JP3063628B2 - Vehicle speed calculation device - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a vehicle speed calculation apparatus suitable for use in detecting the vehicle speed of a four-wheel drive vehicle in which all four wheels are drive wheels, and more particularly, to driving such a vehicle. The present invention relates to a vehicle speed calculation device suitable for use as a vehicle speed as one of parameters for accurately controlling wheel driving force and the like.

[0002]

2. Description of the Related Art In an automobile, a vehicle speed (vehicle speed) is used as a parameter for various controls such as drive control and braking force of the vehicle. Therefore, it is required to accurately detect the vehicle speed. Such a vehicle speed can be determined relatively accurately based on the wheel speeds of the non-driving wheels. It is difficult to determine the vehicle speed.

[0003] That is, since the driving wheels are accompanied by at least some slip when transmitting the driving force to the road surface, if the vehicle speed is calculated from the wheel speeds of the driving wheels, the vehicle speed becomes inevitably higher. Would. In addition, since the degree of the slip of the drive wheels varies depending on the running state and the running environment, it is difficult to calculate the vehicle speed based on the drive wheels (actually, it corresponds to the estimation of the vehicle speed). .

Therefore, as disclosed in JP-A-4-50066 and JP-A-3-54059, the third highest wheel speed among the four wheel speeds of the four drive wheels of a four-wheel drive vehicle is set to the vehicle speed. Techniques for estimating speed have been proposed. Also, in these publications, when the driving wheel slips greatly,
Since the reliability of the vehicle speed estimated based on the driving wheels is greatly reduced, a technique for estimating the vehicle speed by another method has been proposed. In other words, when the drive wheels are slipping significantly, the vehicle speed is calculated based on the acceleration of the vehicle at the time of occurrence of the slip, and thereafter, the vehicle speed based on the wheel speed is returned to the pseudo vehicle speed based on the acceleration. Then, the vehicle speed is calculated based on the wheel speed.

[0005]

However, if the third wheel speed is simply estimated as the vehicle speed as in the above-mentioned prior art, the vehicle speed can be properly estimated when the vehicle is going straight, but when the vehicle turns, the vehicle speed can be appropriately estimated. The vehicle speed cannot be estimated properly. That is, when the vehicle turns, a speed difference occurs between the turning inner wheel and the turning outer wheel. Therefore, if the third wheel speed is simply estimated as the vehicle speed, an appropriate estimation of the vehicle speed cannot be performed.

Means for estimating the average value of a plurality of wheel speeds as the vehicle body speed can be generally considered. However, with such means, the accuracy of the average value of the wheel speed between the front wheels during turning is poor. The wheels may slip during driving or braking, and during a turn, a slip occurs during the turn. Therefore, the reliability of the wheel speed itself is low, and the vehicle speed cannot be properly estimated.

SUMMARY OF THE INVENTION The present invention has been made in view of the above-mentioned problems, and even if a wheel slips due to transmission of a driving force or a braking force or turning, the vehicle speed can be properly adjusted based on the wheel speed of the wheel. It is an object of the present invention to provide a vehicle speed calculation device capable of estimating the vehicle speed.

[0008]

SUMMARY OF THE INVENTION Therefore, a vehicle speed calculating apparatus according to the present invention provides a wheel for calculating wheel speeds of four wheels of a pair of steered wheels and a pair of non-steered wheels provided in a vehicle. A speed sensor, and straight ahead vehicle speed estimating means for estimating a third wheel speed, which is the third highest among the four wheel speeds detected by the wheel speed sensor, as a vehicle speed when the vehicle travels straight,
Steering wheel turning angle detecting means for detecting a turning angle of a steering wheel provided to the vehicle; the third wheel speed estimated by the straight-ahead vehicle speed estimating means; and the steering wheel turning angle detected by the steering wheel turning angle detecting means. And a vehicle body speed estimating means for calculating and estimating the vehicle body speed at the center in the vehicle width direction at the time of turning of the vehicle from the vehicle and a constant unique to the vehicle body of the vehicle.

Preferably, the constant inherent in the vehicle body is
These include vehicle wheelbase, tread width, stability factor, and steering wheel ratio. Also, for example, when turning, a constant specific to the vehicle body and a steering angle (steering wheel turning angle)
And the third wheel speed to calculate the turning radius of the turning inner wheel,
The turning vehicle speed estimating means calculates the vehicle speed based on the turning radius, the third wheel speed, and the tread width of the vehicle.
Based on the wheel speed, the steering angle, and the vehicle-specific constant,
The turning radius on the turning inner wheel side, and
Vehicle based on the turning radius and the third wheel speed at
It is preferable to calculate and estimate the vehicle speed at the center in the width direction .

Further, it is preferable that the third wheel speed is corrected by a longitudinal acceleration by a longitudinal acceleration sensor provided on the vehicle body and a wheel acceleration of the wheel having the third wheel speed.
That is, when the wheel slips or locks significantly, the reliability of the vehicle speed estimated based on the wheel is greatly reduced. Therefore, it is preferable to estimate the vehicle speed by another method. For example, when the acceleration of the wheel (the wheel acceleration of the wheel at the third wheel speed) is equal to or more than a predetermined value or equal to or less than a predetermined value, it is determined that the wheel is largely slipping or locked.
At this time, the vehicle acceleration (longitudinal acceleration) detected by the longitudinal acceleration sensor is assumed to be a constant value, a pseudo vehicle speed is calculated based on the constant value, and this is used as the vehicle speed. When the vehicle speed based on the wheel speed (the vehicle speed corresponding to the wheel speed) returns to the speed, the vehicle speed corresponding to the wheel speed is adopted as the vehicle speed. Therefore,
For example, the turning vehicle speed estimating means calculates the acceleration of the vehicle body in the front-rear direction.
A longitudinal acceleration sensor for detecting the degree and a vehicle having the third wheel speed
The slip of the wheel at the third wheel speed is determined based on the wheel acceleration of the wheel.
To determine whether the vehicle is slipping with respect to the road surface.
It is determined that the wheel at the third wheel speed is in a slip state.
And the front detected by the longitudinal acceleration sensor
At the time of slip when the above-mentioned vehicle body speed is estimated by backward acceleration
It is preferable to have a vehicle speed estimation means. Further, it is preferable that the vehicle body speed estimating means estimates the vehicle body speed as a driving force control parameter in a driving force control device for a vehicle.

[0011]

〔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 2.5 2.4 Output value setting Operation of the present apparatus and effects of the present apparatus [Embodiment] 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 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 such a speed difference, the torque transmitting coupling is made to act 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 a 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).
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 driven 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, a hardware configuration of the power transmission control apparatus for left and right wheels for a vehicle based on the above-described 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 well-known conventional 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 are provided.
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 this embodiment is provided between a differential case 58 of the rear differential 24 for distributing driving force between left and right driving wheels of the rear wheels and a 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 is provided with a speed increasing mechanism 70A for increasing the rotational speed of the input portion of the rear differential 24, that is, the differential case 58, and outputting the rotational speed to one 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 so as 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, and 82B rotate integrally at a constant speed.
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 these 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 which rotates integrally with the right wheel side rotating 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 is switched to the right wheel side clutch plate 90R.
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.

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 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).

The left and right clutches 90L, 90R are set so as not to be 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 reduced as compared with the case where the distribution of the left and right 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 either 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 with the ECU 4
2, the ECU 42 includes a wheel speed sensor 48A, a steering wheel angle sensor (that is, steering wheel angle detecting means for detecting the steering angle of the steering wheel) 48B, a longitudinal acceleration sensor (longitudinal G sensor) 48C, Sensors such as an acceleration sensor (lateral G sensor) 48D, a throttle position sensor (TPS) 48E, and a 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 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, 1
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 through the hydraulic unit 38 needs to be linked with the engine output control, here, 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 Control Overview of the Present Apparatus Here, a control overview 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, 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 (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 processing, each control amount of the motion control of the vehicle is calculated based on the respective calculation results.
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 processing is called control amount calculation means. Among these functions (control amount calculation means), a function of setting a 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 A function of outputting a control signal based on the tuck-in control means, a function of setting a control amount (transient control amount) of steering transient response control related to a steering transient, or a function of this setting function and a control amount obtained by this setting The function of outputting a control signal based on the above is referred to as steering transient response control means. In particular, a means for calculating a control amount (vehicle behavior corresponding control amount) corresponding to the vehicle behavior, such as a target rotation speed difference follow-up control means, an acceleration turning control means, a tack-in correspondence control means, and a vehicle behavior corresponding control amount calculating means, a steering wheel operation A means for calculating a control amount (transient control amount) based on a driving operation state such as a throttle operation or a throttle operation is also referred to as a transient control amount means.

The target .DELTA.N follow-up control has a function of 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 (target value calculating means). 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.

Hereinafter, each of these processes will be described in detail, 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 the 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 and Deceleration At the time of decelerating turning, contrary to the acceleration turning, the steering characteristics of the vehicle are strengthened to the oversteer side, so that the vehicle is likely 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 to the high μ road wheel. 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 to determine the actual rear wheel left / right speed generated during turning or torque transfer control. 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 odvrd: 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) This device has a function of calculating the vehicle speed by calculating (vehicle speed calculating device). ,
The estimated vehicle speed vb is basically calculated on the basis of the third highest wheel speed v3 of 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, since it is conceivable that the detected value of the wheel speed may not be an appropriate value due to noise or the like, in consideration of the reliability of the detected value,
The estimated vehicle speed vb is determined by using the second slowest wheel speed (ie, the third fastest wheel speed) v3 among the four drive wheels.

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) vbd obtained by multiplying the wheel rotation speed by the wheel outer peripheral length is calculated as the vehicle speed. vb. Therefore, the function of calculating the estimated vehicle body speed vb from the third wheel speed (that is, the third fastest wheel speed) v3 is herein referred to as straight-ahead vehicle body speed estimation means.

However, at the time of turning, the wheel speed changes between the turning inner wheel and the turning outer wheel, and such a change in the wheel speed between the inner wheel and the outer wheel differs depending on the turning radius and the vehicle speed. For this reason, at the time of turning, correction according to the turning radius and the like is required. That is, at the time of turning, the third highest wheel speed is the inner wheel of the rear wheel, and the inner wheel side is considered to be the simply calculated vehicle speed vbd.
It is obtained from the geometric relationship as shown in FIG.

Therefore, the third wheel speed vbd (= v3) as the vehicle speed (simple calculated vehicle speed) estimated (calculated) by the above-mentioned straight-ahead vehicle speed estimating means, and a steering wheel angle sensor (a steering wheel turning angle detecting means) From the steering wheel angle (steering wheel turning angle) θh detected at 48B and a constant unique to the vehicle body, that is, the vehicle wheelbase, tread width, stability factor, steering wheel gear ratio, etc., the vehicle width during turning of the vehicle The vehicle speed at the center of the direction is calculated and estimated. The function of calculating and estimating the vehicle speed at the center in the vehicle width direction at the time of turning is referred to as turning vehicle speed estimating means.

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) [ Here, θh: steering wheel angle (steering wheel turning angle)] A: stability factor Lw: wheel base Lt: tread width GR: steering wheel gear ratio The ratio between the vehicle body speed vbd and the vehicle body speed vb is determined by the turning radius RRi on the inner wheel side. And the turning radius RR at the center of the vehicle body, and the turning radius RR is calculated using the turning radius RRi 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) Further, a wheel related to the third wheel speed v3 (= vbd) Is greatly slipped, the third wheel speed v3 deviates significantly from the actual vehicle speed. In such a case,
Since the wheel acceleration of the wheel at the third wheel speed v3 greatly differs from the actual longitudinal acceleration (longitudinal G) of the vehicle, the comparison between the actual wheel acceleration and the actual wheel acceleration results in a large wheel acceleration. Slips and the third wheel speed v3 (= vb
It can be determined that d) cannot be adopted as the vehicle speed.

Therefore, in the present vehicle speed calculating apparatus, the longitudinal G by the longitudinal acceleration sensor 48C provided on the vehicle and the third G
Based on the wheel acceleration d (v3) / dt of the wheel at the wheel speed, it is determined whether or not the wheel is largely slipping.
When determining such a slip, the longitudinal acceleration sensor 48
To estimate the vehicle speed by using a G before and after by C, which instead of the vehicle speed based on the wheel speed, the slip when cars are (this function so as to employ a vehicle speed based on the longitudinal G
The body speed estimation means) .

For example, FIG. 19 is a diagram showing an example in which the wheel speed (third fastest wheel speed v3) and the estimated front and rear G vehicle speed vbs change when the vehicle slips. This shows a case in which the vehicle enters a very low μ road during traveling at a substantially constant level, slips on the tires, and then the slips converge. As shown in the drawing, when a slip occurs in the tire, the wheel speed v3 rapidly increases, and the difference between the wheel acceleration d (v3) / dt of the wheel at the third wheel speed and the longitudinal G by the longitudinal acceleration sensor 48C is larger than a predetermined amount. This is considered to be a non-linear region where the tire is slipping.
At this time, the front and rear G estimated vehicle speed vbs is calculated,
Estimated vehicle speed based on wheel speed v3 (wheel speed corresponding to wheel speed)
vbd (= v3) instead of the front and rear G estimated vehicle speed vb
s is adopted.

The vehicle speed will be described later (item,
2.2.2) is calculated as in the following equation. vbs == gx _SL · t + vb _SL where vb _SL is the vehicle speed vb _SL when the tire slips. gx _SL is the front and back G detected when the tire slips. Due to the increase, the difference between the wheel speed v3 and the estimated front and rear G vehicle speed vbs, that is, the longitudinal slip coefficient dvvbs of the tire described later increases, but as the slip converges, the wheel speed v3 decreases and the front and rear G estimated. As it approaches the vehicle speed vbs,
The longitudinal slip coefficient dvvbs of the tire decreases.

Therefore, the tire longitudinal slip coefficient dvvb
Based on s, it is possible to estimate the slip state of the tire, that is, whether the tire is in a linear region where the tire is not slipping or in a nonlinear region where the tire is slipping. Here, when the longitudinal slip coefficient dvvbs converges below a certain value, the tire returns to the linear region where the tire does not slip.
From the adoption of the estimated vehicle speed vbs, the vehicle speed is returned to the estimated vehicle speed (vehicle speed corresponding to wheel speed) vbd (= v3) based on the wheel speed v3.

2.1.5 Back-and-forth acceleration (gx) First, as shown in the following equation (2.1.5.1), the front-rear acceleration is calculated from the change of the simply calculated vehicle speed vbd calculated at a predetermined cycle. Since the acceleration gxd fluctuates greatly, it is processed by a low-pass filter [see (2.1.5.2)] to obtain the longitudinal acceleration gx.

Gxd = vbd-ovbd (2.1.5.1) where ovbd is 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) The torque movement has a time delay from when its command value is output until it appears as actual vehicle behavior. 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 the present control, as shown in FIG. 16, a coefficient (tire slip coefficient) dgy corresponding to the tire slip state and a coefficient (tire slip coefficient) dvvbs corresponding to the tire slip state are 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 view 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 | gy
f−rgyh | <b [m / s ² ] where a and b are constants, dgy = 0 (2.2.1.4) In general, the actual horizontal G and the calculated horizontal After passing the linear region with G, the actual horizontal G does not increase like the calculated horizontal G, so the above equation (2.2.
1.3) can be transformed as follows.

Gyf = (1 + dgy) rgyh (2.2.1.3.a) When the signal leaves the linear region, dgy gradually increases from 0, and the relation 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 front and rear G estimated vehicle speed vbs calculated as described above, and the third detected simultaneously with this. 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 estimated front and rear G vehicle speed vbs, so that the longitudinal slip coefficient dvvbs of the tire decreases. Therefore, based on the longitudinal slip coefficient dvvbs of the tire, it is possible to estimate the slip state of the tire, that is, whether the tire is in the linear region where the tire is not slipping or in the nonlinear region where the tire is slipping.

Therefore, theoretically, 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 for making a circle. This drift determination coefficient srp is shown in FIG. It can be evaluated by such a drift determination circle (friction circle). FIG.
A value obtained by adjusting the coefficient of the sideslip coefficient dgy (a · dgy) and a value obtained by adjusting the coefficient of the longitudinal slip coefficient dvvbs (b · dvvb)
s) is the horizontal axis and the vertical axis is orthogonal coordinates, and the drift determination coefficient srp is equivalent 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 greater 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. Incidentally, the steering wheel angular velocity Δθh, Δθ _1, respectively Dishitah, referred both d [theta] _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 been terminated. Stop the drift control. In addition, the function of determining the end of the drift corresponding control (slip corresponding control) is referred to as an end determining unit. 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 rear wheel left / right speed difference dvhf by the correction coefficient srp1, and an LSD (limited slip differential) function that responds to the rotation speed difference, that is, a differential limit is performed. . 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 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) is for setting a high μ road gain, and the drift correction coefficient 3 (srp3) is for setting a low μ road gain.

That is, if it is determined that the vehicle is not traveling in the drift 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.

In other words, 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) to cover the reaction delay of 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 transmission torque amount smaller than that of the mechanical LSD during the normal control (other than the drift time). 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 for adjusting 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 correspond to 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.

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 a turning lateral G for such a drift response 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 is generated 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 reason why the calculated lateral G changes suddenly is that 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 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, the 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 (the yaw rate FBC action) and the action as the LSD (the 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 characteristics of the vehicle increase toward the understeer side, and the difference ddvr takes a negative value. Therefore, control is performed to move the torque in a direction to increase the difference ddvr, that is, toward the turning outer wheel. Further, during deceleration turning, the steering characteristic of the vehicle becomes stronger on the oversteer side, and the difference ddvr becomes a positive value. Therefore, control is performed so as to move the torque in a direction to decrease the difference ddvr, that is, to the turning inner wheel side.

(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, a map for a high μ road as shown in FIG. 31 (that is, a map for a high road friction resistance that gives a control amount corresponding to a high road friction resistance) and a map for a low μ road as shown in FIG. A control amount (basic control amount) corresponding to the difference ddvr from a low road surface friction resistance map giving a low road surface friction resistance control amount, that is, a high μ road basic control amount (high road surface friction resistance corresponding control amount) tbhd, A low μ road basic control amount (low road surface frictional resistance corresponding control amount) tbld is set.

As shown in FIGS. 31 and 42, similar differences d
The control amount for the high road frictional resistance map is larger than that for the low road frictional resistance map for dvr (that is, for the similar difference ddvr, the basic control amount for the high μ road (for high road frictional resistance). The control amount tbhd is set to be larger than the low μ road basic control amount (low road surface frictional resistance control amount) tbld. As shown in FIGS. Is provided with a dead zone in which the control amount is 0 in a region where the difference ddvr is small, thereby stabilizing the control.

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) Smoothing processing and limiter type filter processing are performed to suppress control hunting. 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 steering 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 (that is, a map for high road frictional resistance), and FIG. 43 is a map for a low μ road (that is, a map for low road frictional resistance). A correction coefficient (steering transient response coefficient) corresponding to the hold value dstf, that is, a high μ road steering transient response coefficient kbh1, a low μ road steering transient response coefficient kbl
Set 1. Here, these maps (FIGS. 32 and 43) have the same configuration.

(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, so that 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. FIG. 33 is a map for a high μ road (that is, a map for high road surface friction resistance), and FIG. 44 is a map for a low μ road (that is, a map for low road surface friction resistance). The correction coefficient (vehicle speed coefficient) corresponding to vb, that is, the high μ road vehicle speed coefficient kbh2, low μ
The road vehicle speed coefficient kbl2 is set.

(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 larger 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. Note that FIG.
Road Map (ie, High Road Friction Resistance Map), FIG. 45
Is a map for a low μ road (that is, a map for a low road surface frictional resistance).
gy), that is, a basic control amount for a high μ road (a control amount corresponding to a high road surface frictional resistance) tehd,
Basic control amount for low μ road (control amount corresponding to low road friction resistance) te
Set ld.

As shown in FIGS. 34 and 45, the horizontal G (gg
In the region where y is small, the same turning lateral G (ggy)
In contrast, the map for the low road frictional resistance has a larger control amount than the map for the high road frictional resistance, and in the region where the lateral G (ggy) is large, the high lateral frictional resistance is similar to the turning lateral G (ggy). The control map gives a larger control amount than the low road friction resistance map. Further, as shown in FIGS. 34 and 45, each map is provided with a dead zone in which the control amount is 0 in a region where the turning lateral G (ggy) is small, thereby stabilizing the control. Note that the dashed line in FIG. 34 indicates the characteristics of the low road surface frictional resistance map (see FIG. 45).
The chain line in the middle indicates the characteristics of the high road surface frictional resistance map (see FIG. 34).

In this embodiment, the turning lateral G
When (ggy) is equal to or more than a predetermined value, it is determined that the vehicle is turning sharply. However, even if the turning lateral G (ggy) is detected as small, even if the control amount is calculated so as to increase the rotational propulsion force of the turning outer wheel. Good. (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.

FIG. 35 is a map for a high μ road (ie, a map for high road friction resistance), and FIG. 46 is a map for a low μ road (ie, a map for low road friction resistance). Based on this, a correction coefficient (acceleration coefficient) corresponding to the acceleration (throttle opening tps), that is, a high μ road correction coefficient keh1 and a low μ road correction coefficient kel1 is set.

FIG. 36 is a map for a high μ road (ie, a map for high road friction resistance), and FIG. 47 is a map for a low μ road (ie, a map for low road friction resistance). Correction coefficient (acceleration coefficient) corresponding to G (gb) before and after the calculation, that is, a high μ road correction coefficient keh
2. A low-μ road correction coefficient kel2 is set.

Accordingly, the steering response is improved by the gain adjustment based on the longitudinal acceleration, and the delay in gain adjustment by 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 coefficients (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, 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.

FIG. 37 is a map for a high μ road (ie, a map for high road frictional resistance), and FIG. 48 is a map for a low μ road (ie, a map for low road frictional resistance). The correction coefficient (vehicle speed coefficient) corresponding to the estimated vehicle body speed (vb) based on the
The low μ road vehicle speed coefficient kel3 is set.

(4) Switch related to tack-in control As described above, the acceleration turning control is stopped during the tack-in 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, tdh
When d ≠ 0 SW: Off When tdld = 0 SW: On, When tdld ≠ 0 SW: Off (tdgd, tdld: Basic control amounts of tack-in correspondence control described later) The control amount for tack-in correspondence control described later can be alternatively used as described above, but the degree of reflection of the latter control amount is large when the necessity of tack-in correspondence is high depending on the situation of the vehicle. Thus, control data may be obtained by integrating both data.

(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, the steering characteristic of the vehicle tends to be over-steered with an increase in the cornering force of the front wheels, and the vehicle is liable to generate a tuck-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 deceleration turning, tack-in correspondence control for generating or increasing the yaw moment in the turning suppression 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.

FIG. 38 is a map for a high μ road (that is, a map for high road surface friction resistance), and FIG. 49 is a map for a low μ road (that is, a map for low road surface friction resistance). The basic control amount (control amount) corresponding to the turning lateral G (ggy), that is, the basic control amount for high μ road (control amount corresponding to high road surface friction resistance) tdhd, the basic control amount for low μ road (low road surface friction) Resistance control amount) tdld is set.

(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 tack-in control is started by capturing a deceleration command operation by the driver or the start of actual deceleration behavior of the vehicle. 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, ftd
When = 1 SW: ON (4) 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 amounts tdh and tdl are
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.

In the tack-in correspondence control, as in the case of the acceleration turning control, in a region where the turning lateral G (ggy) is small, a low road surface friction resistance map (low μ road map) is used.
May give a larger control amount than the map for high road surface frictional resistance (map for high μ road). Or,
A dead zone is provided in a small area of the turning lateral G (ggy), and a high road surface friction resistance map (high μ road map) gives a larger control amount than a low road surface friction resistance map (low μ road map). You may do so.

2.3.4 Steering Transient Response Control Steering transient response control is control performed during 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. In other words, the ECU 42 has a function of setting a steering transient response control amount (transient control amount), that is, a means steering transient response control amount setting means (transient control amount calculating means). Control amount (steering angular velocity proportional control amount) t
c can be set. For this reason, first, the basic control amounts tchd, tcld according to the steering angular velocity dθh
And a correction according to the vehicle speed, a correction according to the turning or turning back of the steering wheel, and a drift correction coefficient of 4
Drift correction is performed by (srp4), and control is performed based on 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.

FIG. 39 is a map for a high μ road (that is, a map for high road friction resistance), and FIG. 50 is a map for a low μ road (that is, a map for low road friction resistance). The basic control amount (control amount) corresponding to the steering angular velocity dθh, that is, the basic control amount for high μ road (control amount corresponding to high road friction resistance) tchd, the basic control amount for low μ road (low road surface friction resistance control) Volume) tcld.

(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, so that 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 using, for example, the maps shown in FIGS. 40A and 51A at the time of turning, and at the time of turning back, as shown in FIGS. The setting is performed using a map as shown in FIG. FIGS. 40A and 40B are maps for a high μ road (that is, a map for high road surface frictional resistance), and FIGS.
FIG. 51 (B) 0 is a map for a low μ road (ie, a map for low road frictional resistance), and a vehicle speed coefficient (correction coefficient) corresponding to the vehicle speed, that is, a vehicle speed for a high μ road, based on each of these maps. A coefficient kch and a low μ road vehicle speed coefficient kcl are set.

As shown in FIG. 58, for example, as shown in FIG. 58, when the steering wheel angle (steering angle) θh and the steering wheel angular speed (steering angular speed) dθh have the same sign (the same direction), the steering wheel is turned. 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, when the product of the handle angle θh and the handle angular velocity dθh (θh · dθh) is positive (that is, θh · dθh ≧ 0), it is determined that the steering wheel is to be turned back, and when the product is negative (ie, θh · dθh <0), it is determined that the wheel is to be turned back.

(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.

FIGS. 41 (A) and 41 (B) show high μ.
Road Map (ie, High Road Friction Resistance Map), FIG. 52
(A), FIG. 52 (B) 0 is a map for a low μ road (that is, a map for low road frictional resistance), and based on these maps, a gain-up coefficient (correction coefficient) corresponding to the steering wheel angular velocity dθh, That is, the high μ road gain-up coefficient kch and the low μ road gain-up coefficient kcl are set.

(4) Drift Correction Coefficient (srp4) During the drift control, the control amount (basic control amounts tchd and tcld multiplied by the vehicle speed coefficients kch and kcl) is multiplied by the drift correction coefficient srp4 as described above. Thus, the gains of the control amounts tch and tcl are adjusted to be appropriate. 2.4 Road Surface μ Estimation In the torque transfer control, the control effect varies depending on whether or not the road on which the vehicle is traveling is slippery, that is, the state of the road surface frictional resistance. Therefore, the road surface friction coefficient (hereinafter, road surface friction coefficient) representing the road surface frictional resistance. μ) is estimated.

Here, the estimation of the road surface μ is made up of three types: μ estimation during steady turning, μ estimation when starting, and μ estimation when nonlinear.
Perform the steps. 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 function of detecting the road surface μ, that is, the function of determining the road surface μ determination coefficient γ, is also referred to as road surface friction coefficient detecting means.

Output Value Setting As described above, the control amounts include the target ΔN follow-up control amounts tbh, tbl, the acceleration turning control amounts teh, tel, the tuck-in corresponding control amounts tdh, tdl, the steering transient response control amount tch, tcl and a control amount for high μ road (control amount corresponding to high road surface friction resistance) and a control amount for low μ road (control amount corresponding to low road surface friction resistance) are set, respectively. The output control amount tad is calculated while interpolatively reflecting the road control coefficient γ as the road surface friction coefficient calculated by the road surface friction coefficient calculation means.

That is, for each control amount, the road-surface μ determination coefficient γ is determined between those for the high μ road and those for the low μ road.
The value obtained by continuously adjusting the gain in accordance with the value of is used as the 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.

As described above, the degree of reflection of the control amount for the high μ road is set to be larger than the control amount for the low μ road when the control amounts of the high μ and the low μ are interpolatively reflected. Such 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): tc, tuck-in control: td] In this case, fine adjustment of the output value is not substantially performed. The control gain tx is calculated using the above equation (2.4.1.1). 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 is as follows, where txb is the corrected output value and b (b> 1) is the output value fine adjustment coefficient. txb = {b (txl−txh) · (γmax−γ) + γmax · txh} / γmax = [b · (γmax−γ) · txl + {γmax−b · (γmax−γ)} · txh] / γmax (2.4.1.3) Note that txb is clipped at the lower limit by txl, where 0 ≦ b · γ ≦ γmax.

As described above, the degree of reflection of the control amount for the low μ road is set to be greater than the control amount for the high μ road when the control amounts for both the high μ and the low μ are interpolated. Such setting to the low μ side is performed by controlling the control gain te of the acceleration turning control.
About. Output values (output gain) tx, txa, txb obtained by appropriately performing such fine adjustment of the output value
Can be shown as shown in FIG. 70 when the road surface μ is illustrated. In FIG. 70, an alternate long and short dash line indicates an output value txa obtained by finely adjusting the control gain tx to the high μ side.
(I.e., the target ΔN tracking control amount tb), and the solid line indicates the output value tx (ie, the tack-in corresponding control amount td, the steering transient response control amount tc) when the control gain tx is not finely adjusted. The dashed line indicates the control gain t.
It shows the characteristics of the output value txb (that is, the acceleration turning control amount te) obtained by finely adjusting the output value of x 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. 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 reason that the target ΔN following control amount tb is increased on the medium μ road is that the target ΔN following control is a control capable of maintaining the behavior stability of the vehicle even when the road surface μ is relatively low. This is because the target ΔN tracking control is emphasized and it is desired 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 the characteristic of the output value tx in FIG. 70, the inclination can be changed by a parameter such as a constant unique to the vehicle. Thus, control matching, that is, fine adjustment of the output value can be performed according to the vehicle, and more stable control can be performed.
There is also the advantage that the same basic logic can be used depending on the vehicle.

(2) High-pass processing & final output value tad Here, in order to solve the response delay, as shown in FIG. 74, the target ΔN follow-up control amount tb, the tack-in corresponding control amount td, and the acceleration turning control amount te are determined. , High-pass processing is performed. This process is performed, for example, to correct a control delay with respect to a high-frequency input due to fast steering by a high-pass process and to advance a phase of each of these control items.

That is, in driving the actuator (rotational propulsion force distribution adjusting mechanism), it is unavoidable that the response delay of the actuator with respect to the output of the control signal occurs.
Therefore, it is necessary to perform processing so that the response delay of the actuator does not deteriorate control performance. The control signal includes a control amount (transient control amount) calculated based on a driving operation state such as a steering wheel angle or a steering angle (including a steering angular velocity) θh and a throttle opening tp.
For example, a steering transient response control amount (steering angular velocity proportional control amount) t
c, a control amount calculated based on a vehicle behavior such as a difference between left and right wheel rotational speeds and a lateral acceleration generated in the vehicle (vehicle behavior corresponding control amount), for example, a target ΔN follow-up control amount tb, a tack-in corresponding control amount td, acceleration There is a turning control amount te. Driving operation is essentially the main element of the control command, and the delay of the command does not particularly matter for the control amount according to the driving operation, but since the behavior of the vehicle occurs as a result of the control command, the The control amount set based on the delay has already occurred at the time when the control signal is issued, which may cause a problem.

For example, when the vehicle behavior changes abruptly, such a delay in the output of the control amount greatly reduces the control performance. Therefore, in the present apparatus, for example, in order to correct the delay of each control response to the steering input, the control amount corresponding to the vehicle behavior, that is, the target ΔN following control amount tb, the tack-in corresponding control amount td, and the acceleration turning control amount te The high-pass processing is performed to speed up the output of the control signal. As described above, since the steering transient response control amount (steering angular velocity proportional control amount) tc is a control for advancing the phase, there is no need for correction, and the high-pass processing is not performed.

In this control, each control amount tb, td,
The final output value tad is determined by adding te and tc. That is, the ECU 42 functions as an output control amount calculating unit that individually calculates each of the control amounts tb, te, td, and tc based on various parameters, and integrates them to obtain an output value tad. . Therefore, here, the control amounts tb and t necessary for the high-pass processing are described.
For d and te, these are added in advance, and a high-pass process is performed on the added value tfd (= tb + td + te).

High-pass processing The high-pass processing is processing for extracting only high-frequency components from each control output by a high-pass filter. Here, the high-pass processing is performed on the added value tfd of the control amounts tb, td, and te for performing the high-pass processing. To obtain a high-pass processing value tff.

Tfd = tb + td + te (2.4.1.4) tff = HPF [tfd] (2.4.1.5) By this high-pass processing, a control output as shown in FIG. 75 (A) is obtained. A high-pass processing signal tff as shown in FIG. 75B is output from the signal tfd. That is, in the high-pass processing, only a portion having a large magnitude among the differential values of the control output signal tfd is output as a signal (high-pass processing value calculating means).

Further, the processing value tff subjected to the high-pass processing is added to the control output signal tfd (= tb + td + te) subjected to the high-pass processing [FIG.
(C)] to obtain an output control amount (total value) tf (output control amount calculating means). tf = tfd + tff (2.4.1.6) As shown in FIG. 75, the processing value tff is corrected by the gain (high-pass coefficient) kf (that is, tff * kf).
), The balance with other control amounts may be adjusted.

The final output value (tad) The output control amount calculating means calculates the output control amount tf as in the following equation.
Is added to the steering transient response control amount (steering angular velocity proportional control amount) tc that does not perform the high-pass processing to calculate the final output control amount (final output value) tad. tad = tf + tc (2.4.1.7) Limiter In the torque transfer control between the left and right wheels, if the amount of torque transfer is too large, the stability of the behavior of the vehicle may be reduced. The magnitude of the amount of torque movement between the left and right wheels is limited to a maximum value (which is referred to as limit) according to the friction coefficient state of the road surface (road surface μ state).

This limit value, that is, the maximum value limit is determined according to FIG.
As shown in the diagram in block B83 of FIG. 4, the relationship is set to a straight line LIM corresponding to the road surface μ determination coefficient γ. That is, the limit value limit is calculated by the following equation. limit = mg · γ + tal1 (2.4.1.8) where mg is the slope of the straight line LIM, and tal1 is l
This is the minimum value of the limit. As shown in block B83 of FIG. 74, this minimum value tal1 is a limit value limit corresponding to a road surface μ determination coefficient 1 of a low μ road, and
tal2 is a limit value limit corresponding to the road μ determination coefficient γmid of the medium μ road, and tal3 is a limit value limit corresponding to the road μ determination coefficient γmax of the high μ road. In addition,
The road μ determination coefficient γmid of the medium μ road is set to １ ／ of the road μ determination coefficient γmax of the high μ road (γmid = γmax /
2).

With such a limit value limit, the final output value tad is limited as follows. The following equation also takes into account the case where the final output value tad becomes negative depending on the torque movement direction. −limit ≦ tad ≦ limit ・ ・ ・ ・ ・ ・ (2.4.1.9)

2.5 Actuator Driving In the drive processing (actuator drive processing or proportional valve / directional valve switching control processing), as shown in FIG. tad
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)

[0192]

[Table 1]

Determination of Differential State of Left / Right Driving Force Moving Differential As shown in Table 1, here, the differential state of the rear differential relating to right / left driving force movement is determined to be five, and control is performed. 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 a normal running state. In the normal running state, 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 connected, the torque moves to the left, and when the right clutch is connected, the torque moves to the right. This is the case 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 zero, 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 higher than a predetermined value, and when the pressure switch 105 is lower 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.

2.5.3 Indicator Display Control Here, the indicator display control of the display device of the vehicle power transmission control device of the present embodiment will be described. The power transmission control device for left and right wheels for the vehicle has a control state as shown in FIG. 71 so that the driver can grasp the control state when torque transfer control (that is, power transmission control) is performed between the left and right wheels. A display device (hereinafter, referred to as an indicator) 201 having a display unit 202 is provided.

In this example, the display section 202 is configured as a three-dot display composed of three lighting sections 202A, 202B, and 202C made up of LEDs. When the control state is weak (that is, the amount of torque movement is small), only the first lighting unit (LED1) 202A is lit, and when the control state is medium (that is, the amount of torque movement is medium), the first lighting is performed. Section (LED1) 202A and second lighting section (LED2) 20
2B is turned on, and when the control state is strong (that is, the amount of torque movement is large), all of the first lighting section (LED1) 202A to the third lighting section (LED3) 202C are turned on. . Thus, the control state can be grasped as the number of lit dots (the number of LEDs).

Of course, when the control state is weak, only the first lighting section (LED1) 202A is used. When the control state is medium, only the second lighting section (LED2) 202B is used. When the control state is strong, only the third lighting section (LED3) 202C is used. Each of them may be displayed, but it is preferable to increase the number of display dots in accordance with the strength of the control state so that the state can be grasped more quickly.

As such a display device, display means other than LEDs, such as a liquid crystal display and an electric light display, may be used. Also, a graphic display (for example, a bar graph) or a numerical display instead of a dot display may be used. Is also good. Further, a display device for the left wheel and a display device for the right wheel may be provided.
In this case, there is an advantage that the torque movement state can be clearly understood.

The strength of the above-mentioned control state refers to the degree to which the torque transfer control (power transmission control) is performed, and the driver's driving operation is useful for the torque transfer control. This is a control effect such as to what extent (power transmission control) affects the behavior of the vehicle. Therefore, it is desired to display a control state so that the control effect can be grasped.

On the other hand, it is suitable and easy to use the output value tad, which is the final control amount, as an index of the normal control degree. However, as described in the section “Problems to be Solved by the Invention”, such a control amount tad does not always correspond to the control effect. That is, the control amount tad is set according to the road surface μ (road surface friction coefficient), which is a kind of the traveling environment of the vehicle. However, the control effect is higher on the low μ road than on the high μ road, and as described above. The control amount tad is set smaller as the road surface μ becomes lower. Therefore, when the control degree is displayed in such a manner as to be proportional to the control amount tad, on a low μ road,
Although the displayed control degree is low, the control effect that actually appears is large. Conversely, high μ
On a road, a situation in which the displayed control degree is high but the control effect that actually appears is small is caused.

Therefore, in the present display device, the ratio of the set control amount tad to the maximum control amount tadmax in the road surface μ (road surface friction coefficient) (= tad / tadmax) does not directly correspond to the control amount tad itself. The display amount is displayed based on the display amount. The maximum control amount tadmax corresponds to the road surface μ, and the present device uses the road surface μ determination coefficient γ as an amount indicating the road surface μ (the road surface friction coefficient). The display amount is set based on the road surface μ determination coefficient γ. The function of setting the display amount from the control amount tad or the like (converting the display amount) is referred to as a conversion unit.

In this conversion means, as shown in FIG.
Six types of display amount determination reference values for the control amount tad are provided according to the road surface μ determination coefficient γ. In FIG. 72, the horizontal axis is a value (γ / γmax) obtained by dividing the road surface μ determination coefficient γ by the maximum coefficient value γmax, and the vertical axis is the value of the control amount tad. And
The straight lines t _{1 to} t ₆ in the figure are the road surface μ judgment coefficient corresponding values (γ /
γmax) indicates the display amount determination reference value. In particular, on a low μ road (a road surface with a road surface μ determination coefficient γ of 0), the display amount determination reference values are t _L1 to t _L6 , and on a high μ road (a road surface with a road surface μ determination coefficient γ of γmax), the display amount determination is performed. The reference value is t _H1
To t _H6 . The straight lines t _{1 to} t ₆ indicating the display amount determination reference values are respectively represented by points t _{L1 to} t _L6 and points t _{H1 to} t _H6.
And a straight line connecting

Of these display amount determination reference values, t
₂ (including t _L2 and t _H2 ) is a reference value (display value) for lighting LED ₁ , and t ₁ (including t _L1 and t _H1 ) is L
This is a reference value (light-off value) for turning off the ED1. In addition, t ₄ (including t _L4 and t _H4 ) is a reference value (display value) for turning on the LED 2, and t ₃ (including t _L3 and t _H3 ) is a reference value (including the t _L3 and t _H3 ) for turning off the LED 2. Off value). T ₆ (including t _L6 and t _H6 ) is a reference value (display value) for lighting the LED ₃ , and t ₅ (t _L5 and t _H6 ).
_H5 ) is a reference value (light-off value) for turning off the LED 3.

As shown in the figure, the smaller the road μ determination coefficient γ, the larger the display amount even with the smaller control amount tad. This is because the smaller the road surface μ determination coefficient γ, the greater the control effect. Therefore, to obtain the same control effect, the smaller the road surface μ determination coefficient γ, the smaller the control amount tad.
Is set small (of course, the maximum control amount tadmax is also small). Therefore, if attention is paid to the control effect, it is necessary to increase the display amount even with the small control amount tad as the road surface μ determination coefficient γ decreases.

Therefore, the maximum control amount tadmax is set smaller for each reference value as the road surface μ determination coefficient γ is smaller. Conversely, considering the conversion gain for converting the control amount tad to the display amount, The smaller the road surface μ judgment coefficient γ, the larger the setting. Also, the reason why the reference value (display value) for turning on is set to a value larger than the reference value (light-off value) for turning off is that a so-called hysteresis for stabilizing the display is provided. Thereby, even if the control amount tad fluctuates minutely near the reference value, the display state does not change at all and a stable display can be realized.

[0216] 3. Operation of the present apparatus and effects of the present apparatus 3.1 Operation of the present apparatus The present apparatus is configured as described above.
The control is performed as shown in FIG. That is, first, the control is started based on various initial setting inputs. First, in step S10, an input calculation process as shown in FIG. 13 is executed (see item 2.1, input calculation process). Next, in step S20, a drift determination logic as shown in FIG. 16 is executed based on the result of the input calculation processing (item 2.
(See 2 drift determination logic). Further, step S
Proceeding to 30, the vehicle motion control logic is executed based on the results of the input calculation processing and the drift determination (see item 2.3, Vehicle motion control logic).

In this vehicle motion control logic, the target ΔN
Tracking control (see item 2.3.1 Target ΔN tracking control),
Acceleration turning control (refer to item 2.3.2 acceleration turning control),
Control amounts tb, td, te, tc of tack-in response control (see item 2.3.3 Tack-in response control) and steering transient response control (see item 2.3.4 steering transient response control)
Are calculated, and these control amounts tb, td, te, t
c is a high μ road control logic as shown in FIG. 29 and FIG.
By the low μ road control logic as shown in FIG. 0, each control amount tbh, tdh, teh, tch on the high μ road and each control amount tbl, tdl, tel, tcl on the low μ road.
Is calculated as

Then, the process proceeds to a step S40, at which the μ judgment logic is executed (see item 2.4 Road surface μ estimation). In this μ determination logic, a road surface μ determination coefficient γ is set (step S50), and a road surface μ determination is performed (step S50).
60), various output values are set (step S7).
0). Then, the process proceeds to step S80, where the actuator drive logic is executed as shown in FIG. 60 (item 2.
See 5 actuator drive). That is, upon receiving the output value (torque movement amount) tad, a proportional valve control signal is output to the proportional valve 106 according to the torque movement amount according to the output value tad, and the proportional valve control signal is output according to the torque movement direction according to the output value tad. A direction valve control signal is output to a direction valve (direction switching valve) 107 to drive these proportional valve 106 and direction valve 107. At the same time, a display command signal is output to the indicator lamp 110.

Such processing is performed for each required cycle T ₁ through the determination step S90. 3.2 Effects of the present apparatus As described above, according to the target ΔN tracking control, the torque transfer control that tracks the reference wheel speed difference can be appropriately realized, and the steady turning characteristics of the vehicle, that is, the steering Characteristics can be set to a desired state.

According to the acceleration turning control, by moving the torque to the turning outer wheel side, a yaw moment is generated in the turning direction, and the cornering force of the front wheel in the region where the longitudinal acceleration is large is increased to understeer. 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, the torque is moved to the turning inner wheel side to generate a yaw moment in the turning suppressing direction, thereby reducing 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 surface μ, in the μ split state, the torque is moved from the low μ wheel to the high μ road wheel. 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. 73, first, the degree of drift (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 greater 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).

In the step D20, it is determined that the drift control is not being performed, and the 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 process 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 greater 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 rotational 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, the hunting of the control can be prevented by lowering the gain for the high μ road of the present apparatus with the drift correction coefficient 2 (srp2), 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)
In the above, 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 more than a predetermined value, and that the calculation 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.

When control corresponding to vehicle behavior is performed, the output timing of a control signal corresponding to the vehicle behavior is delayed, so that the response of each actuator to the control signal is delayed. Although a hardware response delay occurs with respect to the signal, a high-pass process is performed on the control signal so that the response delay of the actuator does not degrade the control performance (see FIG. 74), so the response delay of the actuator is suppressed. Thus, appropriate control can be realized.

That is, the control amount calculated based on the driving operation state such as the steering wheel angle or the steering angle (including the steering angular velocity) θh and the throttle opening tps, for example, the steering transient response control amount (the steering angular speed proportional control amount) Regarding) tc, the delay of the control command does not particularly matter,
The control amount calculated based on the vehicle behavior, that is, the target ΔN
Regarding the follow-up control amount tb, the tack-in corresponding control amount td, and the acceleration turning control amount te, the output timing of the control signal easily amplifies the response delay of the actuator.

On the other hand, in the present apparatus, the control amount corresponding to such vehicle behavior, that is, the target ΔN follow-up control amount t
b, the control amount td corresponding to the tuck-in, and the acceleration turning control amount te are subjected to high-pass processing to increase the output of the control signal, so that the delay of each control response to the vehicle behavior is corrected. That is, as shown in FIG. 75, high-pass processing is performed on the control output signal tfd, and the high-pass processing value tff corresponding to the high-pass processing value is added to the control output signal tfd. That is, the response delay of the actuator is suppressed by that much. Therefore, even when the behavior of the vehicle changes suddenly, for example, when a sharp turning operation is performed, appropriate torque transfer control is quickly performed, and the turning performance can be improved.

According to the display device, since the control degree corresponding to the control effect that the driver actually wants to grasp is displayed, there is an advantage that the driver can accurately recognize the control situation. In particular, while controlling the control amount carefully according to the road surface μ so that it can be appropriately controlled,
This makes it possible to avoid inconsistencies in the display of the control status, which is likely to accompany the setting of the control amount according to the road surface μ.

That is, in the setting of the control amount according to the road surface μ, the control amount is set lower as the road surface μ is lower (that is, as the road surface μ determination coefficient γ is smaller), but the control amount is lower as the road surface μ is lower. However, the control effect is higher. On the contrary,
When the road surface μ is lower, the display gain of the control situation is increased. Therefore, when the road surface μ is lower, even if the control amount is lower, the display is performed with a higher display amount corresponding to the control effect. However, the control situation can be accurately recognized from the control degree corresponding to the control effect.

Further, since hysteresis is provided between the display value (lighting value) and the light-off value of the indicator 201, there is an advantage that the display becomes stable. 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 the control amount is set to be large when the road surface friction coefficient is high, At the time of the drift control, the correction according to the road surface friction coefficient is performed by performing a correction (gain adjustment) to have an intermediate value as the control amount so as to reduce the control amount change according to the road surface friction coefficient. Although the present invention is not limited to this, the present invention is not limited to this, and during drift control, the control amount correction corresponding to this road surface friction coefficient is prohibited so that the control amount does not change due to the road surface friction coefficient, An appropriate 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 with 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 this apparatus, a control amount for a high μ road (control amount corresponding to high road surface friction resistance) and a control amount for low μ road (control amount corresponding to low road surface friction resistance) are set, and these control amounts are set. The output control amount tad is calculated while reflecting the output control amount tad interpolatively according to the road surface μ determination coefficient γ as the road surface friction coefficient calculated by the road surface friction coefficient calculation means. In particular, for the target ΔN following control amount tb, For the same difference ddvr, the map for the high road friction resistance gives a larger control amount than the map for the low road friction resistance, and the degree of reflection of the high μ road control amount is set in proportion to the road friction coefficient. Since the control amount for the high μ road and the control amount for the low μ road are set to a higher μ side than the proportional distribution value for the road surface μ determination coefficient γ, the target ΔN The control gain tb of the follow-up control becomes relatively high.

The target ΔN follow-up control is performed while keeping the difference between the wheel speeds in between, so that even when the road surface friction coefficient is low, the control effect hardly deviates from the assumed area. Therefore, as described above, in the target ΔN tracking control, a relatively large torque movement is performed while the control amount for the road surface friction coefficient is set relatively large, so that the vehicle behavior can be quickly changed to the target one. The advantage of being able to do it is obtained.

Regarding the acceleration turning control amount te, in the region where the lateral G (ggy) is small, the similar turning lateral G
(Ggy), the map for the low road frictional resistance gives a larger control amount than the map for the high road frictional resistance, and
The control amount for the high μ road and the control amount for the low μ road are compared with the road μ determination coefficient γ so that the degree of reflection of the low μ road control amount is greater than the degree of reflection set in proportion to the road surface friction coefficient. Since the value is set to the μ side which is lower than the proportional distribution value, the control gain te of the acceleration turning control amount becomes relatively low.

In the case where the lateral acceleration such as the acceleration turning control is used as a parameter, particularly when the road surface friction coefficient is low, the relationship between the calculated lateral G and the actual lateral G immediately falls into a nonlinear region as shown in FIG. In other words, the control influence easily deviates from the assumed area. Therefore, as described above, in the acceleration turning control, in the region where the lateral G is small, the low road surface friction resistance map gives a larger control amount than the high road surface friction resistance map, and
The degree of reflection of the low μ road control amount (the control amount corresponding to low road surface friction resistance) is set to be greater than the degree of reflection set in proportion to the road surface friction coefficient, whereby the road surface friction coefficient is low. Even in this case, stable control can be performed quickly.

When the road surface friction coefficient is high, the high μ
By making the degree of reflection of the road control amount (control amount corresponding to high road surface friction resistance) greater than the degree of reflection set in proportion to the road surface friction coefficient, control in the region where the lateral G is small can be reduced as much as possible. Therefore, energy loss can be suppressed accordingly. Also, in the tack-in correspondence control, it is possible to set the same as the acceleration turning control,
In this case, the same effect as in the case of the acceleration turning control can be obtained.

It should be noted that such a fine adjustment of the control amount for the road surface friction coefficient is not limited to the target ΔN tracking control and the acceleration turning control, but may be applied to those that are hardly affected by the road surface friction coefficient and those that are likely to come out. it can. In addition, since the control amount setting map is provided with the dead zone, the control becomes stable.

The fine adjustment of the road friction coefficient need not be limited to the above-described embodiment. For example, as a control map for setting a control amount based on parameters such as a rotational speed difference and a lateral acceleration between left and right wheels, a high road surface friction resistance map that provides a high road surface friction resistance control amount, and a low road surface friction resistance A map for low road frictional resistance that provides a corresponding control amount is provided, and an output control amount is calculated while interpolating the high road surface frictional resistance control amount and the low road surface frictional resistance control amount in an interpolation manner according to the road surface friction coefficient. At the same time, when the two control amounts are interpolatively reflected, the degree of reflection of the high road surface friction resistance corresponding control amount and the low road surface friction resistance corresponding control amount in the middle road surface friction resistance is changed in accordance with parameters such as vehicle-specific constants. The adjustment may be made as described above. In this case, a more appropriate control amount according to the vehicle can be given.

Further, the vehicle speed, which is an important parameter of such control, is estimated in a linear region where the tires (wheels) do not have a large slip when the vehicle is traveling straight ahead and when turning, so that an appropriate estimation is performed. This has the advantage that control accuracy can be improved. That is, when the vehicle is traveling straight in the linear region, the vehicle speed vb is calculated based on the third wheel speed v3, which is the third fastest of the four left and right front wheels and the right and left rear wheels. it can. Also, when turning in the linear region, the third wheel is determined according to the geometric relationship such as the turning radius, based on the steering wheel turning angle and the vehicle body constants such as the wheel base, tread width, stability factor, and steering wheel gear ratio. Since the vehicle speed vb is calculated by correcting the speed v3, the vehicle speed vb can be easily and accurately estimated. This can contribute to an improvement in control accuracy.

In the non-linear region where the tire (wheel) has a large slip, the estimated vehicle speed based on the vehicle speed at the time of occurrence of the slip and the actual acceleration (actual longitudinal G) is used instead of the vehicle speed corresponding to the wheel speed. Thus, the estimation error of the estimated vehicle speed can be suppressed to a small value, and a relatively appropriate vehicle speed can be easily estimated. This can contribute to an improvement in control accuracy.

In the above embodiment, the target ΔN follow-up control amount tb and the acceleration turning control amount te are added to secure 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.

Although the present embodiment has been described for a four-wheel drive vehicle, the power transmission control device for left and right wheels for the vehicle is as follows.
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 to apply between front rear wheel wheel drive vehicle, this case is configured as a power transmission control device for a vehicle. In a two-wheel drive vehicle, it is also possible to use the above-described vehicle speed calculation device, and in this case, the same effect as described above can be obtained.

[0256]

As described in detail above, according to the vehicle speed calculating apparatus of the present invention, the transmission of the driving force and the braking force is performed by estimating the vehicle speed when the vehicle is traveling straight based on the third wheel speed. Therefore, even if a wheel slips, the vehicle speed at the time of going straight can be properly estimated. Further, not only the third wheel speed but also the steering wheel turning angle and a body-specific constant are added to the third wheel speed. ,
By estimating the vehicle speed at the center in the vehicle width direction at the time of turning of the vehicle, the vehicle speed at the time of turning can be properly estimated even if the wheels slip according to the turning, and the vehicle speed at the time of straight traveling and turning can be estimated. There is an advantage that it can be properly estimated (claim 1). Further
The third wheel speed, the steering wheel turning angle and the vehicle body
The turning radius on the turning inner wheel side is calculated based on the
The turning radius on the turning inner wheel side and the third wheel speed
Calculates and estimates the vehicle speed at the center in the vehicle width direction based on
This allows the vehicle speed in the center of the vehicle
It can be obtained exactly (claim 2). Also, the third wheel speed
Wheel is in the slip state (in the slip state according to the turn)
In the forward and backward direction
The third wheel speed is obtained by estimating the vehicle speed from the acceleration.
Even if the wheels are slipping (or locked)
Vehicle speed can be obtained properly (even if
3). Therefore, by using such an estimated vehicle speed, for example, control of the driving force of the vehicle can be more appropriately performed (claim 4) .

[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 according to an 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-right wheel power transmission control device according to 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-right wheel power transmission control device according to one embodiment of the present invention.

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

FIG. 5 is a schematic diagram illustrating an 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 according to 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 according to an embodiment of the present invention.

FIG. 7 is a diagram for explaining target control contents of the vehicle left and right wheel power transmission control device according to one embodiment of the present invention.

FIG. 8 is a diagram for explaining the target control contents of the vehicle left and right wheel power transmission control device according to 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 according to one embodiment of the present invention.

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

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

FIG. 12 is a diagram for explaining the target control content of the power transmission control device for left and right wheels for a vehicle according to 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 according to 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 according to the embodiment of the present invention.

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

FIG. 17 is a diagram illustrating a drift determination process of the vehicle left and right wheel power transmission control device according to the 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 according to the embodiment of the present invention.

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

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

FIG. 23 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 according to 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 a left and right wheel for a vehicle according to the 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 a left and right wheel for a vehicle according to one embodiment of the present invention.

FIG. 26 is a diagram showing a drift correction coefficient related to a drift determination process of the power transmission control device for the left and right wheels for a vehicle according to the embodiment of the present invention.

FIG. 27 is a diagram showing a drift correction coefficient related to a drift determination process of the power transmission control device for a left and right wheel for a vehicle according to the 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 a left and right wheel for a vehicle according to the embodiment of the present invention.

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

FIG. 31 is a map (high μ) related to ΔN tracking control of the vehicle left-right wheel power transmission control device according to one embodiment of the present invention.
It is a figure which shows a road map.

FIG. 32 is a map (high μ) related to ΔN tracking control of the vehicle left-right wheel power transmission control device according to one embodiment of the present invention.
It is a figure which shows a road map.

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 according to one embodiment of the present invention.
It is a figure which shows a road map.

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 according to one embodiment of the present invention.
It is a figure which shows a road map.

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 according to one embodiment of the present invention.
It is a figure which shows a road map.

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 according to one embodiment of the present invention.
It is a figure which shows a road map.

FIG. 37 is a map (high μ) related to acceleration turning control of the power transmission control device for left and right wheels for a vehicle according to one embodiment of the present invention.
It is a figure which shows a road map.

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

FIG. 39 is a view showing a map (high μ road map) related to steering transient response control of the vehicle left-right wheel power transmission control device according to one embodiment of the present invention.

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

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

FIG. 42 is a map (low μ) related to ΔN tracking control of the vehicle left-right wheel power transmission control device according to one embodiment of the present invention.
It is a figure which shows a road map.

FIG. 43 is a map (low μ) related to ΔN tracking control of the vehicle left-right wheel power transmission control device according to one embodiment of the present invention.
It is a figure which shows a road map.

FIG. 44 is a map (low μ) related to ΔN tracking control of the vehicle left-right wheel power transmission control device according to one embodiment of the present invention.
It is a figure which shows a road map.

FIG. 45 is a map (low μ) related to acceleration turning control of the power transmission control device for left and right wheels for a vehicle according to one embodiment of the present invention.
It is a figure which shows a road map.

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 according to one embodiment of the present invention.
It is a figure which shows a road map.

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 according to the embodiment of the present invention.
It is a figure which shows a road map.

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 according to one embodiment of the present invention.
It is a figure which shows a road map.

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

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

FIG. 51 is a view showing a map (low-μ road map) related to steering transient response control of the vehicle left-right wheel power transmission control device according to one embodiment of the present invention.

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

FIG. 53 is a view for explaining drift-responsive control of the power transmission control device for left and right wheels for a vehicle according to 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 according to 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 according to 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 according to one embodiment of the present invention.

FIG. 57 is a diagram illustrating tack-in correspondence control of the vehicle left-right wheel power transmission control device according to 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 according to one embodiment of the present invention.

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

FIG. 60 is a control block diagram relating to drive processing of a power transmission control device for left and right wheels for a vehicle according to an embodiment of the present invention.

FIG. 61 is a diagram showing a map for explaining a drive process of the power transmission control device for left and right wheels for a vehicle according to one embodiment of the present invention.

FIG. 62 is a diagram illustrating a map for describing a drive process of the power transmission control device between the left and right wheels for a vehicle according to the embodiment of the present invention.

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

FIG. 65 is a diagram for explaining differential discrimination in drive processing of the power transmission control device for left and right wheels for a vehicle according to one embodiment of the present invention.

FIG. 66 is a diagram for explaining differential discrimination in drive processing of the power transmission control device for left and right wheels for a vehicle according to one embodiment of the present invention.

FIG. 67 is a diagram for explaining differential discrimination in drive processing of the power transmission control device for left and right wheels for a vehicle according to one embodiment of the present invention.

FIG. 68 is a diagram for explaining differential discrimination in drive processing of the power transmission control device for left and right wheels for a vehicle according to 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 the left and right wheels for a vehicle according to the embodiment of the present invention.

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

FIG. 71 is a schematic view showing a display device of the power transmission control device for left and right wheels for a vehicle according to one embodiment of the present invention.

FIG. 72 is a diagram illustrating display control of the power transmission control device for left and right wheels for a vehicle according to an embodiment of the present invention.

FIG. 73 is a flowchart illustrating drift control of the power transmission control device for left and right wheels for a vehicle according to an embodiment of the present invention.

FIG. 74 is a main part control block diagram relating to output values obtained by road surface μ determination (road surface friction coefficient determination) of the vehicle left and right wheel power transmission control device according to one embodiment of the present invention;

FIG. 75 is a diagram illustrating a high-pass process of the vehicle left-right wheel power transmission control device according to one embodiment of the present invention.

FIG. 76 is a flowchart showing an outline of the operation of the power transmission control device for left and right wheels for a vehicle according to one embodiment of the present invention.

[Explanation of symbols]

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 as limiting means (VCU) 42 control unit the electronic control unit as a (rotational driving force distribution control means) (ECU or controller,) 48A wheel speed sensor 48B steering wheel angle sensor (handle Is angle detecting means) 48C longitudinal acceleration sensor (longitudinal G sensor) 48D lateral acceleration sensor (lateral G sensor) 48E throttle position sensor (TPS) 50 rotational driving force distribution control mechanism (rotational force adjustment means, the torque transfer mechanism) 51 differential carrier 51A Wall 52 Input shaft 54 Drive pinion gear 56 Crown gear 58 Differential case (Diff 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) H) 90R clutch (right clutch) 90AL, 90AR, 90BL, 90BR Clutch plate 92 Clutch case 96 Roller bearing 38 Hydraulic unit 101 Pressure storage unit 102 Control pressure output unit 103 Accumulator 104 Motor pump 105 Pressure switch 106 Electromagnetic proportional pressure control valve (proportional Valve) 107 Electromagnetic directional 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 Motors Corporation (72) Inventor Toshiyuki Manabe 5-33-8 Shiba, Minato-ku, Tokyo Mitsubishi Motors Corporation (56) References JP-A-6-64515 (JP, A) JP JP-A-4-50066 (JP, A) JP-A-3-54059 (JP, A) (58) Fields investigated (Int. Cl. ⁷ , DB name) G01P 3/56 B60R 16/02 B60T 8/66 B60T 8 / 58 F02D 29/02

Claims (4)

(57) [Claims] 1. A wheel speed sensor for obtaining wheel speeds of four wheels of a pair of steerable wheels and a pair of non-steerable wheels provided in a vehicle, and a wheel speed sensor of the four wheels detected by the wheel speed sensor. Straight running vehicle speed estimating means for estimating a third wheel speed among the wheel speeds as the vehicle speed when the vehicle is running straight, steering wheel turning angle detecting means for detecting a turning angle of a steering wheel provided to the vehicle, The vehicle width at the time of turning of the vehicle is obtained from the third wheel speed estimated by the straight vehicle speed estimating means, the steering angle detected by the steering angle detecting means, and a constant unique to the vehicle body. And a turning vehicle speed estimating means for calculating and estimating the vehicle speed at the center of the direction. 2. The turning vehicle speed estimating means is adapted to determine the third wheel speed.
And the steering wheel angle and the body-specific constant
Determine the turning radius on the turning inner wheel side, and
Vehicle width direction based on the turning radius and the third wheel speed
The vehicle speed at the center of the vehicle is calculated and estimated.
The vehicle speed calculation device according to claim 1. 3. The vehicle body speed estimating means according to claim 1, wherein
A longitudinal acceleration sensor for detecting the acceleration of the vehicle, and the third vehicle based on the wheel acceleration of the wheel at the third wheel speed.
The slip state of the wheel at the wheel speed is determined, and the wheel speed at the third wheel speed is determined.
When it is determined that the wheel is in the slip state, the longitudinal acceleration
Increased by the above-mentioned longitudinal acceleration detected by the sensor
A vehicle speed estimating means for slip is provided for estimating the vehicle speed.
The vehicle body according to claim 1 or 2, wherein
Speed calculation device. Wherein said both vehicle speed estimating means, and estimates the vehicle speed as a parameter of the driving force control in the vehicle driving force control apparatus, according to claim 1 to 3 Neu
A vehicle speed calculation device according to any one of claims.