JP2595812B2 - Differential adjustable front and rear wheel torque distribution control device - Google Patents

Description

Description: TECHNICAL FIELD The present invention relates to a four-wheel drive vehicle, and more particularly, to a front-wheel and a rear-wheel by adjusting a differential state between a front wheel and a rear wheel. The present invention relates to a differential adjustment type front and rear wheel torque distribution control device for controlling torque distribution.

2. Description of the Related Art Various types of power transmission devices for automobiles have been known in which the ratio of the transmission torque to the front wheels and the transmission torque to the rear wheels is controlled in accordance with the driving state.

In particular, means for controlling the torque distribution of the front and rear wheels so as to be simply linearly proportional to the longitudinal acceleration acting on the vehicle body has been proposed as control during four-wheel spin.

[Problems to be Solved by the Invention] Incidentally, if the torque distribution of the front and rear wheels is simply controlled to be linearly proportional to the longitudinal acceleration as described above, it is difficult to obtain an optimal torque distribution state.

The present invention has been made in view of the above-described problems, and has as its object to provide a differentially-adjustable front and rear wheel torque distribution control device capable of appropriately controlling torque distribution between front and rear wheels.

[Means for Solving the Problems] For this reason, the differential-adjustable front and rear wheel torque distribution control device of the present invention adjusts the differential state between the front wheel side and the rear wheel side to adjust the front and rear wheels. In the front-rear-wheel differential-adjusting four-wheel-drive vehicle that controls the torque distribution of the front and rear wheels, the differential between the front and rear wheels is limited by applying a frictional force between the front and rear wheels. Front and rear wheel differential adjustment mechanism, front and rear acceleration detection means for detecting the front and rear acceleration of the vehicle, and calculate the front wheel or rear wheel sharing load in the magnitude of the front and rear acceleration detected by the front and rear acceleration detection means The shared load calculating means, the output torque calculating means for calculating a minimum engine output torque required to obtain the magnitude of the longitudinal acceleration detected by the longitudinal acceleration detecting means, and the output torque calculating means calculated by the shared load calculating means. Sharing load and engine output A shared load torque calculating means for calculating the shared load torque from the output torque calculated by the torque calculating means, and a control means for adjusting the differential limiting amount based on the shared load torque are provided. I have.

[Operation] In the above-described differential-adjustment-type front / rear wheel torque distribution control device of the present invention, the shared load torque calculating means calculates the shared load calculated by the shared load calculating means and the engine output torque calculating means. The shared load torque is calculated from the output torque thus obtained. The control means adjusts the differential control amount based on the shared load torque.

[Embodiment] Referring to the drawings, a differentially adjusted front and rear wheel torque distribution control device as one embodiment of the present invention will be described below. FIG. 1 is a block diagram showing the configuration of a main part thereof, and FIG. FIG. 3 is a sectional view showing an essential part of the drive torque transmission system, FIG. 4 is a sectional view of an essential part of the front and rear wheel torque distribution mechanism, and FIG. 5 is a hydraulic supply system of the drive torque transmission system. FIG. 6 is a main part circuit diagram of the hydraulic pressure supply system, FIG. 7 is a diagram showing characteristics of the hydraulic pressure setting duty, and FIG.
FIG. 9 is a block diagram showing details of the steering angle data detecting means, FIG. 9 is a block diagram showing details of the vehicle speed detecting means, FIG. 10 is a diagram showing an ideal rotation speed difference setting map, and FIG. The figure shows the lateral acceleration gain setting map,
12 (a) and 12 (b) are plan views schematically showing wheel states for explaining the ideal rotational speed difference, and FIGS. 13 (a) and 13 (b) are differential clutches respectively. FIG. 14 shows a torque setting map, FIG. 14 is a block diagram showing the longitudinal acceleration corresponding clutch torque setting means, FIG. 15 is an example of the longitudinal acceleration corresponding clutch torque setting map, and FIG. 16 is an example of the engine torque map. FIG. 17 is a diagram showing an example of the transmission torque ratio map, FIG. 18 is a block diagram showing a modification of the engine torque proportional clutch torque setting means, and FIG. 19 is a center differential input torque setting map of FIG. FIG. 20 is a characteristic diagram of the protection control clutch torque, FIG. 21 (a) is a diagram showing a duty characteristic of the first preload learning, and FIG. 21 (b) is a pressure of the first preload learning. Characteristics FIG. 22, FIG. 22 is a diagram showing the pressure characteristic of the second preload learning, FIG. 23 (a) is a diagram showing the duty characteristic of the third preload learning, FIG. (C) is a diagram showing the pressure characteristic of the third preload learning, FIG. 24 is a diagram showing the torque distribution state display means, and FIG. 25 is a diagram for explaining the torque distribution by the torque distribution state estimation means. Characteristic diagram for
26 is a flowchart showing the flow of control of the entire vehicle including the device, FIG. 27 is a flowchart showing the flow of front and rear wheel torque distribution control, and FIG. 28 is a flow of setting the differential corresponding clutch torque. Flowchart, FIG. 29 is a flowchart showing the flow of setting the clutch torque corresponding to the longitudinal acceleration, FIG. 30 is a flowchart showing the flow of setting the clutch torque in proportion to the engine torque, and FIG. 31 is setting of the clutch torque for protection control. 32 is a flowchart showing a flow of the first preload learning, FIG. 33 is a flowchart showing a flow of the second preload learning, and FIG. 34 is a flow of the third preload learning. It is a flowchart which shows.

First, with reference to FIG. 2, an overall configuration of a drive system of a vehicle including the differential-adjustable front and rear wheel torque distribution control device will be described.

In FIG. 2, reference numeral 2 denotes an engine, and the output of the engine 2 is a torque converter 4 and an automatic transmission 6.
Is transmitted to the output shaft 8 via the. The output of the output shaft 8 is a planetary gear differential as a differential that distributes engine torque of the front wheels and the rear wheels to required states via the intermediate gear 10.
It is transmitted to 12.

On the other hand, the output of the planetary gear type differential device 12 is transmitted through a reduction gear mechanism 19 and a front wheel differential gear device 14 to an axle 17.
L and 17R to the left and right front wheels 16 and 18, and on the other hand, from the axles 25L and 25R to the left and right rear wheels 24 via a bevel gear mechanism 15, a propeller shaft 20 and a bevel gear mechanism 21, and a differential gear unit 22 for rear wheels. , 26. Planetary gear differential
12 includes a sun gear 121, a planetary gear 122 disposed outside the sun gear 121, and a ring gear 123 disposed outside the planetary gear 122, similarly to a conventionally known one. The output of the output shaft 8 of the automatic transmission 6 is input to the supporting carrier 125, the sun gear 121 is interlocked with the front wheel differential gear device 14 via the front wheel output shaft 27 and the reduction gear mechanism 19, and the ring gear 123 is Output shaft 2 for rear wheel
Propeller shaft 20 via 9 and bevel gear mechanism 15
It is linked to.

Further, the planetary gear type differential device 14 changes the distribution of the engine output torque between the front wheels and the rear wheels by restricting (or limiting) the differential between the front wheel side output portion and the rear wheel side output portion. A hydraulic multi-plate clutch (hereinafter also referred to as a differential limiting control mechanism) 28 as a differential limiting means or a differential adjusting means which can be provided is additionally provided.

That is, the hydraulic multi-plate clutch 28 is interposed between the sun gear 121 (or the ring gear 123) and the carrier 125, the frictional force changes according to the control pressure applied to its own hydraulic chamber, and the sun gear 121 (or The differential between the ring gear 123) and the carrier 125 is restrained.

Therefore, the planetary gear type differential device 12 appropriately controls the hydraulic multi-plate clutch 28 from a completely free state to a locked state to thereby transmit the torque transmitted to the front wheels and the rear wheels to the front wheels: the rear wheels. Can be controlled between about 32:68 to 50:50. The value of front wheel: rear wheel in a completely free state: about 32:68 can be defined by setting the ratio of the number of teeth of the input gears on the front and rear wheels of the planetary gear,
Here, the pressure is set to be about 32:68 when the pressure in the hydraulic chamber of the hydraulic multi-plate clutch 28 is zero and completely free. In addition, the ratio in the completely free state (about 32:68)
Varies depending on the load balance between the front wheel system and the rear wheel system and the like, but usually takes such a value. When the pressure in the hydraulic chamber is set to the set pressure (9 kg / cm ² ) and the hydraulic multi-plate clutch 28 is in the locked state, and the differential limit becomes substantially zero,
The torque distribution between the front wheels and the rear wheels is 50:50 and is in a directly connected state.

Reference numeral 30 denotes a steering angle sensor for detecting a rotation angle from a neutral position of the steering wheel 32, that is, a steering wheel angle θ, and reference numerals 34a and 34b detect lateral accelerations Gyf and Gyr acting on the front and rear portions of the vehicle body, respectively. In this example, two detection data Gyf and Gyr are averaged to obtain lateral acceleration data. However, only one lateral acceleration sensor 34 is provided near the center of gravity of the vehicle body, and the detected value is calculated. It may be lateral acceleration data. 36 is a longitudinal acceleration sensor that detects the longitudinal acceleration Gx acting on the vehicle body, 38 is a throttle position sensor that detects the throttle opening θt of the engine 2, 39 is an engine key switch of the engine 2,
40, 42, 44, 46 are left front wheel 16, right front wheel 18, left rear wheel respectively
26, wheel speed sensors for detecting the rotation speed of the right rear wheel 28, and the outputs of these switches and each sensor are input to a controller 48 as control means.

Reference numeral 50 denotes an anti-lock brake device, and the anti-lock brake device 50 operates in conjunction with a brake switch 50A. That is, when the brake switch 50A is turned on when the brake pedal 51 is depressed, an operation signal of the anti-lock brake is output in conjunction therewith, and the anti-lock brake device 50 operates. Further, when an operation signal of the antilock brake is output, a signal indicating the state is input to the controller 48 at the same time. Reference numeral 52 denotes a warning light that is turned on based on a control signal from the controller 48.

The controller 48 includes a CPU, a ROM, a RAM, an interface, and the like, which are not shown, but are necessary for control described later.

Reference numeral 54 denotes a hydraulic power source, and 56 denotes a hydraulic power source 54 and a hydraulic multi-plate clutch.
A pressure control valve system (hereinafter abbreviated as a pressure control valve) is provided between the hydraulic chambers 28 and controlled by a control signal from a controller 48.

The vehicle is provided with an automatic transmission. Reference numeral 160 denotes a shift lever position sensor (shift range detecting means) for detecting a selected shift range of a shift lever 160A of the automatic transmission.
Sent to 48.

Further, the engine speed Ne detected by the engine speed sensor (engine speed sensor) 170 and the transmission speed Nt detected by the transmission speed sensor (transmission speed sensor) 180 are also sent to the controller 48.

The details of the hydraulic system related to the hydraulic multi-plate clutch 28 will be described later.

In this example, a traction control system 151 is also provided. That is, the engine 2 includes the main throttle valve 152 whose opening is controlled in accordance with the amount of depression of the accelerator pedal 162A.
Together with the coupling measures, etc., it constitutes an accelerator pedal system engine output adjusting device. An auxiliary throttle valve 153 as engine output control means controlled independently of the accelerator pedal-based engine output adjusting device is provided in series with the main throttle valve 152 in the intake passage of the engine 2. The sub-throttle valve 153 is driven by a motor, and the motor is controlled based on detection results of the rear wheel speed sensors 44, 46, the front wheel speed sensors 40, 42, the engine speed sensor 170, the engine load sensor 172, and the like.

Further, as a sensor, the piston 14 of the clutch 28
A hydraulic pressure sensor 304 for detecting the hydraulic pressure applied to the 1,142 is provided at a predetermined location.

The mechanical part AM of the operation-adjustable front and rear wheel torque distribution control device will be described in more detail. This part is, as shown in FIGS. 3 and 4, an input part to which the driving force of the engine is inputted through the automatic transmission 6, It is divided into a center differential (center differential) 12, an operation restriction mechanism 28, and an output unit for the front wheel side and the rear wheel side.

The input unit includes an input gear 113 that meshes with the intermediate shaft 10A side, and an input case 124 to which the input gear 113 is serrated.
The input case 124 is rotatably mounted on the end cover 115a and the spacer member 115b fixed to the transmission case 115 via bearings 114a and 114b. The input case 124 has an enlarged diameter toward the front (to the left in FIGS. 3 and 4), and has an enlarged portion incorporating the planetary gear element and a rear portion (the fourth enlarged portion) of the enlarged portion. In the figure,
An opening is formed in front of the enlarged diameter portion with a reduced diameter portion formed on the right side). Then, from the rear (to the right in FIGS. 3 and 4) of the rear wheel output shaft 29 described later,
It can be attached to. A plurality of grooves 124d are formed on the outer periphery of the opening.

The center differential 12 is a planetary gear type center differential using a planetary gear mechanism, and includes a sun gear 121, a plurality of planetary pinions (planetary gears) 122 disposed outside the sun gear 121 so as to surround the sun gear 121, and the planetary pinion 122. Ring gear 123 arranged around
And a planet carrier 125 for supporting the planetary pinion 122, and each gear is constituted by a tooth gear.

Of these, the sun gear 121 is provided integrally with the hollow shaft member 27a, and the hollow shaft member 27a and the front wheel output shaft 27a.
Are serration-coupled to the hollow shaft 145, and through this hollow shaft 145, the hollow shaft member 27a and the front wheel output shaft 27
And can rotate together. In addition, hollow shaft
The 145 is provided with a piston accommodating portion 145a described later.

The ring gear 123 is fixed to the connection member 130, and is connected to the rear wheel output portion by the serration connection of the connection member 130 to the rear wheel output shaft 29. As a result, the output of the ring gear 123 is transmitted to the propeller shaft 2 via the connecting member 130, the rear wheel output shaft 29, and the bevel gear mechanism 15.
Input to 0.

The planet carrier 125 has, at its outer periphery, convex portions 125i that can be fitted into the respective grooves 124d of the input case 124, and these fittings are connected so as to rotate integrally with the input case 124. . The sun gear 121 is connected to a front wheel output unit, and the ring gear 123 is connected to a rear wheel output unit.

Further, a stopper 134 is provided to fix each pinion shaft 126.

A plurality of planetary pinions 122 are provided between the sun gear 121 and the ring gear 123. Each of the planetary pinions 122 is mounted on a planet carrier 125 via a pinion shaft 126.

The planet carrier 125 is coupled so as to rotate integrally with the input case 124, but has a flange-shaped base plate portion 125a.
A planetary pinion accommodating portion 125b formed forward and a cylindrical clutch disc supporting portion 125f formed rearward are provided.

These members 121, 122, 123, and 125 can be separately assembled in advance as the planetary gear mechanism unit 12, and after being assembled in this manner, the planetary gear mechanism unit 12 can be mounted in the transmission case 115. I have.

After the input case 124 is mounted in the case 115, the input case 124 is mounted so as to cover the planetary gear mechanism unit 12.

The operation limiting mechanism 28 is configured by a hydraulic multi-plate clutch, and includes an input-side disk plate 28b mounted on the clutch disk support portion 125f of the planet carrier 125,
A plurality of front wheel output side disk plates 28a mounted on a clutch case 146 which rotates integrally with the sun gear 121 and the front wheel output shaft 27 via the hollow shaft 145 are alternately arranged in parallel.

Of these, the front wheel output side disk plate 28a is the first
It is driven by the piston 141 and the second piston 142, and can be connected to the input-side disk plate 28b.

The first piston 141 and the second piston 142 are accommodated in a piston accommodating portion 145a formed on the outer periphery of the hollow shaft 145 so as to be able to move in the axial direction, respectively. , Piston receiving part 145a
A partition plate 143 that is fixed to and that does not move in the axial direction is interposed.

A first oil chamber 144a is provided between the first piston 141 and the piston housing portion 145a, and a second oil chamber 144b is provided between the second piston 142 and the partition plate 143. In these oil chambers 144a and 144b, an oil passage 117a formed in a support member 116 fixed to the transmission case 115 side and an oil passage 117b formed in the hollow shaft 145 pass hydraulic pressure (not shown). Hydraulic pressure is appropriately supplied from a supply system.

These members 28a, 28b, 141, 142, 143, 145, 146 can also be separately assembled in advance as the operation limiting mechanism unit 28, and after being subassembled in this way, the operation limiting mechanism unit 28 can be mounted in the transmission case 115. It has become.

The output section includes a front wheel output section and a rear wheel output section.The front wheel output section includes a front wheel output shaft 27 formed of a hollow shaft, and a front wheel output shaft 27 mounted on the front wheel output shaft 27. Gear 1 meshing with the input gear 19b of the operating gear device (differential) 14
9a, the rear wheel output portion is a rear wheel output shaft 29 provided to penetrate the front wheel output shaft 27,
Bevel gear shaft connected to the tip of this rear wheel output shaft 29
15A and a bevel gear 1 mounted on the bevel gear shaft 15A and meshing with a bevel gear 15b at the tip of the propeller shaft 20.
5a.

The output gear 19a is supported on the transmission case 115 via bearings 114c and 114d, and the bevel gear shaft 15A and the bevel gear 15a are supported on the transmission case 115 via bearings 114e and 114f. The reduction gear mechanism 19 is composed of the output gear 19a and the input gear 19b, and the bevel gear mechanism 15 is composed of the bevel gear 15a and the bevel gear 15b.

In FIG. 3, 101 is a converter housing, 102 is an oil pump, 103 is a front clutch, 104 is a kick down brake, 105 is a rear clutch, 106 is a low reverse brake, 107 is a planetary gear set, 108 is a transfer drive gear, 109 is a rear cover, and 112 is an end clutch.

In FIG. 4, 131 and 132 are plates for supporting each shaft in the axial direction, and 133 is an O-ring.

On the other hand, the hydraulic system relating to the hydraulic multi-plate clutch 28 is configured as shown in FIG. 5 (a schematic hydraulic circuit diagram) and FIG. 6 (a main hydraulic circuit diagram).

That is, as shown in FIG.
A pump 58 for sucking oil in the reservoir 6 is connected to a hydraulic chamber of the hydraulic multi-plate clutch 28 via a check valve 60 and a pressure control valve 62 from its discharge port. The pressure control valve 62 can have a first position in which the hydraulic chamber of the hydraulic multi-plate clutch 28 communicates with the pump 58 and a second position in which the hydraulic chamber communicates with the reservoir of the automatic transmission 6.

The passage between the check valve 60 and the pressure control valve 62 is opened at a set pressure (for example, about 9 kg / cm ² ) to supply oil to the automatic transmission 6.
A relief valve 64 is provided to escape to the reservoir, and an accumulator 66 and a pressure switch 68 are connected to this passage. The detection signal of the pressure switch 68 is input to the controller 48. Note that a motor 58a for driving the pump 58
Is controlled by a control signal of the controller 48.

Of these, the specific configuration of the pressure control valve 62 is
It is as shown in the figure.

In FIG. 6, reference numeral 161 denotes a 4WD control valve. The 4WD control valve 161 is a spool valve, and includes two valve bodies 161 provided on a spool body 161a.
b, 161c. The spool body 161a has a duty pressure (solenoid control pressure) at each end.
Receiving Pd and the reducing pressure Pr, if the duty pressure Pd decreases, it advances to the left in the figure to be in the open state, and if the duty pressure Pd increases, it advances to the right in the figure to be in the closed state. 161d is a spring that biases the spool body 161a in an appropriate direction so that the spool body 161a can move appropriately as described above.

162 is a duty solenoid valve (duty valve).
62a, this solenoid 162a and return spring 162c
The valve element 162b is retracted to open the oil passage 169f when the solenoid 162a is operated, and is advanced by the return spring 162c to close the oil passage 169f when the solenoid 162a is not operated. It has become. The reduty valve 162 is electronically controlled by a controller (computer) 48 based on information from various sensors.

163 is an orifice, 164 is an oil filter, 165
Denotes a reducing valve, an orifice 163 is provided between the reducing valve 165 and the 4WD control valve 161, and an oil filter 164 is provided in an oil passage 169b flowing into the reducing valve 165.

In the reducing valve 165, the valve element 165a is urged at a predetermined pressure by a return spring 165b. With this urging force, when the oil pressure of the valve element 165a falls below the set pressure, the oil pressure is supplied, and the oil pressure is set. When the pressure exceeds the pressure, it automatically moves so as to discharge the hydraulic pressure.

Therefore, for example, when the solenoid 162a operates and the duty valve 162 opens, the 4WD control valve 16
As the hydraulic pressure (duty pressure) Pd on the left end side of 1 decreases and the valve bodies 161b and 161c move leftward by the return spring 161d, the line between the oil passages 169c and 169g is opened, and the line compression P ₁ is the oil chamber 144a of the hydraulic pressure (4WD clutch pressure) P ₄ as the hydraulic multiple disk clutch 28, is supplied to the 144b, the hydraulic multi-plate clutch 28 is configured to be connected.

When the duty valve 162 is closed without operating the solenoid 162a, the hydraulic pressure (duty pressure) Pd on the left end side of the 4WD control valve 161 increases, and the valve body 16
1b and 161c move to the right (the position shown in FIG. 6), the oil passages 169c and 169g are disconnected, and the 4WD clutch pressure P ₄
Is released, and the hydraulic multi-plate clutch 28 is separated.

The duty (Duty) which is a control index of the duty valve 162 and the 4WD clutch pressure P ₄ (= control oil pressure P)
The relationship, for example, as shown in FIG. 7, as shown, duty and the lower the 4WD clutch pressure P ₄ lower, the higher the 4WD clutch pressure P ₄ duty increases are higher. Incidentally, the opposite setting, that is, characteristics become downward-sloping straight line, the duty is the higher the 4WD clutch pressure P ₄ lower, the higher the 4WD clutch pressure P ₄ duty increases conceivable configuration becomes lower.

Next, a planetary gear type differential device is
With reference to the block diagram of FIG. 1, the components of the controller relating to the control for restraining the twelve differentials (hereinafter referred to as driving force distribution control or center differential control) will be described.

In this control, each sensor (wheel speed sensors 40, 42, 44, 46,
Based on the detection information from the steering angle sensors 30a, 30b, 30c, the lateral acceleration sensor 34, the longitudinal acceleration sensor 36, the throttle position sensor 38, the engine speed 170, the transmission speed sensor 180, the shift position sensor 160, etc. The clutch torque of the multi-plate clutch 28 is set, and the differential hydraulic pressure of the hydraulic multi-plate clutch 28 is controlled so as to obtain a target clutch torque.

The data includes ABS information, wheel speed, steering angle, gear position,
Integrated communication between the ABS control unit and the engine control unit (SCI communication: SCI = Serial Communication In)
Data such as terface) is digitally input, and analog input is applied to longitudinal acceleration, lateral acceleration, accelerator opening, hydraulic control to a multi-plate clutch, 4WD control unit control, current to a rear differential electromagnetic clutch, and the like.

The setting of the clutch torque of the hydraulic multi-plate clutch 28 is made so that the ideal differential state is obtained by focusing on the differential state between the front wheels and the rear wheels (the rotational speed difference is also expressed as the rotational speed difference). Differential clutch torque Tv for performing control
And proportional to the engine torque so that a large road surface transmission torque can be obtained in the front-rear-wheel direct-coupled four-wheel drive state at the time of sudden start, etc. One of an engine torque proportional clutch torque Ta and a protection control clutch torque Tpc for protecting the clutch portion of the wet multi-plate clutch. Torque Tv, Tx, Ta, Tpc
Will be described in order.

The differential-compatible clutch torque Tv is a clutch torque that causes the vehicle to behave in accordance with the driver's will when turning, and in order to control the posture of the vehicle body, the rear wheel should be slipped from the rear wheel as a drive base. Is effective, the differential corresponding clutch torque Tv is set to realize such a state.

For this reason, as shown in FIG. 1, the part related to the setting of the differential corresponding clutch torque Tv includes a front and rear wheel actual rotation speed difference detection unit 200, a front and rear wheel ideal rotation speed difference setting unit 210, and a front and rear wheel actual rotation speed rotation. The clutch includes a differential corresponding clutch torque setting unit 220 that sets a clutch torque Tv ′ based on the speed difference ΔVcd and the ideal front and rear wheel rotational speed difference ΔVhc, and a correction unit 246 that corrects the clutch torque Tv ′ for lateral acceleration. The rotation speed is also referred to as a rotation speed.

The front and rear wheel actual rotation speed difference detection unit 200 includes filters 202a to 202
d, a front wheel rotation speed data calculation unit 204a, a rear wheel rotation speed data calculation unit 204b, and a front and rear wheel actual rotation speed difference calculation unit 206.

The filters 202a to 202d respectively include wheel speed sensors 40, 42,
From the rotational speed data signals FL, FR, RL, RR of the left front wheel 16, the right front wheel 18, the left rear wheel 26, and the right rear wheel 28 detected by 44, 46, a minute vibration component of data generated by disturbance or the like. Is to get rid of.

In the front wheel rotation speed data calculation unit 204a, the front wheel rotation speed Vf is obtained by averaging the front wheel rotation speeds obtained from the front wheel rotation speed data signals FL and FR, and the rear wheel rotation speed data calculation unit 204b. , Rear wheel rotation speed data signals RL, R
The rear wheel rotation speed Vr is obtained by averaging the respective wheel speeds of the rear wheels obtained from R.

Further, the front and rear wheel actual rotation speed difference calculation unit 206 subtracts the front wheel rotation speed Vf from the rear wheel rotation speed Vr to obtain the actual rotation speed difference between the front and rear wheels [rotational speed difference between front and rear wheels (front and rear rotation difference: This is also the rotation difference in the center differential.)] Calculate ΔVcd.

The front and rear wheel ideal rotational speed difference setting unit 210 includes a driver request steering angle calculation unit (pseudo steering angle calculation unit) 212 as steering angle data detection means, and a driver request body speed calculation unit (vehicle speed data detection unit). A pseudo vehicle speed calculation unit) 216 and an ideal rotation speed difference setting unit 218 as an ideal operation state setting unit are provided.

As shown in FIG. 8, the driver-requested steering angle calculation unit 212 as the driver-requested steering angle data setting means includes a steering angle sensor 30 (first steering angle sensor 30a, a second steering angle installed on the steering wheel). Sensor 30b, neutral position sensor
Detection data theta ₁ from 30c), θ _2, and the sensor corresponding steering angle _{δh [= f (θ 1,} θ 2, θn)] sensor corresponding steering angle data setting unit 212a for calculating the value of on the basis of .theta.n, horizontal Lateral acceleration data calculator 212b that averages data Gyf, Gyr detected by acceleration sensors 34a, 34b to calculate lateral acceleration data Gy
And a comparison unit 212c that compares the direction of the sensor-assisted steering angle δh with the direction of the lateral acceleration data Gy; and a driver required steering angle setting that sets a driver required steering angle δref according to the comparison result in the comparison unit 212c. (A vehicle speed data setting unit) 212d.

In addition, a function δh = f for obtaining the sensor corresponding steering angle δh.
(Θ ₁ , θ ₂ , θn) correspond to the specifications of the steering wheel angle sensor.

In addition, the sensor corresponding steering angle δh and the lateral acceleration data Gy
Are positive, for example, in the right turning direction.

These sensor corresponding steering angles δh and lateral acceleration data Gy
In order to compare the directions, a function SIG (x) regarding the following direction is set for the detection data x.

When x> 0, SIG (x) = 1 When x = 0, SIG (x) = 0 When x <0, SIG (x) = − 1 Therefore, in the comparison unit 212c, the sensor-related steering angle δh SIG (δh)
And SIG (Gy).

When the direction SIG (δh) of the sensor-associated steering angle δh is equal to the direction SIG (Gy) of the lateral acceleration data Gy, the driver-requested steering angle setting unit 212d sets the sensor-associated steering angle δh to the driver request. When the steering angle (steering angle data) δref is set, and the direction SIG (δh) of the sensor corresponding steering angle δh and the direction SIG (Gy) of the lateral acceleration data Gy are not equal,
0 is set to the driver required steering angle δref.

When the direction SIG (δh) of the sensor-assisted steering angle δh is not equal to the direction SIG (Gy) of the lateral acceleration data Gy, the driver sets the required steering angle δref to 0, for example, when the driver sets a steering wheel such as a counter steer. When the operation is performed, the steering position of the steering wheel may be different from the actual steering angle (turning state) of the vehicle. In such a case, setting the steering angle of the vehicle from the steering position of the steering wheel to perform appropriate control Difficult to do.

Therefore, in order to eliminate such a problem, the direction SIG (δh) of the sensor corresponding steering angle δh and the lateral acceleration data Gy
If the direction SIG (Gy) is not equal, the driver-requested steering angle is set to zero.

As shown in FIG. 9, the driver-requested vehicle speed calculating section 216 outputs the left front wheel 1 detected by the wheel speed sensors 40, 42, 44, 46.
6, rotation speed data signal F of right front wheel 18, left rear wheel 26, right rear wheel 28
A wheel speed selecting unit 216a for selecting the second largest wheel speed data from the bottom (lowest one) of L, FR, RL, and RR, and setting the driver requested vehicle speed from the selected wheel speed data and the like And a driver request vehicle speed calculation unit 216c.

In particular, the driver-requested vehicle speed calculation unit 216c applies the wheel speed data SVW obtained by applying the wheel speed data selected by the wheel speed selection unit 216a to the filter 216b to remove noise components, and the front-rear data detected by the front-rear acceleration sensor 36. Filter acceleration 216d
Longitudinal acceleration data obtained by removing noise components over
Based on Gx, the subsequent vehicle speed is estimated from both data SVW and Gx at a certain point in time. In other words, if the wheel speed data SVW at a certain point in time is V ₂ and the longitudinal acceleration data Gx is a, the theoretical vehicle speed at time t after this point in time
Vref can be calculated by Vref = V ₂ + at.

Also, instead of the longitudinal acceleration data Gx, the wheel speed data SV
W or a driver required vehicle acceleration V′ref obtained by time-differentiating the driver required vehicle speed Vref may be employed.

The second largest wheel speed data among the rotation speed data signals FL, FR, RL, and RR is adopted because each wheel usually slips to the overspeed side in many cases. Then, it is desirable to adopt the wheel speed of the lowest rotation, but the second wheel speed from the bottom is adopted in consideration of data reliability.

Then, in the ideal rotation speed difference setting unit 218, the driver required steering angle δref calculated by the driver required steering angle calculation unit 212
Then, an ideal rotation speed difference ΔVhc is set based on the map shown in FIG. 10 from the driver required vehicle speed Vref calculated by the driver required vehicle speed calculation unit 216. In other words, regarding the vehicle speed, when the vehicle speed is low, the influence of the difference between the starting radii of the front and rear wheels at the time of turning (so-called inner wheel difference) is large, and the rotational speed Vr
Is lower than the rotation speed Vf of the front wheels, but as the vehicle speed increases, the rotation speed Vr of the rear wheels increases with respect to the rotation speed Vf of the front wheels so that the rear wheels can easily slip at high speeds I have to. As a result, the responsiveness of the posture of the vehicle body, which is required at higher speeds, is ensured. Also,
Regarding the steering angle, the larger the steering angle, the larger the rotation difference required for the front and rear wheels, and thus the larger the magnitude | δref | of the steering angle data δref, the larger the value of ΔVhc.

The rotation speed difference ΔVhc between the front and rear wheels due to the difference in the starting radius between the front and rear wheels will be described with reference to FIGS. 12 (a) and 12 (b). FIG. 12 (a) is a diagram schematically illustrating a two-wheeled vehicle including one front wheel and one rear wheel, and FIG. 12 (b).
FIG. 12 is a diagram further schematically showing FIG. 12 (a). As shown in FIGS. 12 (a) and 12 (b), the front wheel speed is Vf, the rear wheel speed is Vr, the vehicle speed at the center of gravity of the vehicle is V, the turning radius Rf of the front wheel, and the turning radius of the rear wheel is Rr, the radius of gyration of the center of gravity of the vehicle is R, the body slip angle is β, the wheelbase is 1, the distance between the center of the front wheel and the center of gravity is 1f, and the distance between the center of the rear wheel and the center of gravity is 1r. The rotation speed difference ΔVhc is expressed as follows.

ΔVhc = Vr−Vf = [(Rr−Rf) / R] · Vref (1.1) Rr = {R ² + 1r ² −2Rlr · cos (π / 2−β)} ^1/2 Rf = ｛ R ² + 1r ² −2 Rlf · cos (π / 2 + β)｝ ¹ / 2β = (1−m / 2l·lf / lr · kr · V) / (1 + A · V ² ) · 1
r / 1 · δ where m is the vehicle weight, kr is the cornering power, and A is the stability factor.

When the front wheel speed Vf and the rear wheel speed Vr are considered to be theoretical, Vf: Vr = Rf: Rr, Vf: V = Rf: R, and the angle βf shown in FIG. , βr is βf-βr
= Is related to AV ^2, from each expression of these relationships as described above,
ΔVhc can be defined as a function of V and δ [ΔVhc = fc (V, δ)]. However, V in this case corresponds to a theoretical value, that is, the driver's required vehicle body speed Vref, and δ also corresponds to a theoretical value, that is, the driver's required steering angle δref. Such a function [ΔV
When hc = fc (Vref, δref)] is mapped, the result is as shown in FIG.

By the way, as for the steering angle, in addition to the actual steering angle (steering angle corresponding to the sensor) δh based on the steering wheel angle θ, there is a turning G equivalent steering angle δy obtained from the lateral acceleration (turning G) Gy during turning.
This turning G equivalent steering angle δy can be calculated by the following equation.

δy = [(1 + A · Vref ² ) / Vref ² ] · 1 · Gy
(1.2) where A is a stability factor, Vref is a theoretical vehicle speed (driver required vehicle speed) described later, and 1 is a wheel base.

For the turning G equivalent rudder angle δy obtained in this way,
The actual steering angle (steering angle corresponding to the sensor) δh is a steering angle that more reflects the driver's intention. That is, if the driver wants to make a larger turn than the current situation, | δh |> | δy |
When the steering angle value | δh | is adopted, the magnitude of the ideal rotation speed difference (slip target value) can be made larger than when the steering angle value | δy | is adopted, while the driver wants to suppress the current turning. | Δh | <| δy |, and the magnitude of the ideal rotational speed difference (slip target value) can be made smaller by employing the steering angle value | δh | than by employing the steering angle value | δy |. is there.

As described above, the front and rear wheel actual rotation speed difference ΔVcd detected by the front and rear wheel actual rotation speed difference detection unit 200 and the front and rear wheel ideal rotation speed difference ΔVhc set by the front and rear wheel ideal rotation speed difference setting unit 210 are: Is subtracted (ΔVcd−ΔVhc) by the subtractor 222,
The obtained difference ΔVc (= ΔVcd−ΔVhc) and the front and rear wheel ideal rotational speed difference ΔVhc are input to the differential-compatible clutch torque setting unit 220 as data.

The differential corresponding clutch torque setting section 220 calculates the difference ΔVc between the front and rear wheel actual rotation speed difference ΔVcd and the front and rear wheel ideal rotation speed difference ΔVhc.
(= ΔVcd−ΔVhc), the clutch torque Tv ′ is set. The clutch torque Tv ′ is set depending on whether the front and rear wheel ideal rotational speed difference ΔVhc is positive or negative.

(I) When .DELTA.Vhc.gtoreq.0 In this case, the speed of the rear wheels is desired to be higher than that of the front wheels, and the clutch torque Tv 'is set as follows.

If .DELTA.Vcd.gtoreq..DELTA.Vhc, the rear wheels are slipping due to excessive rotation, and it is desired to suppress the rear wheels from slipping by transferring part of the engine torque largely distributed to the rear wheels toward the front wheels. At this time, since ΔVcd ≧ 0, when the clutch is connected, the torque moves from the rear wheel side to the front wheel side. Therefore, Tv ′ = a × (ΔVcd−ΔVhc) = a × ΔVc (1.3) so that the clutch torque Tv ′ increases in proportion to the magnitude of the difference ΔVc (ΔVcd−ΔVhc) (where 1.3). , A is a proportional constant).

If ΔVhc>ΔVcd> 0, the front wheels are slipping, but at this time, since ΔVcd> 0, when the clutch is connected, the torque moves from the rear wheels to the front wheels, so if the clutch torque Tv ′ is If it is increased, the engine torque distributed to the front wheels increases, and the slip of the front wheels is promoted. For this reason, it is desired to reduce the engine torque distributed to the front wheels by making the differential limitation free.
Therefore, in this case, the clutch torque Tv 'is set to 0 to set a so-called dead zone.

If 0 ≧ ΔVcd, the front wheels are slipping, and it is desired to increase the distribution of the engine torque to the rear wheels to reduce the slip of the front wheels. At this time, when the clutch is connected, the torque moves from the front wheel side to the rear wheel side, so that Tv ′ = − a × ΔVcd = −a so that the clutch torque Tv ′ increases in proportion to the magnitude of ΔVcd. × (ΔVc + ΔVhc) (1.
4) (where a is a proportional constant).

By mapping such a relationship between Tv ′ and ΔVc,
As shown in FIG. 13 (a), the difference ΔVc
The differential corresponding clutch torque Tv can be obtained from the difference between the front and rear wheel ideal rotational speed difference ΔVhc.

When ΔVhc = 0, there is no dead zone region of ΔVhc>ΔVcd> 0.

(Ii) When ΔVhc <0 In this case, it is desired that the speed of the front wheels be faster than that of the rear wheels, and the clutch torque Tv ′ is set as follows.

If .DELTA.Vcd.gtoreq.0, the rear wheel is over-rotating and slipping, and it is desired to suppress a rear wheel slip by transferring a part of the engine torque largely distributed near the rear wheel to the front wheel side. At this time, when the clutch is connected, the torque moves from the rear wheel side to the front wheel side. So, clutch torque
Tv ′ = a × ΔVcd = a × (ΔVc + ΔVhc) (1.5) so that Tv ′ increases in proportion to the magnitude of ΔVcd (where a is a proportional constant).

If 0>ΔVcd> ΔVhc, it is desirable to promote the rotation of the front wheels more than the rear wheels. At this time, if the clutch is connected, the torque will move from the front wheels to the rear wheels because 0> ΔVcd. When Tv 'is increased, the engine torque distributed to the rear wheels increases, and the slip of the rear wheels is promoted. For this reason, it is desired to reduce the engine torque distributed to the front wheels by making the differential limitation free. Therefore, in this case, the clutch torque Tv ′
Is set to 0 to set a so-called dead zone area.

If .DELTA.Vhc.gtoreq..DELTA.Vcd, the front wheels are slipping excessively. Therefore, it is desired to increase the distribution of the engine torque to the rear wheels to reduce the slip of the front wheels. In this case, 0> ΔVcd
Therefore, when the clutch is connected, the torque moves from the front wheels to the rear wheels. Therefore, the clutch torque Tv ′ is ΔVc (Δ
Tv ′ = − a × (ΔVcd−ΔVhc) = − a × ΔVc (1.
6) (where a is a proportional constant).

By mapping such a relationship between Tv ′ and ΔVc,
As shown in FIG. 13 (b), the difference ΔVc
The differential corresponding clutch torque Tv can be obtained from the difference between the front and rear wheel ideal rotational speed difference ΔVhc.

In this manner, the differential corresponding clutch torque setting section 220
With reference to the map [FIGS. 13 (a) and 13 (b)], the differential corresponding clutch torque Tv 'obtained from ΔVc and ΔVhc
Are corrected by the correction unit 246 in the lateral acceleration.

In the correcting unit 246, the differential G
Subjected to lateral acceleration corrected by multiplying the gain k _1, but is adapted to obtain a differential response clutch torque Tv, the lateral G gain k ₁ is set as follows.

In other words, the detection data Gy from the lateral acceleration sensor 34 is sent to the lateral G gain setting unit 244 after the fine vibration component of the data generated due to disturbance or the like is removed through the filter 242. In the lateral G gain setting unit 244 sets the lateral G gain k ₁ from the lateral acceleration data Gy according to the map shown in the block in the first diagram of the setting portion 244.

The lateral G gain k ₁ is seeks to reflect control the state of the friction coefficient of the road surface mu, as the lateral acceleration Gy increases can determine that the road surface mu is large, as the road surface mu is large, the engine torque I want to be able to give priority to the turning performance of the car body with the rear wheel as the main distribution. Then, when the size of the road surface μ (hence, the size of the lateral acceleration Gy) increases,
The lateral G gain k ₁ is decreased, thereby performing the correction for reducing the set clutch torque Tv. The road surface μ
Even if a large, if not special priority to vehicle turning property, it is conceivable to omit the correction by the lateral G gain k _1.

The longitudinal acceleration-dependent clutch torque Tx is a clutch torque for preventing the vehicle from being strongly understeered and enabling the vehicle to perform a smooth turning operation, and performs control in accordance with the longitudinal acceleration Gx acting on the vehicle. It has become.

The setting of the clutch torque Tx corresponding to the longitudinal acceleration is performed by the clutch torque setting means 254 corresponding to the longitudinal acceleration, and the detection data Gx from the longitudinal acceleration sensor 36 is applied to the filter 252.
After the micro-vibration component of the data generated due to disturbance or the like is removed, the data is sent to the clutch torque setting means 254.

As shown in FIG. 14, the clutch torque setting means 254 includes a front wheel shared load calculating means (shared load calculating means) 254a.
And total output torque calculating means (output torque calculating means) 254b
And front wheel shared torque calculating means (shared torque calculating means) 25
4c and clutch torque calculating means 254d.

In the front wheel shared load calculating means 254a, the longitudinal acceleration data Gx
The front wheel shared load Wf 'at the time of acceleration is obtained from the following equation. The front wheel shared load Wf' is calculated based on known values such as the front wheel shared load Wf at rest, the vehicle weight Wa, the height of the center of gravity h, the wheel base 1, and the longitudinal acceleration. From the data Gx, it is obtained by the following equation.

Wf ′ = Wf− (h / 1) · Wa · Gx (2.1) In the total output torque calculating means 254b, the longitudinal acceleration data Gx
The required total output torque (which is the engine output torque considered on the propeller shaft) Ta is calculated from
Is calculated from the vehicle weight Wa, the tire radius Rt, the final deceleration ratio (the average value in the rear differential and the friction differential) ρ, and the longitudinal acceleration data Gx by the following equation.

Ta = Wa · Gx · Rt / ρ (2.2) In the front wheel shared torque calculating means 254c, the front wheel shared load Wf 'during acceleration obtained by the front wheel shared load calculating means 254a and the total output torque calculating means 254b are obtained. From the required total output torque Ta, the front wheel shared torque Tf is determined by the following equation.

Tf = (Wf ′ / Wa) · Ta (2.3) In the clutch torque calculating means 254d, the required total output torque Ta obtained by the total output torque calculating means 254b and the front wheel shared torque obtained by the front wheel sharing torque calculating means 254c. From Tf, the clutch torque Tx ′ corresponding to the longitudinal acceleration is calculated.

That is, the front wheel torque distribution Tf by the center differential 12 and the differential limiting clutch 28 is expressed as follows, assuming that the rear slip precedes.

Tf = [Zs / (Zs + Zr)] ・ Ta + [(Zr / (Zs + Zr)] ・ T
x '(2.4) where Zs is the number of teeth of the sun gear 12a, and Zr is the number of teeth of the ring gear 12c.

 Equation (2.4) can be modified as follows.

Tx '= Tf- [Zs / (Zs + Zr)] Ta / [Zr / (Zs + Zr)]
・ ・ ・ (2.4 ′) Therefore, the required total output torque Ta and the front wheel shared torque Tf
Thus, the clutch torque Tx 'corresponding to the longitudinal acceleration can be obtained.

On the other hand, from equations (2.1) to (2.4), Wf ', Tf, and Ta are eliminated and Tx' is solved for Gx. First, equations (2.1) and (2.2) are substituted into equation (2.3). , Tf = (Rt / ρ) · (Wf · Gx−h / l · Wa · Gx ² ) (2.
5) From the equations (2.1), (2.4), and (2.5), Tx '= − A · C · (Gx−B / 2C) ² + AB / 4C (2.6) where A = [(Zs + Zr) / Zr] · (Rt / ρ) B = Wf− [Zs / (Zs + Zr)] · Wa C = (h / l) · Wa where the constants related to the constants A, B, and C are Zs = 28, Zr = 60, R
t = 0.296 (m), ρ = 3.6, Wf = 880 (kg), Wa = 1595 (k
g), h = 0.55 (m), 1 = 2.6 (m), Tx '= -4
0.7 (Gx−0.552) ² +12.4, and Tx ′ is the first with respect to Gx.
It can be represented by a quadratic curve as shown in FIG.

However, when Gx ≒ 0.55, Tx ′ takes a maximum value, and Tx ′ decreases in the region of Gx> 0.55. Here, considering the safety of control, the maximum value of Tx ′ is also taken in the region of Gx> 0.55. Is set to a constant value equal to Note that such a setting can also be applied when calculating the clutch torque Tx 'corresponding to the longitudinal acceleration by the clutch torque calculating means 254d.

The longitudinal-acceleration-dependent clutch torque setting means 254 may directly calculate the longitudinal-acceleration-dependent clutch torque Tx ′ from the longitudinal acceleration data Gx based on such a map (see FIG. 15).

The clutch torque corresponding to the longitudinal acceleration set in this way
Tx ′ is corrected by the lateral acceleration correction unit 256. In the correction unit 256, a similar correction and the correction unit 246 described above it is subjected to a lateral acceleration correction by multiplying the lateral G gain k ₁ in longitudinal acceleration corresponding clutch torque Tx ', to obtain the longitudinal acceleration corresponding clutch torque Tx , But this horizontal G gain
k ₁ is omitted is described above, since the aim is also an seeks to reflect control the state of the friction coefficient μ of the road surface in the same manner as described above herein described.

The clutch torque Tx corresponding to the longitudinal acceleration corrected in this manner is output as data when the switch 258a is turned on and off. The switch 258a is turned on when the front wheel speed Vf is higher than the vehicle speed Vref, that is, when the front wheels are slipping (at the time of front slip), based on a signal from the judging means 258.
Is turned on, and turned off in other cases. Therefore,
The clutch torque Tx corresponding to the longitudinal acceleration set only at the time of the front slip is output, and is not output in other cases (in this case, Tx = 0, hereinafter, generally, the switch is turned off and the clutch torque is not output. At times, the value of the clutch torque is set to 0).

The engine torque proportional clutch torque Ta is set in advance in the directly connected four-wheel drive state so that the initial slip of the rear wheels can be prevented when the transmission torque is expected to increase when the vehicle suddenly starts from a stopped state. This is a set torque for setting.

Therefore, the portion for setting the engine torque proportional clutch torque Ta (engine torque proportional clutch torque setting means) includes an engine torque detecting section 264 for detecting the engine torque Te at a certain moment as shown in the lower left portion of FIG.
And a torque converter torque ratio detector 266 for detecting the torque converter torque ratio t at that time, and a transmission reduction ratio detector for detecting the transmission reduction ratio ρm at that time.
276 and an engine torque proportional torque setting unit 268 for obtaining an engine torque proportional torque Ta ′ from the engine torque Te based on a map set in a proportional relationship with the engine torque Te.
And the engine torque proportional torque Ta ′ is added to the torque converter torque ratio t, the transmission reduction ratio ρm, and the final reduction ρ.
_The engine torque proportional clutch torque calculating section 270 that obtains the engine torque proportional clutch torque Ta by multiplying ₁ and the rotation difference gain k ₂ , and the set engine torque proportional clutch torque Ta at low speed (for example, Vref <20 km / h) And a switch 274a that outputs only data.

In the engine torque detector 264, the throttle opening data θth sent from the throttle position sensor 38 through the filter 262a to remove the micro-vibration component of the data generated by disturbance or the like, and the filter 262b sent from the engine speed sensor 170 Engine speed data from which micro-vibration components of data generated by disturbances have been removed
From Ne, the engine torque Te at that time is obtained through an engine torque map as shown in FIG. 16, for example.

In the torque converter torque ratio detection unit 266, the engine speed data Ne sent from the engine speed sensor 170 and the disturbance component is removed through the filter 262b, and the disturbance component sent from the transmission speed sensor 180 and removed through the filter 262c are removed. The transmission torque ratio t at that time is obtained from the transmission speed data Nt through a transmission torque ratio map as shown in FIG. 17, for example.

The transmission reduction ratio detecting section 276 obtains the transmission reduction ratio ρm from the selected shift stage information from the shift position sensor 160 with reference to a shift stage-reduction ratio correspondence map as shown in a block 276 in FIG. It has become.

In the map used for setting of the engine torque proportional torque setting unit 268 (see the block 268 in FIG. 1), the engine torque Te and the engine torque proportional torque Ta ′ indicate the number of teeth Zs and Zr of the sun gear and the ring gear, and the front wheel share. A linear relationship follows a proportionality constant determined from known constants such as the load Wf and the vehicle weight Wa.

In the engine torque proportional clutch torque calculation unit 270,
Engine torque proportional torque T determined as described above
and a ', the torque converter torque ratio t, transmission reduction ratio .rho.m, but operation from a final reduction [rho ₁ and the rotational difference gain k ₂ Metropolitan is performed, the rotational difference gain k ₂ is rotated difference gain setting unit 275
Is set as follows.

That is, the rotational difference gain k ₂ is intended to avoid the tight corner braking phenomenon, the ideal rotational speed difference setting portion 218
Is determined from the ideal rotational speed difference ΔVhc set in accordance with the map shown in FIG. Rotation difference gain k ₂ in this map the relationship between the ideal rotational speed difference ΔVhc is expressed by the following equation.

K ₂ = 0.9 × (| ΔVhcmax || ΔVhc |) /|ΔVhcmax|+0.1
... (3.1) where, ΔVhcmax = MAX | ΔVhc (δ = MAX) | The coefficient 0.9 and constant 0.1 is to the lower limit of k ₂ to 0.1.

As described above, as the ideal rotational speed difference ΔVhc increases, the rotational difference gain k ₂ decreases linearly, and the rotational speed gain k ₂ is multiplied and corrected to obtain the ideal rotational speed difference ΔVhc during turning or the like. Is larger, the engine torque proportional clutch torque Ta is reduced so that turning performance (performance that can prevent the tight corner braking phenomenon) is prioritized over sudden starting performance.

Incidentally, the above-described engine torque proportional torque setting unit 26
As shown in FIG. 18, a part of the engine torque proportional clutch 8 and the engine torque proportional clutch torque calculator 270 is divided into a center differential input torque calculator 267, a clutch torque calculator 269, and a turning corrector 272.
It is also conceivable to change to a configuration consisting of

That is, the center differential input torque calculation unit 267 outputs the engine torque Te transmitted from the engine torque detection unit 264.
The torque converter torque ratio t sent from the torque converter torque ratio detector 266 and the transmission reduction ratio detector 276
From the transmission reduction ratio ρm sent from
The center differential input torque (transmission output torque) Ta is calculated by the following equation.

Ta = t · ρm · ρ ₁ · Te (3.2) where ρ ₁ is the final reduction ratio.

Note that this center differential input torque Ta and engine torque
The relationship with Te is proportional to each set shift. For example, if the torque converter torque ratio t is set to 1.5, it becomes as shown in FIG. However, in practice, this relationship greatly changes depending on the magnitude of the torque converter torque ratio t, and therefore, the torque converter torque ratio t is obtained from the speed ratio i, and based on this,
What is necessary is to seek the relationship between Ta and Te.

The clutch torque calculation unit 269 calculates the clutch torque Tc at which the front-rear drive distribution is equal to the static load distribution from the following equation.

Tc = [(Zs + Zr) / Zr · Wf / Wa-1] · Ta (3.3) where Zs is the number of teeth of the sun gear, Zr is the number of teeth of the ring gear, and Wf
Is the load shared by the front wheels, and Wa is the vehicle weight.

Then, the turning correction unit 272, by correcting the clutch torque Tc obtained in this manner by the rotation difference gain k ₂ described above, the engine torque proportional clutch torque Ta is obtained.

The center differential input torque calculation unit 267 and the clutch torque calculation unit 269 may be integrated to obtain the engine torque Te, the torque converter torque ratio t, and the transmission reduction ratio ρm according to the following equation.

Tc = [(Zs + Zr) / Zr · Wf / Wa−1] · t · ρm · ρ ₁ · T
e (3.4) Further, the switch 274a is turned on at a low vehicle speed (Vref <20km / h in this example) by a signal from the determination means 274, and the engine torque proportional clutch torque Ta is used as data. The output is enabled, but when the vehicle speed becomes higher (Vref ≧ 20 km / h), the output is turned off, and the output is stopped as data of the clutch torque Ta of the engine torque. This is because engine torque proportional control may cause a tight corner braking phenomenon when turning at a certain speed, or may eliminate other control speeds when slip tolerance is required. Thus, a condition is provided that the engine torque proportional control is performed only at the time of constant vehicle speed.

Next, the setting of the protection control clutch torque Tpc for protecting the clutch portion of the wet multi-plate clutch 28 will be described. The setting of the clutch torque Tpc is performed by the protection control unit 230.

That is, in the wet-type multi-plate clutch 28, in general, when the differential rotation between the clutch plates is large, there is a fear that damage such as seizure of the clutch facing and an increase in abrasion may be caused. Is larger, the damage is more likely to occur. On the other hand, in order to avoid such a state and to protect the clutch 28, it is conceivable to make the clutch free (disconnecting the connection between the clutch plates). If performed instantaneously, there is a fear that the attitude of the vehicle may change suddenly. In order to avoid any of these phenomena, the protection control unit 23
With 0, the protection control clutch torque Tpc is set.

The protection control unit 230 receives the front and rear wheel actual rotation speed difference Vcd calculated by the front and rear wheel actual rotation speed difference calculation unit 206, and determines the front and rear wheel actual rotation speed difference Vcd as a reference value (8.6 km / h in this example). 20), the protection control clutch torque Tpc is set in a pattern as shown in FIG. 20 when the state is larger than the reference time (1 second in this example).

That is, when the above-described detection condition is satisfied, the protection control clutch torque Tpc is first reduced for a short time (1 second in this example).
Is set to the upper limit only, and thereafter, it is gradually reduced to 0 (natural release). In this example, the decreasing Tpc and the time tt
Is as follows.

Tpc = 40-14tt (4.1) The time to set the upper limit and the clutch torque Tpc
The rate at which is gradually reduced to 0 (corresponding to the slope in FIG. 20) is
It is desirable to appropriately set the optimum value according to the characteristics of each vehicle.

If the above-described detection condition is not satisfied, the value of the protection control clutch torque Tpc is set to zero.

As described above, the differential corresponding clutch torque Tv, the longitudinal acceleration corresponding clutch torque Tx, the engine torque proportional clutch torque Ta,
Each clutch torque of protection control clutch torque Tpc is
The clutch torques Tv, Tx, Ta, and Tpc that are set for each control cycle that is repeated at an appropriate timing are sent to the maximum value selection unit 280.

In this maximum value selection unit 280, for each control cycle,
The largest one among the clutch torques Tv, Tx, Ta, and Tpc (this clutch torque is referred to as Tc) is selected. However, when the switch 258a or 274a is OFF, the clutch torque Tx
Alternatively, since Ta is not sent, the maximum value selecting section 280 selects the maximum value from the clutch torques sent.

The clutch torque Tc selected in this manner is sent to the torque-pressure converter 282, where the clutch control pressure Pc is set so as to obtain the set clutch torque Tc.

Here, the map (see the block 282 in FIG. 1)
Thus, the clutch control pressure Pc is obtained from the clutch torque Tc. However, since the clutch torque Tc and the clutch control pressure Pc are generally in a proportional relationship, the map is linear as shown in the figure.

Further, the clutch control pressure Pc thus changed is subjected to centrifugal pressure correction and preload correction by an adder / subtractor 284 as preload applying means.

The centrifugal pressure correction is performed by subtracting the centrifugal correction pressure Pv set by the centrifugal correction pressure setting unit 286 from the clutch control pressure Pc. It is obtained from the front wheel speed Vf calculated in 204a using a map as shown. This is because the piston chamber rotates in synchronization with the front wheel side shaft, so the centrifugal hydraulic pressure is
The centrifugal correction pressure Pv is set so as to be proportional to the square of the front wheel vehicle speed Vf.

The preload correction is a correction in which the initial engagement pressure (initial pressure) set by the initial engagement pressure setting unit (preload setting unit) 288 is added to the clutch control pressure Pc as Pi preload.

The purpose of this preload correction is to increase the control response while maintaining the contact state between the clutch plates of the clutch 28 in a marginal contact state (extremely slight contact state) where no drag torque is generated. is there. However, the clearance between the clutch plates of the clutch varies from product to product from the manufacturing stage due to a component error, an assembly error, and the like, and even the same product changes over time. In particular, since the return spring of the clutch plate is generally strong, errors and aging of each part greatly affect the clearance between the clutch plates. For this reason,
It is necessary to always maintain the state of contact between the clutch plates as close as possible while detecting the state of the clean lath between the clutch plates at an appropriate timing.

For this reason, the preload setting unit 288 sets the initial pressure Pi by trying (in this case, learning) an appropriate amount of preload at an appropriate time interval.

There are various methods for this preload learning (setting of the initial pressure Pi from the preload learning value). Here, three types of preload learning will be described.

First, the first preload learning method will be described. To perform the preload learning, the engine must be operated in a steady state (which can be understood from the fact that the oil temperature of the engine has reached a stable temperature state at a predetermined height), and Must be performed under such a condition that the above line pressure is obtained and the control for the other clutches 28 is not affected. For this reason, the conditions of the preload learning are set as follows, for example.

At least 30 minutes have passed since the ignition key was turned on.

The shift selector is selected from one of 1 (first speed), 2 (second speed), D (drive), and N (neutral). The reason why there is no range for P (parking) and R (reverse) is that, in this example, hydraulic pressures different from those in the case of 1, 2, D, and N are output in the case of P and R. .

Vref = 0 km / h (vehicle speed Vref is 0).

Tc ≦ 1kgfm [Clutch torque Tc is a small predetermined value (1kgf
m) The following].

When the above conditions are simultaneously satisfied, the preload learning is executed as follows.

First, as shown in FIG. 21 (a), a pressure greater than the urging pressure of the return spring of the multiple disc clutch 28 and smaller than the designed initial engagement pressure of the clutch 28 [for example, P = 0.4 kgf / cm ² ] equivalent duty for 2 seconds, and then, for example, at an increasing speed of 1.5% / s,
For example, sweep slowly until a duty corresponding to P = 3.0 kgf / cm ² .

Then, the pressure P applied to the hydraulic pistons 141 and 142 becomes 21st.
It changes as shown in FIG. That is, since the clutch plate is initially separated, if the duty rises gently, the hydraulic piston 28 moves accordingly, so that the pressure P also rises gently. When the 141 and 142 move, the clutch plates come into contact with each other, and the pressure P also receives the force of the return spring, so that the pressure P suddenly increases. Further, as the hydraulic pistons 141 and 142 move, the clutch plates come into strong contact and the clutch is completely connected.
This state can be understood from the fact that the increase in the pressure P becomes a limit.

Here, a value (difference) P ″ obtained by second-order differentiation of the detected pressure P with respect to time and a value (difference) P ′ obtained by performing first-order differentiation of the pressure P with time are sometimes calculated in a short cycle. 2
The time when the differential value P ″ becomes maximum is determined to be the time of the start of contact of the clutch plate, the pressure P at this time is determined to be the initial pressure, and the time when the one-stage differential value P ′ becomes maximum is determined. It is determined that the clutch plate is completely engaged.

Specifically, when learning starts and the pressure P increases, the maximum value of the second-order differential value P ″ and the pressure P at this time are stored. The value of the second-order differential value P ″ is It is calculated for each short control cycle and updated as appropriate.

When the first-order differential value P 'becomes maximum (that is,
When the clutch is completely engaged), the calculation of the second order differential value P ″ is terminated, and within the period up to this point, the second order differential value P ″
Is stored as the initial pressure Pi.

During execution of such preload learning, if any of the above conditions of preload learning is not satisfied,
Immediately, the preload learning is interrupted and the mode returns to the normal mode.

Further, the above-described preload learning is performed once, once the ignition key is turned on, and is not executed unless the ignition key is turned off and then turned on once.

Next, a second preload learning method performed by the preload setting unit 288 will be described.

This preload learning is also performed under such conditions that the engine is brought into a stable oil temperature state at a predetermined height, a constant line pressure is obtained, and further, control on other clutches 28 is not affected. Although it is necessary to perform this preload learning many times, it is necessary to slightly relax the preload learning conditions described above and set, for example, the following preload learning conditions.

′ It has been more than 10 minutes since the ignition key was turned on.

The shift selector is selected from one of 1 (first speed), 2 (second speed), D (drive), and N (neutral).

Vref = 0 km / h (vehicle speed Vref is 0).

Tc ≦ 1kgfm [Clutch torque Tc is a small predetermined value (1kgf
m) The following].

A predetermined time (for example, a suitable time of about 5 minutes or less) has elapsed from the previous trial.

When the above conditions are simultaneously satisfied, the preload learning is executed as follows.

First, a preset duty (duty) corresponding to the initial pressure Pi (= P ₁ ) is held for a predetermined time (for example, 2 seconds), and then P = P for a predetermined time (for example, 1 second).
8.8 kgf / cm to ² equivalent duty (almost 100% duty), to sweep.

As a result, the pressure P applied to the hydraulic pistons 141 and 142
Changes two types of patterns as shown by curves L1 and L2 in FIG.

That is, when the clutch is separated in the initial pressure P _1,
As indicated by the curve L1, at a certain point in time when the duty is swept, the clutch comes into contact and starts dragging, so that the hydraulic pistons 141 and 142 are shocked and the pressure P
Is almost 10 while vibrating after overshoot
Full engagement pressure (steady peak pressure) according to 0% duty
Calm down.

Then, when the pressure P overshoots, a peak value (maximum value) Pmax larger than a constant maximum pressure Pc (a known value, here, about 8.8 kgf / cm ² ) by a certain value or more is generated.

On the other hand, when the clutch is in contact with the initial pressure P ₁ and is in a drag state, the pressure P increases almost linearly as the duty is swept as shown by the curve L2, and at a certain point in time the smooth full engagement is achieved. Settles to pressure (steady maximum pressure) Pc.

From such characteristics, the peak value Pmax of the pressure P is stored, and the difference α (= Pmax−) between this value Pmax and the steady maximum pressure Pc is stored.
If Pc) is larger than the predetermined value α ₀ , the initial pressure P ₁
Then it can be determined that the clutch is disengaged.

Therefore, while appropriately increased or decreased from the initial value P ₁ of the starting pressure P, by repeating the trial as described above at appropriate time intervals (e.g., 5 minutes) can be set to detect the correct initial pressure Pi .

That is, this preload learning is desirably performed as many times as long as the above conditions are satisfied, and the initial learning value and the initial pressure Pi set at a certain time point (n-th learning stage) are set.
In general, the initial learning value is represented by PINTG (n)
And the initial pressure Pi as PINT (n). Therefore, the previous initial learning value can be expressed as PINTG (n-1), and the initial pressure can be expressed as PINT (n-1). In the nth learning stage, learning is performed using the previous initial pressure as PINT (n-1). Will be.

Then, by comparing the difference obtained by the predetermined duty sweep alpha and (= Pmax-Pc) and the threshold alpha _1, this initial learning value PINTG (n) and the initial pressure PINT
(N) is set as follows.

When α ≧ α ₀ , PINTG (n) = PINTG (n−1) + β PINT (n) = PINTG (n−1) + β = PINTG (n) When α <α ₀ , PINTG (n) = PINTG ( n-1) -β PINT (n ) = PINTG (n-1) That is, when the alpha ≧ alpha ₀ is the initial learning value PINTG
Regarding (n), the previous initial learning value PINTG (n
-1) is added to the value obtained by adding β (= 1-bit pressure), and the initial pressure PINT (n) is obtained by adding β (= 1-bit pressure) to the previous initial learning value PINTG (n-1). What was subtracted, that is, this initial learning value PINTG
Set to (n).

This means that when the alpha ≧ alpha _0, since it can be determined that the overshoot, the previous initial pressure PINT (n-1), the clutch 28 can be determined not to close until the last minute of contact. Therefore, this initial learning value PI
NTG (n) is obtained by adding β (= 1 pressure for one bit) to the previous initial learning value PINTG (n-1), and the current initial pressure PINT (n) is used as the previous initial learning value PINTG.
It is assumed that β (= 1 pressure for 1 bit) is added to (n−1).

Note that 1 bit is limited by the resolution of a hydraulic pressure sensor signal that detects the hydraulic pressure applied to the piston.
it = 0.05kgf / cm ² or set to a suitable value, such as 1bit = 0.1kgf / cm ^2.

On the other hand, when α <α ₀ , the initial learning value PINTG
Regarding (n), the previous initial learning value PINTG (n
-1) plus β (= 1 bit)
The initial pressure PINT (n) is set to the previous initial learning value PINTG (n-1).

This is because, when α <α ₀ , there is no overshoot, and it can be determined that the clutch 28 is in the last contact state or the excessive contact state with the previous initial pressure PINT (n−1). Therefore, this initial learning value PINTG
(N) is replaced by β for the previous initial learning value PINTG (n-1).
(= 1 bit), but the initial pressure PI
NT (n) is the previous initial learning value PINTG (n-1)
Set as is. This is because it is impossible to judge whether the clutch 28 is in the barely contact state or in the excessive contact state by the result of α <α ₀ alone, and there is a fear that chattering may occur. Instead of immediately using the learning result of this time for the initial pressure Pi, the previous learning value is used.

Thus, in excessive contact, at least 2
It is considered that the state of α <α ₀ continues for consecutive cycles, and the initial pressure Pi is decreased by one cycle, but decreases and approaches an appropriate value.

Note that if any of the above-described conditions of preload learning 〜 is not satisfied during execution of such preload learning, the preload learning is immediately interrupted and the mode returns to the normal mode.

Further, the above-described preload learning is performed under the condition of the above-described preload learning.
As long as is satisfied, the process is continued.

Next, a third preload learning method performed by the preload setting unit 288 will be described.

This preload learning is also set to execute the preload learning when the following preload learning conditions are simultaneously satisfied, similarly to the second preload learning.

′ It has been more than 10 minutes since the ignition key was turned on.

The shift selector is selected from one of 1 (first speed), 2 (second speed), D (drive), and N (neutral).

Vref = 0 km / h (vehicle speed Vref is 0).

Tc ≦ 1kgfm [Clutch torque Tc is a small predetermined value (1kgf
m) The following].

A predetermined time (for example, a suitable time of about 5 minutes or less) has elapsed from the previous trial.

When the above-mentioned conditions ('-) are simultaneously satisfied, the preload learning is executed as follows.

First, the duty is adjusted so that a pressure pattern as shown in FIG. 23 (a) is obtained. In other words, since the holding Introduction duty only 0% for a predetermined time (for example one second), and holds only this predetermined time by those duty initial initial pressure P ₁ equivalent (e.g., 2 seconds), followed by a predetermined P = 8.8kgf / c in time (for example, 1 second)
m ² equivalent duty (almost 100% duty)
And a duty equivalent to P = 8.8 kgf / cm ^{2 is} maintained for a predetermined time (for example, 2 seconds). This pattern is continuously repeated while appropriately changing the initial pressure Pi.

As a result, the pressure P applied to the hydraulic pistons 141 and 142
23 (b), as in the case of the second preload learning,
As shown by curves L1 and L2 in (c), two types of patterns are changed.

Then, from the time t ₀ at which the duty sweep is started (or the time t ₁ at which the pressure P starts to rise), until the steady maximum pressure Pc as shown by the straight line L0 (or a constant pressure value close to this) is reached. Between the straight line L0 and the area S1 and S2 of the portion (hatched in the figure) surrounded by the curve L1 or L2 depicting the change state of the pressure P, the area in the case of the curve L1 with overshoot S1 is clearly larger than area S2 in the case of curve L2 without overshoot.

Therefore, in the third preload learning, similarly to the second preload learning, the above trial is performed at an appropriate time interval (for example, 5 seconds).
(At minute intervals), an appropriate initial pressure Pi can be detected and set.

That is, this preload learning is performed as many times as long as the above condition is satisfied, and the initial learning value and the initial pressure Pi set at a certain time point (n-th learning stage) are set.
In the same manner as above, the initial learning value is generalized to PINTG (n) and the initial pressure Pi is generalized to PINT (n).

Therefore, the previous initial learning value is PINTG (n-
1) The initial pressure can be expressed as PINT (n-1), and at the n-th learning stage, the previous initial pressure learns with PINT (n-1).

Then, by comparing the area S and threshold value S ₀ obtained by a predetermined duty sweep, this initial learning value
PINTG (n) and initial pressure PINT (n) are set as follows.

When the _{S ≧ S 0, PINTG (n} ) = when PINTG (n-1) + β PINT (n) = PINTG (n-1) + β = PINTG (n) S of _{<S 0, PINTG (n)} = PINTG ( n-1) -β PINT (n ) = PINTG (n-1) words, in the case of S ≧ S ₀ corresponds to the case of alpha ≧ alpha ₀ of the second preload learning, in the case of S <S ₀ Α <2 for the second preload learning
This corresponds to the case of α ₀ .

That is, when S ≧ S ₀ , it can be determined that overshoot has occurred. Therefore, with the previous initial pressure PINT (n−1), it can be determined that the clutch 28 is not approaching the barely contact state. Therefore, this initial learning value PINTG
(N) is replaced by β for the previous initial learning value PINTG (n-1).
(= 1 bit pressure), and the current initial pressure PINT (n) is added to the previous initial learning value PINTG (n
-1) is added by β (= 1 pressure for one bit).

On the other hand, when S <S ₀ , there is no overshoot, and the clutch pressure is not increased at the previous initial pressure PINT (n−1).
28 can be determined to be in a barely contact state or an excessive contact state. Therefore, this initial learning value PINTG
(N) is replaced by β for the previous initial learning value PINTG (n-1).
(= 1 bit), but the initial pressure PI
NT (n) is the previous initial learning value PINTG (n-1)
Set as is. The reason for this is also the aforementioned α <
As in the case of alpha _0, it is only the result of S <S _0, the clutch 2
Since it is not possible to judge whether 8 is in the barely contact state or in the excessive contact state, there is a fear that chattering may occur,
In order to avoid this, the previous learning value is used instead of immediately using the current learning result as the initial pressure Pi.

Thus, in excessive contact, at least 2
Is expected to continue the state of S <S ₀ by consecutive cycles, the initial pressure Pi is while delayed by one cycle, is reduced, so that approaches the appropriate one.

Note that if any of the above-described conditions of preload learning 〜 is not satisfied during execution of such preload learning, the preload learning is immediately interrupted and the mode returns to the normal mode.

Further, the above-described preload learning is performed under the condition of the above-described preload learning.
As long as is satisfied, the process is continued.

In the third preload learning, the straight line L0 and the curve L1 or
Instead of the area S (S1, S2) of the part surrounded by L2, the area S '(S1', S2 ') of the part surrounded by a straight line L3 indicating a constant pressure about the initial pressure and the curve L1 or L2. ) Can be considered.

In this case, the calculation of the area S ′ is started at the time t ₀ at which the duty sweep is started (or at the time t ₁ at which the pressure P starts to increase), and the calculation of the area S ′ is completed by a straight line L0. It is a point in time when a steady maximum pressure Pc (or a constant pressure value close to this) is reached. When the determination reference value is S ₀ ′, it can be determined that there is an overshoot when S ′ ≧ S ₀ ′, and it can be determined that there is no overshoot when S ′ <S ₀ ′.

As described above, the clutch control pressure which is the effective hydraulic pressure
The centrifugal pressure is corrected by subtracting the centrifugal correction pressure Pv from Pc, and the control pressure Pcd (= Pc) of the oil chamber supply level corrected by adding the initial pressure (preload) Pi and adding the preload is corrected.
−Pv + Pi) is taken into the peak hold filter 290.

This peak hold filter 290 is a type of limiter that prevents an excessive sudden change in hydraulic pressure so that control does not hunt due to a sudden change in hydraulic pressure. / cm ² /
s), and a slightly lower limit speed (for example, 15.7 kg / cm ² / s) is set for falling of the hydraulic pressure. Then, when a control pressure Pcd at which the speed of the oil pressure change exceeds such a limit is sent, the control pressure is kept at a value corresponding to this limit.

Further, the control pressure Pcd ′ that has passed through the filter 290 is sent to the duty setting unit 295 via the switches 292a and 294a.

The switch 292a is turned off by the signal from the determination means 292 when the ABS control (anti-lock brake control) is being performed (if it is in the ON state), and is turned on when the ABS control is not being performed. . That is, a signal of the control pressure Pcd 'is sent on condition that the ABS control is not performed. This is because it is necessary to reliably operate the ABS during the ABS control, and at this time, controlling the torque distribution state of the front and rear wheels is not preferable because it interferes with the ABS control.

The switch 294a is a control switch for protecting the duty solenoid valve and the clutch plate in accordance with a signal from the determination means 294, and sets the duty to 0 at low speed and when the set clutch torque Tc is small. It is something to try. The low-speed conditions include, for example, Vref ≦ 5 km / h, clutch torque Tc
Can be specified as, for example, Tc ≦ 1 kgfm. Then, when these two conditions are satisfied, the switch 294a is turned off, and the signal of the control pressure Pcd 'is not sent.

The duty setting unit 295 includes the pressure feedback correction unit 2
96 and a pressure-duty converter 298.

The pressure feedback correction unit 296 receives the detection information from the pressure sensor 304 as a fluid pressure detection unit that detects the actual pressure acting on the piston, and corrects the signal of the control pressure Pcd ′. This is for correcting the characteristics of the circuit. The signal sent from the pressure sensor 304 to the pressure feedback correction unit 296 is filtered by the filter 306 to remove noise components due to disturbance or the like.

The pressure-duty converter 298 sets (Duty) corresponding to the control pressure P feedback-corrected by the pressure feedback corrector 296. The map shown in the block of the clutch pressure-duty converter 298 in FIG. The duty increases linearly with respect to the pressure P from the preload state to the maximum pressure state. From such a correspondence, a duty corresponding to the control pressure P is set.

In the hydraulic circuit 300 functioning as the control execution unit, the duty solenoid 302 operates according to the duty set in this way, and the differential limiting clutch 28 of the center differential operates.
Is controlled.

On the other hand, in parallel with such center differential control, the state of torque distribution to the front and rear wheels is displayed in a meter cluster of an instrument panel in a driver's seat.

That is, as shown in FIG. 1 and FIG. 24, a torque distribution display section for graphically displaying (or metering) the state of torque distribution to the front wheels (or rear wheels) in the meter cluster.
312 is provided, and the torque distribution state is displayed by the torque estimation means 310 according to the magnitude of the distribution torque estimated.

The reason for estimating the torque distribution state by the torque estimating means 310 is that it is difficult to actually measure the torque distribution state.

The torque estimating means 310 determines whether the multi-plate clutch 28 has a front wheel output torque (or rear wheel output torque) when a rotational speed difference occurs between the front and rear wheels, and a rotational speed difference between the front and rear wheels. Calculating means 310a for calculating the front wheel output torque (or rear wheel output torque), and the smaller front wheel output torque (or rear wheel output torque) of the front wheel output torque (or rear wheel output torque) in each of these cases. , And these sections 310a and 310b estimate the torque distribution state as described below.

That is, when estimating the torque distribution, the following two cases can be considered. One is a case where it is assumed that the tire and the road surface do not slip and are in the same state as the engagement of the gears, and the center differential always slips. The other one is
In practice, there is always a slip between the tire and the road surface, so the center differential may not slip.

Therefore, the torque distribution in each of these cases and when the state is switched will be considered.

First, as a prerequisite, when the differential limitation is not performed as in the four-wheel drive system, the rear wheel is mainly set (the torque ratio between the front wheel and the rear wheel is, for example, 32:68). The clutch 28 always transmits torque from the rear wheel side to the front wheel side, and is set as follows for simplicity.

ρf / rf <ρr · ρt / rr (5.1) where ρf: front differential ratio ρr: rear differential ratio ρt: transfer ratio rf: front tire radius rr: rear tire radius If the clutch does not slip, Therefore, the front-wheel torque Tf and the rear-wheel torque Tr are as follows.

Tf = Wf / Wa · ｛Tm + kWr · rf / ρ · (rfρrρt / rrρt−
1)｝ (5.2) Tr = Wr / WaWTm-kWfkWrr / ρ ・ (rfrρt / rrρt-
1)｝ (5.3) where Wf: front wheel sharing weight Wr: rear wheel sharing weight Wa: vehicle weight (= Wf + Wr) Tm: mission output torque (= center differential input torque) k: slip ratio coefficient ρ: final deceleration Ratio [= (ρf + ρr · ρt) / 2] When the clutch slips, the front wheel torque Tf ′ and the rear wheel torque Tr ′ are as follows.

Tf ′ = (Tm−Tc) · a / (a + b) + Tc (5.4) Tr ′ = (Tm−Tc) · b / (a + b) (5.5) where Tc: clutch transmission torque capacity a : Number of sun gear teeth b: Number of ring gear teeth When the above-mentioned clutch slips, it means that the clutch allows a front-rear torque difference caused by a weight distribution or a differential ratio difference. Now, Kucci is considering the case of transmitting torque from the rear wheel side to the front wheel side, so that regarding the front wheel torque Tf, Tf ', the smaller one of Tf, Tf' may be considered as the front wheel torque value. it can.

That is, if Tf <Tf ′, the clutch is in the locked state, and the front wheel torque distribution ratio m is: m = Tf / (Tf + Tr) (5.6) If Tf> Tf ′, the clutch is in the slip state and the front wheel torque is The distribution ratio m can be estimated as m = Tf '/ (Tf' + Tr ') (5.7).

FIG. 25 shows the front wheel torque distribution ratio m with respect to the center differential input torque Tm, and the characteristic of the input torque corresponding front wheel torque distribution ratio is a straight line with direct connection when the clutch is in the locked state. When the clutch is in the free state, the clutch has a curved shape according to the magnitude of the control pressure P. In addition,
In the figure, when the pressure P is 2 kgf / cm ² (P = 2) and when the pressure P is 8 kgf / cm ²
For m ² illustrates a (P = 8).

Then, in the characteristic graph, a characteristic line with a smaller m among the straight line attached directly and the curve in the case of a certain control pressure P is adopted.

For example, when P is 2 kgf / cm ^{2, in} a region where the torque Te is smaller than Te ₁ , the straight line attached directly is below the curve of P = 2. The torque distribution ratio is m. The torque Te is a region larger than Te _1, since towards the P = 2 the curve is below the direct, the front wheel torque distribution ratio m in accordance with the curve of P = 2.

On the other hand, when P is 8 kgf / cm ^{2, in} the region shown in the graph, the directly connected straight line is always lower, and thus the front wheel torque distribution ratio m according to the directly connected condition is m.

When the front wheel torque distribution ratio m is set in this way, a signal corresponding to the set value is output to the torque distribution display section 312.
And the torque distribution display section 312 displays the state of torque distribution to the front wheels. In this example, since the torque distribution to the front wheels is about 32% to 50%,
A scale corresponding to this is attached to the torque distribution display section 312,
By turning on the lamp or moving the pointer to the corresponding scale, the display is easily understood.

The display of the torque distribution state may be the torque distribution state to the rear wheels, or the distribution state to the front and rear wheels may be displayed in an analog manner using a graph or the like.

Since the differential-adjustable front and rear wheel torque distribution control device is configured as described above, the differential adjustment is performed as follows.

First, as shown in FIG. 26, the flow of the entire operation of the drive system is as follows. First, each control element is initially set (step a1), learning of the neutral position of the steering angle (step a2), and preloading of the clutch are performed. The learning (step a3) is performed, and then the front and rear wheel driving force distribution control is performed while controlling the clutch 28 according to the set duty (step a4), and further, the rear differential is controlled (step a5).

Then, in steps a7 to a11, slip control, trace control, torque selection, retard control calculation, SCI (Serias C
Engine output control (traction control) such as communication interface communication control is performed, the torque distribution display lamp is turned on (step a12), and failure diagnosis (failure diagnosis) is performed in step a13. Step a14
Then, it is determined whether a predetermined time (15 msec) has elapsed, and
When a predetermined time (15 msec) has elapsed, a runaway check is performed by a watchdog (step a15), and the process returns to step a2 to repeat a series of controls of steps a2 to a13.

That is, the above-described front and rear wheel drive force distribution control, rear differential control, and engine output control are performed at a predetermined cycle (15 msec).

Among them, the front and rear wheel driving force distribution control will be described with reference to the flowchart in FIG.

As shown in FIG. 27, first, wheel speeds FR, FL, RR, RL, steering angles θ ₁ , θ ₂ , θn, lateral acceleration Gy, longitudinal acceleration Gx, throttle opening θth, engine speed Ne, transmission speed Number N
t, data of the selected shift stage and the like are detected and fetched (step b1). From these data, the front wheel speed Vf,
Rear wheel speed Vr, driver required vehicle speed Vref, driver required steering angle δre
f and the like are calculated (step b2).

Then, the driver request vehicle speed Vref, seeking the ideal rotational speed difference ΔVhc the front and rear wheels in accordance with the map from the driver's demand steering angle Derutaref (step b3), by setting the lateral G gain k ₁ according to the map from the lateral acceleration Gy ( step b4), setting the rotational difference gain k ₂ according to the map from the ideal rotational speed difference DerutaVhc (step b5).

Further, in steps b6 to b9, the actual rotational speed difference Δ
Vc, ideal rotational speed difference DerutaVhc, calculates a differential corresponding clutch torque Tv from the lateral G gain k ₁ (in this example determined by converting from the map), and the longitudinal acceleration Gx, lateral G gain from k ₁ longitudinal acceleration corresponding clutch torque Calculate Tx (convert from map)
Throttle opening [theta] th, engine speed Ne, transmission rpm Nt, selected shift stage, and calculates a rotation differential gain k ₂ from the engine torque proportional clutch torque Ta (calculated from the map), in response to the signal of the ideal rotational speed difference ΔVhc Set the protection control clutch torque Tpc.

Then, in step b10, each of these clutch torques T
The maximum value is calculated from v, Tx, Ta, and Tpc as the set clutch torque Tc.

Further, in step b11, the thus determined clutch torque Tc is converted from a map into a clutch engagement pressure Pc.

Subsequently, a preload correction (adding a preload Pi) and a centrifugal pressure correction (decreasing the centrifugal pressure Pv) are applied to this pressure Pc (step
b12), the center differential control pressure Pcd is obtained.

Further, a peak hold filter is passed so that an excessive change in the pressure P can be suppressed (step b1).
3).

Then, it is determined whether the ABS is in operation (step b14) or whether the protection condition of the solenoid valve (Vref ≦ 5km / h, Tc ≦ kgfm) is satisfied (step b15).
If any of these conditions is satisfied, the center differential control pressure Pcd is reset to 0 in step b19.

The center differential control pressure Pcd set in this way is
At step b16, pressure feedback correction is performed.
That is, the difference ΔP between the value of Pcd and the actually measured value of the pressure sensor is calculated, and the proportional correction obtained from the product of the integral correction pressure Pi obtained from the product of the integrated correction gain ki and ΔP (i) and the proportional correction gain kpΔP. The above-described center differential control pressure Pcd is corrected by the corrected pressure Pp to obtain the pressure P.

Further, in step b17, the pressure P is converted into a corresponding duty, and the center differential control, that is, the control of the differential limiting clutch is performed.

The above-described calculation of the differential corresponding clutch torque Tv is performed as shown in FIG.

First, a difference Δ obtained by subtracting the front wheel speed Vf from the rear wheel speed Vr.
Vcd (= Vr-Vf) is calculated (step c1), and from this difference (the actual rotation speed difference between the front and rear wheels) ΔVcd, the ideal rotation speed difference between the front and rear wheels obtained as described above (see step b3). The difference ΔVc (= ΔVcd−ΔVhc) is obtained by subtracting ΔVhc (step c2).

Then, in step c3, it is determined whether or not the above-described ideal rotation speed difference ΔVhc of the front and rear wheels is equal to or greater than 0.

In step c4, the clutch torque Tv 'is set from ΔVc using a map [see FIG. 13 (a)].

Specifically, if ΔVcd ≧ ΔVhc, the clutch torque
Tv ′ = a × (ΔVcd−ΔVhc) = a × ΔVc is set so that Tv ′ increases in proportion to the magnitude of the difference ΔVc (ΔVcd−ΔVhc) (where a is a proportional constant).

If ΔVhc>ΔVcd> 0, the clutch torque T
By setting v 'to 0, a so-called dead zone area is set.

Further, if 0 ≧ ΔVcd, Tv ′ = − a × such that the clutch torque Tv ′ increases in proportion to the magnitude of ΔVcd.
ΔVcd = −a × (ΔVc + ΔVhc) (where a
Is a proportional constant).

When ΔVhc = 0, there is no dead zone region of ΔVhc>ΔVcd> 0.

In step c5, the clutch torque Tv 'is set from ΔVc using a map [see FIG. 13 (b)].

Specifically, if ΔVcd ≧ 0, the clutch torque T
Tv ′ = a × ΔVcd = a × (ΔVc + ΔVhc) is set so that v ′ increases in proportion to the magnitude of ΔVcd (where a is a proportional constant).

If 0>Δcd> ΔVhc, the clutch torque T
By setting v 'to 0, a so-called dead zone area is set.

Further, if ΔVhc ≧ ΔVcd, the clutch torque Tv ′
Is set in such a manner that Tv ′ = − a × (ΔVcd−ΔVhc) = − a × ΔVc (where a is a proportionality constant) so that the value increases in proportion to the magnitude of ΔVc (ΔVcd−ΔVhc).

Thus, at step c4, c5, the obtained differential corresponding clutch torque Tv 'is the lateral acceleration corresponding correction by being integrated with the lateral G gain k ₁ by the correction section 246 (step c6),
The clutch torque Tv corresponding to the differential is obtained.

By setting such differential-compatible clutch torque Tv,
The magnitude of the clutch torque Tv is set appropriately without waste,
It is now possible to properly adjust the attitude control of the vehicle body while setting the rear wheel as the drive base to slip from the rear wheel as appropriate, so that the vehicle can behave in accordance with the driver's will when turning. It becomes.

In other words, when the direction SIG (δh) of the sensor corresponding steering angle δh is not equal to the direction SIG (Gy) of the lateral acceleration data Gy, the driver's requested steering angle is set to 0. Even if the steering position of the steering wheel is different from the actual steering angle (turning state) of the steering wheel when operating the steering wheel or the like, inappropriate data is no longer adopted, contributing to improved control performance. I do.

Furthermore, since the wheel speed data of the second largest magnitude from among the rotation speed data signals FL, FR.RL, RR is adopted as the driver requested vehicle speed Vref, the reliability of the data is ensured.

When the ideal rotation speed difference ΔVhc is set at a low vehicle speed, the difference in the orbit radii of the front and rear wheels at the time of turning (so-called inner wheel difference) has a large effect, and the rotation speed Vr of the rear wheel is smaller than the rotation speed Vf of the front wheel. However, as the vehicle speed increases, the rotational speed Vr of the rear wheel is set to be higher than the rotational speed Vf of the front wheel. Therefore, at high speeds, the rear wheels easily slip, and the responsiveness of the posture of the vehicle body required at higher speeds is ensured. As for the steering angle, the larger the steering angle, the larger the difference in rotation required for the front and rear wheels, and this is appropriately allowed, and there is an advantage that the tight corner braking phenomenon can be avoided.

The above-described calculation of the clutch torque Tx corresponding to the longitudinal acceleration is performed as shown in FIG.

First, the clutch torque Tx 'corresponding to the longitudinal acceleration is read from the map (FIG. 15) based on the detection data Gx from the longitudinal acceleration sensor 36 (step d1).

Then, by performing the lateral acceleration correction by multiplying the lateral G gain k ₁ in the longitudinal acceleration corresponding clutch torque Tx '(step d2), obtaining a longitudinal acceleration corresponding clutch torque Tx.

Further, in step d2, it is determined whether or not the front wheel speed Vf is higher than the vehicle speed Vref.When the front wheel speed Vf is higher than the vehicle speed Vref through the switch 258a, that is, when the front wheels are slipping (During front slip)
The above-mentioned clutch torque Tx corresponding to the longitudinal acceleration is directly used as control data, and the front wheel speed Vf is changed to the vehicle speed Vr.
If it is not larger than ef, that is, if the front wheels are not slipping, the longitudinal acceleration corresponding clutch torque Tx is set to 0 (step d4).

As a result, during acceleration such as during front slip,
While the torque distribution is the same as that of a direct-coupled 4WD, the higher torque is distributed to the base distribution ratio (toward the rear wheel, preventing a strong under-force and enabling smooth turning.

The calculation of the engine torque proportional clutch torque Ta is performed as shown in FIG.

First, the engine torque detector 264 reads the engine torque Te at that time from the throttle opening data θth and the engine speed data Ne through an engine torque map as shown in FIG. 12 (step e1).

Next, the engine torque proportional torque setting unit 268 reads the engine torque proportional torque Ta 'from the engine torque Te through a map (step e2).

Further, the torque converter torque ratio detecting unit 266 obtains the transmission torque ratio t at that time from the engine speed data Ne and the transmission speed data t through a transmission torque ratio map as shown in FIG. 13 (step e3). .

Then, the engine torque proportional clutch torque calculation unit 27
0, the thus obtained engine torque proportional torque Ta ′, the torque converter torque ratio t, and the transmission reduction ratio ρm,
From the final reduction ratio ρ ₁ and the rotation difference gain k ₂ obtained by the rotation difference gain setting unit 275, a center differential input torque (transmission output torque) Ta (= t · ρm · ρ ₁ · Te) is calculated (step e4). ).

Further, at step e5, when the vehicle speed is low (in this example, Vref <
20 km / h), and if the vehicle speed is low, the above-described engine torque proportional clutch torque Ta is output as it is as data, but the vehicle speed increases further (Vref
≧ 20 km / h), and 0 is set as the engine torque proportional clutch torque Ta (step e6), and this is output as control data.

With such an engine torque proportional clutch torque Ta, when starting or when suddenly accelerating from a low speed, etc., a direct connection 4WD state is established, and high torque can be transmitted to the road surface. In this case, tire slip is prevented, driving performance is improved, and driving system durability is also improved.

Further, the above-mentioned calculation of the protection control clutch torque Tpc is performed as shown in FIG.

First, in step f1, it is determined whether the flag FLG is "1". This flag FLG is a control flag that is set to 1 when the protection control is executed, and is set to 0 when the entire control is started.

Therefore, at the start of the control, the process proceeds to step f2, where the actual rotational speed difference Vcd between the front and rear wheels is a reference value (8.6 km / h in this example).
It is determined whether or not this is the case.

If the front and rear wheel actual rotation speed difference Vcd is not equal to or greater than the reference value (8.6 km / h), the process proceeds to step f9. If the timer has been counted, the counting is terminated and the timer is cleared. Then, at step f12, the value of the protection control clutch torque Tpc is set to 0, and at step f14, the flag FLG
Is set to 0.

On the other hand, if it is determined in step f2 that the front / rear wheel actual rotational speed difference Vcd is equal to or greater than the reference value (8.6 km / h), step f3
Then, it is determined whether or not the timer count has been started. If the timer count has not been started, the process proceeds to step f4, and the timer count is started.

When the timer count starts, the step
At f5, it is determined whether or not the timer value is equal to or longer than the reference time (1 sec). If the timer value does not reach the reference time or longer, the process proceeds to step f12, where the value of the protection control clutch torque Tpc is set to 0, At step f14, the flag FLG is set to 0.

If the actual rotational speed difference Vcd between the front and rear wheels is equal to or more than the reference value (8.6 km / h) for several control cycles, the timer count is continued during this time. It is possible to determine that the time has reached the reference time or more. At this time, the process proceeds to step f6.

In step f6, it is determined whether the timer value is equal to or longer than the reference time (2 sec). If the timer value does not reach the reference time or longer, the process proceeds to step f10, in which the value of the protection control clutch torque Tpc is set to 40. I do.

Then, in step f13, the flag FLG is set to 1, and the process proceeds to step f8, where it is determined whether Tpc is 0 or more. When the process proceeds from step f10 to step f8, the timer count is continued because Tpc is naturally 0 or more.

Then, the state of Tpc = 40 continues, and the timer value becomes 2
If sec or more, the process proceeds from step f6 to step f7, and the value of the protection control clutch torque Tpc is gradually reduced in a relationship of Tpc = 40−14 × (timer value−2).

In this way, after several control cycles, when the protection control clutch torque Tpc does not become 0 or more, the process proceeds from step f8 to step f11, the timer count is terminated, the timer is cleared, and At f12,
The value of the protection control clutch torque Tpc is set to 0, and the flag FLG is set to 0 in step f14.

As a result, the front-rear wheel actual rotational speed difference Vcd becomes the reference value (8.6
If the condition required for clutch protection is satisfied that the state of km / h or more continues for the reference time (1 second) or more, the characteristics shown in FIG. 20 will be obtained, that is, first, only for a short time (1 second in this example) The clutch torque Tpc for protection control is set so as to be set to the upper limit value and thereafter gradually decreased to 0 (natural release).
Is set.

This clutch torque Tpc for protection control protects the clutch plate, which contributes to the improvement of the durability of the device and the effect of preventing the spin of the vehicle.

Here, the above-described preload correction will be described with reference to FIGS.

First, in the first preload learning method, as shown in FIG. 32, in steps g1 to g4, it is determined whether 30 minutes or more have elapsed since the ignition key was turned on, and whether the shift selector is 1 (1 Speed), 2 (second speed), D (drive), N
(Neutral), whether the vehicle speed Vref is 0 km / h (stop state), whether the clutch torque set value Tc is a small predetermined value (1k
gfm) or less.

Then, if all of these conditions are satisfied, the process proceeds to step g5, and if any of these conditions is not satisfied, the learning control is not performed.

Proceeding to step g5, it is determined whether or not the preload learning has been performed since the ignition key was turned on.If the preload learning has already been performed, the learning control is not performed, and if the preload learning has not been performed, Proceed to step g6.

In step g6, the hydraulic pressure is raised, the maximum value (MAX) of the two-stage differential value of the hydraulic pressure is detected, and the hydraulic pressure P at that time is stored.

That is, first, as shown in FIG.
= 0.4 kgf / cm ² for 2 seconds, and then, for example, at an increasing speed of 1.5% / s, for example, P =
3.0kgf / cm up to ² equivalent of duty, to slowly sweep.

On the other hand, from the pressure P applied to the hydraulic pistons 141 and 142, which changes as shown in FIG. Is stored.

Then, the stored pressure P is set to the initial pressure.

Specifically, when learning starts and the pressure P increases, the maximum value of the second-order differential value P ″ and the pressure P at this time are stored, and the value of the second-order differential value P ″ is stored. Is calculated for each control cycle and is updated as appropriate. When the first-order differential value P ′ becomes maximum (that is, when the clutch is completely engaged), the calculation of the second-order differential value P ″ is terminated. Is stored as the initial pressure Pi when the maximum value of the second-order differential value P ″ is obtained.

Then, during execution of such preload learning, if any of the above conditions for preload learning is no longer satisfied, immediately stop preload learning and return to the normal mode.
Once this preload learning is performed once the ignition key is turned on, it is not executed unless the ignition key is turned off and then on once.

In addition, in the second preload learning method, as shown in FIG. 33, in steps h1 to h5, it is determined whether or not 10 minutes or more have elapsed since the ignition key was turned on.
Whether a predetermined time (for example, about 5 minutes or an appropriate shorter time) has elapsed from the previous trial, whether the shift selector is 1 (first speed), 2 (second speed), D (drive), N (neutral) Is determined, whether Vref = 0 km / h, and whether Tc ≦ 1 kgfm. Then, when all of these conditions are satisfied, the process proceeds to step h6. If any of these conditions is not satisfied, the learning control is not performed.

In step h6, the hydraulic pressure is raised and an overshoot value of the hydraulic pressure is detected.

In other words, start-up of oil pressure, only holds preset initial initial pressure P ₁ corresponding duty (duty) for a predetermined time (e.g., 2 seconds), then a predetermined time (e.g., 1
P = 8.8kgf / cm ² in duty (for almost 100%)
).

Then, the hydraulic pistons 141 and 142 which change correspondingly
The overshoot value α of the pressure P applied to is detected.

Further, in the next step h7, it is determined whether or not this α is larger than a threshold value.

That is, the peak value (maximum value) Pmax of the pressure P is detected,
The difference (Pmax−Pc) between the maximum value Pmax and the steady maximum pressure Pc (here, about 8.8 kgf / cm ² ) is defined as an overshoot value α.
If this alpha is larger than the threshold (alpha _0), there is an overshoot, thus, it can be determined that the initial initial pressure P ₁ in the clutch 28 is separated, is not greater than the alpha threshold, overshoot never been, i.e., the initial initial pressure P ₁ in the clutch 28 can be determined that the marginal contact or excessive contact.

If α is larger than the threshold value, the process proceeds to step h8, where PINTG (n) = PINTG (n−1) + β PINT (n) = PINTG (n−1) + β = PINTG (n) That is, the initial learning value PINTG (n) is set to a value obtained by adding β (= 1-bit pressure) to the previous initial learning value PINTG (n−1), and the initial pressure PINT
(N) is the previous initial learning value PINTG (n-
1) plus β (= pressure for 1 bit), that is,
The current initial learning value PINTG (n) is set.

On the other hand, if α is not larger than the threshold, the process proceeds to step h9, where PINTG (n) = PINTG (n−1) −β PINT (n) = PINTG (n−1) That is, the initial learning value PINTG (n) about,
Β (= 1 bit) in the previous initial learning value PINTG (n-1)
Minutes), but the initial pressure PINT
(N) is the previous initial learning value PINTG (n-
Set to 1).

Note that if any of the above-described conditions of preload learning 〜 is not satisfied during execution of such preload learning, the preload learning is immediately interrupted and the mode returns to the normal mode.

Further, the above-described preload learning is performed under the condition of the above-described preload learning.
As long as is satisfied, the process is continued.

In the third preload learning method, as shown in FIG. 34, it is determined whether the conditions are the same as those in the third preload learning. That is, in steps h1 to h5, it is determined whether or not 10 minutes or more have elapsed since the ignition key was turned on, and a predetermined time (for example, an appropriate time of about 5 minutes or less) has elapsed since the previous trial. Whether or not
Whether the shift selector is selected from among 1 (first speed), 2 (second speed), D (drive), N (neutral), whether Vref = 0 km / h, Tc ≦ 1
Whether it is kgfm or not is determined respectively.

When all of these conditions are satisfied, the process proceeds to step h16. If any of these conditions is not satisfied, the learning control is not performed.

In step h16, the hydraulic pressure is raised, and the difference between the predetermined pressure and the hydraulic pressure value is integrated.

In other words, start-up of oil pressure, only holds preset initial initial pressure P ₁ corresponding duty (duty) for a predetermined time (e.g., 2 seconds), then a predetermined time (e.g., 1
P = 8.8kgf / cm ² in duty (for almost 100%)
).

Then, the hydraulic pistons 141 and 142 which change correspondingly
The difference between the pressure P applied to the pressure and a predetermined pressure (a pressure close to the maximum pressure) is integrated. That is, as shown in FIGS. 23 (b) and (c), from the time t ₀ at which the duty sweep is started (or the time t ₁ at which the pressure P starts increasing), the steady maximum pressure as shown by the straight line L0 Until Pc (or a constant pressure value close to Pc) is reached, the area S (shaded in the figure) of a portion surrounded by the straight line L0 and the curve L1 or L2 depicting the change state of the pressure P ( S
1, S2) is calculated.

Further, in the next step h17, determines whether the calculated area S is larger than the threshold S _0. In other words, towards the area S1 in the case of a overshoot curve L1 is, since obviously larger than the area S2 of the case of overshoot of curved lines L2, by comparing the area S with the threshold S _0, overshoot Is determined.

Therefore, if the area S is greater than the threshold S _0, step
Proceeding to h8, PINTG (n) = PINTG (n-1) + β PINT (n) = PINTG (n-1) + β = PINTG (n) That is, for the initial learning value PINTG (n),
Β (= 1 bit) in the previous initial learning value PINTG (n-1)
Min pressure) and set the initial pressure PINT
(N) is the previous initial learning value PINTG (n-
1) plus β (= pressure for 1 bit), that is,
The current initial learning value PINTG (n) is set.

On the other hand, it is not greater than the area S is the threshold value S _0, the process proceeds to step h9, PINTG (n) = PINTG (n-1) -β PINT (n) = PINTG (n-1) words, the initial learning value PINTG For (n),
Β (= 1 bit) in the previous initial learning value PINTG (n-1)
Minutes), but the initial pressure PINT
(N) is the previous initial learning value PINTG (n-
Set to 1).

If any of the preload learning conditions 〜 to 満 た is not satisfied even during the execution of the third preload learning, the preload learning is immediately interrupted to return to the normal mode.

Also, in this case, the process is continued as long as the preload learning condition 'is satisfied.

By such first to third preload learning,
Appropriate initial pressure Pi can be set, which greatly contributes to improvement of control response.

In particular, in the first preload learning, there is an advantage that the initial pressure Pi can be set by one learning, which is extremely simple.

Further, in the second and third preload learning, the initial pressure Pi is set by several learnings, but there is an advantage that setting accuracy is high and a response improving effect is great.

In particular, in the determination based on the integral value (area), it can be relatively appropriately determined whether the initial pressure Pi is appropriate,
Initial pressure Pi without relying heavily on the capabilities of the pressure sensor
Can be set.

Further, the control performed through the switch 294a protects the duty solenoid valve and the clutch plate, and contributes to the improvement of the reliability and durability of the device.

Further, a torque distribution display section 312 is provided in the meter cluster.
Is provided, and the torque distribution state to the front wheels (or the rear wheels) is graphically displayed (or displayed by a meter), so that the driver can drive while recognizing the torque distribution state of the vehicle, and can be used effectively for driving, while driving Can be made more enjoyable, and the merchantability is greatly improved.

Furthermore, the result of the torque distribution estimation performed at this time may be used as feedback for control of each unit.

[Effects of the Invention] As described above in detail, according to the differentially adjusted front and rear wheel torque distribution control device of the present invention, the front and rear wheels are adjusted by adjusting the differential state between the front and rear wheels. In a front-rear wheel differential-adjustable four-wheel drive vehicle that controls the distribution of torque to the wheels, the differential between the front and rear wheels is limited by applying a frictional force between the front and rear wheels. A front-rear wheel differential adjusting mechanism for adjusting the dynamic state, a front-rear acceleration detecting means for detecting the front-rear acceleration of the vehicle, and a front-wheel sharing load or a rear-wheel sharing load for the magnitude of the front-rear acceleration detected by the front-rear acceleration detecting means. Sharing load calculating means for calculating, output torque calculating means for calculating a minimum engine output torque required to obtain the magnitude of the longitudinal acceleration detected by the longitudinal acceleration detecting means, and calculation by the sharing load calculating means. Load and engine output Necessary due to the configuration in which the shared load torque calculating means for calculating the shared load torque from the output torque calculated by the force torque calculating means and the control means for adjusting the differential limiting amount based on the shared load torque are provided. The torque for maintaining a high acceleration is distributed in the same manner as in the direct-coupled four-wheel drive, and more torque can be distributed to the base side, so that strong understeer is prevented, and the vehicle can easily turn smoothly.

In a region where the differential limiting amount set based on the shared load torque decreases as the longitudinal acceleration increases, the control unit corrects the differential limiting amount so as to suppress the reduction. Providing the means improves control safety.

[Brief description of the drawings]

1 to 34 show a differential-adjustment type front and rear wheel torque distribution control device as one embodiment of the present invention. FIG. 1 is a block diagram showing a configuration of a main part thereof, and FIG. FIG. 3 is a sectional view showing a main part of the drive torque transmitting system, FIG. 4 is a sectional view of a main part of the front and rear wheel torque distribution mechanism, and FIG. 5 is a schematic diagram of the hydraulic supply system. FIG. 6 is a circuit diagram of a main part of the hydraulic pressure supply system, FIG. 7 is a diagram showing characteristics of the hydraulic pressure setting duty, FIG. 8 is a block diagram showing details of the total steering angle data detecting means, 9 is a block diagram showing details of the vehicle speed detecting means, FIG. 10 is a diagram showing the map for setting the ideal rotational speed difference, FIG. 11 is a diagram showing the map for setting the lateral acceleration gain, and FIG. (A), (b) is a plan view schematically showing a wheel state for explaining the ideal rotational speed difference. Figure 13 (a), (b) is a diagram showing the respective differential corresponding clutch torque setting map, 14
FIG. 15 is a block diagram showing the longitudinal acceleration corresponding clutch torque setting means, FIG. 15 is a longitudinal acceleration corresponding clutch torque setting map, FIG. 16 is an example of an engine torque map thereof, and FIG. 17 is a transmission torque thereof. FIG. 18 shows an example of a ratio map. FIG. 18 is a block diagram showing a modification of the engine torque proportional clutch torque setting means.
FIG. 19 is a center differential input torque setting map, FIG. 20 is a characteristic diagram of the protection control clutch torque, and FIG. 21 (a).
FIG. 21B is a diagram showing a duty characteristic of the first preload learning, FIG. 21B is a diagram showing a pressure characteristic of the first preload learning, and FIG. 22 is a diagram of a pressure characteristic of the second preload learning. FIG. 23 (a) is a diagram showing the characteristic, FIG. 23 (a) is a diagram showing the duty characteristic of the third preload learning, FIG. 23 (b),
(C) is a diagram showing the pressure characteristic of the third preload learning, FIG. 24 is a diagram showing the torque distribution state display means, and FIG. 25 is a diagram for explaining the torque distribution by the torque distribution state estimation means. 26 is a flowchart showing the control flow of the entire vehicle including the device, FIG. 27 is a flowchart showing the flow of the front and rear wheel torque distribution control, and FIG. 28 is the differential corresponding clutch torque FIG. 29 is a flowchart showing the flow of setting the clutch torque corresponding to the longitudinal acceleration, and FIG.
FIG. 31 is a flowchart showing a flow of setting the engine torque proportional clutch torque, FIG. 31 is a flowchart showing a flow of setting the protection control clutch torque, FIG. 32 is a flowchart showing a flow of the first preload learning, No.
FIG. 33 is a flowchart showing the flow of the second preload learning, and FIG. 34 is a flowchart showing the flow of the third preload learning. 2 ... Engine, 4 ... Torque converter, 6 ... Automatic transmission, 8 ... Output shaft, 10 ... Intermediate gear (transfer idler gear), 12 ... Center differential (Center differential), 14 ... For front wheels Differential gearing, 15 ……
Bevel gear mechanism, 15A ... bevel gear shaft, 15a ... bevel gear, 16, 18 ... front wheel, 17L, 17R ... front wheel side axle, 19 ...
... reduction gear mechanism, 19a ... output gear, 20 ... propeller shaft, 21 ... bevel gear mechanism, 22 ... differential gear device for rear wheel, 24, 26 ... rear wheel, 25L, 25R ... rear Wheel axle, 2
7 ... front wheel output shaft, 27a ... hollow shaft member, 28 ... differential limiting mechanism, 28a ... front wheel output side disk plate, 28b ...
Input side disk plate, 29 …… Output shaft for rear wheel, 30,30
a, 30b, 30c ... steering wheel angle sensor, 32 ... steering wheel, 34, 34a, 34b ... lateral acceleration sensor, 36 ... longitudinal acceleration sensor, 38 ... throttle sensor, 39 ... engine key switch, 40, 42, 44, 46 ... Wheel speed sensor, 48
…… Controller, 50 …… Anti-lock brake device,
50A …… Brake switch, 51 …… Brake pedal, 52
… Warning light, 54… Hydraulic source, 56… Pressure control system valve (pressure control valve), 58… Pump, 60… Check valve, 62… Pressure control valve, 64… Relief valve, 66… …accumulator,
68 ... Pressure switch, 68a ... Motor, 113 ... Input gear, 114a to 114f ... Bearing, 115 ... Transmission case, 115a ... End cover, 115b ... Spacer member, 116 ... Support member, 117a, 117b Oil passage, 121 Sun gear, 122 Planetary pinion (planetary gear), 123 Ring gear, 124 Input case, 125
… Planet carrier, 125a …… Base plate part, 12
5b planetary pinion storage section, 125f clutch disk support section, 126 pinion shaft, 130 connection member, 141 first piston, 142 second piston, 14
3 ... Partition plate, 144a ... First oil chamber, 144b ... Second
Oil chamber, 145: Hollow shaft, 145a: Piston housing, 160 ...
… Shift lever position sensor (shift range detecting means),
160A: Shift lever of automatic transmission, 161: Control valve as 4WD control means, 162A: Accelerator pedal, 162: Duty solenoid valve (duty valve), 163: Orifice, 164: Oil filter, 165 …… Reducing valve, 170 …… Engine rotation speed sensor, 180 …… Transmission rotation speed sensor, 200 …… Actual front / rear axis rotation speed difference detection unit, 202a-202d…
… Filter, 204a …… Front wheel rotation speed data calculation unit,
204b: rear wheel rotational speed data calculating unit, 206: front and rear wheel actual rotational speed difference calculating unit, 210: front and rear wheel ideal rotational speed difference setting unit, 212: driver required steering as steering angle data detecting means Angle calculation section (pseudo steering angle calculation section), 212a: sensor-responsive steering angle data setting section, 212b: lateral acceleration data calculation section, 212c: comparison section, 212d: driver requested steering angle setting section (vehicle speed data) Setting section), 216: driver required vehicle speed calculating section (pseudo vehicle speed calculating section) as vehicle speed data detecting means, 216a: wheel speed selecting section, 216c: driver requested vehicle speed calculating section, 216d: ... Filter, 218 ... Ideal rotational speed difference setting section as ideal differential state setting section, 220 ... Differential clutch torque setting section, 222 ... Subtractor, 230 ... Protection control section, 242 ... Filter, 244: Horizontal G gain setting section, 24
6... Correction unit, 254, clutch torque setting means corresponding to longitudinal acceleration, 254 a, front wheel shared load calculating means, 254 b, total output torque calculating means, 254 c, front wheel shared torque calculating means,
254d: Cutch torque calculating means, 256: Lateral acceleration correspondence correction section, 258a: Switch, 258: Judgment means, 264:
Engine torque detector, 266: Torque torque ratio detector, 267: Center differential input torque calculator, 268: Engine torque proportional torque setting unit, 269: Clutch torque calculator, 270: Engine torque proportional clutch torque Section, 272: turning correction section, 274a: switch, 274 ...
Determination means, 275: rotational difference gain setting section, 276: transmission reduction ratio detecting section, 280: maximum value selecting section, 2
82: Torque-pressure conversion unit, 286: Centrifugal correction pressure setting unit, 288: Initial engagement pressure setting unit (preload setting unit), 290:
Peak hold filter, 292a, 294a ... Switch, 295 ...
… Duty setting unit, 292, 294… Judging means, 296… Pressure feedback correction unit, 298… Pressure-duty conversion unit, 300…, 302… Duty solenoid, 304…
... pressure sensor, 306 ... filter, 310 ... torque estimation means, 310a ... calculation means, 310b ... selection means, 312 ... torque distribution display section, AM ... differential adjustment type front and rear wheel torque distribution control device Mechanical part.

Claims (2)

(57) [Claims] 1. A front-rear wheel-adjustable four-wheel drive vehicle that controls a torque distribution to a front wheel and a rear wheel by adjusting a differential state between a front wheel side and a rear wheel side. A front and rear wheel differential adjustment mechanism that limits the differential between the front and rear wheels by applying a frictional force between the front and rear wheels to adjust the differential state,
Longitudinal acceleration detecting means for detecting longitudinal acceleration of the vehicle, shared load calculating means for calculating a front wheel shared load or a rear wheel shared load in the magnitude of the longitudinal acceleration detected by the longitudinal acceleration detecting means, and the longitudinal acceleration detecting means The output torque calculating means for calculating the minimum engine output torque necessary for obtaining the magnitude of the longitudinal acceleration detected in the above, and the shared load calculated by the shared load calculating means and the engine output torque calculated by the engine output torque calculating means. A shared load torque calculating means for calculating the shared load torque from the output torque, and a control means for adjusting the differential limiting amount based on the shared load torque. Front and rear wheel torque distribution control device. 2. In a region where the differential limiting amount set based on the shared load torque is reduced by the control means as the longitudinal acceleration increases, the differential limiting amount is controlled to suppress this reduction. 2. The differential-adjustment-type front / rear wheel torque distribution control device according to claim 1, further comprising means for correcting the torque.