JPH0939762A - Controller for behavior of vehicle - Google Patents

Abstract

PROBLEM TO BE SOLVED: To stabilize behavior of a vehicle by determining an ideal steering angle during a braking force control and changing the characteristics of a steering mechanism such that the difference between an actual steering angle and the ideal steering angle during the braking force control is reduced. SOLUTION: Output signals from wheel speed sensors 30FL to 30RR of wheels 12FL to 12RR, a steering angle sensor 34, a transverse G sensor 36 and a yaw rate sensor 38 are input into ECU10 to judge if a braking force control is conducted. When it is judged as YES, an estimated vehicle speed yaw rate is calculated from the wheel speed of a inside front wheel of swiveling and an ideal steering angle is calculated form a vehicle slip angle and an actual steering angle. A degree of over-steering is calculated from the ideal steering angle and the actual steering angle and the assist rate of steering force to be realized for the degree of over-steering is set. A map determining the assist rate to be realized for the degree of over-steering K is memorized in the ECU10.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a vehicle behavior control device, and more particularly to a vehicle behavior control device for stabilizing the vehicle behavior by controlling the braking force of each wheel according to the running state of the vehicle. Regarding

[0002]

2. Description of the Related Art Conventionally, for example, Japanese Patent Laid-Open No. 2-70561.
As disclosed in the publication, there is known a device for stabilizing the vehicle behavior by controlling the braking force of each wheel according to the running state of the vehicle. When a braking force is applied to the front wheels located on the outer side of the turning wheel when the vehicle turns, the braking force acts on the vehicle as a torque that hinders the turning of the vehicle. On the other hand, if a braking force is applied to the rear wheels located on the inner wheel side of the turn, the braking force acts on the vehicle as a torque for assisting the turning of the vehicle.

As described above, the braking force generated on each wheel is
Affects the turning performance of the vehicle. Therefore, by controlling the braking force of each wheel according to the turning state of the vehicle, when the turning speed is excessive, torque in the direction to suppress the turning is provided, and when the turning speed is insufficient, The vehicle behavior can be stabilized at the time of turning by generating a torque in the direction of assisting the turning.

In the above conventional apparatus, the deviation Δγ between the actual yaw rate γ of the vehicle (the turning angular velocity of the vehicle) γ and the target yaw rate γ ₀ corresponding to the vehicle speed V and the steering angle δ is calculated, and Δγ is “0”. The braking force of each wheel is controlled so that According to such control, the actual yaw rate γ that is substantially equal to the target yaw rate γ ₀ can be generated when the vehicle turns, and stable vehicle behavior can be maintained.

[0005]

In the above-mentioned conventional device, the braking force control for stabilizing the vehicle behavior is presumed to be unstable if no vehicle control is performed. Below, it is executed to achieve a stable turning state. At this time, in addition to the turning torque corresponding to the steering angle of the steered wheels, the turning torque resulting from the braking force appropriately generated on each wheel acts on the vehicle. Therefore, during execution of the braking force control, a situation is formed in which a change in the steering angle is difficult to be reflected in the vehicle behavior. Under such a situation, the driver unnecessarily increases or decreases the steering angle, and as a result, the actual steering angle is changed from the ideal steering angle that is originally required for the vehicle to stably turn. There may be cases where the corners are excessive or insufficient.

The braking force control for stabilizing the vehicle behavior is stopped when the vehicle behavior stabilizes. At this time, if there is a deviation between the actual steering angle and the ideal steering angle, the braking force control is stopped, and at the same time, the turning torque acting on the vehicle suddenly changes and the vehicle behavior becomes unstable again.

Further, if the road surface μ changes during execution of the braking force control, the turning torque acting on the vehicle due to the braking force control also changes. When the turning torque due to the braking force control changes, the contribution of the cornering force of the steered wheels to the turning behavior of the vehicle changes. For this reason,
When there is a difference between the actual steering angle and the ideal steering angle, when the turning torque changes due to the braking force control, the vehicle behavior changes before and after that.

As described above, when a deviation occurs between the actual steering angle and the ideal steering angle during the execution of the braking force control, the vehicle behavior becomes unstable when the braking force control is stopped or when the road surface μ changes. It is easy to become. On the other hand, the above-described conventional device has a characteristic that it is easy to generate a deviation between the actual steering angle and the ideal steering angle during execution of the braking force control as described above. In this respect, the above-mentioned conventional device still leaves room for improvement in stabilizing the vehicle behavior.

The present invention has been made in view of the above points, and during execution of the braking force control for the purpose of stabilizing the vehicle behavior, the actual steering angle becomes the ideal steering angle for the running state of the vehicle. An object of the present invention is to provide a vehicle behavior control device that solves the above problems by realizing steering characteristics that are easily matched.

[0010]

SUMMARY OF THE INVENTION The above object is to provide a steering angle detecting means for detecting a steering angle in a vehicle behavior control device for stabilizing the vehicle behavior by controlling the braking force of each wheel during traveling of the vehicle. And an ideal steering angle detecting means for obtaining an ideal steering angle during the braking force control, and a steering for changing the characteristics of the steering mechanism so that the deviation between the actual steering angle during the braking force control and the ideal steering angle becomes small. This is achieved by a vehicle behavior control device including a characteristic changing unit.

In the present invention, the ideal steering angle detecting means detects the ideal steering angle with respect to the running state of the vehicle during execution of the braking force control. Further, the steering characteristic changing means changes the characteristic of the steering mechanism so that the deviation between the ideal steering angle and the actual steering angle detected by the steering angle detecting means becomes small. With this configuration, the actual steering angle during the braking force control can be easily matched with the ideal steering angle.

[0012]

FIG. 1 shows a system configuration diagram of an embodiment of the present invention. The system of this embodiment is controlled by an electronic control unit (ECU) 10 described later. In FIG. 1, 12FL, 12FR, 12RL, 12
RR indicates the left front wheel, right front wheel, left rear wheel, and right rear wheel of the vehicle, respectively. Four wheels 12FL, 12FR, 12RL,
A wheel cylinder (not shown) is incorporated in each of the 12RRs.

When the hydraulic pressure is supplied to each wheel cylinder, each wheel cylinder generates a braking torque corresponding to the hydraulic pressure.
The wheel cylinders of the wheels 12FL, 12FR, 12RL, 12RR have hydraulic control valves 14FL, 14F, respectively.
R, 14RL, and 14RR (hereinafter, when these are collectively referred to by reference numeral 14) are connected. A hydraulic passage 16 and a reservoir tank 18 communicate with the hydraulic control 14. The hydraulic control valve 14 is a two-position valve that operates in response to a signal supplied from the outside, and has a pressure increasing position that communicates the wheel cylinder with the hydraulic passage 16, and a pressure reducing position that communicates the wheel cylinder with the reservoir tank 18. And realize.

A hydraulic power source switching valve 20 is provided in the hydraulic passage 16.
Are in communication. The hydraulic source switching valve 20 is also in communication with a high pressure source including a hydraulic pump 22 and an accumulator 24 and a master cylinder 26. The hydraulic power source switching valve 20 is a two-position valve that operates in response to a signal supplied from the outside, and connects the hydraulic passage 16 and the master cylinder 26 to a control execution position that connects the hydraulic passage 16 and the hydraulic pump 22. To achieve the normal position.

The hydraulic pump 22 pumps up the brake fluid from the reservoir tank 22 and feeds it to the accumulator 24 side under the condition that the hydraulic pressure switching valve 20 is in the control execution position. The accumulator 24 stores the hydraulic pressure generated at that time and supplies a stable hydraulic pressure with little pulsation to the hydraulic pressure switching valve 20. Therefore, when the hydraulic pressure switching valve 20 is at the control execution position, the hydraulic pump 2 is provided in the hydraulic passage 16.
A predetermined hydraulic pressure corresponding to the discharge capacity of 2 is introduced. The master cylinder 26 generates a hydraulic pressure according to the brake pedal force applied to the brake pedal 28. Therefore, when the hydraulic pressure switching valve 20 is in the normal position, the hydraulic pressure according to the brake pedal force is guided to the hydraulic passage 16.

In this embodiment, the hydraulic control valve 1 described above is used.
4 and the hydraulic pressure source switching valve 20 are controlled by the ECU 10. The ECU 10 has wheels 12FL, 12
Wheel speed V _WFL, V of FR, 12RL, 12RR respectively
_WFR, V _WRL, V _WRR (Hereinafter, when these are collectively referred to,
A wheel speed sensor 30FL for detecting the wheel speed V _W ),
30FR, 30RL, 30RR (hereinafter, these are collectively denoted by reference numeral 30), a steering angle sensor 34 for detecting the steering angle δ of the steering wheel 32, and a lateral acceleration Gy acting on the vehicle. Lateral G sensor 36,
And a turning angular velocity generated around the center of gravity of the vehicle, that is,
A yaw rate sensor 38 that detects the yaw rate γ of the vehicle is connected.

The steering wheel 32 is connected to a hydraulic reaction force type power steering device (hereinafter referred to as a PS device) 40. Hereinafter, referring to FIG. 2, the PS device 40
The configuration of will be described. FIG. 2 shows a system configuration diagram of the PS device 40. As shown in FIG. 2, the steering wheel 32 is fixed to the upper end of the steering shaft 42. The lower end of the steering shaft 42 is connected to the valve mechanism 44. A pinion gear 46 engaged with a steering rack (not shown) is fixed to the lower end of the valve mechanism 44. When the steering wheel is steered, the steering force is applied to the steering shaft 42,
It is transmitted to the steering rack via the valve mechanism 44 and the pinion gear 46. Therefore, a twisting torque corresponding to the steering force acts on the valve mechanism 44.

Inside the valve mechanism 44, the valve mechanism 4
A throttle mechanism 44a for changing the throttle opening degree according to the magnitude of the twist angle generated inside the valve mechanism 44 when a twisting torque is applied to the valve 4 is incorporated. The function of the diaphragm mechanism 44a can be represented by using the four diaphragms X, X ', Y, Y'indicated by a chain line in FIG. These throttles X, X ', Y, Y'are all throttles X, X, when steering torque is not input to the valve mechanism 44.
X ′, Y, Y ′ have the same opening area, and the valve mechanism 4
When the steering torque in the left steering direction is input to 4, the apertures X and Y are enlarged and the apertures X ′ and Y ′ are reduced,
Further, when the steering wheel in the right steering direction is input to the valve mechanism 44, the apertures X'and Y'are enlarged and the apertures X and Y are reduced.

The upstream side of the throttle mechanism 44a (the upstream side of the throttles X and X ') communicates with the oil liquid discharge port of the oil pump 48. The downstream side of the throttle mechanism 44 a (downstream side of the throttles Y and Y ′) communicates with the reservoir tank 50. The reservoir tank 50 communicates with the oil liquid suction port of the oil pump 48. Further, the first portion of the power cylinder 52 is provided in a portion of the diaphragm mechanism 44a that connects the diaphragm X and the diaphragm Y '.
The power cylinder 5 is provided at a portion where the hydraulic passage 54a communicating with the hydraulic chamber 52a and the throttle X'and the throttle Y communicate with each other.
The hydraulic passages 54b communicating with the second hydraulic chambers 52b communicate with each other.

According to this structure, when no steering torque is applied to the steering shaft 42,
The hydraulic pressure supplied to the hydraulic passage 54a and the hydraulic pressure supplied to the hydraulic passage 54b become equal pressure, and as a result, the equal hydraulic pressure is guided to the first hydraulic chamber 52a and the second hydraulic chamber 52b of the power cylinder 52. Get burned. In this case, no thrust is generated in the power cylinder 52.

On the other hand, the steering shaft 42
When a steering torque is applied to the apertures, the apertures X, X ', Y, Y'
Is not uniform, and different hydraulic pressures are supplied to the hydraulic passages 54a and 54b. In this case, the power cylinder 52
A pressure difference is generated between the first hydraulic chamber 52a and the second hydraulic chamber 52b, and the power cylinder 52 generates a thrust force corresponding to the pressure difference.

The power cylinder 52 is provided in series in the axial direction of the steering rack. Therefore, the horizontal thrust generated in the power cylinder 52 acts as a force for displacing the steering rack in the steering direction, that is, as a steering assist force. In the steering control device of the present embodiment, the steering assist force when the steering wheel 32 is steered is generated in this way.

The valve mechanism 44 includes the valve mechanism 4
In order to make the torsional rigidity of 4 variable, a hydraulic reaction chamber is built in. The hydraulic reaction force chamber communicates with the reaction force hydraulic passage 58 that communicates the oil pump 48 with the reservoir tank 50 via the hydraulic passage 56. A variable throttle 60 is provided in the reaction force hydraulic passage 58 at the downstream side of the connection portion with the hydraulic passage 56. With such a configuration, the larger the opening of the variable throttle, the higher the hydraulic pressure guided to the hydraulic passage 56, that is, the hydraulic pressure introduced to the hydraulic reaction force chamber, while the smaller the opening of the variable throttle, the larger the hydraulic pressure. The hydraulic pressure introduced into the force chamber is constant.

The valve mechanism 44 is constructed so that the higher the hydraulic pressure introduced into the hydraulic reaction chamber, the higher the torsional rigidity. Therefore, the valve mechanism 44 exhibits low torsional rigidity when the opening of the variable throttle 60 is large, and exhibits high torsional rigidity when the opening of the variable throttle 60 is small.
When the valve mechanism 44 exhibits high torsional rigidity, the amount of change in the opening area of the apertures X, X ', Y, Y'per unit steering force becomes small. Therefore, in this case, the steering characteristic in which the steering assist force by the power cylinder 52 is difficult to obtain, that is, the steering characteristic having a high sense of rigidity is realized. On the other hand, when the torsional rigidity of the valve mechanism 44 is low, the throttles X, X ', Y, per unit steering force,
The amount of change in the opening area of Y'becomes large. Therefore, in this case, a state where the steering assist force by the power cylinder 52 is easily obtained, that is, a steering characteristic with a light feeling is realized. Therefore, according to the PS device 40 of this embodiment, the variable diaphragm 60
By adjusting the opening degree of the steering wheel, it is possible to realize both a steering characteristic having a high sense of rigidity and a steering characteristic giving priority to a feeling of lightness.

Next, the content of the braking force control for stabilizing the vehicle behavior, which is executed in the system of this embodiment, will be described with reference to FIGS. 3 to 9. FIG.
The figure which showed the vehicle in the left turn in planar view is shown. In FIG. 3, “C” represents the center of gravity of the vehicle. As shown in the figure, when the vehicle is making a left turn, a yaw rate γ is generated around the center of gravity C of the vehicle in the counterclockwise direction. If an appropriate yaw rate γ according to the vehicle speed V and the steering angle δ is realized while the vehicle is traveling, it can be estimated that the vehicle is turning in a stable state. On the other hand, if γ is excessive with respect to V and δ, the turning speed of the vehicle is excessive, that is,
It can be estimated that the vehicle has a tendency to spin, and if γ is insufficient, it can be estimated that the vehicle is not turning properly, that is, the vehicle has a tendency to drift out.

By the way, when the front wheel FR located on the outer side of the turning wheel generates a braking force F _BRK during turning of the vehicle as indicated by a solid arrow in FIG. 3, the braking force F _BRK is _relative to the center of gravity C. And acts as a torque in a direction that prevents the vehicle from turning. Therefore, while the vehicle is turning, the front wheel 12FL on the turning outer wheel side is
Alternatively, if the braking force is generated in 12FR, the yaw rate γ generated in the vehicle can be suppressed.

On the other hand, when the braking force F _BRK is generated on the rear wheels 12RL, 12RR while the vehicle is turning, as shown by the broken line arrow in FIG. 3, the center of gravity of the vehicle shifts to the front wheels 12FL, 12FR side. , Centripetal force inward in the turning direction increases.
Further, the braking force F _BRK generated by the rear wheel RL located on the turning inner wheel side acts on the center of gravity C as a torque in a direction for assisting the turning of the vehicle. Therefore, while the vehicle is turning, the rear wheels 1
If the braking force is generated in 2RL and 12RR, the yaw rate γ can be assisted.

Therefore, in the system of the present embodiment, behavior estimation is performed during turning of the vehicle, and according to the estimated behavior,
When it is determined that the vehicle has a tendency to drift out, an appropriate hydraulic pressure is supplied to the wheel cylinders of the rear wheels 12RL and 12RR, and when the vehicle has a tendency to spin, the front wheel 12FL on the outer wheel turning side or 12FR (hereinafter fo
The vehicle behavior is stabilized by supplying an appropriate hydraulic pressure to a wheel cylinder (referred to as ut).

In this embodiment, the concept of the spin degree SV and the drift degree DV is introduced as a criterion for determining whether or not the behavior of the vehicle is stable. Spin degree S
V is the degree of oversteer tendency during turning, and the drift degree DV is the degree of understeer tendency during turning. Hereinafter, a method for obtaining them will be described with reference to FIG.

FIG. 4 shows an equivalent two-wheel vehicle model of a four-wheel vehicle used for estimating the behavior of the vehicle during turning. In FIG. 4, C is the center of gravity of the vehicle, V is the vehicle speed, and β
Is the angle of travel of the center of gravity with respect to the axle (hereinafter referred to as the vehicle body slip angle), βf is the slip angle of the front wheels, βr is the slip angle of the rear wheels, γ is the yaw rate around the center of gravity, and Ff is the front wheel 1.
2FL, 12FR cornering force, Fr
Is the resultant force of the cornering forces of the rear wheels 12RL and 12RR, δ is the steering angle, a is the distance between the front wheel axle and the center of gravity C, and b is the distance between the rear wheel axle and the center of gravity C.

In the two-wheeled vehicle model shown in FIG. 4, when the vehicle weight is m, the following equation of motion is established on the Y axis passing through the center of gravity C. mV (dβ / dt + γ) = Ff + Fr (1) The first term (mV · dβ / dt) in the left side of the above equation (1) is the translational acceleration (V · dβ / acting on the center of gravity C of the vehicle. dt)
And the vehicle weight (m). Further, the second term (mVγ) on the left side of the equation (1) is a centrifugal force acting on the vehicle. The total value of them becomes the total value of the lateral force acting on the vehicle and is balanced with Ff + Fr shown on the right side.

The total value of the lateral forces acting on the vehicle is Ff + Fr.
Then, the lateral acceleration Gy acting on the vehicle can be expressed by the following equation. Gy = (Ff + Fr) / m (2) When the above equations (1) and (2) are arranged, the translational acceleration (V · dβ / dt) can be expressed as the following equation.

V · dβ / dt = Gy−V · γ (3) Therefore, the rate of change dβ / dt of the vehicle slip angle β and the slip angle β can be expressed as follows. dβ / dt = (Gy / V) −γ (4) β = ∫ {(Gy / V) −γ)} dt (5) Used in the above equations (4) and (5) Parameter G
y, V, and γ can be measured by the lateral G sensor 36, the wheel speed sensor 30, and the yaw rate sensor 38, respectively. Therefore, according to the system of the present embodiment, the vehicle slip angle β and its change rate dβ / dt can be accurately calculated.

By the way, the slip angle β of the vehicle is a parameter having a larger value as the turning speed of the vehicle is higher. Therefore, it is possible to determine that the vehicle behavior tends to have a spin tendency as the value increases. Also, the slip angle β
The rate of change dβ / dt of is a parameter that takes a large value when the turning speed of the vehicle is rapidly increased. Therefore, it can be determined that the vehicle has a tendency to spin as the value increases. Therefore, in the present embodiment, the spin degree S
V is defined by the following equation using constants k ₁ and k ₂ .

SV = k ₁ · β + k ₂ · dβ / dt (6) On the other hand, the drift degree DV is defined based on the yaw rate γ. That is, when the vehicle is stably turning in the neutral steer state, the yaw rate γ according to the steering angle δ and the vehicle speed V is generated around the center of gravity C. Therefore,
Actual yaw rate γ measured by the yaw rate sensor 38
However, if the yaw rate is smaller than the yaw rate estimated from the steering angle δ and the vehicle speed V, it can be determined that the vehicle behavior tends to drift. Therefore, in the present embodiment, using the deviation Δγ between the target yaw rate γ ₀ determined by the relationship between the steering angle δ and the vehicle speed V and the actual yaw rate γ actually acting on the vehicle, and the constant k ₃ , The drift degree DV is defined as in the equation.

DV = k ₃ · Δγ (7) In the present embodiment, the ECU 10 calculates the spin degree SV and the drift degree DV according to the above method, and the braking force of each wheel based on the calculation result. Is controlled to stabilize the vehicle behavior during turning. 5 and 6
Shows an example of a flowchart of a braking force control routine executed by the ECU 10 so as to realize such a function.

As shown in FIG. 5, when this routine is started, first, at step 100, various parameters required for executing this routine are read. Specifically, the lateral acceleration Gy acting on the vehicle and the yaw rate γ,
The vehicle speed V and the steering angle δ are read.

In step 102, the rate of change of the vehicle body slip angle β is dβ / dt = (Gy / V) according to the above equation (4).
−γ is calculated. Further, in step 104, the vehicle body slip angle β = ∫ {(Gy
/ V) -γ} dt is calculated. Then, Step 106
In, the spin degree SV = k ₁ · β + k ₂ · dβ / dt is calculated by substituting the calculated values into the above equation (6).

In step 108, the processing for obtaining the target yaw rate γ ₀ corresponding to the vehicle speed V and the steering angle δ is executed. The ECU 10 stores a map that defines the target yaw rate γ _{0 in} relation to V and δ, and in this step, γ ₀ is calculated by searching the map. Next, at step 110, a deviation Δγ = γ ₀ −γ between the target yaw rate γ ₀ obtained as described above and the actual yaw rate γ read at step 100 is calculated.
Then, in step 112, by substituting Δγ into the equation (7), the degree of drift DV = k ₃ · Δγ is calculated.

After the spin degree SV and the drift degree DV of the vehicle are calculated as described above, the braking force control for stabilizing the vehicle behavior during turning is executed based on these values. That is, in step 114, assuming that the vehicle has a tendency to spin, the target slip ratio S ₀ fout to be realized by the front outer wheel fout on the turning outer wheel side is calculated based on the spin degree SV.

As described above, the braking force generated by the front wheel fout on the turning outer wheel side acts on the vehicle as a torque for suppressing the yaw rate γ. Therefore, if a braking force corresponding to the spin degree SV is generated in the front wheel fout on the outer turning wheel side, the spin tendency of the vehicle can be appropriately suppressed.

By the way, at the time of braking, a braking force corresponding to the slip ratio of the wheel is generated on the wheel. That is, the braking force of the wheel is generated when the brake torque acting on the wheel causes slip between the tire and the road surface. Then, the braking force exhibits a maximum value at a predetermined slip ratio (hereinafter referred to as a limit slip ratio) according to the characteristics of the tire, and becomes a value substantially proportional to the slip ratio in a region below the limit slip ratio. Therefore, when the braking force control is performed, by controlling the brake hydraulic pressure so that the slip ratio does not exceed the limit slip ratio, the grip state of the wheels can always be properly maintained. Further, in a region where the slip ratio does not exceed the limit slip ratio, by controlling the brake hydraulic pressure so that the slip ratio becomes the target value, it is possible to accurately control the braking force generated between the tire and the road surface. .

Therefore, in this embodiment, the braking force of each wheel is controlled based on the slip ratio of each wheel. For the above reason, in step 114, the target slip ratio S ₀ fout to be realized by the front outer wheel fout on the turning outer wheel side is calculated based on the spin degree SV.

In step 114, specifically, the target slip ratio S ₀ fout is calculated by searching the map shown in FIG. 7 with the spin degree SV. Since the spin degree SV may be calculated with a small value even when the vehicle is traveling in a stable turn, the target slip ratio map is a predetermined value SV.
_The dead zone is the area below ₀ . Further, in order to prevent the target slip ratio S ₀ fout exceeding the limit slip ratio of the tire from being calculated, the target slip ratio map is set so that the target slip ratio S ₀ fout is saturated in the region of the predetermined value SV ₁ or more. It is set.

When the target slip ratio S ₀ fout is set according to the map shown in FIG. 6 and the slip ratio is realized by the front outer wheel fout on the turning outer wheel side, in the region of SV ₀ <SV, the degree of the tendency of spin of the vehicle is determined. Accordingly, an appropriate large braking force is generated in the direction in which the spin tendency is suppressed.

Further, in step 116, assuming that the vehicle is in a tendency to drift, the rear wheels 12R on the turning outer wheel side are assumed.
L or 12RR (hereinafter referred to as rout) and target slip ratios S ₀ rout and S ₀ rin to be realized by the rear wheel 12RL or 12RR (hereinafter referred to as rin) on the turning inner wheel side are calculated based on the drift degree DV. To be done. As mentioned above, the rear wheels 12R
The braking force generated by L, 12RR acts on the vehicle as a force that increases the centripetal force of the vehicle during turning. Therefore, the rear wheel rout on the turning outer wheel side and the rear wheel rin on the turning inner wheel side
If a braking force corresponding to the drift degree DV is generated for each of them, the drift tendency of the vehicle can be appropriately suppressed.

In this step 116, the rear wheel on the turning outer wheel side is
Target slip ratio S that should be achieved by rout ₀rout is shown in Figure 8
Follow the map and realize with the rear wheel rin on the turning inner wheel
Power target slip ratio S₀rin follows the map shown in Figure 9
Each is calculated. The maps shown in FIGS. 8 and 9 are
For the same reason as the map shown in FIG. 7, DV ≦ DV
₀Area is set as the dead zone and DV₁
<Slip ratio S in the DV range₀rout, S₀rin gets tired
It is set to be harmonized.

According to these maps, the target slip ratio S
_{When 0} rout and S ₀ rin are set, and the slip ratios are realized at the rear wheel rout on the outer turning wheel side and the rear wheel rin on the inner turning wheel side, respectively, in the region of DV ₀ <DV, the drift tendency of the vehicle Depending on the degree, an appropriate amount of braking force is generated in the direction in which the drift tendency is suppressed.

When the above processing is completed, next step 1
At 18, the turning direction of the vehicle is identified. The yaw rate sensor 38 outputs a yaw rate signal having a different sign depending on the turning direction of the vehicle. In this step, the turning direction of the vehicle is specified based on the code. When the turning direction is specified in this way, the turning outer wheel and the turning inner wheel are determined based on the result.

When the processing of step 118 is completed, the processing of step 120 shown in FIG. 6 is then executed. In step 120, the estimated vehicle speed V is calculated based on the output signal of the wheel speed sensor 30. As described above, in the system of the present embodiment, when the vehicle is turning, braking force is generated on the front wheel fout on the turning outer wheel side and the left and right rear wheels rout, rin to stabilize the vehicle behavior. Therefore, these turning outer side of the front wheels fout and left and right rear wheels rout, and the wheel speed V _W of rin, is little difference between the vehicle speed occurs. On the other hand, the front wheel 12FL or 12FR (hereinafter, referred to as fin) on the turning inner wheel side does not generate the braking force. Therefore, the wheel speed V _W of the front fin of the turning inner wheel side is always a value corresponding with the vehicle speed. Therefore, in step 120, the estimated vehicle speed V is _calculated based on the wheel speed V _WFL or V _WFR of the front wheel fin on the turning inner wheel side.
Is calculated.

After the above processing is completed, next step 12
In 2, it is determined whether or not the target slip ratio S ₀ fout> 0 set for the front wheel fout on the outer turning wheel side is satisfied. When the slip tendency of the vehicle is strong and the slip degree SV exceeding the predetermined value (SV ₀ shown in FIG. 7) is detected, the condition of this step is satisfied. In this case, the process of step 124 is executed thereafter. On the other hand, when the slip tendency of the vehicle is weak and the SV is less than the predetermined value, the condition of this step is not satisfied. In this case, steps 124 and 126 are jumped thereafter, and the process of step 128 is executed.

At step 124, the front wheel fo on the side of the turning outer wheel fo
The theoretical wheel speed V ₀ fout of ut is calculated. Theoretical wheel speed V ₀ f
out is a wheel speed estimated to occur at the front wheel fout on the outer wheel side of the turning when the vehicle is turning at the estimated wheel speed V, and is calculated based on the estimated wheel speed V. When the wheel speed V ₀ fout is generated on the front wheel fout on the outer turning wheel side, the slip ratio of the wheel is “0”.

After the above processing is completed, next step 12
At 6, a process for controlling the braking force of the front wheels fout is executed. Specifically, first, the theoretical wheel speed V ₀ f
based on the wheel speed Vfout of out and reality, the slip ratio of the turning outer side of the front wheel _{fout Sfout = (1-V 0} fout / Vfou
t) × 100 is calculated. Then, the slip ratio Sf
The brake hydraulic pressure supplied to the front wheel fout on the turning outer wheel side is controlled so that out matches the target slip ratio S ₀ fout. When such control is executed, the front wheel fout on the turning outer wheel side
In, the braking state with the target slip ratio S ₀ fout is realized.

When the processing of step 126 is completed, it is then determined in step 128 whether or not the target slip ratio S ₀ rout> 0 set for the rear wheel rout on the turning outer wheel side is satisfied. When the drift tendency of the vehicle is strong and the drift degree DV exceeding the predetermined value (DV ₀ shown in FIG. 8) is detected, the condition of this step is satisfied. In this case, the process of step 130 is executed thereafter. On the other hand, when the drift tendency of the vehicle is weak and DV is less than the predetermined value, the condition of this step is not satisfied. In this case, steps 130 and 132 are jumped thereafter, and step 1
34 is executed.

In step 130, the rear wheel ro on the turning outer wheel side is rotated.
The theoretical wheel speed V ₀ rout of ut is calculated. Theoretical wheel speed V ₀ r
out is a wheel speed estimated to occur at the rear wheel rout on the outer wheel side when the vehicle is turning at the estimated wheel speed V, and is calculated based on the estimated wheel speed V. When such a wheel speed V ₀ rout is generated on the rear wheel rout on the turning outer wheel side, the slip ratio of the wheel is “0”.

After the above processing is completed, next step 13
At 2, the process for controlling the braking force of the rear wheel rout is executed. Specifically, first, the theoretical wheel speed V ₀ r
out and the actual wheel speed Vrout, the slip ratio Srout of the rear wheel rout on the turning outer wheel side is Srout = (1−V ₀ rout / Vrou
t) × 100 is calculated. Then, the slip ratio Sr
The brake hydraulic pressure supplied to the rear wheel rout on the turning outer wheel side is controlled so that out matches the target slip ratio S ₀ fout. When this control is executed, the rear wheel rout on the turning outer wheel side
In, the braking state with the target slip ratio S ₀ rout is realized.

When the processing of step 132 is completed, the routine proceeds to step 134, where it is judged if the target slip ratio S ₀ rin> 0 set for the rear wheel rin on the turning inner wheel side is satisfied. As a result, if the above conditions are not satisfied, steps 136 and 138 are jumped, and the routine of this time is ended. On the other hand, the target slip ratio S ₀ rin
If the condition> 0 is satisfied, then in steps 136 and 138, the above steps 130 and 132 are performed.
A process similar to the above is executed, and the braking state with the target slip ratio S ₀ rin is realized on the rear wheel rin on the turning inner wheel side.

As described above, when the ECU 10 executes the routines shown in FIGS. 5 and 6, when the vehicle has a tendency to spin, an appropriate braking force is generated on the front wheel fout on the outer wheel turning side, and the vehicle Spin tendency is suppressed. On the other hand, if the vehicle tends to drift, the left and right rear wheels rout, rin
An appropriate braking force is generated in the vehicle, and the drift tendency of the vehicle is suppressed. Therefore, in a vehicle equipped with the system of the present embodiment, stable vehicle behavior can be realized during turning.

By the way, the above braking force control realizes a steady circular turning state, that is, a state in which the front wheels and the rear wheels generate a cornering force with an appropriate ratio when the turning behavior of the vehicle becomes unstable. Will be executed. Therefore, even if an excessive steering angle δ is given to the steered wheels,
Even if the steering angle δ is turned back improperly, the vehicle tries to maintain a steady circular turning state. Therefore, during the execution of the braking force control, the change in the steering angle δ is difficult to be reflected in the vehicle behavior. In such a situation where the change in the steering angle δ is difficult to be reflected in the vehicle behavior, an excessive actual steering angle δ is generated with respect to the ideal steering angle originally required for turning the vehicle in a stable state, Or make the actual steering angle δ insufficient,
Unnecessary steering is easy to perform.

Then, in a situation where there is a deviation between the ideal steering angle and the actual steering angle, for example, the vehicle behavior is stabilized, that is, both the spin degree SV and the drift degree DV are suppressed below a predetermined value. When it is determined that the braking force control is ended, or when the road surface μ changes, the vehicle behavior is likely to be unstable as described above.

Therefore, in the system of this embodiment, the ideal steering angle δ ^* corresponding to the running state of the vehicle is estimated during the execution of the braking force control, and the actual steering angle δ substantially matches the ideal steering angle δ ^*. In the region where the steering angle is increased, a large steering assist force is generated in the PS device 40, and the more the actual steering angle δ departs from the ideal steering angle δ ^*
The steering assist force of the S device 40 is to be reduced.
According to this configuration, after the execution of the braking force control is started, it is necessary to input a large steering torque to the steering wheel 32 in order to realize the steering angle δ that is largely separated from the ideal steering angle δ ^*. Become. Therefore, according to the system of the present embodiment, it is possible to effectively prevent the steering angle δ from being unnecessarily increased or turned back during the execution of the braking force control.

Hereinafter, the specific processing contents for realizing the above-mentioned function, that is, the processing contents for obtaining the theoretical steering angle δ ^* corresponding to the running state of the vehicle during the execution of the braking force control, and its processing The PS device 40 based on the theoretical steering angle δ ^*
The contents of processing for changing the characteristics of will be described.

As shown in the vehicle model shown in FIG. 4, the cornering force of the front wheels is Ff, the cornering force of the rear wheels is Fr, the distance between the front wheel axle and the center of gravity C is a, and the rear wheel axle and the center of gravity C are If the distance is b, the relationship shown in the following equation holds when the vehicle makes a steady circle turn. Since the braking force control is executed to realize the steady circular turning state as described above, it is estimated that the relationship shown in the following equation (8) is established during the execution of the control.

Ff: Fr = b: a (8) By the way, the cornering force generated on the wheel is almost proportional to the slip angle β until the slip angle β of the wheel reaches the limit slip angle β _0. Increase. If the constant of proportionality is called cornering power, the cornering power of the front wheels is Cf, and the cornering power of the rear wheels is Cr, then Ff = Cf · βf and Fr = Cr · βr are established.

When the vehicle is turning while generating the yaw rate γ around the center of gravity C as in the vehicle model shown in FIG. 4, the velocity vector V _{f of the} front wheels is equal to the magnitude a · γ toward the inside of the turn. It can be grasped as a composite vector of the vector and the vector of the vehicle speed V. Further, the rear wheel speed vector V _r can be grasped as a composite vector of a vector of the magnitude b · γ toward the outside of the turn and a vector of the vehicle speed V.

In this case, the angle formed by the direction of the vehicle speed V and the advancing direction of the front wheels and the angle formed by the direction of the vehicle speed V and the advancing direction of the rear wheels are "a.γ / V" and "b.γ", respectively. / V "
It can be expressed as. Therefore, the front wheel slip angle β _f ,
And the slip angle β _f of the rear wheels, using the steering angle δ of the front wheels and the vehicle body slip angle β, the following equations (9) and (1
0).

Β _f = β + a · γ / V−δ (9) β _r = β−b · γ / V (10) Therefore, the cornering force Ff of the front wheels and the cornering force Fr of the rear wheels are obtained. Are respectively expressed by the following equations (11),
It can be expressed as (12).

Ff = Cf (β + a · γ / V−δ) (11) Fr = Cr (β−b · γ / V) (12) Formulas (11) and (12) Is applied to the relation of the above equation (8), the relational expression of the following equation (13) is established. Further, by solving the equation (13) for δ, the following equation (14) is obtained.

A · Cf (β + a · γ / V−δ) = b · Cr (β−b · γ / V) (13) δ = {a + (Cr / Cf) · (b ² / a) } Γ / V + {1- (Cr / Cf) · (b / a)} β ... (14) δ shown in the above equation (14) is a steering that satisfies the condition for realizing a steady circular turning state. It is a horn. Therefore, after the braking force control is started and the vehicle is controlled to the steady circular turning state,
When the steering angle δ satisfying the relationship of the above equation (14) is realized, the vehicle can maintain the steady circular turning state even if the braking force control is stopped thereafter or the road surface μ changes. Therefore, in the present embodiment, δ that satisfies the relationship of the above formula (14) is used as the theoretical steering angle δ ^* during turning.

FIG. 10 is a diagram showing the relationship between the ideal steering angle δ ^* and the actual steering angle δ on a two-dimensional coordinate system. The δ ^* = δ straight line shown in FIG. 10 represents an ideal state that is realized when the actual steering angle δ matches the ideal steering angle δ ^* . Also,
In FIG. 10, an area below the δ ^* = δ straight line in the first quadrant and an area above the δ ^* = δ straight line in the third quadrant are areas where the actual steering angle δ is insufficient, and δ ^* = in the first quadrant. The area above the straight line δ and the area below the straight line δ ^* = δ in the third quadrant are the actual steering angle δ.
In the second quadrant and the fourth quadrant, a region where the steering direction and the direction of the ideal steering angle δ ^* are reversed, that is, a counter steer region, respectively, are shown.

In FIG. 10, δ^*= Δ holds
Region, ideal region is δ / δ^*Area where = 1 holds
It is. Further, the excess region shown in FIG.^*> 1
Is an area where Furthermore, the shortage area shown in FIG.
And the counter steer area is δ / δ^*Area where <1 holds
Area. Therefore, during execution of the braking force control, the actual steering angle
δ and ideal steering angle δ^*And obtain δ / δ^*=
1, δ / δ^*> 1, δ / δ ^*Matches any of the conditions <1
If it is, it is judged whether the steering angle is excessive or insufficient.
be able to.

In the present embodiment, in order to facilitate the subsequent processing, the characteristic value of the overcut degree K expressed by the following equation is introduced to judge whether the steering angle is excessive or insufficient. K = (δ / δ ^* )-1 (15) That is, according to the overcut degree K, K = in the ideal region.
0, K> 1 in the excess area, and K <1 in the insufficient area and the counter steer area. Therefore, when the excessive cut degree K is used, it is possible to more easily determine whether the steering angle is excessive or insufficient, depending on which of K = 0, K> 1, and K <1 holds.

FIG. 11 shows the ECU 10 for calculating the ideal steering angle δ ^* during execution of the braking force control and changing the assist ratio of the PS device 40 so as to match the actual steering angle δ with the theoretical steering angle δ ^*. 3 is a flowchart of an example of a control routine executed by the. When the routine shown in FIG. 11 is started, first, at step 200, it is judged which of the wheels is under the braking force control. As a result, when it is determined that the braking force control is not executed for any of the wheels, it is determined that no special process needs to be performed, and the current process is ended. On the other hand, if it is determined that the braking force control is being executed on any of the wheels, step 202 is performed in order to realize an appropriate steering characteristic.
Proceed to.

At step 202, various parameters necessary for executing this routine are read. Specifically, the estimated vehicle body speed V calculated in the process of step 120 shown in FIG. 6, the yaw rate γ detected by the yaw rate sensor 38, and the vehicle body slip angle β calculated in the process of step 104 shown in FIG. , And the actual steering angle δ detected by the steering angle sensor 34 are read.

When the reading of the various parameters described above is completed, then in step 204, the above step 2
The ideal steering angle δ ^* is calculated by substituting V, γ, and β read in 02 into the equation (14). Then
In step 206, the ideal steering angle δ ^* and the actual steering angle δ read in step 202 are substituted into the above equation (15) to calculate the overcut degree K.

Thereafter, in step 208, the assist ratio of the steering force to be realized with respect to K is set, and then this routine is ended. The ECU 10 determines the overcut degree K
A map that defines the magnitude of the assist rate to be realized is stored. In this step, the assist rate is set by searching the map.

FIG. 12 shows an example of the assist rate map stored in the ECU 10. As shown in FIG. 12, in the present embodiment, when the overcut degree K is “0”, that is, when the actual steering angle δ matches the theoretical steering angle δ ^* , the assist ratio has a peak value “1”. Is set to 0, and the assist ratio is reduced as K departs from “0”, that is, as the deviation between the actual steering angle δ and the theoretical steering angle δ ^* increases.

The ECU 10 sends a drive signal corresponding to the assist ratio set as described above to the variable diaphragm 6 of the PS device 40.
Supply 0. As a result, the hydraulic pressure corresponding to the assist ratio set in step 208 is introduced into the hydraulic reaction chamber of the valve mechanism 44, and the steering characteristic corresponding to the set assist ratio is realized in the PS device 40. .

When the assist ratio of the PS device 40 is set as described above, when the driver performs the correction steering in the direction in which the deviation between the actual steering angle δ and the theoretical steering angle δ ^* decreases, the assist is gradually assisted. Power becomes high. Therefore, the correction steering in such a direction can be easily performed. On the other hand, when the driver performs the correction steering in the direction in which the deviation between the actual steering angle δ and the theoretical steering angle δ ^* increases, the assist force gradually decreases, and in order to continue the correction steering, the large steering angle is required. Requires torque. Therefore, unnecessary correction steering in this direction is effectively prevented.

[0080] Thus, according to the system of the present embodiment, the steering angle during the execution of the braking force control [delta] is liable to be matched to the theoretical steering angle [delta] ^*, and hardly separates from the theoretical steering angle [delta] ^*. Therefore, according to the system of the present embodiment, the vehicle behavior can always be maintained in a stable state, including before and after the braking force control is stopped and before and after the change of the road surface μ.

In the above embodiment, by controlling the hydraulic pressure introduced into the hydraulic reaction chamber, the torsional rigidity of the valve mechanism is changed to obtain the desired steering characteristic, but the steering characteristic is changed. The method is not limited to this. 13 and 14 show another embodiment of the mechanism for changing the steering characteristic. 13 and 14, the same parts as those shown in FIG.
The same reference numerals are given and the description is omitted.

FIG. 13 shows P suitable for the system of this embodiment.
The 2nd Example of S apparatus is shown. The PS device shown in FIG.
It is a flow control PS device that changes the steering characteristics by changing the amount of oil flowing through the throttle mechanism 44a. The PS device according to the present embodiment includes only a hydraulic passage 62 and a variable throttle 64 provided in the hydraulic passage 62 as a mechanism for changing the steering characteristic.

When the aperture area of the variable diaphragm 64 is large,
A large amount of oil liquid flowing out from the oil pump 48 is supplied to the hydraulic passage 6.
2 and flows into the reservoir tank 50. in this case,
When a twisting torque acts on the throttle mechanism 44a, a pressure difference between the first hydraulic chamber 52a and the second hydraulic chamber 52b of the power cylinder 52 is unlikely to occur, and a steering characteristic with a low assist ratio is realized in the PS device.

On the other hand, when the opening area of the variable throttle 64 is small, a large amount of oil liquid flowing out from the oil pump 48 flows into the reservoir tank 50 via the hydraulic passage 62. In this case, when a torsion torque acts on the throttle mechanism 44a, the first hydraulic chamber 52a and the second hydraulic chamber 52 of the power cylinder 52 are
A pressure difference with b is easily generated, and a steering characteristic with a high assist ratio is realized in the PS device.

As described above, according to the PS device shown in FIG. 13, the assist ratio is reduced by increasing the opening area of the variable diaphragm 64, and the assist area is reduced by decreasing the opening area of the variable diaphragm 64. The rate can be increased. Therefore, the PS device shown in FIG. 13 can also realize the same function as the PS device 40 shown in FIG.

FIG. 14 shows P suitable for the system of this embodiment.
The 3rd Example of S apparatus is shown. The PS device shown in FIG.
The PS device changes the steering characteristics by changing the amount of oil flowing into the power cylinder 52 and the amount of oil flowing out of the power cylinder 52. The PS device of this embodiment has hydraulic passages 54a and 5a as mechanisms for changing the steering characteristics.
4b is provided with variable diaphragms 66 and 68, respectively.

When the opening areas of the variable throttles 66 and 68 are large, the amount of oil that flows into the power cylinder 52 and flows out from the power cylinder 52 when a pressure difference occurs between the hydraulic passages 54a and 54b. Is relatively large. In this case, when a torsion torque acts on the throttle mechanism 44a, the first hydraulic chamber 52a and the second hydraulic chamber 52b of the power cylinder 52 are
And a steering characteristic with a high assist ratio is realized in the PS device.

On the other hand, when the opening areas of the variable throttles 66 and 68 are small, they flow into and out of the power cylinder 52 when a pressure difference is generated between the hydraulic passages 54a and 54b. The amount of oil is relatively small. In this case, when a twisting torque acts on the diaphragm mechanism 44a,
The first hydraulic chamber 52a and the second hydraulic chamber 5 of the power cylinder 52
A differential pressure with 2b is unlikely to occur, and a steering characteristic with a low assist rate is realized in the PS device.

As described above, according to the PS device shown in FIG. 14, the assist ratio is lowered by increasing the aperture areas of the variable diaphragms 66 and 68, and the variable diaphragms 66 and 68 are also reduced.
The assist ratio can be increased by reducing the opening area of. Therefore, the PS device shown in FIG. 14 can also realize the same function as the PS device 40 shown in FIG.

In the above embodiment, the steering angle sensor 34 corresponds to the steering angle detecting means described above. Also,
By the ECU 10 executing the process of step 204, the ideal steering angle detecting means described above, the ECU 10 executing the process of step 208, and the PS device realizing the steering characteristic according to the assist ratio, The above-mentioned steering characteristic changing means are realized respectively.

By the way, in the above-described embodiment, the ideal steering angle during the execution of the braking force control is the steering angle for realizing the stable turning state, but the present invention is not limited to this, and for example, It is also possible to set a steering angle or the like that should be realized in order to travel the planned traveling route as the ideal steering angle. Such a function is, for example, an optimum steering for detecting the position of the own vehicle using a navigation system using GPS, reading lane data extending in front of the vehicle from map data, and driving in the lane with stable behavior. The function of the ideal steering angle detection means can be realized by calculating the angle.

[0092]

As described above, according to the present invention, the actual steering angle during execution of the braking force control easily matches the ideal steering angle with respect to the running state of the vehicle. When the actual steering angle and the ideal steering angle match, the vehicle behavior does not change suddenly when the braking force control is stopped, or when the road surface μ changes during execution of the braking force control. Therefore, according to the vehicle behavior control device of the present invention, the vehicle behavior can be stably maintained even when the contribution of the cornering force of the steered wheels to the vehicle behavior changes.

[Brief description of drawings]

FIG. 1 is a system configuration diagram of an embodiment of the present invention.

FIG. 2 is a system configuration diagram of an example of a power steering device used in an embodiment of the present invention.

FIG. 3 is a plan view of a vehicle turning left.

FIG. 4 is a diagram showing a two-wheel model used for vehicle behavior analysis.

FIG. 5 is a flowchart (part 1) of an example of a braking force control routine executed in an embodiment of the present invention.

FIG. 6 is a flowchart (part 2) of an example of a braking force control routine executed in an embodiment of the present invention.

FIG. 7 is a first example of a map used for executing a braking force control routine.

FIG. 8 is a second example of a map used for executing a braking force control routine.

FIG. 9 is a third example of a map used for executing a braking force control routine.

FIG. 10 is a diagram showing a relationship between an ideal steering angle δ ^* and an actual steering angle δ in two-dimensional coordinates.

FIG. 11 is a flowchart of an example of a steering characteristic setting routine executed in an embodiment of the present invention.

FIG. 12 is an example of a map used for executing a steering characteristic setting routine.

FIG. 13 is a system configuration diagram of a second example of the power steering device used in the embodiment of the present invention.

FIG. 14 is a system configuration diagram of a third example of the power steering device used in the embodiment of the present invention.

[Explanation of symbols]

10 Electronic Control Unit (ECU) 12 (12FL, 12FR, 12RL, 12RR) Wheels 14 (14FL, 14FR, 14RL, 14RR) Hydraulic Control Valve 20 Hydraulic Source Switching Valve 30 (30FL, 30FR, 30RL, 30RR) Wheel Speed Sensor 32 Steering wheel 34 Steering angle sensor 36 Lateral acceleration sensor 38 Yaw rate sensor 40 Power steering device (PS device) 44 Valve mechanism 44a Throttle mechanism 60, 64, 66, 68 Variable throttle

Claims (1)

[Claims] 1. In a vehicle behavior control device for stabilizing a vehicle behavior by controlling a braking force of each wheel when the vehicle is traveling, a steering angle detection means for detecting a steering angle, and an ideal steering angle control means for controlling the braking force. An ideal steering angle detecting means for obtaining a steering angle; and a steering characteristic changing means for changing the characteristics of the steering mechanism so that the deviation between the actual steering angle during the braking force control and the ideal steering angle becomes small. Characteristic vehicle behavior control device.