JPH106804A - Controller for transmitting power between right and left wheels of vehicle - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a stable turn performance by detecting the degree of slip of wheels when a vehicle is turning and by controlling a rotational force adjusting mechanism for adjusting the rotational force of the right and left wheels in response to the state of turning of the vehicle estimated based on the detected results. SOLUTION: A controller for transmitting power between the right and left wheels which is mounted on the rear wheel of a four-wheel-drive vehicle comprises a rotational force adjusting mechanism (torque transmitting mechanism) 50 provided in a differential gear carrier, a hydraulic unit for controlling the mechanism 50 and an ECU 42. The torque transmitting mechanism 50 comprises a speed change mechanism and a torque transmitting mechanism of variable transmitting capacity control type and is interposed between the differential gear case of a rear differential gear and a right wheel side rotation shaft. The ECU 42 detects, based on the output signals of a wheel speed sensor, steering wheel angle sensor, 21 transverse G sensor and a back-and-forth G sensor, the degree of slip of the wheel when the vehicle is turning and controls the amount of slip of the right and left clutches 90L, 90R of the torque transmitting mechanism of variable transmitting capacity control type in response to the turning state of the vehicle estimated based on the detected results.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention is provided between left and right wheels of a vehicle. When the left and right wheels are driving wheels, the driving force is transmitted to the left and right wheels while controlling the distribution of the driving force. The present invention relates to a power transmission control device for left and right wheels for transmitting and receiving torque between the left and right wheels when the left and right wheels are rotated when the is a driven wheel, and in particular, to enhance the stability of the vehicle body posture during turning of the vehicle. The present invention relates to a power transmission control device between left and right wheels for a vehicle, which is suitable for power transmission control between left and right wheels.

[0002]

2. Description of the Related Art In a motor vehicle, a differential mechanism is provided between left and right wheels, which are driving wheels, for allowing a differential generated at the time of turning or the like. If one of the wheels spins, for example, if it gets stuck in the sand, there may be a case where only one of the wheels rotates and the other wheel hardly rotates, and the driving torque cannot be transmitted to the road surface.

Therefore, in such a case, a differential limiting mechanism (LSD = Limited Slip Differential) capable of limiting the differential has been developed. Such a left and right wheel differential limiting mechanism includes a type that is proportional to the rotational speed difference between the left and right wheels and a type that is proportional to the input torque. The right and left wheel rotation speed difference proportional type uses VC that utilizes the viscosity of the liquid.
(Viscous coupling) type LSD, etc.
There is an advantage that the running stability of the vehicle can be improved. On the other hand, the input torque proportional type includes a mechanical type such as a friction type such as a general LOM (lock automatic) type LSD, and has an advantage that the turning performance of the vehicle can be improved.

However, in the various differential limiting mechanisms described above, the differential control characteristics are determined by physical properties and the like, and the differential control characteristics can always be adjusted so that the differential control can always be appropriately performed. Has not become. There is also a system called an electronically controlled LSD in which the LSD is electronically controlled. However, even in such a system, the torque transfer between the wheels is limited only from a high speed side to a low speed side. It is considered that the traveling performance cannot be sufficiently improved during the turning traveling of the vehicle.

The applicant of the present invention has proposed, for example, Japanese Patent Application Laid-Open No. Hei 5-131855, which discloses that torque can be distributed between left and right wheels in various running states of a vehicle without causing a large torque loss or energy loss. Kaihei 7-108840
JP-A-7-108841, JP-A-7-108842,
Japanese Patent Laying-Open Nos. Hei 7-108843 and Hei 7-156681 have proposed a torque transfer control device for left and right wheels for a vehicle, as disclosed in Japanese Patent Application Laid-Open Nos. Hei 7-108843 and Hei 7-156681.

In this torque transfer control device between left and right wheels, when two rotating bodies arranged coaxially are slid in contact with each other at different rotating speeds, the rotating body having a higher rotating speed has a lower rotating speed. This utilizes the characteristic that torque is transmitted to the rotating body. That is, for example, this device is provided with a transmission mechanism that changes the rotation speed input to the differential device or the rotation speed of one of the wheel shafts to high and low speeds and outputs the same, and receives the respective outputs of the transmission mechanism to make a difference. A plurality of speed change interlocking members that rotate at a rotation speed different from the driving device or one of the wheel shafts, a constant speed interlocking member that rotates at the same speed as the other wheel axle of the left and right wheels, and these speed change interlocking members. It is provided with a plurality of torque transmission couplings such as a wet multi-plate clutch provided between the transmission and the speed interlocking member.

In such a device, even if the left and right wheels are rotating at a constant speed, the torque transmission coupling requires:
Since the rotational speed is different between the speed change interlocking member side and the constant speed interlocking member side, if the torque transmission coupling is actuated by engaging a wet multi-plate clutch or the like, the speed change interlocking member side and the constant speed interlocking member side become The torque is transmitted from the higher speed to the lower speed. If the degree of speed change by the speed change mechanism is set to be greater than or equal to a certain value, torque can be transmitted from the inner wheel having a lower rotation speed to the outer wheel having a higher rotation speed during turning.

In a torque transmission coupling such as a wet multi-plate clutch, for example, by controlling the switching of the engagement of each wet multi-plate clutch and the degree of engagement, the transmission torque to one wheel axle is controlled. Can be increased or decreased, and the transmission torque to the other wheel axle can be increased or decreased. Therefore, since the transmission torque capacity can be variably controlled, the torque can be transmitted to the left and right wheels in a desired direction with a desired transmission torque capacity.

Such a device can be applied whether the left and right wheels are driving wheels or driven wheels, and if the left and right wheels are driving wheels, the distribution of the driving force from the engine to the left and right wheels is adjusted. If the left and right wheels are driven wheels, the wheels on the torque transmitting side will receive the driving force by the torque transmission, and the wheels on the torque transmitting side will receive the braking force.

In any case, the magnitude of the driving force or the braking force exerted between the left and right wheels and the road surface is imbalanced in the left and right directions, thereby generating a yaw moment in the vehicle and reducing the behavior of the vehicle. Can be controlled.

[0011]

By the way, in a device capable of transmitting torque between the left and right wheels of a vehicle as described above,
In controlling the behavior of the vehicle by generating a yaw moment in the vehicle, the direction and strength of the yaw moment to be applied to the vehicle differ according to the running state of the vehicle, so the direction of torque transmission between the left and right wheels and An important issue is how to control the size.

That is, in a normal running state of the vehicle, for example, it is considered that it is suitable to control the torque so that a speed difference occurs between the left and right wheels according to the vehicle speed, the steering angle, and the like during turning. However, in a running state where the vehicle drifts while skidding the wheels, such torque control is by no means suitable. Therefore, for example, when the vehicle is drifting, it is desired that control different from that when the vehicle is not drifting can be performed. For this purpose, it is necessary to determine whether or not to drift.

Japanese Patent Application Laid-Open No. 7-101262 discloses an apparatus capable of controlling the distribution of torque to the left and right driving wheels by a control type orifice coupling.
In particular, when the turning direction determined from the lateral acceleration and the turning direction determined from the steering angle are opposite, it is determined that the vehicle is in the counter steer state. When the counter steer state is not in the high lateral acceleration state, the left and right drive wheels are directly connected. Along with
A control method for increasing the driving force of the wheel inside the turn when the vehicle is in the counter-steering state and the high lateral acceleration state is shown.

However, such a prior art (Japanese Unexamined Patent Publication No. 7-101262) has a low reliability in the method of judging countersteer. Further, in the high lateral acceleration state, the driving force of the wheel on the inner side of the turn is increased. However, according to such control, the actual lateral acceleration direction and the lateral acceleration direction acting by this control are opposite to each other. May become unstable and control may diverge.

SUMMARY OF THE INVENTION The present invention has been made in view of the above-described problems. In estimating the behavior of a vehicle, the state of the wheels is used as a determination element to determine the drift turning state, so that the running state of the vehicle can be more reliably determined. It is an object of the present invention to provide a power transmission control device for left and right wheels for a vehicle, which is capable of comprehending and solving various problems actually occurring at the time of drift turning to obtain a stable running characteristic of the vehicle. I do.

[0016]

According to a first aspect of the present invention, there is provided a vehicle power transmission control apparatus for a left and right wheel according to the present invention, wherein at least one of the left and right wheels of a front wheel and a rear wheel has a power. A driving force from the power unit is provided on a vehicle configured as a driving wheel to which a driving force is applied from the device, in relation to the left and right driving wheels or between the front and rear driving wheels according to a traveling state of the vehicle. Controlling the transmission of the driving force, or controlling the differential between the left and right driven wheels to which the driving force is not transmitted, or controlling the braking force by a braking device in relation to the front, rear, left and right wheels, Rotational force adjusting means for adjusting the rotational force of the left and right wheels or front and rear wheels or front and rear left and right wheels, slip degree detecting means for detecting the slip degree of the wheels during turning of the vehicle, and detection information of the slip degree detecting means Can be estimated based on It is characterized in that and a control unit capable of carrying out the slip corresponding control for controlling the rotational force adjusting mechanism according to both of the turning state.

When controlling the transmission of the driving force from the power unit in relation to the left and right driving wheels, or controlling the differential between the left and right driven wheels to which the driving force is not transmitted, When the degree of slip is equal to or more than a set value, it is preferable that the rotational force adjusting mechanism is controlled so as to be in a differential inhibition (diff lock) state. According to a second aspect of the present invention, there is provided a vehicle power transmission control apparatus according to the first aspect of the present invention, wherein the degree of slip detecting means detects vertical slip and side slip of the wheel and detects the slip. And the degree of slippage is detected by integrating the slippage and the skidding.

The longitudinal slip is preferably calculated based on a difference between an estimated vehicle speed that can be estimated from the vehicle speed at the time of occurrence of the slip and the longitudinal acceleration of the vehicle, and a vehicle speed that can be estimated only from the wheel speed. . Further, it is preferable that the side slip is calculated based on a dimensionless value of a difference between a calculated lateral acceleration that can be calculated from the steering angle and the vehicle speed and an actually detected actual lateral acceleration.

According to a third aspect of the present invention, there is provided a power transmission control device for a vehicle between left and right wheels according to the first aspect of the present invention, wherein the control means controls the degree of slip detected by the degree of slip detection and other parameters. A start determining means for determining whether or not to start the slip response control based on the slip control is started when the start determination means makes a start determination.

Preferably, the other parameters are a calculated lateral acceleration which can be calculated from the steering angle and the vehicle speed, an actually detected actual lateral acceleration, and a steering angular velocity.
It is preferable that the slip control is started when the degree of slip is equal to or more than a set value, the direction of the calculated lateral acceleration and the direction of the actual lateral acceleration are in the opposite direction, and the steering angular velocity is equal to or more than the set value.

According to a fourth aspect of the present invention, there is provided a power transmission control apparatus for a vehicle between left and right wheels according to the first aspect of the present invention, wherein the control means is configured such that at least the degree of slip detected by the degree of slip detection means is greater than or equal to a set value. At the time of, the start determination means for determining the start of the slip response control, and end determination means for determining the end of the slip response control based on parameters other than the degree of slip, and during the slip response control the Even if the degree of slip falls below the set value, the slip-response control is continued unless the end of the control is determined by the end determination means.

Preferably, when the calculated lateral acceleration and the actual lateral acceleration are in the same direction, it is determined to end the slip control. The power transmission control device for left and right wheels of the present invention according to claim 5 is provided in association with left and right wheels of a front wheel or a rear wheel, and transmits driving force from a power device to the left and right wheels, or A rotational propulsion distribution adjusting mechanism for transmitting and receiving the rotational force associated with the rotation of the left and right wheels between the left and right wheels to change the rotational propulsion of the left and right wheels; and the rotational propulsion distribution adjustment according to the turning state of the vehicle. A control means for controlling a mechanism to adjust the rotational force of the left and right wheels, and a power transmission control device for the left and right wheels for a vehicle, comprising: A target value calculating means for calculating a target value; and a turning state determining means for determining whether the vehicle is in a drift turning state or a non-drift turning state, wherein a difference between the target value and an actual left and right wheel rotational speed is provided. And the latter is compared with Controlling the rotary propulsion force distribution adjusting mechanism by feedback control so as to approach the target value, and setting the target value to zero when the turning state determination means determines that the vehicle is in a drift turning state. Features.

Note that the target value may be set to a value close to zero when the vehicle is in a running state close to the drift turning state. According to a sixth aspect of the present invention, there is provided a vehicle power transmission control apparatus for a vehicle according to the fifth aspect, further comprising a road surface friction coefficient detecting unit for detecting a road surface friction coefficient, wherein the control unit comprises the road surface friction coefficient detecting device. When the road friction coefficient detected by the coefficient detection means is low, the feedback gain is set low. When the road friction coefficient is low and the turning state determination means determines that the vehicle is in the drift turning state, the feedback gain is set. Is increased.

Preferably, when the road surface friction coefficient is high, the feedback gain is set high,
When the road surface friction coefficient is high and the vehicle is in a drift turning state, the feedback gain is reduced. This suppresses response delay due to feedback control during normal control,
Hunting of the control can be prevented during the drift control.

According to a seventh aspect of the present invention, there is provided a power transmission control apparatus for a vehicle between left and right wheels, which is provided in association with left and right wheels of a front wheel or a rear wheel, and transmits a driving force from a power device to the left and right wheels, Alternatively, a rotation propulsion distribution adjusting mechanism for changing the rotation propulsion of the left and right wheels by transmitting and receiving the rotation force accompanying the rotation of the left and right wheels between the left and right wheels, and the rotation propulsion according to the traveling state of the vehicle. Control amount calculating means for calculating an output control amount of the force distribution adjusting mechanism; and control means for controlling the rotary propulsion force distribution adjusting mechanism based on the output control amount calculated by the control amount calculating means. In the power transmission control device for the left and right wheels for a vehicle, the vehicle has a turning state determination unit that determines whether the vehicle is in a drift state or a non-drift state, and a road surface friction coefficient detection unit that detects a road surface friction coefficient, Above control amount The output means corrects the output control amount in accordance with the road surface friction coefficient detected by the road surface friction coefficient detection means when the turning state determination means determines that the vehicle is in the non-drift state. When the state is determined, the correction is prohibited or suppressed.

Preferably, in the non-drift turning state, when the road surface friction coefficient is low, the output control amount is low, and when the road surface friction coefficient is high, the output control amount is increased so that the correction is made. In the turning state, the correction is prohibited or suppressed so that the output control amount takes an intermediate value without corresponding to the road surface friction coefficient.

The power transmission control device for a vehicle according to the present invention is provided in connection with the left or right wheel of a front wheel or a rear wheel, and transmits the driving force from the power device to the left and right wheels. Alternatively, a rotational force accompanying the rotation of the left and right wheels is transmitted and received between the left and right wheels, and a rotational propulsion force distribution adjusting mechanism for changing the rotational propulsion force of the left and right wheels, and a transient control signal based on at least an artificial operation. And a control means for transmitting a control signal to the rotational propulsion force distribution adjusting mechanism in accordance with a running state of the vehicle, wherein the vehicle wheel-side power transmission control device comprises: The slip degree detection means and the control means inhibit transmission of the transient control signal or weaken the transient control in an area where the slip degree detected by the slip degree detection means is larger than a predetermined value. Up It is characterized by modifying the transient control signal.

Preferably, the control means has a steady control means and a transient control means for performing control according to the running state of the vehicle, respectively, and when the slippage exceeds a set value, the transient control means is provided. And the steady control means controls the differential mechanism between the wheels in a direction in which the differential is prohibited (ie, differential lock).

According to a ninth aspect of the present invention, there is provided a power transmission control device for a vehicle between left and right wheels, which is provided in association with left and right wheels of a front wheel or a rear wheel, and transmits a driving force from a power device to the left and right wheels; Alternatively, a rotation propulsion force distribution adjusting mechanism for changing the rotation propulsion force of the left and right wheels by transmitting and receiving the rotation force accompanying the rotation of the left and right wheels between the left and right wheels, and the rotation corresponding to at least the lateral acceleration of the vehicle. A control means for transmitting a control signal to the propulsion distribution adjusting mechanism, wherein a turning state for determining whether the vehicle is in a drift turning state or in a non-drift turning state is provided in a power transmission control device for left and right wheels for a vehicle. When the turning state determining means determines that the vehicle is in a non-drift turning state, the control means based on a reference lateral acceleration calculated from a steering angle and a vehicle speed of the vehicle. System Sends a signal, when the vehicle is determined to be drift turning state is characterized by sending a control signal based on the actual lateral acceleration of the vehicle.

[0030]

〔table of contents〕

1. 1. System outline of this device 1.1 Concept of hardware configuration of this device 1.2 Hardware configuration of this device 1.3 Outline of control of this device 1.4 Functions and effects to be obtained by control of this device 2. 2. Control details of this device 2.1 Input calculation processing 2.2 Drift determination logic 2.3 Vehicle motion control logic 2.3.1 Target ΔN tracking control 2.3.2 Acceleration turning control 2.3.3 Tack-in correspondence control 2 3.3.4 Steering transient response control 2.4 Estimation of road surface μ 2.5 Actuator drive 2.5.1 Proportional valve / directional valve control 2.5.2 Hydraulic pump motor control 2.5.3 Indicator display control 3. Effects of the present apparatus [Embodiment] 1. 1. System overview of this device 1.1 Concept of hardware configuration of this device First, the concept of hardware configuration of this device will be described.
The power transmission control device for left and right wheels for a vehicle, when two rotating bodies disposed coaxially are slid in contact with each other at different rotating speeds, the rotating body having the higher rotating speed has the lower rotating speed. This utilizes the characteristic that torque is transmitted to the rotating body.

That is, as shown in FIG. 5, the rotational speed NL of one (here, the left wheel side) of the left wheel side rotating member and the right wheel side rotating member installed coaxially with each other is the other (here, the right wheel side rotating member). Side), the speed difference at this time, namely, the slip speed (= NL-NR), and the sliding contact force, ie, the pressing force P between the left wheel side rotating member and the right wheel side rotating member. The torque of the magnitude Tc corresponding to the left wheel side (FIG. 5)
It is transmitted from the middle and left sides to the right wheel side (right side in FIG. 5).

As described above, if the left wheel side is larger than the right wheel side between the left wheel side and the right wheel side, it is possible to easily transmit the torque from the left wheel side to the right wheel side. If the right wheel side is larger than the left wheel side, torque transmission from the right wheel side to the left wheel side can be easily performed. Therefore, in order to realize a state in which the left wheel side is larger than the right wheel side even in a region where the left and right wheels rotate at a constant speed, for example, a shift in which the left wheel side rotational speed VL is shifted to the left wheel side at a high speed. If the mechanism is provided, even if the left and right wheels rotate at a constant speed, the left wheel side between the left wheel side member receiving the output of the transmission mechanism and the right wheel side member rotating at the same speed VR as the right wheel. Can be realized in a state where the rotation speed of the wheel is higher than that of the right wheel. Also,
For example, if a transmission mechanism for shifting the rotation speed VR of the right wheel side to a low speed is provided on the right wheel side, even if the left and right wheels are rotating at a constant speed,
A state in which the rotation speed on the left wheel side is higher than that on the right wheel side can be realized between the left wheel side member rotating at the same speed VL as the left wheel and the right wheel side member receiving the output of the transmission mechanism.

If the right wheel side is also configured symmetrically to this, a state where the right wheel side is larger than the left wheel side can always be realized. When turning the vehicle, the turning inner wheel rotates at a lower speed than the turning outer wheel.However, depending on the speed ratio setting of the transmission mechanism, the turning member on the inner wheel side also has a higher speed than the rotating member on the outer wheel side during turning of the vehicle. Can be shifted.

If a torque transmission coupling is provided between the left wheel side rotating member and the right wheel side rotating member provided with the speed difference as described above, the torque transmitting coupling can be operated appropriately. Under certain traveling conditions, torque can be constantly transmitted from the left wheel side to the right wheel side and from the right wheel side to the left wheel side. Of course, even when turning at the maximum steering angle, the drive torque on the inner wheel side is transmitted to the outer wheel side,
By setting the gear ratio by the transmission mechanism, under all driving conditions,
Torque can be constantly transmitted from the left wheel side to the right wheel side and from the right wheel side to the left wheel side.

In a coupling of variable torque transmission capacity such as a wet multi-plate clutch mechanism, the amount of transmission torque can be adjusted according to the engagement pressure (pressing force P) and the like.
Incidentally, the transmission mechanism and the coupling interposed between the right wheel side and the left wheel side are provided directly between the right wheel side and the left wheel side. By providing these transmission mechanisms and couplings between the wheel side (right wheel side or left wheel side), power transmission between the left wheel side and right wheel side via an input part of the differential ( Torque transfer).

The power transmission (torque transfer) between the left and right wheels of the vehicle according to such a principle can be applied whether the left or right wheel is a driving wheel or a driven wheel.
The distribution of the driving force from the engine to the left and right wheels will be adjusted. If the left and right wheels are driven wheels, the torque transmitting wheels will receive the driving force by torque transmission, and the torque transmission The wheel on the side that performs the braking receives the braking force. In any case, the magnitude of the driving force or the braking force exerted between each of the left and right wheels and the road surface is imbalanced in the left and right directions, thereby generating a yaw moment in the vehicle and thereby changing the behavior of the vehicle. Can be controlled.

1.2 Hardware Configuration of the Apparatus Next, the hardware configuration of the power transmission control apparatus for left and right wheels for a vehicle based on such a theory will be described with reference to FIGS. 1.2.1 Configuration of Power Transmission System for Vehicle According to Apparatus The power transmission control apparatus for left and right wheels for a vehicle according to the present embodiment is shown in FIG.
As shown in the figure, it is provided on the rear wheel of a four-wheel drive vehicle.

In FIG. 1, reference numeral 2 denotes an engine,
The output of the engine 2 is transmitted to a differential gear mechanism (= center differential, hereinafter, referred to as a center differential) 8 via a transmission 4 and an intermediate gear mechanism 6. The output of the center differential 8 is transmitted to the axle 1 via a front differential gear mechanism (= front differential, hereinafter referred to as front differential) 10 on the one hand.
2L, 12R to the left and right front wheels 14, 16, while the bevel gear mechanism 18, the propeller shaft 20
The transmission is transmitted from the axles 26L, 26R to the left and right rear wheels 28, 30 via a bevel gear mechanism 22, and a rear wheel differential gear device (= rear differential, hereinafter, referred to as a rear differential) 24. The rotary propulsion force distribution adjusting mechanism (or torque adjusting means, hereinafter, referred to as a torque transfer mechanism) 50 of the power transmission control device between the left and right wheels is provided in the rear differential 24.

The center differential 8 is composed of differential pinions 8A and 8B and side gears 8C and 8D meshed with the differential pinions 8A and 8B, similarly to the conventionally known ones. The torque is applied to the side gear 8
C, 8D, from the side gear 8C to the front wheel side,
The differential is transmitted from the side gear 8D to the rear wheel side while being allowed.

Here, from the side gear 8C to the front differential 10 on the front wheel side via the front wheel output shaft 32, from the side gear 8D, the rear wheel output shaft 34 and the bevel gear mechanism 1
8, torque is transmitted from the propeller shaft 20 to the rear wheel side. The center differential 8 restricts (or restricts) the differential between the front-wheel output portion and the rear-wheel output portion to distribute the output torque (rotational propulsion) of the engine between the front and rear wheels. A viscous coupling unit (VCU) 36 is provided as a controllable differential limiting means (ie, a limited slip differential (LSD)).

This VCU 36 is interposed between the front wheel output shaft 32 and the rear wheel output shaft 34, and limits the differential between the front wheel side and the rear wheel side with a force according to the differential state. by doing,
It is possible to avoid a situation in which only the light load side of the front and rear wheels idles and the rotational torque is not transmitted to the heavy load side. 1.2.2 Configuration of Rotational Propulsion Force Distribution Adjusting Mechanism of Present Device The present left and right wheel power transmission control device includes a rotational propulsion force distribution adjusting mechanism (torque moving mechanism) 50 provided in a differential carrier 51, It comprises a hydraulic unit 38 and an electronic control unit (hereinafter referred to as ECU) 42 as control means (or rotational propulsion force distribution control means). Here, the rear differential 24 and the rear differential 24 and the axle 26
The configuration of the torque transfer mechanism 50 inserted between the L and 26R will be described with reference to FIG.

As shown in FIG. 2, an input shaft 52 is connected to the rear end of the propeller shaft 20, and a drive pinion gear 54 is connected to the input shaft 52 so as to rotate integrally. The drive pinion gear 54 meshes with a crown gear 56 provided on the outer periphery of a differential case (differential case) 58. The output of the engine is transmitted from the input shaft 52 to the rear differential 24 via the drive pinion gear 54 and the crown gear 56. It is being conveyed.

The rear differential 24 has two pairs of pinions provided in a differential case 58, that is, differential pinions 60A and 60B, and side gears 62 and 64 meshed with these differential pinions 60A and 60B, similarly to the conventional one. The rotational torque input from the differential pinions 60A and 60B is
The differential gears are transmitted to the side gears 62 and 64 and transmitted from the side gear 62 to the left-wheel rotating shaft 66 and from the side gear 64 to the right-wheel rotating shaft 68 while allowing the respective differentials to be allowed. . Also, left and right rotating shafts 66,
1, 68 is connected to axles 26L, 26R connected to the left and right rear wheels 28, 30 as shown in FIG.

The torque transfer mechanism 50 of the present embodiment is provided between the differential case 58 of the rear differential 24 for distributing the driving force between the left and right driving wheels of the rear wheel and the right wheel side rotating shaft 68.
Transmission mechanism 70 and transmission capacity variable control torque transmission mechanism 90
The transmission of rotational propulsion between the left wheel side and the right wheel side, that is, power transmission (torque movement), is performed via the differential case 58.

The speed change mechanism 70 includes a speed increasing mechanism 70A for increasing the rotation speed of the input portion of the rear differential 24, that is, the rotation speed of the differential case 58, and outputting the rotation to one side of the left and right wheels (here, the right wheel side).
Since it is integrally provided with a speed reduction mechanism 70B for decelerating and outputting to one side (right wheel side), it is also referred to as an acceleration / deceleration mechanism. The transmission capacity variable control type torque transmission mechanism 90 uses a wet hydraulic multi-plate clutch mechanism (hereinafter, also referred to as a clutch) capable of adjusting the transmission capacity according to the control oil pressure.
A clutch (left clutch) 90L provided between the output side of the speed reduction mechanism 70B of the transmission mechanism 70 and the right wheel side to transmit torque to the left wheel side, and the output side of the speed increasing mechanism 70A of the transmission mechanism 70 and the right wheel. A clutch (right clutch) 90 </ b> R provided between the right side and the right side to transmit torque is integrally formed. Such a transmission capacity variable control type torque transmission mechanism 90 is also referred to as an integral coupling or simply a coupling.

The acceleration / deceleration mechanism 70 will be described. The acceleration / deceleration mechanism 70 includes a hollow intermediate shaft 72 connected to rotate integrally with the differential case 58, a hollow intermediate shaft 74 connected to the right clutch 90R, It is interposed between a hollow intermediate shaft 76 connected to the left clutch 90L. Each of the intermediate shafts 72, 74, 76 is a hollow shaft, and the intermediate shafts 72, 74 are mounted on the outer periphery of the right wheel side rotation shaft 68 so as to be able to relatively rotate. 7
4 is also provided on the outer circumference so that it can also be rotated relative to each other.

Gears 78A, 80A, 82A are provided on these intermediate shafts 72, 74, 76, respectively, and a counter shaft 84 is provided on the outer periphery of these intermediate shafts 72, 74, 76. This counter shaft 84
Is provided with a triple gear 86. Triple gear 86
Is composed of gears 78B, 80B, 82B,
8B is the gear 78A of the intermediate shaft 72, the gear 80B is the gear 80A of the intermediate shaft 74, and the gear 82B is the gear 8 of the intermediate shaft 76.
2A, respectively.

The acceleration / deceleration mechanism 70 is provided with such a gear 78
Intermediate shafts 72, 74, 76 having A, 80A, 82A
, Counter shaft 84, gears 78B, 80B, 8
And a triple gear 86 having 2B. As shown in FIG. 3, a plurality (three in this case) of countershafts 84 are provided on the outer periphery of the intermediate shafts 72, 74, 76 with a phase shifted from that of the drive pinion 54.
Thus, although the ring gear is not provided, the gear 78A,
80A and 82A as sun gears and gears 78B and 80B,
The arrangement is similar to that of a triple planetary gear mechanism using 82B as a planetary pinion.

Each counter shaft 84 is fixed to a wall 51A provided on the differential carrier 51.
Therefore, the gears 78B, 80B, and 82B only rotate about the counter shaft 84 as the axis. This allows
The intermediate shafts 72, 74, 76 are supported in the radial direction by gears 78A, 80A, 82A and gears 78B, 80B, 82.
Through the engagement with B, it is also performed by the plurality of counter shafts 84 fixed to the wall 51A as described above.

In FIG. 3, reference numeral 96 denotes a roller bearing. Then, the gears 78A, 80A, when 82A number of teeth of the with Z _1, Z _2, Z ₃ respectively, are set to satisfy the relationship of _{Z 2 <Z 1 <Z 3} . Also, the gears 78B, 80B, 82
Assuming that the number of teeth of B is Z ₄ , Z ₅ , and Z ₆ , Z ₆ <
Z ₄ <Z ₅ is set.

Thus, the transmission mechanism (acceleration / deceleration mechanism) 70
Then, gear 78A, gear 78B, gear 80A, gear 80
The combination of gears B constitutes a speed increasing mechanism 70A for increasing the rotation input to the rear differential 24 and outputting the rotation to the right wheel side. The combination of the gears 78A, 78B, 82A, and 82B is input to the rear differential 24. A deceleration mechanism 70B that decelerates the applied rotation and outputs the decelerated rotation to the right wheel side is configured.

That is, in the transmission mechanism (acceleration / deceleration mechanism) 70, when the differential case 58 is rotated by the rotational torque input to the rear differential 24, the rotation of the differential case 58
The power is transmitted from the gear 78A to the plurality of gears 78B on the outer periphery of the gear 78A via the intermediate shaft 72. And each gear 78
B together with each gear 80B, 82B
, The gears 80A and 82A meshing with the gears 80B and 82B rotate.

At this time, the gears 78B, 80B, 82B rotate at a constant speed integrally, but these gears 78B, 80B,
The gears 78A, 80A, and 82A that mesh with the gear 82B rotate at mutually different speeds depending on the setting of the fraction as described above. That is, the gears 78A, 80 related to the speed increasing mechanism 70A.
A, 78B, and 80B, the numbers of teeth Z ₁ and Z ₂ of the gears 78A and 80A are in a relation of Z ₁ > Z ₂ , and the gear 78
Since the numbers of teeth Z ₄ and Z ₅ of B and 80B satisfy the relationship of Z ₄ <Z ₅ , the gear 80A is rotated at a higher speed than the gear 78A.

Considering the speed increase ratio in this case, that is, the rotational speed ratio of the gear 80A to the gear 78A,
The ratio of the rotation speed of the gear 80B (that is, the rotation speed of the triple gear 86) to the rotation speed of the gear 78A (the value of the rotation speed of the gears 78B and 80B when the gear 78A makes one rotation) is Z ₁ / Z _4. , The rotation speed of the gear 80A and the rotation speed of the gears 78B, 80B (ie,
The ratio (the number of rotations of the gear 80A when the gears 78B and 80B make one rotation) is Z ₅ / Z.
₂ , and the rotation speed ratio of the gear 80A to the gear 78A is (Z ₁ · Z ₅ ) / (Z ₂ · Z ₄ ).

The gear 78 relating to the speed reduction mechanism 70B
A, 82A, 78B, and 82B are gears 78A,
Number of teeth Z _1, Z ₃ of 82A are in relation of Z ₁ <Z _3,
Since the numbers of teeth Z ₄ and Z ₆ of the gears 78B and 82B satisfy the relationship of Z ₄ > Z ₆ , the gear 82A rotates at a lower speed than the gear 78A. The reduction ratio in this case, that is, the gear 8
Considering the rotation speed ratio of the gear 2A to the gear 78A, the ratio between the rotation speed of the gears 78B and 82B (that is, the rotation speed of the triple gear 86) and the rotation speed of the gear 78A (the gear 78B when the gear 78A makes one rotation). , 82B) is Z ₁ /
Z ₄ , the ratio of the rotation speed of the gear 82A to the rotation speed of the gears 78B and 82B (ie, the rotation speed of the triple gear 86) (the gear 78)
(The value of the rotation speed of the gear 82A when B and 82B make one rotation)
Is Z ₆ / Z ₃ , and the rotation speed ratio of the gear 82A to the gear 78A is (Z ₁ · Z ₆ ) / (Z ₃ · Z ₄ ).

By the way, the transmission capacity variable control type torque transmission mechanism 90 to which the output of the acceleration / deceleration mechanism 70 is input,
That is, the left clutch 90L and the right clutch 90R are
As shown in the figure, the differential gear 51 is installed in a space on the right wheel side of the acceleration / deceleration mechanism 70 in the differential carrier 51. These hydraulic multi-plate clutches 90L, 90R are connected to clutch plates 90AL, 90AR, which are connected to a clutch case 92 so as to rotate integrally with the right wheel side rotation shaft 68, and to rotate integrally with the intermediate shafts 74, 76. Clutch plates 90BR, 90B
L, and two pistons (not shown) for applying hydraulic pressure (clutch pressure) to the clutches 90L and 90R, respectively. The drive hydraulic pressures of the two hydraulic pistons are adjusted through the hydraulic unit 38 by electronic control of the controller 42. The engagement state of the clutches 90L and 90R is adjusted.

The left clutch 90L includes a right wheel side clutch plate 90AL that rotates integrally with the right wheel side rotation shaft 68, and an intermediate shaft 76.
And a clutch plate 90BL on the output side of the speed reduction mechanism 70B that is coupled so as to rotate integrally. Since the clutch plate 90BL rotates integrally with the gear 82A reduced by the reduction mechanism 70B together with the intermediate shaft 76, the clutch plate 90B is rotated unless the speed ratio of the left wheel to the right wheel increases.
L rotates at a lower speed than the right wheel side clutch plate 90AL that rotates integrally with the right wheel side rotation shaft 68.

Therefore, if the clutch 90L is engaged, even if the right wheel is rotating at a lower speed than the left wheel during a right turn, the clutch plate 90AL is
The torque is transmitted to the BL, that is, from the right wheel to the input side of the rear differential, and the amount of torque from the engine to the right wheel is reduced, and the amount of torque to the left wheel is increased. Can be done.

The right clutch 90R has a right wheel side clutch plate 90AR which rotates integrally with the right wheel side rotation shaft 68, and a speed increasing mechanism 70A which is connected so as to rotate integrally with the intermediate shaft 74.
And an output side clutch plate 90BR. The clutch plate 90BR includes the intermediate shaft 74 and the speed increasing mechanism 70A.
As a result, the clutch plate 90BR rotates at a higher speed than the right wheel side clutch plate 90AR which rotates integrally with the right wheel side rotation shaft 68 unless the speed ratio of the right wheel to the left wheel increases. I do.

Therefore, if the clutch 90R is engaged, even if the left wheel is rotating at a lower speed than the right wheel during a left turn, the clutch plate 90BR side to the right wheel side clutch plate 90
The torque is transmitted to the AR side, that is, from the input portion side of the rear differential to the right wheel side, so that the amount of torque from the engine to the right wheel side is increased, and the amount of distribution to the left wheel side is increased. Can be reduced.

The transmission capacity variable control type torque transmission mechanism may be any mechanism capable of variably controlling the transmission torque capacity. In addition to the mechanism of this embodiment, another wet type such as an electromagnetic hydraulic multi-plate clutch mechanism may be used. In addition to the multi-plate clutch mechanism, these multi-plate clutch mechanisms, a hydraulic or electromagnetic friction clutch, a hydraulic or electromagnetic type controllable VCU (Viscous Coupling Unit), a hydraulic or electromagnetic type A controllable HCU (Hydric coupling unit = differential pump type hydraulic coupling), and other couplings such as an electromagnetic fluid type or an electromagnetic powder type clutch can also be used.

Incidentally, the wet hydraulic multi-plate clutch mechanism 90
When L and 90R are engaged, torque is transmitted according to the magnitude of the differential amount between the engaged clutch plates (90AL and 90BL, 90AR and 90BR) and the strength of engagement.
That is, if the control oil pressure is adjusted in consideration of the differential amount between the clutch plates to adjust the strength of engagement of the clutch mechanisms 90L and 90R, the amount of torque movement can be reliably controlled.

Therefore, the hydraulic pressure adjusting units for the clutches 90L and 90R in the hydraulic unit 38 are also controlled through the ECU 42 so that the torque distribution to the left and right rear wheels is in a desired state. In this case, the ECU 42 controls required parts of the hydraulic unit 38 based on engine information, wheel speed information, steering wheel angle information (steering angle information), information on the lateral acceleration and longitudinal acceleration of the vehicle body, and the like.

For example, when it is desired to distribute the drive torque from the input shaft 52 to the left wheel rotating shaft 66 more, the left clutch 90L can be engaged with an appropriate control pressure in accordance with the degree of the distribution (distribution ratio). When it is desired to distribute the drive torque from the input shaft 52 to the right wheel rotating shaft 68 more, the right clutch 90R may be engaged with an appropriate control pressure in accordance with the degree of distribution (distribution ratio).

Further, the clutches 90L and 90R are set so that they are not completely engaged at the same time.
When one of the left and right clutches 90L, 90R is completely engaged, the other is not engaged. That is, the operation mode of the clutches 90L and 90R is
There is a mode in which only the right clutch 90R is engaged, and a neutral mode in which neither is engaged.

As described above, in the torque transfer mechanism 50, the distribution of the left and right torques can be adjusted by moving the torque, so that the torque loss is extremely small compared to the case where the distribution of the right and left torques is adjusted simply by braking one wheel. There is a feature that the torque distribution can be adjusted over a wider range, and a yaw moment can be generated in the vehicle without discomfort, for example.

1.2.3 Configuration of Hydraulic Unit According to the Present Apparatus Here, the configuration of the hydraulic unit 38 will be described with reference to FIG. As shown in FIG. 4, the hydraulic unit 38 includes a pressure accumulating unit 101 for accumulating hydraulic oil and a hydraulic oil accumulated in the pressure accumulating unit 101 by appropriately adjusting the pressure of the clutch 90L, 90R.
And a control pressure output unit 102 for supplying the oil pressure to an oil chamber (not shown).

The accumulator 101 includes an accumulator 103
And a motor pump 104 for pressurizing the working oil in the accumulator 103 to a predetermined pressure, and a pressure switch 105 for monitoring the differential oil pressure pressurized by the motor pump 104. Further, the control pressure output unit 102 is connected to the motor pump 1
An electromagnetic proportional pressure control valve (abbreviated as a proportional valve) 106 for adjusting the pressure of the hydraulic oil in the accumulator 103 whose pressure has been adjusted through the pressure valve 04, and a hydraulic oil adjusted by the proportional valve 106 are connected to the left or right clutch 90L, 90R. And an electromagnetic directional control valve (directional switching valve) 107 for switching between supply to an oil chamber (not shown).

The hydraulic unit 38 is provided by the ECU 4
2, the ECU 42 includes a wheel speed sensor 48A, a steering wheel angle sensor 48B, a longitudinal acceleration sensor (longitudinal G sensor) 48C, a lateral acceleration sensor (lateral G sensor) 48D, and a throttle position sensor (TPS).
48E and sensors such as the pressure switch 105 are connected.

Based on the information from these sensors, the ECU 42 changes the motor pump 104 and the proportional valve of the hydraulic unit 38 according to the running state of the vehicle, that is, the vehicle speed, the steering state, and the vehicle body movement state. 106 and the direction switching valve 107 are controlled. This proportional valve 1
06 and differential limiting control through the control of the direction switching valve 107,
That is, the details of the torque transfer control will be described later.

In FIG. 4, reference numeral 108 denotes a battery,
Reference numeral 09 denotes a motor relay, which controls the motor pump 104 by controlling the supply of electric power from the battery 108 through the motor relay 109, and manages the accumulated pressure by the accumulator 101 based on the detection information of the pressure switch 105. The operation of the motor pump 104 is controlled through the relay 109 while being controlled. Reference numeral 110 denotes an indicator lamp that indicates whether or not differential limiting control by the hydraulic unit 38, that is, torque movement control is being performed.

Since the differential limiting control via the hydraulic unit 38 needs to be linked with the engine output control, the ECU 42 outputs a control command to the hydraulic unit 38 and controls the engine output control. The output reduction information is also sent to an engine ECU (not shown). The ECU 42 includes a CPU (not shown), a ROM, a RAM,
It has an interface.

1.3 Outline of Control of the Present Apparatus Here, an outline of control of the present apparatus will be described with reference to a control block diagram showing a functional configuration relating to control of the present apparatus in FIG. As shown in FIG. 6, the process according to the present control includes a sensor input process for receiving a sensor input, a calculation process for performing various calculations based on these sensor input values, and a vehicle motion based on the calculation process result. Control amount calculation processing for calculating each control amount of control and drive processing for driving each actuator based on each calculated control amount can be divided.

Of these, in the sensor input processing, four wheel speed sensors 48A, steering wheel angle sensors 48B, longitudinal acceleration sensors (longitudinal G sensors) 48C, lateral acceleration sensors (lateral G sensors) 48D, and throttle position sensors (TP)
S) Receive a sensor input from 48E or the like. In the arithmetic processing, an actual measurement value and a theoretical value of the speed difference between the right and left rear wheels are calculated. The actual measurement value (actual rotation speed difference) is based on the wheel speed values from the four wheel speed sensors 48A,
The theoretical value (target value, theoretical rotational speed difference) is calculated by the steering wheel angle sensor 4
Based on the steering angle from 8B and the vehicle speed (vehicle speed) obtained from the wheel speed values from the four wheel speed sensors 48A,
Each is calculated. The front and rear G sensor 48C, the horizontal G
G (g) before and after the calculation based on the detection value from the sensor 48D
b), calculation horizontal G (gy) is calculated. In the arithmetic processing, drift determination and road surface μ estimation are further performed.

In the control amount calculation process, each control amount of the motion control of the vehicle is calculated based on each calculation result.
In this control, the control amount (target ΔN follow-up control amount) of the target rotation speed difference follow-up control (target ΔN follow-up control) related to the control during normal turning, and the control amount (acceleration turn control amount) of the acceleration turning control during the acceleration turn And the control amount of the tuck-in corresponding control at the time of tack-in of the vehicle (tack-in corresponding control amount)
And a control amount (steering transient response control amount) of the steering transient response control relating to the steering transition time are provided, and the respective control amounts are added to output the added control amount (output control amount). ing.

The function of performing the control amount calculation process is called control amount calculation means. Among these functions (control amount calculation means), the function of setting the control amount related to the target ΔN follow-up control,
Alternatively, this setting function and a function of outputting a control signal based on the control amount obtained by this setting are performed by ΔN following control means or target rotational speed difference following control means, a function of setting a control amount of acceleration turning control, or The setting function and the function of outputting a control signal based on the control amount obtained by this setting include acceleration / turn control means, the function of setting the control amount of tack-in correspondence control, or the setting function and the control amount obtained by this setting The function of outputting a control signal on the basis of the tack-in correspondence control means, the function of setting the control amount (transient control amount) of the steering transient response control related to the steering transient, or the setting function and the control amount obtained by this setting The function of outputting a control signal based on this is referred to as steering transient response control means. In particular, the steering transient response control means is also referred to as transient control means, and the control function other than during transient is also referred to as steady control means.

The target ΔN tracking control has a function (target value calculating means) for calculating or storing a target value of a yaw angle or a difference between left and right wheel rotational speeds corresponding to a turning state of the vehicle. It has a function of calculating a control amount according to the target value at the time (steady turning control means). In the drive processing, a proportional valve output for outputting a command signal to the proportional valve 106 to adjust the amount of torque movement, and a command signal to a directional valve (direction switching valve) 107 for setting the direction of torque movement. A direction valve output and an indicator display output for outputting a display command signal to the indicator lamp 110 are performed.

[0086] Each of these processes will be described in detail below, but before that, the functions and effects to be obtained by controlling the present apparatus will be described. 1.4 Actions and Effects to be Obtained by Control of the Apparatus The present apparatus is capable of (1) improving the turning performance, (2) ensuring the stability of the vehicle at the time of turning deceleration, and (3) improving the starting performance and acceleration performance. It was developed with the aim of improvement, and its control principle will be described from these viewpoints.

1.4.1 Improvement of Turning Performance During acceleration turning, the steering characteristics of the vehicle increase toward the understeer side, so that it is difficult to make a sharp turn intended by the driver. In particular, when the driving force of the engine is also transmitted to the front wheels, the cornering force generated on the front wheels according to the driving force load on the front wheels is reduced, and understeering is likely to increase.

Therefore, at the time of turning, particularly at the time of accelerating turning,
By moving the torque toward the turning outer wheel, a yaw moment is generated in the turning direction to increase the cornering force of the front wheel in a region where the longitudinal acceleration is large, thereby suppressing understeering. Thus, when the traveling locus of the vehicle at the time of the acceleration turning is compared when the same acceleration turning operation is performed, as shown in FIG. 7, the state is improved from the state without the control to the state with the control.

FIG. 8 shows an example of a steering ratio (= θh / θh0, θh: actual steering wheel angle, θh0: steering wheel angle theoretically required) according to the lateral G (turning G) generated at the time of turning. The region where the steering ratio sharply increases in the figure corresponds to the turning limit region. As shown in the figure, it is understood that the turning limit is improved by controlling the torque transfer to the turning outer wheel.
In addition, the start of control for improving the turning performance is as follows.
By performing the control when the steering ratio non-linearly increases without control, the control frequency can be reduced.

1.4.2 Ensuring Vehicle Stability During Turning Deceleration At the time of decelerating turning, contrary to the acceleration turning, the steering characteristics of the vehicle are strengthened toward the oversteer side, so that the vehicle is liable to cause tack-in. Therefore, at the time of decelerating turning, contrary to the accelerating turning, the torque is moved to the turning inner wheel side, thereby generating a yaw moment in the turning suppressing direction, reducing the cornering force of the front wheel and suppressing oversteering. . As a result, the tack-in of the vehicle is suppressed, as shown in FIG. 9 from the state indicated as no control to the state indicated as control.

FIG. 10 shows an example of the time change of the yaw rate. The solid line indicates control and the chain line indicates no control. As shown in the figure, immediately after the accelerator was released, it was found that the yaw rate increased without control and the vehicle attitude changed suddenly.With control, the yaw rate converged smoothly without increasing, and the stability of the vehicle attitude was maintained. It is understood that it is done.

1.4.3 Improvement of Turning Performance During Drift Driving In this control, it is determined whether or not the vehicle is about to drift, and this determination result is used for turning control of the vehicle through torque movement control of the left and right wheels. Used. That is, when the vehicle drifts, the wheels are skidding, and the traveling state of the vehicle is a traveling state that is significantly different from the normal state. Therefore, the target value [rear wheel reference rotation speed difference (dvh) of the rotation speed difference (rotation speed difference) of the left and right wheels, which is the torque movement control capable of adjusting the steering characteristic as described above or the more basic torque movement control.
f)], so that more appropriate control can be performed by reflecting the result of the determination as to whether or not the vehicle is drifting in the turning control such as the torque transfer control according to the above. That is,
When drifting, control is performed on the side that limits the differential between the left and right wheels, so that the generation of turning force due to torque transfer control is suppressed, and the grip force of the wheel with reduced grip force is restored. You do it. This allows
The purpose is to improve turning performance during drift running.

1.4.4 Improvement of Startup Performance and Acceleration Performance In this device, torque control is performed so that the start performance and acceleration performance can be improved when the frictional resistance of the road surface differs between the left and right wheels. In other words, when the road surface contacted by one of the left and right wheels is lower μ than the road surface contacted by the other wheel, the wheel on the lower μ side during starting (this is called μ split start) or during acceleration Since the load on the wheel becomes smaller, the differential mechanism works, and the driving force on the wheel on the low μ side increases, while the driving force on the wheel on the high μ road side decreases.

As a result, the wheels on the low μ side slip and cannot sufficiently transmit the driving force to the road surface while being supplied with most of the driving force from the engine, and the wheels on the high μ road side receive the driving force from the engine. Since almost no driving force is supplied, as shown in FIG. 11 without control, the driving force cannot be sufficiently transmitted to the road surface, and the vehicle does not travel easily.

Therefore, in such a μ split state, the torque is moved from the low μ wheel side to the high μ road wheel side. As a result, as shown in FIG. 11 as having control, the driving force transmitted from the wheels on the high μ road side to the road surface is increased, and the vehicle can be started and accelerated more quickly and efficiently. it can. For example, FIG. 12 shows an example of a time change of the acceleration G due to the μ-split start. The solid line indicates control and the chain line indicates no control. As shown, it can be seen that the acceleration G is improved by the torque transfer control.

2. Control Contents of the Apparatus Here, the contents of the above-described torque control will be described more specifically in the order of input calculation processing, drift determination logic, vehicle motion control logic, road surface μ estimation, and actuator drive. 2.1 Input Calculation Processing In the input calculation processing, as shown in FIG. 13, the rear left wheel speed vrl, the rear right wheel speed vrr, the steering wheel angle θh, the vehicle body speed vb, the steering wheel angular speed dθh, and the front left wheel speed vf
1, a detection signal relating to the front right wheel speed vfr is received from each sensor, and a previous calculated value (torque movement amount ta,
In response to a road surface μ determination coefficient γ) and detection signals from a pressure switch, an idle switch, a lateral G sensor, a TPS (throttle position sensor) and the like, the following numerical values are calculated.

2.1.1 Rear wheel left / right speed difference (dvrd) The difference between the rear left wheel speed vrl and the rear right wheel speed vrr is calculated, and the actual rear wheel left / right speed generated during turning or torque transfer control is calculated. The speed difference dvrd (= vrl-vrr) is obtained. 2.1.2 Digital filter value (d
vrf) Since the actual speed difference dvrd is used to determine the operating state of the torque transfer control, the actual speed difference dvrd is filtered by a digital filter to remove noise effects. Here, filter processing is performed as shown in equation (2.1.2.2) for performing smoothing processing as in equation (2.1.2.1).

Dvrf1 = (dvrd + odvrd) / 2 (2.1.2.1) dvrf = LPF [dvrd] = LPF [dvrf1, dvrf] (2.1.2.2) where dvrd: holds the previous dvrd Value dvrfl: Smoothed value 2.1.3 Average rear wheel speed (vr) By averaging rear left wheel speed vrl and rear right wheel speed vrr, average rear wheel speed vr [= (vrl + vrr)
/ 2] to determine the operating state of the torque transfer control.

2.1.4 Estimated Vehicle Speed (vb), Turning Radius (RR) The estimated vehicle speed vb is calculated based on the third highest wheel speed v3 among the four left and right front wheels and the right and left rear wheels. . This is because this vehicle is a four-wheel drive vehicle, so that each wheel becomes a drive wheel, and such a drive wheel slides on the road surface when transmitting driving force to the road surface. If the vehicle speed is obtained, even if it is slight, it will be faster than the actual vehicle speed. Therefore, the slowest wheel speed among the four drive wheels most corresponds to the actual vehicle speed. However, it is conceivable that the detected value of the wheel speed may not be an appropriate value due to noise or the like. Therefore, in consideration of the reliability of the detected value, the second slowest wheel speed (ie, the third wheel speed) among the four drive wheels is considered. The estimated vehicle speed vb is obtained by adopting the faster wheel speed v3.

When the vehicle is traveling straight, the wheel speed and the vehicle speed correspond at a fixed ratio. For example, the vehicle speed (simple calculated vehicle speed) v obtained by multiplying the wheel rotation speed by the wheel outer peripheral length is obtained.
Although bd can be the vehicle speed vb, a correction according to the turning radius is required at the time of turning. That is, at the time of turning, the third highest wheel speed is considered to be the inner wheel of the rear wheel, and the inner wheel side is considered to be the simply calculated vehicle speed vbd. Is required from the relationship. That is, the turning radius RRi on the inner wheel side can be calculated by the following equation (2.1.3.1) based on the vehicle speed vbd on the inner wheel side.

RRi = (1 + A * vbd ² ) * Lw / δ = (1 + A * vbd ² ) * Lw * GR / θh (2.1.3.1) where δ: actual steering angle (= θh / GR) A : Stability factor Lw: Wheel base Lt: Tread width GR: Handle gear ratio The ratio between the vehicle body speed vbd and the vehicle body speed vb is equal to the ratio between the inner wheel side turning radius RRi and the vehicle center center turning radius RR. The turning radius RR is calculated by the following equation (2.1.
4.1), the vehicle speed vb is divided into the following formula (2.1.
As in 4.2) to (2.1.4.4), it can be obtained from the vehicle speed vbd and the steering wheel angle θh.

RR = RRi + Lt / 2 (2.1.4.1) Right turn vb = (RRi + Lt / 2) / RRi * vbd (2.1.4.2) Straight ahead vb = vbd (2.1. 4.3) When turning left vb = (RRi−Lt / 2) / RRi * vbd (2.1.4.4) Note that the turning radius RR at the center of the vehicle body is such a vehicle speed vb
Based on the following equation (2.1.4.5) can be expressed.

RR = (1 + A * vb ² ) * Lw / δ = (1 + A * vb ² ) * Lw * GR / θh (2.1.4.5) 2.1.5 longitudinal acceleration (gx) As described in (2.1.5.1), the longitudinal acceleration gxd calculated from the change in the simply calculated vehicle body speed vbd calculated in a predetermined cycle is greatly fluctuated. .5.2)] to obtain the longitudinal acceleration gx.

Gxd = vbd-ovbd (2.1.5.1) where ovbd: a simple calculated vehicle speed vbd before one cycle or a predetermined cycle gx = LPF [gxd] (2.1.5.2) 2.1. 6. Reference lateral acceleration (gy) The reference lateral acceleration (gy) can be calculated from the radius RRi and the estimated vehicle speed vb, assuming the centrifugal force acting on the vehicle when turning, and the radius RRi is calculated from the steering wheel angle θh as described above. Therefore, the reference lateral acceleration (gy) can be calculated by the following equation (2.1.6.1).
Is calculated from the steering wheel angle θh and the estimated vehicle speed vb. This reference lateral acceleration (gy) is also referred to as a calculated lateral G.

Gy = vb ² / RR = vb ² * θh / [(1 + A * vb ² ) * Lw * GR] (2.1.6.1) 2.1.7 Rear wheel reference rotational speed difference (dvhf) Rear wheel The reference rotational speed difference dvhf is determined by a turning radius RR during turning.
The rotational speed difference of the rear wheel can be geometrically calculated from the relationship shown in FIG. 15 according to the equation (2.1.4.5). The rotation speed difference dvh is calculated by the function of the estimated vehicle speed vb and the steering wheel angle θh.
Find r. The above-mentioned rear left wheel speed vrl and rear right wheel speed vrr have been subjected to low-pass filter processing, and in order to match the phases thereof, the rotation speed difference dvhr is processed by a low-pass filter [see (2.1.7.2)]. , The rear wheel reference rotational speed difference dvrf is obtained. The function of calculating the rear wheel reference rotation speed difference dvrf is referred to as target value calculation means.

Dvhr = Lt * vb / RR = Lt * vb * θh / [(1 + A * vb ² ) * Lw * GR] (2.1.7.1) dvhf = LPF [dvhr] (2.1.7.2) 2.1.8 Front Wheel Reference Rotation Speed Difference (dvhff) The front wheel reference rotation speed difference dvrfff is determined by the turning radius R at the time of turning.
The rotational speed difference between the front wheels, which can be geometrically calculated from the relationship shown in FIG.
Using the relationship of 5), first, as in the following equation (2.1.8.1),
A rotation speed difference dvh is obtained from a function of the estimated vehicle speed vb and the steering wheel angle θh, and is processed by a low-pass filter [see (2.1.8.2)] to obtain a front wheel reference rotation speed difference dvr.
ff.

Dvrf = Lt * vb * cos (θh / GR) / RR = Lt * vb * cos (θh / GR) * [θh / [(1 + A * vb ² ) * Lw * GR] (2.1. 8.1) dvrfff = LPF [dvrf] (2.1.8.2) 2.1.9 Front wheel left / right speed difference (dvfd) The difference between front left wheel speed vfl and front right wheel speed vfr is calculated, and the vehicle is turning. And the actual speed difference dvf between the left and right rear wheels
d (= vfl-vfr) is obtained.

2.1.10 Amount of Torque Movement (taf: Temporary Delay Value) Since the torque movement has a time delay from when its command value is output to when it appears as actual vehicle behavior, the torque movement command value A low-pass filter is applied to ta to adjust the phases (see (2.1.10.1)) to obtain a torque movement amount taf. taf = LPF [ta] (2.1.10.1) 2.2 Drift determination logic In this control, it is determined whether or not the vehicle is about to drift, and the result of this determination is determined by the vehicle through the torque movement control of the left and right wheels. It is used for motion control. Therefore, in this control,
Drift determination is performed by each processing as shown in FIG. The function of determining whether the vehicle is in the drift state or the non-drift state is referred to as drift determination means (turning state determination means).

That is, in this control, it is determined that drift occurs when the tire slips or slips. Tire sideslip can be determined when the relationship between the calculated lateral G and the actual lateral G becomes non-linear, and the vertical slip of the tire becomes non-linear when the relationship between the estimated vehicle speed vb and the longitudinal G estimated vehicle speed vbs described later is non-linear. It can be determined when it has become. Usually, when the vehicle drifts, the vehicle is accompanied by skidding or vertical skidding.

Therefore, in this control, as shown in FIG. 16, a coefficient (tyre coefficient of tire slip) dgy corresponding to the side slip state of the tire and a coefficient (longitudinal coefficient of tire) dvvbs corresponding to the vertical slip state of the tire, as shown in FIG. Based on the drift determination coefficient srp, a drift determination coefficient srp as a degree of wheel slip is set (calculated) and output. The drift correction coefficient srp1 to srp is further determined based on the drift determination coefficient srp.
Set 5. The function of setting the drift determination coefficient srp as the degree of slip of the wheel is referred to as slip degree detecting means. Also, such a drift determination coefficient s
Control that reflects rp in the torque transfer control is also referred to as slip control.

2.2.1 Tire Side Slip Coefficient (dgy) In this control, as described above, the calculated lateral G, that is, the reference lateral acceleration gy is calculated from the steering wheel angle θh and the estimated vehicle speed vb. On the other hand, the actual lateral acceleration (actual lateral G) rgy is detected by the lateral G sensor. When the vehicle is traveling without skidding, the relationship between the calculated lateral G and the actual lateral G becomes linear. Therefore, in order to make a drift determination,
The calculated horizontal G and the actual horizontal G are compared.

However, the calculated lateral G (gy) is calculated from the input information such as the steering wheel angle θh, and a phase delay occurs before the lateral G occurs in the vehicle according to the steering wheel. In the control, the calculated horizontal G is filtered by a low-pass filter to perform phase adjustment (see (2.2.1.1)). gyf = LPF [gy] (2.2.1.1) Also, due to the influence of tires and differences in gear ratio, etc., an error exists between the calculated lateral G (gy) and the actual lateral G (rgy) even in the linear region. Therefore, the actual horizontal G (rgy) is corrected by the coefficient k as shown in the following equation (2.2.1.2) to perform coefficient adjustment.

Rgyh = k * rgy (2.2.1.2) Thus, it is possible to compare the calculated horizontal G (gyf) with the matched phase and the actual horizontal G (rgyh) with the matched coefficient. Then, the calculated horizontal G calculated by the following equation (2.2.1.3)
(Gyf) and the actual lateral G (rgyh) based on the dimensionless value (tire slip coefficient) dgy, which is non-linear between the calculated lateral G and the actual lateral G, ie, occurs in the lateral direction of the tire. Determine non-linearity.

FIG. 17 is a diagram showing an example of the correspondence between the actual lateral G (rgy) and the calculated lateral G.
Like the straight line A, the actual lateral G (rgy) and the calculated lateral G have a linear relationship. However, actually, when the tire exceeds the grip limit of the tire, a skid or the like occurs. Does not increase. On the high μ road, the linearity is maintained up to the region where the lateral G is high as shown by the curve B, but on the low μ road, the lateral G is
Can not maintain the linearity in the low region.

The tire side slip coefficient dgy is calculated by the following equation (2.2.
Define as 1.3). dgy = | (gyf-rgyh) / rgyh | (2.2.1.3) However, in the calculation of the side slip coefficient dgy of such a tire, a calculation start condition such as the following equation (2.2.1.4) and A clear condition such as the following equation (2.2.1.5) is provided. This is due to the size of the actual horizontal G (rgyh), the calculated horizontal G and the actual horizontal G
If the magnitude of the difference (gyf-rgyh) does not increase beyond a certain value, there is no danger that the vehicle will drift. In such a case, the calculation of the side slip coefficient dgy is not performed, and the calculation frequency is reduced. It is decreasing.

| Rgyh | <a [m / s ² ] and | gyf-rgyh | <b [m / s ² ] where a and b are constants, and dgy = 0. (2.2.1.4) In general, the actual lateral G does not increase like the computed lateral G beyond the linear region between the computed lateral G and the computed lateral G.
1.3) can be transformed as follows.

Gyf = (1 + dgy) rgyh (2.2.1.3.a) When the linear region is exited, dgy gradually increases from 0, and the relationship of the above equation (2.2.1.3.a) is For example, it can be shown as a straight line D in FIG. So, in theory,
If the sideslip coefficient dgy becomes a value other than 0, it can be determined that the linearity has been distorted. However, in practice, it is difficult to always perfectly match even if the phase adjustment and the coefficient adjustment are performed for the actual lateral G and the calculated lateral G, Even in the actual linear region, the slip coefficient dgy often occurs (becomes other than 0) in many cases. For this reason, in this control, as shown in FIG.
If the sideslip coefficient dgy is equal to or less than a first predetermined value (dgy1), the linear slip region dgy is set to a second predetermined value (dgy2).
In this case, if the sideslip coefficient dgy is between the first predetermined value and the second predetermined value, the degree of non-linearity increases as the value approaches the second predetermined value.

2.2.2 Tire longitudinal slip coefficient (dvvb)
s) In this control, as described above, the estimated vehicle speed vb is calculated based on the third highest wheel speed v3 of the four wheels, but if the tire slips significantly, the vehicle speed based on such wheel speed v3 is calculated. vb becomes larger than the actual vehicle speed. Therefore, when the occurrence of tire slip is estimated, based on the vehicle speed at this time and the front and rear G instead of the wheel speed, the front and rear G
The estimated vehicle speed vbs is calculated.

The estimated vehicle speed vbs before and after G is
Based on the front and rear G of the vehicle detected by the sensor, the vehicle speed vb _SL when the tire slips and the detected value of the front and rear G (gx) _SL are calculated by the following equation (2.2.2.1). Note that t
Is the elapsed time since the slip occurred, and the wheel speed (for example,
It can be estimated that a slip has occurred when the third fastest wheel speed v3) increases sharply.

Vbs == gx _SL · t + vb _SL (2.2.2.1) The longitudinal slip coefficient dvvbs of the tire is the estimated front and rear G vehicle body speed vbs calculated as described above, and the third detected simultaneously therewith. The following equation (2.2.2.
The calculated value dvvbsd is taken into account in consideration of the noise effect and the like, and this is further filtered with a low-pass filter [see (2.2.2.3)] to determine the tire longitudinal slip coefficient dvvbs.

Dvvbsd = v3-vbs (2.2.2.2) dvvbs = LPF [dvvbsd] (2.2.2.3) FIG. 19 shows the wheel speed (third fastest wheel speed v3) when the vehicle slips. FIG. 7 is a diagram showing an example in which the estimated vehicle body speed vbs changes before and after. Here, a case is shown in which the actual vehicle speed VR is substantially constant, the vehicle enters the extremely low μ road during traveling, a slip occurs on the tire, and the slip converges thereafter. As shown in the figure, when a slip occurs in the tire, the wheel speed v3 rapidly increases, and the estimated vehicle speed v
bs is calculated.

Immediately after the occurrence of the slip, the wheel speed v3
Increases, the difference between the wheel speed v3 and the estimated front and rear G vehicle body speed vbs, that is, the tire longitudinal slip coefficient dvvbs increases. When the slip converges, the wheel speed v
3 decreases and approaches the longitudinal G estimated vehicle speed vbs, so that the tire longitudinal slip coefficient dvvbs decreases. Therefore, based on the longitudinal slip coefficient dvvbs of the tire, it is possible to estimate a slip state of the tire, that is, a linear region where the tire is not slipping or a nonlinear region where the tire is slipping.

Therefore, in theory, the longitudinal slip coefficient dvvbs
If becomes non-zero, it can be determined that it has become nonlinear,
Actually, the slip occurrence estimation and the longitudinal G estimated vehicle speed v
Since an error also occurs in the estimation of bs, in this control, FIG.
As shown in the figure, the longitudinal slip coefficient dvvbs is equal to the first predetermined value (d
vvbs1) if less than, linear region, longitudinal slip coefficient dvvb
If s is equal to or larger than a second predetermined value (dvvbs2), the longitudinal slip coefficient dvvbs is set to the first predetermined value and the second predetermined value.
When the value is between the predetermined value and the second predetermined value, the degree of nonlinearity is assumed to increase.

2.2.3 Drift determination coefficient (srp) In the present apparatus, drift determination is performed in consideration of both the side slip coefficient dgy and the vertical slip coefficient dvvbs as described above.
Therefore, a value obtained by combining the sideslip coefficient dgy and the longitudinal slip coefficient dvvbs (this is referred to as a drift determination coefficient) srp (= srpd ² ) is calculated by the following equation (2.2.3.1).
Used for drift judgment.

Srp = (a · dgy) ² + (b · dvvbs) ² (2.2.3.1) where a and b are coefficient adjustments to make a circle. This drift determination coefficient srp is shown in FIG. It can be evaluated by such a drift determination circle (friction circle). FIG. 21 shows a value (a) obtained by coefficient-adjusting the sideslip coefficient dgy.
• dgy) and the value (b · dvvbs) obtained by adjusting the longitudinal slip coefficient dvvbs on the horizontal axis and the vertical axis, respectively, indicating orthogonal coordinates, and the drift determination coefficient srp corresponds to the square of the distance from the origin in these coordinates.

The drift determination circle is a circle centered on the origin of such coordinates, and includes a first radius r ₁ and a second radius r ₂ (r ₁ <r ₂ ). And the radius r ₁
The linear region (the region where the tire is not slipping) is inside the circle, the nonlinear region (the tire is slipping) outside the radius r ₁ , and the drift region is outside the radius r ₂ of the nonlinear region. Is set.

That is, the square root (srp ^1/2 ) of the drift determination coefficient srp is within the radius r ₁ (that is, srp ^1/2 ≦ r
₁ ) is a linear region, if srp ^1/2 is larger than radius r ₁ (ie, srp ^1/2 > r ₁ ), it is a non-linear region, and further, srp ^1/2 is larger than radius r ₂ ( That is, srp ^1/2
> R ₂ ), it is in the drift region. In the non-linear region, the region where r ₁ <srp ^1/2 ≦ r ₂ is not a perfect drift, but the drift determination coefficient sr
It is assumed that there is a drift tendency of a degree corresponding to p.

For example, FIG. 22 shows the drift determination coefficient srp
Intended to show the drift determination for, srp is the radius r ₁ ² or less (i.e., srp ≦ r ₁ ²⁾ the linear region if,
srp than the radius r ₂ ² large (i.e., srp> r ₂ ²⁾ drift region if, srp is ^{_{r 1 2 <srp ≦ r 2}} 2 regions is not a full drift, drift determination coefficient sr
It is assumed that the degree of drift corresponds to p.

(Drift Corresponding Control Start Condition) If the drift determination coefficient srp is equal to or more than a predetermined value, and if the countersteer is turned off and the steering wheel angular velocity of the countersteer is higher than a predetermined speed, it is determined that the vehicle is drifting (drift determination). Means or turning state determination means). It is determined that the countersteer is turned off when the steering angle exceeds the neutral position, that is, when the direction of the calculated lateral G is opposite to the direction of the actual lateral G. That is, when all of the following three conditions are satisfied at the same time, it is determined that the vehicle is traveling in drift, and the drift-based control (slip-based control) is started. Note that the function of determining the start of such drift correspondence control (slip correspondence control) is referred to as start determination means. • The drift determination coefficient srp is equal to or more than a predetermined value. Srp> srp0 (2.2.6.1) • The direction of the calculated lateral G (gy) and the direction of the actual lateral G (rgyh) are opposite gy · rgyh <0 ··· (2.2.6.2) also handle angular velocity Derutashitah is Δθh ≧ Δθ ₁ is at a predetermined speed [Delta] [theta] ₁ or more (deg / s) ··· (2.2.6.3 ), the three equations conditions are simultaneously Even when the condition is not satisfied, it may be determined that the vehicle is drifting when the drift determination coefficient srp is equal to or greater than a predetermined value. Note that Δθh is dθh
In both cases, Δθ ₁ is also referred to as dθ ₁ .

(Drift-Compatible Control Termination Condition) When the steering angle returns to the neutral position again, that is, when the calculated lateral G direction is equal to the actual lateral G direction, it is determined that drift traveling has ended. Stop the drift control. The end (end) of such drift correspondence control (slip correspondence control)
Is referred to as termination determination means. The direction of the calculated lateral G (gy) and the direction of the actual lateral G (rgyh) are the same direction. Gy · rgyh> 0 (2.2.6.4) Also, it is determined that the vehicle is not drifting by the above conditional expression. Then, the drift determination coefficient srp is set to zero (sr
p = 0).

2.2.4 Drift correction coefficient 1 (srp
1) The drift correction coefficient 1 (srp1) is set based on the drift determination coefficient srp using a map as shown in FIG. Then, at the time of drift determination, correction is performed by multiplying the target ΔN tracking control by the drift correction coefficient 1 (srp1). Thus, the yaw rate feedback control is stopped by correcting the reference left-right speed difference dvhf between the right and left rear wheels by the correction coefficient srp1, and an LSD (limited slip differential) function that responds to the rotational speed difference, that is, a differential limit is set. . Further, the steering transient response coefficient (kdh) for decreasing the gain according to the steering speed is corrected so as not to decrease the gain. Further, if it is determined that the vehicle is not drifting due to the start / end conditions of the drift response control, the drift determination coefficient srp becomes zero (srp = 0). Therefore, as shown in FIG. 23, the drift correction coefficient 1 (the feedback gain srp1) ) Is srp1 = srp1max
(That is, 1).

2.2.5 Drift correction coefficients 2 to 5 (sr
p2-5) Drift correction coefficients 2 and 3 (feedback gain sr
p2,3) sets the strength of the LSD effect by the target ΔN tracking control, and is set based on the drift determination coefficient srp using a map as shown in FIGS. Among them, the drift correction coefficient 2 (srp2) sets the gain for the high μ road and the drift correction coefficient 3 (srp2).
3) is for setting each of the low μ road gains.

That is, if it is determined that the vehicle is not drifting according to the start / end conditions of the drift response control, the drift determination coefficient srp becomes zero (srp = 0).
As shown in FIG. 25, the drift correction coefficients 2 and 3 (feedback gains srp2 and srp3) are equal to srp2
= Srp2max = 1, srp3 = 1, and both are set to 1, and the gain of the target ΔN control is not adjusted. Also,
At the time of drift, the high μ road gain, that is, the drift correction coefficient 2 (feedback gain srp2) is decreased, and the low μ road gain, that is, the drift correction coefficient 3 (feedback gain srp3) is increased.

FIG. 28 shows the rotational speed difference (dvhf-dvrd).
5 shows the amount of torque movement with respect to. As shown in FIG. 28, the drift correction coefficient 2 (srp2) is reduced by lowering the high μ road gain of the present apparatus, and the drift correction coefficient 3 (srp3) is increased by lowering the low μ road gain. Also attempts to approximate the characteristics of a mechanical LSD.

That is, the control amount for the high μ road is made to have a larger transmission torque amount than that of the mechanical LSD during normal control (other than at the time of drift), thereby covering the reaction delay of the feedback control. When the vehicle drifts, the hunting of the control is likely to occur because the left and right wheels are slipping. At this time, the hunting of the control can be performed by reducing the gain.

Further, with respect to the control amount for the low μ road, the slip of the wheels can be prevented by making the amount of transmission torque smaller than that of the mechanical LSD during normal control (other than drift). By increasing the gain at the time of drift, the difference between the rotational speeds of the left and right wheels can be eliminated promptly, whereby stable turning can be performed.

The drift correction coefficient 4 (srp4) is used to adjust the control gain of the steering angular velocity proportional control term to an appropriate control gain at the time of drift determination, and uses a map as shown in FIG. It is set based on the coefficient srp.

The drift correction coefficient 5 (srp5) is used to adjust the control gains of the acceleration turning control term and the tack-in control term to appropriate control gains at the time of drift determination. It is set based on the drift determination coefficient srp. 2.2.6 Turning lateral G (drift correspondence, ggy) By the way, in this control, there is a torque movement control based on a lateral acceleration (turning lateral G) applied to the vehicle at the time of turning. For example, tack-in response control and acceleration turning control are performed. This corresponds to this.
The turning side G corresponds to the calculated side G and the actual side G described above.
When there is no difference between the calculated lateral G and the actual lateral G when the tire is gripping the road surface (during grip traveling), the computed lateral G and the actual lateral G correspond to the behavior of the vehicle. Thus, the calculated lateral G having a higher processing speed than the actual lateral G can be used as the turning lateral G. However, when the vehicle is drifting, a large difference occurs between the calculated lateral G and the actual lateral G. Therefore, the calculated lateral G cannot be used. In this case, the actual lateral G corresponding to the behavior of the vehicle is used as the turning lateral G. Must be used.

Therefore, the present apparatus normally uses the calculated horizontal G, and uses the calculated horizontal G when the calculated horizontal G cannot cope with the actual situation. For this reason, when it is determined that the vehicle is drifting under the drift-dependent control start condition, the turning lateral G
Then, the mode is switched from the use of the calculated lateral G to the use of the actual lateral G, and if it is determined that the drift traveling has been completed in the drift corresponding control end condition, the setting is made to return from the adoption of the actual lateral G to the use of the calculated lateral G. ing.

Note that the selection of the horizontal G is performed by selecting the horizontal G selection coefficient dor.
i, and when calculation horizontal G is selected, dori = 0,
When the actual horizontal G is selected, dori = dori1 (constant). The turning side G: ggy corresponding to the drift can be expressed by the following equation by the side G selection coefficient dori. ggy = [dori · rgyh + (dori1-dor)
i) .gy] / dori1 where gy: calculated lateral G, rgyh: actual lateral G. Further, an example of selecting the turning lateral G according to such drift will be described with reference to FIG. In FIG. 53, the solid line indicates the calculated horizontal G (gy), the broken line indicates the real horizontal G (rgyh),
As shown in the figure, when the vehicle turns, a lateral G occurs in the vehicle, and there is no difference between the calculated lateral G and the actual lateral G during the grip traveling. G changes suddenly. The sudden change in the calculated lateral G is because the driver applies the steering wheel operation in the drift state. When the steering wheel operation is applied, the calculated lateral G is calculated based on the steering wheel angle θh as in the equation (2.1.6.1). That is, the calculated lateral G greatly changes. In particular, when the counter steer is turned off during drift, the calculated lateral G becomes the actual lateral G.
Changes in the opposite direction. The calculation is performed only until the calculated lateral G changes in the direction opposite to the actual lateral G and the calculated lateral G becomes the same direction as the actual lateral G, that is, only during the period indicated as “drift control” in FIG. Instead of the horizontal G, an actual horizontal G input is adopted.

2.3 Vehicle Motion Control Logic As described above, in the present power transmission control device, as control modes, target rotation speed difference follow-up control (target ΔN follow-up control),
Acceleration turning control, tack-in response control, and steering transient response control are provided. Here, each of these controls will be described in detail.

2.3.1 Target ΔN Tracking Control The target ΔN tracking control is a control aiming at both the action as the yaw rate feedback control (yaw rate FBC action) and the action as the LSD (LSD action). The gain adjustment is performed so as to eliminate the difference between the rear wheel reference rotational speed difference (theoretical value, dvhf) obtained as described above according to 2.1.7.2) and the speed difference (actual speed difference, dvrd) between the right and left rear wheels. Do. Therefore, the respective control amounts (the control amount tbh for the high μ road and the control amount tbl for the low μ road) are set as shown in the dashed block B31 in FIGS.

For example, at the time of acceleration / deceleration turning, the stability of the vehicle becomes nonlinear, and the rear wheel reference rotational speed difference (theoretical value, dv)
hf) and the actual speed difference between the right and left rear wheels (actual speed difference, dv)
rd), the difference ddvr (= dvhf−dvrd) is different from the theoretical value dvhf and the actual speed difference dvrd.
Control is performed so as to eliminate vr. That is, during acceleration turning, the steering characteristic of the vehicle increases toward the understeer side, and the difference ddvr takes a negative value. Therefore, control is performed so as to move the torque in a direction to increase the difference ddvr, that is, toward the turning outer wheel. In addition, during deceleration turning, the steering characteristic of the vehicle increases toward the oversteer side, and the difference ddvr takes a positive value. Therefore, control is performed so as to move the torque in a direction to decrease the difference ddvr, that is, toward the turning inner wheel.

(1) Rotation speed difference proportional control (tbhd, t
bld) First, a difference ddvr between a rear wheel reference rotational speed difference (theoretical value, dvhf) and an actual speed difference between the right and left rear wheels (actual speed difference, dvrd) is obtained by the following equation (2.3.1.1). This difference ddv
Regarding r, the basic control amount corresponding to the map shown in FIGS. 31 and 42, that is, the basic control amount tbh for the high μ road.
d, the basic control amount tbld for the low μ road is set.

Ddvr = dvhf−dvrd (2.3.1.1) However, during the drift control, the drift correction coefficient 1 is added to the reference rear wheel reference rotational speed difference (dvhf) as shown in the following equation. (Srp1), the yaw rate FBC
The operation is eliminated, and only the LSD operation that responds to the rotational speed difference is controlled. dvhf = dvhf · srp1 / srp1max (2.3.1.2) Then, the basic control amount tbd (tbhd, tbld)
29 and 30, smoothing processing,
Filter processing and steering transient response coefficients (kbh, kb
l), a vehicle speed coefficient (kbh2, kbl2), and a drift correction coefficient 2, 3 (srp2, srp3) are multiplied to make a final control amount tb (tbh, tb).
1).

(2) Smoothing processing and filter processing (tbhdf2, tbldf2) In order to suppress control hunting, smoothing processing and limiter type filter processing are performed. In the smoothing process, for example, the following control amount tb (n
-1), that is, tbh (n-1) or tbl (n-1)
And the current basic control amount tbd (n), that is, tbhd
(N) or tbld (n) is averaged.

Tbdf1 (n) = [tbd (n) + tb (n−1)] / 2 (2.3.1.3) Also, as the limiter type filter, one having a different gradient between when increasing and when decreasing is used. use. Then, the filtered value is set to tbdf2 (that is, tbhdf for the high μ road).
2, tbldf2) for low μ roads.

(3) Steering transient response coefficient (kbh1, kb
l1) When the vehicle is steered suddenly, in the target ΔN tracking control, a signal is output with a delay, which adversely affects the behavior of the vehicle.
Therefore, using the peak hold value of the steering wheel angular velocity,
After the fast steering, the control gain for the target ΔN tracking control is reduced for a while according to the steering wheel angular velocity.

The peak hold value dst of the steering wheel angular velocity
f is the absolute value of the detected steering angular velocity dθh (| d
θh |) is set as in the following equation (2.3.1.5) if the condition as in the following equation (2.3.1.4) is satisfied, and set as in the following equation (2.3.1.6) if this condition is not satisfied. Is done. Condition: dstf ≦ | dθh | +2 (2.3.1.4) dstf = | dθh | (2.3.1.5) dstf ≦ dstf-2 (2.3.1.6) Also, during the drift control, the value dstf of the peak hold is multiplied by a drift correction coefficient 1 (srp1) as follows in order to remove the influence of the steering wheel.

Dstf = dstf · srp1 / srp1max (2.3.1.7) The transition of the handle angular velocity dθh according to the peak hold coefficient relating to the peak hold value dstf is shown in FIG.
The steering wheel angular velocity dθh gradually returns to 0 as the peak hold coefficient is smaller.
As for the steering transient response coefficient (kbh, kbl), a corresponding value is set from a map as shown in FIGS. 32 and 43 based on the peak hold value dstf corrected by the drift correction coefficient 1 (srp1) in this manner. . Note that FIG.
2 is a map for a high μ road, and FIG. 43 is a map for a low μ road. Here, these maps are configured similarly.

(4) Vehicle speed coefficient (kbh2, kbl2) Also, as the vehicle speed increases, the influence of the target ΔN tracking control on the vehicle behavior increases. Therefore, when the vehicle speed increases, the stability of the vehicle behavior is emphasized. The gain of the target ΔN tracking control is reduced. Here, the gain is set to a vehicle speed coefficient (kbh2, kbl2). The vehicle speed coefficient (kbh2, kbl2) is set based on the estimated vehicle speed vb from a map as shown in FIGS. 33 and 44, for example. Then, the set vehicle speed coefficient (kbh2, kbl
2) is multiplied.

(5) Drift correction coefficients 2, 3 (srp
2, drift 3) The drift correction coefficients 2 and 3 (srp2, 3) set the strength of the LSD effect by the target ΔN tracking control as described above. Drift correction coefficient 2 (srp
2) is for setting the gain for the high μ road, and lowers the gain when drifting. Drift correction coefficient 3 (s
rp3) is for setting a low-μ road gain, and increases the gain when drifting.

2.3.2 Acceleration Turn Control Acceleration turn control is a control for suppressing an increase in the understeer tendency during a sharp turn as described above. This control is necessary because the stability of the vehicle is non-linear. This is the case. That is, as shown in FIG.
If the relationship between the turning G and the steering ratio deviates from the linear region (see the broken line portion), the turning radius of the vehicle increases as shown by the broken line in FIG. This is because during a sharp turn, the steering characteristics of the vehicle increase toward the understeer side.

As described above, during a sharp turn, the target ΔN
In the following control, the torque is moved to the turning outer wheel to generate a moment in the turning direction to increase the cornering force of the front wheels. However, the target ΔN following control has a slight reaction delay due to feedback control. Therefore,
At the time of such a sharp turn, acceleration turning control for moving the torque to the turning outer wheel side is performed to generate or increase the yaw moment in the turning direction, and to increase the cornering force of the front wheels in a region where the longitudinal acceleration is large. This is to prevent understeering.

(1) Acceleration turning control amount (teh, tel) In this control, as shown in a block B32 in FIGS. 29 and 30, when the turning lateral G (ggy) is equal to or more than a predetermined value, the acceleration turning control is performed. The basic control amounts tehd, teld are set,
In addition to this, correction according to acceleration or accelerator opening (that is,
Multiplication of acceleration coefficients keh1, keh2, kel1, kel2) and correction according to vehicle speed (ie, vehicle speed coefficient keh)
3, multiplication by kel3), and a correction by a drift correction coefficient [drift correction coefficient 5 (srp5)] is performed to obtain a final control amount at the time of acceleration turning, teh, tel. This control amount is output on condition that the tack-in correspondence control is not being performed. Note that the basic control amounts tehd and teld are set based on the turning side G (ggy) by determining that the turning side G (ggy) is a sharp turn when the turning side G (ggy) is equal to or more than a predetermined value, based on maps as shown in FIGS. Therefore, the steering response is improved.

In this embodiment, the turning side G (g)
gy) is determined to be a sharp turn when it is equal to or greater than a predetermined value,
The control amount may be calculated so as to increase the rotation propulsion force of the turning outer wheel when the turning lateral G (ggy) is detected even if it is minute. (2) acceleration coefficient (keh1, keh2, kel1, k
el2) At the time of rapid acceleration, the cornering force of the front wheels decreases and the vehicle understeers strongly. Therefore, the gains of the control amounts teh and tel related to the acceleration turning control are increased, and the cornering forces of the front wheels are increased. Therefore, the acceleration coefficient k
eh1, keh2, kel1, kel2 are set, and the acceleration coefficient keh is set in the basic control amounts tehd, teld.
The gain of the control amounts teh and tel is increased by multiplying 1, keh2, kel1 and kel2.
Acceleration coefficient (throttle opening coefficient) keh1, kel1
Is set based on the throttle opening (tps) using a map as shown in FIGS. 35 and 46, and the acceleration coefficient ke
h2 and kel2 are set based on G (gb) before and after the calculation using maps as shown in FIGS.

Therefore, the steering response is improved by the gain adjustment based on the longitudinal acceleration, and the response delay of the gain adjustment based on the longitudinal acceleration can be covered by the gain adjustment based on the throttle opening, and the steering response is further improved. Needless to say, the steering response is improved only by adjusting the gain by either the longitudinal acceleration or the throttle opening.

(3) Vehicle speed coefficient (keh3, kel3) During high-speed running, the stability of the behavior of the vehicle is likely to be reduced by the torque transfer control, so that the acceleration turning control amounts teh and tel are controlled so as to maintain the stability of the vehicle. Decrease the gain. Therefore, the vehicle speed coefficients keh3, k
el3 is set, and the control amount (basic control amount tehd, te
In ld, acceleration coefficients keh1, keh2, kel1, ke
l2) multiplied by this vehicle speed coefficient keh3, kel
By multiplying by 3, the gain of the control amounts teh and tel is reduced. The vehicle speed coefficients keh3 and kel3 are
The setting is made based on the estimated vehicle speed (vb) using maps as shown in FIGS. Therefore, the oversteer phenomenon can be suppressed by prohibiting or weakening the control at or above the predetermined vehicle speed.

(4) Switch related to tack-in control As described above, the acceleration turning control is stopped during the tack-in response control. This is to ensure that the tack-in response control is given priority over the acceleration turning control, and that more urgent control of tack-in prevention is performed. Therefore, as shown in FIGS. 29 and 30, a switch (switch function) that operates as follows is provided.

When tdhd = 0, SW: ON, when tdhd ≠ 0, SW: OFF When tdld = 0, SW: ON, when tdld ≠ 0, SW: OFF (tdgd, tdld: Basic control of tack-in correspondence control described later) The control amount for the acceleration turning control and the control amount for the tack-in correspondence control described later can be used alternatively as described above, but the degree of necessity of the tack-in correspondence depends on the situation of the vehicle. When it is high, the control data may be obtained by integrating both data so that the degree of reflection of the control amount of the school building becomes large.

(5) Drift correction coefficient 5 (srp5) During the drift control, the input of the turning lateral G is switched from the calculated lateral G to the actual lateral G input, so that the control amount (basic control amount teh)
acceleration coefficients keh1, keh2, kel for d and teld
1, kel2, the vehicle speed coefficient keh3, and kel3) are multiplied by the above-described drift correction coefficient 5 to adjust the gains of the control amounts teh and tel to appropriate values.

2.3.3 Tack-in Correspondence Control As described above, during deceleration turning, contrary to acceleration turning, as the cornering force of the front wheels increases, the steering characteristics of the vehicle tend to oversteer, and the vehicle tends to cause tack-in. Become. As described above, during deceleration turning, the target ΔN
In the follow-up control, the torque is moved to the turning inner wheel side to generate a yaw moment in the turning restraining direction, and thereby,
Although oversteering is suppressed, the target ΔN tracking control has a slight reaction delay due to feedback control.

Therefore, at the time of decelerating turning, tack-in correspondence control for generating or increasing the yaw moment in the turning suppressing direction is performed by moving the torque to the turning inner wheel side, and the cornering force of the front wheels is reduced to suppress oversteering. I do. As a result, the turning behavior of the vehicle is controlled from the state shown by the dashed line in FIG. 57 to the state shown by the solid line, thereby avoiding the tuck-in of the vehicle and the spin caused by the tuck-in.

(1) Tack-in correspondence control (tdh, td
l) Therefore, at the time of such a decelerating turn, the control of the tack-in suppression for moving the torque to the turning inner wheel side is performed. In this control, as shown in a block B32 in FIGS. Control amounts tdhd, tdl
The basic control amounts tdhd and tdld are adopted as they are if the operation according to the tack-in corresponding control condition is satisfied, that is, if the tack-in corresponding control condition is satisfied, and the basic control amount tdh if the tack-in corresponding control condition is not satisfied.
d and tdld are set to 0. Further, various corrections including a drift correction coefficient [drift correction coefficient 5 (srp5)] are performed, and a final tack-in corresponding control amount td is obtained.
h, tdl. This control amount is output in response to a set / clear command. Note that the basic control amounts tdhd and tdld are set based on the turning side G (ggy) using maps as shown in FIGS.

(2) Start condition and end condition (kdhd, kdld) of tack-in control The tack-in control starts when the start condition is satisfied and ends when the end condition is satisfied. Correction coefficients kdhd, kd according to conditions
By multiplying the basic control amounts tdhd and tdld by ld, a control amount corresponding to a condition is output.

(Start Condition) Tack-in control is necessary when deceleration is started during a turn, so that the driver starts deceleration command operation or actual start of deceleration behavior of the vehicle, and starts the tack-in control. It is a condition. As an example, a throttle position sensor (TPS) 48
The time change dtps of the throttle opening along with the throttle opening tp detected at E is set to each of the threshold values tps1, dtps1.
(Dtps1 <0), and when the following condition is satisfied,
Assuming that the start condition of the tack-in correspondence control is satisfied, the correction coefficients kdhd and kdld are set to 1.

Tps <tps1 and dtps <dtps1 (Termination condition) Tack-in correspondence control is not required when deceleration is completed at the time of turning or when the turning itself is ended. The absence of the lateral G is used as an end condition of the tack-in correspondence control. Whether or not the deceleration operation has been completed can be determined based on whether or not the throttle opening has become larger than a predetermined opening tps2 (this opening tps2 is larger than the start condition opening tps1).

Therefore, as shown in the following equation, whether the throttle opening tp detected by the throttle position sensor (TPS) 48E is larger than the threshold value tps2, or
If any of the turning laterals G (ggy) becomes 0, the end of the tack-in correspondence control is satisfied, and the correction coefficients kdh1 and kdl1 are set to 0. tps> tps2 or ggy = 0 (3) Switch (f) related to prohibition of tack-in correspondence control
td) A switch is provided for the case where the tack-in correspondence control is not desired to be effective, and a switch command is issued as follows according to the switch signal ftd. If the switch SW is on, the control amounts tdh and tdl are output. No output if off.

When ftd = 0, SW: OFF, when ftd = 1, SW: ON (4) Drift correction coefficient 5 (srp5) During the drift control, the input of the turning lateral G is input from the calculated lateral G to the actual lateral G input. , So that the control amounts tdh and tdl
The control amount td is obtained by multiplying the drift correction coefficient 5 described above.
The gains of h and tdl are adjusted to be appropriate.

2.3.4 Steering Transient Response Control Steering transient response control is a control performed at the time of steering transition.
As shown in a block B33 in FIGS. 29 and 30, control is performed so as to be proportional to the change in the steering angle, that is, the steering angular velocity. That is, the basic control amount tc according to the steering angular velocity dθh
hd, tcld, and a correction according to the vehicle speed,
Drift correction is performed by the correction according to the steering wheel turning or turning back, and the drift correction coefficient 4 (srp4), and the control is performed by the control amounts tch and tcl thus obtained.

(1) Steering angular velocity proportional control Basic control amounts tchd, tcl according to steering angular velocity dθh
Here, d is set. Here, the basic control amounts tchd and tc are set in accordance with the steering angular velocity dθh so as to be substantially proportional to the steering angular velocity dθh using maps as shown in FIGS.
Set ld. As a result, torque transfer control according to the steering angular velocity dθh can be performed, and the transient characteristics, that is, the initial response to steering wheel steering, are improved.

(2) Vehicle speed coefficient (kch, kcl) At low speeds, control responsiveness does not need to be particularly enhanced. Therefore, the degree of demand for steering transient response control is low. Therefore, the gain relating to the steering transient response control amount is reduced. , At high speeds, when turning the steering wheel or turning back, the steering angular velocity d
If the torque transfer control according to θh is performed, the behavior of the vehicle may be destabilized. Therefore, in order to avoid such instability, the vehicle speed coefficient kch, Set kcl.

Such vehicle speed coefficients kch and kcl are set, for example, by using maps as shown in FIGS. 40A and 51A at the time of turning, and at the time of turning back, by using FIGS. The setting is performed using a map as shown in FIG. For example, as shown in FIG. 58, when the steering wheel angle (steering angle) θh and the steering wheel angular velocity (steering angular velocity) dθh have the same sign (the same direction), it is determined that the steering wheel is to be cut and returned. When the steering wheel angle θh and the steering wheel angular velocity dθh have different signs (different directions), it is determined that the steering wheel is returned. That is, the product (θh · dθh) of the handle angle θh and the handle angular velocity dθh is positive [ie,
When θh · dθh ≧ 0], the cut is made negative (that is, θ
When h · dθh <0], it is determined that the vehicle is returned.

(3) Gain up coefficient (kch2, kc
l2) If the gain is increased with respect to the incision or return of the steering wheel, the responsiveness at the time of turning increases, and the steering wheel angular velocity d
The gains kch2 and kcl2 are set according to θh to correct the control amount. Gain up coefficient kch2, kc
l2 is, for example, at the time of cutting, as shown in FIG.
The setting is performed using a map as shown in FIG. 2 (A), and at the time of switching back, the setting is performed using a map as shown in FIGS. 41 (B) and 52 (B). At the time of turning back, a response delay of the yaw rate occurs, so that the gain up coefficient is set to be larger at the time of turning back than at the time of turning so as to eliminate the response delay. Further, since the yaw rate delay is larger on the low μ road, the gain-up coefficient is set to be larger on the low μ road than on the high μ road.

(4) Drift Correction Coefficient (srp4) During drift control, the control amount (basic control amounts tchd and tcld multiplied by vehicle speed coefficients kch and kcl) is multiplied by drift correction coefficient srp4 as described above. Thus, the gains of the control amounts tch and tcl are adjusted to be appropriate. 2.4 Estimation of road μ Here, estimation of road μ is performed by estimating μ during steady turning,
The estimation is performed at three stages: μ estimation at the time of starting and μ estimation at the time of nonlinearity. Each stage of the steady turning, the starting, and the non-linear turning exists in a region shown in FIG. 59 with respect to the turning lateral G and the vehicle speed. Note that the μ estimation at the time of starting sets an initial value relating to the road surface μ. The non-linear time is a case where the vehicle is non-linear with respect to the steering of the steering wheel. Here, in each of these cases, a road surface μ determination coefficient (road surface friction coefficient, that is, a coefficient representing the degree of road surface μ) γ is obtained, and an output gain value (output value) of each control amount is obtained from the road surface μ determination coefficient γ value. ). The road surface μ
, That is, the function of obtaining the road surface μ determination coefficient γ is also referred to as road surface friction coefficient detection means.

Output Value Setting As described above, the control amounts include the target ΔN follow-up control amounts tbh and tbl, the acceleration turning control amounts teh and tel,
Tack-in correspondence control tdh, tdl, steering transient response control amount tch, tcl, high μ road control amount and low μ
A road control amount is set, and each of the control amounts is steplessly gained between the high μ road and the low μ road according to the value of the road surface μ determination coefficient γ. The adjusted value is used as an output value (output gain).

For example, assuming that the control amount for high μ road (control gain for high μ road) is txh and the control amount for low μ road (control gain for low μ road) is txl, the output value (output gain) tx is
It is calculated from the road surface μ judgment coefficient γ by the following equation. The road surface μ determination coefficient γ is a value of 0 to γmax. Here, the case where the road surface μ determination coefficient γ is 0 is a low μ road, and the road surface μ determination coefficient γ
Is γmax, a high μ road, and between the low μ road and the high μ road,
That is, a case where the road surface μ determination coefficient γ has an intermediate value between 0 and γmax is referred to as a medium μ road.

Tx = {γ · txh + (γmax−γ) · txl} / γmax = {(txh−txl) · γ + γmax · txl} / γmax = {(txl−txh) · (γmax−γ) + γmax · txh} / Γmax (2.4.1.1) Here, the control gain (control amount) tx is set so as to shift to the high μ road side or to the low μ road side. Then, the output value is finely adjusted.

The control gain tx is set to the high μ side (the target Δ
N follow-up control: tb) The output value fine adjustment formula for the high μ side uses the corrected output value as tx
a, if the output value fine adjustment coefficient is a (a> 1), the following equation is obtained. txa = {a (txh−txl) · γ + γmax · txl} / γmax = {a · γ · txh + (γmax−a · γ) · txl} / γmax where 0 ≦ a · γ ≦ γmax (2.4.1.2) Note that txa is clipped at the upper limit by txh due to 0 ≦ a · γ ≦ γmax. The setting on the high μ side is performed with respect to the control gain tb of the target ΔN tracking control.

The control gain tx is set between the high μ side and the low μ side. [Steering angular velocity proportional control (transient response control): t
c, tuck-in control: td] In this case, the control gain tx is calculated using the above equation (2.4.1.1) without substantially performing the output value fine adjustment. Such a calculation is performed on the control gain (control amount) tc of the steering angular velocity proportional control (transient response control) and the control gain (control amount) td of the tack-in response control.

The control gain tx is set to the low μ side (acceleration turning control: te). The output value fine-adjustment formula for the low μ side calculates the corrected output value as tx
b, when the output value fine adjustment coefficient is b (b> 1), the following equation is obtained. txb = ｛b (txl−txh) · (γmax−γ) + γ
max · txh｝ / γmax = [b · (γmax−γ) · tx
l + {γmax−b · (γmax−γ)} · txh] / γma
x However, assuming that 0 ≦ b · γ ≦ γmax, txb is clipped at the lower limit by txl. The setting on the low μ side is performed with respect to the control gain te of the acceleration turning control.

The output values (output gains) tx, txa, txb obtained by appropriately performing such fine adjustment of the output value can be shown as shown in FIG. 70 with respect to the road surface μ. In FIG. 70, the one-dot chain line indicates the characteristic of the output value txa (ie, the target ΔN tracking control control amount tb) obtained by finely adjusting the control gain tx to the high μ side, and the solid line indicates the control gain tx. The characteristics of the output value tx when no adjustment is made (that is, the tack-in response control td, the steering transient response control amount tc) are shown.
Shows the characteristics of the output value txb (that is, the acceleration turning control amount te) obtained by finely adjusting the output value to the low μ side.

As shown in FIG. 70, the lower the road surface μ (the smaller the road surface μ determination coefficient γ), the smaller the control amount (output value) tx. This is because the lower the road surface μ, the higher the control effect. Therefore, in order to obtain the same control effect, it is necessary to reduce the control amount (output value) tx as the road surface μ decreases. Further, the target ΔN tracking control control amount tb
Is increased on the medium μ road because the target ΔN tracking control is a control that can maintain the stability of the behavior of the vehicle even if the road surface μ is relatively low. This is because it is desirable to actively stabilize the behavior of the vehicle. The reason why the acceleration turning control amount te is reduced on the medium μ road is that the acceleration turning control amount te has a property that it is difficult to secure the behavior stability of the vehicle when the road surface μ is low.

In this way, the target ΔN follow-up control amount tb and the acceleration turning control amount t are determined according to the road surface μ determination coefficient γ.
e, tack-in response control td, steering transient response control amount tc
Is calculated, as shown in FIG. 6, the target ΔN follow-up control control amount tb, the acceleration turning control amount te, and the tack-in corresponding control t
After the high-pass processing is performed to add the sum and the differential value of the sum, a steering transient response control amount tc is further added to obtain an output value ta. This is further filtered by a low-pass filter as described below to obtain a final output value tad.

Tad = LPF [ta] That is, the ECU 42 individually calculates each of the control amounts tb, te, td and tc based on various parameters, and integrates these to obtain an output value tad (or t). (Notation) is obtained.

The high-pass processing of the added value of the target ΔN follow-up control amount tb, the acceleration turning control amount te, and the tack-in correspondence control td is for solving the response delay on the hardware. This is to advance the phase of the control term. 2.5 Actuator drive In the drive processing (actuator drive processing or proportional valve / directional valve switching control processing), as shown in FIG. 60, the output value (torque moving amount) tad is received, and the output value tad is received.
To the actuator drive signal for performing the torque movement in the direction and the amount according to the output value tad, and outputs the proportional valve control signal to the proportional valve 106 according to the torque movement amount, and the direction according to the torque movement direction. Valve (directional switching valve) 10
7, a directional valve control signal is output to these proportional valves 10
6, The direction valve 107 is driven. At the same time, a display command signal is output to the indicator lamp 110 (reference numeral 1).
06, 107 and 110 are shown in FIG. 4).

2.5.1 Proportional Valve / Directional Valve Control (1) Proportional Valve Output From the final output value ta, a torque transfer-current map (FIG. 61)
Using the current correction map (see FIG. 62) and the current correction map (see FIG. 62). (2) Directional valve output The left and right driving force moving mechanism is such that when the left and right rotation speed difference (dvrf) becomes a speed equal to or higher than the speed increasing / decreasing gear ratio in low-speed large steering angle turning or running on a μ split road, torque is applied to the high speed side wheels. Can not move. Therefore, as shown in Table 1 later, and if divided in accordance with the differential differential state, the torque transfer by combining it select the appropriate clutch.

Here, the information necessary for the control includes the speed ratio Sm indicating the gear ratio, the rotational speed vr of the differential case,
Rear wheel left and right speed difference (digital filtered value) d
vhf, as shown in FIG. 63, the left wheel speed and the right wheel speed of the rear wheel are vrl and vrr, respectively, and the number of teeth of the gears 78A, 80A and 82A are Z1 and Z, respectively.
2, Z3, and the gears 78B, 80B, and 82B have Z4, Z5, and Z6, respectively, the following equations (2.5.2.1) to (2.5.
It can be calculated from the relationship in 2.3).

Sm = 1−Z6 · Z1 / (Z3 · Z4) = Z5 · Z1 / (Z2 · Z4) −1 (2.5.2.1) vr = (vrl + vrr) / 2 (2.5.2.2) ) dvhf = digital filter [vrl-vrr] (2.5.2.3)

[0182]

[Table 1]

Determination of Differential State of Left / Right Driving Force Moving Diff As shown in Table 1, control is performed by determining five differential states of the rear differential relating to left / right driving force movement. The determination of the differential state of the differential will now be described.

First, a description will be given with reference to the velocity diagram of FIG. In this velocity diagram, the abscissa indicates the gears 78B and 78A, 8
0B and 80A, 82B and 82A (Z6 /
Z3, Z5 / Z2, Z4 / Z1), and the vertical axis indicates the speed. Here, the left and right wheel speeds vrl, vrr of the rear wheels, the speed A of the speed increasing gear 80B, the speed B of the speed reducing gear 82B, and the differential case 58. Is shown.

FIG. 64 shows the state of normal running. During normal running, the difference dvrf between the left and right wheel velocities vrl and vrr of the rear wheels is shown.
Is small (| dvrf |), the left clutch 9
When 0L is brought into contact, torque is moved from the right wheel 30 to the left wheel 28, and when the right clutch 90R is brought into contact, torque is moved from the left wheel 28 to the right wheel 30. Recognize.

However, the magnitude of the difference dvrf (| dvr
When f |) becomes large and exceeds 2 Sm · vr, torque movement to one side cannot be performed. That is, | dvrf
It is necessary to change the content of the torque transfer control at the boundary of | = 2Sm · vr. Here, the rotational speed difference dvrf is
When increasing or decreasing around -2Sm · vr or 2Sm · vr, hunting occurs in the control, which is not preferable.
A fixed dead zone region (± δ, δ is a small coefficient) is provided in | = 2Sm · vr, and the cases are classified as follows.

Dvrf> 2 · Sm · vr + δ 2 · Sm · vr + δ ≧ dvrf ≧ 2 · Sm · vr−δ 2 · Sm · vr-δ>dvrf> −2 · Sm · vr +
δ-2 · Sm · vr + δ ≧ dvrf ≧ −2 · Sm · vr
−δ−2 · Sm · vr−δ ≧ dvrf These differential conditions can be shown as in FIG. In FIG. 65, the abscissa represents the speed difference dvrf between the left and right rear wheels, and the ordinate represents the rotational speed vr of the differential case.

FIG. 6 shows a case where the speed relationship between the clutch plates at the left clutch speed is reversed as compared with the normal case.
This is the case as shown in the velocity diagram of FIG. In this velocity diagram (the same applies to FIGS. 67 to 69 below), the left wheel speed is represented by L, and the right wheel speed is represented by R. In such a case, regardless of whether the left or right clutch is connected, the torque moves from the left wheel to the right wheel. Therefore, as shown in Table 1, when the final output value tad is negative (tad <0), If it is desired to move the torque to the right wheel, the left clutch is connected (in Table 1, L
Shown). This is because the left clutch has a smaller speed difference between the clutch plates, so that a smooth torque movement with less clutch connection shock can be realized. The final output value tad is 0 (tad = 0) or positive (tad>
0), both the left and right clutches are not connected (Table 1).
In the figure, it is indicated as N). In particular, when the final output value tad is positive (t
The reason why none of the clutches is connected in the case of ad> 0) is that torque movement to the left wheel is not realized. In this case, control is prohibited.

FIG. 67 shows the case where the speed relationship between the clutch plates at the left clutch speed is almost 0, as shown in the speed diagram of FIG. In such a case, since the torque can be moved from the left wheel to the right wheel by connecting the right clutch, as shown in Table 1, the final output value ta
If d is negative (tad <0), that is, if it is desired to move the torque to the right wheel, the right clutch is connected (shown as R in Table 1). In other cases, that is, when the final output value tad is 0 (tad = 0) or positive (tad> 0), neither the left or right clutch is connected (indicated as N in Table 1). In particular, the reason why none of the clutches is connected when the final output value tad is positive (tad> 0) is to prevent control hunting from occurring, and in this case, control is prohibited.

In the normal state, that is, when the left clutch is engaged, the torque moves to the left, and when the right clutch is engaged, the torque moves to the right, as shown in the speed diagram of FIG. In such a case, as shown in Table 1, when the final output value tad is positive (tad> 0), the torque is moved from the right wheel to the left wheel by connecting the left clutch, and the final output value When tad is 0 (tad = 0), the left and right clutches are not connected, and when the final output value tad is negative (tad <0), the right clutch is connected and the torque from the left wheel to the right wheel is increased. Move.

FIG. 68 shows the case where the speed relationship between the clutch plates at the right clutch speed is almost 0, as shown in the speed diagram of FIG. In such a case, the torque can be moved from the right wheel to the left wheel by connecting the left clutch, and as shown in Table 1, the final output value ta
When d is positive (tad <0), that is, when it is desired to move the torque to the left wheel, the left clutch is connected, and the other, that is, the final output value tad is 0 (tad = 0) or negative (td).
In the case of ad <0), both the left and right clutches are not connected. In particular, the reason why none of the clutches is connected when the final output value tad is positive (tad> 0) is to prevent control hunting from occurring, and in this case, control is prohibited.

FIG. 6 shows a case where the speed relationship between the clutch plates at the right clutch speed is reversed as compared with the normal case.
This is the case as shown in the velocity diagram of FIG. In such a case, regardless of whether the left or right clutch is engaged, the torque moves from the right wheel to the left wheel. Therefore, as shown in Table 1, when the final output value tad is positive (tad> 0), that is, If it is desired to move the torque to the left wheel, the right clutch is connected. This is because the right clutch has a smaller speed difference between the clutch plates, so that a smooth torque movement with less clutch connection shock can be realized. In other cases, that is, when the final output value tad is 0 (tad = 0) or negative (ta
In the case of d <0), both the left and right clutches are not connected.
In particular, when the final output value tad is positive (tad> 0), no clutch is connected because torque movement to the right wheel is not realized, and in this case, control is prohibited.

2.5.2 Hydraulic pump motor control The hydraulic pump motor is provided with a motor relay operating condition and a stopping condition. (Operating conditions) When the pressure switch 105 (see FIG. 4) is lower than a predetermined value, the motor relay is operated except in the following cases. -When the hydraulic pump system fails-Within a predetermined time after the ignition switch is turned on-When not controlling (that is, tad = 0) The condition that the pressure switch 105 is equal to or more than the predetermined value, and the condition other than when the hydraulic pump system fails As a result, it is possible to prevent the excessive hydraulic pressure from acting on the clutch, and from a condition that a predetermined time after the ignition switch is turned on, control is performed after a sufficient hydraulic pressure is generated in the hydraulic pump immediately after the start of the engine. Therefore, reliable control can be realized. In addition, when the motor relay is not operated when the control is not being performed, unnecessary operation of the hydraulic pump motor can be avoided and efficient operation can be performed.

(Stop Condition) When the pressure switch 105 is equal to or more than a predetermined value and when the pressure switch 105 is less than the predetermined value, the motor relay is stopped. When a predetermined time or more has elapsed since the operation of the motor relay, and when a failure of the hydraulic pump system has been determined.When the pressure switch 105 stops the motor relay at a predetermined value or more, excessive hydraulic pressure acts on the clutch. It is possible to avoid it and stop when a predetermined time or more has elapsed since the operation of the motor relay, so that the load on the hydraulic pump can be reduced. It is possible to avoid that a high hydraulic pressure acts on the clutch and control confusion.

[0195] 3. As described above, according to the target ΔN follow-up control, the torque transfer control that follows the reference wheel speed difference can be appropriately realized, and the steady turning characteristic of the vehicle, that is, the steering characteristic is preferred. Can be set to the state. According to the acceleration turning control, the torque is moved to the turning outer wheel side to generate a yaw moment in the turning direction, thereby increasing the cornering force of the front wheel in a region where the longitudinal acceleration is large, thereby suppressing understeering. Thus, as shown in FIG. 7, the state is improved from the state without control to the state with control.

Further, according to the tack-in suppression control, contrary to the acceleration turning, by moving the torque toward the turning inner wheel, a yaw moment is generated in the turning suppression direction to reduce the cornering force of the front wheels. By suppressing over-steering, the tack-in of the vehicle is suppressed as shown in FIG. According to the steering transient control,
The steering operation by the driver, that is, the torque transfer control according to the steering angular velocity can be appropriately realized, and the turning performance of the vehicle can be improved.

Further, by the torque transfer control according to the road μ, in the μ split state, the torque is moved from the low μ wheels to the high μ wheels. As a result, as shown in FIG. 11, the driving force transmitted from the wheels on the high μ road side to the road surface increases, and the vehicle can be started and accelerated more quickly and efficiently. Also, the target Δ
The final control amount tad is determined by adding the control amounts of the N following control, the acceleration turning control, the tack-in suppression control, and the steering transient control, and the control is performed. Therefore, the target ΔN following control, the acceleration turning control, and the tack-in Torque transfer control in which both the suppression control and the steering transient control are reflected is realized.

By the way, focusing on the drift control, the control amount is set as shown in the flowchart of FIG. That is, as shown in FIG. 71, first, a drift degree (drift determination coefficient) srp is calculated (step D10), and it is determined whether or not the drift control is being performed (step D20). From the No route, the process proceeds to steps D30, D40, and D50, and the control condition for starting the drift response is determined.

That is, in step D30, it is determined whether or not the degree of drift (drift determination coefficient) srp is larger than a predetermined value srp0 (ie, whether or not srp> srp0). If the drift degree srp is not larger than the predetermined value srp0, the condition for starting the drift-related control is not satisfied, and the normal control step (step D7) starts from the No route.
Go to 0).

On the other hand, the drift degree srp is equal to the predetermined value srp.
If it is larger than 0, the process proceeds from the Yes route to step D40.
To the direction of the calculated lateral G (gyf) and the actual lateral G (rgy).
h) is opposite to the direction (ie, gyf · rgy)
h <0). If the direction of the calculated lateral G (gyf) and the direction of the actual lateral G (rgyh) are not opposite, the condition for starting the drift-responsive control is not satisfied, and the process proceeds from the No route to the normal control step (step D70).

On the other hand, the direction of the calculated horizontal G (gyf) and the actual horizontal G
If the reverse is the direction (Rgyh), the process proceeds from Yes route in step D50, determines whether the steering wheel angular velocity Dishitah is the predetermined speed d [theta] ₁ or more (i.e., whether dθh ≧ dθ _1). The steering wheel angular velocity dθh is equal to the predetermined velocity dθ
If it is not ₁ or more, the condition for starting the drift-responsive control is not satisfied, and the process proceeds from the No route to the normal control step (Step D70).

On the other hand, when the steering wheel angular velocity dθh is equal to the predetermined velocity d
If theta ₁ or more, will be a drift corresponding control start conditions are satisfied from Yes route, the process proceeds to step D60. If it is determined in step D20 that the drift control is being performed, the process proceeds to step D60 from the Yes route.
Proceed to. In step D60, a condition for terminating the drift control is determined. That is, it is determined whether the direction of the calculated lateral G (gyf) and the direction of the actual lateral G (rgyh) are the same (that is, whether gyf · rgyh> 0).

At step D20, it is determined that the drift control is not being performed, so that steps D30, D40, D
If the process proceeds to step D60 via 50, naturally, No
Drift control step from route (Step D130)
Proceed to. When it is determined in step D20 that the drift control is being performed and the process proceeds to step D60, the calculation horizontal G
If the direction of (gyf) and the direction of the actual lateral G (rgyh) are the same, the process proceeds to the normal control step (step D70), the drift control is stopped, and the calculated lateral G (g)
If the direction of yf) is not the same as the direction of the actual lateral G (rgyh), the process proceeds to step D130 to continue the drift control.

When the routine proceeds to the normal control in step D70, first, in step D80, the transient control amount tc is set according to the steering angular velocity dθh [tc = f (dθh)].
Then, in step D90, the rear wheel reference rotation speed difference dvh
f can be calculated from the tread width Lt of the vehicle, the vehicle speed vb, and the turning radius RR of the vehicle center by the following equation.

Dvhf = f (Lt, vb, RR) × srp1 = f (Lt, vb, RR) (∵srp1) Then, in step D100, the control gain of the target ΔN tracking control, that is, the drift correction coefficients 2 and 3 (Srp
2, srp3) is calculated. At this time, as described above, the control gains srp2 and srp3 are both set to 1 (srp2 = 1, srp3 = 1).

When the process proceeds to the drift control in step D130, first, in step D140, the transient control amount tc
Is set to 0 [tc = f (0)]. This is because the drift determination coefficient srp5 becomes 0 as shown in FIG. 27, and the transient control amount tc becomes 0 due to this coefficient srp5 as shown in the block B33 in FIGS. 29 and 30. Then, in step D150, the reference rotational speed difference dv
hf is corrected by the drift correction coefficient 1 (srp1) [dvhf = f (Lt, vb, RR) × srp
1].

Further, as shown in FIG. 23, when the drift determination coefficient srp is equal to or larger than a predetermined value srp0, the drift determination coefficient 1 (srp1) becomes 0, as shown in the block B31 in FIGS. 29 and 30. The reference rotational speed difference dvhf is set to 0 by the coefficient srp1. Then, in step D170, the control gain of the target ΔN tracking control, that is, the drift determination coefficients 2, 3 (srp2, srp3)
Is calculated. That is, first, as shown in FIG.
Control gain s for high μ road according to drift determination coefficient srp
rp2 is calculated. That is, as shown below, the control gain during drift is the control gain (sr
p2 = 1). Further, as shown in FIG. 25, the low μ road control gain srp3 is calculated according to the drift determination coefficient srp. That is, as described below, the control gain during drift is set to be larger than the control gain during non-drift (srp2 = 1). -When not drifting Srp2 = 1, srp3 = 1-When drifting 0 <srp2 <1 <srp3 Thus, the transient control amount tc and the reference rotation speed difference dvh
f, the control gains (drift determination coefficients) srp2 and srp3 of the target ΔN tracking control are set, and then step D1 is performed.
90, the reference rotational speed difference dvhf and the actual rotational speed difference d
Difference ΔN from vrd (= dvhf−dvrd = ddvr)
Is calculated. Then, the process proceeds to step D200, where the difference ΔN
(= Ddvr) control gain (drift determination coefficient) sr
By multiplying p2 and srp3, a control amount tbh for high μ road and a control amount tbl for low μ road of the ΔN following control are calculated in a stepless manner according to the value of the road surface μ determination coefficient γ as described above. The gain is adjusted and output as a ΔN tracking control amount tb (see the following equation).

Tbh = f (ΔN) × srp2 tbl = f (ΔN) × srp3 tb = f (tbh, tbl) Finally, the process proceeds to step D210, where the ΔN following control amount tb and the transient control amount tc are calculated as described above. Other control amount tot
Her (that is, tack-in response control td, acceleration turning control amount te) is added to calculate the total control amount t.

Therefore, at the time of drift determination, the target ΔN tracking control,
Since the steering angular velocity proportional control (steering transient response control), the acceleration turning control, and the tack-in correspondence control are performed, the following advantages are provided. That is, the drift correction coefficient 1 (srp
According to 1), if the degree of drift traveling is high, the effect of the LSD by the target ΔN tracking control is enhanced, and the recovery of the grip force of the wheels is accelerated, so that stable turning traveling can be performed.

Further, with the drift correction coefficient 2 (srp2), the hunting of the control can be prevented by lowering the high μ road gain of the present apparatus, and the drift correction coefficient 3 (srp3) for the low μ road By increasing the gain, the difference between the rotational speeds of the left and right wheels can be quickly eliminated, and the characteristics can be approached to those of the mechanical LSD, and stable turning can be performed.

Further, a drift correction coefficient 4 (srp4)
According to the embodiment, at the time of drift determination, the control gain of the steering angular velocity proportional control term (steering transient response control term) can be adjusted to an appropriate control gain, and stable turning can be performed even during steering transition. During the drift control, the input of the turning lateral G for the acceleration turning control and the tack-in correspondence control is switched from the calculated lateral G to the actual lateral G input,
By correcting the gain of the control amount with the drift correction coefficient 5 (srp5), the gain can be adjusted to an appropriate value.

The drift start determination condition, that is, the start condition of the drift response control (slip response control) is that the drift determination coefficient srp is equal to or larger than a predetermined value, and that the calculated lateral G (g
It the direction of direction and actual lateral G (rgyh) of y) is reversed, that wheel angular velocity Δθh is the predetermined speed [Delta] [theta] ₁ or more, is the fact that three conditions are satisfied at the same time, drift end determination condition, i.e., The drift corresponding control end condition is that when the direction of the calculated lateral G is equal to the direction of the actual lateral G, the start determination and the end determination of the slip corresponding control are performed based on different parameters. Once the slip response control is started, the slip response control is continued until the end determination condition is satisfied, so that the control is stabilized and the turning performance of the vehicle is further stabilized.

In the present embodiment, during normal control (other than during drift control), the control amount is set to be small when the road surface friction coefficient is low, and to be large when the road surface friction coefficient is high. In addition, at the time of drift control, a correction (gain adjustment) is performed so that the control amount becomes an intermediate value so that a change in the control amount corresponding to the road surface friction coefficient is reduced, so that the road surface friction coefficient can be adjusted. However, the present invention is not limited to this, and during drift control, the control amount correction according to the road surface friction coefficient is prohibited so that the control amount does not change due to the road surface friction coefficient. Alternatively, an intermediate control amount may be given.

For example, a basic control amount when the road surface friction coefficient is high is set, and during normal control other than the drift control, the basic control amount is corrected by a correction coefficient α such that the basic control amount becomes small when the road surface friction coefficient is low. , During drift control, a value unrelated to the road surface friction coefficient (correction coefficient β)
The basic control amount is corrected so as to be an intermediate value. In this case, the correction coefficient α continuously changes from 1 to a minimum value αmax (αmax <1) according to the road surface friction coefficient, and the correction coefficient β is a fixed value smaller than 1 and larger than the minimum value αmax. .

In the above embodiment, the target .DELTA.N follow-up control amount tb and the acceleration turning control amount te are added to ensure the turning performance of the rapid acceleration turning. Immediately after the start, the control may be temporarily performed based on the acceleration turning control amount, and thereafter, the control may be switched to the target ΔN tracking control amount for the steady control. In short, it is important to control so that the turning force of the turning outer wheel is increased immediately after the start of the sharp turning.

Further, in the present embodiment, the description has been given for a four-wheel drive vehicle.
It can be provided between the left and right driven wheels and left and right driven wheels of two-wheel drive vehicles such as front wheel drive vehicles and rear wheel drive vehicles.
It is conceivable that the present invention is applied between front wheels of a wheel drive vehicle.

[0219]

As described above in detail, according to the power transmission control device for left and right wheels according to the first aspect of the present invention, according to the turning state of the vehicle which can be estimated based on the degree of wheel slip. Since the rotational force adjusting mechanism is controlled for slippage, stable turning performance can be obtained. According to the power transmission control device for left and right wheels for a vehicle according to the second aspect of the present invention, an appropriate degree of slip is calculated because the degree of slip of the wheel is calculated from the longitudinal slip and the side slip of the wheel. As a result, the slip control can be appropriately performed.

According to the power transmission control device for left and right wheels according to the third aspect of the present invention, since the start determination means determines the start of the slip response control, the necessity of the slip response control becomes clear. By starting the slip response control based on the start determination, more stable turning performance can be obtained. According to the power transmission control device for left and right wheels of the present invention described in claim 4, the start determination and the end determination of the slip response control are performed based on different parameters, and in particular, once the slip response control is started. Until the termination determination condition is satisfied, the slip control is continued, so that the control is stabilized and the turning performance of the vehicle is further stabilized.

According to the power transmission control apparatus for left and right wheels of the present invention, the vehicle drifts in the control for making the difference between the left and right wheel rotational speeds close to the target value (target left and right wheel rotational speed difference follow-up control). When the vehicle is in a turning state, the target value is set to zero, so that the left and right wheels can be directly connected, and stable drift turning can be realized.
Since feedback control is also used in this case, stable control can be performed.

According to the power transmission control apparatus for a vehicle according to the sixth aspect of the present invention, when the road surface friction coefficient is low, the feedback gain is set low, and the feedback gain is increased in a drift turning state. By doing so, the slip of the wheels during normal control (during non-drift running) can be suppressed, and the right and left wheels can be brought closer to the directly connected state during the drift control, so that more stable drift turning can be realized. .

According to the seventh aspect of the present invention, when the road surface friction coefficient is low, the gain is adjusted so as to reduce the output control amount of the rotary propulsion force distribution adjusting mechanism when the road surface friction coefficient is low. By prohibiting or suppressing the gain adjustment in the drift turning state, it is possible to suppress the slip of the wheels during normal control (during non-drift running) and to bring the left and right wheels closer to the directly connected state during the drift control, More stable drift turning traveling can be realized.

According to the power transmission control device for left and right wheels of the vehicle according to the eighth aspect of the present invention, the transient control by the rotary propulsion force distribution adjusting mechanism is performed based on an artificial operation. By prohibiting or weakening the control, stable drift turning traveling can be realized. According to the power transmission control device for left and right wheels for a vehicle according to the ninth aspect, the control is performed by a rotation propulsion force distribution adjusting mechanism corresponding to the lateral acceleration of the vehicle.
When the vehicle is in a non-drift turning state, control is performed using the reference lateral acceleration calculated from the steering angle and the vehicle body speed, and when the vehicle is in a drift turning state, control is performed using the actual lateral acceleration. Control can be performed, which contributes to improvement of the turning performance of the vehicle.

[Brief description of the drawings]

FIG. 1 is a schematic overall configuration diagram of a drive system of a vehicle including a vehicle left and right wheel power transmission control device as one embodiment of the present invention.

FIG. 2 is a schematic configuration diagram showing a rotation propulsion force distribution adjusting mechanism (torque moving mechanism) of the vehicle left and right wheel power transmission control device as one embodiment of the present invention.

FIG. 3 is a schematic layout diagram showing a shaft configuration of a rotational propulsion force distribution adjusting mechanism (torque moving mechanism) of the vehicle left and right wheel power transmission control device as one embodiment of the present invention.

FIG. 4 is a schematic diagram showing the configuration of a hydraulic unit and a control system of a rotational propulsion force distribution adjusting mechanism (torque moving mechanism) of a vehicle power transmission control device between left and right wheels as one embodiment of the present invention.

FIG. 5 is a schematic diagram for explaining the operation principle of a rotational propulsion force distribution adjusting mechanism (torque moving mechanism) of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 6 is a control block diagram of a power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 7 is a diagram for explaining the target control content of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 8 is a diagram for explaining the target control content of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 9 is a diagram for explaining the target control content of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 10 is a diagram for explaining target control contents of a power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 11 is a diagram for explaining the target control content of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 12 is a diagram for explaining target control contents of a power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 13 is a control block diagram relating to input calculation processing of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 14 is a diagram illustrating an input calculation process of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 15 is a diagram illustrating an input calculation process of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 16 is a control block diagram relating to a drift determination process of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 17 is a diagram illustrating a drift determination process of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 18 is a diagram illustrating a drift determination process of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 19 is a diagram illustrating a drift determination process of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 20 is a diagram illustrating a drift determination process of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 21 is a diagram illustrating a drift determination process of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 22 is a diagram illustrating a drift determination process of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 23 is a diagram illustrating a drift correction coefficient relating to a drift determination process of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 24 is a diagram showing a drift correction coefficient relating to a drift determination process of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 25 is a diagram showing a drift correction coefficient relating to a drift determination process of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 26 is a diagram illustrating a drift correction coefficient relating to a drift determination process of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 27 is a diagram showing a drift correction coefficient relating to a drift determination process of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 28 is a diagram illustrating a drift correction coefficient relating to a drift determination process of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 29 is a control block diagram relating to a control amount calculation process (process for a high μ road) of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 30 is a control block diagram relating to a control amount calculation process (process for a low μ road) of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 31 is a map (high μ) related to ΔN tracking control of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention;
FIG.

FIG. 32 is a map (high μ) related to ΔN tracking control of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.
FIG.

FIG. 33 is a map (high μ) related to ΔN tracking control of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.
FIG.

FIG. 34 is a map (high μ) related to acceleration turning control of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.
FIG.

FIG. 35 is a map (high μ) related to acceleration turning control of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.
FIG.

FIG. 36 is a map (high μ) related to acceleration turning control of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.
FIG.

FIG. 37 is a map (high μ) relating to acceleration turning control of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.
FIG.

FIG. 38 is a diagram showing a map (high μ road) relating to tack-in correspondence control of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 39 is a diagram showing a map (high μ road) relating to steering transient response control of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 40 is a diagram showing a map (high μ road) related to steering transient response control of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 41 is a diagram showing a map (high μ road) relating to steering transient response control of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 42 is a map (low μ) related to ΔN tracking control of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention;
FIG.

FIG. 43 is a map (low μ) related to ΔN tracking control of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.
FIG.

FIG. 44 is a map (low μ) related to ΔN tracking control of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.
FIG.

FIG. 45 is a map (low μ) relating to acceleration turning control of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.
FIG.

FIG. 46 is a map (low μ) related to acceleration turning control of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.
FIG.

FIG. 47 is a map (low μ) related to acceleration turning control of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.
FIG.

FIG. 48 is a map (low μ) related to acceleration turning control of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.
FIG.

FIG. 49 is a diagram showing a map (low μ road) related to tack-in correspondence control of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 50 is a diagram showing a map (low μ road) related to steering transient response control of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 51 is a diagram showing a map (low μ road) relating to steering transient response control of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 52 is a diagram showing a map (low μ road) relating to steering transient response control of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 53 is a diagram for explaining drift control of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 54 is a diagram illustrating ΔN tracking control of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 55 is a diagram illustrating acceleration turning control of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 56 is a diagram illustrating acceleration turning control of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 57 is a diagram illustrating tack-in response control of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 58 is a diagram illustrating steering transient response control of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 59 is a diagram illustrating road surface μ determination of the vehicle left-right wheel power transmission control device as one embodiment of the present invention.

FIG. 60 is a control block diagram relating to drive processing of a vehicle left-right wheel power transmission control device as one embodiment of the present invention.

FIG. 61 is a diagram illustrating a map for describing a drive process of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 62 is a diagram illustrating a map for explaining a driving process of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 63 is a diagram for explaining a differential determination of a driving process of the vehicle left and right wheel power transmission control device as one embodiment of the present invention.

FIG. 64 is a diagram for explaining differential discrimination in drive processing of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 65 is a diagram illustrating a differential determination of a driving process of the power transmission control device for the left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 66 is a diagram for explaining a differential determination of a driving process of the power transmission control device for a left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 67 is a diagram illustrating a differential determination of a driving process of the power transmission control device for the left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 68 is a diagram illustrating a differential determination of a driving process of the vehicle left-right wheel power transmission control device as one embodiment of the present invention.

FIG. 69 is a diagram illustrating a differential determination of a driving process of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 70 is a diagram illustrating fine adjustment of an output value of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

FIG. 71 is a flowchart illustrating drift control of the power transmission control device for left and right wheels for a vehicle as one embodiment of the present invention.

[Explanation of symbols]

Reference Signs List 2 engine 4 transmission 6 intermediate gear mechanism 8 differential gear mechanism [center differential (center differential)] 8A, 8B differential pinion 8C, 8D side gear 10 front wheel differential gear mechanism [front differential (front differential)] 12L, 12R axle 14, Reference Signs List 16 front wheel 18 bevel gear mechanism 20 propeller shaft 22 bevel gear mechanism 24 differential gear device for rear wheel [rear differential (rear differential)] 26L, 26R axle 28, 30 rear wheel 32 output shaft for front wheel 34 output shaft for rear wheel 36 differential Viscous coupling unit (VCU) as limiting means 42 Electronic control unit (ECU or controller) as controlling means (rotational propulsion force distribution controlling means) 48A Wheel speed sensor 48B Handle angle sensor 48C Front and rear Acceleration sensor (front-rear G sensor) 48D Lateral acceleration sensor (lateral G sensor) 48E Throttle position sensor (TPS) 50 Rotational propulsion force distribution control mechanism (rotational force adjusting means, torque transfer mechanism) 51 Differential carrier 51A Wall 52 Input shaft 54 Drive pinion gear 56 Crown gear 58 Differential case (Differential case) 60A, 60B Differential pinion 62, 64 Side gear 66 Left wheel side rotating shaft 68 Right wheel side rotating shaft 70 Transmission mechanism 70A Speed increasing mechanism 70B Reduction mechanism 72, 74, 76 Intermediate shaft 78A, 80A, 82A Gear (sun gear) 78B, 80B, 82B Gear (planetary pinion) 84 Counter shaft 86 Triple gear 90 Transmission capacity variable control torque transmission mechanism 90L Clutch (left clutch) 90R Clutch (right clutch) 90AL, 90AR, 90BL, 90BR Clutch plate 92 Clutch case 96 Roller bearing 38 Hydraulic unit 101 Accumulator 102 Control pressure output 103 Accumulator 104 Motor pump 105 Pressure switch 106 Electromagnetic proportional pressure control valve (proportional valve) 107 Electromagnetic direction Control valve (directional switching valve) 108 Battery 109 Motor relay 110 Indicator lamp

 ──────────────────────────────────────────────────続 き Continued on the front page (72) Inventor Hiroyuki Suzuki 5-33-8 Shiba, Minato-ku, Tokyo Inside Mitsubishi Motors Corporation (72) Inventor Kazuki Ishiguro 5-33-8 Shiba, Minato-ku, Tokyo Mitsubishi (72) Inventor Toshiyuki Manabe 5-33-8 Shiba, Minato-ku, Tokyo Mitsubishi Motors Corporation

Claims (9)

[Claims] At least one of the left and right wheels of a front wheel and a rear wheel is provided in a vehicle configured as a drive wheel to which a driving force is applied from a power unit, and is provided in accordance with a traveling state of the vehicle Controlling the transmission of the driving force from the power unit in relation to the left and right driving wheels or the front and rear driving wheels, or controlling the differential between the left and right driven wheels to which the driving force is not transmitted. Or a turning force adjusting means for controlling the turning force of the left and right wheels or the front and rear wheels or the front and rear left and right wheels by controlling the braking force of the braking device in relation to the front and rear and left and right wheels; And a slip corresponding control for controlling a turning force adjusting mechanism according to a turning state of the vehicle which can be estimated based on the detection information of the slip degree detecting means. Control means and A power transmission control device for left and right wheels for a vehicle, comprising: 2. The slippage detecting means according to claim 1, wherein the slippage detecting means detects the slippage by detecting the vertical slippage and the skidding of the wheel and integrating the detected vertical slippage and the skidding. Power transmission control device for left and right wheels for vehicles. 3. The control means includes start determination means for determining whether or not to start the slip response control based on the degree of slip detected by the degree of slip detection means and other parameters. 2. The power transmission control device for left and right wheels for a vehicle according to claim 1, wherein the slip response control is started when a start determination is made in (1). 4. The control means includes: start determination means for determining the start of the slip response control when at least the degree of slip detected by the degree of slip detection means is equal to or greater than a set value; and a parameter other than the degree of slip. End judgment means for judging the end of the slip response control, and even if the degree of slip falls below the set value during the slip response control, the end judgment means does not judge the control end. 2. The power transmission control device between left and right wheels for a vehicle according to claim 1, wherein the slip control is continued as long as possible. 5. A driving force from a power unit is transmitted to the left and right wheels, or a rotational force accompanying rotation of the left and right wheels is provided between the left and right wheels. A rotary propulsion force distribution adjusting mechanism for changing the rotational propulsion force of the left and right wheels, and a control for controlling the rotary propulsion force distribution adjustment mechanism in accordance with the turning state of the vehicle to adjust the rotational force of the left and right wheels. Means,
A power transmission control device for left and right wheels for a vehicle, the control means comprising: target value calculation means for calculating a target value of a left and right wheel rotation speed difference corresponding to a traveling state of the vehicle; And a turning state determining means for determining whether the vehicle is in a non-drift turning state. The target value is compared with an actual left and right wheel rotation speed difference, and the rotation is determined by feedback control so that the latter approaches the former. Controlling the thrust distribution adjusting mechanism, and setting the target value to zero when the turning state determining means determines that the vehicle is in a drift turning state, between the left and right wheels for the vehicle. Power transmission control device. 6. A road surface friction coefficient detecting unit for detecting a road surface friction coefficient, wherein the control unit sets a feedback gain to be low when the road surface friction coefficient detected by the road surface friction coefficient detecting unit is low. 6. The power transmission control according to claim 5, wherein the feedback gain is increased when the road surface friction coefficient is low and the turning state determining means determines that the vehicle is in the drift turning state. apparatus. 7. A driving force from a power unit is transmitted to the left and right wheels, or a rotational force accompanying rotation of the left and right wheels is provided between the left and right wheels. A rotary propulsion force distribution adjusting mechanism that changes the rotational propulsion force of the left and right wheels, and a control amount calculation unit that calculates an output control amount of the rotary propulsion distribution adjustment mechanism according to the traveling state of the vehicle; A control means for controlling the rotational propulsion force distribution adjusting mechanism based on the output control amount calculated by the control amount calculation means; and a power transmission control device for left and right wheels for a vehicle, wherein the vehicle is in a drift state. A turning state determining means for determining whether the vehicle is in a non-drift state or not, and a road surface friction coefficient detecting means for detecting a road surface friction coefficient. Size Then, the output control amount is corrected according to the road surface friction coefficient detected by the road surface friction coefficient detection unit, and the correction is prohibited or suppressed when the turning state determination unit determines that the vehicle is in the drift state. A power transmission control device for left and right wheels for vehicles. 8. A driving force from a power unit is transmitted to the left and right wheels, or a rotational force accompanying rotation of the left and right wheels is provided between the left and right wheels. Giving and receiving, a rotary propulsion force distribution adjusting mechanism that changes the rotary propulsion force of the left and right wheels, including at least a transient control signal based on an artificial operation,
A control means for transmitting a control signal to the rotational propulsion force distribution adjusting mechanism in accordance with a running state of the vehicle; and a power transmission control device for left and right wheels for a vehicle, wherein the slip detecting means detects a degree of slip of the wheel. A degree detecting means, wherein the control means inhibits the transmission of the transient control signal in a region where the slip degree detected by the slip degree detecting means is larger than a predetermined value or weakens the transient control so as to weaken the transient control. A power transmission control device for left and right wheels for a vehicle, wherein the control signal is corrected. 9. A driving force from a power unit is transmitted to the left and right wheels, or a rotational force accompanying rotation of the left and right wheels is provided between the left and right wheels. A rotary propulsion force distribution adjusting mechanism for changing the rotational propulsion force of the left and right wheels, and control means for transmitting a control signal to the rotary propulsion force distribution adjustment mechanism at least in accordance with the lateral acceleration of the vehicle. A vehicle-side power transmission control device for left and right wheels, comprising: a turning state determining unit that determines whether the vehicle is in a drift turning state or a non-drift turning state; When the vehicle is determined to be in the non-drift turning state, a control signal based on the reference lateral acceleration calculated from the steering angle of the vehicle and the vehicle speed is transmitted, and it is determined that the vehicle is in the drift turning state. And transmitting a control signal based on the actual lateral acceleration of the vehicle when the power transmission is performed.