JP6314616B2 - Vehicle turning control device and vehicle turning control method - Google Patents

Description

  The present invention relates to a vehicle turning control device and a vehicle turning control method.

There is an index called a stability factor that represents the turning behavior of a vehicle.
In the prior art described in Patent Document 1, a stable stability is made by setting a target stability factor, calculating an actual stability factor, and controlling the braking / driving force of the vehicle according to these deviations. I am trying.

JP2011-236810A

However, it is only possible to grasp the static turning behavior at a certain point in time simply by calculating the actual stability factor. In other words, it is not possible to provide driving assistance for the next turning behavior simply by paying attention to the current turning behavior, so the driving assistance is performed after grasping the dynamic change characteristics of the turning behavior that changes every moment. It was hoped that.
The subject of this invention is performing the driving assistance according to the dynamic change characteristic of turning behavior.

A turning control apparatus for a vehicle according to an aspect of the present invention defines a relationship between a tire slip angle and a lateral force, which are reference characteristics, in a predetermined coordinate system as a reference characteristic line, and the tire slip angle or lateral Get power. Then, referring to the reference characteristic line, the slope of the tangent at the coordinate position corresponding to the acquired slip angle or lateral force is calculated as the first cornering power. Then, a target stability factor of the vehicle is set, and the turning traveling of the vehicle is driven and controlled in accordance with the target stability factor. Further, when the first cornering power is smaller, the target stability factor is set so that the understeer characteristic of the vehicle becomes stronger, or when the first cornering power is in a steady region larger than a predetermined threshold, the understeer characteristic of the vehicle The target stability factor is set so as to maintain a predetermined reference value.

  According to the present invention, since the slope of the tangent at the coordinate position corresponding to the slip angle or the lateral force is calculated as the first cornering power in the reference characteristic line, the first cornering power is determined next. You can see how it changes. Therefore, by using this first cornering power, a target stability factor of the vehicle is set, and driving control of the vehicle's turning travel is performed, so that driving support corresponding to the dynamic change characteristic of the turning behavior can be performed. it can.

It is a figure which shows the structure of a turning control apparatus. It is a figure which shows the relationship between the slip ratio (lambda) of a tire, and the braking / driving force Fx of a tire when the friction coefficient (mu) of a road surface differs. It is a figure which shows the relationship between the slip angle (beta) of a tire, and the lateral force Fy of a tire when the friction coefficient (micro | micron | mu) of a road surface differs. It is a figure which shows the relationship between the points Pc and Pd. It is a figure which shows the relationship between the slip angle (beta) of a tire, and the lateral force Fy of a tire. It is a map for calculating the first cornering power K1 from the lateral force Fy. It is a map for calculating the first cornering power K1 from the slip angle β. It is a map for calculating the second cornering power K2 from the lateral force Fy. It is a map for calculating the second cornering power K2 from the slip angle β. It is a map for setting the target stability factor A ^* from the first cornering power K1. It is a map for setting the distribution ratio Rd for yaw moment control from the first cornering power K1. It is a map for setting the distribution ratio Rd for deceleration control from the first cornering power K1. It is a figure which shows the structure of an electronically controlled throttle. It is a figure which shows the structure of a brake actuator. It is a flowchart which shows a traveling control process. It is a figure which shows the relationship between slip angle (beta) and lateral force Fy when a slip angle is (beta) 1. It is a figure which shows the relationship between slip angle (beta) and lateral force Fy when a slip angle is (beta) 2 larger than (beta) 1. It is a figure which shows the relationship between slip angle (beta) and lateral force Fy when a slip angle is (beta) 3 larger than (beta) 2. It is a figure which shows the relationship between the moving speed of a horizontal direction, and the moving speed of the front-back direction by acceleration / deceleration.

<< First Embodiment >>
"Constitution"
In the present embodiment, not only the current turning behavior but also dynamic change characteristics of the turning behavior that changes every moment are grasped, and then driving assistance for the next turning behavior is performed. In addition, not only when the driver is driving, but even if there is no human driving operation, the surrounding environment is recognized by the radar and camera mounted on the vehicle, and the vehicle traveling system is the main autonomous It can also be applied to those that can travel automatically (automatic traveling).

The configuration of the travel control device 11 will be described with reference to FIG.
The travel control device 11 according to the present embodiment includes a controller 12 and an actuator 13.
The controller 12 is composed of, for example, a microcomputer and inputs detection signals from various sensors such as an acceleration sensor, a wheel speed sensor, a lateral acceleration sensor, a yaw rate sensor, and a steering angle sensor.

  The acceleration sensor detects an acceleration / deceleration Gx in the longitudinal direction of the vehicle. The acceleration sensor detects, for example, the displacement of the movable electrode relative to the fixed electrode as a change in capacitance, and converts it into a voltage signal proportional to the acceleration / deceleration and outputs it to the controller 12. The controller 12 determines the acceleration / deceleration Gx from the input voltage signal. The controller 12 processes acceleration as a positive value and processes deceleration as a negative value.

The wheel speed sensor detects wheel speeds Vw _{FL to} Vw _RR of each wheel. The wheel speed sensor includes, for example, a sensor rotor that rotates together with a wheel and has a protrusion (gear pulser) formed on the circumference thereof, and a detection circuit having a pickup coil provided to face the protrusion of the sensor rotor. Then, the change in the magnetic flux density accompanying the rotation of the sensor rotor is converted into a voltage signal by the pickup coil and output to the controller 12. The controller 12 determines the wheel speeds Vw _{FL to} Vw _RR from the input voltage signal, and calculates, for example, a wheel speed average value of non-driven wheels (driven wheels) or a wheel speed average value of all wheels as a vehicle speed.

  The lateral acceleration sensor detects the lateral acceleration Gy of the vehicle. The lateral acceleration sensor detects, for example, the displacement of the movable electrode with respect to the fixed electrode as a change in capacitance, and converts it into a voltage signal proportional to the lateral acceleration and the direction, and outputs it to the controller 12. The controller 12 determines the lateral acceleration Gy from the input voltage signal. The controller 12 processes the right turn as a positive value and the left turn as a negative value.

  The yaw rate sensor detects the yaw rate γ of the vehicle body. This yaw rate sensor is provided on a body that is on a spring, for example, a vibrator made of a crystal tuning fork is vibrated by an alternating voltage, and the distortion amount of the vibrator caused by the Coriolis force at the time of angular velocity input is converted into an electrical signal. 12 is output. The controller 12 determines the yaw rate γ of the vehicle from the input electric signal. The controller 12 processes the right turn as a positive value and the left turn as a negative value.

  The steering angle sensor is composed of a rotary encoder and detects the steering angle θ of the steering shaft. The steering angle sensor detects light transmitted through the slit of the scale with two phototransistors when the disk-shaped scale rotates together with the steering shaft, and outputs a pulse signal accompanying the rotation of the steering shaft to the controller 12. . The controller 12 determines the steering angle θ of the steering shaft from the input pulse signal. The steering angle sensor processes a right turn as a positive value and a left turn as a negative value.

The controller 12 includes a turning characteristic estimation unit 14, a target behavior setting unit 15, and a drive control unit 16.
The turning characteristic estimation unit 14 includes a tire lateral force estimation unit 21, a friction coefficient estimation unit 22, a correction unit 23, a reference characteristic defining unit 24, a first cornering power calculation unit (K1 calculation unit) 25, a first One stability factor calculation unit (A1 calculation unit) 26, a second cornering power calculation unit (K2 calculation unit) 27, a second stability factor calculation unit (A2 calculation unit) 28, and a characteristic estimation unit 29 And comprising.

The tire lateral force estimator 21 estimates the front wheel lateral force Fy _F and the rear wheel lateral force Fy _R according to the lateral acceleration Gy and the yaw rate γ based on the following equations. Where m is the vehicle mass, Fy _F is the lateral force of the front wheel, Fy _R is the lateral force of the rear wheel, I is the moment of inertia, γ ′ is the differential value of γ (dγ / dt), and L _F is the vehicle center of gravity. The distance to the front wheel axle, _LR, is the distance from the vehicle center of gravity to the rear wheel axle.
mGy = Fy _F + Fy _R
Iγ ′ = Fy _F × L _F −Fy _R × _LR

The friction coefficient estimation unit 22 estimates the road friction coefficient μ in accordance with the acceleration / deceleration Gx and the wheel speeds Vw _{FL to} Vw _RR . For the estimation of the friction coefficient μ, for example, a technique described in Japanese Patent Publication No. 5206490 is used.
Here, estimation of the friction coefficient μ will be described.
FIG. 2 shows the relationship between the tire slip ratio λ and the tire braking / driving force Fx when the road surface friction coefficient μ is different.

  The horizontal axis represents the tire slip ratio λ, the vertical axis represents the tire braking / driving force Fx, and the relationship between the slip ratio λ and the braking / driving force Fx on the road surface A where the friction coefficient is μa is indicated by a solid line. This is indicated by a characteristic line La. Further, the relationship between the slip ratio λ and the braking / driving force Fx on the road surface B where the friction coefficient is μb is indicated by a broken characteristic line Lb. Since the characteristic lines La and Lb have the same aspect ratio, their shapes are similar.

  Here, an arbitrary straight line that passes through the coordinate origin [0, 0] and intersects with each of the characteristic lines La and Lb is indicated by an alternate long and short dash line Ls. The point where the straight line Ls and the characteristic line La intersect is Pa, the linear distance connecting the coordinate origin [0, 0] and the point Pa is a1, and the distance in the vertical axis direction from the coordinate origin [0, 0] to the point Pa Is a2, and the horizontal distance from the coordinate origin [0, 0] to the point Pa is a3. The point where the straight line Ls and the characteristic line Lb intersect is Pb, the straight line distance connecting the coordinate origin [0, 0] and the point Pb is b1, and the vertical axis direction from the coordinate origin [0, 0] to the point Pb Is b2, and the horizontal distance from the coordinate origin [0, 0] to the point Pb is b3.

  If the ratio (Fx / λ) between the slip ratio λ and the braking / driving force Fx is the same at the point Pa and the point Pb, the ratio between μa and μb (μa / μb) is the ratio between a1 and b1 ( a1 / b1). Further, assuming that the ratio (Fx / λ) between the slip ratio λ and the braking / driving force Fx is the same at the points Pa and Pb, the ratio (μa / μb) between μa and μb is the difference between a2 and b2. It becomes equal to the ratio (a2 / b2). Further, if the ratio (Fx / λ) of the slip ratio λ and the braking / driving force Fx is the same at the points Pa and Pb, the ratio (μa / μb) between μa and μb is the difference between a3 and b3. It becomes equal to the ratio (a3 / b3).

  This can be proved geometrically as follows. That is, a triangle having three sides of a1, a2, and a3 is similar to a triangle having three sides of b1, b2, and b3. Therefore, the ratio of a1 and b1, the ratio of a2 and b2, and the ratio of a3 and b3 are all equal (a1: b1 = a2: b2 = a3: b3). The ratio between a2 and b2 (a2 / b2) and the ratio between a3 and b3 (a3 / b3) are equal to the ratio between μa and μb (μa / μb).

As described above, if any one of the ratio between a1 and b1, the ratio between a2 and b2, and the ratio between a3 and b3 can be known, the ratio between μa and μb can be known. Therefore, for example, when estimating the friction coefficient μb of the road surface B, as shown in the following formula, any one of the ratio of a1 and b1, the ratio of a2 and b2, and the ratio of a3 and b3 is calculated as follows. By multiplying the friction coefficient μa of A, the friction coefficient μb of the road surface B can be estimated.
μb = (b1 / a1) × μa
μb = (b2 / a2) × μa
μb = (b3 / a3) × μa

Therefore, the friction coefficient estimation unit 22 uses, for example, a dry pavement road surface as a reference road surface A, and stores the friction coefficient μa and the characteristic line La in advance. Then, the current traveling road surface is set as the estimation target road surface B, and the current slip ratio λ and braking / driving force Fx are acquired and set as a point Pb. At this time, the slip ratio λ corresponding to the horizontal axis coordinate of the point Pb is b3, the braking / driving force Fx corresponding to the vertical axis coordinate of the point Pb is b2, and b1 is obtained from these b3 and b2 (b1 = √ {b3 ² + b2 ² }). A point where a straight line Ls connecting the coordinate origin [0, 0] and the point Pb intersects with the characteristic line La is defined as a point Pa. From the coordinates of this point Pa, a2 and a3 are obtained, and a1 is obtained from these a2 and a3 (a1 = √ {a3 ² + a2 ² }). Then, the friction coefficient μb of the road surface B is estimated using any one of the ratio of a1 and b2, the ratio of a2 and b2, the ratio of a3 and b3, and the friction coefficient μa of the road surface A.

For convenience, the friction coefficient of the reference and .mu.a, has been described a friction coefficient estimated as .mu.b, the friction coefficient of the reference is expressed as mu _R, the estimated friction coefficient is expressed as mu _E.
The braking / driving force Fx is determined according to the acceleration / deceleration Gx, the wheel speeds Vw _{FL to} Vw _RR , the motor current value, the motor rotation speed, the reduction ratio, etc. for an electric vehicle. In the case of an engine vehicle, the acceleration / deceleration Gx, the wheel speeds Vw _{FL to} Vw _RR , the engine speed, the reduction ratio, and the like are obtained.
The slip ratio λ is obtained according to the wheel speeds Vw _{FL to} Vw _RR and the vehicle speed (body speed) V.
The above is the description of the friction coefficient estimation unit 22.

The correction unit 23 corrects the lateral force Fy of the tire according to the road surface friction coefficient μ.
Here, correction of the lateral force Fy will be described.
FIG. 3 shows the relationship between the tire slip angle β and the tire lateral force Fy when the road surface friction coefficient μ is different.
The coordinate abscissa represents the tire slip angle β, the coordinate abscissa represents the tire lateral force Fy, and the relationship between the slip angle β and the lateral force Fy on the road surface C where the friction coefficient is μc = 1.0, This is indicated by a solid characteristic line Lc. The relationship between the slip angle β and the lateral force Fy on the road surface D where the friction coefficient is μd = 0.5 is indicated by a broken characteristic line Ld. Further, the relationship between the slip angle β and the lateral force Fy on the road surface E where the friction coefficient is μ _E = 0.2 is indicated by a broken characteristic line Le. Since the characteristic lines Lc, Ld, and Le have the same aspect ratio, their shapes are similar.

  Any of the characteristic lines Lc, Ld, and Le transitions from linear to nonlinear as the slip angle β increases from zero. That is, while the slip angle β is within a predetermined range from 0, the increase rate of the lateral force Fy is constant with respect to the increase of the slip angle β, that is, the slope of the characteristic line is constant. When the slip angle β exceeds a predetermined range, the increase rate of the lateral force Fy gradually decreases with an increase in the slip angle β. That is, the inclination of the characteristic line decreases and finally becomes substantially zero, that is, substantially parallel to the coordinate horizontal axis. When the vertical load of the tire is W, the lateral force Fy is saturated at μ × W. Thus, the relationship between the tire slip angle β and the lateral force Fy includes a linear region and a non-linear region.

  Here, an arbitrary straight line that passes through the coordinate origin [0, 0] and intersects with the characteristic lines Lc, Ld, and Le is indicated by an alternate long and short dash line Ls. A point where the straight line Ls and the characteristic line Lc intersect is Pc, and a tangent to the characteristic line Lc at the point Pc is LTc. Further, a point where the straight line Ls and the characteristic line Ld intersect is Pd, and a tangent line of the characteristic line Ld at the point Pd is LTd. In addition, a point where the straight line Ls and the characteristic line Le intersect is defined as Pe, and a tangent line of the characteristic line Le at the point Pe is defined as LTe. As described above, since the characteristic lines Lc, Ld, and Le are similar, the tangents LTc, LTd, and LTe all have the same inclination, that is, become parallel. Therefore, for example, in order to estimate the inclination of the tangent LTd, the inclination of the tangent LTc or the tangent LTe may be estimated.

Here, the relationship between the points Pc and Pd is shown in FIG.
A straight line distance along the straight line Ls connecting the coordinate origin [0, 0] and the point Pc is c1, and a distance in the vertical axis direction from the coordinate origin [0, 0] to the point Pc is c2, and the coordinate origin [0, The horizontal distance from 0] to the point Pc is c3. Further, the linear distance along the straight line Ls connecting the coordinate origin [0, 0] and the point Pd is d1, the distance in the vertical axis direction from the coordinate origin [0, 0] to the point Pd is d2, and the coordinate origin [ The distance in the horizontal direction from [0, 0] to the point Pd is d3. Since the characteristic lines Lc and Ld have the same aspect ratio, their shapes are similar.

  If the ratio of the slip angle β and the lateral force Fy (Fy / β) is the same at the point Pc and the point Pd, the ratio of μc to μd (μc / μd) is the ratio of c1 to d1 (c1 / D1). If the ratio of the slip angle β and the lateral force Fy (Fy / β) is the same at the point Pc and the point Pd, the ratio of μc to μd (μc / μd) is the ratio of c2 to d2. It becomes equal to (c2 / d2). If the ratio of the slip angle β and the lateral force Fy (Fy / β) is the same at the points Pc and Pd, the ratio of μc to μd (μc / μd) is the ratio of c3 to d3. It becomes equal to (c3 / d3).

  This can be proved geometrically as follows. That is, a triangle having three sides of c1, c2, and c3 is similar to a triangle having three sides of d1, d2, and d3. Therefore, the ratio between c1 and d1, the ratio between c2 and d2, and the ratio between c3 and d3 are all equal (c1: d1 = c2: d2 = c3: d3). The ratio between c2 and d2 (c2 / d2) and the ratio between c3 and d3 (c3 / d3) are equal to the ratio between μc and μd (μc / μd).

Thus, if the ratio between μc and μd can be known, the ratio between c1 and d1, the ratio between c2 and d2, and the ratio between c3 and d3 can be known. Therefore, for example, when estimating the vertical coordinate c2 of the point Pc having the same inclination as the point Pd, the vertical coordinate d2 of the point Pd is multiplied by the ratio of μc and μd as shown in the following equation. Thus, the vertical coordinate c2 of the point Pc can be estimated.
c2 = d2 × (μc / μd)
Also, when estimating the horizontal coordinate c3 of the point Pc having the same inclination as the point Pd, the horizontal axis coordinate d3 of the point Pd is multiplied by the ratio of μc and μd as shown in the following equation. The horizontal axis coordinate c3 of the point Pc can be estimated.
c3 = d3 × (μc / μd)

Therefore, the correction unit 23 obtains the coordinates [c3 or c2] of the point Pc where the inclination of the tangent line is the same as the point Pd from the coordinates [d3 or d2] of the point Pd. That is, using the ratio of μc and μd, d3 is corrected to c3, or d2 is corrected to c2.
Therefore, for example, a dry paved road surface is set as a reference road surface C, and the friction coefficient μc is stored in advance. Then, the current traveling road surface is set as the road surface D, and the friction coefficient μd and the current lateral force Fy are acquired. At this time, the lateral force Fy corresponding to the vertical coordinate of the point Pd is d2. A point where the straight line Ls connecting the coordinate origin [0, 0] and the point Pd intersects the characteristic line Lc is defined as a point Pc, and a lateral force Fy corresponding to the vertical coordinate c2 of the point Pc is obtained. That is, as shown in the following equation, the lateral forces Fy _F and Fy _R are corrected by multiplying the current lateral forces Fy _F and Fy _R by the ratio (μc / μd) of μc and μd.
Fy _F ← Fy _F × (μc / μd)
Fy _R ← Fy _R × (μc / μd)

Although the configuration for correcting d2 to c2 has been described here, the configuration is not limited to this, and a configuration in which d3 is corrected to c3 may be used. That is, the current slip angle β is acquired as the horizontal coordinate d3 of the point Pd. A point where the straight line Ls connecting the coordinate origin [0, 0] and the point Pd intersects the characteristic line Lc is defined as a point Pc, and a slip angle β corresponding to the horizontal coordinate c3 of the point Pc is obtained. That is, as shown in the following equation, the slip angles β _F and β _R may be corrected by multiplying the current slip angle β by the ratio (μc / μd) of μc and μd.
β _F ← β _F × (μc / μd)
β _R ← β _R × (μc / μd)
For convenience, the friction coefficient of the reference and [mu] c, has been described a friction coefficient estimated as [mu] d, the friction coefficient of the reference is expressed as mu _R, the estimated friction coefficient is expressed as mu _E. Note that the friction coefficient μ _E used in the correction unit 23 is an average value of the friction coefficient μ _E estimated by the friction coefficient estimation unit 22 over a predetermined period.
The above is the description of the correction unit 23.

The reference characteristic defining unit 24 defines, as a reference characteristic line Lr, a relationship between the tire slip angle β and the lateral force Fy, which are reference characteristics, in a predetermined coordinate system.
Here, the reference characteristic line Lr will be described.
FIG. 5 shows the relationship between the tire slip angle β and the tire lateral force Fy.
Taking a slip angle β of the tire coordinate horizontal axis, taking a lateral force Fy of the tire coordinate vertical axis, in the reference road surface of the reference friction coefficient mu _R, the relationship between the slip angle β and the lateral force Fy of the tire as a reference characteristic Is defined as a solid reference characteristic line Lr. Here, although one reference characteristic line Lr will be described, it is assumed that the reference characteristic line Lr _F for the front wheels is actually defined and the reference characteristic line Lr _R for the rear wheels is defined.

  The reference characteristic line Lr transitions from linear to non-linear when the slip angle β increases from zero. That is, while the slip angle β is in a predetermined range from 0, the increase rate of the lateral force Fy is constant with respect to the increase of the slip angle β, that is, the inclination of the reference characteristic line Lr is constant. When the slip angle β exceeds a predetermined range, the increase rate of the lateral force Fy gradually decreases with an increase in the slip angle β. That is, the inclination of the reference characteristic line Lr decreases and finally becomes substantially zero, that is, substantially parallel to the coordinate horizontal axis. Thus, the relationship between the tire slip angle β and the lateral force Fy includes a linear region and a non-linear region.

Here, the cornering power K will be described.
The cornering power K is represented by the ratio (Fy / β) of the lateral force Fy to the tire slip angle β, that is, the inclination.
An arbitrary point P is taken in the non-linear region on the reference characteristic line Lr, and the tangent at this point P is assumed to be Lk1. Further, if the coordinates at which the tangent Lk1 and the vertical axis intersect are [0, Fy1], the coordinates of the point P are [β0, Fy2], and the difference between Fy2 and Fy1 is ΔFy (= Fy2−Fy1), the tangent Lk1 Is expressed by ΔFy / β0. The slope ΔFy / β0 of this tangent Lk1 is taken as the first cornering power K1. If the straight line connecting the coordinate origin [0, 0] and the point P is Lk2, the slope of the straight line Lk2 is represented by Fy2 / β0, and the slope Fy2 / β0 of the straight line Lk2 is set as the second cornering power K2. And Further, if the tangent of the linear region in the reference characteristic line Lr is Lk3 and the vertical axis coordinate corresponding to the horizontal axis β0 of this tangent Lk3 is Fy3, the slope of the tangent Lt3 is expressed by Fy3 / β0. The slope Fy3 / β0 of the tangent Lk3 is set as a third cornering power K3.

Each cornering power K1 to K3 will be described.
The third cornering power K3 indicates the tire characteristics in the linear region. That is, since the cornering power is when the relationship of the lateral force Fy to the slip angle β is in a steady state, it can be said to be a steady cornering power.
The second cornering power K2 indicates static tire characteristics at a certain point in time. That is, since it becomes the cornering power when the slip angle β and the lateral force Fy at a certain point in time are maintained, it can be said to be a static cornering power. It can also be said to be quasi-stationary cornering power in the sense that it does not change constantly.
The first cornering power K1 indicates dynamic tire characteristics at a certain point in time. That is, it can be said that the cornering power is dynamic when the slip angle β and the lateral force Fy change from a certain point in time as the vehicle speed V or the steering angle θ increases. Moreover, it can be said that it is a transitional cornering power in the sense of being in a transitional state.

Each cornering power K1 to K3 will be further described.
When the cornering power as a reference is K0 [N / rad], the cornering power Kb [N / rad] considering the wheel load can be expressed by the following equation. Here, Ww is a wheel load, and W0 is a vehicle weight per wheel (W / 4) [kg].
Kb = K0 × (Ww / W0)
Further, the cornering force Fc can be expressed by the following formula.
Fc = − {(Kb × β−2 × μ × Ww) ² / (4 × μ × Ww)} + (μ × Ww)
Further, when the cornering force Fc and the slip angle β [deg] are used, the cornering power Kf can be expressed by the following equation.
Kf = Fc / β
Further, when the cornering power Kb, the lateral acceleration dy [g], and the friction coefficient μ between the road surface and the tire are used, the cornering power Km can be expressed by the following equation.
Km = Kb × √ {1- (dy / μ)}

The cornering power Km approximates the first cornering power K1, the cornering power Kf approximates the second cornering power K2, and the cornering power Kb approximates the third cornering power K3.
Therefore, the cornering power Km may be estimated and used instead of the first cornering power K1. Further, the cornering power Kf may be estimated and used instead of the second cornering power K2. Further, the cornering power Kb may be estimated and used instead of the third cornering power K3.
The above is the description of the reference characteristic defining unit 24.

The first cornering power calculation unit 25 refers to the reference characteristic line Lr, and calculates the slope of the tangent line Lk1 at the coordinate position corresponding to the tire lateral force Fy as the first cornering power K1. Here, although one cornering power K1 will be described, it is assumed that the first cornering power K1 _F at the front wheels is actually calculated and the first cornering power K1 _R at the rear wheels is calculated.
Here, based on the reference characteristic line Lr, a map for calculating the first cornering power K1 according to the lateral force Fy is prepared. With reference to this map, the first cornering power K1 is determined according to the lateral force Fy. A cornering power K1 is calculated.

A map for calculating the first cornering power K1 from the lateral force Fy is shown in FIG.
This map is set such that the lateral force Fy is taken on the coordinate horizontal axis, the first cornering power K1 is taken on the coordinate vertical axis, and the first cornering power K1 is reduced as the lateral force Fy is increased. According to the reference characteristic line Lr, the slope of the tangent line Lk1 is the same as the slope of the tangent line Lk3 while in the linear region. Therefore, when the lateral force Fy is 0, the first cornering power K1 is the third cornering power K1. Cornering power K3. Further, according to the reference characteristic line Lr, the slope of the tangent line Lk1 is finally substantially 0, that is, substantially parallel to the coordinate horizontal axis, so that the first cornering is performed when the lateral force Fy reaches the maximum value Fy _MAX. The power K1 becomes zero. When the vertical load of the tire is W, the maximum value Fy _MAX is represented by μ × W.

Here, the configuration for calculating the first cornering power K1 according to the lateral force Fy is described, but the present invention is not limited to this, and the first cornering power K1 is calculated according to the slip angle β. It is good also as a structure.
That is, a map for calculating the first cornering power K1 according to the slip angle β is prepared based on the reference characteristic line Lr, and the first cornering according to the slip angle β is prepared by referring to this map. The power K1 is calculated.

A map for calculating the first cornering power K1 from the slip angle β is shown in FIG.
This map is set such that the horizontal axis represents the slip angle 、, the coordinate vertical axis represents the first cornering power K1, and the larger the slip angle Β, the smaller the first cornering power K1. According to the reference characteristic line Lr, the slope of the tangent line Lk1 is the same as the slope of the tangent line Lk3 while in the linear region. Therefore, when the slip angle ０ is 0, the first cornering power K1 is the third cornering power K1. Cornering power K3. Further, according to the reference characteristic line Lr, since substantially 0, that is, substantially parallel to the coordinate abscissa to the inclination of the tangent line Lk1 Eventually, when the slip angle beta reaches the maximum value beta _MAX, first cornering The power K1 becomes zero.

The first stability factor calculation unit 26 calculates the first stability factor A1 using the first cornering powers K1 _F and K1 _R as shown in the following equation. Here, m vehicle mass, L is the wheel base, L _F is the distance from the vehicle center of gravity to the front wheel axle, is L _R is the distance to the rear wheel axle from the vehicle center of gravity.
^{A1 = - (m / 2L 2} ) × {(L F × K1 F -L R × K1 R) / (K1 F × K1 R)}
The first stability factor A1 indicates a dynamic change characteristic of the turning behavior at a certain time point. In other words, it becomes a stability factor when the turning behavior changes from a certain point as the vehicle speed V or the steering angle θ increases. It can also be said to be a transient stability factor in the sense of being in a transient state.

The second cornering power calculation unit 27 refers to the reference characteristic line Lr, determines the coordinate position corresponding to the lateral force Fy of the tire and the slope of the straight line Lk2 connecting the coordinate origin [0, 0] as the second cornering power. Calculated as K2. Here is a description of one of the cornering power K2, actually shall calculates the second cornering power K2 _F in the front wheels, and calculates the second cornering power K2 _R at the rear wheels.
Here, based on the reference characteristic line Lr, a map for calculating the second cornering power K2 according to the lateral force Fy is prepared, and this map is referred to and the second cornering power Ky is determined according to the lateral force Fy. A cornering power K2 is calculated.

A map for calculating the second cornering power K2 from the lateral force Fy is shown in FIG.
This map is set so that the lateral force Fy is taken on the coordinate horizontal axis, the second cornering power K2 is taken on the coordinate vertical axis, and the second cornering power K2 is reduced as the lateral force Fy is increased. According to the reference characteristic line Lr, the slope of the straight line Lk2 is the same as the slope of the tangent line Lk3 while in the linear region. Therefore, when the lateral force Fy is 0, the second cornering power K2 is the third cornering power K2. Cornering power K3. Further, according to the reference characteristic line Lr, since one end of the straight line Lk2 is the coordinate origin [0, 0], and the inclination thereof is always greater than 0, the second cornering power K2 is always greater than 0.

Here, the configuration for calculating the second cornering power K2 according to the lateral force Fy has been described. However, the present invention is not limited to this, and the second cornering power K2 is calculated according to the slip angle β. It is good also as a structure.
That is, a map for calculating the second cornering power K2 according to the slip angle β is prepared based on the reference characteristic line Lr, and the second cornering according to the slip angle β is prepared by referring to this map. The power K2 is calculated.

A map for calculating the second cornering power K2 from the slip angle β is shown in FIG.
This map is set so that the slip angle β is taken on the coordinate horizontal axis, the second cornering power K2 is taken on the coordinate vertical axis, and the second cornering power K2 becomes smaller as the slip angle β becomes larger. According to the reference characteristic line Lr, the slope of the straight line Lk2 is the same as the slope of the tangent line Lk3 while in the linear region, so when the slip angle β is 0, the second cornering power K2 is the third cornering power K2. Cornering power K3. Further, according to the reference characteristic line Lr, since one end of the straight line Lk2 is the coordinate origin [0, 0], and the inclination thereof is always greater than 0, the second cornering power K2 is always greater than 0.

The second stability factor calculation unit 28 calculates the second stability factor A2 using the second cornering powers K2 _F and K2 _R as shown in the following equation. Here, m vehicle mass, L is the wheel base, L _F is the distance from the vehicle center of gravity to the front wheel axle, is L _R is the distance to the rear wheel axle from the vehicle center of gravity.
^{A2 = - (m / 2L 2} ) × {(L F × K2 F -L R × K2 R) / (K2 F × K2 R)}
The second stability factor A1 indicates a static turning behavior at a certain time point. That is, since it becomes a stability factor when the vehicle speed V or the steering angle θ at a certain point in time is maintained, it can be said to be a static stability factor. It can also be said to be a quasi-stationary stability factor in the sense that it does not change constantly.

  The characteristic estimation unit 29 estimates a dynamic change characteristic of the turning behavior according to the sign and the size of the first stability factor A1. That is, whether the turning behavior that changes as the vehicle speed V or the steering angle θ increases is an understeer characteristic (US characteristic), an oversteer characteristic (OS characteristic), or a neutral steer characteristic (NS characteristic). Is estimated. Specifically, when the first stability factor A1 is a positive value, the turning behavior that changes as the vehicle speed V or the steering angle θ increases is an understeer characteristic, and the first stability factor A1 It is estimated that the larger the absolute value, the stronger the understeer characteristic. Further, when the first stability factor A1 is a negative value, the turning behavior that changes as the vehicle speed V or the steering angle θ increases is an oversteer characteristic, and the absolute value of the first stability factor A1. It is estimated that the larger the is, the stronger the oversteer characteristic is. When the first stability factor A1 is 0, it is estimated that the turning behavior that changes as the vehicle speed V or the steering angle θ increases is the neutral steer characteristic.

The characteristic estimation unit 29 estimates a static turning behavior according to the sign and size of the second stability factor A2. That is, whether the turning behavior when the vehicle speed V or the steering angle θ at a certain time is maintained is an understeer characteristic (US characteristic), an oversteer characteristic (OS characteristic), or a neutral steer (NS characteristic). Is estimated. Specifically, when the second stability factor A2 is a positive value, the turning behavior when the vehicle speed V or the steering angle θ is maintained is an understeer characteristic, and the absolute value of the second stability factor A2 It is estimated that the larger the is, the stronger the understeer characteristic is. When the second stability factor A2 is a negative value, the turning behavior when the vehicle speed V or the steering angle θ is maintained is an oversteer characteristic, and the absolute value of the second stability factor A2 is large. It is estimated that the oversteer characteristic is stronger. When the second stability factor A2 is 0, it is estimated that the turning behavior when the vehicle speed V or the steering angle θ is maintained is the neutral steer characteristic.
The above is the configuration of the turning characteristic estimation unit 14.

The target behavior setting unit 15 sets, for example, a target stability factor A ^* as the target behavior according to the first cornering power K1. The target stability factor A ^* indicates a static target turning behavior at a certain time point. That is, since it becomes the target stability factor when the vehicle speed V or the steering angle θ at a certain time is maintained, it can be said to be a static target stability factor. It can also be said to be a quasi-stationary target stability factor in the sense that it does not change constantly. The target stability factor A ^* indicates a static target turning behavior according to its sign and size.

That is, the target turning behavior when the vehicle speed V or the steering angle θ at a certain time point is maintained is an understeer characteristic (US characteristic), an oversteer characteristic (OS characteristic), or a neutral steer (NS characteristic). Indicate. Specifically, when the target stability factor A ^* is a positive value, the target turning behavior when the vehicle speed V or the steering angle θ is maintained is an understeer characteristic, and the absolute value of the target stability factor A ^* is A larger value indicates stronger understeer characteristics. When the target stability factor A ^* is a negative value, the target turning behavior when the vehicle speed V or the steering angle θ is maintained has an oversteer characteristic, and the larger the absolute value of the target stability factor A ^* is, the larger the target stability factor A ^* is. This indicates that the oversteer characteristic is strong. Further, when the target stability factor A ^* is 0, it indicates that the target turning behavior when the vehicle speed V or the steering angle θ is maintained is the neutral steer characteristic.

Here, a map for setting the target stability factor A ^* is prepared according to the first cornering power K1, and the target stability factor is determined according to the first cornering power K1 by referring to this map. Set A ^* . As the first cornering power K1, one of the first cornering power K1 _{F at} the front wheels and the first cornering power K1 _{R at} the rear wheels is used.
A map for setting the target stability factor A ^* from the first cornering power K1 is shown in FIG.
This map has the first cornering power K1 on the coordinate horizontal axis and the target stability factor A ^* on the coordinate vertical axis. For the first cornering power K1, a value Kt1 that is greater than 0 and near 0 and a value Kt2 that is greater than Kt1 are determined in advance. A range in which the first cornering power K1 is larger than Kt2 is defined as a steady region, and a range in which the first cornering power K1 is smaller than Kt2 is defined as an unsteady region. Further, the target stability factor A ^* is predetermined and the maximum value A _MAX is greater than 0, the reference value A _R is large and near zero values than 0, the.

According to this map, when the first cornering power K1 is in the large range than Kt2, the target stability factor A ^* maintains the reference value A _R. Further, when the first cornering power K1 is in the range of Kt1 from Kt2, the more first cornering power K1 is small, increases in the range of the maximum value A _MAX target stability factor A ^* from the reference value A _R. Further, when the first cornering power K1 is in the range of 0 from Kt1, the target stability factor A ^* maintains the maximum value A _MAX . Thus, the target stability factor A ^* increases in the positive (+) region as the first cornering power K1 decreases.

The drive control unit 16 includes a control amount setting unit 17, a distribution ratio setting unit 18, and a drive unit 19.
Control amount setting unit 17, a second stability factor A2, and using the target stability factor A ^*, target yaw rate gamma ^* target yaw moment for achieving Mz ^*, and for realizing the target vehicle speed V ^* Target deceleration Gx ^* is set as a target control amount.
Here, as shown in the following equation, the target yaw rate γ ^* is set according to the target stability factor A ^* , the vehicle speed V, and the steering angle θ. Here, L is a wheel base. Note that the target yaw rate γ ^* is preferably low-pass filtered.
γ ^* = {V / (1 + A ^* × V ² )} × (θ / L)
Then, as shown in the following equation, the target yaw moment Mz ^* is set according to the deviation (γ ^* −γ) between the target yaw rate γ ^* and the yaw rate γ. Here, Δt is a unit time. The target yaw moment Mz ^* is a positive value for a right turn and a negative value for a left turn.
Mz ^* = (γ ^* −γ) / Δt

Further, as shown in the following equation, the target vehicle speed V ^* is set according to the second stability factor A2, the target stability factor A ^* , and the vehicle speed V.
V ^* = {√ (A ^* / A2)} × V
Then, as shown in the following equation, the target deceleration Gx ^* is set according to the target vehicle speed V ^* and the deviation (V ^* −V) between the vehicle speed V. Here, Δt is a unit time.
Gx ^* = (V ^* −V) / Δt
The above is the description of the control amount setting unit 17.

The distribution ratio setting unit 18 sets a distribution ratio Rd between yaw moment control based on the target yaw moment Mz ^* and deceleration control based on the target deceleration Gx ^* . The distribution ratio Rd is a ratio of the yaw moment control when the entire yaw moment control and deceleration control is 1.0. Therefore, the rate of deceleration control is “1.0-Rd”.
Here, a map for setting the distribution ratio R is prepared according to the first cornering power K1, and the distribution ratio Rd is set according to the first cornering power K1 at the front wheels with reference to this map. To do. As the first cornering power K1, one of the first cornering power K1 _{F at} the front wheels and the first cornering power K1 _{R at} the rear wheels is used.

A map for setting the distribution ratio Rd for yaw moment control from the first cornering power K1 is shown in FIG.
This map takes the first cornering power K1 on the coordinate horizontal axis and the distribution ratio Rd on the coordinate vertical axis. For the first cornering power K1, a value Kt3 larger than 0 and a value Kt4 larger than Kt3 are determined in advance. According to this map, when the first cornering power K1 is in a range greater than Kt4, the distribution ratio Rd is maintained at 1.0. Further, when the first cornering power K1 is in the range from Kt4 to Kt3, the distribution ratio Rd is smaller in the range from 0 to 1.0 as the first cornering power K1 is smaller. Further, when the first cornering power K1 is in the range of 0 from Kt3, the distribution ratio Rd is maintained at 0. Thus, the smaller the first cornering power K1, the smaller the distribution ratio Rd.

Here, a configuration has been described in which the ratio of yaw moment control when the entire yaw moment control and deceleration control is 1.0 is set as the distribution ratio Rd, but the present invention is not limited to this. A configuration may be adopted in which the ratio of deceleration control when the entire yaw moment control and deceleration control is 1.0 is set as the distribution ratio Rd. In this case, the rate of yaw moment control is “1.0-Rd”.
Here, a map for setting the distribution ratio R is prepared according to the first cornering power K1, and the distribution ratio Rd is set according to the first cornering power K1 at the front wheels with reference to this map. To do.

FIG. 12 shows a map for setting the distribution ratio Rd for deceleration control from the first cornering power K1.
This map takes the first cornering power K1 on the coordinate horizontal axis and the distribution ratio Rd on the coordinate vertical axis. For the first cornering power K1, a value Kt3 larger than 0 and a value Kt4 larger than Kt3 are determined in advance. According to this map, when the first cornering power K1 is in a range greater than Kt4, the distribution ratio Rd is maintained at 0. Further, when the first cornering power K1 is in the range of Kt4 to Kt3, the distribution ratio Rd increases in the range of 0 to 1.0 as the first cornering power K1 decreases. Further, when the first cornering power K1 is in the range of 0 from Kt3, the distribution ratio Rd is maintained at 1.0. Thus, the distribution ratio Rd increases as the first cornering power K1 decreases.

The first cornering power K1 corresponds to a tire grip margin rate. The tire grip margin rate α can be expressed by the following equation. Here, 0.25 and 0.8 are reference values on the dry paved road surface.
α = 1−√ {(Gx ² /0.25 ² ) + (Gy ² /0.8 ² )}
Therefore, the tire grip margin rate α may be estimated and used instead of the first cornering power K1. For example, when the margin rate α is in a range larger than 0.5, the yaw moment control distribution ratio Rd is maintained at 1.0. When the margin rate α is in the range of 0.5 to 0.1, the yaw moment control distribution ratio Rd is smaller in the range of 1.0 to 0 as the margin rate α is smaller. When the margin rate α is in the range of 0.1 to 0, the yaw moment control distribution ratio Rd is maintained at 0. Thus, the yaw moment control distribution ratio Rd may be reduced as the tire grip margin rate α is smaller.
The above is the description of the distribution ratio setting unit 18.

The drive unit 19 executes yaw moment control and deceleration control in accordance with the yaw moment control distribution ratio Rd. That is, as shown in the following equation, the target yaw moment Mz ^* is corrected by multiplying the target yaw moment Mz ^* by the distribution ratio Rd, and the target deceleration Gx ^* is multiplied by (1.0−Rd). As a result, the target deceleration Gx ^* is corrected. In order to realize the corrected target yaw moment Mz ^* and target deceleration Gx ^* , drive control of the actuator 13 is performed.
Mz ^* ← Mz ^* × Rd
Gx ^* ← Gx ^* × (1.0−Rd)

When the ratio of deceleration control is set as the distribution ratio Rd, the target yaw moment Mz is multiplied by (1.0−Rd) by multiplying the target yaw moment Mz ^* as shown in the following equation. ^* Shall be corrected. Further, the target deceleration Gx ^* is corrected by multiplying the target deceleration Gx ^* by the distribution ratio Rd. In order to realize the corrected target yaw moment Mz ^* and target deceleration Gx ^* , drive control of the actuator 13 is performed.
Mz ^* ← Mz ^* × (1.0-Rd)
Gx ^* ← Gx ^* × Rd

The target yaw moment Mz ^* is realized by, for example, braking force control that generates a braking force difference between the left and right wheels.
Specifically, when the absolute value of the target yaw moment Mz ^* is smaller than a predetermined set value Ms, the braking force difference ΔP _F between the left and right wheels at the front wheel and the left and right wheels at the rear wheel are calculated as shown in the following formula. to set the braking force difference ΔP _R. d is a tread.
ΔP _F = 0
ΔP _R = | Mz ^* | × 2 / d
Thus, when the absolute value of the target yaw moment Mz ^* is less than the set value Ms, a braking force difference is generated between the left and right wheels only for the rear wheels.

Further, when the absolute value of the target yaw moment Mz ^* is equal to or larger than the set value Ms, the braking force difference ΔP _F between the left and right wheels in the front wheel and the braking force difference ΔP _R between the left and right wheels in the rear wheel as shown in the following equations. Set.
ΔP _F = (| Mz ^* | −Ms) × 2 / d
ΔP _R = Ms × 2 / d
Thus, when the absolute value of the target yaw moment Mz ^* is equal to or greater than the set value Ms, a braking force difference is generated between both the front wheels and the rear wheels.

The target deceleration Gx ^* is realized by driving force control and braking force control.
That is, if the rotational drive source is an engine, the target deceleration Gx ^* is generated by engine braking, and if the rotational drive source is a motor, the target deceleration Gx ^* is generated by regenerative braking. The driving force control includes a shift control for shifting down the gear ratio.
Specifically, the deceleration generated by the braking force difference [Delta] P _F and [Delta] P _R and Gm, as shown below, by subtracting the deceleration Gm from target deceleration Gx ^*, corrects the target deceleration Gx ^* To do.
Gx ^* ← Gx ^* -Gm

The maximum deceleration obtained by the driving force control of the rotational drive source is Gd _MAX . When this maximum deceleration Gd _MAX is larger than the corrected target deceleration Gx ^* , the target is obtained only by the driving force control of the rotational drive source. A deceleration Gx ^* is generated. On the other hand, when the maximum deceleration Gd _MAX is smaller than the target deceleration Gx ^* , the difference (Gx ^* −Gd _MAX ) is generated by the braking force control. Thus, the driving force control is executed with priority, and the braking force control is executed to compensate for the shortage.
The configuration for generating the target yaw moment Mz ^{* based} on the difference in braking force between the left and right wheels has been described. However, the present invention is not limited to this. The moment Mz ^* may be generated. Further, in a vehicle equipped with electric power steering or steering-by-wire, the target yaw moment Mz ^* may be generated by applying steering torque.
The above is the configuration of the drive control unit 16.

The actuator 13 includes, for example, a driving force control device 30 and a brake control device 50.
The driving force control device 30 controls the driving force of the rotational drive source. For example, if the rotational drive source is an engine, the engine output (the number of revolutions and the engine torque) is controlled by adjusting the opening of the throttle valve, the fuel injection amount, the ignition timing, and the like. If the rotational drive source is a motor, the motor output (number of revolutions and motor torque) is controlled via an inverter.

As an example of the driving force control device 30, FIG. 13 shows the configuration of an electronically controlled throttle that controls the opening of the throttle valve.
A throttle shaft 32 extending in the radial direction is supported in the intake pipe 31 (for example, an intake manifold), and a disk-like throttle valve 33 having a diameter less than the inner diameter of the intake pipe 31 is supported on the throttle shaft 32. Is fixed. A throttle motor 35 is connected to the throttle shaft 32 via a speed reducer 34.

  Therefore, when the throttle motor 35 is rotated to change the rotation angle of the throttle shaft 32, the throttle valve 33 closes or opens the intake pipe 31. That is, when the surface direction of the throttle valve 33 is along the direction perpendicular to the axis of the intake pipe 31, the throttle opening becomes the fully closed position, and when the surface direction of the throttle valve 33 is along the axis direction of the intake pipe 31, The throttle opening is the fully open position. When an abnormality occurs in the throttle motor 35, the motor drive system, the accelerator sensor 36 system, the throttle sensor 39 system, etc., the throttle shaft 32 is opened in the opening direction so that the throttle valve 33 opens a predetermined amount from the fully closed position. It is mechanically energized.

  The accelerator sensor 36 has two systems, and detects a pedal opening degree PPO that is a depression amount (operation amount) of the accelerator pedal 37. The accelerator sensor 36 is a potentiometer, for example, and converts the pedal opening of the accelerator pedal 37 into a voltage signal and outputs the voltage signal to the engine controller 38. The engine controller 38 determines the pedal opening PPO of the accelerator pedal 37 from the input voltage signal. The pedal opening PPO is 0% when the accelerator pedal 37 is in the non-operating position, and the pedal opening PPO is 100% when the accelerator pedal 37 is in the maximum operating position (stroke end).

  The throttle sensor 39 has two systems and detects the throttle opening SPO of the throttle valve 33. The throttle sensor 39 is, for example, a potentiometer, and converts the throttle opening of the throttle valve 33 into a voltage signal and outputs the voltage signal to the engine controller 38. The engine controller 38 determines the throttle opening SPO of the throttle valve 33 from the input voltage signal. The throttle opening SPO is 0% when the throttle valve 33 is in the fully closed position, and the throttle opening SPO is 100% when the throttle valve 33 is in the fully open position.

The engine controller 38 normally sets the target throttle opening SPO ^* according to the pedal opening PPO, and sets the motor control amount according to the deviation ΔPO between the target throttle opening SPO ^* and the actual throttle opening SPO. Set. The motor control amount is converted into a duty ratio, and the throttle motor 35 is driven and controlled by a pulsed current value. Further, when the engine controller 38 receives a drive command from the controller 12, the engine controller 38 controls the throttle motor 35 by giving priority to the drive command. For example, when a driving command for reducing the driving force is received, the throttle motor 35 is driven and controlled by reducing the target throttle opening SPO ^* corresponding to the pedal opening PPO.
The above is the description of the driving force control device 30.

Next, the brake control device 50 will be described.
The brake control device 50 controls the braking force of each wheel. For example, the hydraulic pressure of a wheel cylinder provided in each wheel is controlled by a brake actuator used for anti-skid control (ABS), traction control (TCS), stability control (VDC: Vehicle Dynamics Control), and the like.

The configuration of the brake actuator is shown in FIG.
The brake actuator 51 is interposed between the master cylinder 52 and the wheel cylinders 53FL to 53RR.
The master cylinder 52 is a tandem type that creates two hydraulic pressures according to the driver's pedaling force. The master cylinder 52 transmits the primary side to the front left and rear right wheel cylinders 53FL and 53RR, and the secondary side transmits the right front wheel and A diagonal split system is used for transmission to the wheel cylinders 53FR and 53RL for the left rear wheel.

Each wheel cylinder 53FL to 53RR is built in a disc brake that generates a braking force by clamping a disc rotor with a brake pad, or a drum brake that generates a braking force by pressing a brake shoe against the inner peripheral surface of the brake drum. is there.
The primary side includes a first gate valve 61A, an inlet valve 62FL (62RR), an accumulator 63, an outlet valve 64FL (64RR), a second gate valve 65A, a pump 66, and a damper chamber 67.

  The first gate valve 61A is a normally open valve that can close the flow path between the master cylinder 52 and the wheel cylinder 53FL (53RR). The inlet valve 62FL (62RR) is a normally open valve that can close the flow path between the first gate valve 61A and the wheel cylinder 53FL (53RR). The accumulator 63 is communicated between the wheel cylinder 53FL (53RR) and the inlet valve 62FL (62RR). The outlet valve 64FL (64RR) is a normally closed valve that can open a flow path between the wheel cylinder 53FL (53RR) and the accumulator 63.

  The second gate valve 65A is a normally closed valve that can open a flow path that connects the master cylinder 52 and the first gate valve 61A and the accumulator 63 and the outlet valve 64FL (64RR). The pump 66 communicates the suction side between the accumulator 63 and the outlet valve 64FL (64RR), and communicates the discharge side between the first gate valve 61A and the inlet valve 62FL (62RR). The damper chamber 67 is provided on the discharge side of the pump 66, suppresses pulsation of the discharged brake fluid, and weakens pedal vibration.

Similarly to the primary side, the secondary side also has a first gate valve 61B, an inlet valve 62FR (62RL), an accumulator 63, an outlet valve 64FR (64RL), a second gate valve 65B, a pump 66, A damper chamber 67.
The first gate valves 61A and 61B, the inlet valves 62FL to 62RR, the outlet valves 64FL to 64RR, and the second gate valves 65A and 65B are two-port, two-position switching, single solenoid, and spring offset type electromagnetic operations, respectively. It is a valve. The first gate valves 61A and 61B and the inlet valves 62FL to 62RR open the flow path at the non-excited normal position, and the outlet valves 64FL to 64RR and the second gate valves 65A and 65B are at the non-excited normal position. The flow path is closed.

The accumulator 63 is a spring-type accumulator in which a compression spring faces the piston of the cylinder.
The pump 66 is a positive displacement pump such as a gear pump or a piston pump that can ensure a substantially constant discharge amount regardless of the load pressure.
With the above configuration, the primary side will be described as an example. When the first gate valve 61A, the inlet valve 62FL (62RR), the outlet valve 64FL (64RR), and the second gate valve 65A are all in the non-excited normal position. Then, the hydraulic pressure from the master cylinder 52 is transmitted as it is to the wheel cylinder 53FL (53RR) and becomes a normal brake.

  Further, even when the brake pedal is not operated, the first gate valve 61A is excited and closed while the inlet valve 62FL (62RR) and the outlet valve 64FL (64RR) are in the non-excited normal position. The second gate valve 65A is excited and opened, and the pump 66 is further driven to suck the hydraulic pressure in the master cylinder 52 through the second gate valve 65A and discharge the hydraulic pressure to the inlet valve 62FL (62RR). ) To the wheel cylinder 53FL (53RR) to increase the pressure.

  When the first gate valve 61A, the outlet valve 64FL (64RR), and the second gate valve 65A are in the non-excited normal position, if the inlet valve 62FL (62RR) is excited and closed, the wheel cylinder 53FL (53RR) ) To the master cylinder 52 and the accumulator 63 are blocked, and the hydraulic pressure of the wheel cylinder 53FL (53RR) is maintained.

  Further, when the first gate valve 61A and the second gate valve 65A are in the non-excited normal position, the inlet valve 62FL (62RR) is excited and closed, and the outlet valve 64FL (64RR) is excited and opened. The hydraulic pressure of the wheel cylinder 53FL (53RR) flows into the accumulator 63 and is reduced. The hydraulic pressure flowing into the accumulator 63 is sucked by the pump 66 and returned to the master cylinder 52.

Also on the secondary side, the normal braking, pressure increasing, holding, and pressure reducing operations are the same as the operations on the primary side, and detailed description thereof will be omitted.
The brake controller 54 controls each wheel cylinder by drivingly controlling the first gate valves 61A and 61B, the inlet valves 62FL to 62RR, the outlet valves 64FL to 64RR, the second gate valves 65A and 65B, and the pump 66. Increase, hold, or reduce the fluid pressure of 53FL to 53RR.

In the present embodiment, a diagonal split method is used in which the brake system is divided into front left / rear right and front right / rear left, but the present invention is not limited thereto. The front / rear split method may be adopted.
Further, in the present embodiment, the spring-shaped accumulator 63 is adopted, but the present invention is not limited to this, and the brake fluid extracted from each of the wheel cylinders 53FL to 53RR is temporarily stored to efficiently reduce the pressure. Therefore, any type such as a weight type, a gas compression direct pressure type, a piston type, a metal bellows type, a diaphragm type, a bladder type, and an in-line type may be used.

  In the present embodiment, the first gate valves 61A and 61B and the inlet valves 62FL to 62RR open the flow path at the non-excited normal position, and the outlet valves 64FL to 64RR and the second gate valves 65A and 65B are non-excited. Although the flow path is closed at the normal excitation position, the present invention is not limited to this. In short, since it is only necessary to open and close each valve, the first gate valves 61A and 61B and the inlet valves 62FL to 62RR open the flow path at the excited offset position, and the outlet valves 64FL to 64RR and the second gate are opened. The valves 65A and 65B may close the flow path at the excited offset position.

The brake controller 54 normally controls the hydraulic pressures of the wheel cylinders 53FL to 53RR by driving and controlling the brake actuator 51 in accordance with anti-skid control, traction control, and stability control. When the brake controller 54 receives a drive command from the controller 12, the brake controller 54 controls the brake actuator 51 by giving priority to the drive command. For example, when a drive command for increasing the pressure of a predetermined wheel cylinder among the four wheels is received, the brake actuator 51 is driven and controlled by increasing the normal target hydraulic pressure.
The above is the description of the brake control device 50.

Next, a travel control process executed by the controller 12 every predetermined time (for example, 10 msec) will be described with reference to FIG.
First, in step S101, various data are read. Specifically, acceleration / deceleration Gx, wheel speeds Vw _{FL to} Vw _RR , lateral acceleration Gy, yaw rate γ, vehicle speed V, steering angle θ, and the like are read.
In the subsequent step S102, the lateral force Fy _F of the front wheels and the lateral force Fy _{R of the} rear wheels are estimated by the processing of the tire lateral force estimating unit 21.
In step S103, the processing of the friction coefficient estimation unit 22 estimates the friction coefficient mu _E of the road surface.

In subsequent step S104, the lateral forces Fy _F and Fy _R are obtained by multiplying the lateral forces Fy _F and Fy _R by the ratio of μ _R and μ _E (μ _R / μ _E ) by the processing of the correction unit 23. to correct.
In the subsequent step S105, the first cornering power calculation unit 25 refers to the map and calculates the first cornering powers K1 _F and K1 _R according to the lateral forces Fy _F and Fy _R.
In subsequent step S106, the first stability factor A1 is calculated using the first cornering powers K1 _F and K1 _R by the processing of the first stability factor calculation unit 26.

In step S107, the processing of the second cornering power calculation unit 27, by referring to a map, and calculates a second cornering power K2 _F and K2 _R in accordance with the lateral force Fy _F and Fy _R.
In step S108, the processing of the second stability factor calculating section 28, using the second cornering power K2 _F and K2 _R, to calculate a second stability factor A2.

  In subsequent step S109, a dynamic change characteristic of the turning behavior is estimated according to the sign and magnitude of the first stability factor A1 by the process of the characteristic estimation unit 29. That is, whether the turning behavior that changes as the vehicle speed V or the steering angle θ increases is an understeer characteristic (US characteristic), an oversteer characteristic (OS characteristic), or a neutral steer characteristic (NS characteristic). Is estimated. Further, the static turning behavior is estimated according to the sign and size of the second stability factor A2. That is, whether the turning behavior when the vehicle speed V or the steering angle θ at a certain time is maintained is an understeer characteristic (US characteristic), an oversteer characteristic (OS characteristic), or a neutral steer (NS characteristic). Is estimated.

In subsequent step S110, a target stability factor A ^* is set as the target behavior of the vehicle by the processing of the target behavior setting unit 15.
In step S111, the processing of the control amount setting unit 17, a second stability factor A2, and using the target stability factor A ^*, target yaw moment for achieving the target yaw rate gamma ^* Mz ^*, and the target A target deceleration Gx ^* for realizing the vehicle speed V ^* is set.
In subsequent step S112, the distribution ratio Rd between the yaw moment control and the deceleration control is set by the processing of the distribution ratio setting unit 18.
In the subsequent step S113, at least one of yaw moment control and deceleration control is executed according to the distribution ratio Rd by the processing of the drive unit 19, the actuator 13 is driven and controlled, and then the process returns to a predetermined main program.
The above is the traveling control process.

<Action>
Next, the operation of the first embodiment will be described.
First, FIG. 16 shows the relationship between the slip angle β and the lateral force Fy when the slip angle is β1.
The coordinate horizontal axis represents the slip angle β, the coordinate vertical axis represents the lateral force Fy, the relationship between the front wheel slip angle β and the lateral force Fy is indicated by a solid characteristic line Ln _F , and the rear wheel slip angle β It shows the relationship between the force Fy by the dashed characteristic line Ln _R. Further, a point that the slip angle on the characteristic line Ln _F is β1 and P1 _F, a tangential line at this point P1 _F and Lk1 _F and a Lk2 _F a straight line connecting the point P1 _F origin of coordinates [0, 0] To do. Further, a point that the slip angle on the characteristic line Ln _R is β1 and P1 _R, a tangent line at this point P1 _R and Lk1 _R is a Lk2 _R a straight line connecting the point P1 _R origin of coordinates [0, 0] To do. At this time, since the lateral force Fy of the rear wheel is larger than the lateral force Fy of the front wheel, the inclination of the straight line Lk2 _R becomes larger than the inclination of the straight line Lk2 _F. Therefore, it can be determined that the turning behavior is understeer characteristics. Further, there is no significant difference between the inclination of the tangent line Lk1 _{F and} the inclination of the tangent line Lk1 _R. Therefore, even if the slip angle β of the tire increases, the increase amount of the lateral force Fy does not greatly differ between the front wheels and the rear wheels, and the understeer characteristic is still maintained.

Next, FIG. 17 shows the relationship between the slip angle β and the lateral force Fy when the slip angle is β2 larger than β1.
Here, the point at which the slip angle is β2 on characteristic line Ln _F and P2 _F, a tangential line at this point P2 _F and Lk1 _F, Lk2 a straight line connecting the point P2 _F origin of coordinates [0,0] _F And Also, the point at which the slip angle is β2 on characteristic line Ln _R and P2 _R, a tangent line at this point P2 _R and Lk1 _R is a Lk2 _R a straight line connecting the point P2 _R origin of coordinates [0, 0] To do. At this time, since the lateral force Fy of the rear wheel is larger than the lateral force Fy of the front wheel, the inclination of the straight line Lk2 _R becomes larger than the inclination of the straight line Lk2 _F. Therefore, it is determined that the turning behavior is understeer characteristics. However, the inclination of the tangent line Lk1 _R is smaller than the inclination of the tangent line Lk1 _F and is substantially zero. Therefore, when the tire slip angle β increases, the lateral force Fy of the front wheels increases, but the lateral force Fy of the rear wheels starts to decrease. Therefore, while the turning behavior is in the understeer characteristic, it approaches the oversteer characteristic.

Next, FIG. 18 shows the relationship between the slip angle β and the lateral force Fy when the slip angle is β3 larger than β2.
Here, the point on the characteristic line Ln _F where the slip angle is β3 is P2 _F , the tangent at this point P3 _F is Lk1 _F, and the straight line connecting the coordinate origin [0, 0] and the point P3 _F is Lk2 _F And Also, the point at which the slip angle becomes β3 on the characteristic line Ln _R and P2 _R, a tangent line at this point P3 _R and Lk1 _R is a Lk2 _R a straight line connecting the point P3 _R origin of coordinates [0, 0] To do. In this case, since equal and the lateral force Fy of the lateral force Fy and the rear wheels of the front wheel, also equal inclinations of the straight line Lk2 _R linear Lk2 _F. Therefore, it is determined that the turning behavior is a neutral steer characteristic. However, the slope of the tangent line Lk1 _R is smaller than the slope of the tangent Lk1 _F, the soaring positive value, if the rightward descending a negative value, while the slope of the tangent line Lk1 _F is greater than 0 The slope of the tangent Lk1 _R is less than zero. Therefore, when the tire slip angle β increases, the lateral force Fy of the front wheels increases, but the lateral force Fy of the rear wheels further decreases. Therefore, it is immediately before the oversteer characteristic appears while the turning behavior is in the neutral steer characteristic.

As described above, the inclination of the straight line Lk2 _{F and} the inclination of the straight line Lk2 _R merely indicate static tire characteristics at a certain time, that is, the current turning behavior. Further, it is not possible to provide support for the next turning behavior only by paying attention to the current turning behavior, so it is required to grasp the dynamic change characteristics of the turning behavior that changes every moment.
In addition, as a method for achieving stable turning, a target yaw rate γ ^* is set and control intervention is performed according to a deviation Δγ from the actual yaw rate γ. This can suppress the limit behavior of the vehicle, but does not necessarily contribute to the ease of handling of the vehicle for the driver. That is, since the human internal model is constructed from the history of driving operations and vehicle behavior, it can be said that the behavior expected for a certain operation is an extension of the previous operation. Since the yaw rate γ constantly changes depending on the vehicle speed V and the steering angle θ, when the control intervention is performed according to the yaw rate deviation Δγ, the driver can recognize the causal relationship with the driving operation even if the driver can identify the change in the vehicle behavior. Hateful. Therefore, it is difficult to relate to the ease of handling of the vehicle for the driver.

In the present embodiment, attention is paid not only to the inclination of the straight line Lk2 _{F and} the inclination of the straight line Lk2 _R , but also to the inclination of the tangent Lk1 _{F and} the inclination of the tangent Lk1 _R. That is, in a predetermined coordinate system, the relationship between the tire slip angle β and the lateral force Fy serving as the reference characteristic is defined as the reference characteristic line Lr. Then, with reference to the reference characteristic line Lr, the inclinations of the tangents Lk1 _F and Lk1 _R at the coordinate positions corresponding to the tire lateral forces Fy _F and Fy _R are calculated as first cornering powers K1 _F and K1 _R. The first cornering powers K1 _F and K1 _R indicate dynamic tire characteristics at a certain point in time. That is, the cornering power becomes dynamic cornering power when the slip angle β and the lateral force Fy change from a certain point in time as the vehicle speed V or the steering angle θ increases. The first stability factor A1 is calculated using the first cornering powers K1 _F and K1 _R.

Then, the dynamic change characteristic of the turning behavior is estimated according to the sign and size of the first stability factor A1. That is, whether the turning behavior that changes as the vehicle speed V or the steering angle θ increases is an understeer characteristic (US characteristic), an oversteer characteristic (OS characteristic), or a neutral steer characteristic (NS characteristic). Is estimated.
As described above, since the inclinations of the tangents Lk1 _F and Lk1 _R at the coordinate position corresponding to the lateral force Fy are calculated as the first cornering powers K1 _F and K1 _R in the reference characteristic line Lr, the next cornering is performed. You can see how the power changes. Therefore, by calculating the first stability factor A1 using the first cornering powers K1 _F and K1 _R , it is possible to grasp how the stability factor changes next, and thus The dynamic change characteristics of the turning behavior can be grasped.

On the other hand, the inclination of the straight line Lk2 _{F and} the inclination of the straight line Lk2 _R are calculated as the second cornering powers K2 _F and K2 _R. The second cornering powers K2 _F and K2 _R indicate static tire characteristics at a certain point in time. That is, the cornering power is obtained when the slip angle β and the lateral force Fy at a certain point in time are maintained, so that the cornering power is static. Using this second cornering power K2 _F and K2 _R, to calculate a second stability factor A2.
Then, a static turning behavior is estimated according to the sign and size of the second stability factor A2. That is, whether the turning behavior when the vehicle speed V or the steering angle θ at a certain time is maintained is an understeer characteristic (US characteristic), an oversteer characteristic (OS characteristic), or a neutral steer (NS characteristic). Is estimated.

In this way, on the reference characteristic line Lr, the slopes of the straight lines Lk2 _F and Lk2 _R connecting the coordinate position corresponding to the lateral force Fy and the coordinate origin [0, 0] are set as the second cornering powers K2 _F and K2 _R. Since it is calculated, it is possible to grasp the current state of the cornering power. Therefore, by calculating the second stability factor A2 using the second cornering powers K2 _F and K2 _R , it is possible to grasp the current state of the stability factor, Thus, the static turning behavior can be grasped.

Further, a target stability factor A ^* of the vehicle is set by using any one of the first cornering powers K1 _F and K1 _R , and the turning of the vehicle is driven and controlled in accordance with the target stability factor A ^*. To do. Thereby, driving assistance according to the dynamic change characteristic of turning behavior can be performed.
For target stability factor A ^* settings, smaller the first cornering power K1, sets a target stability factor A ^* as understeer characteristic of the vehicle becomes stronger. This means that the smaller the first cornering power K1 is, the closer the lateral force Fy of the front wheel or the rear wheel approaches the saturated state. In this case, it is desirable to strengthen the understeer characteristic in advance. In this way, by setting the target stability factor A ^* so that the understeer characteristic of the vehicle becomes stronger as the first cornering power K1 is smaller, driving assistance corresponding to the dynamic change characteristic of the turning behavior is performed. Can do.

More specifically, when the first cornering power K1 is in the larger constant region than threshold Kt2 a predetermined so as to maintain the reference value A _R understeer characteristic of the vehicle is predetermined, target stability factor A ^* Set. Thus, when the first cornering power K1 is in the constant region, by setting the target stability factor A ^* to the reference value A _R, maintaining the weak understeer characteristics, to maintain a good steering characteristic it can.
On the other hand, when the first cornering power K1 is in the unsteady region smaller than the threshold value Kt2 is understeer characteristic of the vehicle is so stronger than the reference value A _R, sets the target stability factor A ^*. Thus, when the first cornering power K1 is in the non-stationary region by a target stability factor A ^* than the reference value A _R is set to a large value, prevented from becoming likely transition to over-steer characteristic In addition, good steering characteristics can be maintained.

  By the way, the smaller the first cornering power K1 means that the lateral force Fy of the front wheel or the rear wheel is approaching a saturated state, and therefore the vehicle does not have energy for bending the vehicle. Therefore, even when the yaw moment control is executed when the first cornering power K1 is small, the turning behavior of the vehicle may not be effectively controlled. On the other hand, as the first cornering power K1 is larger, it means that the front wheel or the rear wheel has a sufficient margin for exerting a sufficient lateral force Fy, so that the vehicle has sufficient energy to bend the vehicle. Therefore, when the first cornering power K1 is large, the turning behavior of the vehicle can be effectively controlled by executing the yaw moment control.

Therefore, in order to realize the target stability factor A ^* , when performing at least one of the yaw moment control and the deceleration control of the vehicle, either one of the first cornering powers K1 _F and K1 _R is used. The distribution ratio Rd for control and deceleration control is set. That is, the greater the first cornering power K1, the greater the yaw moment control weight than the deceleration control by increasing the yaw moment control distribution ratio Rd. On the other hand, the smaller the first cornering power K1, the smaller the yaw moment control distribution ratio Rd, thereby relatively increasing the weight of the deceleration control than the yaw moment control. Thereby, driving assistance according to the dynamic change characteristic of turning behavior can be performed.

FIG. 19 shows the relationship between the moving speed in the horizontal direction and the moving speed in the front-rear direction due to acceleration / deceleration.
(A) in the drawing shows a case where the first cornering power K1 is large and the tire grip has a sufficient margin. Assume that a target stability factor A ^* is on the characteristic line L _A indicated by a dotted line. Now, there is a turning traveling state, the relationship between the lateral moving speed and front-rear direction of the moving speed, and at point P0 on the characteristic line L _A. When there is the driver's acceleration demand from this state, if there is a sufficient margin to the tire grip, the yaw moment control to increase the moving speed of the lateral, the transition to point P1 on the characteristic line L _A Can do. Thereby, the target stability factor A ^* can be maintained and a desired understeer characteristic can be obtained.

(B) in the figure shows a case where the first cornering power K1 is small and the tire grip does not have a sufficient margin. Now, there is a point P0 on the characteristic line L _A, when there is an acceleration request of the driver from this state, as described above, if there is a sufficient margin to the tire grip, the yaw moment control, lateral movement Speed can be increased. However, because there is no enough room to the tire grip, this case is the deceleration control, to suppress the movement speed of the front-rear direction, it continues to stay at a point P0 on the characteristic line L _A. Thereby, the target stability factor A ^* can be maintained and a desired understeer characteristic can be obtained.
As described above, by setting the distribution ratio Rd of the yaw moment control and the deceleration control using one of the first cornering powers K1 _F and K1 _R , the operation according to the dynamic change characteristic of the turning behavior is set. Can provide support.

In the yaw moment control and the deceleration control, the target yaw moment Mz ^* and the target vehicle speed V ^* for realizing the target yaw rate γ ^* using the second stability factor A2 and the target stability factor A ^* are used ^. The target deceleration Gx ^* for realizing is set. Then, the target yaw moment Mz ^* is multiplied by the distribution ratio Rd to correct the target yaw moment Mz ^* , and the target deceleration Gx ^* is multiplied by (1.0−Rd) to obtain the target deceleration Gx ^*. Correct. In order to realize the corrected target yaw moment Mz ^* and target deceleration Gx ^* , drive control of the actuator 13 is performed, and control intervention is performed so that a desired understeer characteristic can be maintained. Thus, the controllability of the vehicle is ensured compared with the case where the control intervention is performed according to the yaw rate deviation Δγ and the turning behavior is forcibly controlled, and the vehicle can be driven along the intended line. Therefore, the handling of the vehicle for the driver is improved.

  Further, referring to the reference characteristic line Lr, there is a model (map) that defines the relationship between the lateral force Fy of the tire and the first cornering power K1 that is the inclination of the tangent line Lk1, and this model (map) is referred to Then, the first cornering power K1 is calculated using the lateral force Fy. Thus, the first cornering power K1 is easily calculated by calculating the first cornering power K1 using the model that defines the relationship between the lateral force Fy of the tire and the first cornering power K1. be able to.

  In addition, referring to the reference characteristic line Lr, there is a model (map) that defines the relationship between the lateral force Fy of the tire and the second cornering power K2 that is the inclination of the straight line Lk2, and this model (map) is referred to Then, the second cornering power K2 is calculated using the lateral force Fy. Thus, the second cornering power K2 is easily calculated by calculating the second cornering power K2 using the model that defines the relationship between the lateral force Fy of the tire and the second cornering power K2. be able to.

The reference characteristic line Lr, since defining the assumption that the reference coefficient of friction mu _R, and the different road surface mu _R of the reference road surface friction coefficient mu is the calculation accuracy of the first cornering power K1 In addition, the calculation accuracy of the second cornering power K2 is affected. Therefore, to estimate the friction coefficient mu _E of the road surface, to correct the lateral force Fy _F and Fy _R using the friction coefficient mu _E. More specifically, the lateral force Fy _F and Fy _R, the friction coefficient mu _R and the estimated friction coefficient criterion by multiplying the ratio (μ _R / μ _E) and mu _E, lateral force Fy _F and Fy _R is corrected. Thus, by the reference coefficient of friction mu _R and the estimated friction coefficient for correcting the lateral force Fy _F and Fy _R using the ratio (μ _{R /} μ _E) and mu _E, the friction coefficient of the road surface mu is also the reference of mu _R a different road, i.e. be running scene is changed, it is possible to calculate the first cornering power K1 and second cornering power K2 accurately.

Further, the estimated friction coefficient mu _E of the friction coefficient estimation unit 22, which may be affected by noise or disturbance contains estimation errors. Therefore, the friction coefficient μ _E used for correcting the lateral forces Fy _F and Fy _R is an average value of the friction coefficient μ _E estimated by the friction coefficient estimation unit 22 over a predetermined period. In this manner, by using the average value of the estimated friction coefficient mu _E over a predetermined period to the correction to suppress the influence of noise or disturbance, it is possible to improve the accuracy of correcting the lateral force Fy _F and Fy _R .

《Correspondence relationship》
In the present embodiment, the reference characteristic defining unit 24 corresponds to a “reference characteristic defining unit”, and the tire lateral force estimating unit 21 corresponds to a “tire condition obtaining unit”. The first cornering power calculation unit 25 corresponds to the “first cornering power calculation unit”, the target behavior setting unit 15 corresponds to the “target behavior setting unit”, and the drive control unit 16 corresponds to the “drive control unit”. Corresponding to Further, the friction coefficient estimation unit 22 corresponds to a “friction coefficient estimation unit”, the correction unit 23 corresponds to a “correction unit”, and the maps in FIGS. 6 and 7 correspond to a “first model”.

"effect"
Next, the effect of the main part in 1st Embodiment is described.
(1) The vehicle turning control apparatus according to the present embodiment defines, as a reference characteristic line Lr, a relationship between the tire slip angle β and the lateral force Fy serving as reference characteristics in a predetermined coordinate system. Then, with reference to the reference characteristic line Lr, the slope of the tangent at the coordinate position corresponding to the slip angle β or the lateral force Fy is calculated as the first cornering power K1. Then, the target stability factor A ^* of the vehicle is set using the first cornering power K1, and the turning traveling of the vehicle is driven and controlled according to the target stability factor A ^* .
Thus, since the slope of the tangent at the coordinate position corresponding to the slip angle β or the lateral force Fy is calculated as the first cornering power K1 in the reference characteristic line Lr, the first cornering power K1 is next. Can understand how changes occur. Therefore, by using this first cornering power K1, a target stability factor A ^* of the vehicle is set, and driving control of the vehicle's turning travel is performed, thereby providing driving support in accordance with the dynamic change characteristics of the turning behavior. It can be carried out.

(2) The turning control device for a vehicle according to the present embodiment sets the target stability factor A ^* so that the understeer characteristic of the vehicle becomes stronger as the first cornering power K1 is smaller.
In this way, by setting the target stability factor A ^* so that the understeer characteristic of the vehicle becomes stronger as the first cornering power K1 is smaller, driving assistance corresponding to the dynamic change characteristic of the turning behavior is performed. Can do.

(3) vehicle turning control apparatus according to this embodiment, when the first cornering power K1 is a threshold Kt2 larger constant region determined in advance, keep the reference value A _R understeer characteristic of the vehicle is predetermined As shown, the target stability factor A ^* is set.
Thus, it when the first cornering power K1 is in the constant region, by setting the target stability factor A ^* to the reference value A _R, which maintains a weak understeer characteristics, to maintain good steering characteristic Can do.

(4) The vehicle turning control apparatus according to this embodiment, when the first cornering power K1 is a threshold Kt2 smaller unsteady region, as understeer characteristic of the vehicle is stronger than the reference value A _R, the target Set stability factor A ^* .
Thus, when the first cornering power K1 is in the non-stationary region by a target stability factor A ^* than the reference value A _R is set to a large value, prevented from becoming likely transition to over-steer characteristic In addition, good steering characteristics can be maintained.

(5) The turning control device for a vehicle according to the present embodiment refers to the reference characteristic line Lr, and includes either one of the tire slip angle β or the lateral force Fy and the first cornering power K1 that is the tangential slope. A first model that defines the relationship is provided. With reference to the first model, the first cornering power K1 is calculated using the slip angle β or the lateral force Fy.
Thus, the first cornering power K1 is easily calculated by calculating the first cornering power K1 using the model that defines the relationship between the lateral force Fy of the tire and the first cornering power K1. be able to.

(6) The vehicle turning control apparatus according to the present embodiment estimates the friction coefficient mu _E of the road surface, one of the slip angle β or the lateral force Fy, corrected in accordance with the friction coefficient mu _E of the road surface.
Thus, by correcting the lateral force Fy using the estimated friction coefficient mu _E, also the mu _R of the friction coefficient mu is the reference of the road of a different road, first cornering power K1 and second The cornering power K2 can be calculated with high accuracy.

(7) The vehicle turning control device according to the present embodiment, the coefficient of friction of the road surface is assumed to define the reference characteristic line Lr and the reference value mu _R, and estimates the estimated road friction coefficient mu and _E, according to the ratio of the reference value mu _R and estimates mu _E, corrects one of the slip angle β or the lateral force Fy.
Thus, by the reference coefficient of friction mu _R and the estimated friction coefficient for correcting the lateral force Fy _F and Fy _R using the ratio (μ _{R /} μ _E) and mu _E, the friction coefficient of the road surface mu is also the reference of mu _R a different road surface, it is possible to calculate the first cornering power K1 and second cornering power K2 accurately.

(8) The vehicle turning control device according to the present embodiment, in accordance with the average value of the friction coefficient mu _E of the road surface estimated over a predetermined period, to correct one of the slip angle β or the lateral force Fy.
In this manner, by using the average value of the estimated friction coefficient mu _E over a predetermined period to the correction to suppress the influence of noise or disturbance, it is possible to improve the correction accuracy of the lateral force Fy.

(9) In the vehicle turning control method according to the present embodiment, the relationship between the tire slip angle β and the lateral force Fy, which are reference characteristics, is defined in advance as a reference characteristic line Lr in a predetermined coordinate system. . Then, with reference to the reference characteristic line Lr, the slope of the tangent at the coordinate position corresponding to the tire slip angle β or the lateral force Fy is calculated as the first cornering power K1. Then, the target stability factor A ^* of the vehicle is set using the first cornering power K1, and the turning traveling of the vehicle is driven and controlled according to the target stability factor A ^* .
Thus, since the slope of the tangent at the coordinate position corresponding to the slip angle β or the lateral force Fy is calculated as the first cornering power K1 in the reference characteristic line Lr, the first cornering power K1 is next. Can understand how changes occur. Therefore, by using this first cornering power K1, a target stability factor A ^* of the vehicle is set, and driving control of the vehicle's turning travel is performed, thereby providing driving support in accordance with the dynamic change characteristics of the turning behavior. It can be carried out.

  Although the present invention has been described with reference to a limited number of embodiments, the scope of rights is not limited thereto, and modifications of the embodiments based on the above disclosure are obvious to those skilled in the art.

DESCRIPTION OF SYMBOLS 11 Traveling control apparatus 12 Controller 13 Actuator 14 Turning characteristic estimation part 15 Target behavior setting part 16 Drive control part 17 Control amount setting part 18 Distribution ratio setting part 19 Drive part 21 Tire lateral force estimation part 22 Friction coefficient estimation part 23 Correction part 24 Reference characteristic defining unit 25 Cornering power calculating unit 26 Stability factor calculating unit 27 Cornering power calculating unit 28 Stability factor calculating unit 29 Characteristic estimating unit 30 Driving force control device 50 Brake control device

Claims (9)

In a predetermined coordinate system, a reference characteristic defining unit that defines the relationship between the slip angle and lateral force of the tire, which is a reference characteristic, as a reference characteristic line;
A tire condition acquisition unit for acquiring the slip angle or lateral force of the tire;
First, the slope of the tangent at the coordinate position corresponding to the slip angle or lateral force acquired by the tire condition acquisition unit is calculated as a first cornering power with reference to the reference characteristic line defined by the reference characteristic defining unit. Cornering power calculation unit,
A target behavior setting unit that sets a target stability factor of the vehicle using the first cornering power calculated by the first cornering power calculation unit;
A drive control unit that drives and controls turning of the vehicle according to the target stability factor set by the target behavior setting unit ,
The target behavior setting unit
The vehicle turning traveling control apparatus , wherein the target stability factor is set such that the understeer characteristic of the vehicle becomes stronger as the first cornering power is smaller . In a predetermined coordinate system, a reference characteristic defining unit that defines the relationship between the slip angle and lateral force of the tire, which is a reference characteristic, as a reference characteristic line;
A tire condition acquisition unit for acquiring the slip angle or lateral force of the tire;
First, the slope of the tangent at the coordinate position corresponding to the slip angle or lateral force acquired by the tire condition acquisition unit is calculated as a first cornering power with reference to the reference characteristic line defined by the reference characteristic defining unit. Cornering power calculation unit,
A target behavior setting unit that sets a target stability factor of the vehicle using the first cornering power calculated by the first cornering power calculation unit;
A drive control unit that drives and controls turning of the vehicle according to the target stability factor set by the target behavior setting unit ,
The target behavior setting unit
When the first cornering power is in a steady region greater than a predetermined threshold, the target stability factor is set so that the vehicle's understeer characteristic maintains a predetermined reference value . A turning control device. The target behavior setting unit
Wherein when the first cornering power is in the small non-stationary region than the threshold, as understeer characteristic of the vehicle it is stronger than the reference value, according to claim 2, characterized in that sets the target stability factor Vehicle turning control device. The first cornering power calculator is
With reference to the reference characteristic line, the tire has a first model that defines a relationship between one of a tire slip angle or a lateral force and the first cornering power that is the inclination of the tangent line,
Referring to the first model, using the slip angle or the lateral force acquired by the tire condition acquisition unit, any one of claim 1 to 3, characterized in that calculating the first cornering power The turning travel control device for a vehicle according to 1. A friction coefficient estimator for estimating the friction coefficient of the road surface;
Claim 1-4, characterized in that it comprises a correction unit for correcting in accordance with the friction coefficient of the road surface with one of the slip angle or the lateral force acquired by the tire condition acquisition unit, which is estimated by the friction coefficient estimating section The turning travel control device for a vehicle according to any one of the above. The correction unit is
According to the ratio between the reference value and the estimated value, the road surface friction coefficient premised on defining the reference characteristic line is a reference value, and the road surface friction coefficient estimated by the friction coefficient estimating unit is an estimated value. 6. The vehicle turning control device according to claim 5 , wherein one of the slip angle and the lateral force acquired by the tire state acquisition unit is corrected. The correction unit is
In response to said average value of the friction coefficient of the road surface estimated friction coefficient estimating unit for a predetermined period, claims and correcting one of the obtained slip angle or the lateral force in the tire-state obtaining section 5 Or the turning control apparatus for vehicles as described in 6 . In a predetermined coordinate system, the relationship between the slip angle and lateral force of the tire, which is the reference characteristic, is defined in advance as a reference characteristic line, and the tire corresponds to the tire slip angle or lateral force with reference to the reference characteristic line. Calculate the slope of the tangent at the coordinate position as the first cornering power,
Using the first cornering power, a target stability factor of the vehicle is set, and according to the target stability factor, driving control of turning of the vehicle is performed ,
The turning stability control method for a vehicle , wherein the target stability factor is set so that the understeer characteristic of the vehicle becomes stronger as the first cornering power is smaller . In a predetermined coordinate system, the relationship between the slip angle and lateral force of the tire, which is the reference characteristic, is defined in advance as a reference characteristic line, and the tire corresponds to the tire slip angle or lateral force with reference to the reference characteristic line. Calculate the slope of the tangent at the coordinate position as the first cornering power,
Using the first cornering power, a target stability factor of the vehicle is set, and according to the target stability factor, driving control of turning of the vehicle is performed ,
When the first cornering power is in a steady region greater than a predetermined threshold, the target stability factor is set so that the vehicle's understeer characteristic maintains a predetermined reference value . Turning control method.

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