JP2009185672A - Driving force controller for vehicle - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To improve stability of a vehicle by accurately and exactly controlling driving force in any travel scene such as starting, straight travelling, and cornering. <P>SOLUTION: A second torque down quantity operation part 4 operates body slip angle speed dβ/dt based on lateral acceleration d<SP>2</SP>y/dt<SP>2</SP>at a torque down base quantity setting part 4b and sets torque down base quantity ΔT2mover based on the body slip angle speed dβ/dt at the torque down base quantity setting part 4b. A first collection coefficient setting part 4c operates target yaw rate γt and operates yaw rate deviation Δγ, and sets first collection coefficient Kcant based on the yaw rate deviation Δγ. A second correction coefficient setting part 4d sets a second correction coefficient Kvs based on vehicle speed V. The second torque down quantity setting part 4e operates second torque down quantity T2over by multiplying ΔT2mover by Kcant and Kvs , and outputs the same to a controlled variable setting part 5. <P>COPYRIGHT: (C)2009,JPO&INPIT

Description

  The present invention relates to a driving force control device for a vehicle that controls driving force so as to appropriately maintain a grip force of a wheel in various running scenes of starting, going straight, and turning.

  Conventionally, various traction control devices that restrict and control the driving force to suppress wheel slip have been proposed and put into practical use.

For example, in Japanese Patent Laid-Open No. 10-310042, an estimated value of the friction circle radius of each wheel is obtained, and each wheel estimated according to the running state of the vehicle is generated within a range not exceeding the estimated value of the friction circle radius. A technique for adjusting the resultant force between the lateral force and the longitudinal force is disclosed.
JP 10-310042 A

  By the way, since it is difficult to obtain the information necessary for estimating the road surface friction coefficient and the tire force used for traction control as disclosed in the above-mentioned Patent Document 1 with high accuracy particularly at low speed, There is a possibility that the traction control in can not exert a sufficient effect. For example, when estimating the road surface friction coefficient using self-aligning torque, there is a tendency to detect the road surface friction coefficient higher at low speeds, which may cause the torque reduction amount to be set too low to perform sufficient torque reduction. There is. Therefore, in order to avoid this, it is unavoidable that the traction control is not activated unless the wheel speed exceeds a certain speed. However, this does not activate the traction control when turning on a snowy road. Problems arise.

  The present invention has been made in view of the above circumstances, and is a vehicle driving force control device capable of accurately and accurately controlling a driving force and improving the stability of the vehicle in all starting, straight traveling, and turning traveling scenes. The purpose is to provide.

  The present invention relates to a vehicle driving force control apparatus that appropriately maintains the grip force of a wheel by reducing the output of a driving source, and a lateral acceleration detecting means for detecting lateral acceleration and a steering angle detecting means for detecting a steering angle. A yaw rate detecting means for detecting a yaw rate; a target yaw rate calculating means for calculating a target yaw rate based on the detected steering angle; a yaw rate comparing means for comparing the detected yaw rate with the target yaw rate; A torque down amount setting means for setting a torque down amount using the comparison result between the acceleration and the yaw rate, and an output control means for reducing the output of the drive source based on the torque down amount are provided.

  According to the driving force control apparatus for a vehicle according to the present invention, it is possible to improve the stability of the vehicle by accurately and accurately controlling the driving force in every driving scene of starting, going straight, and turning.

Hereinafter, embodiments of the present invention will be described with reference to the drawings.
1 to 22 show an embodiment of the present invention, FIG. 1 is a functional block diagram showing the configuration of a driving force control device, FIG. 2 is a flowchart of a driving force control program, and FIG. 3 is a road surface friction coefficient estimating device. FIG. 4 is an explanatory diagram of the steering angle-steering torque characteristic, FIG. 5 is a characteristic diagram of the return speed set according to the vehicle speed and the front wheel slip angle, and FIG. 6 is the steering capacity and vehicle speed. FIG. 7 is a time chart showing an example of road surface friction coefficient estimation according to this embodiment, and FIG. 8 is a diagram showing the relationship between the accelerator opening and the throttle opening for generating the required engine torque. FIG. 9 is an explanatory view showing an example, FIG. 9 is an explanatory view showing an example of a required engine torque set by the engine speed and the throttle opening degree, and FIG. 10 is a functional block diagram showing a configuration of a first torque down amount calculation unit. 11 is a flowchart of a first torque-down amount calculation processing routine, FIG. 12 is a flowchart continuing from FIG. 11, FIG. 13 is a flowchart of an additional yaw moment calculation routine, FIG. 14 is an explanatory diagram of a lateral acceleration saturation coefficient, and FIG. FIG. 16 is an explanatory diagram of the difference in value of the additional yaw moment between the high μ road and the low μ road, FIG. 17 is a functional block diagram showing the configuration of the second torque-down amount calculation unit, and FIG. FIG. 19 is an explanatory diagram of the characteristic of the torque down basic amount with respect to the vehicle slip angular velocity, FIG. 20 is an explanatory diagram of the characteristic of the first correction coefficient with respect to the yaw rate deviation, and FIG. FIG. 22 is an explanatory diagram showing the characteristics of the second correction coefficient with respect to the vehicle speed. In this embodiment, a four-wheel drive vehicle with a center differential is taken as an example of the vehicle, and a vehicle in which the longitudinal driving force distribution can be varied from the base torque distribution Rf_cd by the center differential by a differential limiting clutch or the like (fastening torque TLSD). Explained as an example.

In FIG. 1, reference numeral 1 denotes a vehicle driving force control device that is mounted on a vehicle and appropriately suppresses the driving force. The driving force control device 1 includes a wheel speed sensor 11 for each wheel, a handle angle sensor 12, A yaw rate sensor 13, a lateral acceleration sensor 14, an accelerator opening sensor 15, an engine control unit 16, a transmission control unit 17, and a road surface μ estimation device 18 are connected, and wheel speeds ωfl, ωfr, ωrl, ωrr of each wheel (subscript “ fl ”is the left front wheel,“ fr ”is the right front wheel,“ rl ”is the left rear wheel, and“ rr ”is the right rear wheel), steering wheel angle θH, yaw rate γ, lateral acceleration (d ² y / dt ² ), The accelerator opening θACC, the engine speed Ne, the engine torque Teg, the main transmission gear ratio i, the turbine speed Nt of the torque converter, the engagement torque TLSD of the differential limiting clutch, and the road surface friction coefficient μ are input.

  Based on these input signals, the driving force control device 1 calculates an appropriate driving force in accordance with a driving force control program described later, and outputs it to the engine control unit 16. As will be described later, the engine control unit 16 outputs a control signal to a throttle control unit (not shown) as an output control means to drive the motor and operate the throttle valve.

First, the road surface μ estimation device 18 that estimates the road surface friction coefficient μ input to the driving force control device 1 will be described with reference to FIGS. The road surface μ estimation device 18 estimates the road surface friction coefficient μ based on the self-aligning torque Tsa, and the relationship between the rack thrust Fr and the self-aligning torque Tsa is expressed by the following equation (1). Therefore, the road surface friction coefficient μ is estimated by detecting the change of the self-aligning torque Tsa by the rack thrust Fr.
Fr = (2 / Ln) · Tsa (1)
Here, Ln is the knuckle arm length.

As shown in FIG. 3, the road surface μ estimation device 18 includes a wheel speed sensor 11, a handle angle sensor 12, a yaw rate sensor 13, a lateral acceleration sensor 14, an engine control unit 16, a transmission control unit 17, and a steering torque sensor 21 for each wheel. , Electric power steering motor 22 is connected, wheel speeds ωfl, ωfr, ωrl, ωrr of each wheel, steering wheel angle θH, yaw rate γ, lateral acceleration (d ² y / dt ² ), engine speed Ne, engine torque Teg, The main transmission gear ratio i, the turbine speed Nt of the torque converter, the engagement torque TLSD of the differential limiting clutch, the driver steering force Fd, and the assist force FEPS by electric power steering are input.

  Then, the road surface μ estimation device 18 estimates the road surface friction coefficient μ based on each of the input signals described above and outputs it to the driving force control device 1. That is, the road surface μ estimation device 18 includes a vehicle speed calculation unit 18a, a front wheel slip angle calculation unit 18b, a front wheel ground load calculation unit 18c, a front wheel longitudinal force calculation unit 18d, a front wheel lateral force calculation unit 18e, and a front wheel friction circle utilization factor calculation unit 18f. The estimation rack thrust calculation unit 18g, the reference rack thrust calculation unit 18h, the rack thrust deviation calculation unit 18i, and the road surface friction coefficient calculation unit 1j are mainly configured.

  The vehicle speed calculation unit 18a receives the wheel speeds ωfl, ωfr, ωrl, and ωrr of each wheel from the four-wheel wheel speed sensor 11, and calculates the vehicle speed V (= (ωfl + ωfr + ωrl + ωrr) / 4) by calculating the average of these. And output to the front wheel slip angle calculation unit 18b and the road surface friction coefficient calculation unit 18j.

  The front wheel slip angle calculation unit 18b receives the handle angle θH from the handle angle sensor 12, the yaw rate γ from the yaw rate sensor 13, and the vehicle speed V from the vehicle speed calculation unit 18a. Then, as described below, the front wheel slip angle βf is calculated based on the vehicle motion model, and is output to the reference rack thrust calculation unit 18h and the road surface friction coefficient calculation unit 18j.

The equation of motion related to the translational movement in the lateral direction of the vehicle is that the cornering force (one wheel) of the front and rear wheels is Cf, Cr, and the vehicle body mass is m.
2 · Cf + 2 · Cr = m · (d ² y / dt ² ) (2)
It becomes.

  On the other hand, the equation of motion related to the rotational motion around the center of gravity is expressed by the following equation (3), where the distances from the center of gravity to the front and rear wheel axes are Lf and Lr, the yaw moment of inertia of the vehicle body is Iz, and the yaw angular acceleration is (dγ / dt). Indicated by

  2 · Cf · Lf−2 · Cr · Lr = Iz · (dγ / dt) (3)

Further, when the vehicle slip angle is β and the vehicle slip angular velocity (dβ / dt), the lateral acceleration (d ² y / dt ² ) is
(D ² y / dt ² ) = V · ((dβ / dt) + γ) (4)
It is represented by
Therefore, the above equation (2) becomes the following equation (5).
2 · Cf + 2 · Cr = m · V · ((dβ / dt) + γ) (5)

The cornering force responds close to the first-order lag with respect to the tire slip angle, but this response lag is ignored and the suspension characteristics are linearized using equivalent cornering power that incorporates tire characteristics. .
Cf = Kf · βf (6)
Cr = Kr · βr (7)
Here, Kf and Kr are equivalent cornering powers of the front and rear wheels, and βf and βr are slip angles of the front and rear wheels.

In consideration of the influence of rolls and suspensions among the equivalent cornering powers Kf and Kr, the front and rear wheel slip angles βf and βr are calculated using the equivalent cornering powers Kf and Kr as the front and rear wheel steering angles δf, δr, The steering gear ratio can be simplified as follows, where n is n.
βf = δf− (β + Lf · γ / V)
= (ΘH / n)-(β + Lf · γ / V) (8)
βr = δr− (β−Lr · γ / V) (9)

ａ11＝−２・（Ｋｆ＋Ｋｒ）／（ｍ・Ｖ）
ａ12＝−１．０−２・（Ｌｆ・Ｋｆ−Ｌｒ・Ｋｒ）／（ｍ・Ｖ^２）
ａ21＝−２・（Ｌｆ・Ｋｆ−Ｌｒ・Ｋｒ）／Ｉｚ
ａ22＝−２・（Ｌｆ^２・Ｋｆ＋Ｌｒ^２・Ｋｒ）／（Ｉｚ・Ｖ）
ｂ11＝２・Ｋｆ／（ｍ・Ｖ・ｎ）
ｂ12＝２・Ｋｒ／（ｍ・Ｖ）
ｂ21＝２・Ｌｆ・Ｋｆ／Ｉｚ
ｂ22＝−２・Ｌｒ・Ｋｒ／Ｉｚ Summarizing the above equations of motion, the following equation of state is obtained.
(Dx (t) / dt) = A · x (t) + B · u (t) (10)
x (t) = [β γ] ^T
u (t) = [θH δr] ^T

a11 = −2 · (Kf + Kr) / (m · V)
a12 = −1.0−2 · (Lf · Kf−Lr · Kr) / (m · V ² )
a21 = -2. (Lf.Kf-Lr.Kr) / Iz
a22 = −2 · (Lf ² · Kf + Lr ² · Kr) / (Iz · V)
b11 = 2 · Kf / (m · V · n)
b12 = 2 · Kr / (m · V)
b21 = 2 · Lf · Kf / Iz
b22 = −2 · Lr · Kr / Iz

  That is, the vehicle slip angle β is calculated by solving the above equation (10), and the vehicle slip angle β is substituted into the above equation (8) to calculate the front wheel slip angle βf.

  The front wheel contact load calculation unit 18c receives the engine torque Teg and the engine speed Ne from the engine control unit 16, and receives the main transmission gear ratio i and the turbine speed Nt of the torque converter from the transmission control unit 17.

Then, according to the following equation (11), the front wheel ground load Fzf is calculated and output to the front wheel longitudinal force calculation unit 18d and the front wheel friction circle utilization factor calculation unit 18f.
Fzf = Wf − ((m · Ax · h) / L) (11)
Here, Wf is the front wheel static load, h is the height of the center of gravity, L is the wheelbase, and Ax is the longitudinal acceleration (= Fx / m). Fx in the calculation formula of the longitudinal acceleration Ax is a total driving force, and is calculated by, for example, the following equation (12) and is also output to the front wheel longitudinal force calculation unit 18d.
Fx = Tt · η · if / Rt (12)
Here, η is drive system transmission efficiency, if is the final gear ratio, and Rt is the tire radius. Tt is a transmission output torque, which is calculated by the following equation (13), for example, and this transmission output torque Tt is also output to the front wheel longitudinal force calculation unit 18d.
Tt = Teg · t · i (13)
Here, t is a torque ratio of the torque converter, and is obtained by referring to a preset map of the rotational speed ratio e (= Nt / Ne) of the torque converter and the torque ratio of the torque converter.

  The front wheel front / rear force calculation unit 18d receives the engagement torque TLSD of the differential limiting clutch from the transmission control unit 17, and the front wheel contact load Fzf, the total driving force Fx, and the transmission output torque Tt from the front wheel contact load calculation unit 18c. . Then, for example, according to the procedure described later, the front wheel longitudinal force Fxf is calculated and output to the front wheel friction circle utilization factor calculation unit 18f.

Hereinafter, an example of a calculation procedure for calculating the front wheel longitudinal force Fxf will be described.
First, the front wheel load distribution ratio WR_f is calculated by the following equation (14).
WR_f = Fzf / W (14)
Here, W is a vehicle weight (= m · G; G is a gravitational acceleration).

Next, the minimum front wheel front / rear torque Tfmin and the maximum front wheel front / rear torque Tfmax are calculated by the following equations (15) and (16).
Tfmin = Tt · Rf_cd−TLSD (≧ 0) (15)
Tfmax = Tt · Rf_cd + TLSD (≧ 0) (16)

Next, the minimum front wheel longitudinal force Fxfmin and the maximum front wheel longitudinal force Fxfmax are calculated by the following equations (17) and (18).
Fxfmin = Tfmin · η · if / Rt (17)
Fxfmax = Tfmax · η · if / Rt (18)

Then, the state is determined as follows.
When WR_f ≦ Fxfmin / Fx, it is assumed that the differential limiting torque is increased on the rear wheel side, and the determination value I = 1.
When WR_f ≧ Fxfmax / Fx, it is assumed that the differential limiting torque is increased on the front wheel side, and the determination value I = 3.
In cases other than the above, it is determined as normal time, and the determination value I = 2.

Next, according to the determination value I described above, the front wheel longitudinal force Fxf is calculated as follows.
When I = 1 Fxf = Tfmin · η · if / Rt (19)
When I = 2 ... Fxf = Fx.WR_f (20)
When I = 3 Fxf = Fxfmax.η.if / Rt (21)

The front wheel lateral force calculation unit 18 e receives the yaw rate γ from the yaw rate sensor 13 and the lateral acceleration (d ² y / dt ² ) from the lateral acceleration sensor 14. Then, the front wheel lateral force Fyf is calculated by the following equation (22), and is output to the front wheel friction circle utilization factor calculating unit 18f.
Fyf = (Iz · (dγ / dt)
+ M · (d ² y / dt ² ) · Lr) / L (22)

The front wheel friction circle utilization factor calculation unit 18f receives the front wheel ground load Fzf from the front wheel ground load calculation unit 18c, the front wheel longitudinal force Fxf from the front wheel longitudinal force calculation unit 18d, and the front wheel lateral force Fyf from the front wheel lateral force calculation unit 18e. The Then, the front wheel friction circle utilization factor rf is calculated by the following equation (23) and output to the road surface friction coefficient calculating unit 18j.
rf = (Fxf ² + Fyf ² ) ^1/2 / Fzf (23)

The estimated rack thrust calculation unit 18g receives the handle angle θH from the handle angle sensor 12, the driver steering force Fd from the steering torque sensor 21, and the assist force FEPS by the electric power steering from the electric power steering motor 22. Then, the estimated rack thrust FE is calculated by the following equation (24), and is output to the rack thrust deviation calculator 18i.
FE = Fd + FEPS−FFRI (24)
Here, FFRI is a force generated by friction or the like in the steering system, and is set by referring to a map set in advance, for example. An example of this map is shown in FIG. In this example, FFRI is given by the steering angle-steering torque characteristic, and is given by a hysteresis function based on the steering angle and the steering angular velocity. Note that the map characteristic shown in FIG. 4 is a map that takes into account the lateral acceleration (d ² y / dt ² ) and the value of the driver steering force Fd (specifically, the hysteresis interval between the ascending side and the descending side, The lateral acceleration (d ² y / dt ² ) and the driver steering force Fd may be changed to wider characteristics), and the FFRI may be obtained more accurately. By considering the FFRI in this manner, the estimated rack thrust FE can be accurately calculated not only when the steering is increased, but also when the steering is returned, and the road surface friction coefficient μ can be estimated over a wide range. It has become.

The reference rack thrust calculation unit 18h receives the front wheel slip angle βf from the front wheel slip angle calculation unit 18b. Then, the reference rack thrust FR is calculated by the following equation (25), and is output to the rack thrust deviation calculator 18i.
FR = −2 · Kf · ((ζc + ζn) / Ln) · βf (24)
Here, ζc is a caster trail, and ζn is a pneumatic trail.

The rack thrust deviation calculator 18i receives the estimated rack thrust FE from the estimated rack thrust calculator 18g and the reference rack thrust FR from the reference rack thrust calculator 18h. And rack thrust deviation (DELTA) FR is calculated by the following (26) Formula, and it outputs to the road surface friction coefficient calculating part 18j.
ΔFR = | FE−FR | (26)

  The road surface friction coefficient calculation unit 18j receives the vehicle speed V from the vehicle speed calculation unit 18a, the front wheel slip angle βf from the front wheel slip angle calculation unit 18b, and the front wheel friction circle use rate rf from the front wheel friction circle use rate calculation unit 18f. The rack thrust deviation ΔFR is input from the thrust deviation calculator 18i.

  Then, the rack thrust deviation ΔFR is compared with a preset maximum value determination threshold μmaxa. If the rack thrust deviation ΔFR is equal to or greater than the maximum value determination threshold μmaxa, it is determined that the tire is slipping, The front wheel friction circle utilization factor rf at that time is set as the road surface friction coefficient μ. When the rack thrust deviation ΔFR is smaller than the maximum value determination threshold μmaxa, the return speed Vμ for returning the road surface friction coefficient μ to a predetermined setting value (for example, 1.0) based on the vehicle speed V and the front wheel slip angle βf. Is determined with reference to a preset map (an example is shown in FIG. 5), and the road surface friction coefficient μ is calculated and output while returning the road surface friction coefficient at the return speed Vμ. Yes.

Note that the maximum value determination threshold μmaxa may be set to a large value according to the absolute value of the lateral acceleration (d ² y / dt ² ).

  As shown in FIG. 5, the map for determining the return speed Vμ has a characteristic that the return speed Vμ decreases as the vehicle speed V increases and as the front wheel slip angle βf increases.

When the vehicle is modeled by a two-wheel model, a so-called steering capacity ωn · ζ that determines the steering stability of the vehicle is obtained by the following equation (27). The higher the steering capacity ωn · ζ, the higher the convergence of the vehicle.
ωn · ζ = (a11 + a22) / 2 (27)
The above-mentioned a11 and a22 are those described in the above-mentioned equation (10), and the term a11 is known to contribute to the convergence property of the vehicle slip angle, and this term is linear. The greater the change is, the more the stability of the vehicle is improved and the response is in line with the driver's feeling. The term a22 is a system matrix element that affects the convergence of the yaw motion, and serves as a negative feedback gain of the yaw rate.

These terms a11 and a22 can be described by the following equations (28) and (29) when the nonlinearity of the tire is simply taken into account by a quadratic equation.
a11 = (1 / (m · V)) · (2 · (Kf + Kr)
− ((Kf ² / (μ ² · Fzf ² −Fxf ² ) ^1/2 ) · | βf |
+ (Kr ² / (μ ² · Fzr ² −Fxr ² ) ^1/2 ) · | βr |)) (28)
a22 = (1 / (Iz · V)) · (2 · (Lf ² · Kf + Lr ² · Kr)
− ((Lf ² · Kf ² / (μ ² · Fzf ² −Fxf ² ) ^1/2 ) · | βf |
+ (Lr ² · Kr ² / (μ ² · Fzr ² −Fxr ² ) ^1/2 ) · | βr |)) (29)
Here, Fzr is the rear wheel ground contact load, and Fxr is the rear wheel longitudinal force.

  As is clear from the above equations (28) and (29), the higher the vehicle speed V, the smaller the terms a11 and a22, and hence the steering capacity ωn · ζ becomes smaller (FIG. 6 (a)). reference). Therefore, in consideration of this, the characteristic is set such that the return speed Vμ decreases as the vehicle speed V increases, thereby suppressing a large rapid change.

  Similarly, as is clear from the above equations (28) and (29), the higher the front wheel slip angle βf, the smaller the terms a11 and a22, and hence the steering capacity ωn · ζ becomes smaller ( (Refer FIG.6 (b)). Therefore, in consideration of this, the characteristic is set such that the return speed Vμ decreases as the front wheel slip angle βf increases, thereby suppressing a large rapid change.

  In the present embodiment, the return speed Vμ is set using the front wheel slip angle βf. However, as is clear from the above equations (28) and (29), the rear wheel slip angle βr is set. May be used to set the return speed Vμ.

An example of this road surface friction coefficient estimation will be described with reference to the time chart of FIG.
Until time t1, ΔFR <μmaxa, and the road surface friction coefficient μ is stably set to 1.0.

  From time t1 to time t2, ΔFR ≧ μmaxa is satisfied, and it is determined that the tire is slipping, and the front wheel friction circle utilization rate rf at that time is set as the road surface friction coefficient μ.

  From time t2 to time t3, ΔFR <μmaxa again, and the road surface friction coefficient μ is set while returning the road surface friction coefficient to 1.0 at the return speed Vμ.

  Thus, according to the present embodiment, the rack thrust deviation ΔFR is compared with the preset maximum value determination threshold μmaxa, and if the rack thrust deviation ΔFR is greater than or equal to the maximum value determination threshold μmaxa, the tire Is determined to be slipping, and the front wheel friction circle utilization ratio rf at that time is set as the road surface friction coefficient μ. When the rack thrust deviation ΔFR is smaller than the maximum value determination threshold μmaxa, the return speed Vμ for returning the road surface friction coefficient μ to a predetermined setting value (for example, 1.0) based on the vehicle speed V and the front wheel slip angle βf. Is determined with reference to a map set in advance, and the road surface friction coefficient μ is calculated and output while returning the road surface friction coefficient at the return speed Vμ. For this reason, when it can be determined that the tire is slipping, the road surface friction coefficient μ is set appropriately low by the front wheel friction circle utilization rate rf at that time, and in other cases, the road surface friction coefficient is set to 1. Since the road surface friction coefficient μ is set while returning to 0, the road surface friction coefficient μ is set appropriately without being maintained at a low value, even if it is difficult to estimate the road surface friction coefficient. Even in such a case, it is possible to set the road surface friction coefficient appropriately and maximize the potential of the vehicle behavior control.

  The return speed at which the road surface friction coefficient is returned to 1.0 is variably set in accordance with the vehicle speed V and the front wheel slip angle βf in consideration of the steering capacity ωn · ζ related to the convergence of the vehicle behavior. Therefore, a smooth road friction coefficient μ is set with a natural feeling while maintaining a high level of vehicle stability.

  In the embodiment of the present invention, the road surface friction coefficient is estimated based on the friction circle utilization factor rf or a predetermined set value (for example, 1) according to the relationship of the deviation ΔFR between the estimated rack thrust FE and the reference rack thrust FR. 0.0), it goes without saying that it can be applied to other road surface friction coefficient estimation devices. For example, it is impossible to estimate the road surface friction coefficient with respect to the technique for estimating the road surface friction coefficient from the steering angle, the vehicle speed, the yaw rate, etc., using the adaptive control theory proposed by the present applicant in Japanese Patent Laid-Open No. 8-2274. When a situation (for example, when the steering angle is 0) is detected, the return speed at which the road surface friction coefficient μ is restored to a preset value (for example, 1.0) based on the vehicle speed V and the front wheel slip angle βf. Vμ may be determined with reference to a map set in advance, and the road surface friction coefficient μ may be calculated and output while returning the road surface friction coefficient at the return speed Vμ. Similarly, it is needless to say that the present invention can be applied to other estimation methods, for example, a method for estimating a road surface friction coefficient using an observer as disclosed in Japanese Patent Application Laid-Open No. 2000-71968.

  Then, as shown in FIG. 1, the driving force control device 1 includes a required engine torque calculation unit 2, a first torque down amount calculation unit 3, a second torque down amount calculation unit 4, and a control amount setting unit 5. It is configured.

  The requested engine torque calculation unit 2 receives the accelerator opening θACC from the accelerator opening sensor 15 and receives the engine speed Ne from the engine control unit 16. Then, a throttle opening θth is obtained from a preset map (for example, a map of θACC-θth relationship as shown in FIG. 8), and based on the throttle opening θth and the engine speed Ne, FIG. The engine torque is obtained from the map shown, and this engine torque is output to the control amount setting unit 5 as the required engine torque Tdrv. The required engine torque Tdrv may be obtained from a map corresponding to a preset accelerator opening θACC, or may be read from the engine control unit 16 and used.

The first torque reduction amount calculation unit 3 is connected to a wheel speed sensor 11, a handle angle sensor 12, a yaw rate sensor 13, a lateral acceleration sensor 14, an engine control unit 16, a transmission control unit 17, and a road surface μ estimation device 18 for each wheel. Wheel speed ωfl, ωfr, ωrl, ωrr, steering wheel angle θH, yaw rate γ, lateral acceleration (d ² y / dt ² ), engine speed Ne, engine torque Teg, main transmission gear ratio i, torque converter Turbine rotational speed Nt, differential limiting clutch engagement torque TLSD, and road surface friction coefficient μ are input. Then, the first torque-down amount calculation unit 3 calculates the first torque-down amount T1over based on these input signals and outputs it to the control amount setting unit 5. Hereinafter, the first torque-down amount calculation unit 3 will be described with reference to FIGS.

  That is, as shown in FIG. 10, the first torque down amount calculation unit 3 includes a transmission output torque calculation unit 3a, a total driving force calculation unit 3b, a front / rear ground load calculation unit 3c, a left wheel load ratio calculation unit 3d, Contact load calculator 3e, each wheel longitudinal force calculator 3f, each wheel required lateral force calculator 3g, each wheel lateral force calculator 3h, each wheel friction circle limit value calculator 3i, each wheel required tire resultant force calculator 3j, Each wheel tire resultant force calculation unit 3k, each wheel required over tire resultant force calculation unit 3l, each wheel over tire resultant force calculation unit 3m, an over tire force calculation unit 3n, and a first torque down amount setting unit 3o are mainly configured. .

The transmission output torque calculation unit 3a receives the engine speed Ne and the engine torque Teg from the engine control unit 16, and the main transmission gear ratio i and the turbine speed Nt of the torque converter from the transmission control unit 17. Then, for example, the transmission output torque Tt is calculated by the above-described equation (13), and is output to the total driving force calculation unit 3b and each wheel longitudinal force calculation unit 3f.
The total driving force calculation unit 3b receives the transmission output torque Tt from the transmission output torque calculation unit 3a. Then, for example, the total driving force Fx is calculated by the above-described equation (12), and is output to the front / rear ground load calculating unit 3c and the front / rear wheel front / rear force calculating unit 3f.
The front / rear ground load calculator 3c receives the total driving force Fx from the total driving force calculator 3b.

Then, the front wheel ground load Fzf is calculated by the above-described equation (11), and is output to each wheel ground load calculation unit 3e and each wheel longitudinal force calculation unit 3f, and the rear wheel ground load Fzr is calculated by the following equation (30). Calculate and output to each wheel ground load calculation unit 3e.
Fzr = W−Fzf (30)

Lateral acceleration (d ² y / dt ² ) is input from the lateral acceleration sensor 14 to the left wheel load ratio calculation unit 3d. Then, the left wheel load ratio WR_l is calculated by the following equation (31), and is output to each wheel ground load calculating unit 3e, each wheel required lateral force calculating unit 3g, and each wheel lateral force calculating unit 3h.
WR — 1 = 0.5 − ((d ² y / dt ² ) / G) · (h / Ltred) (31)
Here, Ltred is the average tread value of the front and rear wheels.

Each wheel ground load calculation unit 3e receives the front wheel ground load Fzf and the rear wheel ground load Fzr from the front and rear ground load calculation unit 3c, and the left wheel load ratio WR_l from the left wheel load ratio calculation unit 3d. Then, the left front wheel ground load Fzf_l, the right front wheel ground load Fzf_r, the left rear wheel ground load Fzr_l, and the right rear wheel ground load Fzr_r are calculated by the following equations (32), (33), (34), and (35), respectively. And output to each wheel friction circle limit value calculation unit 3i.
Fzf_l = Fzf · WR_l (32)
Fzf_r = Fzf · (1-WR_l) (33)
Fzr_l = Fzr · WR_l (34)
Fzr_r = Fzr · (1−WR_l) (35)

  Each wheel longitudinal force calculation unit 3f receives the engagement torque TLSD of the differential limiting clutch in the center differential from the transmission control unit 17, receives the transmission output torque Tt from the transmission output torque calculation unit 3a, and outputs the total driving force calculation unit 3b. The total driving force Fx is input from the front, and the front wheel ground load Fzf is input from the front and rear ground load calculation unit 3c. Then, for example, according to the procedure described later, the left front wheel front / rear force Fxf_l, the right front wheel front / rear force Fxf_r, the left rear wheel front / rear force Fxr_l, and the right rear wheel front / rear force Fxr_r are calculated. It outputs to the wheel tire resultant force calculating part 3k.

  Hereinafter, an example of a procedure for calculating the left front wheel longitudinal force Fxf_l, the right front wheel longitudinal force Fxf_r, the left rear wheel longitudinal force Fxr_l, and the right rear wheel longitudinal force Fxr_r will be described.

  First, the front wheel front / rear force Fxf is calculated by any of the above-described equations (19), (20), and (21), and the rear wheel front / rear force Fxr is calculated by the following equation (36).

            Fxr = Fx−Fxf (36)

By using the front wheel longitudinal force Fxf and the rear wheel longitudinal force Fxr, the left front wheel longitudinal force Fxf_l, the right front wheel longitudinal force Fxf_r, the left rear wheel longitudinal force Fxr_l, and the right according to the following equations (37) to (40) The rear wheel longitudinal force Fxr_r is calculated.
Fxf_l = Fxf / 2 (37)
Fxf_r = Fxf_l (38)
Fxr_l = Fxr / 2 (39)
Fxr_r = Fxr_l (40)

  Note that the calculation of the front-rear force of each wheel shown in the present embodiment is merely an example, and is appropriately selected according to the drive type / drive mechanism of the vehicle.

Each wheel required lateral force calculation unit 3g has a lateral acceleration (d ² y / dt ² ) from the lateral acceleration sensor 14, a yaw rate γ from the yaw rate sensor 13, a handle angle θH from the handle angle sensor 12, and (4 Wheel) The wheel speeds ωfl, ωfr, ωrl, and ωrr of each wheel are input from the wheel speed sensor 11, and the left wheel load ratio WR_l is input from the left wheel load ratio calculation unit 3d.

  Then, the additional yaw moment Mzθ is calculated according to the procedure described later (in accordance with the flowchart shown in FIG. 13), and based on this additional yaw moment, the required front wheel lateral force Fyf_FF is calculated according to the following equation (41). ) To calculate the required rear wheel lateral force Fyr_FF. Based on these required front wheel lateral force Fyf_FF and requested rear wheel lateral force Fyr_FF, the left front wheel required lateral force Fyf_l_FF, the right front wheel required lateral force Fyf_r_FF, the left rear wheel required lateral force Fyr_l_FF, and the right by the equations (43) to (46) The rear wheel required lateral force Fyr_r_FF is calculated and output to each wheel required tire resultant force calculating unit 3j.

Fyf_FF = Mzθ / L (41)
Fyr_FF = (− Iz · (dγ / dt)
+ M · (d ² y / dt ² ) · Lf) / L (42)
Here, Iz is the yaw moment of inertia of the vehicle, and Lf is the distance between the front axis and the center of gravity.

Fyf_l_FF = Fyf_FF · WR_l (43)
Fyf_r_FF = Fyf_FF · (1-WR_l) (44)
Fyr_l_FF = Fyr_FF · WR_l (45)
Fyr_r_FF = Fyr_FF · (1-WR_l) (46)

Further, as shown in FIG. 13, the additional yaw moment Mzθ is first calculated in step (hereinafter abbreviated as “S”) 301 (for example, V = (ωfl + ωfr + ωrl + ωrr) / 4), and proceeds to S302. The lateral acceleration / steering wheel angle gain Gy is calculated by the following equation (47).
Gy = (1 / (1 + A · V ² )) · (V ² / L) · (1 / n) (47)
Here, A is a stability factor, and n is a steering gear ratio.

Next, the process proceeds to S303, and a map set in advance according to the value obtained by multiplying the lateral acceleration / handle angle gain Gy and the handle angle θH (Gy · θH) and the lateral acceleration (d ² y / dt ² ) is referred to. Then, the lateral acceleration saturation coefficient Kμ is calculated. As shown in FIG. 14A, the map for obtaining the lateral acceleration saturation coefficient Kμ is obtained by multiplying the lateral acceleration / handle angle gain Gy and the handle angle θH (Gy · θH) and the lateral acceleration (d ² y / dt ² ), and is set to a smaller value as the lateral acceleration (d ² y / dt ² ) increases when the handle angle θH is equal to or greater than a predetermined value. This lateral acceleration as is a high μ road when Gy · .theta.H large ^(d 2 y / dt ²⁾ is larger, the lateral acceleration ^(d 2 y / dt ²⁾ is less likely to occur in the low μ road It expresses. As a result, the value of the reference lateral acceleration (d ² yr / dt ² ), which will be described later, is large when the lateral acceleration (d ² y / dt ² ) is large when Gy · θH is large, as shown in FIG. When the road is considered to be a road, the value is set to a low value so that the correction amount for the additional yaw moment Mzθ is small.

Next, in S304, the lateral acceleration deviation sensitive gain Ky is calculated by the following equation (48).
Ky = Kθ / Gy (48)
Here, Kθ is a steering angle sensitive gain, and is calculated by the following equation (49).
Kθ = (Lf · Kf) / n (49)
Kf is the equivalent cornering power of the front shaft.

That is, according to the above equation (48), the lateral acceleration deviation sensitive gain Ky is set as a guideline (maximum value) in a state where the rudder does not work at all on an extremely low μ road (γ = 0, (d ² y / dt). ² ) = 0) and the case where the additional yaw moment Mzθ (steady value) is 0 is considered.

Next, proceeding to S305, the reference lateral acceleration (d ² yr / dt ² ) is calculated by the following equation (50).

(D ² yr / dt ² ) = Kμ · Gy · (1 / (1 + Ty · s)) · θH (50)
Here, s is a differential operator, Ty is a first-order lag time constant of lateral acceleration, and this first-order lag time constant Ty is calculated by, for example, the following equation (51), where Kr is the equivalent cornering power of the rear axis. Is done.
Ty = Iz / (L · Kr) (51)

Next, in S306, the lateral acceleration deviation (d ² ye / dt ² ) is calculated by the following equation (52).
(D ² ye / dt ² ) = (d ² y / dt ² ) − (d ² yr / dt ² ) (52)

In step S307, the yaw rate / handle angle gain Gγ is calculated by the following equation (53).
Gγ = (1 / (1 + A · V ² )) · (V / L) · (1 / n) (53)

Next, the process proceeds to S308, and considering the case where the additional yaw moment Mzθ (steady value) becomes 0, for example, when the grip travels ((d ² ye / dt ² ) = 0) by the following equation (54), the yaw rate sensitivity The gain Kγ is calculated.

    Kγ = Kθ / Gγ (54)

  In step S309, a vehicle speed sensitive gain Kv is calculated from a preset map. This map is set in order to avoid an unnecessary additional yaw moment Mzθ in a low speed region, for example, as shown in FIG. In FIG. 15, Vc1 is 40 km / h, for example.

In S310, the additional yaw moment Mzθ is calculated by the following equation (55).
Mzθ = Kv · (−Kγ · γ + Ky · (d ² ye / dt ² ) + Kθ · θH) (55)

That is, as shown in the equation (55), the term −Kγ · γ is a yaw moment that is sensitive to the yaw rate γ, the term Kθ · θH is a yaw moment that is sensitive to the handle angle θH, and Ky · (d ² ye / dt The term ² ) is the correction value for the yaw moment. Therefore, as shown in FIG. 16, when the lateral acceleration (d ² y / dt ² ) is operated on a high μ road, the additional yaw moment Mzθ also becomes a large value, and the motion performance is improved. On the other hand, when traveling on a low μ road, the additional yaw moment Mzθ reduces the additional yaw moment Mzθ by the above-described correction value, so that the turning performance does not increase and stable traveling performance can be obtained. It has become.

Each wheel lateral force calculation unit 3h receives a lateral acceleration (d ² y / dt ² ) from the lateral acceleration sensor 14, a yaw rate γ from the yaw rate sensor 13, and a left wheel load ratio WR_l from the left wheel load ratio calculation unit 3d. Then, the front wheel lateral force Fyf_FB is calculated by the following equation (56), and the rear wheel lateral force Fyr_FB is calculated by the following equation (57). Based on these front wheel lateral force Fyf_FB and rear wheel lateral force Fyr_FB, the left front wheel lateral force Fyf_l_FB, right front wheel lateral force Fyf_r_FB, left rear wheel lateral force Fyr_l_FB, right rear wheel lateral force Fyr_r_FB Is output to each wheel tire resultant force calculation unit 3k.
Fyf_FB = (Iz · (dγ / dt)
+ M · (d ² y / dt ² ) · Lr) / L (56)
Fyr_FB = (− Iz · (dγ / dt)
+ M · (d ² y / dt ² ) · Lf) / L (57)
Here, Lr is the distance between the rear axis and the center of gravity.
Fyf_l_FB = Fyf_FB · WR_l (58)
Fyf_r_FB = Fyf_FB · (1-WR_l) (59)
Fyr_l_FB = Fyr_FB · WR_l (60)
Fyr_r_FB = Fyr_FB · (1-WR_l) (61)

  Each wheel friction circle limit value calculation unit 3i receives the road surface friction coefficient μ from the road surface μ estimation device 18, and each wheel ground load calculation unit 3e receives a left front wheel ground load Fzf_l, a right front wheel ground load Fzf_r, and a left rear wheel ground load. Fzr_l and right rear wheel ground load Fzr_r are input.

Then, the left front wheel friction circle limit value μ_Fzfl, the right front wheel friction circle limit value μ_Fzfr, the left rear wheel friction circle limit value μ_Fzrl, and the right rear wheel friction circle limit value μ_Fzrr are calculated by the following equations (62) to (65). , Output to each wheel required overtire force calculation unit 3l and each wheel overtire force calculation unit 3m.
μ_Fzfl = μ · Fzf_l (62)
μ_Fzfr = μ · Fzf_r (63)
μ_Fzrl = μ · Fzr_l (64)
μ_Fzrr = μ · Fzr_r (65)

Each wheel required tire resultant force calculation unit 3j receives the left front wheel front / rear force Fxf_l, right front wheel front / rear force Fxf_r, left rear wheel front / rear force Fxr_l, and right rear wheel front / rear force Fxr_r from each wheel front / rear force calculation unit 3f. A left front wheel required lateral force Fyf_l_FF, a right front wheel required lateral force Fyf_r_FF, a left rear wheel required lateral force Fyr_l_FF, and a right rear wheel required lateral force Fyr_r_FF are input from the lateral force calculation unit 3g. Then, by calculating the left front wheel required tire resultant force F_fl_FF, the right front wheel required tire resultant force F_fr_FF, the left rear wheel required tire resultant force F_rl_FF, and the right rear wheel required tire resultant force F_rr_FF by the following formulas (66) to (69), It outputs to the overtire force calculation part 3l.
F_fl_FF = (Fxf_l ² + Fyf_l_FF ² ) ^1/2 (66)
F_fr_FF = (Fxf_r ² + Fyf_r_FF ² ) ^1/2 (67)
F_rl_FF = (Fxr_l ² + Fyr_l_FF ² ) ^1/2 (68)
F_rr_FF = (Fxr_r ² + Fyr_r_FF ² ) ^1/2 (69)

Each wheel tire resultant force calculation unit 3k receives the left front wheel front / rear force Fxf_l, right front wheel front / rear force Fxf_r, left rear wheel front / rear force Fxr_l, and right rear wheel front / rear force Fxr_r from each wheel front / rear force calculation unit 3f. A left front wheel lateral force Fyf_l_FB, a right front wheel lateral force Fyf_r_FB, a left rear wheel lateral force Fyr_l_FB, and a right rear wheel lateral force Fyr_r_FB are input from the calculation unit 3h. Then, the left front wheel tire resultant force F_fl_FB, the right front wheel tire resultant force F_fr_FB, the left rear wheel tire resultant force F_rl_FB, and the right rear wheel tire resultant force F_rr_FB are calculated by the following equations (70) to (73), Output to 2o.
F_fl_FB = (Fxf_l ² + Fyf_l_FB ² ) ^1/2 (70)
F_fr_FB = (Fxf_r ² + Fyf_r_FB ² ) ^1/2 (71)
F_rl_FB = (Fxr_l ² + Fyr_l_FB ² ) ^1/2 (72)
F_rr_FB = (Fxr_r ² + Fyr_r_FB ² ) ^1/2 (73)

Each wheel required over-tyre force calculation unit 31 receives from each wheel friction circle limit value calculation unit 3i left front wheel friction circle limit value μ_Fzfl, right front wheel friction circle limit value μ_Fzfr, left rear wheel friction circle limit value μ_Fzrl, right rear wheel friction The circle limit value μ_Fzrr is input, and the left front wheel required tire resultant force F_fl_FF, the left rear wheel required tire resultant force F_rl_FF, and the right rear wheel required tire resultant force F_rr_FF are input from each wheel required tire resultant force calculation unit 3j. . Then, the left front wheel required overtire force ΔF_fl_FF, the right front wheel required overtire force ΔF_fr_FF, the left rear wheel required overtire force ΔF_rl_FF, and the right rear wheel required overtire force ΔF_rr_FF are calculated by the following equations (74) to (77). And output to the over-tyre force calculator 3n.
ΔF_fl_FF = F_fl_FF−μ_Fzfl (74)
ΔF_fr_FF = F_fr_FF−μ_Fzfr (75)
ΔF_rl_FF = F_rl_FF−μ_Fzrl (76)
ΔF_rr_FF = F_rr_FF−μ_Fzrr (77)

Each wheel over-tyre force calculation unit 3m, from each wheel friction circle limit value calculation unit 3i, left front wheel friction circle limit value μ_Fzfl, right front wheel friction circle limit value μ_Fzfr, left rear wheel friction circle limit value μ_Fzrl, right rear wheel friction circle The limit value μ_Fzrr is inputted, and the left front wheel tire resultant force F_fl_FB, the right front wheel tire resultant force F_fr_FB, the left rear wheel tire resultant force F_rl_FB, and the right rear wheel tire resultant force F_rr_FB are inputted from each wheel tire resultant force calculation unit 3k. Then, by calculating the left front wheel over-tyre force ΔF_fl_FB, the right front wheel over-tyre force ΔF_fr_FB, the left rear wheel over-tyre force ΔF_rl_FB, and the right rear wheel over-tyre force ΔF_rr_FB by the following equations (78) to (81), It outputs to the calculating part 3n.
ΔF_fl_FB = F_fl_FB−μ_Fzfl (78)
ΔF_fr_FB = F_fr_FB−μ_Fzfr (79)
ΔF_rl_FB = F_rl_FB−μ_Fzrl (80)
ΔF_rr_FB = F_rr_FB−μ_Fzrr (81)

The over-tyre force calculating unit 3n receives the left front wheel required over-tyre force ΔF_fl_FF, the right front wheel required over-tyre force ΔF_fr_FF, the left rear wheel required over-tyre force ΔF_rl_FF, and the right rear wheel required over-tyre force from each wheel required over-tyre force calculating unit 3l. ΔF_rr_FF is input, and the left front wheel overtire force ΔF_fl_FB, the right front wheel overtire force ΔF_fr_FB, the left rear wheel overtire force ΔF_rl_FB, and the right rear wheel overtire force ΔF_rr_FB are input from each wheel overtire force calculation unit 3m. Then, the sum of each wheel required overtire force ΔF_fl_FF, ΔF_fr_FF, ΔF_rl_FF, ΔF_rr_FF and the sum of each wheel overtire force ΔF_fl_FB, ΔF_fr_FB, ΔF_rl_FB, ΔF_rr_FB are compared, and the larger value is set as the overtire force Fover. To do. That is,
Fover = MAX ((ΔF_fl_FF + ΔF_fr_FF + ΔF_rl_FF + ΔF_rr_FF)
, (ΔF_fl_FB + ΔF_fr_FB + ΔF_rl_FB + ΔF_rr_FB)) (82)

In the first torque-down amount setting unit 3o, the engine speed Ne from the engine control unit 16, the main transmission gear ratio i from the transmission control unit 17, and the turbine speed Nt of the torque converter are exceeded from the over tire force calculation unit 3n. The tire force Fover is input. Then, the first torque-down amount T1over is calculated by the following equation (83) and output to the control amount setting unit 5.
T1over = Fover · Rt / t / i / η / if (83)

Next, the first torque reduction amount calculation processing program executed by the first torque reduction amount calculation unit 3 will be described with reference to the flowcharts of FIGS.
First, in S201, the transmission output torque calculation unit 3a calculates the transmission output torque Tt by the above-described equation (13).

  Next, it progresses to S202 and the total driving force Fx is calculated by the total driving force calculation part 3b by the above-mentioned (12) Formula.

  Next, the process proceeds to S203, where the front / rear ground load calculation unit 3c calculates the front wheel ground load Fzf by the above-described equation (11), and calculates the rear wheel ground load Fzr by the above-described equation (30).

  Next, proceeding to S204, the left wheel load ratio calculation unit 3d calculates the left wheel load ratio WR_l by the above-described equation (31).

  Next, the process proceeds to S205, where each wheel ground load calculation unit 3e uses the above-described equations (32), (33), (34), and (35) to respectively determine the left front wheel ground load Fzf_l, the right front wheel ground load Fzf_r, and the left rear. The wheel contact load Fzr_l and the right rear wheel contact load Fzr_r are calculated.

  Next, the process proceeds to S206, and the front / rear wheel force calculation unit 3f calculates the left front wheel front / rear force Fxf_l, the right front wheel front / rear force Fxf_r, the left rear wheel front / rear force Fxr_l, and the right rear wheel by the equations (37) to (40) described above. The longitudinal force Fxr_r is calculated.

  Next, the process proceeds to S207, where each wheel required lateral force calculation unit 3g calculates the left front wheel required lateral force Fyf_l_FF, the right front wheel required lateral force Fyf_r_FF, the left rear wheel required lateral force Fyr_l_FF, the right according to the above-described equations (43) to (46). The rear wheel required lateral force Fyr_r_FF is calculated.

  Next, proceeding to S208, each wheel lateral force calculation unit 3h uses the above-described equations (58) to (61) to calculate the left front wheel lateral force Fyf_l_FB, the right front wheel lateral force Fyf_r_FB, the left rear wheel lateral force Fyr_l_FB, and the right rear wheel. The lateral force Fyr_r_FB is calculated.

  Next, in S209, each wheel friction circle limit value calculation unit 3i calculates the left front wheel friction circle limit value μ_Fzfl, the right front wheel friction circle limit value μ_Fzfr, the left rear wheel friction circle according to the above-described equations (62) to (65). The limit value μ_Fzrl and the right rear wheel friction circle limit value μ_Fzrr are calculated.

  Next, proceeding to S210, each wheel required tire resultant force calculation unit 3j uses the above-described equations (66) to (69) to calculate the left front wheel required tire resultant force F_fl_FF, the right front wheel required tire resultant force F_fr_FF, and the left rear wheel required tire resultant force F_rl_FF. The right rear wheel required tire resultant force F_rr_FF is calculated.

  Next, the process proceeds to S211, and each wheel tire resultant force calculation unit 3k calculates the left front wheel tire resultant force F_fl_FB, the right front wheel tire resultant force F_fr_FB, the left rear wheel tire resultant force F_rl_FB, and the right rear wheel tire according to the above formulas (70) to (73). The resultant force F_rr_FB is calculated.

  Next, the process proceeds to S212, and each wheel required overtire force calculation unit 3l calculates the left front wheel required overtire force ΔF_fl_FF, the right front wheel required overtire force ΔF_fr_FF, and the left rear wheel required according to the above-described equations (74) to (77). The overtire force ΔF_rl_FF and the right rear wheel required overtire force ΔF_rr_FF are calculated.

  Next, the process proceeds to S213, and each wheel over-tyre force calculation unit 3m calculates the left front wheel over-tyre force ΔF_fl_FB, the right front wheel over-tyre force ΔF_fr_FB, the left rear wheel over-tyre force ΔF_rl_FB, according to the equations (78) to (81) described above. The right rear wheel over-tyre force ΔF_rr_FB is calculated.

  Next, the process proceeds to S214, and the over tire force calculation unit 3n calculates the over tire force Fover by the above-described equation (82).

  Next, the process proceeds to S215, where the first torque-down amount setting unit 3o calculates the first torque-down amount T1over by the above-described equation (83), and outputs it to the control amount setting unit 5 to exit the program.

  As described above, according to the embodiment of the present invention, the first torque-down amount calculation unit 3 generates a torque value that causes the tire force generated on the wheel to exceed the friction circle limit value based on the driver request, and a current value generated on the wheel. The tire force is compared with the torque value exceeding the frictional circle limit value, and the larger value is subtracted from the driving force required by the driver to be corrected. For this reason, not only the present but also the future excessive torque state is properly corrected, the spin and plow are properly controlled, and the tire grip force is properly maintained to maintain the vehicle running stability. It is possible to improve.

  In addition, the value to be corrected by subtracting the driving force required by the driver is a torque value that causes the tire force to exceed the friction circle limit value, so the driving force in the front-rear direction is not suddenly suppressed, and the driver There is no unnatural feeling or dissatisfaction such as lack of acceleration.

On the other hand, in FIG. 1, the second torque-down amount calculation unit 4 includes a wheel speed sensor 11 for each wheel, a handle angle sensor 12 as a steering angle detection means, a yaw rate sensor 13 as a yaw rate detection means, and a lateral acceleration detection means. The lateral acceleration sensor 14 is connected, and wheel speeds ωfl, ωfr, ωrl, ωrr, steering wheel angle θH, yaw rate γ, and lateral acceleration (d ² y / dt ² ) of each wheel are input. Then, the second torque-down amount calculation unit 4 calculates a second torque-down amount T2over based on these input signals and outputs it to the control amount setting unit 5. Hereinafter, the second torque-down amount calculation unit 4 will be described with reference to FIGS.

  That is, as shown in FIG. 17, the second torque down amount calculation unit 4 includes a vehicle speed calculation unit 4a, a torque down basic amount setting unit 4b, a first correction coefficient setting unit 4c, and a second correction coefficient setting unit 4d. The second torque-down amount setting unit 4e is mainly configured.

  The vehicle speed calculation unit 4a receives the wheel speeds ωfl, ωfr, ωrl, and ωrr of each wheel from the four-wheel wheel speed sensor 11, and calculates a vehicle speed V (= (ωfl + ωfr + ωrl + ωrr) / 4) by calculating an average of these. And output to the torque reduction basic amount setting unit 4b, the first correction coefficient setting unit 4c, and the second correction coefficient setting unit 4d.

The torque-down basic amount setting unit 4b receives the yaw rate γ from the yaw rate sensor 13, the lateral acceleration (d ² y / dt ² ) from the lateral acceleration sensor 14, and the vehicle speed V from the vehicle speed calculation unit 4a. First, the vehicle slip angular velocity (dβ / dt) is calculated from the following equation (84), and based on the vehicle slip angular velocity (dβ / dt), the vehicle slip angular velocity previously set by experiment or the like is calculated. With reference to the characteristic map of the torque-down basic amount (FIG. 19), the torque-down basic amount ΔT2mover is set and output to the second torque-down amount setting unit 4e. That is, the torque-down basic amount setting unit 4b is provided as torque-down amount setting means.
(Dβ / dt) = γ − ((d ² y / dt ² ) / V) (84)

  As shown in FIG. 19, the characteristic map of the torque down basic amount with respect to the vehicle slip angular velocity is set so that the torque down basic amount ΔT2mover increases as the vehicle slip angular velocity (dβ / dt) increases. This is because when the vehicle body angular velocity (dβ / dt) increases, the possibility of spin and plow increases, so the amount of torque reduction is increased to surely prevent such unstable vehicle behavior.

The first correction coefficient setting unit 4c receives the handle angle θH from the handle angle sensor 12, the yaw rate γ from the yaw rate sensor 13, and the vehicle speed V from the vehicle speed calculation unit 4a. Then, the steering angle based yaw rate (target yaw rate) γt is calculated by the following equation (85), the yaw rate deviation Δγ is calculated by the following equation (86), and the absolute value | Δγ | The first correction coefficient Kcant is set with reference to the characteristic map (FIG. 20) of the first correction coefficient with respect to the yaw rate deviation set in advance by experiments or the like, and the second torque down amount setting unit 4e. Output to.
γt = (1 / (1 + A · V ² )) · (V / L) · (θH / n) (85)
Δγ = γt−γ (86)
As will be described later, the first correction coefficient Kcant is a coefficient for correcting the torque-down basic amount ΔT2mover by multiplying the torque-down basic amount ΔT2mover in the second torque-down amount setting unit 4e.

That is, the basic torque-down amount ΔT2mover is calculated based on the lateral acceleration (d ² y / dt ² ) from the lateral acceleration sensor 14 as shown in the above equation (84). The lateral acceleration (d ² y / dt ² ) from 14 is influenced by the cant as shown in FIG.

For example, as shown in FIG. 21B, in the case of a cant that tilts the vehicle body toward the inside of the turn, the lateral acceleration (d ² y / dt ² ) from the lateral acceleration sensor 14 is the actual lateral acceleration. Is Gr, gravitational acceleration is G, and cant tilt angle is θ, (Gr · cos (θ) −G · sin (θ)), which is smaller than the actual value.

Conversely, as shown in FIG. 21C, in the case of a cant that tilts the vehicle body toward the outside of the turn, the lateral acceleration (d ² y / dt ² ) from the lateral acceleration sensor 14 is (Gr · cos (θ) + G · sin (θ)), which is larger than the actual value.

  In order to eliminate such an error due to the influence of the cant on the output value from the lateral acceleration sensor 14, a correction coefficient (first correction coefficient) is used by using the steering angle-based yaw rate (target yaw rate) γt that is not affected by the cant. ) Is set.

  As shown in FIG. 20, the characteristic map of the first correction coefficient for the yaw rate deviation is a region where the absolute value | Δγ | of the yaw rate deviation is small, in other words, the difference between the target yaw rate γt and the actual yaw rate is small. When the steering wheel operation and the vehicle behavior substantially coincide with each other, it is determined that it is not necessary to set the torque reduction amount, and is set to substantially zero. Thus, the first correction coefficient setting unit 4c is provided as a target yaw rate calculation means, a yaw rate comparison means, and a torque down amount setting means.

  The second correction coefficient setting unit 4d receives the vehicle speed V from the vehicle speed calculation unit 4a. Based on the vehicle speed V, a second correction coefficient Kvs is set with reference to a characteristic map (FIG. 22) of the second correction coefficient with respect to the vehicle speed that has been set in advance through experiments or the like, and the second torque is set. It outputs to the down amount setting part 4e.

  As will be described later, the second correction coefficient Kvs is a coefficient for correcting the torque-down basic amount ΔT2mover by multiplying the torque-down basic amount ΔT2mover by the second torque-down amount setting unit 4e.

  As shown in FIG. 22, the characteristic map of the second correction coefficient with respect to the vehicle speed is set to be high in the low vehicle speed region. This is a value obtained by calculating the first torque down amount T1over by the first torque down amount calculating unit 3 based on the road surface friction coefficient μ from the road surface μ estimating device 18, and this road surface friction coefficient. Since μ is estimated based on the self-aligning torque in particular, it tends to increase in the low vehicle speed region. Accordingly, in the low vehicle speed region, the first torque down amount T1over becomes small, and there is a possibility that sufficient torque down cannot be performed. Therefore, the second torque down amount T2over is given so that accurate traction control can be performed. It is.

The second torque down amount setting unit 4e includes a torque down basic amount ΔT2mover from the torque down basic amount setting unit 4b, a first correction coefficient Kcant from the first correction coefficient setting unit 4c, and a second correction coefficient setting unit. The second correction coefficient Kvs is input from 4d. Then, the second torque-down amount T2over is calculated by the following equation (87) and output to the control amount setting unit 5. That is, the second torque down amount setting unit 4e is provided as a torque down amount setting unit together with the first correction coefficient setting unit 4c.
T2over = Kcant · Kvs · ΔT2mover (87)

Next, the second torque-down amount calculation processing program executed by the above-described second torque-down amount calculation unit 4 will be described with reference to the flowchart of FIG.
First, in S401, the vehicle speed calculation unit 4a calculates the vehicle speed V, and in S402, the torque down basic amount setting unit 4b calculates the vehicle slip angular velocity (dβ / dt) according to the above-described equation (84).

  Next, in S403, the torque-down basic amount setting unit 4b sets the torque-down basic amount ΔT2mover with reference to the characteristic map (FIG. 19) of the torque-down basic amount with respect to the vehicle slip angular velocity.

  Next, the process proceeds to S404, and the first correction coefficient setting unit 4c calculates the target yaw rate γt by the above-described equation (85).

  Next, the process proceeds to S405, and the first correction coefficient setting unit 4c calculates the yaw rate deviation Δγ by the above-described equation (86).

  Next, in S406, the first correction coefficient setting unit 4c sets the first correction coefficient Kcant with reference to the characteristic map (FIG. 20) of the first correction coefficient with respect to the yaw rate deviation.

  Next, the process proceeds to S407, and the second correction coefficient setting unit 4d sets the second correction coefficient Kvs with reference to the characteristic map (FIG. 22) of the second correction coefficient with respect to the vehicle speed.

  Then, the process proceeds to S408, and the second torque down amount setting unit 4e calculates the second torque down amount T2over according to the above-described equation (87), and outputs it to the control amount setting unit 5 to exit the program.

  As described above, according to the embodiment of the present invention, the second torque-down amount calculation unit 4 performs the second operation based on the vehicle slip angular velocity (dβ / dt) obtained from the lateral acceleration without using the road surface friction coefficient μ. Since the torque down amount T2over is calculated, an accurate torque down amount can be obtained even in the low speed range.

Further, the second torque reduction amount T2over calculates the vehicle slip angular velocity (dβ / dt) based on the lateral acceleration (d ² y / dt ² ) from the lateral acceleration sensor 14, and the vehicle slip angular velocity (dβ / dt). ) Based on the torque-down basic amount ΔT2mover, the lateral acceleration (d ² y / dt ² ) from the lateral acceleration sensor 14 may be affected by a cant and include an error. Since this is corrected by the first correction coefficient Kcant that is set based on the rudder angle-based yaw rate that is not affected by cant, it is possible to obtain a precise torque down amount that is not affected by cant. It has become.

Next, in FIG. 1, the control amount setting unit 5 includes a request engine torque Tdrv from the request engine torque calculation unit 2, a first torque down amount T1over from the first torque down amount calculation unit 3, and a second torque. A second torque down amount T2over is input from the down amount calculation unit 4. Then, the total torque down amount Tover, which is the final torque down amount, is calculated by the following equation (88), the control amount Treq is calculated by the following equation (89), and is output to the engine control unit 16. That is, the control amount setting unit 5 constitutes a torque down amount setting means.
Tover = T1over + T2over (88)
Treq = Tdrv−Tover (89)

Next, the driving force control program executed by the above-described driving force control apparatus 1 will be described with reference to the flowchart of FIG.
First, parameters required in S101, that is, wheel speeds ωfl, ωfr, ωrl, ωrr of each wheel, steering wheel angle θH, yaw rate γ, lateral acceleration (d ² y / dt ² ), accelerator opening θACC, engine speed Ne. , The engine torque Teg, the main transmission gear ratio i, the turbine speed Nt of the torque converter, the engagement torque TLSD of the differential limiting clutch, and the road surface friction coefficient μ are read. In S102, the required engine torque calculation unit 2 Calculate Tdrv.

  Next, proceeding to S103, the first torque-down amount computing unit 3 computes the first torque-down amount T1over by the above-described equation (83), and proceeding to S104, the second torque-down amount computing unit 4 Calculates the second torque-down amount T2over by the above-described equation (87).

  Next, the process proceeds to S105, where the control amount setting unit 5 calculates the total torque down amount Tover using the above-described equation (88), and proceeds to S106, where the control amount Treq is calculated using the above-described equation (89). Output to the control unit 16 and exit the program.

  As described above, according to the present embodiment, in the low speed range, torque reduction is executed by the second torque down amount T2over set based on the vehicle body slip angular velocity (dβ / dt) obtained from the lateral acceleration. In this area, torque reduction is executed by the first torque down amount T1over set based on the road surface friction coefficient μ. Therefore, accurate torque reduction can be executed in all vehicle speed areas, and all of starting, going straight, and turning are possible. It is possible to improve the stability of the vehicle by accurately controlling the driving force in the traveling scene.

In the present embodiment, the vehicle slip angular velocity (dβ / dt) when calculating the second torque-down amount T2over is the above-described equation (84), that is, yaw rate γ, lateral acceleration (d ² y / dt). ² ) The calculation is made from the vehicle speed V. Alternatively, the output from the vehicle slip angle sensor may be obtained by differentiation, based on position information from GPS (Global Positioning System). It may be obtained or may be obtained by processing image information from the camera.

  In the present embodiment, the vehicle slip angular velocity is calculated from the yaw rate γ from the yaw rate sensor, the lateral acceleration from the lateral acceleration sensor and the vehicle speed, and the basic torque reduction amount is set based on the vehicle slip angular velocity. ing. However, in setting the basic torque reduction amount, the yaw rate γ, the vehicle speed, and the vehicle slip angular velocity are not necessarily required, and other parameters are not particularly limited as long as the lateral acceleration affected by the cant is at least used.

  Further, in the present embodiment, the basic torque down amount is set from the lateral acceleration sensor, and then the basic torque down amount is corrected based on the yaw rate deviation between the detected yaw rate and the target yaw rate, and the influence of cant Is eliminated. However, the present invention is not limited to the method of performing correction after calculating the torque down amount as a reference, and even if the torque down amount is set using the lateral acceleration and the yaw rate deviation that eliminates the influence of cant from the beginning. The problems of the present invention can be solved.

  Furthermore, in this embodiment, the yaw rate deviation between the detected yaw rate and the target yaw rate is used. However, the present invention is not limited thereto, and may be a comparison between the detected yaw rate and the target yaw rate, such as a ratio of the detected yaw rate to the target yaw rate, and is not limited to the yaw rate deviation.

Functional block diagram showing the configuration of the driving force control device Driving force control program flowchart Functional block diagram showing the configuration of the road surface friction coefficient estimation device Illustration of steering angle-steering torque characteristics Characteristic chart of return speed set according to vehicle speed and front wheel slip angle Explanatory drawing showing the relationship between steering capacity, vehicle speed and wheel slip angle Time chart showing an example of road surface friction coefficient estimation according to this embodiment Explanatory drawing which shows an example of the relationship between the accelerator opening and throttle opening for generating required engine torque Explanatory drawing showing an example of required engine torque set by engine speed and throttle opening Functional block diagram showing the configuration of the first torque-down amount calculation unit Flowchart of first torque down amount calculation processing routine Flowchart continuing from FIG. Flow chart of additional yaw moment calculation routine Illustration of lateral acceleration saturation coefficient Characteristic map of vehicle speed sensitive gain Explanatory diagram of the difference in the value of the additional yaw moment on the high and low μ roads Functional block diagram showing the configuration of the second torque-down amount calculation unit Flow chart of second torque down amount calculation processing routine Explanatory diagram of characteristics of basic torque-down amount with respect to vehicle slip angular velocity Explanatory drawing of the characteristic of the first correction coefficient for the yaw rate deviation Explanatory diagram showing the effect of Kant on lateral acceleration sensor Explanatory diagram of characteristics of second correction coefficient for vehicle speed

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Driving force control apparatus 2 Required engine torque calculating part 3 1st torque down amount calculating part 4 2nd torque down amount calculating part 4a Vehicle speed calculating part 4b Torque down basic amount setting part (torque down amount setting means)
4c 1st correction coefficient setting part (target yaw rate calculation means, yaw rate comparison means, torque down amount setting means)
4d 2nd correction coefficient setting part 4e 2nd torque down amount setting part (torque down amount setting means)
5 Control amount setting part (torque down amount setting means)
11 Wheel speed sensor 12 Handle angle sensor (steering angle detection means)
13 Yaw rate sensor (yaw rate detection means)
14 Lateral acceleration sensor (lateral acceleration detection means)
15 Accelerator opening sensor 16 Engine control unit (output control means)
17 Transmission Control Unit 18 Road Surface μ Estimator

Claims (3)

In the vehicle driving force control device that appropriately maintains the grip force of the wheels by reducing the output of the driving source,
Lateral acceleration detecting means for detecting lateral acceleration;
Steering angle detection means for detecting the steering angle;
A yaw rate detection means for detecting the yaw rate;
Target yaw rate calculating means for calculating a target yaw rate based on the detected steering angle;
A yaw rate comparison means for comparing the detected yaw rate with the target yaw rate;
Torque down amount setting means for setting a torque down amount using at least a comparison result between the lateral acceleration and the yaw rate;
Output control means for reducing the output of the drive source based on the torque reduction amount;
A driving force control apparatus for a vehicle, comprising:   The torque down amount setting means corrects the torque down amount set based on the lateral acceleration based on the comparison result of the yaw rate, and the output control means outputs the output of the drive source based on the corrected torque down amount. The driving force control apparatus for a vehicle according to claim 1, wherein:   The yaw rate comparing means calculates a deviation between the detected yaw rate and the target yaw rate, and the torque down amount setting means corrects the torque down amount to be smaller as the absolute value of the yaw rate deviation is smaller. The vehicle driving force control device according to claim 1, wherein the driving force control device is a vehicle driving force control device.

Images