JP4876924B2 - Roll control device for vehicle - Google Patents

Description

  The present invention relates to a roll control device for a vehicle, and more specifically, a roll control device for a vehicle including a damping force generating means with variable damping force and an anti-roll moment generating means for increasing or decreasing the anti-roll moment when the vehicle turns. Concerning.

  In vehicles such as automobiles equipped with variable damping force type shock absorbers, it is well known that the roll of the vehicle body is suppressed by increasing / decreasing the damping force of the left and right shock absorbers when the vehicle is turning. . For example, in the following Patent Document 1 relating to the application of the present applicant, by calculating the damping force difference between the inner and outer wheels during a transient turn based on a vehicle model in which a virtual damper is set inside the turn when the vehicle turns. A damping coefficient control device is described that suppresses the roll of the vehicle body and lowers the center of gravity of the vehicle body to improve the running performance of the vehicle during a transient turn. If the damping force of the shock absorber is controlled as described above, the roll of the vehicle body during the transient turning of the vehicle can be reduced and the running stability of the vehicle can be improved as compared with the case where the damping force of the shock absorber is not controlled. Can do.

In vehicles such as automobiles equipped with anti-roll moment generating means such as an active stabilizer device, the roll of the vehicle body is suppressed by increasing / decreasing the anti-roll moment generated by the anti-roll moment generating means when the vehicle turns. It is also well known to do. If the anti-roll moment generated by the anti-roll moment generating means during turning of the vehicle is controlled to increase / decrease, the roll of the vehicle body at the time of transient turning of the vehicle and steady turning will be less than when the increase / decrease of the anti-roll moment is not controlled. It is possible to reduce the driving stability of the vehicle.
Japanese Patent No. 3509544

  Since the control of the damping force of the shock absorber when turning the vehicle and the control of the anti-roll moment generated by the anti-roll moment generating means are both controls for suppressing the roll of the vehicle body, the shock absorber and the anti-roll moment generating means In the case of a vehicle having both of the above, the damping force of the shock absorber and the anti-roll moment generated by the anti-roll moment generating means must be appropriately controlled in relation to each other.

  In addition, since the damping force that can be generated by the shock absorber and the anti-roll moment that can be generated by the anti-roll moment generating means are both finite, the anti-roll moment generating means is not suitable for situations where the roll moment acting on the vehicle is large. It is conceivable that an anti-roll moment that opposes an excessive roll moment that cannot be dealt with is generated by damping forces acting in opposite directions on the contraction side and the extension side of the shock absorber.

  However, in general, the maximum damping force on the contraction side of the shock absorber is smaller than the maximum damping force on the expansion side. Therefore, when the magnitude of the anti-roll moment that should be generated by the damping force of the shock absorber is large, the damping required on the contraction side is required. The force exceeds its maximum damping force, and not only can the anti-roll moment required for the shock absorber not be generated, but also the control interference problem that the original damping force control of the shock absorber is disturbed occurs. .

For example, if the vehicle tread is T, the damping force of the shock absorber of the turning outer ring and the turning inner wheel by the shock absorber's original damping force control is Fsout and Fsin, respectively, and the upward direction is the positive direction of the damping force, the turning outer wheel side The damping force Fsout on the contraction side acts upward and has a positive value, whereas the damping force on the turning inner ring side (extension side) acts downward and has a negative value, so that the shock absorber's natural damping The anti-roll moment Mara by force control is expressed by the following formula A.
Mara = Fsout · T / 2−Fsin · T / 2 (A)

Further, if the additional damping forces of the outer and inner wheels required for the shock absorber to generate an anti-roll moment against the excessive roll moment are ΔFsout and ΔFsin (both positive values), the additional damping force of the shock absorber The anti-roll moment Marb based on (the correction amount of the damping force) is expressed by the following formula B.
Marb = ΔFsout · T / 2 + ΔFsin · T / 2 (B)

When the sum Fsout + ΔFsout of the damping force Fsout of the original damping force control of the turning outer wheel and the additional damping force ΔFsout is equal to or less than the maximum damping force Fscmax on the contraction side of the shock absorber, the shock absorber attenuates the sum Fsout + ΔFsout as the damping force on the contraction side. Since a force can be generated, the anti-roll moment Marab due to the damping force of the original damping force control and the additional damping force is expressed by the following formula C.
Marab = (Fsout + ΔFsout−Fsin + ΔFsin) · T / 2 (C)

However, when the sum Fsout + ΔFsout of the damping force Fsout of the original damping force control of the turning outer wheel and the additional damping force ΔFsout exceeds the maximum damping force Fscmax on the contraction side of the shock absorber, the shock absorber uses the maximum damping force Fscmax as the damping force on the contraction side. However, since the damping force of the sum Fsout + ΔFsout cannot be generated, the anti-roll moment Marab ′ in this case is expressed by the following formula D.
Marab ′ = (Fscmax−Fsin + ΔFsin) · T / 2 (D)

  Since the maximum damping force Fscmax is smaller than the sum Fsout + ΔFsout, the anti-roll moment Marab ′ is smaller than the necessary anti-roll moment Marab, and the roll of the vehicle body cannot be reliably suppressed.

Further, the sum Fstotal of the damping forces of the left and right wheels when the original damping force control of the shock absorber is executed is expressed by the following equation E.
Fstotal = Fsout + Fsin (E)

In addition to the damping forces Fsout and Fsin of the original damping force control, the shock absorbers of the turning outer wheel and the turning inner wheel generate additional damping forces ΔFsout and ΔFsin, respectively, so as to generate an anti-roll moment against the excessive roll moment. Sometimes, the damping forces of the outer turning wheel and the inner turning wheel are Fsout + ΔFsout and Fsin−ΔFsin, respectively. Therefore, the sum Fstotal of the damping forces of the left and right wheels is expressed by the following equation F.
Fstotal = Fsout + ΔFsout + Fsin−ΔFsin (F)

  Therefore, if the magnitudes of the additional damping forces ΔFsout and ΔFsin are the same, the value of the above equation E becomes the same as the value of the above equation E, and therefore the original damping force control of the shock absorber generates an excessive roll moment. Are not inhibited by the additional damping forces ΔFsout and ΔFsin.

However, when the sum Fsout + ΔFsout of the damping force Fsout of the original roll restraining control and the additional damping force ΔFsout of the turning outer wheel exceeds the maximum damping force Fscmax on the contraction side of the shock absorber, the sum Fstotal of the damping forces of the left and right wheels is Since the value is expressed by the formula G and is not the same as the value of the formula E, the original damping force control of the shock absorber is inhibited by the additional damping forces ΔFsout and ΔFsin for generating an excessive roll moment. End up.
Fstotal = Fscmax + Fsin−ΔFsin (G)

The present invention suppresses the roll of the vehicle body by controlling both the damping force by the damping force generating means of variable damping force type and the anti-roll moment by the anti-roll moment generating means, and the necessary anti-roll moment is reduced to the anti-roll moment. When the anti-roll moment that can be generated by the generating means is exceeded, the above-mentioned problems in the case where the insufficient anti-roll moment is generated by the damping force of the damping force generating means have been made. The main problem is that when the damping force required for at least one of the left and right damping force generating means exceeds the maximum damping force of the damping force generating means, the damping force of the opposite left and right damping force generating means is also corrected. Therefore, it is necessary as much as possible without obstructing the original damping force control of the damping force generating means. Is to control the damping force of the damping force generating means to inch roll moment is generated.
[Means for Solving the Problems and Effects of the Invention]

  According to the present invention, the main problem described above includes the structure of claim 1, that is, the damping force generating means of variable damping force type disposed between the sprung and unsprung parts corresponding to each wheel. A damping force control means for controlling the damping force generating means so that the damping force becomes a target damping force for setting the vehicle posture to a target posture when the vehicle turns, and acts between the sprung and unsprung portions. The anti-roll moment generating means for increasing / decreasing the anti-roll moment by increasing / decreasing the force to be applied so that the anti-roll moment becomes the target anti-roll moment for reducing the roll of the vehicle when the vehicle turns. An anti-roll moment control means for controlling the anti-roll moment, and the target anti-roll moment is generated by the anti-roll moment control means. When the maximum anti-roll moment of the means is exceeded, the target damping force of the left and right wheels is corrected so as to reduce the roll of the vehicle, and the corrected target damping force of at least one of the left and right wheels is the damping force of the wheel. When the maximum damping force of the generating means is exceeded, for the wheel with the larger excess damping force, the target damping force of the wheel is reduced and corrected with the larger excess damping force, and the right and left sides are reversed. This wheel is achieved by a vehicle roll control device characterized in that the target damping force of the wheel is reduced and corrected with the larger excess damping force.

  According to the configuration of claim 1, when the target anti-roll moment exceeds the maximum anti-roll moment of the anti-roll moment generating means, the target damping force of the left and right wheels is corrected so as to reduce the roll of the vehicle. The anti-roll moment can be generated at least partly by the damping force, which ensures that the body roll during transient turning is suppressed compared to when the target damping force is not modified to reduce the vehicle roll. can do.

  According to the configuration of claim 1, when the corrected target damping force of at least one of the left and right wheels exceeds the maximum damping force of the damping force generating means of the wheel, the excess damping force is large. For the other wheel, the correction is reduced by the excess damping force of the larger target damping force of the wheel, and for the opposite wheel, the target damping force of the wheel is larger than the larger excess of the target damping force. Since it is reduced and corrected by the damping force, an insufficient anti-roll moment can be generated as large as possible by the damping force as described in detail later, and the adverse effect on the control of the original damping force is reduced as much as possible. Thus, the roll of the vehicle body can be effectively and appropriately suppressed.

  According to the present invention, in order to effectively achieve the above main problem, in the configuration of claim 1, the roll control device is configured such that the target anti-roll moment is a maximum anti-roll moment generating means. When the roll moment is exceeded, the target damping force is modified so that an excess anti-roll moment is generated by an increase in the damping force.

  According to the configuration of the second aspect, when the target anti-roll moment exceeds the maximum anti-roll moment of the anti-roll moment generating means, the target damping force is corrected so that the excess anti-roll moment is generated by the increase of the damping force. Therefore, the anti-roll moment that is insufficient due to the fact that the anti-roll moment generation means cannot be generated can be reliably generated by the damping force without excess or deficiency.

  According to the present invention, in order to effectively achieve the above-mentioned main problem, in the configuration of claim 1 or 2, the roll control device is sized with the upward direction as the positive direction of the damping force. Is configured to increase and correct the target damping force of the outer turning wheel and to reduce and correct the target damping force of the inner turning wheel with the same correction amount (configuration of claim 3).

  According to the configuration of the third aspect, the target damping force of the outer turning wheel is corrected to be increased and the target damping force of the inner turning wheel is corrected to be reduced by a correction amount having the same magnitude, so that it will be described in detail later. Thus, the correction of the target damping force can be reliably prevented from adversely affecting the control of the original damping force.

  According to the present invention, in order to effectively achieve the main problems described above, in the configuration according to any one of claims 1 to 3, the damping force generating means is a shock absorber capable of increasing or decreasing a damping coefficient. And the damping force control means controls the damping coefficient of the shock absorber so that the damping coefficient becomes a target damping coefficient corresponding to the target damping force, and the roll control device controls the target damping force corresponding to the corrected target damping force. The corrected target damping force is estimated based on the damping coefficient and the relative vertical speed of the sprung and unsprung parts, and based on the maximum value of the damping coefficient and the relative vertical speed of the sprung and unsprung parts. It is comprised so that the maximum damping force of the said damping force generation means may be estimated (structure of Claim 4).

  According to the configuration of the fourth aspect, the damping coefficient of the shock absorber is controlled so that the damping coefficient becomes the target damping coefficient corresponding to the target damping force, and the target damping coefficient corresponding to the corrected target damping force and the sprung and The corrected target damping force is estimated based on the unsprung relative vertical speed, so that the damping force of the shock absorber can be reliably controlled to the target damping force, and the corrected target damping force can be accurately determined. Since the maximum damping force of the damping force generating means can be estimated based on the maximum value of the damping coefficient and the relative vertical speed of the sprung and unsprung parts, the maximum damping force of the damping force generating means can be accurately determined. Can be estimated.

  According to the present invention, in any one of the first to fourth aspects, the anti-roll moment generating means is configured to be an active stabilizer device capable of increasing or decreasing the stabilizer force. ).

  According to the configuration of the fifth aspect, since the anti-roll moment generating means is an active stabilizer device capable of increasing / decreasing the stabilizer force, the excess anti-roll moment can be reliably estimated according to the operating state of the active stabilizer device. .

  According to the present invention, in order to effectively achieve the main problem described above, in the configuration of claim 2, the roll control device is configured to reduce the target damping force based on the excess anti-roll moment. The correction amount is calculated (configuration of claim 6).

According to the configuration of claim 6, since the correction amount of the target damping force is calculated based on the excess anti-roll moment, the anti-roll moment that is insufficient due to the fact that the anti-roll moment generating means cannot be generated. The amount of correction of the target damping force necessary for generating the roll moment can be accurately calculated.
[Preferred embodiment of problem solving means]

  According to one preferable aspect of the present invention, in the configuration according to any one of the first to fifth aspects, the corrected target damping force of one of the left and right wheels is the maximum damping force of the damping force generating means of the wheel. And the corrected target damping force of the opposite wheel is less than or equal to the maximum damping force of the wheel damping force generating means for one wheel, the excess damping of the target damping force of that wheel is exceeded for one wheel A reduction correction is performed with force, and the opposite right and left wheels are configured to reduce and correct the target damping force of the wheel with an excess damping force (preferred aspect 1).

  According to another preferred aspect of the present invention, in the configuration according to any one of the first to fifth aspects or the preferred aspect 1, the corrected target damping forces of both the left and right wheels are respectively applied to the wheels. When exceeding the maximum damping force of the damping force generating means, for the wheel with the larger excess damping force, the target damping force of the wheel is reduced and corrected with the larger excess damping force, The wheel on the opposite side is configured to reduce and correct the target damping force of the wheel with the larger excess damping force (preferred aspect 2).

  According to another preferred aspect of the present invention, in the structure according to any one of the first to fifth aspects or the preferred aspect 1 or 2, the anti-roll moment generating means is an anti-roll moment generating means on the front wheel side. It is comprised so that it may comprise the anti-roll moment generation | occurrence | production means of the rear wheel side (Preferable aspect 3).

  According to another preferred aspect of the present invention, in the configuration of the preferred aspect 3 described above, the roll control device calculates a target anti-roll moment of the entire vehicle for reducing the roll of the vehicle. The target anti-roll moment on the front wheel side and the rear wheel side is calculated based on the anti-roll moment and the front / rear wheel distribution ratio of roll rigidity, and the target anti-roll moment on the front wheel side is the maximum anti-roll moment of the anti-roll moment generating means on the front wheel side. If the target damping force of the left and right front wheels is corrected so as to reduce the roll of the vehicle and the corrected target damping force of at least one of the left and right front wheels exceeds the maximum damping force of the damping force generating means of the wheel, the excess is exceeded. For the front wheel with the larger damping force of the minute, the target damping force of the front wheel is set to the damping force of the larger excess. The front wheel on the opposite left and right sides is corrected to reduce the magnitude of the target damping force of the front wheel with the larger excess damping force, and the target anti-roll moment on the rear wheel side is adjusted to the rear wheel side. When the maximum anti-roll moment of the anti-roll moment generating means is exceeded, the target damping force of the left and right rear wheels is corrected so as to reduce the roll of the vehicle, and the corrected target damping force of at least one of the left and right rear wheels is the damping of the relevant wheel. When exceeding the maximum damping force of the force generating means, for the rear wheel having a larger excess damping force, the target damping force of the rear wheel is reduced and corrected by the larger excess damping force, The rear wheel on the opposite side is configured to reduce and correct the target damping force of the rear wheel with the larger excess damping force (preferred aspect 4).

  According to another preferred embodiment of the present invention, in the configuration of the preferred embodiment 4, the corrected target damping force of one of the left and right front wheels exceeds the maximum damping force of the damping force generating means of the wheel. When the corrected target damping force of the front wheel on the opposite side is less than or equal to the maximum damping force of the wheel damping force generating means, the target damping force of the wheel is reduced by the excess damping force for one front wheel It corrects, and it is comprised so that the magnitude | size of the target damping force of the said wheel may be reduced and corrected with the excess damping force about the front wheel on the opposite left and right sides (preferable aspect 5).

  According to another preferred aspect of the present invention, in the configuration of the preferred aspect 4 or 5, the corrected target damping force of one of the left and right rear wheels is the maximum damping force of the damping force generating means of the wheel. If the target damping force after correction of the rear wheel on the opposite left and right side is less than or equal to the maximum damping force of the damping force generating means of the wheel, the target damping force of the wheel is exceeded for one of the rear wheels. A reduction correction is made with the damping force, and the rear wheels on the opposite left and right sides are configured to reduce and correct the magnitude of the target damping force of the wheel with the excess damping force (preferred aspect 6).

  According to another preferred aspect of the present invention, in the structure according to any one of claims 2 to 5 or preferred aspects 1 to 6, the anti-roll moment and the turning depending on the correction amount of the target damping force of the turning outer wheel. The target damping force of the left and right wheels is corrected so that the sum of the anti-roll moment based on the correction amount of the target damping force of the inner ring becomes an excess anti-roll moment (preferred aspect 7).

The present invention will now be described in detail with reference to a few preferred embodiments with reference to the accompanying drawings.
[First embodiment]

  FIG. 1 is a schematic configuration diagram showing a first embodiment of a roll control device for a vehicle according to the present invention applied to a vehicle having an active stabilizer device on the front wheel side and the rear wheel side.

  In FIG. 1, 10FL and 10FR respectively indicate left and right front wheels of the vehicle 12, and 10RL and 10RR respectively indicate left and right rear wheels of the vehicle 12. An active stabilizer device 16 is provided between the left and right front wheels 10FL and 10FR, and an active stabilizer device 18 is provided between the left and right rear wheels 10RL and 10RR. The active stabilizer devices 16 and 18 suppress the roll of the vehicle body by applying an anti-roll moment to the vehicle (vehicle body) when the vehicle turns.

  The active stabilizer device 16 is integrally connected to a pair of torsion bar portions 16TL and 16TR extending coaxially with each other along an axis extending in the lateral direction of the vehicle, and the outer ends of the torsion bar portions 16TL and 16TR, respectively. And a pair of arm portions 16AL and 16AR. The torsion bar portions 16TL and 16TR are respectively supported by a vehicle body not shown in the drawing via brackets not shown in the drawing so as to be rotatable around its own axis. The arm portions 16AL and 16AR extend in the vehicle front-rear direction so as to intersect the torsion bar portions 16TL and 16TR, respectively, and the outer ends of the arm portions 16AL and 16AR are suspension members 14FL such as suspension arms of the left and right front wheels 10FL and 10FR, respectively. It is linked to 14FR.

  The active stabilizer device 16 has an actuator 20F between the torsion bar portions 16TL and 16TR. Actuator 20F has a built-in electric motor, and rotationally drives a pair of torsion bar portions 16TL and 16TR as necessary, so that torsional stress is generated when the left and right front wheels 10FL and 10FR bounce and rebound in opposite phases to each other. Thus, the force for suppressing the bounce and rebound of the wheel is changed, thereby increasing or decreasing the anti-roll moment applied to the vehicle at the position of the left and right front wheels, thereby suppressing the roll of the vehicle body on the front wheel side.

  Similarly, the active stabilizer device 18 includes a pair of torsion bar portions 18TL and 18TR extending coaxially with each other along an axis extending in the lateral direction of the vehicle, and outer ends of the torsion bar portions 18TL and 18TR, respectively. It has a pair of arm portions 18AL and 18AR connected together. The torsion bar portions 18TL and 18TR are respectively supported by a vehicle body not shown in the drawing via brackets not shown in the drawing so as to be rotatable around its own axis. The arm portions 18AL and 18AR extend in the vehicle longitudinal direction so as to intersect the torsion bar portions 18TL and 18TR, respectively, and the outer ends of the arm portions 18AL and 18AR are suspension members 14RL such as suspension arms of the left and right rear wheels 10RL and 10RR, respectively. And 14RR.

  The active stabilizer device 18 has an actuator 20R between the torsion bar portions 18TL and 18TR. The actuator 20R has a built-in electric motor, and if necessary, the pair of torsion bar portions 18TL and 18TR are driven to rotate relative to each other, so that the left and right rear wheels 10RL and 10RR bounce and rebound in opposite phases. The force for suppressing the bounce and rebound of the wheel is changed by the stress, thereby increasing or decreasing the anti-roll moment applied to the vehicle at the positions of the left and right rear wheels, thereby suppressing the roll of the vehicle body on the rear wheel side.

  Since the structures of the active stabilizer devices 16 and 18 do not form the gist of the present invention, as long as the rolls of the front and rear wheels can be suppressed by functioning as anti-roll moment generating means. Any configuration known in the art may be used.

  The actuators 20F and 20R of the active stabilizer devices 16 and 18 are controlled by the electronic control device 22 controlling the control current for the motor. Although not shown in detail in FIG. 1, the electronic control unit 22 includes a CPU, a ROM, a RAM, and an input / output port device, which are connected to each other via a bidirectional common bus and a drive circuit. It may be better.

  FIG. 2 shows the suspension of the right front wheel as a whole. The wheel 10FR is supported by a wheel support member 26 so as to be rotatable around a rotation axis 28. The outer ends of the upper arm 34 and the lower arm 36 are pivotally attached to the upper end and the lower end of the wheel support member 26 by ball joints 30 and 32, respectively. The inner ends of the upper arm 34 and the lower arm 36 are pivotally attached to the vehicle body 42 by rubber bush devices 38 and 40, respectively.

  As shown in FIGS. 1 and 2, each wheel is provided with variable damping force type shock absorbers 44FL to 44RR as damping force generating means. Each of the shock absorbers 44FL to 44RR has a cylinder 46 and a piston 48 that is reciprocally fitted to the cylinder 46, and is connected to the vehicle body 42 by an upper support 50 at the upper end of the piston 48. To the lower arm 36.

  Although not shown in FIG. 2, each of the shock absorbers 44FL to 44RR has a plurality of actuators 48C provided with the opening degree (throttle degree) of the damping force generating valves 48A and 48B provided on the piston 48 corresponding to each other. The damping coefficient is changed by increasing / decreasing in multiple stages over the control stage. Each of the shock absorbers 44FL to 44RR generates a damping force according to the relative speed and damping coefficient of the piston 48 with respect to the cylinder 46 associated with the bounding and rebounding of the wheels 10FL to 10RR. Suppresses changes in body posture during acceleration / deceleration and turning.

  Note that the higher the control stage, the smaller the opening of the damping force generation valves 48A and 48B and the higher the damping coefficient. Further, when the relative speeds of the pistons 48 are the same, the expansion side damping force is larger than the contraction side damping force. Further, the shock absorber may continuously change the damping coefficient by continuously increasing or decreasing the opening degree of the damping force generating valve.

  In FIG. 2, reference numeral 50 denotes a suspension spring disposed in parallel with the shock absorbers 44FL to 44RR at each wheel. The suspension spring 50 is fixed to an upper support 56 attached to the vehicle body 52. It is elastically mounted between the upper seat 58 and the lower seat 60 fixed to the lower support 58 attached to the lower arm 36 in a compressed state.

  Thus, the vehicle body 42 is a spring on the shock absorbers 44FL to 44RR and the suspension spring 50, and the wheel support member 26, the upper arm 34, and the lower arm 36 cooperate with each other to form the suspension members 14FL to 14RR, and the shock absorbers 44FL to 44RR and It is unsprung with respect to the suspension spring 50. Although not shown in FIGS. 1 and 2, the outer ends of the arm portions 16AL, 16AR, 18AL, and 18AR of the active stabilizer devices 16 and 18 are connected to the upper arm 34 or the lower arm 36 through connection links.

  The suspension shown in FIG. 2 is a double wishbone suspension. However, the suspension of the vehicle to which the roll control device of the present invention is applied is not limited to the double wishbone suspension, and is a McPherson strut. It may be any type of suspension known in the art, such as a suspension of the type or a trailing arm type.

  The control stage of each shock absorber 24FL to 24RR, and thus the opening of the damping force generating valve, is controlled by controlling the control current for the corresponding actuator 48C by the electronic control unit 62. Although not shown in detail in FIG. 1, the electronic control unit 62 also has a CPU, a ROM, a RAM, and an input / output port device, and these are connected to each other by a bidirectional common bus and a driving circuit. It may be better. Further, the electronic control units 22 and 62 communicate with each other to exchange necessary signals.

  As shown in FIG. 1, the electronic control unit 22 includes a signal indicating the vehicle lateral acceleration Gy detected by the lateral acceleration sensor 64, a signal indicating the vehicle speed V detected by the vehicle speed sensor 66, a rotation angle sensor 68F, Signals indicating the actual rotation angles φF and φR of the actuators 20F and 20R detected by 68R are input.

  On the other hand, the electronic control unit 62 includes a signal indicating the longitudinal acceleration Gx of the vehicle detected by the longitudinal acceleration sensor 70, a signal indicating the steering angle θ detected by the steering angle sensor 72, and the stroke of each wheel detected by the stroke sensor 74i. A signal indicating Xi (i = fl, fr, rl, rr) and a signal indicating the vertical acceleration Gzi (i = fl, fr, rl, rr) of the vehicle body at the position corresponding to each wheel detected by the vertical acceleration sensor 76i Is entered.

  The lateral acceleration sensor 40 and the steering angle sensor 46 detect the lateral acceleration Gy and the steering angle θ with positive values generated when the vehicle turns to the right, and the longitudinal acceleration sensor 70 detects the longitudinal acceleration Gx with the acceleration direction of the vehicle as positive. Is detected. The rotation angle sensors 68F and 68R detect the rotation angles φF and φR with a positive value in the direction of reducing the roll of the vehicle body when the vehicle turns left, and the stroke sensor 74i sets the stroke in the bounce direction of the wheel as positive. The stroke Xi of the wheel is detected with the stroke in the rebound direction being negative. The vertical acceleration sensor 76i detects the vertical acceleration Gzi of the vehicle body with the upward acceleration as positive and the downward acceleration as negative.

  The electronic control unit 22 calculates the target anti-roll moment Mat of the vehicle so as to increase the anti-roll moment in the direction to cancel the roll moment acting on the vehicle based on the lateral acceleration Gy of the vehicle according to the flowchart shown in FIG. Based on the vehicle speed V, a target roll stiffness distribution ratio Rmf for the front wheels is calculated. Then, the electronic control unit 22 calculates the front anti-roll moment Maft and the rear anti-roll moment Mart based on the anti-roll moment Mat and the front roll stiffness distribution ratio Rmf. Regarding the anti-roll moment, the direction of the anti-roll moment that should be generated when the vehicle turns left is positive.

  In addition, the electronic control unit 22 determines that the target anti-roll moments Maft and Mart are the maximum anti-roll moments Mafmax and Marmax of the active stabilizer devices 16 and 18 (the anti-roll moments when the actuators 20F and 20R are rotated to the maximum, respectively). ), The magnitudes of the target anti-roll moments Maft and Mart are corrected to Mafmax and Marmax, respectively, and the deviations between the target anti-roll moments Maft and Mart and the maximum anti-roll moments Mafmax and Marmax before correction are exceeded. Roll moments ΔMaf and ΔMar are calculated.

  Further, the electronic control unit 22 calculates target rotation angles φFt and φRt of the actuators 20F and 20R of the active stabilizer devices 16 and 18 based on the target anti-roll moments Maft and Mart, respectively, and the rotation angles φF and φR of the actuators 20F and 20R are calculated. Control is performed so that the respective target rotation angles φFt and φRt correspond to each other, whereby the roll of the vehicle at the time of turning or the like is reduced with a roll rigidity with a preferable front-rear distribution ratio. Further, the electronic control unit 22 outputs signals indicating the excess roll moments ΔMaf and ΔMar to the electronic control unit 62.

  On the other hand, the electronic control unit 62 performs the basic target damping force Fsbi (i) of each shock absorber 44FL to 44RR in a manner known in the art based on the longitudinal acceleration Gx of the vehicle according to the flowcharts shown in FIGS. = fl, fr, rl, rr), the required increase / decrease damping force Fsdf for the left and right front wheel shock absorbers 44FL and 44FR as the increase / decrease amount of the damping force required by the electronic control unit 22 based on the excess roll moments ΔMaf and ΔMar, The required increase / decrease damping force Fsdr for the left and right rear wheel shock absorbers 44RL and 44RR is calculated. Regarding the damping force, the upward direction is positive.

  The electronic control unit 62 basically sets the value obtained by correcting the basic target damping force Fsbi by the required damping force Fsdi as the target damping force Fsti (i = fl, fr, rl, rr) of each shock absorber 44FL to 44RR. When at least one of the target damping force Fsti exceeds the maximum damping force that can be generated by each of the shock absorbers 44FL to 44RR, the target damping force of the wheel is larger for the wheel having the larger excess damping force. Fsti is corrected to be reduced by the larger excess damping force, and for the opposite left and right wheels, the target damping force Fsti of the wheel is increased and corrected by the larger excess damping force.

  Further, the electronic control unit 62 calculates the differential value Xdi (i = fl, fr, rl, rr) of the stroke Xi detected by the stroke sensor 74i as the stroke speed of each wheel, and obtains the target damping force Fsti and the stroke speed Xdi. Based on this, the target control stages Sti (i = fl, fr, rl, rr) of the shock absorbers 44FL to 44RR are calculated, and control is performed such that the control stages Si of the shock absorbers 44FL to 44RR respectively correspond to the corresponding target control stages Sti.

  Next, the control of the active stabilizer device in the illustrated first embodiment will be described with reference to the flowchart shown in FIG. Note that the flowchart control shown in FIG. 2 is executed by the electronic control unit 22, is started by closing an ignition switch not shown in the figure, and is repeatedly executed every predetermined time.

  In step 10, a signal indicating the vehicle lateral acceleration Gy is read. In step 20, the larger the vehicle lateral acceleration Gy, the larger the target anti-roll moment Mat. A target anti-roll moment Mat is calculated from a map corresponding to the graph shown in FIG. 10 based on the lateral acceleration Gy of the vehicle. In step 30, the target roll stiffness distribution ratio Rmf of the front wheels is increased so that the vehicle speed V increases. Is calculated as a value greater than 0 and less than 1. The target roll stiffness distribution ratio Rmf for the front wheels may be a constant.

In step 40, based on the target anti-roll moment Mat and the front-wheel target roll stiffness distribution ratio Rmf, the front-wheel target anti-roll moment Maft and the rear-wheel target anti-roll moment Mart are calculated according to the following equations 1 and 2, respectively. The
Maft = Rmf · Mat (1)
Mart = (1-Rmf) Mat (2)

  In step 50, it is determined whether or not the target anti-roll moment Maft of the front wheel exceeds the maximum anti-roll moment Mafmax (positive constant) that can be generated by the active stabilizer device 16 on the front wheel side, and a negative determination is made. If YES, the routine proceeds to step 70. If an affirmative determination is made, the target anti-roll moment Maft of the front wheels is set to the maximum anti-roll moment Mafmax in step 60, and the excess anti-roll moment ΔMaf is set to the target anti-roll moment. Calculated as a deviation between Maft and the maximum anti-roll moment Mafmax, and then proceeds to step 90.

  In step 90, it is determined whether or not the target anti-roll moment Maft of the rear wheel is smaller than the sign inversion value −Mafmax of the maximum anti-roll moment Mafmax that can be generated by the active stabilizer device 16 on the rear wheel side. When a negative determination is made, the routine proceeds to step 90. When an affirmative determination is made, the target anti-roll moment Maft of the rear wheel is set to -Mafmax at step 80, and the excess anti-roll moment ΔMaf is set to the target anti-roll. The deviation between the moment Maft and the sign inversion value -Mafmax of the maximum anti-roll moment Mafmax, and thus the sum of the target anti-roll moment Maft and the maximum anti-roll moment Mafmax is calculated.

  In step 90, it is determined whether or not the rear wheel target anti-roll moment Mart exceeds the maximum anti-roll moment Marmax (positive constant) that can be generated by the active stabilizer device 18 on the rear wheel side. When the determination is made, the routine proceeds to step 110. When the determination is affirmative, at step 100, the target anti-roll moment Mart of the rear wheel is set to the maximum anti-roll moment Marmax and the excess anti-roll moment ΔMar is set to the target. Calculated as a deviation between the anti-roll moment Mart and the maximum anti-roll moment Marmax, and then proceeds to step 130.

  In step 110, it is determined whether or not the rear wheel target anti-roll moment Mart is smaller than the sign-inverted value -Marmax of the maximum anti-roll moment Marmax that can be generated by the active stabilizer device 18 on the rear wheel side. When a negative determination is made, the routine proceeds to step 130. When an affirmative determination is made, the rear wheel target anti-roll moment Mart is set to -Marmax in step 120, and the excess anti-roll moment ΔMar is set to the target anti-roll. The deviation between the moment Mart and the sign-inverted value -Marmax of the maximum anti-roll moment Marmax is calculated as the sum of the target anti-roll moment Mart and the maximum anti-roll moment Marmax.

  In step 130, the target rotational angles φft and φrt of the actuators 20F and 20R of the active stabilizer devices 16 and 18 are calculated based on the front wheel target anti-roll moment Maft and the rear wheel target anti-roll moment Mart, respectively. In this case, the rotation angles φf and φr of the actuators 20F and 20R are controlled to be the target rotation angles φft and φrt, respectively.

  Next, the damping force control of the shock absorber in the illustrated first embodiment will be described with reference to the flowcharts shown in FIGS. Note that the control according to the flowchart shown in FIG. 4 is executed, is started by closing an ignition switch not shown in the figure, and is repeatedly executed every predetermined time. FIG. 5 shows a subroutine of step 240 in the flowchart shown in FIG.

In step 210, a signal indicating the lateral acceleration Gy of the vehicle is read, and in step 220, each shock absorber 44FL when the vehicle turns according to the flowcharts shown in FIGS. A basic target damping force Fsbi (i = fl, fr, rl, rr) is calculated as a target value for the original damping force control of .about.44RR. In step 230, the required increase / decrease damping force Fsdf for the front wheel and the required increase / decrease damping force Fsdr for the rear wheel are respectively required to achieve the excess roll moments ΔMaf and ΔMar by the damping force. Calculation is performed according to the following equations 3 and 4 based on the moments ΔMaf and ΔMar. In the following formulas 3 to 6, Tf and Tr are front and rear treads, respectively.
Fsdf = | ΔMaf | / Tf (3)
Fsdr = | ΔMar | / Tr (4)

  In step 240, in accordance with the flowchart shown in FIG. 5, the right front wheel and the left front wheel are attenuated as correction amounts for the damping force to satisfy the request from the electronic control unit 62 without adversely affecting the original damping force control. Force correction amounts ΔFsfr and ΔFsfl (both values of 0 or more) are calculated as described later.

  In step 260, for example, the turning direction of the vehicle is determined in a manner known in the art based on the sign of the lateral acceleration Gy or the steering angle θ of the vehicle. The left and right front wheel target damping forces Fstfl and Fstfr are calculated according to the following formulas 5 and 6. When the vehicle is not turning left, the left and right front wheel target damping forces Fstfl and Fstfr are calculated according to the following formulas 7 and 8.

Fstfl = Fsbfl−Fsdf + ΔFsfl (5)
Fstfr = Fsbfr + Fsdf-ΔFsfr (6)
Fstfl = Fsbfl + Fsdf−ΔFsfl (7)
Fstfr = Fsbfr−Fsdf + ΔFsfr (8)

  In step 270, the left front wheel target control stage Stfl is calculated based on the left front wheel stroke speed Xdfl and the target damping force Fstfl, and the right front wheel target control based on the right front wheel stroke speed Xdfr and the target damping force Fstfr. The stage Stfr is calculated, and the control stages Sfl and Sfr of the shock absorbers 44FL and 44FR for the left and right front wheels are controlled to be the corresponding target control stages Stfl and Stfr, respectively.

  In step 280, correction amounts ΔFsrl and ΔFsrr (both values of 0 or more) of the damping force of the left and right rear wheels are calculated in the same manner as in step 240, and in step 290, the above steps are calculated. The turning direction of the vehicle is determined in the same manner as in the case of 260, and when the vehicle is in a left turn state, the target damping forces Fstrl and Fstrr of the left and right rear wheels are calculated according to the following equations 9 and 10, and the vehicle is When the vehicle is not in the left turn state, the target damping forces Fstrl and Fstrr for the left and right rear wheels are calculated according to the following equations 11 and 12.

Fstrl = Fsbrl−Fsdr + ΔFsrl (9)
Fstrr = Fsbrr + Fsdr−ΔFsrr (10)
Fstrl = Fsbrl + Fsdr−ΔFsrl (11)
Fstrr = Fsbrr−Fsdr + ΔFsrr (12)

  In step 300, the left rear wheel target control stage Strl is calculated based on the left rear wheel stroke speed Xdrl and the target damping force Fstrl, and the right rear wheel stroke speed Xdrr and the target damping force Fstrr are calculated based on the right rear wheel. The wheel target control stage Strr is calculated, and the control stages Srl and Srr of the shock absorbers 44RL and 44RR for the left and right rear wheels are controlled to be the corresponding target control stages Strl and Strr, respectively.

  In step 241 of the front wheel damping force correction amount calculation routine shown in FIG. 5, the turning direction of the vehicle is determined in the same manner as in the case of step 260 described above, so that the vehicle is turned left. If a negative determination is made, the process proceeds to step 244. If an affirmative determination is made, the process proceeds to step 242.

  In step 242, based on the stroke speed Xdfl of the left front wheel and the current control stage Sfl of the shock absorber 44FL of the left front wheel, the front inner wheel at the turn at the current control stage Sfl is obtained from the map corresponding to the graph shown in FIG. Of the right front wheel stroke speed Xdfr and the current control stage Sfr of the right front wheel shock absorber 44FR based on the map corresponding to the graph shown in FIG. 11 in the current control stage Sfr. The maximum damping force Fsfcmax of the turning outer front wheel is calculated.

In step 243, the excess damping forces Fsflo and Fsfro of the left and right front wheels are calculated according to the following equations 13 and 14, respectively.
Fsflo = | Fsfl−Fsdfl | − | Fsfemax | (13)
Fsfro = | Fsfr + Fsdfr | − | Fsfcmax | (14)

  In step 244, based on the stroke speed Xdfl of the left front wheel and the current control stage Sfl of the shock absorber 44FL of the left front wheel, the turning front wheel at the current control stage Sfl from the map corresponding to the graph shown in FIG. Of the right front wheel stroke speed Xdfr and the current control stage Sfr of the right front wheel shock absorber 44FR based on the map corresponding to the graph shown in FIG. 11 in the current control stage Sfr. The maximum damping force Fsfemax of the turning inner front wheel is calculated.

In step 245, the excess damping forces Fsflo and Fsfro of the left and right front wheels are calculated according to the following equations 15 and 16, respectively.
Fsflo = | Fsfl + Fsdfl | − | Fsfcmax | (15)
Fsfro = | Fsfr−Fsdfr | − | Fsfemax | (16)

In step 246, it is determined whether or not any of the following conditions (1) or (2) is satisfied. If a negative determination is made, the process proceeds to step 252 and an affirmative determination is made. If YES, go to step 247.
(1) Fsfro> 0 and Xdfr> 0
(2) Fsfro <0 and Xdfr <0

In step 247, it is determined whether any of the following conditions (3) or (4) is satisfied. If an affirmative determination is made, the process proceeds to step 249, where a negative determination is made. In step 248, the correction amounts ΔFsfl and ΔFsfr of the damping force of the left and right front wheels are set to the excess damping force Fsfro of the right front wheel, and then the process proceeds to step 250.
(3) Fsflo> 0 and Xdfl> 0
(4) Fsflo <0 and Xdfl <0

  In step 249, it is determined whether or not the deviation between the absolute value of the excess damping force Fsfro of the right front wheel and the absolute value of the excess damping force Fsflo of the left front wheel is a positive value, that is, the excess damping force Fsfro of the right front wheel. Is determined to be greater than the magnitude of the excess damping force Fsflo of the left front wheel. If an affirmative determination is made, the correction amounts ΔFsfl and ΔFsfr of the right and left front wheels are corrected to the right in step 250. After the front wheel excess damping force Fsfro is set, the routine proceeds to step 250, and when a negative determination is made, in step 251, the left and right front wheel damping force correction amounts ΔFsfl and ΔFsfr are set to the left front wheel excess damping force Fsflo. Then go to step 250.

  In step 252, as in the case of step 247, it is determined whether or not any of the above conditions (3) or (4) is satisfied, and if an affirmative determination is made, the process proceeds to step 253. Then, after the correction amounts ΔFsfl and ΔFsfr of the left and right front wheels are set to the excess damping force Fsflo of the left front wheel, the routine proceeds to step 250, and when a negative determination is made, the correction of the damping forces of the left and right front wheels is performed in step 254. After the quantities ΔFsfl and ΔFsfr are set to 0, the routine proceeds to step 250.

  Next, the basic target damping force Fsbi calculation routine executed in step 220 will be described with reference to FIGS.

  First, in step 310, based on the vehicle lateral acceleration Gy, a map corresponding to the graph shown in FIG. 12 is used so that the target roll angle θrt of the vehicle body increases as the lateral acceleration Gy of the vehicle increases. The target roll angle θrt is calculated, and in step 320, the actual roll angle θr of the vehicle body is calculated according to the flowchart shown in FIG.

In step 322 of the flowchart shown in FIG. 7, the wheel base of the vehicle is L, and the distance in the vehicle front-rear direction between the front and rear wheel axles and the center of gravity of the vehicle body is Lf and Lr, respectively. According to 17 and 18, the sprung acceleration Gozl and Gozr at the center of gravity of the left and right wheels of the vehicle body are calculated.
Gozl = (Gzfl·Lr + Gzrl·Lf) / L (17)
Gozr = (Gzfr · Lr + Gzrr · Lf) / L (18)

In step 324, the average value of the front wheel tread Tf and the rear wheel tread Tr is taken as the vehicle wheel tread Tf, and the roll angular acceleration θrdd around the longitudinal axis of the vehicle is calculated according to the following equation 19, In 326, the actual roll angle θr of the vehicle body is calculated by second-order integration of the roll angular acceleration θrdd.
θrdd = (Gozl−Gozr) / T (19)

Returning to the flowchart shown in FIG. 6, in step 330, the corrected roll angle Δθr of the vehicle body is calculated as a deviation (θrt−θr) between the target roll angle θrt of the vehicle body and the actual roll angle θr of the vehicle body, In 340, the corrected roll angular acceleration Δθrdd (= d ² (Δθr) / dt ² ) is calculated by differentiating the corrected roll angle Δθr of the vehicle body by second order time.

In step 350, the roll stiffness Kr of the vehicle is estimated based on the actual rotational angles φF and φR of the actuators 20F and 20R of the active stabilizer devices 16 and 18 transmitted from the electronic control unit 22, and the vehicle longitudinal direction is also determined. The corrected roll moment ΔMr for setting the actual roll angle θr of the vehicle body to the target roll angle θrt is calculated according to the following equation 20, where Ir is the moment of inertia around the axis. The sign of ΔMr is positive when it is in the same direction as the roll angle increasing side and negative when it is in the same direction as the roll angle decreasing side.
ΔMr = Ir · Δθrdd + Kr · Δθr (20)

The corrected damping forces required for the shock absorbers of the turning inner front wheel and the turning outer front wheel are ΔFfin and ΔFfout, respectively. The corrected damping forces required for the shock absorbers of the turning inner rear wheel and the turning outer rear wheel are ΔFrin and ΔFrout, respectively. Then, the corrected roll moment ΔMr is expressed by the following equation (21).
ΔMr = (ΔFfin−ΔFout) · Tf / 2 + (ΔFrin−ΔFrout) · Tr / 2 (21)

Assuming that the magnitudes of the corrected damping forces of all the shock absorbers are the same now, assuming that the value is ΔFs, and the corrected damping force of the shock absorber outside the turn is a positive value, the above formula 21 is expressed by the following formula 22 or 23: Is transformed as follows. Accordingly, the magnitude ΔFs of the correction damping force of the shock absorber is expressed by the following equation 24.
ΔMr = {(ΔFs) − (− ΔFs)} · Tf / 2
+ {(ΔFs) −− ΔFs} · Tr / 2 (22)
−ΔMr = {(− ΔFs) − (− ΔFs)} · Tf / 2
+ {(− ΔFs) −ΔFs} · Tr / 2 (23)
ΔFs = ΔMr / (Tf + Tr) (24)

  For example, when the corrected roll moment ΔMr is positive when the vehicle is turning left, for example, the expression 22 above indicates that the left and right front and rear wheels that are the turning inner wheel are further allowed to roll in the right direction on the roll angle increasing side. This is applied when a positive correction damping force ΔFs is required for each shock absorber, and a negative correction damping force (−ΔFs) is required for each of the shock absorbers of the right front and rear wheels that are turning outer wheels. Also, for example, when the correction roll moment ΔMr is positive when the vehicle is turning right, the shock absorbers of the right front and rear wheels that are turning inner wheels are further allowed to roll leftward, which is the increase side of the roll angle. This is also applied to the case where a positive correction damping force ΔFs is required for each of the left and right, and a negative correction damping force (−ΔFs) is required for each of the left and right front and rear shock absorbers.

  On the other hand, for example, when the corrected roll moment ΔMr is negative when the vehicle is turning left, the above-described Expression 23 is the left front and rear wheels that are the turning inner wheels so that the roll in the right direction on the roll angle increasing side is restricted. This is applied when a negative correction damping force (−ΔFs) is required for each of the shock absorbers, and a positive correction damping force ΔFs is required for each of the shock absorbers of the right front and rear wheels that are turning outer wheels. In addition, for example, when the corrected roll moment ΔMr is negative when the vehicle is turning right, the left and right shock absorbers that are the turning inner wheels are restricted so that the roll in the left direction on the roll angle increasing side is restricted. This is also applied to cases where a negative correction damping force (−ΔFs) is required for each, and a positive correction damping force ΔFs is required for each of the left and right shock absorbers that are turning outer wheels.

  Accordingly, in this first embodiment, the corrected damping force ΔFs of the shock absorber is calculated according to the above equation 24 in step 360 so that the magnitude of the corrected damping force of all the shock absorbers is the same.

  In step 430, the stroke speed Xid (i = fl, fr, rl, rr) is calculated by differentiating the stroke Xi of each wheel with respect to time, and the stroke speed Xid and the control stage Si of each shock absorber are calculated. Based on the above, the current damping force Fsi (i = fl, fr, rl, rr) of each shock absorber is calculated. Specifically, the damping force Fsi of each shock absorber that varies according to the stroke speed Xid is calculated with reference to the damping force map provided in the ROM of the electronic control unit 62 and corresponding to the graph shown in FIG. Is done.

  As shown in FIG. 11, the damping force map includes a shock absorber damping force Fsi that increases as the stroke speed Xid increases from “0” to a positive predetermined value for each of the plurality of control stages Si, and the stroke. The damping force Fsi of the shock absorber, which decreases as the speed Xid decreases from “0” to a negative predetermined value, is stored. Note that the magnitude of the damping force Fsi gradually increases from the soft (low damping force) side to the hard (high damping force) side as the throttle amount of the damping force generating valve increases, with the same stroke speed Xid.

In step 440, based on the current damping force Fsi of each shock absorber calculated in step 430 and the corrected damping force ΔFs calculated in step 360, the basic target of each shock absorber according to the following equation 25: A damping force Fsbi is calculated.
Fsbi = Fsi ± ΔFs (25)

In the above formula 25, when the corrected roll moment ΔMr is positive, the basic target damping force Fsbi on the turning outer wheel side is calculated as (Fsi + ΔFs), and the basic target damping force Fsbi on the turning inner wheel side is (Fsi−ΔFs). ). On the contrary, when the corrected roll moment ΔMr is negative, the basic target damping force Fsbi on the turning outer wheel side is calculated as (Fsi−ΔFs), and the basic target damping force Fsbi on the turning inner wheel side is calculated as (Fsi + ΔFs). Is done.
[First Modification]

  In this first variant embodiment, the damping coefficients of all shock absorbers are corrected equally. That is, when the magnitude of the modified attenuation coefficient is ΔC and the above equation 20 is replaced by the following equation 26, the modified attenuation coefficient ΔC is expressed by the following equation 27, and is calculated according to the following equation 27. Accordingly, the corrected damping forces ΔFsfin to ΔFsrout of each shock absorber are expressed by the following equations 28 to 31.

ΔMr = (ΔC · Xfind−ΔC · Xfoutd) · Tf / 2
+ (ΔC · Xrind−ΔC · Xroutd) · Tr / 2 (26)
ΔC = 2ΔMr / {(Xfind−Xfoutd) · Tf + (Xrind−Xroutd) · Tr} (27)
ΔFsfin = ΔC · Xfind (28)
ΔFsfout = ΔC · Xfoutd (29)
ΔFsrin
= ΔC ・ Xrind (30)
ΔFsrout
= ΔC ・ Xroutd (31)

  In the above equations, Xfind, Xfoutd, Xrind, and Xroutd representing the stroke speeds of the turning inner front wheel, the turning outer front wheel, the turning inner rear wheel side, and the turning outer rear wheel are the lateral acceleration Gy and steering of the vehicle. It is calculated based on the stroke speed Xid (i = fl, fr, rl, rr) of each wheel according to the turning direction of the vehicle determined based on the angle θ.

Thus, in this first modified embodiment, in step 360, the corrected damping forces ΔFsfin to ΔFsrout of the turning inner front wheel, the turning outer front wheel, the turning inner rear wheel side, and the turning outer rear wheel are determined according to the above equations 27 to 31. The corrected damping force ΔFsi (i = fl, fr, rl, rr) of each wheel is calculated based on these values. In step 430, the basic target damping force Fsbi of each shock absorber is calculated according to the following equation 32 corresponding to equation 25 above. The selection of the code in the following equation 32 is the same as in the case of the first embodiment described above. The other steps of the first modified embodiment are performed in the same manner as in the first embodiment described above.
Fsbi = Fsi ± ΔFsi (32)
[Second modified embodiment]

In this second modified embodiment, the damping coefficient of the shock absorber for the left and right wheels on the front wheel side and the rear wheel side so that the roll damping rate ζ on the front wheel side and the rear wheel side of the vehicle is the same. Will be corrected equally. When the roll damping rate ζ on the front wheel side and the rear wheel side of the vehicle is the same, the following Expression 33 is established.

  Here, ΔMrf and ΔMrr are correction roll moments on the front wheel side and the rear wheel side, respectively, and have a relationship of ΔMrf + ΔMrr = ΔMr. Irf and Irr are roll inertia moments on the front wheel side and the rear wheel side, respectively, and there is a relationship of Irf + Irr = Ir. Krf and Krr are roll rigidity on the front wheel side and the rear wheel side, respectively, and these roll rigidity changes depending on the operating conditions of the active stabilizer devices 16 and 18, but there is a relationship of Krf + Krr = Kr.

Assuming that the ratio of the corrected roll moment ΔMrf on the front wheel side to the corrected roll moment ΔMrr on the front wheel side is the corrected roll moment ratio λ, the corrected roll moment ratio λ is expressed by the following equation 34. Here, hf and hr are the roll arm lengths on the front wheel side and the rear wheel side, respectively, and mf and mr are the masses of the vehicle bodies on the front wheel side and the rear wheel side, respectively.

From the relationship of the above equation 34 and Mrf + ΔMrr = ΔMr, the corrected roll moment ΔMrf on the front wheel side is expressed by the following equation 35, and the corrected roll moment ΔMrr on the rear wheel side is expressed by the following equation 36.
ΔMrf = ΔMr · λ / (λ + 1)
...... (35)
ΔMrr = ΔMr / (λ + 1) (36)

Therefore, the corrected damping force ΔFsf of the shock absorber for the left and right wheels on the front wheel side is expressed by the following formula 37, and the corrected damping force of the shock absorber for the left and right wheels on the rear wheel side is expressed by the following formula 38.
ΔFsf = ΔMrf / Tf
= ΔMr · λ / {(λ + 1)
Tf} …… (37)
ΔFsr = ΔMrr / Tr
= ΔMr / {(λ + 1) Tr} (38)

  Thus, in the second modified embodiment, in step 360, the corrected damping force ΔFsf of the left and right front wheels is calculated according to the above equation 37, and the corrected damping force ΔFsr of the left and right rear wheels is calculated according to the above equation 38. In step 430, the basic target damping force Fsbi of each shock absorber is calculated according to the following equations 39 to 42 corresponding to the above equation 25. The selection of symbols in the following equations 39 to 42 is the same as in the case of the first embodiment described above. The other steps of the second modified embodiment are performed in the same manner as in the first embodiment described above.

Fsbfl = Fsfl ± ΔFsf (39)
Fsbfr = Fsfr ± ΔFsf (40)
Fsbrl = Fsrl ± ΔFsr (41)
Fsbrr = Fsrr ± ΔFsr (42)
[Third Modification]

In the third modified embodiment, as in the second modified embodiment, the front wheels according to the above equations 33 to 36 are set so that the roll damping rate ζ is the same on the front wheel side and the rear wheel side of the vehicle. The corrected roll moment ΔMrf on the side and the corrected roll moment ΔMrr on the rear wheel side are calculated. However, in the third modified embodiment, unlike the second modified embodiment, the damping coefficients of the left and right shock absorbers are equally modified on the front and rear wheels of the vehicle. Is done. That is, the corrected damping coefficient of the shock absorber for the left and right wheels on the front wheel side is ΔCf, the corrected damping coefficient of the shock absorber for the left and right wheels on the rear wheel side is ΔCr, and the above equation 21 is replaced by the following equations 43 and 44: It is done.
ΔMrf = ΔCf · (Xfind−Xfoutd) · Tf / 2 (43)
ΔMrr = ΔCr · (Xrind−Xroutd) · Tr / 2 (44)

  From the above equations 43 and 44, the left and right front wheel shock absorber modified damping coefficient ΔCf and the left and right rear wheel shock absorber modified damping coefficient ΔCr are expressed by the following equations 45 and 46, respectively. The forces ΔFsfin to ΔFsrout are expressed by the following equations 47 to 50.

ΔCf = 2ΔMrf / {(Xfind−Xfoutd) · Tf} (45)
ΔCr = 2ΔMrr / {(Xrind−Xroutd) · Tr} (46)
ΔFsfin = ΔCf · Xfind (47)
ΔFsfout = ΔCf · Xfoutd (48)
ΔFsrin = ΔCr · Xrind (49)
ΔFsrout = ΔCr · Xroutd (50)

Thus, in this third modified embodiment, in step 360, the turning inner front wheel, the turning outer front wheel, the turning inner rear wheel side, and the turning outer rear wheel are corrected in accordance with the expressions 35, 36 and 45-50. Damping forces ΔFsfin to ΔFsrout are calculated, and a corrected damping force ΔFsi (i = fl, fr, rl, rr) of each wheel is calculated based on these values. In step 430, the basic target damping force Fsbi of each shock absorber is calculated according to the above equation 32. The other steps of the third modified embodiment are executed in the same manner as in the first embodiment.
[Second Example]

  FIG. 8 is a flowchart showing the main part of the subroutine for calculating the basic target damping force Fsbi in the second embodiment, and FIG. 9 shows the subroutine for calculating the actual pitch angle θp of the vehicle body executed in step 370 of FIG. It is a flowchart to show. In FIG. 8, the same step numbers as those shown in FIG. 6 are assigned to the same steps as those shown in FIG.

  In the second embodiment, the control of the active stabilizer devices 16 and 18 by the electronic control device 22 is executed in the same manner as in the first embodiment described above according to the flowchart shown in FIG. The control of the damping force of the shock absorbers 44FL to 44RR by the electronic control unit 62 is basically executed in the same manner as in the first embodiment described above according to the flowcharts shown in FIGS. The pitching of the vehicle body is controlled in addition to the rolling control of the vehicle body so that the phase difference between the roll angle θrt and the pitching generated in the vehicle body becomes zero, which is shown in FIG. 8 instead of step 360 of FIG. Steps 370 to 420 in the flowchart are executed.

  In step 370 executed after step 350, the target pitch of the vehicle body is determined from the map corresponding to the graph shown in FIG. 13 based on the absolute value of the target roll angle θrt of the vehicle body calculated in step 310. The angle θpt is calculated. The calculation of the target pitch angle θpt is specifically performed with reference to a target pitch angle table provided in the ROM of the electronic control unit 62. As shown in FIG. 13, the target pitch angle table stores a target pitch angle θpt so that the posture of the vehicle body slightly leans forward when the vehicle turns, and the target pitch angle θpt is the target pitch angle θpt. It is uniquely determined by the roll angle θrt, and increases nonlinearly as the absolute value of the target roll angle θrt increases. The target pitch angle θpt may be calculated by a function having the target roll angle θrt as a variable.

  In step 380, the actual pitch angle θp of the vehicle body is calculated according to the flowchart shown in FIG.

In step 382 of the flowchart shown in FIG. 9, average values Gzf and Gzr of the vertical acceleration on the front-wheel and rear-wheel springs are calculated according to the following equations 51 and 52, respectively.
Gzf = (Gzfl + Gzfl) / 2 (51)
Gzr = (Gzrl + Gzrr) / 2 (52)

In step 384, the vehicle wheel base is set to L, and the pitch angular acceleration θpdd of the vehicle body is calculated according to the following equation 53. In step 386, the actual pitch of the vehicle body is obtained by second-order integration of the pitch angular acceleration θpdd. The angle θp is calculated.
θpdd = (Gzr−Gzf) / L (53)

Returning to the flowchart shown in FIG. 8, in step 390, the corrected pitch angle Δθp of the vehicle body is calculated as a deviation (θpt−θp) between the pitch angle θpt of the vehicle body and the actual pitch angle θp of the vehicle body. In this case, the corrected pitch angle acceleration Δθpdd (= d ² (Δθp) / dt ² ) is calculated by differentiating the corrected pitch angle Δθp of the vehicle body by the second order time. When the vehicle turns, the actual pitch angle θp is normally a positive value, that is, a value corresponding to the forward leaning posture of the vehicle body.

In step 410, the corrected pitch moment ΔMp necessary for correcting the pitch angle of the vehicle body is calculated according to the following equation 54, where Ip is the moment of inertia about the left-right axis of the vehicle and Kp is the pitch stiffness of the vehicle. The sign of the corrected pitch moment ΔMp is positive when the vehicle is leaning forward and negative when the vehicle is leaning backward.
ΔMp = Ip · Δθpdd + Kp · Δθp (54)

In this second embodiment, the corrected damping coefficient of the shock absorber for the left and right wheels on the front wheel side is ΔCf, and the corrected damping coefficient of the shock absorber for the left and right wheels on the rear wheel side is ΔCr, 21 is replaced by the following expression 55, and the above expression 54 is replaced by the following expression 56.
ΔMr = (ΔCf · Xfind−ΔCf · Xfoutd) · Tf / 2
+ (ΔCr · Xrind−ΔCr · Xroutd) · Tr / 2 (55)
ΔMp = − (ΔCr · Xfind + ΔCf · Xfoutd) · Lf
+ (ΔCr · Xrind + ΔCr · Xroutd) · Lr (56)

  Since the above formulas 55 and 56 are expressed by the determinant of the following formula 57, the correction damping coefficients ΔCf and ΔCr of the shock absorbers for the left and right front wheels and the left and right rear wheels of the vehicle are calculated according to the following formula 58.

  In order for the inverse matrix of Equation 58 to hold, each stroke speed must satisfy the conditions of Xfind ≠ Xfoutd and Xrind ≠ Xroutd. The reason is that when Xfind = Xfoutd and Xrind = Xroutd, no roll is generated in the vehicle body, so that the corrected damping coefficients ΔCf and ΔCr in the situation where the roll is generated cannot be calculated by the above equation 58. is there. Each stroke speed must also satisfy the conditions of Xfind ≠ −Xfoutd and Xrind ≠ −Xroutd. The reason is that, when Xfind = −Xfoutd and Xrind = −Xroutd, no pitching occurs in the vehicle body. Therefore, the corrected attenuation coefficients ΔCf and ΔCr in the situation where the pitching occurs in the above equation 37 are expressed by the above equation 58. This is because it cannot be calculated.

  Accordingly, in step 420, the shock absorbers of the left and right front wheels and the left and right rear wheels according to the above equation 58 based on the corrected roll moment ΔMr calculated in step 350 and the corrected pitch moment ΔMp calculated in step 410. Modified attenuation coefficients ΔCf and ΔCr are calculated.

  In step 430, the corrected damping forces ΔFsfin to ΔFsrout of the turning inner front wheel, the turning outer front wheel, the turning inner rear wheel side, and the turning outer rear wheel are calculated in accordance with the above formulas 45 to 50 in the same manner as in the third modified embodiment. Based on these values, the corrected damping force ΔFsi (i = fl, fr, rl, rr) of each wheel is calculated. In step 440, the basic target damping force Fsbi of each shock absorber according to the above equation 32. Is calculated.

  As will be understood from the above description, in the control of the active stabilizer device in each of the above-described embodiments and each of the modified embodiments, the target anti-roll moment is increased as the lateral acceleration Gy of the vehicle increases in step 20. The target anti-roll moment Mat is calculated based on the lateral acceleration Gy of the vehicle so as to increase the magnitude of Mat, and in steps 30 and 40, the front anti-roll moment is calculated based on the target anti-roll moment Mat and the front roll target roll stiffness distribution ratio Rmf. A target anti-roll moment Maft and a rear wheel target anti-roll moment Mart are calculated.

  When the magnitude of the target anti-roll moment Maft of the front wheel exceeds the maximum anti-roll moment Mafmax that can be generated by the active stabilizer device 16 on the front wheel side in steps 50 to 80, the target anti-roll moment Maft of the front wheel is increased. The magnitude is corrected to the magnitude of the maximum anti-roll moment Mafmax, and the excess anti-roll moment ΔMaf on the front wheel side is calculated as the difference between the target anti-roll moment Maft of the front wheels and the maximum anti-roll moment Mafmax.

  Similarly, when the magnitude of the rear wheel target anti-roll moment Mart exceeds the magnitude of the maximum anti-roll moment Marmax that can be generated by the rear wheel side active stabilizer device 18 in steps 90 to 120, the rear wheel target anti-roll moment Mart is exceeded. The magnitude of the anti-roll moment Mart is corrected to the magnitude of the maximum anti-roll moment Marmax, and the excess anti-roll moment ΔMar on the rear wheel side is calculated as a difference between the target anti-roll moment Mart of the rear wheel and the maximum anti-roll moment Marmax. Calculated.

  Further, at step 130, the target rotational angles φft and φrt of the actuators 20F and 20R of the active stabilizer devices 16 and 18 are calculated based on the front wheel target anti-roll moment Maft and the rear wheel target anti-roll moment Mart. Thus, the rotation angles φf and φr of the actuators 20F and 20R are controlled to be the target rotation angles φft and φrt, respectively, whereby the anti-roll moments due to the stabilizer forces on the front wheels and the rear wheels are respectively changed to the target anti-roll moments Maft of the front wheels. And the rear wheel target anti-roll moment Mart.

  In the control of the damping force of the shock absorber when the vehicle turns in each of the above-described embodiments and modified embodiments, the basic target damping force Fsbi of each of the shock absorbers 44FL to 44RR is calculated in step 220. In step 230, the required increase / decrease damping force Fsdf for the front wheels and the required increase / decrease damping force Fsdr for the rear wheels are values based on the excess roll moments ΔMaf and ΔMar as values required to achieve the excess roll moments ΔMaf and ΔMar in step 230, respectively. Calculated.

  In step 240, correction amounts ΔFsfl and ΔFsfr of the damping force of the left and right front wheels are calculated based on the required increase / decrease damping force Fsdf of the front wheels so that the damping force of the shock absorbers of the left and right front wheels does not exceed the maximum damping force. In step 270, the basic target damping forces Fsbfl and Fsbfr of the left and right front wheels are corrected by the damping force correction amounts ΔFsfl and ΔFsfr, respectively, so that the target damping forces Fstfr and Fstfl of the right front wheel and left and right front wheels are calculated. Thus, the damping forces Fsfl and Fsfr of the left front wheel are controlled to become the target damping forces Fstfl and Fstfr, respectively.

  Similarly, the correction amounts ΔFsrl and ΔFsrr for the left and right rear wheels are set based on the required increase / decrease damping force Fsdr for the rear wheels so that the damping force of the shock absorbers for the left and right rear wheels does not exceed the maximum damping force in step 280. In step 290, the basic target damping forces Fsbrl and Fsbrr of the left and right rear wheels are respectively corrected by the correction amounts ΔFsrl and ΔFsrr of the damping force, whereby the target damping forces Fstrfl and Fstrr of the left and right rear wheels are calculated. In step 300, the left and right rear wheel damping forces Fsrl and Fsrr are controlled to be the target damping forces Fstrl and Fstrr, respectively.

  In particular, in the calculation of the correction amounts ΔFsfl and ΔFsfr of the left and right front wheels at step 240, the maximum damping force Fsfcmax of the turning outer front wheel and the maximum damping force Fsfemax of the turning inner front wheel are calculated at steps 242 to 245. In addition, the excess damping forces Fsflo and Fsfro of the target damping forces Fstfl and Fstfr of the left and right front wheels with respect to the maximum damping forces Fsfcmax and Fsfemax are calculated.

  In steps 246, 247, and 252 it is determined whether or not the target damping forces Fstfl and Fstfr of the left and right front wheels exceed the corresponding maximum damping forces Fsfcmax and Fsfemax, respectively, and the target damping forces Fstfl and Fstfr of the left and right front wheels are determined. If none of the corresponding maximum damping forces Fsfcmax and Fsfemax are exceeded, the correction amounts ΔFsfl and ΔFsfr of the damping forces of the left and right front wheels are set to 0 in step 254.

  When either one of the target damping forces Fstfl and Fstfr for the left and right front wheels exceeds the corresponding maximum damping forces Fsfcmax and Fsfemax, the correction amounts ΔFsfl and ΔFsfr for the left and right front wheels are maximized in step 248 or 253, respectively. The excess amount of the target damping force with respect to the force is set to Fsflo or Fsfro.

  Further, when both of the target damping forces Fstfl and Fstfr of the left and right front wheels exceed the corresponding maximum damping forces Fsfcmax and Fsfemax, the correction amounts ΔFsfl and ΔFsfr of the left and right front wheels damping force with respect to the maximum damping force in steps 249 to 251. The larger value Fsflo or Fsfro of the excess amount of the target damping force is set. The above setting control is similarly performed for the left and right rear wheels.

  Therefore, according to each of the above-described embodiments and modified embodiments, the maximum anti-roll that the magnitudes of the target anti-roll moments Maft and Mart can be generated by the active stabilizer devices 16 and 18 in both the front wheels and the rear wheels. When the magnitudes of the moments Mafmax and Marfmax are exceeded, the target damping force of the shock absorbers 44FL to 44RR is corrected and the excess anti-roll moments ΔMaf and ΔMar are generated by the corrected damping force. Even when the magnitudes of the moments Maft and Mart exceed the maximum anti-roll moments Mafmax and Marfmax that can be generated by the active stabilizer devices 16 and 18, compared to the case where the damping force of the shock absorbers 44FL to 44RR is not corrected, The roll of the vehicle body can be effectively suppressed.

  Further, according to each of the above-described embodiments and modified embodiments, in any case of the front wheels and the rear wheels, the corrected target damping force of at least one of the left and right wheels is the maximum damping force of the shock absorber of the wheel. When exceeding, the wheel with the larger excess damping force is corrected for reduction by the excess damping force with the larger target damping force of the wheel, and the wheel target with respect to the left and right opposite wheels is corrected. Since the reduction force is corrected by the excess damping force with the larger damping force, the excess anti-roll moments ΔMaf and ΔMar are generated as effectively as possible without hindering the original damping force control of the shock absorber. Thus, as compared with the case where the target damping force is not corrected according to the present invention, the roll of the vehicle body is effectively and appropriately suppressed without hindering the original damping force control. It is possible.

  For example, as shown in FIG. 14, the target damping force of the original damping force control, that is, the basic target damping forces Fsbout and Fsbin, on both the outer turning wheel side (contraction side) and the inner turning side wheel side (extension side). When the value (Fsbout + Fsd and Fsbin−Fsd) corrected by the required increase / decrease damping force Fsd for generating the excess anti-roll moment does not exceed the maximum damping force, step 246 of the flowchart shown in FIG. And 252 are negatively determined, and correction amounts ΔFsfl and ΔFsfr are set to 0 in step 254. Therefore, in steps 260, 270, 290, and 300, the damping forces are the basic target damping forces Fsbout and Fsbin. It is controlled to be the same value as.

  On the other hand, as shown in FIG. 15, the difference between the basic target damping force Fsbin and the required increase / decrease damping force Fsd does not exceed the maximum damping force on the extension side on the turning inner ring side. When the sum of the basic target damping force Fsbout and the required increase / decrease damping force Fsd exceeds the contraction-side maximum damping force, an affirmative determination is made at step 246 of the flowchart shown in FIG. When a negative determination is made at step 247, or a negative determination is made at step 246 and an affirmative determination is made at step 252, the correction amount ΔFsfl and the correction amount ΔFsfl and ΔFsfr is set to the excessive damping force ΔFsout of the turning outer wheel (Fsfro when the vehicle is turning left, and Fsflo when the vehicle is turning right), and decreases at steps 260 and 290. For the turning outer wheel where the damping force exceeds, the target damping force of the wheel is reduced and corrected by the excess damping force ΔFsout. For the turning inner wheel, the target damping force of the wheel is increased and corrected by the excess damping force ΔFsout of the turning outer wheel. (The magnitude is reduced by ΔFsout).

  Accordingly, assuming that the direction of the anti-roll moment that generates the damping force is positive, the anti-roll moments Masf and Masr applied to the vehicle by the damping force on the front wheel side and the rear wheel side are expressed by the following equations 59 and 60, respectively. The excess roll moments ΔMaf and ΔMar can be reliably achieved by the damping force, thereby effectively suppressing the roll of the vehicle body.

Masf = (Fsbout + Fsd−ΔFsout) T / 2 + (− Fsbin + Fsd−ΔFsout) T / 2
= (Fsbout + Fsd) T / 2 + (-Fsbin + Fsd) T / 2 (59)
Masr = (Fsbout + Fsd−ΔFsout) T / 2 + (− Fsbin + Fsd−ΔFsout) T / 2
= (Fsbout + Fsd) T / 2 + (-Fsbin + Fsd) T / 2 (60)

Further, the sum Fstotal of the damping forces of the left and right wheels is a value expressed by the following formula 61 for both the front wheel side and the rear wheel side, and this value is the same as when the increase / decrease correction by the excess damping force ΔFsout is not performed. Is the value of Therefore, even when the increase / decrease correction by the excess damping force ΔFsout is performed, the original damping force control of the shock absorber can be performed without being affected by the required increase / decrease damping forces Fsdf and Fsdr.
Fstotal = (Fsbout + Fsd−ΔFsout) + (Fsbin−Fsd + ΔFsout)
= Fsbout + Fsbin (61)

  Further, as shown in FIG. 16, on the turning inner ring side, the difference between the basic target damping force Fsbin and the required increase / decrease damping force Fsd exceeds the maximum damping force on the expansion side, and on the turning outer ring side. When the sum of the basic target damping force Fsbout and the required increase / decrease damping force Fsd exceeds the contraction-side maximum damping force, an affirmative determination is made in steps 246 and 247 of the flowchart shown in FIG. In steps 249 to 251, the larger value of the excess damping forces Fsfro and Fsflo (the value on the normal turning outer wheel side, which is ΔFsmax in FIG. 16) is set as the correction amount ΔFsfr. Is done.

  Therefore, in this case, the target damping force of the wheel with the larger excess damping force (normal turning outer wheel) is controlled to the maximum damping force, and the wheel on the opposite side (normal turning inner wheel) has the maximum damping amount with the maximum damping force. The damping force is controlled to be smaller than the force, so that the excess anti-roll moments ΔMaf and ΔMar can be generated as effectively as possible without significantly impairing the original damping force control of the shock absorber.

  In particular, according to the first embodiment and the first to third modified embodiments described above, the basic target damping force Fsbi of each of the shock absorbers 44FL to 44RR calculated in step 220 is shown in FIGS. In order to reduce the corrected roll angle Δθr, which is a deviation between the target roll angle θrt of the vehicle body calculated based on the lateral acceleration Gy of the vehicle and the actual roll angle θ of the vehicle body, to 0 by being calculated according to the flowchart shown. Therefore, the basic target damping force Fsbi can be calculated so that the phase difference between the lateral acceleration Gy of the vehicle and the roll generated in the vehicle body becomes zero. Thereby, the damping force of the shock absorber can be controlled so that the roll angle of the vehicle body becomes an appropriate value and phase with respect to the magnitude and phase of the lateral acceleration Gy of the vehicle.

  Further, according to the first embodiment described above, the correction damping force ΔFs of each wheel is a common value for all the wheels, so that the correction of each wheel is performed as compared with the first to third modification examples. The damping force can be easily calculated. Further, according to the second modification example described above, the front wheel side correction damping force ΔFsf and the rear wheel side correction damping force ΔFsr are calculated as the correction damping force, which is compared with the case of the first embodiment described above. Thus, the damping force of the shock absorber on the front wheel side and the rear wheel side can be controlled more appropriately. Further, according to the first and third modification examples described above, the correction damping forces ΔFsfin, ΔFsfout, ΔFsrin, ΔFsrout of each wheel are calculated as the correction damping forces. As compared with the modified embodiment, the damping force of the shock absorber of each wheel can be controlled more appropriately.

  Further, according to the second embodiment described above, the basic target damping force Fsbi of each of the shock absorbers 44FL to 44RR calculated in step 220 is calculated according to the flowcharts shown in FIGS. Thus, the vehicle body calculated based on the corrected roll moment ΔMr for reducing the corrected roll angle Δθr, which is a deviation between the target roll angle θrt of the vehicle body and the actual roll angle θ of the vehicle body, to 0 and the target roll angle θrt of the vehicle body. Is calculated as a value necessary to generate both of the corrected pitch moment ΔMp for reducing the corrected pitch angle Δθp, which is a deviation between the target pitch angle θpt of the vehicle and the actual pitch angle θp of the vehicle body, to zero.

  Therefore, the basic target damping force Fsbi is calculated so that the phase difference between the lateral acceleration Gy of the vehicle and the roll generated on the vehicle body becomes zero, and the phase difference between the target roll angle θrt of the vehicle body and the pitch generated on the vehicle body becomes zero. Thus, the damping force of the shock absorber of each wheel can be controlled so that both the roll angle and the pitch angle of the vehicle body have appropriate values.

  Although the present invention has been described in detail with reference to specific embodiments, the present invention is not limited to the above-described embodiments, and various other embodiments are possible within the scope of the present invention. It will be apparent to those skilled in the art.

  For example, in each of the above-described embodiments and modified embodiments, the anti-roll moment generating means is an active stabilizer device that increases or decreases the stabilizer force, but between the sprung and unsprung positions at the left and right wheels. Any device known in the art may be used as long as the applied force can be increased or decreased, for example, an air suspension device, a hydropneumatic suspension device, or an active suspension device.

  In each of the above-described embodiments and modified embodiments, the front wheel required increase / decrease damping force Fsdf and the rear wheel required increase / decrease attenuation force Fsdr for achieving the excess roll moments ΔMaf and ΔMar are the same between the left and right wheels. However, the required increase / decrease damping force may be set to a different value between the left and right wheels so that the inner turning wheel is larger than the outer turning wheel, for example.

  Further, in each of the above-described embodiments and modified embodiments, the actual roll angle θr of the vehicle body is calculated based on the sprung acceleration Gzi at the position of each wheel, and in the above-described second embodiment. The actual pitch angle θp of the vehicle body is also calculated based on the sprung acceleration Gzi at the position of each wheel. The actual roll angle θr and the actual pitch angle θp of the vehicle body are, for example, the stroke of each wheel. It may be calculated based on Xi.

  In each of the above-described embodiments and modified embodiments, the shock absorbers of all the wheels have the same damping force characteristic (FIG. 11), but the damping force characteristics of the left and right wheel shock absorbers are the same. As long as the shock absorber for the front wheel and the shock absorber for the rear wheel may have different damping force characteristics, respectively.

  Further, in each of the above-described embodiments and modified embodiments, the target roll angle θrt is set to the same value for the lateral acceleration Gy of the vehicle regardless of whether the target roll angle θrt increases or decreases. However, for example, the target roll angle θrt is set to a different value with respect to the lateral acceleration Gy of the vehicle when increasing and when decreasing, and there is a slight hysteresis within a range in which a good roll feeling is ensured. It may be modified to exist.

1 is a schematic configuration diagram showing a first embodiment of a vehicle roll control device according to the present invention applied to a vehicle having an active stabilizer device on a front wheel side and a rear wheel side. FIG. FIG. 2 is an explanatory view showing a suspension of a right front wheel of the vehicle shown in FIG. 1. 3 is a flowchart showing a control routine of the active stabilizer device in the first embodiment. 3 is a flowchart showing a main routine of damping force control of the shock absorber in the first embodiment. FIG. 6 is a flowchart showing a subroutine for calculating correction amounts ΔFsfl and ΔFsfr of the damping force of the left and right front wheels, which is executed in step 240 of FIG. 4. FIG. 5 is a flowchart showing a subroutine of basic target damping force Fsbi calculation executed in step 220 of FIG. 4. FIG. It is a flowchart which shows the subroutine of the actual roll angle (theta) r calculation of the vehicle body performed in step 320 of FIG. It is a flowchart which shows the subroutine of the basic target damping force Fsbi calculation in a 2nd Example. FIG. 9 is a flowchart showing a subroutine for calculating the actual pitch angle θp of the vehicle body executed in step 370 of FIG. 8. It is a graph which shows the relationship between the lateral acceleration Gy of a vehicle, and the target anti-roll moment Mat. It is a graph which shows the relationship between the wheel stroke speed Xdi, the control stage Sai of a shock absorber, and the damping force Fsi of a shock absorber. It is a graph which shows the relationship between the lateral acceleration Gy of a vehicle, and the target roll angle (theta) rt of a vehicle body. 5 is a graph showing a relationship between an absolute value of a target roll angle θrt of a vehicle body and a target pitch angle θpt of the vehicle body. In each of the turning outer ring side (contraction side) and turning inner ring side (extension side), each damping is performed when the sum of the basic target damping forces Fsbout and Fsbin and the required increase / decrease damping force Fsd does not exceed the maximum damping force. It is explanatory drawing which shows force. The sum of the basic target damping force Fsbin and the required increase / decrease damping force Fsd does not exceed the maximum damping force on the inner turning wheel side, but the basic target damping force Fsbout and the required increase / decrease damping force Fsd on the outer turning wheel side It is explanatory drawing which shows each damping force before correction | amendment (A) by this invention, and (B) after correction | amendment about the case where the sum of is exceeding the maximum damping force. The sum of the basic target damping force Fsbin and the required increase / decrease damping force Fsd exceeds the maximum damping force on the inner turning wheel side, and the basic target damping force Fsbout and the required increase / decrease damping force Fsd on the outer turning wheel side. It is explanatory drawing which shows each damping force before correction | amendment (A) by this invention, and (B) after correction | amendment about the case where the sum exceeds the maximum damping force.

Explanation of symbols

  16, 18 ... Active stabilizer device, 20F, 20R ... Actuator, 22 ... Electronic control device, 44FL-44RR ... Shock absorber, 48C ... Actuator, 62 ... Electronic control device, 64 ... Lateral acceleration sensor, 66 ... Vehicle speed sensor, 68F, 68R: rotation angle sensor, 70: longitudinal acceleration sensor, 72: steering angle sensor, 74i: stroke sensor, 76i: vertical acceleration sensor

Claims (6)

  A damping force variable-type damping force generating means disposed between the sprung and unsprung portions corresponding to each wheel, and the damping force for the damping force to bring the vehicle posture to the target posture when the vehicle turns. Damping force control means for controlling the damping force generation means to become force, and anti-roll moment generation means for increasing or decreasing the anti-roll moment by increasing or decreasing the force acting between the sprung and unsprung An anti-roll moment control means for controlling the anti-roll moment generating means so that the anti-roll moment becomes a target anti-roll moment for reducing the roll of the vehicle when the vehicle turns. When the target anti-roll moment exceeds the maximum anti-roll moment of the anti-roll moment generating means, the vehicle When the target damping force of the left and right wheels is corrected to reduce the roll, and the corrected target damping force of at least one of the left and right wheels exceeds the maximum damping force of the damping force generating means of the wheel, the excess amount For a wheel having a larger damping force, the target damping force of the wheel is corrected by reducing the larger excess damping force, and for a wheel on the opposite side, the target damping force of the wheel is increased. A roll control device for a vehicle, wherein the vehicle is controlled to be reduced by the larger excess damping force.   The roll control device corrects the target damping force so that when the target anti-roll moment exceeds the maximum anti-roll moment of the anti-roll moment generating means, an excess anti-roll moment is generated by an increase in damping force. The roll control apparatus for a vehicle according to claim 1.   The roll control device has the upward direction as the positive direction of the damping force, and increases and corrects the target damping force of the outer turning wheel with the same amount of correction, and reduces and corrects the target damping force of the inner turning wheel. The roll control apparatus for a vehicle according to claim 1 or 2, characterized by the above-mentioned.   The damping force generating means is a shock absorber capable of increasing or decreasing a damping coefficient, and the damping force control means controls the damping coefficient of the shock absorber so that the damping coefficient becomes a target damping coefficient corresponding to the target damping force, The roll control device estimates the corrected target damping force based on the target damping coefficient corresponding to the corrected target damping force and the relative vertical speeds of the sprung and unsprung springs, and determines a maximum damping coefficient. The vehicle roll control device according to any one of claims 1 to 3, wherein a maximum damping force of the damping force generating means is estimated based on a relative vertical speed of the sprung and unsprung portions.   The vehicle roll control device according to any one of claims 1 to 4, wherein the anti-roll moment generating means is an active stabilizer device capable of increasing or decreasing a stabilizer force.   The vehicle roll control device according to claim 2, wherein the roll control device calculates a correction amount of the target damping force based on the excess anti-roll moment.

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