According to this configuration, when the steering control by the steer-by-wire system becomes insufficient, the caster trail increases the pivot point of at least one kingpin shaft of the lower link structure or the upper link structure at least when the steered wheels are steered. Direction, that is, the lower pivot point of the lower link structure is moved forward in the vehicle longitudinal direction, the upper pivot point of the upper link structure is moved backward in the vehicle longitudinal direction, or the lower pivot point of the lower link structure is moved forward in the vehicle longitudinal direction. And the upper pivot point of the upper link structure can be moved rearward in the vehicle front-rear direction. For this reason, it is possible to increase the caster angle by tilting the kingpin shaft backward to increase the caster trail and to ensure the straight traveling performance, and to exhibit the fail-safe function.
At this time, by moving the moving direction of the virtual upper pivot point also inward in the vehicle width direction, the king pin inclination angle can be reduced, the positive scrub radius can be increased inward in the vehicle width direction, and the force that enters the steering rack member can also be increased. It can be reduced at the same time.

(5) A steered control device comprising a steer-by-wire system that steers the steered wheels mechanically separated from the steering wheel according to the steering state of the steering wheel by operating an actuator, and the steered wheels are rotatable. And a vehicle suspension device to be supported. The vehicle suspension device includes an axle carrier having an axle that rotatably supports the steered wheels, a lower link structure that connects a vehicle body side support portion and the axle carrier below the axle, and An upper link structure for connecting the vehicle body side support portion and the axle carrier on the upper side of the axle, and a kingpin shaft passing through the lower pivot point of the lower link structure and the upper pivot point of the upper link structure is connected to the steering wheel. It is set to pass through the tire ground contact surface at the neutral position. Further, the steered control device operates the actuator to steer the steered wheels so as to generate a restoring force for self-alignment on the steered wheels, thereby ensuring straightness of the suspension device, A pivot that moves at least one pivot point of the lower link structure and the upper link structure so that the kingpin shaft tilts backward in a side view of the vehicle when steering control of the steering control device is insufficient. A point moving part is provided.

  According to this configuration, the steering control device configured by the steer-by-wire system has a kingpin shaft passing through the lower pivot point of the lower link structure and the upper pivot point of the upper link structure of the suspension device at the neutral position of the steering wheel. It is possible to ensure straightness by compensating for the decrease in straightness due to setting so as to pass through the tire contact surface. For this reason, in this embodiment, it is possible to improve the maneuverability and stability of the vehicle while reducing the weight of the suspension device.

  And, when the turning control by the turning control device constituted by the steer-by-wire system is insufficient, the pivot point of at least one of the lower link structure and the upper link structure can be moved to tilt the kingpin shaft backward It is possible to ensure straightness and to exhibit a fail-safe function, and to ensure maneuverability and stability.

(6) The straightness ensuring part is provided in the steering control device, and the straightness ensuring part ensures the straightness of the vehicle suspension device.
Therefore, for example, by using the electric actuator in the steering control device configured by the steer-by-wire system, it is possible to perform the straightness ensuring control corresponding to the setting of the kingpin axis in the present invention. For this reason, in this embodiment, it is possible to improve the maneuverability and stability of the vehicle while reducing the weight of the suspension device.

(7) The lower link structure includes the first link member and the second link member that intersect each other, and is a virtual lower pivot point that is represented by an intersection of the first link member and the second link member. Is arranged so as to move at least forward in the vehicle front-rear direction when the steered wheels are steered.
As a result, the virtual lower pivot point moves forward in the vehicle front-rear direction when the steered wheels are steered, so that the caster angle can be increased and the caster trail can be increased, and straightness can be ensured. For this reason, maneuverability and stability can be ensured at the time of failure of the steer-by-wire system.
In this case, the virtual lower pivot point represented by the intersection of the first lower link member and the second lower link member in plan view can be set on the inner side in the vehicle width direction. For this reason, the scrub radius can be set large within the range of the positive scrub in relation to the virtual upper pivot point.

(8) The upper link structure includes a first upper link member and a second upper link member, and a virtual upper pivot point of the first upper link member and the second upper link member is set as a steered wheel. It is set as the link arrangement | positioning which moves at least to the vehicle front-back direction back at the time of steering.
As a result, the virtual upper pivot point moves rearward in the vehicle front-rear direction when the steered wheels are steered, so that the caster angle can be increased and the caster trail can be increased, thereby ensuring straightness. For this reason, it is possible to improve the maneuverability and stability at the time of failure of the steer-by-wire system.

(9) The steer-by-wire system includes a steering actuator that steers the steered wheel, and an actuator control device that controls the steered actuator so that the steered angle of the steered wheel corresponds to the steering angle of the steering wheel. And the steering actuator is controlled so as to compensate for the decrease in straightness due to the importance of the steering response of the suspension device.
As a result, the kingpin shaft of the suspension device can be steered with a smaller rack axial force by reducing the moment around the kingpin shaft by passing through the tire contact surface at the straight traveling position of the steered wheel. Steer-by-wire system that reduces the straightness by ensuring the steering response of the suspension device when the steering response that can control the direction of the wheel with a smaller force is emphasized compared to the straightness It can be complemented with.

(10) The straightness guaranteeing part estimates a self-aligning torque and guarantees the straightness of the suspension device.
By adopting this configuration, the straightness guaranteeing section can secure the straightness reduced by securing the high responsiveness of the suspension device with the self-aligning torque, and the steering and stability can be improved. .

(11) When the steering control device starts steering the steering wheel from the neutral position, the steering control device adjusts the straight travel performance ensuring control by the straight travel performance securing section to change the initial turning response to the suspension device itself. A steering response setting unit for setting the steering response is provided.
According to this configuration, when turning is started from the neutral position of the steering wheel, the initial response characteristic can be set to high turning response. Then, ideal steering response can be ensured by adjusting the steering response of the suspension device itself by the straight travel guarantee control by the straight travel guarantee section.

(12) The turning control device includes a turning angle control unit that estimates the compliance steer and corrects the displacement of the turning wheels.
According to this configuration, therefore, the rigidity of the bush inserted in the vehicle body side support portion of the lower arm that constitutes the suspension device can be reduced, and the riding comfort of the vehicle can be improved.

(13) When the steering control device starts steering at least from the neutral position, the steering control device sets a high steering response in the steering response of the suspension device in the initial steering state. When the vehicle is in the steered state after the initial steered state, a steered response setting unit is provided that sets the steered response required by the straight travel guarantee control by the straight travel guarantee unit.
According to this configuration, a high steering response characteristic of the suspension device is ensured for the initial turning, and control is performed to ensure the straightness of the suspension device itself of the steered actuator in the straightness ensuring portion after the initial set time has elapsed. And ideal steering response characteristics can be obtained.

(14) The steering response setting unit includes a delay control unit that delays the straightness ensuring control by the straightness ensuring unit when the steering wheel is steered from a neutral position.
According to this configuration, since the delay control unit delays the start of the straightness ensuring control by the straightness ensuring unit, the initial turning response characteristic can be set to the high turning response of the suspension device itself.

(15) The delay control unit includes a gain adjustment unit that adjusts the start of straightness ensuring control by the straightness ensuring unit.
According to this configuration, by this, the gain adjustment unit sets the gain for the straightness guarantee control value in the straightness guarantee control, for example, to “0”, so that the straightness guarantee control is not performed and the gain is set to “0”. By setting a large value, for example, “1”, it is possible to start straightness guarantee control. For this reason, by providing the gain adjustment unit, it is possible to easily adjust the start of the straight travel performance ensuring control.

(16) The delay control unit delays the rectilinearity ensuring control by the rectilinearity ensuring unit by 0.1 second from the steering start timing at which the steering wheel is steered right or left from the neutral position. Let it begin.
According to this configuration, the initial turning response characteristic can effectively use the high turning response characteristic of the suspension device itself, and the straightness guaranteeing control by the straightness guaranteeing part is performed after the initial period of 0.1 seconds has elapsed. The ideal steering response characteristic can be obtained by starting.

(17) The turning control device includes a target turning angle calculation unit that calculates a target turning angle according to a steering angle, and the straight travel performance guarantee unit at a target turning angle calculated by the target turning angle calculation unit. An adder for adding the straightness ensuring control value, a turning motor control unit for forming a motor command current for matching the addition output of the adder and the rotation angle of the turning motor constituting the turning actuator, and A current control unit that forms a motor drive current supplied to the steered motor that matches the motor command current.

  According to this configuration, the target turning angle calculation unit calculates the target turning angle corresponding to the steering angle of the steering wheel, and adds the straight travel guarantee ensuring control value to the target turning angle by the adder, and the turning motor The control unit forms a target motor current that matches the rotation angle of the steering motor that constitutes the actuator with the addition output of the adder, and the motor current control unit forms a motor drive current that matches the target motor command current. Can be driven and controlled in accordance with the steering angle of the steering wheel. Here, since the target turning angle output from the target turning angle calculation part is adjusted by the turning response control part, optimal turning control can be performed.

(18) The steering control device includes at least a steering angle detection unit that detects a steering angle of the steering wheel, a steering operation detection unit that detects a steering operation of the steered wheels by the steering actuator, and the steering angle. An abnormality control unit that monitors the steering angle detected by the detection unit and the steering operation detected by the steering operation detection unit, performs abnormality determination, and stops control of the steering actuator when it is determined as abnormal. ing.
According to this configuration, the steering control device performs self-diagnosis to determine abnormality, and when it is determined to be abnormal, control of the steering actuator can be stopped, and malfunction can be reliably prevented.

(Application 1)
In the first embodiment, the kingpin axis is set in the tire ground contact surface. As an example, the case where the caster trail is set to a value close to zero and the road landing point of the kingpin shaft coincides with the tire ground contact surface center point will be described. did.
On the other hand, in this application example, the setting condition of the kingpin axis is limited to the range from the center point of the tire contact surface to the front end of the tire contact surface.
(effect)
If the road landing point of the kingpin shaft is set from the center of the tire contact surface to the front end of the tire contact surface, it is possible to ensure both straightness and a reduction in the weight of the steering operation. That is, maneuverability and stability can be improved.

(Application example 2)
In the first embodiment, in the coordinate plane shown in FIG. 11, a region surrounded by a one-dot chain line is taken as an example of a region suitable for setting. On the other hand, an isoline of the rack axial force of interest is used as a boundary line, and an area inside the range indicated by the boundary line (in the decreasing direction of the kingpin tilt angle and the increasing direction of the scrub radius) is set as an area suitable for setting. it can.
(effect)
Assuming the maximum value of the rack axial force, the suspension geometry can be set within the range below the maximum value.

(Application 3)
In the first and second embodiments and each application example, the case where the suspension device 1B is applied to a vehicle including a steer-by-wire type steering device has been described as an example. However, instead of the steer-by-wire method, a mechanical steering mechanism The suspension device 1B can be applied to a vehicle including the steering device.
In this case, the kingpin axis is determined according to the conditions based on the above examination results, the caster trail is set in the tire contact surface, and the link arrangement of the mechanical steering mechanism is configured accordingly.
(effect)
Even in a steering mechanism having a mechanical structure, it is possible to reduce the moment around the kingpin and reduce the steering force required by the driver, and to improve maneuverability and stability.

(Second Embodiment)
Next, a second embodiment of the present invention will be described with reference to FIGS.
In the second embodiment, the upper link structure of the first embodiment described above is replaced with a compression-type structure in which link members do not intersect with each other, and a compression-type structure in which link members intersect with each other is employed. .
That is, in the second embodiment, of the steered wheels 17FL and 17FR, the steered wheel 17FL will be described. As shown in FIGS. 22 to 27, the lower link structure connected to the lower end side from the axle of the axle carrier has been described above. Similar to the first embodiment, a compression link (first lower link member) 37 and a tension link (second lower link member) 38 are crossed.

  On the other hand, the upper link structure connected to the upper end of the axle carrier has a transverse link as a first upper link that extends in the axial direction slightly behind the tire center axis in the vehicle longitudinal direction in plan view. (Transverse link member) 71 and a compression link (compression link member) 72 serving as a second upper link disposed on the rear side in the vehicle front-rear direction of the transverse link 71 are crossed. Each of the transverse link 71 and the compression link 72 is composed of an I arm, and is individually supported on the upper end side of the axle carrier.

  Here, as schematically shown in FIG. 26 and FIG. 27 (a), the transverse link 71 has a wheel center support point TBUa at the upper end of the axle carrier at the center of the tire outside the inner end in the vehicle width direction of the steered wheels 17FL. The vehicle body side support point TBUb is disposed slightly on the rear side in the vehicle front-rear direction of the tire center axis on the inner side in the vehicle width direction.

  Further, as schematically shown in FIGS. 26 and 27 (a), the compression link 72 has a wheel side support point CPUa at the upper end of the axle carrier on the wheel side of the transverse link 71 on the inner side in the vehicle width direction of the steered wheel 17FL. The vehicle body side support point CPUb is disposed on the vehicle longitudinal direction rear side of the transverse link 71 with respect to the vehicle body side support point TBUb of the transverse link 71 from the support point TBUa. Therefore, the compression link 72 extends obliquely rearward from the wheel side support point CPUa.

  In the upper link structure, the virtual upper pivot point PU is an intersection of the transverse link 71 and the compression link 72 unlike the upper link structure of the first embodiment described above in plan view. As shown in FIG. 26, the virtual upper pivot point PU is set slightly rearward in the vehicle longitudinal direction from the axle center line on the inner side in the vehicle width direction of the steered wheels 17FL.

On the other hand, in the lower link structure, as shown in FIG. 27 (b), the virtual lower pivot point PL is near the inner end in the vehicle width direction of the steered wheels 17FL and from the axle centerline as in the first embodiment described above. It is set slightly forward in the front-rear direction.
Therefore, as shown in FIG. 25, the kingpin axis KS connecting the virtual upper pivot point PU and the virtual lower pivot point PL has its road landing point on the front side of the vehicle with respect to the tire ground contact surface center point within the tire contact surface. For this reason, the kingpin tilt angle can be set to 15 degrees or less, and the caster angle is close to zero degrees, so that the moment around the kingpin axis can be further reduced as in the first embodiment. For this reason, the steering can be performed with a smaller rack axial force, and the direction of the wheel can be controlled with a smaller force. Thereby, maneuverability and stability can be improved.

In the second embodiment, the bush rigidity of the vehicle body side support point CPUb of the compression link 72 of the upper link structure is the bush rigidity of the wheel side support point CPUa, the wheel side support point TBUa of the transverse link 71 and the vehicle body side support point TBUb. It is set smaller than the bush rigidity.
Therefore, at the time of turning as a turning outer wheel during non-braking, there is no movement of the vehicle body side support point CPUb of the compression link 72, and the compression link 72 rotates clockwise around the normal support point Pu shown by the solid line to show a one-dot chain line. The position is shown. Similarly, the transverse link 71 also rotates clockwise around the vehicle body side support point TBUb to the position shown in the dashed line. For this reason, the virtual upper pivot point PU, which is the intersection of both the links 71 and 72, moves obliquely forward, which is the front side in the vehicle front-rear direction on the inner side in the vehicle width direction as shown by the one-dot chain line.

At this time, in the lower link structure, as described in the first embodiment, the virtual lower pivot point PL slightly tilts outward in the vehicle width direction and moves forward in the vehicle front-rear direction.
For this reason, the kingpin axis KS connecting the virtual lower pivot point PL and the virtual upper pivot point PU has a larger kingpin inclination angle, but an increase in inclination in a side view of the vehicle is suppressed. For this reason, the scrub radius is reduced in the positive scrub region, and the straight running performance is reduced by this amount.

  On the other hand, since both the virtual lower pivot point PL and the virtual upper pivot point PU move forward in the vehicle front-rear direction when viewed from the side of the vehicle, the change in caster angle is suppressed and the increase in caster trail is also suppressed. Therefore, the rack axial force can be maintained in a small state, and as described above in the first embodiment, an abnormality occurs in the steer-by-wire system SBW and the driving of the steering actuator 8 is stopped, and the input side steering shaft Even when 3 and the output side steering shaft 10 are directly connected by the mechanical backup 27, a relatively light steering can be performed.

  However, when the vehicle is in a braking state in a steered state that is a turning outer wheel, a force in the vehicle front-rear direction (rearward force) acts on the wheel in addition to lateral force on the steered wheel 17FL. In this state, since the bush rigidity is large in the lower link structure, there is almost no rearward movement amount of the body support points CPLb and TSLb of the compression link 37 and the tension link 38. Will affect the structure. When the vehicle longitudinal force acts on the upper link structure, the compression link 72 mainly takes charge of the vehicle longitudinal force.

  Since the bushing rigidity of the compression link 72 is set lower than the bushing rigidity of the other support points of the vehicle body side support point CPUb of the compression link 72, the vehicle body side support point CPUb of the compression link 72 is shown enlarged in FIG. In this manner, the vehicle moves from the state shown in the solid line to the rear in the vehicle longitudinal direction as shown in the dotted line. In other words, the compression link 72 moves from the state shown by the one-dot chain line to the state shown by the dotted line, and accordingly, the transverse link 71 also rotates counterclockwise around the vehicle body side support point TBUb and travels straight. It becomes a position that almost overlaps the solid line of time.

  Accordingly, the virtual upper pivot point PU moves diagonally leftward and rearward from PU1 shown in the alternate long and short dash line to the rear side in the vehicle longitudinal direction on the outer side in the vehicle width direction. For this reason, the caster angle is increased in a side view of the vehicle and the caster trail is increased, so that the straight traveling performance is ensured and an understeering tendency occurs. As a result, as described above in the first embodiment, when the failure occurs in the steer-by-wire system SBW and the control of the steered actuator 8 by the steer-by-wire system SBW is stopped, the steered wheel 17FL is set as the turning outer wheel. When the vehicle is in a braking state at the time of turning, the straightness can be ensured on the suspension device 1B side, so that the maneuverability and stability of the vehicle can be improved and a good fail-safe function can be exhibited. Further, since the vehicle body side support point CPUb of the compression link 72 having the upper link structure moves rearward in the vehicle front-rear direction even during braking in a straight traveling state of the vehicle, the caster angle can be increased and the caster trail can be increased. For this reason, it is possible to further improve the straightness, and to ensure the maneuverability and stability of the vehicle.

  In the present embodiment, the link arrangement of the compression link 37 and the tension link 38 constituting the lower link structure, the link arrangement of the transverse link 71 and the compression link 72 constituting the upper link structure, and the compression link having reduced bush rigidity. 72 bushes correspond to the pivot point moving part.

(Effect of 2nd Embodiment)
(1) In the upper link structure, the first upper link member is a transverse link member, the second upper link member is a compression link member, and the second upper link member is on the vehicle body side. The rigidity of the connection point is set smaller than the rigidity of the connection point on the vehicle body side of the first upper link member.
Therefore, it is possible to ensure straightness during turning braking and to ensure straightness even in the case where an abnormality occurs in the steer-by-wire system SBW and the straightness of the suspension device 1B cannot be complemented. Property and stability can be ensured, and a good fail-safe function can be exhibited.

(Third embodiment)
Next, a third embodiment of the present invention will be described with reference to FIGS.
In the third embodiment, when the abnormality occurs in the steer-by-wire system, the vehicle body side support point of the compression link 72 in the second embodiment described above is forcibly moved to the vehicle rear side to ensure straight traveling performance. It is.
That is, in the third embodiment, as shown in FIG. 29, the vehicle body side support point TBUb of the upper link structure transverse link 71 is slightly more than the vehicle body side support point TBUb of the transverse link 71 of the second embodiment described above. It is set on the vehicle rear side.

Further, as shown in FIG. 30, the pivot point constituting the pivot point moving unit for forcibly moving the vehicle body side support point CPUb of the compression link 72 of the upper link structure to the rear and moving the virtual upper pivot point PU to the rear. A forced movement unit 79 is provided.
The pivot point forced moving portion 79 includes a support position variable bush 80 disposed at the vehicle body side support point CPUb of the compression link 72 having an upper link structure. This support position variable bush 80 forms a pair of straights 80a and 80b at a longitudinally symmetrical position sandwiching the vehicle body side support shaft 81 in the vehicle longitudinal direction.

  Further, the curls 80 a and 80 b of the support position variable bush 80 are connected to a hydraulic pressure source 83 via a four-port electromagnetic switching valve 82. Here, one input port of the four-port electromagnetic switching valve 82 is connected to the hydraulic pressure source 83, the other input port is connected to the oil tank 84, and one output port is connected to the straight 80 a of the support position variable bush 80. The other output port is connected to the curb 80b of the support position variable bush 80. The 4-port electromagnetic switching valve 82 supplies the hydraulic pressure from the hydraulic pressure source 83 to both 80a and 80b at the first switching position when the control signal CS is in the OFF state, and the control signal CS is in the ON state. At a certain second switching position, the hydraulic pressure source 83 is connected to the curb 80 b, and the curb 80 a is connected to the oil tank 84.

  Therefore, in the state where the hydraulic pressure is supplied to the slips 80a and 80b, the outer circumference circle of the support position variable bush 80 maintains a concentric circle with respect to the center axis of the vehicle body side support shaft 85 as shown by the solid line in FIG. However, when the hydraulic pressure of the straightening 80a is returned to the oil tank 84, the central axis of the outer circumference circle of the support position variable bush 80 is on the vehicle body side as in the case of braking in the second embodiment described above as shown by the dotted line in FIG. It moves to the vehicle rear side from the central axis of the support shaft 85.

Then, the 4-port electromagnetic switching valve 82 is controlled to be switched by the abnormality determination process shown in FIG.
As shown in FIG. 31, in this abnormality determination process, step S33 for outputting the control signal CS in the ON state to the 4-port electromagnetic switching valve 82 is added after step S27 in the abnormality determination process of FIG. In addition, after step S32, the control signal CS output to the 4-port electromagnetic switching valve 82 is in an OFF state and step S34 is added except that step S34 is added. The corresponding steps are denoted by the same step numbers, and detailed description thereof will be omitted.

(Operation of Third Embodiment)
Now, in the abnormality determination process of FIG. 31 executed by the controller / drive circuit unit 26, the steer-by-wire in which the steering angle θs, the steering input angle θi, the first turning angle δr1, and the second turning angle δr2 are normal. In a state in which the system SBW is normal, the process proceeds to steps S25 through steps S21 to S24, the process proceeds to steps S31 through steps S28 to S30, and the mechanical backup 27 is set in the non-operating state so that the input side steering shaft 3 and the output are output. The side steering shaft 10 is separated.

Next, the process proceeds to step S32, where the steering reaction force actuator 6 and the steering actuator 8 are driven to generate a steering reaction force corresponding to the steering angle θs of the steering wheel 2 and the turning angle δr1 corresponding to the steering angle θs. And δr2.
Next, the process proceeds to step S34, and the control signal CS is turned off. For this reason, the 4-port electromagnetic switching valve 82 maintains the first switching position, and the hydraulic pressure of the hydraulic pressure source 83 is supplied to both the straights 80a and 80b of the support position variable bush 80, so that the compression link 72 is supported on the vehicle body side. The point CPUb maintains the position indicated by the solid line in FIG. 29 and FIG.
Therefore, as in the second embodiment described above, when the vehicle is traveling straight, the virtual upper pivot point PU of the upper link structure is the vehicle width of the steered wheels 17FL as shown in FIGS. 29 and 32 (a). It is set to the rear side in the vehicle front-rear direction from the axle center line inside the direction.

As for the lower link structure, similarly to the first and second embodiments described above, the virtual lower pivot point PL is near the inner end in the vehicle width direction of the steered wheels 17FL and on the rear side in the vehicle longitudinal direction from the axle centerline. Is set.
Accordingly, as shown in FIG. 29, the kingpin axis KS connecting the virtual upper pivot point PU and the virtual lower pivot point PL has its road landing point PT on the front side of the vehicle from the tire ground contact surface center point within the tire ground contact surface. For this reason, the kingpin tilt angle can be set to 15 degrees or less, and the caster angle is close to zero degrees, so that the moment around the kingpin axis can be further reduced as in the first embodiment. For this reason, the steering can be performed with a smaller rack axial force, and the direction of the wheel can be controlled with a smaller force. Thereby, maneuverability and stability can be improved.

Further, at the time of turning as a turning outer wheel, the vehicle body side support point CPUb of the compression link 72 does not move, and the compression link 72 rotates clockwise around the normal support point Pu shown by the solid line to the position shown by the one-dot chain line. Become. Similarly, the transverse link 71 also rotates clockwise around the vehicle body side support point TBUb to the position shown in the dashed line. For this reason, the virtual upper pivot point PU, which is the intersection of both the links 71 and 72, moves obliquely forward, which is the front side in the vehicle front-rear direction on the inner side in the vehicle width direction as shown by the one-dot chain line.
At this time, in the lower link structure, as described in the first embodiment, the virtual lower pivot point PL slightly tilts outward in the vehicle width direction and moves forward in the vehicle front-rear direction.

  For this reason, the kingpin axis KS connecting the virtual lower pivot point PL and the virtual upper pivot point PU has a larger kingpin inclination angle, but an increase in inclination in a side view of the vehicle is suppressed. For this reason, the scrub radius decreases in the positive scrub region, and the straight running performance is reduced by this amount. However, the straight running performance can be ensured by increasing the kingpin tilt angle of the kingpin axis KS in the vehicle front view.

  If an abnormality occurs in any of the steering angle sensor 4, the steering reaction force control system, and the turning control system from the normal state of the steer-by-wire system SBW, the process proceeds to step S26 in the abnormality detection process of FIG. Is operated, and the input side steering shaft 3 and the output side steering shaft 10 are connected. At substantially the same time, the driving of the steering reaction force actuator 6 and the turning actuator 8 is stopped (step S27).

Further, the control signal CS is turned on (step S33), the four-port electromagnetic switching valve 82 is switched to the second switching position, the straight 80a of the support position variable bush 80 is connected to the oil tank 84, and to the straight 80b. The supply of hydraulic pressure from the hydraulic source 83 is continued. For this reason, when the hydraulic pressure of the curb 80a is reduced, the support axis variable bush 80 moves to the rear side in the vehicle front-rear direction with respect to the center axis of the vehicle body side support shaft 85.
For this reason, the vehicle-body-side support point CPUa of the compression link 72 is shown by the dotted line in FIG. 32 (a) and an enlarged view of FIG. 33 in the same manner as during braking when turning as a turning outer wheel in the second embodiment described above. To the vehicle body side support point CPUb 'on the vehicle rear side.

  As the compression link 72 moves rearward, the transverse link 71 also rotates counterclockwise in plan view around the vehicle body side support point TBUb. At this time, since the vehicle body side support point CPUb of the transverse link 71 is arranged behind the wheel side support point CPUa in the vehicle front-rear direction, the wheel side support point CPUa is as shown in FIGS. 32 (a) and 33. Furthermore, it moves slightly outward in the vehicle width direction. For this reason, as shown in FIG. 32A and FIG. 33, the steered wheels 17FL are slightly inclined in the toe-in direction, so that straightness can be ensured.

At the same time, the virtual upper pivot point PU moves to the PU1 shown by the dotted line in the backward and forward direction of the vehicle and diagonally leftward and outward in the vehicle width direction.
For this reason, as in the second embodiment described above, the inclination of the kingpin shaft KS in a side view is increased, the caster angle is increased, the caster trail is increased, and the straight travel performance is increased. It is possible to compensate for the decrease in straightness due to the stop of the straightness complementation processing by the rudder control.

  Then, when the steered wheel 17FL is in a steered state where the turning wheel 17FL becomes a turning outer wheel in a state where the abnormality of the steered control system continues, the compression link 72 is shown in FIG. 32 (a) and FIG. It rotates clockwise around the vehicle body side support point CPUb '. Accordingly, the transverse link 71 also rotates in the clockwise direction in plan view around the vehicle body side support point TBUb, and the virtual upper pivot point PU moves from PU1 to PU2 marked with ● in front of the vehicle in the vehicle width direction. Move forward diagonally to the right.

On the other hand, when the abnormality of the steering control system occurs, the transverse link 71 rotates about the vehicle body side support point TBUb in the clockwise direction in plan view as indicated by the broken line at 29. Further, the compression link 72 rotates as shown by the broken line around the vehicle body side support point CPUb shown by the solid line.
For this reason, the virtual upper pivot point PU represented by the intersection of the transverse link 71 and the compression link 72 moves from PU to PU3 inward in the vehicle width direction and obliquely forward to the right in the vehicle front-rear direction PU3.
Therefore, in the steered state where the steered wheel 17FL becomes a turning outer wheel when the steered control system becomes abnormal and the steered control cannot be performed, the virtual upper pivot point PU2 is malfunctioning in the steered control system. Compared to the virtual upper pivot point PU3 when there is not, the vehicle is moving left diagonally rearward to the vehicle longitudinal direction rear side outside in the vehicle width direction.

  Therefore, as in the second embodiment described above, the virtual upper pivot point PU is moved outward in the vehicle width direction and rearward in the vehicle front-rear direction in a state where the straightness is deteriorated in which the abnormality of the steering control system has occurred. To increase the inclination of the kingpin shaft KS in a side view of the vehicle, to increase the caster angle, and to increase the caster trail to compensate for the decrease in straightness due to the stop of control of the steer-by-wire system SBW to ensure straightness Can do.

  In the third embodiment, the case where the pivot point forcibly moving portion 79 is constituted by the support position variable bush 80 having the straights 80a and 80b, the 4-port electromagnetic switching valve 82, and the hydraulic source 83 has been described. However, the present invention is not limited to this, and the vehicle body side support point CPUb of the compression link 72 may be moved rearward by a linear movement mechanism such as an electric linear actuator or a hydraulic cylinder.

Further, the movement direction of the vehicle body side support point CPUb of the compression link 72 is not limited to the rear in the vehicle front-rear direction, and may be moved rearward in the extension direction of the compression link 72.
Further, instead of moving the vehicle body side support point CPUb of the compression link 72 of the upper link structure by the pivot point forced moving portion 79, the vehicle body side support point CPLb of the tension link 38 of the lower link structure is replaced by the pivot point forced moving portion. You may make it move to the vehicle front-back direction front side.

(Effect of the third embodiment)
(1) A pivot that forcibly moves at least one pivot point of the lower link structure and the upper link structure in a direction in which the straightness is increased in a state in which the steered wheel cannot be steered by the steer-by-wire system. A point compulsory moving part is provided.
According to this configuration, when the steer-by-wire system is abnormal and the steered wheels cannot be steered, the pivot point is increased, that is, the caster angle is increased by increasing the caster angle, that is, the caster angle. It is possible to compensate for the decrease in straightness due to the steer-by-wire system.

(2) The steer-by-wire system includes a steering angle detector that detects a steering angle of the steering wheel, a turning angle detector that detects a turning angle of the steered wheels by the turning actuator, the steering angle, and the steering angle When the deviation of the turning angle is greater than or equal to a predetermined value, an abnormality control unit is provided that determines an abnormality and stops the control of the steering actuator, and the abnormality control unit forcibly moves the pivot point when the abnormality is determined to be abnormal Actuate the part.

According to this configuration, the steer-by-wire system performs self-diagnosis to determine abnormality, and when it is determined to be abnormal, the control of the steering actuator is stopped, and at this time the pivot point forced movement unit is activated, The decrease can be compensated for surely. In this case, the straightness can be increased regardless of whether or not the steered wheels are steered.
In the third embodiment, the case where the pivot point forcibly moving unit 79 is operated when an abnormality is detected in the abnormality detection process has been described. However, the present invention is not limited to this. The pivot point forced moving unit 79 may be operated when the steering control becomes insufficient, particularly when the straight travel performance is insufficient.

(Fourth embodiment)
Next, a fourth embodiment of the present invention will be described with reference to FIGS.
In the fourth embodiment, for example, an electric power steering device is applied instead of the steer-by-wire system as a steering control device in the first embodiment.
That is, in the fourth embodiment, as shown in FIG. 34, the automobile 1 includes a vehicle body 1A. The vehicle body 1A is provided with a suspension device 1B that supports the wheels WFR, WFL, WRR, and WRL and a steering device SS that steers the steered wheels WFR and WFL on the front wheel side.

The steering device SS includes a steering mechanism SM and an electric power steering device EP that applies a steering assist force to the steering mechanism SM.
The steering mechanism SM includes an input side steering shaft SSi, an output side steering shaft SSo, a steering wheel SW, a pinion gear PG, a rack shaft LS, and a tie rod TR.
A steering wheel SW is mounted on the input-side steering shaft SSi at the tip on the vehicle rear side. The input side steering shaft SSi and the output side steering shaft SSo are rotatably supported by the vehicle body 1A and are connected to each other via a torsion bar (not shown).

  A pinion gear PG is connected to the vehicle front end side of the output side steering shaft SSo, and this pinion gear PG meshes with a rack gear formed on the rack shaft LS to form a pinion and rack mechanism. In this pinion and rack mechanism, the rotational motion of the steering wheel SW is converted into a linear motion in the vehicle width direction. Tie rods TR are coupled between both ends of the rack shaft LS and the steered wheels WFR and WRL. These tie rods TR connect both ends of the rack shaft LS and the knuckle arms of the wheels WFR and WFL via ball joints.

  On the other hand, the electric power steering device EP is based on the difference in rotation angle between the steering angle sensor AS that detects the steering angle of the steering wheel SW mounted on the input side steering shaft SSi, and the input side steering shaft SSi and the output side steering shaft SSo. A steering torque sensor TS that detects a steering torque based on the electric actuator EA that transmits a steering control force to the output-side steering shaft SSo, and a rotation angle sensor RS that detects a rotation angle of the electric actuator EA. Yes. Here, the electric actuator EA is composed of an electric motor, and a gear that rotates integrally with the motor shaft meshes with a gear that is formed in a part of the output side steering shaft SSo, and rotates the output side steering shaft SSo. .

The electric power steering device EP includes a steering control device CT for driving and controlling the electric actuator EA, wheel speed sensors WSFR to WSRL for detecting the wheel speeds of the wheels WFR to WRL, and a vehicle state parameter acquisition unit CP. I have.
The vehicle state parameter acquisition unit CP acquires the vehicle speed based on a pulse signal indicating the rotational speed of the wheels output from the wheel speed sensors WFR to WRL. Moreover, vehicle state parameter acquisition part CP acquires the slip ratio of each wheel based on a vehicle speed and the rotational speed of each wheel. And vehicle state parameter acquisition part CP outputs each acquired parameter to steering control apparatus CT.

In addition to the parameters input from the vehicle state parameter acquisition unit CP, the steering control device CT includes a steering angle θs detected by the steering angle sensor 4, a steering torque Ts detected by the steering torque sensor TS, and a rotation angle sensor RS. The detected actuator rotation angle θa is input.
The steering control device CT includes a power steering control unit PC and a straight travel guarantee unit SG. The power steering control unit PC calculates a target auxiliary steering torque based on the steering torque Ts and the vehicle speed V, calculates a driving current for driving the electric actuator EA based on the calculated target steering auxiliary torque, and calculates the driving current. The electric actuator EA is supplied and controlled to drive the electric actuator EA.

As will be described later, the rectilinearity ensuring unit SG performs rectilinear complementation control that complements the rectilinearity of the suspension device 1B.
The wheels WFR, WFL, WRR, WRL are configured by attaching tires to the wheel hub mechanism WH, and are installed in the vehicle body 1A via the suspension device 1B of the first to third embodiments described above. Among these, the front wheels (wheels WFR, WFL) serving as steered wheels change the direction of the wheels WFR, WFL relative to the vehicle body 1 </ b> A when the knuckle arm is swung by the tie rod 15.

Next, a specific configuration of the steering control device CT will be described with reference to FIGS.
As shown in FIG. 35, the power steering control unit PC includes a target auxiliary torque current command value calculation unit TO and an actuator current control unit AC. The target auxiliary torque current command value calculation unit TO calculates a target auxiliary torque current command value It ^* corresponding to the steering torque Ts with reference to the control map based on the steering torque Ts detected by the steering torque sensor TS and the vehicle speed V. Then, the calculated target auxiliary torque current command value It ^* is output to the adder AD. In the adder AD, a target actuator current Ia ^* is calculated by adding a straight travel guarantee current command value Isa ^* to be described later to the target auxiliary torque current command value It ^* , and the calculated target actuator current Ia ^* is subtracted by the subtractor SB. Output to. An actuator current Iad supplied to the electric actuator EA detected by the actuator current sensor CS is fed back to the subtractor SB. Accordingly, the subtractor SB subtracts the actuator current Iad from the target actuator current command value Ia ^* to calculate the current deviation ΔI.
The actuator current control unit AC calculates the actuator current Iad by performing, for example, PID control on the current deviation ΔI input from the subtractor SB, and outputs the calculated actuator current Iad to the electric actuator EA.

On the other hand, in the straight travel guarantee unit SG, as with the straight travel complementing unit 53 in the first embodiment described above, the left and right wheel drive forces TR output from the drive force control device DC that controls the left and right drive wheel drive forces. And TL are input, and the steering torque Ts detected by the steering torque sensor ST is input, and based on these, the self-aligning torque Tsa is calculated. That is, in the straight travel guarantee part SG, first, the driving force difference ΔT (= TL−TR) between the driving forces TR and TL of the left and right wheels is calculated, and the generated torque shown in FIG. 12 is based on the calculated driving force difference ΔT. With reference to the estimated control map, the generated torque Th generated at the time of turning by the torque steer phenomenon is estimated.
Further, the straight travel guarantee unit SG multiplies the calculated self-aligning torque Tsa by a predetermined current gain Ki to calculate a straight travel guarantee current command value Isa ^* (= Ki · Tsa).

Here, the calculation of the self-aligning torque Tsa in the straight travel guarantee part SG first calculates the driving force difference ΔT (= TL−TR) between the driving forces TR and TL of the left and right wheels, and calculates the calculated driving force difference ΔT. Referring to the generated torque estimation control map shown in FIG. 17 described above, the generated torque Th generated at the time of turning is estimated by the torque steer phenomenon, and the generated torque Th is calculated from the steering torque Ts detected by the steering torque sensor 5. Subtract to calculate the self-aligning torque Tsa.
The self-aligning torque Tsa can be calculated by any method as in the first embodiment described above.

Then, the straightness ensuring current command value Isa ^* calculated by the straightness ensuring unit SG is supplied to the adder AD described above. In this adder AD, the target actuator current command value Ia ^* is calculated by adding the straight travel guarantee current command value Isa ^* to the target assist torque current command value I ^* calculated by the target assist torque current command value calculation unit TO. Then, the calculated target actuator current command value Ia ^* is supplied to the subtractor SB.
Accordingly, in the power steering control unit PC of the steering control device CT, the straight traveling performance ensuring unit SG uses the target auxiliary torque current command value It ^* calculated according to the steering torque Ts and the vehicle speed V input to the steering wheel SW. The calculated straight current ensuring current command value Isa ^* is added to calculate the target actuator current command value Ia ^* .
Since the electric actuator EA is controlled based on the target actuator current command value Ia ^* , in addition to the steering assist torque corresponding to the steering force transmitted to the steering wheel SW by the electric actuator EA, the straightness of the suspension device 1B Steering torque that secures the above is generated and transmitted to the output-side steering shaft SSo.

Further, in the steering control device CT, the abnormality of the steering torque sensor 5, the abnormality of the power steering control unit PC itself, and the abnormality of the straight travel guarantee unit SG are always diagnosed, and the steering torque sensor 5 and the power steering control unit PC itself are detected. When an abnormality is detected, for example, an actuator relay (not shown) disposed between the actuator current control unit AC and the electric actuator EA is turned off to stop the output of the actuator current Iad to the electric actuator EA.
When the output of the actuator current Iad to the electric actuator EA is stopped, the turning of the steered wheels WFR and WFL by the steering of the steering wheel 2 of the driver is continued, but the straightness by the straightness guaranteeing part SG is ensured. The generation of turning torque is stopped.

  However, since the suspension device 1B is configured in the same manner as in the first embodiment described above, when the driver enters a steered state in which the steering wheel 2 is turned to the right or left from the neutral position, In the suspension device 1B, as in the first embodiment, the virtual lower pivot point PL moves forward in the vehicle longitudinal direction with the lower link structure, and the virtual upper pivot point PU moves backward in the vehicle longitudinal direction with the upper link structure. For this reason, the kingpin axis KS tilts backward, the caster angle is reduced, the caster trail is increased, and straight travel performance can be ensured.

  Moreover, in the said embodiment, the caster angle of the suspension apparatus 1B is set to substantially zero. As shown in FIG. 36 (a), the relationship between the caster angle, the steering response, and the steering stability is high when the caster angle is zero. It cannot be ensured, that is, there is a trade-off relationship between steering response to caster angle and steering stability.

  On the other hand, the relationship between the road surface contact point position of the kingpin axis KS, the lateral force reduction allowance, and the straightness is as shown in FIG. That is, when the grounding point of the kingpin axis KS is at the center of the tire grounding surface, the lateral force reduction allowance is maximized as shown by the solid line. However, the straight travel performance cannot be ensured as shown by the broken line. When the grounding point of the kingpin axis KS is moved forward from the center of the tire grounding surface, the lateral force reduction margin is gradually reduced as the grounding point of the kingpin shaft KS moves away from the tire grounding surface center, and the straightness is gradually increased. To improve.

Thereafter, when the contact point of the kingpin shaft KS reaches the front end of the tire contact surface, the lateral force reduction allowance is reduced to about half of the maximum value, and on the contrary, the straight traveling performance is in a good state. Further, when the grounding point of the kingpin axis KS moves forward beyond the front end of the tire grounding surface, the lateral force reduction allowance is further reduced from about half of the maximum value, and on the contrary, the straightness is further improved.
In the above-described embodiment, in order to increase the lateral force reduction allowance, the kingpin shaft KS is set so as to pass through the tire ground contact surface with the steering wheel SW in the neutral position. For this reason, the linearity of the suspension device 1B is in a reduced state, and this reduction in linearity can be supplemented by controlling the electric actuator EA with the straightness guaranteeing part SG. For this reason, in the said 1st Embodiment, sufficient linearity can be ensured by supplementing the fall of the linearity in the suspension apparatus 1B by the linearity ensuring part SG.
Also in the fourth embodiment, as in the first embodiment described above, the lower link structure or the upper link structure can be configured such that the pivot point of the A arm or the like does not move.

  In the fourth embodiment, the suspension device 1B is not limited to the configuration of the first embodiment, and the upper link structure of the suspension device of the second embodiment or the third embodiment can be applied. In this case, when the upper link structure of the suspension device of the third embodiment is applied, the pivot point is detected when an abnormality that stops the driving of the electric actuator EA occurs by the abnormality diagnosis process of the steering control device CT. In addition to actuating the forcible moving part 79, the pivot point forcible moving part 79 is also actuated to tilt the kingpin axis KS backward when the straight running guarantee torque generated by the straight running guarantee part SG is insufficient. Is preferably ensured.

(Effect of 4th Embodiment)
(1) The vehicle suspension includes: a steering control device that controls the steering wheel by operating an actuator according to a steering state of the steering wheel; and a vehicle suspension device that rotatably supports the steering wheel. The apparatus includes an axle carrier having an axle that rotatably supports the steered wheels, a lower link structure that couples a vehicle body side support portion and the axle carrier below the axle, and a vehicle body side above the axle. An upper link structure for connecting the support portion and the axle carrier, and at least one of the king pins of the lower link structure and the upper link structure when the steering control of the steering control device is insufficient. A pivot point moving unit is provided for moving the pivot point of the shaft so that the kingpin shaft tilts backward in a vehicle side view.
As a result, when the steering control of the steering control device is insufficient, it is possible to move the pivot point of at least one of the lower link structure and the upper link structure to tilt the kingpin shaft backward, ensuring straightness It is possible to ensure maneuverability and stability.

(2) A steering control device that turns an steered wheel by operating an actuator according to a steering state of the steering wheel, and a vehicle suspension device that rotatably supports the steered wheel. The vehicle suspension device includes an axle carrier having an axle that rotatably supports the steered wheels, a lower link structure that connects a vehicle body side support portion and the axle carrier below the axle, and An upper link structure for connecting the vehicle body side support portion and the axle carrier on the upper side of the axle. Further, the kingpin shaft passing through the lower pivot point of the lower link structure and the upper pivot point of the upper link structure is set so as to pass through the tire ground contact surface at a neutral position of the steering wheel, and the steering control device includes: The steered wheels are steered so as to generate self-aligning restoring force on the steered wheels by actuating an actuator to ensure straightness of the suspension device, and at least steer control of the steered control device A pivot point moving part that moves the pivot point of at least one kingpin shaft of the lower link structure and the upper link structure so that the kingpin shaft tilts backward in a side view of the vehicle. Yes.

  As a result, when the straightness guarantee control by the steering control device is insufficient, the pivot point of the lower link structure and the upper link structure can be moved to tilt the kingpin shaft backward, thereby improving the straightness. It can be ensured to exhibit a fail-safe function, and the maneuverability and stability can be ensured.

In addition, since the kingpin shaft is set to pass through the tire ground contact surface at the neutral position of the steering wheel, the moment around the kingpin shaft can be made smaller, so steering can be performed with a smaller rack axial force. And can control the direction of the wheels with a smaller force.
Therefore, in this embodiment, it is possible to improve the maneuverability and stability of the vehicle while reducing the weight of the suspension device.

(3) The straightness ensuring part is provided in the steering control device, and the straightness ensuring part ensures the straightness of the vehicle suspension device.
Therefore, the straightness ensuring control corresponding to the setting of the kingpin axis in the present invention can be performed using the electric actuator in the electric power steering apparatus. For this reason, in this embodiment, it is possible to improve the maneuverability and stability of the vehicle while reducing the weight of the suspension device.

(4) The straightness guaranteeing part estimates the self-aligning torque to ensure straightness.
Therefore, the straight travel performance guaranteeing section can secure the straight travel performance reduced by securing the high responsiveness of the suspension device with the self-aligning torque, and the steering and stability can be improved.
Therefore, in this embodiment, it is possible to improve the maneuverability and stability of the vehicle while reducing the weight of the suspension device.

(Modification)
In addition, in the said 4th Embodiment, steering control apparatus CT demonstrated the case where it comprised by power steering control part PC and the straight travel guarantee part SG. However, the present invention is not limited to the above-described configuration, and as the steering control device CS, the power steering control unit PC may be omitted and only the straight travel guarantee unit SG may be provided. In this case, in the configuration of FIG. 11, the target auxiliary torque current command value calculation unit TO and the adder AD are omitted, and the straightness ensuring current command value Isa ^* output from the straightness ensuring unit SG is directly subtracted. What is necessary is just to make it input into the device SB.
Further, in the fourth embodiment, the straight travel performance ensuring unit SG calculates the straight travel performance ensuring current command value Isa ^* based on the self aligning torque Tsa. However, the present invention is not limited to this. A fixed value represented by an average value of the torque Tsa may be set.

DESCRIPTION OF SYMBOLS 1 ... Automobile, 1A ... Vehicle body, 1B ... Suspension device, 2 ... Steering wheel, 3 ... Input side steering shaft, 4 ... Steering angle sensor, 5 ... Steering torque sensor, 6 ... Steering reaction force actuator, 7 ... Steering reaction force actuator Angle sensor, 8 ... steering actuator, 9 ... steering actuator angle sensor, 10 ... output side steering shaft, 11 ... steering torque sensor, 12 ... pinion gear, 14 ... steering rack member, 15 ... tie rod, 17FR, 17FL, 17RR , 17RL ... wheels, 21 ... vehicle state parameter acquisition unit, 24FR, 24FL, 24RR, 24RL ... wheel speed sensor, 26 ... control / drive circuit unit, 27 ... mechanical backup, 32 ... axle, 33 ... axle carrier, 34 ... spring Member, 37 ... Compression Link (compression link member), 38 ... tension link (tension link member), 39 ... transverse link (transverse link member), 40 ... compression link (compression link member), 41 ... shock absorber, 50 ... steering control unit , 51 ... Target turning angle calculation unit, 52 ... Steering angle control unit, 53 ... Straightness complementation unit, 54 ... Disturbance compensation unit, 55 ... Adder, 56 ... Delay control unit, 56a ... Steering start detection unit, 56b ... monostable circuit, 56c ... gain adjustment unit, 56d ... multiplier, 56e ... adder, 58 ... steering angle deviation computing unit, 59 ... steering motor control unit, 60 ... current deviation computing unit, 61 ... motor current detection , 62 ... Motor current control unit, 63 ... Actuator control device, 64 ... Driving force control device, 70 ... Knuckle, 71 ... Transverse Click, 72 ... compression links 79 ... pivot point forced movement section, 80 ... support position variable bush, 82 ... 4-port solenoid switch valve, 83 ... hydraulic source, 84 ... oil tank

Claims (18)

A steering control device that controls the steered wheels by operating an actuator according to a steering state of the steering wheel;
A vehicle suspension device that rotatably supports the steered wheel,
The vehicle suspension apparatus includes an axle carrier having an axle that rotatably supports the steered wheel, a lower link structure that connects a vehicle body side support portion and the axle carrier below the axle, and the axle. An upper link structure for connecting the vehicle body side support portion and the axle carrier on the upper side;
When the steered wheel is a turning outer wheel, at least when the steered wheel is in a steered state from a straight traveling state and in a braking state , at least one kingpin shaft of the lower link structure and the upper link structure A pivot point moving part that moves the pivot point so that the kingpin shaft tilts backward in a vehicle side view ,
In the upper link structure, the first upper link member is constituted by a transverse link member, the second upper link member is constituted by a compression link member, and a vehicle body side connection point of the second upper link member is formed. A vehicle steering apparatus, wherein the rigidity is set to be smaller than the rigidity of a vehicle body side connection point of the first upper link member . A steering control device for operating the actuator according to the steering state of the steering wheel to steer the steered wheels;
A vehicle suspension device that rotatably supports the steered wheel,
The vehicle suspension apparatus includes an axle carrier having an axle that rotatably supports the steered wheel, a lower link structure that connects a vehicle body side support portion and the axle carrier below the axle, and the axle. An upper link structure for connecting the vehicle body side support portion and the axle carrier on the upper side, and a kingpin shaft passing through the lower pivot point of the lower link structure and the upper pivot point of the upper link structure is set to a neutral position of the steering wheel And set it to pass through the tire ground contact surface,
The steered control device operates the actuator to steer the steered wheels so that the steered angle of the steered wheels is restored, thereby ensuring straightness of the vehicle,
When the steered wheel is a turning outer wheel, at least when the steered wheel is in a steered state from a straight traveling state and in a braking state , at least one kingpin shaft of the lower link structure and the upper link structure A pivot point moving part that moves the pivot point so that the kingpin shaft tilts backward in a vehicle side view ,
In the upper link structure, the first upper link member is constituted by a transverse link member, the second upper link member is constituted by a compression link member, and a vehicle body side connection point of the second upper link member is formed. A vehicle steering apparatus, wherein the rigidity is set to be smaller than the rigidity of a vehicle body side connection point of the first upper link member . A steered wheel control system comprising a steer-by-wire system that steers and controls steered wheels mechanically separated from the steering wheel according to the steering state of the steering wheel;
A vehicle suspension device that rotatably supports the steered wheel,
The vehicle suspension apparatus includes an axle carrier having an axle that rotatably supports the steered wheel, a lower link structure that connects a vehicle body side support portion and the axle carrier below the axle, and the axle. An upper link structure for connecting the vehicle body side support portion and the axle carrier on the upper side;
When the steered wheel is a turning outer wheel, at least when the steered wheel is in a steered state from a straight traveling state and in a braking state , at least one kingpin shaft of the lower link structure and the upper link structure A pivot point moving part that moves the pivot point so that the kingpin shaft tilts backward in a vehicle side view ,
In the upper link structure, the first upper link member is constituted by a transverse link member, the second upper link member is constituted by a compression link member, and a vehicle body side connection point of the second upper link member is formed. The vehicle steering apparatus, wherein rigidity is set to be smaller than rigidity of a vehicle body side connection point of the first upper link member . A steering control device comprising a steer-by-wire system that steers the steered wheels mechanically separated from the steering wheel according to the steering state of the steering wheel by actuating an actuator;
A vehicle suspension device that rotatably supports the steered wheel,
The vehicle suspension apparatus includes an axle carrier having an axle that rotatably supports the steered wheel, a lower link structure that connects a vehicle body side support portion and the axle carrier below the axle, and the axle. An upper link structure for connecting the vehicle body side support portion and the axle carrier on the upper side, and a kingpin shaft passing through the lower pivot point of the lower link structure and the upper pivot point of the upper link structure is set to a neutral position of the steering wheel And set it to pass through the tire ground contact surface,
The steered control device operates the actuator to steer the steered wheels so that the steered angle of the steered wheels is restored, thereby ensuring straightness of the vehicle,
When the steered wheel is a turning outer wheel, at least one pivot point of the lower link structure and the upper link structure is set when at least the steered wheel changes from a straight traveling state to a steered state and enters a braking state. , Provided with a pivot point moving part that moves so that the kingpin shaft tilts backward in a vehicle side view ,
In the upper link structure, the first upper link member is constituted by a transverse link member, the second upper link member is constituted by a compression link member, and a vehicle body side connection point of the second upper link member is formed. A vehicle steering apparatus, wherein the rigidity is set to be smaller than the rigidity of a vehicle body side connection point of the first upper link member . The lower link structure has a first link member and a second link member intersecting each other,
The pivot point moving unit has a link arrangement that moves a virtual lower pivot point represented by an intersection of the first link member and the second link member at least forward in the vehicle front-rear direction when the steered wheels are steered. The vehicle steering apparatus according to claim 1 , wherein: The upper link structure has a first upper link member and a second upper link member,
The pivot point moving unit is configured to be a link arrangement that moves the virtual upper pivot point of the first upper link member and the second upper link member at least rearward in the vehicle front-rear direction when turning the steered wheels. The vehicular steering apparatus according to claim 1 or 5 .   A pivot point forcible moving unit is provided for forcibly moving the pivot point so that the kingpin shaft tilts backward in a side view of the vehicle when the steering control of the steering control device is abnormal. Item 5. The vehicle steering device according to any one of Items 1 to 4. The vehicle steering apparatus according to any one of claims 1 to 7 , wherein the steering control apparatus includes a steering angle control unit that estimates a compliance steer and corrects a displacement of a steered wheel. . The steering control device includes a straightness ensuring unit that ensures straightness of the vehicle by the steering control according to any of disturbance, self-aligning torque, and compliance steer. The vehicle steering apparatus according to any one of 1 to 8 . The vehicular steering device according to claim 9 , wherein the straight travel guarantee unit estimates a self-aligning torque to secure straight travel of the vehicle. When the steering control device starts steering the steering wheel from a neutral position, the steering control device adjusts the straightness guaranteeing control by the straightness guaranteeing unit to adjust the initial steering response to the steering response of the suspension device itself. The vehicle steering system according to claim 9 or 10 , further comprising a steering response setting unit that sets the steering response. The steering control device sets a high steering response with the steering response of the suspension device itself in the initial steering state when at least the steering wheel starts to be steered from the neutral position. The steering response setting unit is provided for setting a steering response required by the straight travel guarantee control by the straight travel guarantee unit when the steered state has passed. The vehicle steering apparatus according to any one of 9 to 11 . The steering response setting unit, wherein when the steer the steering wheel from the neutral position, to claim 12, characterized in that it comprises a delay control section that delays straightness secured control of the straightness secured portion Vehicle steering system. 14. The vehicle steering apparatus according to claim 13 , wherein the delay control unit includes a gain adjustment unit that adjusts the start of straightness ensuring control by the straightness ensuring unit. The delay control unit starts the straightness ensuring control by the straightness ensuring unit after delaying by 0.1 second from the steering start timing when the steering wheel is steered to the right or left from the neutral position. The vehicle steering apparatus according to claim 13 or 14 , characterized in that: The steering control device includes a target turning angle calculation unit that calculates a target turning angle according to a steering angle, and the straightness of the straight travel guarantee unit at the target turning angle calculated by the target turning angle calculation unit. An adder that adds a collateral control value, a turning motor control unit that forms a motor command current that matches an addition output of the adder and a rotation angle of a turning motor that constitutes the turning actuator, The vehicle steering apparatus according to any one of claims 8 to 15 , further comprising: a current control unit that forms a motor drive current supplied to the steering motor that coincides with the steering motor. The steering control device includes at least a steering angle detection unit that detects a steering angle of the steering wheel, a steering operation detection unit that detects a steering operation of the steered wheels by the steering actuator, and the steering angle detection unit. An abnormality control unit that monitors the steering angle detected by the steering operation and the steering operation detected by the steering operation detection unit, performs abnormality determination, and stops control of the steering actuator when it is determined to be abnormal. The vehicle turning device according to any one of claims 1 to 16 , wherein The steering control device includes a steering angle detection unit that detects a steering angle of the steering wheel, a steering angle detection unit that detects a turning angle of the steered wheel by the steering actuator, the steering angle and the turning angle. When the deviation of the steering angle is greater than or equal to a predetermined value, it is provided with an abnormality control unit that determines an abnormality and stops the control of the steering actuator. The vehicle steering device according to claim 7 , wherein the vehicle steering device is operated.

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