JP4806930B2 - Vehicle steering system - Google Patents

Description

  The present invention relates to a vehicle steering apparatus, and more particularly to a vehicle steering apparatus having excellent steering response.

Conventionally, as a steering device capable of independently controlling the steering angles of the left and right wheels, for example, a device described in Patent Document 1 is known. The steering device described in Patent Document 1 variably controls the steering angle ratio of the left and right front wheels, the means for turning the rear wheels in the same phase or in the opposite phase with respect to the front wheels, and the variable control of the steering angle ratio of the rear left and right wheels. Means. As a result, the Ackerman steering characteristic, in which the steering angle of the turning inner wheel is larger than that of the turning outer wheel, is changed to the parallel steering characteristic in which the turning inner and outer wheels have the same steering angle.
JP 9-39827 A

However, only by using the parallel steering characteristic, it is possible to solve the problem that the lateral force average generated in the left and right wheels as cornering force during turning is saturated in a relatively low lateral acceleration region (for example, around 0.6 G). could not. When the above-mentioned saturation state is reached, if the lateral acceleration increases further, even if the steering angle is increased and the tire side slip angle is increased, the lateral force average does not increase, the effectiveness of the rudder decreases, and the vehicle response Decreases. This is because the wheel load of the turning inner wheel decreases as the lateral acceleration increases, and if the lateral force generated on the turning inner wheel exceeds the specified lateral acceleration, it tends to decrease even if the side slip angle is increased. When the decrease cannot be compensated by the lateral force generated by the turning outer wheel, the saturation state occurs.
About the above-mentioned problem, it is easy to receive the influence, so that a steering characteristic approaches the Ackerman steering characteristic side.
The present invention has been made in view of the above-described problems, and provides a vehicle steering apparatus that can increase the turning limit.

In order to solve the above-described problems, a vehicle steering apparatus according to the present invention includes steering angle control means that can independently adjust the actual steering angle of a pair of left and right steering wheels that are steered in response to a steering input from a steering wheel. In the vehicle steering apparatus provided, the steering angle control means makes the actual steering angle of the outer turning wheel larger than the actual steering angle of the inner turning wheel when the vehicle is turning.
At this time, it is preferable that the caster rail has a size capable of suppressing the transmission of torque to the steering wheel due to the self-aligning torque acting in the direction of increasing the side slip angle of the steering wheel. In addition, when a power steering device for adding auxiliary torque to the steering system is provided, the auxiliary torque is caused by self-aligning torque that acts in the direction of increasing the side slip angle of the steered wheel when the vehicle turns. The torque transmitted to the steering wheel is preferably a torque that cancels out the torque.

  According to the present invention, the turning limit is improved and the steering response is excellent even when the lateral acceleration is high.

Next, embodiments of the present invention will be described with reference to the drawings.
(First embodiment)
FIG. 1 shows a schematic configuration of the vehicle steering apparatus according to the first embodiment.
A front suspension member 9 is attached to the front portion of the vehicle 1 in FIG. 1, and a front right wheel 2 and a front left wheel 3 are attached to the front suspension member 9 via suspension links 20. The front suspension member 9 is mounted with a steering gear box 7 connected to the steering wheel 10 and actuators 6 and 8 connected to the left and right of the steering gear box 7.

The configuration of the steering gear box 7 and the actuators 6 and 8 will be described with reference to FIG.
The steering input from the steering wheel 20 is transmitted to the rack and pinion rack 21 via the steering shaft 10a. Thereby, the rack 21 is displaced left and right by an amount corresponding to the steering angle.

At both ends of the rack 21, actuators 6 and 8 that can be expanded and contracted are provided coaxially with the rack 21. And the end part of the actuator rod 22 which is a movable part of the actuators 6 and 8 is connected to the tie rod, and further connected to the front left wheel via the knuckle arm. Accordingly, the tie rod and the knuckle arm swing according to the amount of displacement of the rack 21 and the actuator 8, and the front wheel connected to one end of the knuckle arm rotates to be steered.
The actuators 6 and 8 are driven by hydraulic pressure from the hydraulic circuit 11, and the hydraulic pressure from the hydraulic circuit 11 is controlled according to a command value from the control unit 16. The hydraulic circuit 11 includes various hydraulic sources and a hydraulic control valve.

The operation of the actuator 8 provided on the left wheel side will be described with reference to FIG. FIG. 2A shows a state in which the steering angle δ _H input from the steering wheel 10 is zero (0) and the command value from the control unit 16 is also zero (0). In this state, the displacement of the rack 21 is zero. At (0), the actuator 6 does not operate either. Accordingly, since the total amount of displacement of the rack 21 and the actuator 8 to the left is zero (0), the actual steering angle of the front left wheel is also zero (0), and the vehicle is in a straight traveling state. FIG. 2B shows a state in which the steering angle δ _H is input but there is no command for adjusting the actual steering angle from the control unit 16, and the displacement amounts of the rack 21 and the actuator 8 are determined according to the steering angle. Displacement amount ΔL2 (displaced to the right in the figure). Accordingly, the front left wheel is steered left, for example, as it is according to the steering angle. FIG. 2C shows a state in which the steering angle δ _H is input and an instruction for adjusting the actual steering angle is given from the control unit 16, and the displacement amounts of the rack 21 and the actuator 8 are the total displacement to the left. This is the sum of the amount of displacement ΔL2 of the rack 21 and the amount of displacement ΔL1 of the actuator 8. Therefore, the actual steering angle of the front left wheel is greatly adjusted by the command value from the control unit 16 in addition to the steering input.

Although not shown, the right actuator 6 is arranged symmetrically with the actuator 6. In this case, the displacement of the rack 21 based on the steering angle δ _H input is common to both the left end and the right end, but the pair of actuators 6 and 8 are independently controlled.
The pair of actuators 6 and 8 described above constitute the steering angle control means of the present invention.

Further, the vehicle 1 includes a vehicle speed sensor 14 that detects a vehicle speed V as a vehicle state quantity detection unit that detects a vehicle state quantity, a lateral acceleration sensor 13 that detects a lateral acceleration Ay generated in the vehicle 1 during a turn, A yaw rate sensor 12 for detecting the yaw rate γ generated in the vehicle 1 during turning and a steering angle sensor 15 as a steering angle detecting means for detecting the steering angle δ _H of the steering wheel 10 are provided, and these detection signals are controlled. Input to unit 16.

  Based on the vehicle state quantity and the steering angle, the control unit 16 moves the actuators 6 and 8 so that the actual steering angle of the outer turning wheel becomes larger than the actual steering angle of the inner turning wheel during vehicle turning. The command value is calculated and output to the hydraulic circuit 11. The control unit 16 includes an arithmetic processing device such as a microcomputer.

In the configuration as described above, the actuators 6 and 8 are not limited to the hydraulic type, and an electric motor type may be used. In this case, the hydraulic circuit 11 serves as an electric actuator driver.
In the above configuration, the steering input from the steering wheel 10 is input to the steering gear box 7 by mechanical joining, but the application of the present invention is not limited to such a configuration. For example, the steering gear box 7 is not provided, and the steering angle from the steering wheel 10 is input to the control unit 16 to calculate an appropriate actual steering angle, and the actuators 6 and 8 are operated based on this to operate the left and right front wheels. The actual rudder angle may be adjusted.

Next, the detailed configuration of the control unit 16 and the calculation processing of the actual steering angle will be described according to the block chart shown in FIG.
As shown in FIG. 3, the control unit 16 includes a lateral force sum estimating unit 16 a that is a lateral force sum estimating unit of the present invention, each wheel lateral force calculating unit 16 b that is each wheel lateral force calculating unit of the present invention, Each functional block includes a wheel load estimation unit 16c that is a wheel load detection unit of the present invention, a maximum lateral force estimation unit 16d that is a maximum lateral force estimation unit of the present invention, and an actual wheel steering angle calculation unit 16e. , Real steering angle calculation processing is realized. These are configured as, for example, one or more subroutine programs, and a series of processing by these programs is executed by timer interrupt processing at predetermined time intervals (for example, every 10 msec).

In the lateral force sum estimating unit 16a, based on the yaw rate γ, lateral acceleration Ay, vehicle speed V, and steering angle δ _H detected by the sensors, the total lateral force that can be generated by the front and right tires at that moment, that is, The lateral force sum is estimated, and the result is output to each wheel lateral force calculation unit 16b. Specifically, the yaw rate γ, the lateral acceleration Ay, the vehicle speed V, and the steering angle δ _H detected by the sensors are first input from the sensors. Based on this input, for example, the steering angle [delta] _H is determined whether the vehicle is turning by determining whether the predetermined or more, when the vehicle is turning, based on the vehicle state quantity and the steering angle To estimate the lateral force sum.

At this time, the vehicle is replaced with a two-wheel model including a front wheel and a rear wheel (hereinafter also referred to as a “two-wheel model”). Then, it calculates the front wheel lateral force average Fy _f based on the front wheel lateral force average Fy _f in equation (1) below.
Fy _f = K _f (Ay) · {N · δ _H −β− (l _f / V) · γ} (1)
Where K _f (Ay) is the front wheel cornering power as a function of lateral acceleration, N is the gear ratio of the steering gear box, β is the vehicle side slip angle, and l _f is from the front wheel in the two-wheel model. This is the distance to the center of gravity of the vehicle.

In the above equation (1), the side slip angle of the front wheel in the two-wheel model is calculated based on the vehicle state quantity and the steering angle that change from moment to moment by the portion {Nδ _H −β− (l _f / V) γ}. The front wheel average lateral force Fy _f can be estimated from the product of the side slip angle of the front wheel in the two-wheel model and K _f (Ay) indicating the tire characteristics.
The K _f (Ay) is a model of the front wheel cornering power stored in advance in the memory of the control unit 16 as a function of the lateral acceleration, and is calculated as the average cornering power of the left and right wheels when the present invention is applied. Is a characteristic value. FIG. 4 shows the lateral force average of the left and right wheels when the present invention is applied (thick line a in the figure). This is the average of the lateral force of the outer turning wheel and the lateral force of the inner turning wheel obtained at each lateral acceleration when the present invention is applied. In the figure, the lateral force of the outer turning wheel and the turning force at each lateral acceleration are obtained. The lateral force of the inner ring is shown connected by a solid line b. As shown by the solid line c in FIG. 5, the inclination of the lateral force average at each lateral acceleration corresponds to the front wheel cornering power.

The vehicle side slip angle β is a side slip angle of the center of gravity of the vehicle in the two-wheel model, and is obtained by a well-known calculation method obtained from a vehicle equation of motion based on the yaw rate γ, the vehicle speed V, and the steering angle δ _H.
The wheel load estimator 16c receives the lateral acceleration Ay from the lateral acceleration sensor 13, calculates the wheel loads of the left and right front wheels based on the input, and outputs the result to the maximum lateral force estimator 16d. The calculation of the wheel load is performed using a primary expression of a lateral acceleration multiple as shown in FIG. 6 derived from the balance of moments during turning. In FIG. 6, W is the vehicle weight, W _out shows the before turning outer wheel load, W _in the front turning inner wheel load.

In the maximum lateral force estimating unit 16d, the front inner ring maximum lateral force (Fy _fin ) _max that is the maximum lateral force that can be generated by the left and right front wheels based on the wheel loads of the left and right front wheels input from the wheel load estimating unit 16c. And the front outer wheel maximum lateral force (Fy _fout ) _max is calculated, and the result is output to each wheel lateral force calculation unit 16b. The estimation of the maximum lateral force is performed by obtaining the maximum lateral force within the allowable range of the tire side slip angle at the wheel load from the relationship between the tire side force and the side slip angle as shown in FIG. The maximum lateral force at each wheel load estimated in this way is shown in FIG. 7 (points in the figure). Thus, the maximum lateral force increases as the wheel load increases. In the memory of the control unit 16, information on the tire characteristic values of the left and right front wheels is stored in advance for calculating the maximum lateral force. The tire characteristic value may be obtained from an approximate mathematical model called a magic formula or from a map. In this embodiment, the tire characteristic value is determined only by the wheel load. However, a more complicated model is provided, and the camber angle, the slip ratio, the road surface friction coefficient, etc. are measured or estimated to estimate the maximum lateral force. The effect of the present invention can be enhanced by using for the above.

In each wheel lateral force calculation unit 16b, the front wheel average lateral force Fy _f input from the lateral force wheel estimation unit 16a and the front inner ring maximum lateral force (Fy _fin ) _max input from the maximum lateral force specifying unit 16d. The front inner wheel lateral force Fy _fin and the front outer wheel lateral force Fy _fout , which are tire lateral forces to be generated by the front wheels, are calculated based on the front outer wheel maximum lateral force (Fy _fout ) _max . And the result is output to each wheel steering angle calculating part 16e. The calculation of the tire lateral force that each front wheel should generate is performed using the following equations.

The relationship between the front inner ring lateral force Fy _fin and the front outer ring lateral force Fy _out is expressed by equation (2).
2Fy _f = Fy _fin + Fy _out (2)
In the present embodiment, the relationship between the front inner ring maximum lateral force (Fy _fin ) _max and the front outer ring maximum lateral force (Fy _fout ) _max is defined as in the following equation (3).
Fy _fin Fy _fout
――――――――― = ――――――――― (3)
(Fy _fin ) _max (Fy _fout ) _max

  This expression (3) is an expression for determining the ratio of the sum of lateral forces to be generated to be borne by the turning inner and outer wheels. By defining as in Expression (3), the ratio of the lateral force generated by each tire to the maximum lateral force that can be generated by each tire in the turning inner and outer wheels can be made the same in the turning inner and outer wheels. As a result, even if the turning lateral acceleration increases, the load that both wheels should generate lateral force becomes even, and it is possible to efficiently use the lateral force that can generate the left and right tires up to the limit acceleration.

From the expressions (2) and (3), the front inner ring lateral force Fy _fin and the front outer ring lateral force Fy _fout are
2Fy _fin
Fy _fin = ―――――――――――――――――― Fy _f (4)
(Fy _fin ) _max + (Fy _fout ) _max
2Fy _fout
Fy _fout = ―――――――――――――――――― Fy _f (5)
(Fy _fin ) _max + (Fy _fout ) _max
It is possible to determine how to distribute the front wheel lateral force sum calculated by the lateral force wheel estimator 16a to the left and right.

In each wheel actual steering angle calculation unit 16e, based on the front inner wheel lateral force Fy _fin and the front outer wheel lateral force Fy _fout input from each wheel lateral force calculation unit 16b, the adjustment steering angles Δδ _in and Δδ _out by the actuator are calculated. The target actuator position for the calculation is output to the hydraulic circuit 11.
Specifically, first, tire side slip angles α _in and α _out that generate front inner wheel lateral force Fy _fin and front outer wheel lateral force Fy _fout are obtained. That is, the front inner wheel lateral force Fy _fin and the front outer wheel lateral force Fy _fout are calculated with respect to the wheel loads of the front left and right wheels calculated by the wheel load estimating unit 16c based on the tire characteristic values similar to those used in the wheel load estimating unit 16c. The generated tire side slip angles α _in and α _out are specified.

Next, based on the calculated tire side slip angles α _in and α _out , the actual steering angle δ _in of the turning inner wheel and the actual steering angle δ _out of the turning outer wheel are calculated using the following formula (6) or (7).
δ _in = 1 / N {α _in −β− (l _f / V) · γ} (6)
δ _out = 1 / N {α _out −β− (l _f / V) · γ} (7)
Further, the difference between the actual steering angles δ _in and δ _out and the steering angle adjusted by the steering input from the steering wheel 10 is calculated, and the adjustment steering angles Δδ _in and Δδ _out by the actuator are calculated.

Here, the relationship between the calculated tire side slip angles α _in and α _out and the front inner wheel lateral force Fy _fin and the front outer wheel lateral force Fy _fout is shown in FIG. When the tire side slip angles α _in and α _out are calculated such that the ratio of the lateral force to the maximum lateral force is the same between the turning inner wheel and the turning outer wheel, the tire side slip angle α _out of the turning outer wheel becomes larger. Therefore, the actual rudder angles δ _in and δ _out are also larger than the actual rudder angle δ _out of the outer turning wheel.
Then, a target actuator position for adding the adjustment steering angles Δδ _in and Δδ _out by the actuator to a predetermined wheel is calculated.
The above is the calculation processing of the actual steering angle by the control unit 16.

Next, the operation of the first embodiment will be described.
When a steering input is made to the steering wheel 10 while the vehicle is running, the rack 21 is displaced to the right, for example, and the left and right front wheels are steered to the left. In Con DOO roll unit 16, detects this pivoting, on the basis of the input vehicle state quantity and the steering angle, it performs arithmetic processing of the actual steering angle. As a result, the front right wheel 2 having a difference between the calculated actual steering angle δ _out and the current steering angle is adjusted to the calculated actual steering angle δ _out by the operation of the actuator 6. On the other hand, since there is no difference between the actual rudder angle δ _in and the current rudder angle, the front left wheel 3 is not instructed to adjust the actual rudder angle from the control unit 16, and the actuator 8 does not operate. Accordingly, as shown in FIG. 9, the vehicle turns in a state where the actual steering angle of the front right wheel 2 that is the outer turning wheel is adjusted to be larger than the actual steering angle of the front left wheel 3 that is the inner turning wheel.

Next, the effect of the present invention will be described.
FIG. 5 (a) is a graph showing the relationship between the tire lateral force average, the tire side slip angle and the lateral acceleration generated in the steered wheels. The case where the present invention is applied is a solid line a, and the conventional parallel steering method is adopted. Is indicated by a dotted line d. As is apparent from the figure, when the present invention is applied, it does not saturate until the lateral acceleration increases compared to the prior art, and for example, in the prior art, the tire lateral force average is higher than 0.6 G which starts to saturate. Even in the lateral acceleration region, it is confirmed that the slope of the tire lateral force average is still large. FIG. 5 (b) is a graph showing the tire lateral force average slope at a lateral acceleration exceeding 0.6G (indicated by a one-dot chain line f) in order to clarify this difference. The inclination of the tire lateral force average corresponds to the cornering power. Therefore, it can be seen that when the present invention is applied, the vehicle exhibits excellent steering response.

The difference in the effect is as follows.
In general, in the relationship between the tire lateral force and the tire side slip angle, the tire side force generated as the wheel load increases, and the tire side slip angle generating the maximum side force also increases. Conversely, when the wheel load is small, the maximum lateral force is reached (saturates) at a small tire side slip angle. The present invention utilizes this, and as shown in FIG. 4, the tire side slip angle of the turning outer wheel where the wheel load increases due to load movement as the lateral acceleration increases is increased, and the turning of the lateral force to be generated on both wheels is turned. By increasing the ratio of the lateral force generated by the outer ring, cornering force can be generated corresponding to higher lateral acceleration. Furthermore, the lateral force is distributed so that the ratio of the lateral force generated at the wheel to the maximum lateral force that can be generated at the left and right steering wheels is equal or substantially equal between the left and right steering wheels. As the acceleration increases, the ratio of the lateral force generated in the turning outer wheel can be increased, and the saturation of the turning inner wheel can be delayed. Thereby, the effect of this invention is remarkable in the area | region where a lateral acceleration is high.

  On the other hand, as shown in FIG. 15, the conventional parallel steering method is a method of turning with the same turning angle of the turning outer wheel and the turning inner wheel, and the tire side slip angle is the same for the turning inner and outer wheels. In this case, the lateral slip angle of the turning inner and outer wheels is increased by the same amount so as to generate a corresponding lateral force as the lateral acceleration increases. For this reason, when the lateral acceleration increases, the lateral force of the turning inner wheel is saturated, and the lateral force of the turning inner wheel decreases even if the side slip angle is increased to increase the lateral force. Therefore, it can be seen that the lateral force average of both wheels is saturated (see FIG. 16). In addition, the Ackerman steering method that has been conventionally employed is a method in which the actual steering angle of the turning inner wheel is set larger than that of the turning outer wheel, as shown in FIG. Although not shown, as is apparent from the above reasons, this method is more susceptible to the problem of lateral force saturation in the low lateral acceleration region.

(Second embodiment)
Next, a second embodiment of the present invention will be described.
As shown in FIG. 10, the second embodiment has substantially the same configuration as the first embodiment, but differs in that the left and right rear wheels 5 and 4 can also determine the actual steering angle independently on the left and right. . A description of the same parts as in the first embodiment will be omitted, and different parts will be described.
The left and right rear wheels 5 and 4 are attached to the rear suspension member 17 via the suspension link 20. The rear suspension member 17 has an actuator 18 for adjusting the actual steering angle of the rear right wheel 4 and the actual steering angle of the rear left wheel 5. An actuator 19 for adjusting the angle is attached. The actuators 19 and 18 are connected to the left and right rear wheels 5 and 4 via tie rods and knuckle arms, respectively, and are driven by the hydraulic pressure from the same hydraulic circuit 11 as the left and right front wheels 3 and 2, respectively. Steer 2 respectively. For example, when the left rear actuator 19 extends, the rear left wheel 5 is steered to the right, and conversely, when it contracts, the rear left wheel 5 is steered to the right. The hydraulic circuit 11 is controlled in accordance with a command value from the same control unit 16 as the left and right front wheels 3 and 3.

The control unit 16 has the same configuration as that of the first embodiment, but the rear wheel is also generated in each part by the sum of the lateral forces of the rear wheels, the maximum lateral force that can be generated by the left and right rear wheels, and the left and right rear wheels. The difference is that the lateral force and the actual steering angle of the left and right rear wheels are determined. Hereinafter, different points will be described.
In the lateral force sum estimating unit 16a, the lateral force sum Fy _r of the rear wheels is replaced by a two-wheel model as described above, is calculated using the following equation (8).
Fy _r = K _r (Ay) {δ _r −β− (l _r / V) · γ} (8)
Here, K _r (Ay) is the rear wheel cornering power (a function of lateral acceleration), δ _r is the rear wheel average actual steering angle, β is the vehicle side slip angle estimated from the vehicle state quantity, and l _r is in the two-wheel model. This is the distance from the rear wheel to the center of gravity of the vehicle.

The rear wheel average steering angle δ _r is determined by vehicle state quantities such as the steering angle δ _H , the lateral acceleration Ay, the yaw rate γ, the vehicle side slip angle β, and the vehicle speed V as conventional techniques.
δ _r = δ _r (δ _H , Ay, γ, β, V) (9)
The wheel load estimator 16c and the maximum lateral force estimator 16d calculate the wheel load and the maximum lateral force in the same manner as the front wheels, and each wheel lateral force calculator 16b calculates the equations (2) to (5). By applying to the rear wheel, the rear wheel inner wheel lateral force Fy _rin and the rear wheel outer wheel lateral force Fy _rout are calculated.

In each wheel actual steering angle calculation unit 16e, based on the inputs of the rear wheel inner wheel lateral force Fy _rin and the rear wheel outer wheel lateral force Fy _rout , the tires of the rear turning inner and outer wheels using the tire characteristic values are used in the same manner as the front wheels. The sideslip angles α _in and α _out are calculated. Then, by using the tire side slip angles α _in and α _out of the rear turning inner and outer wheels, the actual steering angles δ ′ _in and δ ′ _out of the rear turning inner and outer wheels are obtained by the following equations (10) and (11).
δ ′ _in = 1 / N {α ′ _in −β− (l _r / V) · γ} (10)
δ ′ _out = 1 / N {α ′ _out −β− (l _r / V) · γ} (11)

Maximum lateral force as the ratio of the lateral force is the same relationship turning inner and outer turning wheel for, the actual steering angle Deruta' _in the rear turning inner and outer rings, the direction of the actual steering angle after turning outer calculating the Deruta' _out growing.
FIG. 11 shows a state in which the present invention is applied to a rear-wheel in-phase steered vehicle (FIG. 11A) and a rear-wheel reverse-phase steered vehicle (FIG. 11B) that are generally used in rear-wheel steered vehicles. Show. In any case, the actual rudder of the turning outer wheel is larger. As described above, when steering the front wheels and the rear wheels, the effect of both-wheel steering can be enhanced by increasing the actual steering angle of the outer turning wheel to the actual steering angle of the inner turning wheel for both the front and rear wheels. For this reason, it is possible to reduce the steering angle of one wheel, for example, the rear wheel, by the amount improved by the steering of both wheels, and the unit can be reduced in weight and size.

(Third embodiment)
Finally, a third embodiment of the present invention will be described.
In the third embodiment, a negative self-aligning torque (hereinafter referred to as “SAT”) is further transmitted to the steering wheel 10 in a vehicle having the same configuration as the first embodiment or the second embodiment. Means for preventing are provided.
Here, SAT has characteristics as shown in FIG. Positive SAT works in a direction to reduce the tire side slip angle when the tire side slip angle increases (see FIG. 12B). Therefore, if no force is applied at that moment, the steering wheel is returned to the original position. There are characteristics. However, when a tire side slip angle exceeds a certain size, the sign becomes negative. Since the negative SAT works in a direction to increase the tire side slip angle as shown in FIG. 12, if the force is not applied at that moment, the steering wheel does not return to the original position. It will increase. In addition, changing the direction of the SAT makes it difficult for the driver to match the desired steering angle.

FIG. 13 shows an example of a configuration that prevents negative SAT from being transmitted as steering torque due to the geometry of the suspension.
FIG. 13 is a view of the front right wheel 2 as viewed from the lateral direction of the vehicle. A caster that is a vertical distance between a point at which the kingpin shaft that is a turning shaft intersects the ground and a tire ground contact point as viewed from the lateral direction of the vehicle. The trail is sized to suppress the generation of steering torque that acts in the direction of increasing the side slip angle of the front wheels due to negative self-aligning torque. Thereby, the controllability of the steering angle is improved.

FIG. 14 shows another example for preventing the negative SAT from being transmitted to the steering wheel.
The vehicle shown in FIG. 14 is equipped with an electric power steering device 23 that applies auxiliary torque to the steering shaft 10a. Further, a torque sensor 24 is provided closer to the steering gear box 7 than the portion to which the auxiliary torque of the steering shaft 10a is applied, and the torque at the portion is detected. When the torque sensor 24 detects the torque _TF , it outputs the value to the control unit 25. Control unit 25, on the basis of the torque T _F detected by the torque sensor 24, in the torque in the opposite direction, and a command to the electric power steering apparatus 23 so as to impart the same or more auxiliary torque T _P Output.

Accordingly, the torque acts in a direction to increase the slip angle due to the negative SAT by assist torque T _P is canceled, the torque T _H, which is transmitted to the steering wheel acts in a direction to return the steering. With such a configuration, there is no restriction to adjust the handle torque by the suspension geometry, and there is an advantage that the degree of freedom in suspension design is increased.
Although the electric power steering apparatus has been described as an example, the power steering apparatus is not limited to the electric type, and may be, for example, a hydraulic type.

It is a figure showing the schematic structure of the steering device for vehicles concerning a first embodiment. It is a figure explaining operation | movement of the actuator which concerns on 1st embodiment. It is a block chart for demonstrating the process of the control unit which concerns on 1st embodiment. It is a graph which shows the relationship of the lateral force average of a right-and-left wheel at the time of applying this invention, a lateral acceleration, and a tire side slip angle. It is a figure explaining the cornering power at the time of applying this invention. It is a graph which shows the relationship between a lateral acceleration and a wheel load. It is a graph for demonstrating the estimation process of the maximum lateral force. It is a figure explaining the relationship between the tire side slip angle of the turning inner and outer wheels and the tire lateral force. It is a figure explaining the relationship of the actual steering angle of the turning inner and outer wheels in the first embodiment. It is a figure which shows schematic structure of the steering apparatus for vehicles which concerns on 2nd embodiment. It is a figure explaining the relationship of the actual steering angle of the turning inner and outer wheels of the front wheels and rear wheels in the second embodiment. It is a graph which shows the characteristic of SAT. It is a figure explaining an example of 3rd embodiment. It is a figure explaining the other example of 3rd embodiment. It is a figure explaining a parallel steering system. It is a graph which shows the relationship between the tire side slip angle and tire lateral force in a parallel steering system. It is a figure explaining an Ackerman steering system.

Explanation of symbols

1 Vehicle 2 Front Right Wheel 3 Front Left Wheel 4 Rear Right Wheel 5 Rear Left Wheel 6, 8 Actuator (For Front Wheel)
7 Steering gear box 9 Front suspension member 10 Steering wheel 10a Steering shaft 16 Stove roll unit (for hydraulic circuit)
17 Rear suspension members 19, 18 Actuator (for rear wheels)
20 Suspension link 20 Steering wheel 21 Rack 22 Actuator rod 23 Electric power steering device 24 Torque sensor 25 Control unit (for electric power steering device)

Claims (7)

In a vehicle steering apparatus provided with a steering angle control means capable of independently adjusting left and right actual steering angles of a pair of left and right steering wheels steered in response to a steering input from a steering wheel, and provided with a caster rail ,
The rudder angle control means makes the actual rudder angle of the outer turning wheel larger than the actual rudder angle of the inner turning wheel when turning the vehicle ,
The caster rail has a size capable of suppressing transmission of torque to a steering wheel due to self-aligning torque acting in a direction to increase a side slip angle of the steering wheel . Steering device. Provided with steering angle control means that can independently adjust the actual steering angle of the pair of left and right steering wheels that are steered in response to steering input from the steering wheel, and a power steering device that adds auxiliary torque to the steering system In a vehicle steering system,
The rudder angle control means makes the actual rudder angle of the outer turning wheel larger than the actual rudder angle of the inner turning wheel when turning the vehicle ,
The auxiliary torque is a torque that cancels a torque transmitted to the steering wheel due to a self-aligning torque that acts in a direction that increases a side slip angle of the steering wheel when the vehicle turns. apparatus. The steering angle control means, when the vehicle turns, than the actual steering angle of the turning inner wheel, the actual steering angle of the turning outer wheel based on the wheel load of the turning inner and outer, according to claim 1 or claim, characterized in that to increase vehicle steering apparatus according to 2. Wheel load detection means for detecting the wheel load of each of the left and right steering wheels; and maximum lateral force estimation means for estimating the maximum lateral force that can be generated by each of the left and right steering wheels in the wheel load detected by the wheel load detection means;
Lateral force sum estimating means for estimating a lateral force sum that is a sum of lateral forces to be generated by the left and right steered wheels;
The lateral force generated by the left and right steering wheels with respect to the maximum lateral force estimated by the maximum lateral force estimating means for each of the left and right steering wheels is equal or substantially equal between the left and right steering wheels. Each wheel lateral force calculating means for calculating the lateral force generated by the left and right steering wheels by distributing the lateral force sum estimated by the force sum estimating means to the left and right steering wheels,
The steering angle control means adjusts the actual steering angle of each of the left and right steering wheels so that the left and right steering wheels generate the lateral force calculated by the wheel lateral force calculation means. Or the steering device for vehicles of Claim 2 . Vehicle state quantity detection means for detecting the vehicle state quantity, and steering angle detection means for detecting the steering angle of the steering wheel,
The lateral force sum estimating means estimates the lateral force sum based on a vehicle state quantity detected by the vehicle state quantity detecting means and a steering angle detected by the steering angle detecting means. Item 5. The vehicle steering device according to Item 4 . The vehicle steering apparatus according to claim 5 , wherein the vehicle state quantities are a lateral acceleration, a vehicle side slip angle, a yaw rate, and a vehicle speed. The vehicle steering system according to any one of claims 1 to 6, each of the front and rear wheels is equal to or is the steering wheel.

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