JP2653303B2 - Hydraulic adjustment type driving force distribution control device for vehicles - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a driving force distribution control device for a vehicle which is suitable for controlling driving force distribution to right and left driving wheels in a vehicle and controlling driving force distribution to front and rear wheels in a four-wheel drive vehicle. In particular, the present invention relates to a hydraulically-adjustable driving force distribution control device for a vehicle that performs driving force distribution control by hydraulic pressure.

[0002]

2. Description of the Related Art Power transmission of a motor vehicle configured to control the ratio of the torque transmitted to the front wheels (also referred to as driving torque or driving force) to the torque transmitted to the rear wheels in accordance with the driving state. Various devices are known. For example, in a so-called full-time four-wheel drive automobile, a center differential for appropriately distributing the torque from the engine to the front wheel side and the rear wheel side, and a viscous coupling for limiting a differential in the center differential are used. It is conceivable to provide a differential limiting mechanism and adjust the differential limiting mechanism to control the torque ratio in accordance with the driving state of the vehicle.

[0003] In addition to such drive force distribution control to the front and rear wheels, drive force distribution control to the left and right wheels may be considered.

[0004]

However, in each of the above-described differential limiting mechanisms, the differential control characteristics are determined by physical properties and the like, and the differential control mechanism does not actively perform the differential control. No torque distribution control can be performed. Therefore, as a differential limiting mechanism, between two rotating members that cause a differential (for example, between a front wheel side rotating member and a rear wheel side rotating member, or between a right wheel side rotating member and a left wheel side rotating member), For example, a mechanism is conceivable in which a hydraulic wet multi-plate clutch is interposed, the hydraulic pressure is adjusted, and the engagement state of the multi-plate clutch is adjusted to limit the differential.

That is, an oil chamber (piston chamber) and a piston driven in accordance with the oil pressure in the oil chamber are provided, and the engagement state of the wet multi-plate clutch is changed through the piston. Thereby, when the multi-plate clutch is engaged by adjusting the oil pressure in the oil chamber, the driving torque is transmitted from the member having the higher rotation speed to the member having the lower rotation speed between the two rotating members that generate the differential, Torque distribution control can be performed. By finely adjusting the engagement state of the multi-plate clutch, it is possible to adjust the torque transmission state from the member having the higher rotation speed to the member having the lower rotation speed, and to perform the torque distribution control positively. It can also contribute to performing appropriate torque distribution control.

In order to realize a driving force distribution control device based on such differential limitation, it is assumed that hydraulic control can be performed with high accuracy. Incidentally, the oil chamber and the piston that generate the hydraulic pressure need to be provided adjacent to the wet-type multi-plate clutch, and are generally provided in a rotating portion such as one of the above-mentioned two rotating members. However, if the oil chamber is provided in a rotating part, a hydraulic pressure (centrifugal hydraulic pressure or centrifugal pressure) is generated in the oil chamber by centrifugal force, and it becomes difficult to control the hydraulic pressure with high precision as described above.

SUMMARY OF THE INVENTION The present invention has been made in view of the above-mentioned problems, and has been developed in view of the above problem. It is an object to provide a control device.

[0008]

SUMMARY OF THE INVENTION Accordingly, a hydraulically-adjustable driving force distribution control device for a vehicle according to the present invention is a vehicle driving force distribution control device for distributing and controlling the driving force of an engine in a vehicle to each driving wheel. An oil chamber provided in the rotating part,
A hydraulic piston built in the oil chamber, driving force distribution adjusting means driven by the hydraulic piston to adjust the driving force distribution state of the engine, and hydraulic pressure adjusting means for adjusting the oil pressure in the oil chamber; Control means for setting the oil pressure in the oil chamber so that the driving force distribution adjusting means performs a required driving force distribution state, and controlling the hydraulic pressure adjusting means based on the set oil pressure. In addition, a centrifugal oil pressure correction unit is provided for offsetting and correcting the centrifugal oil pressure generated according to the rotation speed of the oil chamber when setting the oil pressure.

[0009]

In the above-described driving force distribution control device for a hydraulically adjusted vehicle according to the present invention, the control means sets the oil pressure in the oil chamber so that the driving force distribution adjusting means performs a required driving force distribution state. In setting the oil pressure, the centrifugal oil pressure generated by the centrifugal oil pressure correction unit of the oil pressure setting unit is offset and corrected in accordance with the rotation speed of the oil chamber. The hydraulic pressure adjusting means is controlled in accordance with the magnitude of the hydraulic pressure set in this way, the hydraulic pressure in the oil chamber is adjusted through the hydraulic pressure adjusting means, and the driving force distribution adjusting means is provided through the hydraulic piston. It is driven to perform the driving force distribution state.

[0010]

BRIEF DESCRIPTION OF THE DRAWINGS FIG. 1 is a block diagram showing the construction of a main part of a hydraulically-adjustable driving force distribution control apparatus for a vehicle according to an embodiment of the present invention; FIG. FIG. 3 is a cross-sectional view showing the main part of the drive torque transmission system, FIG. 4 is a cross-sectional view of the main part of the front and rear wheel torque distribution mechanism, and FIG. 5 is a schematic circuit diagram of the hydraulic supply system. 6, 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 steering angle data detecting means, and FIG. FIG. 10 is a block diagram showing details of the detection means.
FIG. 11 is a diagram showing the ideal rotation speed difference setting map, FIG. 11 is a diagram showing the rotation difference gain setting map, and FIG. 12 is a plan view schematically showing wheel states for explaining the ideal rotation speed difference. 13 is a diagram showing a map for setting a clutch torque for differential acceleration, FIG. 14 is a block diagram showing a clutch torque setting means for longitudinal acceleration, FIG. 15 is a map for setting clutch torque for longitudinal acceleration, and FIG. FIG. 17 is a diagram showing an example of the transmission torque ratio map, FIG. 18 is a block diagram showing no modification of the engine torque proportional clutch torque setting means, and FIG. 19 is a center differential input torque setting map thereof. 20 is a characteristic diagram of the protection control clutch torque, FIG. 21 is a diagram for explaining the first preload learning, and FIG. FIG. 23 is a diagram showing the pressure characteristic of the second preload learning, FIG. 23 is a diagram for explaining the third preload learning, FIG. 24 is a diagram showing the torque distribution state display means, and FIG. FIG. 26 is a characteristic diagram for explaining torque distribution by the estimating means. FIG. 26 is a flowchart showing a flow of control of the entire vehicle including the device, and FIG.
FIG. 28 is a flowchart showing the flow of the front / rear wheel torque distribution control, FIG. 28 is a flowchart showing the flow of setting the differential corresponding clutch torque, FIG. 29 is a flowchart showing the flow of setting the longitudinal acceleration corresponding clutch torque, and FIG.
0 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, and FIG.
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 flowchart showing a flow of the third preload learning.

First, with reference to FIG. 2, an overall configuration of a drive system of a vehicle including a differentially adjusted front and rear wheel drive force distribution control device as the hydraulically adjusted vehicle drive force distribution control device will be described. In FIG. 2, reference numeral 2 denotes an engine. The output of the engine 2 is transmitted to an output shaft 8 via a torque converter 4 and an automatic transmission 6. The output of the output shaft 8 is output from a center differential (hereinafter abbreviated as center differential) 12 as an operating device for distributing the engine torque (= driving torque or driving force) of the front wheels and the rear wheels to a required state via the intermediate gear 10. Is transmitted to

The output of the center differential 12 is transmitted from the axles 17L and 17R to the left and right front wheels 16 and 18 via a reduction gear mechanism 19 and a front differential gear device (front differential) 14 on the one hand, and bevel gears on the other hand. Mechanism 1
5, transmitted from the axles 25L, 25R to the right and left rear wheels 24, 26 via a propeller shaft 20, a bevel gear mechanism 21, and a rear wheel differential gear device 22.

The center 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 the conventionally known gear. The automatic transmission 6 is mounted on a carrier 125 that supports the planetary gear 122.
The output of the output shaft 8 is input, 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 connected to the rear wheel output shaft 29 and the bevel gear mechanism. 15 is linked to the propeller shaft 20.

In the center differential 12, the distribution of the engine output torque between the front wheels and the rear wheels can be changed 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 28 as a differential limiting mechanism or a differential adjusting mechanism is provided. That is, the hydraulic multi-plate clutch 28 is connected to the sun gear 121 (or the ring gear 1).
23) and the carrier 125, the frictional force is changed by the control pressure applied to its own hydraulic chamber, and the differential between the sun gear 121 (or the ring gear 123) and the carrier 125 is restrained. It has become.

The hydraulic multi-plate clutch 28 may be interposed between the sun gear 121 and the ring gear 123. Accordingly, the center differential 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 about 32:68 for the front wheels and the rear wheels. The control can be performed at a ratio (for example, 60:40) corresponding to the ground load of the front and rear wheels from the degree.

Front wheel in completely free state: rear wheel value: about 3
2:68 can be defined by setting the tooth ratio of the input gears on the front wheel side and the rear wheel side of the planetary gears. In this case, when the pressure in the hydraulic chamber of the hydraulic multi-plate clutch 28 is zero and completely free, It is set to be about 32:68.
The ratio in this completely free state (about 32:68) is
This value usually varies depending on the load balance between the front wheel system and the rear wheel system, and the like. Also, the pressure in the hydraulic chamber is set to the set pressure (about 9 kg / cm ² ), and
8 is in the locked state, and when the differential limit becomes substantially zero, the torque distribution between the front wheels and the rear wheels becomes a ratio (for example, 60:40) according to the ground load of the front and rear wheels.

Reference numeral 30 denotes the steering wheel 3
2, steering wheel angle sensors (steering wheel angle sensors) 34a and 34b for detecting a rotation angle from a neutral position, that is, a steering wheel angle (steering wheel angle) θ, are lateral accelerations G acting on the front and rear portions of the vehicle body, respectively.
This is a lateral acceleration sensor that detects yf and Gyr. 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. The detected value may be used as the lateral acceleration data.

Reference numeral 36 denotes a longitudinal acceleration G acting on the vehicle body.
x is a longitudinal acceleration sensor that detects x, 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, 46, and 44 are the left front wheel 16, the right front wheel 18,
These are wheel speed sensors for detecting the rotational speeds of the left rear wheel 24 and the right rear wheel 26. The outputs of these switches and the respective sensors are input to the controller 48.

Reference numeral 50 denotes an anti-lock brake device, which 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 antilock brake operation signal is output in conjunction therewith, and the antilock brake device 50 is operated. Further, when the operation signal of the anti-lock brake is output, a signal indicating the state is simultaneously output to the controller 4.
8 is input. Reference numeral 52 denotes an indicator 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 the control described later. Reference numeral 54 denotes a hydraulic pressure source, and 56 denotes a pressure control valve system (hereinafter, referred to as a pressure control valve) interposed between the hydraulic pressure source 54 and the hydraulic chamber of the hydraulic multi-plate clutch 28 and controlled by a control signal from a controller 48. (Abbreviated).

The vehicle is provided with an automatic transmission. Reference numeral 160 denotes a shift lever 1 of the automatic transmission.
This is a shift lever position sensor (shift range detecting means) for detecting the selected shift range of 60A, and this detection information is also sent to the controller 48. Further, the controller 4 also controls 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.
8 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. In other words, the engine 2 includes the main throttle valve 152 whose opening is controlled in accordance with the amount of depression of the accelerator pedal 53, and constitutes an accelerator pedal-based engine output adjusting device together with the accelerator pedal 53 and the coupling measures. . An auxiliary throttle valve 153 as engine output control means controlled independently of the accelerator pedal system engine output adjusting device.
Is the main throttle valve 1 in the intake passage of the engine 2.
52 in series. This sub throttle valve 1
The motor 53 is driven by a motor, and the motor is controlled based on detection results of the rear wheel speed sensors 44 and 46, the front wheel speed sensors 40 and 42, the engine speed sensor 170, the engine load sensor 172, and the like.

Further, as a sensor, the clutch 28
A hydraulic pressure sensor 304 for detecting a hydraulic pressure applied to the pistons 141 and 142 is provided at a predetermined position. The mechanical part AM of the hydraulically-adjustable front and rear wheel torque distribution control device will be described in further detail. As shown in FIGS.
An input unit for inputting an engine torque (driving force) through the automatic transmission 6, a center differential 12, and a differential limiting mechanism 2
8 and an output unit for the front wheel side and the rear wheel side.

The input section comprises an input gear 113 meshing with the intermediate shaft 10A side, and an input case 124 to which the input gear 113 is serrated.
The bearing 114 is attached to the end cover 115a and the retainer 115b fixed to the transmission case 115.
a and 114b so as to be freely rotatable. The input case 124 has a shape whose diameter is increased toward the front (to the left in FIGS. 3 and 4), and includes an enlarged portion having a planetary gear element therein and a rear portion of the enlarged portion (in FIG. 4). , Right), and an opening is formed in front of the enlarged diameter portion. Then, it can be mounted on the output shaft 29 from behind (rightward in FIGS. 3 and 4) a rear wheel output shaft 29 described later. Also, on the outer periphery of the opening,
A plurality of grooves 124d are formed.

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

The sun gear 121 is provided integrally with the hollow shaft member 27a.
The output shaft 27 for the front wheel and the output shaft 27 for the front wheel are both serrated and connected to the hollow shaft 145, and the hollow shaft member 27a and the output shaft 27 for the front wheel can rotate integrally through the hollow shaft 145. The hollow shaft 145 is provided with a piston housing 145a described later.

The ring gear 123 is connected to the connecting member 13.
0, and the connecting member 130 is connected to the rear wheel output shaft 2.
9 and serration-coupled to the rear wheel output portion. Thus, the output of the ring gear 123 is input to the propeller shaft 20 via the connecting member 130, the output shaft 29 for the rear wheel, and the bevel gear mechanism 15.

The planet carrier 125 has, on the outer periphery thereof, convex portions 125i that can be fitted into the respective grooves 124d of the input case 124, and these fittings allow the planet carrier 125 to rotate integrally with the input case 124. Have been.
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. Also,
In order to fix each pinion shaft 126, a stopper 134 is provided.

The sun gear 121 and the ring gear 12
3, a plurality of planetary pinions 122 are provided.
2 are 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. The planet carrier 125 has a flange-shaped base plate portion 125a and a planetary pinion accommodating portion 12 formed rearward of the base plate portion 125a.
5b, and a cylindrical clutch disc supporting portion 125f formed forward.

Then, each of these members 121, 122,
The 123 and 125 can be separately assembled as the center differential unit 12 in advance, and after being sub-assembled in this way, the center differential unit 12 can be mounted in the transmission case 115.
After the input case 124 is mounted in the case 115, the input case 124 is mounted so as to cover the center differential unit 12.

Hydraulic multi-plate clutch 2 as differential limiting mechanism
8 is an input-side disk plate 28 attached to the clutch disk support 125 f of the planet carrier 125.
b and a plurality of front wheel output-side disk plates 28a mounted on a clutch case 146 that rotates integrally with the sun gear 121 and the front wheel output shaft 27 via the hollow shaft 145, and are alternately arranged in parallel.

Of these, the front wheel output side disk plate 2
8a is driven by the first piston 141 and the second piston 142, and can be joined 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 movable in the axial direction, respectively. Between the piston 141 and 142, the piston accommodating portion 145 is provided.
Partition plate 143 fixed to a and not moving in the axial direction
Is interposed.

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

Each of these members 28a, 28b, 141,
142, 143, 145, 146 can also be separately assembled in advance as the hydraulic multiple disc clutch unit 28,
After the sub-assembly, the hydraulic multi-plate clutch unit 28 can be mounted in the transmission case 115. The output unit includes a front wheel output unit and a rear wheel output unit, and the front wheel output unit includes:
It comprises an output shaft 27 for a front wheel formed of a hollow shaft, and an output gear 19a mounted on the output shaft 27 for the front wheel and meshing with an input gear 19b of a differential gear device (differential) 14 for a front wheel. The rear wheel output portion includes a rear wheel output shaft 29 provided so as to penetrate through the front wheel output shaft 27, a bevel gear shaft 15A coupled to a tip end of the rear wheel output shaft 29, and Bevel gear 15b attached to bevel gear shaft 15A and at the tip of propeller shaft 20
And a bevel gear 15a that meshes with it.

The output gear 19a has a bearing 114c,
114d, the bevel gear shaft 15A and the bevel gear 1
5a is supported on the transfer case 115c side via bearings 114e and 114f. Output gear 1
The reduction gear mechanism 19 is constituted by 9a and the input gear 19b, and the bevel gear mechanism 15 is constituted by 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, and 108 is a transfer drive. A gear, 109 is a rear cover, and 112 is an end clutch.

In FIG. 4, reference numerals 132a, b and c denote plates for axially supporting the shafts, and reference numeral 133 denotes a circlip. 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. 5, the reservoir 6a also serves as the automatic transmission 6, and the pump 58 for sucking oil in the reservoir 6a is supplied from its discharge port via the check valve 60 and the pressure control valve 62. And is connected to a hydraulic chamber of the hydraulic multi-plate clutch 28.
The pressure control valve 62 can take a first position that connects the hydraulic chamber of the hydraulic multi-plate clutch 28 and the pump 58 and a second position that connects the pump 58 and the reservoir 6a of the automatic transmission 6. The position during this time is controlled by the controller 48.

In the passage between the check valve 60 and the pressure control valve 62, there is provided a reducing valve 165 that opens at a set pressure (for example, about 9 kg / cm ² ) to release oil to the reservoir of the automatic transmission 6. 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. The motor 58a for driving the pump 58 is controlled by a control signal from the controller 48.

The specific configuration of the pressure control valve 62 is as shown in FIG. 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. Spool body 161a
Receives the duty pressure (solenoid control pressure) Pd and the reducing pressure Pr at both ends thereof. When the duty pressure Pd decreases, it advances to the left in the figure to be in the open state, and when the duty pressure Pd increases, it increases to the right in the figure. The operation proceeds to the closed state. The spring 161d biases the spool body 161a in an appropriate direction so that the spool body 161a can move appropriately as described above.

Reference numeral 162 denotes a duty solenoid valve (duty valve).
Is a valve body 16 driven by a solenoid 162a and the solenoid 162a and a return spring 162c.
2b, and the valve element 162b is provided with a solenoid 162a.
When the solenoid 162a is not operated, the oil passage 169f is opened, and when the solenoid 162a is not operated, the oil passage 169f is advanced by the return spring 162c to close the oil passage 169f. The reduty valve 162 is connected to the controller (computer) 4 based on information from various sensors.
8 is electronically controlled.

Reference numeral 163 denotes an orifice, 164 denotes an oil filter, 165 denotes a reducing valve, and the orifice 163 includes a reducing valve 165 and a 4WD.
Between the control valve 161 and the oil filter 1
64 is an oil passage 16 flowing into the reducing valve 165.
9b. The reducing valve 165 includes a valve body 165a having a return spring 165b.
The valve body 165a is automatically supplied with hydraulic pressure by the urging force when the hydraulic pressure becomes equal to or less than the set pressure, and is discharged when the oil pressure becomes equal to or more than the set pressure. To move to.

Therefore, for example, the solenoid 162a
When the duty valve 162 is opened by operating, 4WD
The hydraulic pressure at the left end of the control valve 161 (duty
Pressure) Pd decreases and is returned by the return spring 161d.
By moving the valve body 161b, 161c to the left,
The passage between the oil passages 169c and 161g is opened, and the line pressure P ₁
Is operating hydraulic pressure (4WD clutch pressure) P_FourAs hydraulic multi-board
The oil is supplied to each of the oil chambers 144a and 144b of the latch 28.
So that the hydraulic multi-plate clutch 28 is connected.
It is configured.

When the duty valve 162 is closed without operating the solenoid 162a, the hydraulic pressure (duty pressure) at the left end of the 4WD control valve 161 is set.
Pd rises, and the valve body parts 161b and 161c move rightward (to the position shown in FIG. 6), and the oil passages 169c and 161c are moved.
Together and the 9g are disconnected so 4WD clutch pressure P ₄ is released, the hydraulic multi-plate clutch 28 is configured so as to be separated.

The relationship between the duty (Duty), which is a control index of the duty valve 162, and the 4WD clutch pressure P ₄ (= control oil pressure P) is, for example, as shown in FIG. Is 4WD
Clutch pressure P ₄ becomes 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, with reference to the block diagram of FIG. 1, the components of the controller relating to the control for restraining the differential of the center differential 12 by the hydraulic multi-plate clutch 28 (hereinafter referred to as driving force distribution control or center differential control) will be described. I will explain.
In this control, each sensor (wheel speed sensors 40, 42, 4)
4, 46, steering angle sensors 30a, 30b, 30c, lateral acceleration sensor 34, longitudinal acceleration sensor 36, throttle position sensor 38, engine speed sensor 170, transmission speed sensor 180, shift position sensor 160, etc.) , The clutch torque of the hydraulic multi-plate clutch 28 is set, and the differential hydraulic pressure of the hydraulic multi-plate clutch 28 is controlled so as to obtain the target clutch torque.

The ABS information, wheel speed,
Steering angle, gear position, total communication between ABS control unit and engine control unit (SCI communication: SCI = S
data such as erial Communication Interface) is input digitally, and the longitudinal acceleration, lateral acceleration, accelerator opening,
Analog inputs are made regarding the hydraulic control for the multi-plate clutch, the control of the 4WD control unit, the current to the electromagnetic clutch of the rear wheel differential gear device (rear differential) 22, and the like.

The setting of the clutch torque of the hydraulic multi-plate clutch 28 is performed in consideration of the differential state between the front wheels and the rear wheels (a rotational speed difference, which is also referred to as a rotational speed difference). And a longitudinal-acceleration-responsive clutch torque Tb for performing control in accordance with the longitudinal acceleration acting on the vehicle.
And the engine torque proportional clutch torque Ta that is set in proportion to the engine torque so that a large road surface transmission torque can be obtained in the front and rear wheels directly connected four-wheel drive state at sudden start, etc., and the clutch part of the wet multi-plate clutch One of the clutch torques Tpc, Tb, Ta, and Tpc will be described in order.

The differential corresponding clutch torque Tv is a clutch torque that causes the vehicle to behave in accordance with the driver's intention during turning. To control the posture of the vehicle, the rear wheel is used as a drive base and the rear wheel is driven. Since it is effective to set to slip, the clutch torque T corresponding to the differential
v is set to achieve such a state.

For this reason, the portion related to the setting of the differential corresponding clutch torque Tv (front and rear wheel rotational speed difference proportional calculating means 20)
1), as shown in FIG. 1, a front and rear wheel actual rotation speed difference detection unit 200, a front and rear wheel ideal rotation speed difference setting unit 210, a front and rear wheel actual rotation speed difference ΔVcd, and a front and rear wheel ideal rotation speed difference ΔVhc.
Thus, a differential-corresponding clutch torque setting section 220 for setting the clutch torque Tv 'from the above, and a correction section 246 for correcting the clutch torque Tv' for the lateral acceleration.

The front and rear wheel actual rotational speed difference detecting section 200 includes filters 202a to 202d, a front wheel rotational speed data calculating section 204a, and a rear wheel rotational speed data calculating section 204.
b and an actual front and rear wheel rotational speed difference calculation unit 206. The filters 202a to 202d provide 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 the wheel speed sensors 40, 42, 44, 46, respectively. From among
This is for removing a micro-vibration component of data generated by disturbance or the like.

The front wheel rotational speed data calculating section 20
In 4a, 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 outputs the rear wheel rotation speed data signal RL. , RR, the rear wheel rotation speed Vr is obtained by averaging the respective wheel speeds of the rear wheels.

Further, the front and rear wheel actual rotational speed difference calculating section 206
Then, by subtracting the front wheel rotation speed Vf from the rear wheel rotation speed Vr, the actual rotation speed difference between the front and rear wheels [the rotation speed difference between the front and rear wheels (the rotation difference between the front and rear wheels, which is also the rotation difference in the center differential)] ΔVcd Is calculated. 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 a steering angle data detection unit, and a driver request body speed calculation unit (a vehicle speed data detection unit). An estimated vehicle speed calculation unit or a pseudo vehicle speed calculation unit) 216;
Ideal rotation speed difference setting unit 21 as ideal operating state setting unit
8 is provided.

As shown in FIG. 8, the driver-requested steering angle calculation unit 212 as the driver-requested steering angle data setting means is provided as follows.
Steering angle sensor 30 installed on steering handle
(The first steering angle sensor 30a, the second steering angle sensor 30b,
Based on the detection data θ ₁ , θ ₂ , θn from the neutral position sensor 30 c), the sensor corresponding steering angle δh [=
f (θ ₁ , θ ₂ , θ n)] and a sensor-based steering angle data setting unit 212a for calculating the value of the lateral acceleration sensor 34a,
A lateral acceleration data calculator 212b for averaging the data Gyf and Gyr detected at 34b to calculate lateral acceleration data Gy.
And the direction of the sensor corresponding steering angle δh and the lateral acceleration data Gy
And a comparing unit 212c for comparing the
A driver required steering angle setting unit (steering angle data setting unit) 21 that sets a driver required steering angle δref according to the comparison result in
2d.

The function δh = f (θ ₁ , θ ₂ , θn) for obtaining the sensor-corresponding steering angle δh depends on the specifications of the steering wheel angle sensor. Also, the sensor-compatible steering angle δh
In each of the lateral acceleration data Gy, for example, the right turning direction is positive. In order to compare the directions of the sensor-corresponding steering angle δh and the lateral acceleration data Gy, a function SIG (x) relating to the direction with respect to the detected data x is as follows.
Set. 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 is compared with the direction of the lateral acceleration data Gy.
(Δh) is compared with SIG (Gy).

Then, the driver required steering angle setting section 212d
When the direction SIG (δh) of the sensor-assisted steering angle δh is equal to the direction SIG (Gy) of the lateral acceleration data Gy, the sensor-assisted steering angle δh is set to the driver-requested steering angle (steering angle data) δref. The direction SIG (δh) of the sensor corresponding steering angle δh and the direction SIG of the lateral acceleration data Gy
If (Gy) is not equal, 0 is set to the driver-requested steering angle δref.

The direction SIG (δh of the steering angle δh corresponding to the sensor
) Is not equal to the direction SIG (Gy) of the lateral acceleration data Gy, the reason why the driver-requested steering angle δref is set to 0 is that, for example, when the driver performs a steering operation such as counter-steering, the steering position of the steering wheel and the steering position are determined. In some cases, the actual steering angle (turning state) of the vehicle may be different. In such a case, if the steering angle of the vehicle is set from the steering position of the steering wheel, it is difficult to perform appropriate control.

In order to eliminate such inconvenience, if 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, the driver's required steering angle is set. Is set to 0. As shown in FIG. 9, the estimated vehicle speed calculation unit 216 includes rotation speed data of the left front wheel 16, the right front wheel 18, the left rear wheel 26, and the right rear wheel 28 detected by the wheel speed sensors 40, 42, 44, 46. A wheel speed selector 216a for selecting the second largest wheel speed data from the signals FL, FR, RL, and RR from the bottom (from the smaller one), and an estimated vehicle speed from the selected wheel speed data and the like. Estimated vehicle speed calculator 216
c.

In particular, in the estimated vehicle speed calculating section 216c, the wheel speed data SVW obtained by filtering the wheel speed data selected by the wheel speed selecting section 216a through the filter 216b to remove noise components, and the longitudinal speed sensor 36 detect the wheel speed data SVW. Based on the longitudinal acceleration data Gx obtained by removing the noise component by applying the longitudinal acceleration to the filter 216d, the subsequent vehicle speed is estimated from both the data SVW and Gx at a certain point in time. That is, the wheel speed data S at a certain point in time
Assuming that VW is V ₂ and longitudinal acceleration data Gx is a (acceleration data at one time point to be calculated), a theoretical vehicle speed Vref after a time t after this time point can be calculated by Vref = V ₂ + at. Using the changing longitudinal acceleration data Gx, the theoretical vehicle speed V after time t from this point
ref can be calculated by Vref = V ₂ + ∫Gx dt.

The reason why the wheel speed data having the second largest wheel speed data among the rotation speed data signals FL, FR, RL, and RR is adopted as the wheel speed data SVW is that each wheel is normally on the overspeed side. In many cases, the vehicle slips and it is normally desirable to use the wheel speed of the lowest rotation. However, the second wheel speed from the bottom is used in consideration of data reliability.

Then, the ideal rotational speed difference setting section 218 calculates the driver's required steering angle δref calculated by the driver's required steering angle calculating section 212 and the estimated vehicle speed Vref calculated by the estimated vehicle speed calculating section 216. An ideal rotation speed difference ΔVhc is set in accordance with a map as shown in FIG. In other words, with respect to the vehicle speed, at low vehicle speeds, the influence of the difference in the orbit radii of the front and rear wheels at the time of turning (so-called inner wheel difference) is large, and the rotation speed Vr of the rear wheel is smaller than the rotation speed Vf of the front wheel, but the vehicle speed is high. As the rotation speed Vr of the rear wheel increases with respect to the rotation speed Vf of the front wheel, the rear wheel can easily slip at high speed. As a result, the responsiveness of the posture of the vehicle body, which is required at higher speeds, is ensured. As for the steering angle, the larger the steering angle is, the larger the rotation difference required for the front and rear wheels is. Therefore, the larger the magnitude | δref | of the steering angle data δref, the larger the value of ΔVhc.

The difference between the rotational speeds ΔVhc of the front and rear wheels due to the difference in the orbit radii of the front and rear wheels will be described with reference to FIGS. 12A is a diagram schematically illustrating a two-wheeled vehicle including one front wheel and one rear wheel, and FIG. 12B is a diagram schematically illustrating FIG. 12A.
As shown in FIGS. 12A and 12B, the front wheel speed is V
f, the rear wheel speed is Vr, the vehicle speed at the center of gravity of the vehicle is V,
The turning radius of the front wheel is Rf, the turning radius of the rear wheel is Rr, the turning radius of the vehicle center of gravity is R, the vehicle body slip angle is β, the wheelbase is 1, the distance between the front wheel center and the center of gravity is If, the rear wheel center is Assuming that the distance from the center of gravity is lr, the rotational speed difference ΔVhc between the front and rear wheels is expressed as follows. ΔVhc = Vr−Vf = [(Rr−Rf) / R] · Vref (1.1) where Rr = ｛R ² + lr ² -2Rlr · cos (π / 2-β)｝ ^1/2 Rf = ｛R ² + lf ² -2Rlf · cos (π / 2 + β)｝ ^1/2 β = (1-m / 2l · lf / lr · kr · V) / (1 + A · V ² ) · lr / l · δ where m is vehicle weight, kr is rear cornering power, A
Is the stability factor.

The front wheel speed Vf and the rear wheel speed Vr
Is considered theoretical, Vf: Vr = Rf: Rr,
Vf: V = Rf: R, and the further angle .beta.f shown in FIG. 12 (b), the .beta.r, there are relationships βf-βr = AV ^2, and from the equations of these relationships as described above, the DerutaVhc V It can be defined as a function of δ [ΔVhc = fc (V, δ)]. However, V in this case is a theoretical value, that is, the estimated vehicle speed Vref.
Corresponds to the theoretical value, that is, the driver's required steering angle δ.
ref is equivalent. Such a function [ΔVhc = fc (Vre
f, δref)] is mapped as shown in FIG.

Incidentally, the steering angle θ
In addition to the actual steering angle (steering angle corresponding to the sensor) δh based on the above, 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 ] · l · Gy (1.2) where A is a stability factor, Vref is a theoretical vehicle speed (estimated vehicle speed) described later, l is a wheel base.

The turning G equivalent steering angle δy thus obtained
On the other hand, the above-described actual steering angle (steering angle corresponding to the sensor) δh is a steering angle that more reflects the driver's will. That is, if the driver wants to make a larger turn than the current situation, | δh |> |
δy |, and the magnitude of the ideal rotational speed difference (slip target value) can be made larger by employing the steering angle value | δh | than by employing the steering angle value | δy |. | Δh | <| δy |
By employing the steering angle value | δh |, it is possible to reduce the magnitude of the ideal rotation speed difference (slip target value) as compared with employing the steering angle value | δy |.

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 is:
The difference between the front and rear wheel ideal rotation speed difference ΔVhc set by the front and rear wheel ideal rotation speed difference setting unit 210 is subtracted by the subtractor 222 (ΔVhc).
cd−ΔVhc), and the obtained difference ΔVc (= ΔVcd−Δ)
Vhc) and the ideal rotational speed difference ΔVhc between the front and rear wheels are input to the differential corresponding clutch torque setting unit 220 as data.

The differential corresponding clutch torque setting section 220
Front and rear wheel actual rotation speed difference ΔVcd and front and rear wheel ideal rotation speed difference ΔV
The clutch torque Tv ′ is set in accordance with the difference ΔVc (= ΔVcd−ΔVhc) from the clutch torque Tc.
v ′ is set. (i) When .DELTA.Vhc.gtoreq.0, in this case, it is desired that the speed of the rear wheels 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 a rear wheel slip by transferring part of the engine torque largely distributed to the rear wheels toward the front wheels. . Therefore, the clutch torque Tv '
Tv '= a * ([Delta] Vcd- [Delta] Vhc) = a * [Delta] Vc (1.3) so that is increased in proportion to the difference [Delta] Vc ([Delta] Vcd- [Delta] Vhc) (where a is proportional. constant).

If ΔVhc>ΔVcd> 0, the actual rotational speed of the rear wheel is higher than that of the front wheel even though the front wheel is slipping.
If the value 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.gtoreq..DELTA.Vcd, the front wheels are slipping, and it is desired to reduce the distribution of the engine torque to the front wheels to reduce the slip of the front wheels. Therefore, Tv ′ = − a × ΔVcd = −a × (ΔVc + ΔVhc) (1.4) is set so that the clutch torque Tv ′ increases in proportion to the magnitude of ΔVcd (where a is Proportional constant).

FIG. 13 (a) shows a map of the relationship between Tv 'and .DELTA.Vc. The map shows the relationship between the difference .DELTA.Vc and the ideal rotational speed difference .DELTA.Vhc between the front and rear wheels. Tv can be determined.
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 higher than that of the rear wheels, and the clutch torque Tv ′ is set as follows.

If .DELTA.Vcd.gtoreq.0, the rear wheels are slipping due to excessive rotation, and it is desired to suppress a rear wheel slip by transferring a part of the engine torque largely distributed to the rear wheels toward the front wheels. . Therefore, when the clutch torque Tv 'is Δ
Tv '= a.times..DELTA.Vcd = a.times. (. DELTA.Vc + .DELTA.Vhc) (1.5) so as to increase in proportion to the magnitude of Vcd (where a is a proportional constant).

If 0>ΔVcd> ΔVhc, the actual rotational speed of the rear wheel is higher than that of the front wheel even though the rear wheel is slipping.
Increases the engine torque distributed to the rear wheels, and the slip of the rear wheels is promoted. For this reason, it is desired to reduce the engine torque distributed to the rear wheels by setting the differential limit to free. Therefore, in this case, the clutch torque Tv 'is set to 0 to set a so-called dead zone.

If .DELTA.Vhc.gtoreq..DELTA.Vcd, the front wheels are slipping. Therefore, it is desired to reduce the distribution of the engine torque to the front wheels to reduce the slip of the front wheels. Therefore, Tv '=-a.times. (. DELTA.Vcd-.DELTA.Vhc) =-a.times..DELTA.Vc (1.6) so that the clutch torque Tv' increases in proportion to the magnitude of .DELTA.Vc (.DELTA.Vcd-.DELTA.Vhc). (Where a is a proportional constant).

FIG. 13B shows the relationship between Tv 'and .DELTA.Vc as shown in FIG. 13 (b). The map shows the relationship between the difference .DELTA.Vc and the ideal rotational speed difference .DELTA.Vhc between the front and rear wheels. Tv can be determined.
In this manner, the differential corresponding clutch torque setting unit 220
Then, referring to the map [FIGS. 13A and 13B], ΔVc
And differential clutch torque Tv obtained from ΔVhc
'Is corrected for lateral acceleration.

[0075] The correction unit 246 is subjected to a lateral acceleration correction by multiplying the lateral G gain k ₁ to the differential corresponding clutch torque Tv ', but so as to obtain a differential response clutch torque Tv, the lateral G gain k ₁ is set as follows. That is, 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 of the setting portion 244 of FIG.

The lateral G gain k ₁ is determined by the friction coefficient μ of the road surface.
Is to be reflected in the control, and the lateral acceleration G
As y becomes larger, it can be determined that the road surface μ is larger, and as the road surface μ becomes larger, it is desired that the distribution of the engine torque is made mainly of the rear wheels and that the turning performance of the vehicle body is given priority. Therefore,
The size of the road μ (therefore, the size of the lateral acceleration Gy)
If larger, the lateral G gain k ₁ is decreased, thereby performing the correction for reducing the set clutch torque Tv. It should be noted that even when the road surface μ is large, it is conceivable to omit the correction by the lateral G gain k ₁ unless special priority is given to the turning performance of the vehicle body.

The wheel slip corresponding clutch torque Tb is
In the above-described wheel slip corresponding clutch torque control,
When traveling on a low μ road (a road with a low friction coefficient μ),
When all four wheels slip and control hunting is likely to occur, clutch torque for preventing deterioration of the steering characteristics such as strong understeering of the vehicle and enabling the vehicle to perform a smooth turning operation. The control is performed in accordance with the wheel slip amount Ev.

This clutch torque Tb for wheel slip is
Is set in the clutch torque setting means 2 corresponding to the wheel slip.
First, the estimated vehicle speed Vref and the wheel speed S
From VW, the wheel slip amount Ev (= SVW-Vre)
f) is calculated, and the clutch torque Tb is set based on the slip amount Ev. For this reason,
As shown in FIG. 14, the wheel slip corresponding clutch torque setting means 254 includes a slip amount calculation unit 254a and a clutch torque setting unit 254b.

In the slip amount calculating section 254a, the estimated vehicle speed Vref calculated by the estimated vehicle speed calculating section 216d of the estimated vehicle speed calculating section 216 and the wheel speed selecting section 216a selected by the wheel speed selecting section 216a of the estimated vehicle speed calculating section 216. The wheel speed SVW, which is the second smallest wheel speed among the wheel speeds of the wheels, is input, and the difference between these wheel speeds (= SVW-Vref) is used as the wheel slip amount Ev.
Is calculated.

That is, if the wheels do not slip, the estimated vehicle speed Vref becomes equal to the second lowest wheel speed SVW, which is originally assumed to be substantially equal to the actual vehicle speed. The estimated vehicle speed Vref is smaller than the wheel speed SVW. And this wheel speed SV
The difference between W and the estimated vehicle speed Vref is determined by the slip amount Ev of the wheel.
You can think.

In the clutch torque setting section 254b, FIG.
5, the clutch torque Tb corresponding to the wheel slip is set based on the slip amount Ev. That is, as shown in the map of FIG. 15, when a slip occurs, the clutch torque Tb is increased in proportion to the increase in the slip amount Ev, and when the slip amount Ev becomes an appropriate magnitude, the clutch torque Tb becomes the maximum value ( M
AX).

That is, when a slip occurs, it is desired to immediately connect the clutch immediately to ensure that the driving force can be transmitted to the road surface. However, when the clutch is directly connected, the behavior of the vehicle body suddenly changes, and a shock or a sense of discomfort occurs during traveling. This may cause a hunting of the control in a state on the boundary line of whether or not the slip has occurred. Therefore, even if a slip occurs, the clutch torque Tb is gradually increased according to the slip amount Ev, and the control is smoothly shifted from the other controls to the clutch torque control corresponding to the wheel slip. The hunting does not seem uncomfortable.

The slope (Tb / Ev) at the time of increase in the map of FIG. 15 is determined to an appropriate size according to the characteristics of the vehicle. That is, the gradient (Tb
/ Ev) can be applied to different vehicle types. It is also conceivable that the inclination (Tb / Ev) can be changed according to other factors such as the vehicle speed even for one vehicle type (see a chain line portion in FIG. 15).

When the four-wheel slip is continuing, a four-wheel slip-only map as shown by a broken line in FIG. 15 is prepared so that the clutch is always directly connected even if the slip amount Ev is small. It is conceivable to set the clutch torque Tb based on this map. Although not shown, the lateral G is applied to the wheel slip corresponding clutch torque Tb set in this manner, similarly to the correction unit 246 described above.
It may be obtained a wheel slip corresponding clutch torque Tb by performing a lateral acceleration corrected by multiplying the gain k _1.

[0085] The lateral G gain k ₁ will be omitted here and described above, because 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. Here, the second smallest wheel speed is represented as the wheel speed as the wheel speed data SVW.
If at least the second lowest wheel speed is higher than the vehicle speed Vref, it can be determined that the wheel is slipping, and the slip amount is at least the second lowest wheel speed S.
From the idea that there is only the portion corresponding to VW, such a second lowest wheel speed is adopted.

On the other hand, it is conceivable to use the average value of the three wheel speed data excluding the smallest wheel speed data among the rotation speed data signals FL, FR, RL and RR as the wheel speed data SVW. In order to judge the slip state of each wheel, it is originally preferable to perform the determination based on all of the wheel speed data of the four wheels. However, small wheel speed data is excluded in consideration of the reliability of the data. is there.

The clutch torque Tb corresponding to the wheel slip set as described above is output to the maximum value selecting section 280 as data. Engine torque proportional clutch torque Ta
The set torque for setting the direct-coupled four-wheel drive state in advance 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. It is.

Therefore, this engine torque proportional clutch
Set the torque Ta (engine torque proportional crack)
The chit torque setting means 260 is shown in the lower left part of FIG.
The engine that detects the engine torque Te at a certain moment
Torque detector 264 and the torque converter torque ratio t at that time.
The torque converter torque ratio detection unit 266 for detecting the
Transmitter that detects the transmission reduction ratio ρm
And the engine torque Te.
Engine torque based on a map set in proportion to
Engine torque proportional to engine torque Ta '
Gin torque proportional torque setting section 268 and this engine
The torque converter torque ratio t, torque
Transmission reduction ratio ρm, final reduction ρ₁And rotation difference
Gain k _TwoMultiply by the engine torque proportional clutch
Engine torque proportional clutch torque calculation to obtain torque Ta
Section 270 and the set engine torque proportional clutch
Only when the speed is low (for example, Vref <20 km / h)
And a switch 274a for outputting data as data.
I have.

In the engine torque detector 264, the filter 262 sent from the throttle position sensor 38
The throttle opening degree data θth from which the minute vibration component of the data generated by the disturbance or the like is removed through a, and the engine from which the minute vibration component of the data generated by the disturbance and the like is transmitted from the engine speed sensor 170 through the filter 262b. The engine torque Te at that time is obtained from the rotation speed data Ne through an engine torque map as shown in FIG. 16, for example.

The torque converter torque ratio detector 266 receives the filter 262b from the engine speed sensor 170
For example, the transmission torque ratio Ne shown in FIG. 17 is obtained from the engine speed data Ne from which the disturbance component has been removed through the transmission speed sensor N and the transmission speed data Nt sent from the transmission speed sensor 180 and having the disturbance component removed through the filter 262c. The transmission torque ratio t at that time is obtained through a map.

Transmission reduction ratio detecting section 276
In this embodiment, the transmission reduction ratio ρm is determined from the selected shift stage information from the shift lever position sensor 160 with reference to a shift stage-reduction ratio correspondence map as shown in a block 276 in FIG. A map used for setting the engine torque proportional torque setting unit 268 (FIG. 1)
), The engine torque Te and the engine torque proportional torque Ta ′ are proportional constants determined from known constants such as the number of teeth Zs and Zr of the sun gear and the ring gear, the front wheel shared load Wf, and the vehicle weight Wa. Is a linear relationship.

The engine torque proportional clutch torque calculating section 270 calculates the engine torque proportional torque Ta ′ determined as described above, the torque converter torque ratio t, the transmission reduction ratio ρm, the final reduction ρ _1, and the rotation difference gain k.
_The rotation difference gain k ₂ is set by the rotation difference gain setting section 275 as follows. That is, the rotation difference gain k ₂ is to avoid the tight corner braking phenomenon, and the ideal rotation speed difference setting unit 218
Is determined according to 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) , however, ΔVhcmax = MAX | ΔVhc (δ = MAX) | The coefficient 0.9 and constant 0.1, the k ₂ This is to set the lower limit to 0.1.

As described above, the rotation difference gain k ₂ decreases linearly as the ideal rotation speed difference ΔVhc increases, and the rotation difference gain k ₂ is corrected by multiplying the rotation difference gain k _2. When the speed difference ΔVhc becomes large, turning performance (performance that can prevent the tight corner braking phenomenon) is prioritized over sudden starting performance,
The engine torque proportional clutch torque Ta is reduced.

As shown in FIG. 18, the engine torque proportional torque setting section 268 and the engine torque proportional clutch torque calculating section 270 are replaced with a center differential input torque calculating section 268 and a clutch torque calculating section 270 as shown in FIG.
It is also conceivable to change the configuration to include the swing correction unit 70 and the turning correction unit 270a. That is, in the center differential input torque calculation section 268, the engine torque Te sent from the engine torque detection section 264 and the torque converter torque ratio detection section 26
6, a center differential input torque (transmission output torque) Td is calculated by the following equation from the torque converter torque ratio t sent from 6 and the transmission reduction ratio ρm sent from the transmission reduction ratio detector 276. _{Td = t · ρm · ρ 1} · Te ··· (3.2) However, ρ ₁ is the final reduction ratio.

The relationship between the center differential input torque Td and the engine torque Te is proportional to each set shift. For example, when the torque converter torque ratio t is set to 1.5, the relationship becomes as shown in FIG. However, actually,
Since this relationship greatly changes depending on the magnitude of the torque converter torque ratio t, the torque converter torque ratio t may be obtained from the speed ratio i, and the relationship between Ta and Te may be obtained based on this.

The clutch torque calculating section 270 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-Zs/Zr] .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. , Wa are vehicle weights.

[0097] Then, the turning correction unit 270a, 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 by the following equation. Tc = [(Zs + Zr) / Zr · Wf / Wa-Zs / Zr ] _{· t · ρm · ρ 1 ·} Te ··· (3.4) In addition, the switch 274a is the signal from the determination unit 274, At low vehicle speed (Vref <20km / h in this example)
Is turned on to output the engine torque proportional clutch torque Ta as data. However, when the vehicle speed further increases (Vref ≧ 20 km / h), it is turned off and the data of the engine torque proportional clutch torque Ta is output. To stop the output. 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 this engine torque proportional control is performed only at low 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 section 230. .
That is, in the wet-type multi-plate clutch 28, in general, when the differential rotation (= ΔN) between the clutch plates is increased, there is a fear that damage such as seizure of the clutch facing and an increase in the wear amount may be caused. Damage is more likely to occur as the duration of the state is longer. This slip between the clutch plates is a level that does not cause any problem under general driving conditions,
Under special conditions, for example, when running resistance becomes extremely large when running on sandy ground, when mounting different diameter tires, or when mounting a tire chain, it becomes more likely to occur, and the load on the clutch 28 becomes excessive under such special conditions. Prone.

Therefore, under such severe operating conditions, a control for stopping the clutch slip (clutch protection control) is performed by detecting this and setting the clutch 28 in a directly connected state. In this apparatus, the maximum torque capacity of the clutch 28 is set so that the clutch 28 can be directly connected even under the above-mentioned severe conditions.
By applying the maximum oil pressure to the clutch 8, the clutch 28 is brought into the directly connected state.

Further, when the clutch protection control is ended and the control is shifted to another control, if the connection state of the clutch 28 is instantaneously switched to the free state, the posture of the vehicle may be suddenly changed. Therefore, it is necessary to set the protection control clutch torque Tpc by the protection control unit 230 so that these phenomena can be avoided. This protection control unit 230
First, the state of occurrence of ΔN is estimated from the working surface pressure of the clutch 28 and the running conditions from the ΔN proportional gain of the differential limiting clutch 28, and the absorbed energy q of the clutch 28 is obtained. 28 protection control is performed.

That is, the absorbed energy q (here,
q ₂ ) is equal to the threshold value q ₀ [for example, 0.6 (kgfm / c)
m ² )], the clutch protection control (that is, the control for bringing the clutch 28 into the direct connection state) is executed. The absorbed energy q of the clutch 28 is the same as that of the clutch 28 because the clutch 28 is a continuous slip clutch.
From the unit absorbed energy q ′.

This unit absorbed energy q 'is
Of the product of the surface pressure (clutch torque per unit area) P and the differential rotation speed V [ie, q ′ = f (P ·
V)]. Therefore, the unit absorbed energy q 'can be obtained from the clutch torque Tc, the clutch slip speed Δωc (Δωc is obtained from ΔVcd) and the total area A of the clutch by the following equation. q ′ = Tc · Δωc / A Then, the absorbed energy q ₂ is obtained by integrating the unit absorbed energy q ′ with time as in the following equation. q ₂ = ∫
q'dt The integration in this case is performed for a predetermined time (for example, 2
(T) (seconds). For example, when the calculated time is t ₀ (unit of second), the definite integration is performed from time (t ₀ -2) to time t ₀ .

Actually, such calculation is performed in a computer such as a microcomputer as a control means. For example, the following calculation of the unit absorbed energy q 'is performed periodically to obtain the wheel speed difference ΔV And pressure sensor value (clutch 2
The unit absorbed energy q 'is periodically obtained from the measured value of the surface pressure applied to the pressure P) and the preload learning value (to be described in detail later) Pi. q ′ = C · ΔV · (P−Pi) Here, for example, assuming that the control sampling period is T seconds, the absorbed energy q is calculated as in the following equation. q ₂ = TΣq ′ Here, Σq ′ is the sum of the values of q ′ sampled during a predetermined time (for example, 2 seconds).

The absorption energy q thus obtained
₂ is equal to or greater than a threshold value q ₀ (q ₂ ≧ q ₀ ), protection control is started. At this time, the protection control clutch torque Tpc
Is set as shown in the graph of FIG. That is,
When the control condition is satisfied, the protection control clutch torque Tpc
Is first set to the maximum value (for example, 40 kgfm), the clutch 28 is directly connected, and this is held for a fixed time (1 second in this example), and then the clutch torque Tpc is set at a fixed gradient (at a fixed rate). It is set to decrease to the initial pressure.

However, at the time of this decrease, the deviation ΔVc of the actual wheel speed difference Vcd from the target value (target wheel speed difference) Vhc
While monitoring (= Vcd−Vhc), if the magnitude of the deviation ΔVc is equal to or larger than a set value ΔV ₀ (eg, ΔV ₀ = 1.0),
That is, if | ΔVc | ≧ ΔV ₀ , the maximum pressure is set again. This is because, when the clutch torque Tpc is reduced, the direct connection of the clutch 28 is released, and the clutch 28 can slip. At this time, if the road surface, the tires, and the like are easily slipped,
The slip of the clutch 28 increases. In this case, the clutch protection may be required again after this, and the clutch torque Tpc is set to the maximum value again before the clutch protection condition (here, the condition that q ₂ ≧ q ₀ ) is satisfied. To make the clutch 28 directly connected,
The hunting of the clutch control is prevented to ensure the protection of the clutch and to perform the stable clutch control.

Here, the slip of the clutch 28 becomes large, and the wheel speed difference Vcd corresponding to this clutch slip is larger than the target wheel speed difference Vhc by a set value (ΔV _0).
= 1.0) or more (that is, | ΔVc |
≧ ΔV ₀ ), the clutch torque Tpc is again
Is set to the maximum value, and the clutch 28 is brought into the directly connected state.

This is because if the traveling environment such as the road surface and the tires becomes less severe and the wheel speed difference Vcd can be made closer to the target wheel speed difference Vhc, the differential corresponding clutch torque Tv corresponding to the target wheel speed difference Vhc is used. This is because the control is performed so that the control for improving the traveling performance of the original vehicle is performed. Therefore, if the wheel speed difference Vcd cannot quickly approach the target wheel speed difference Vhc after the direct connection of the clutch 28 is released, it is determined that the clutch protection is necessary again, and the clutch protection condition (here, q ₂ ≧ q
_The clutch 28 is brought into the directly connected state before the condition ( ₀ ) is satisfied.

Incidentally, a minute wheel speed difference (a constant value) is set, and if the wheel speed difference Vcd becomes larger than this set value while the clutch torque Tpc is decreasing, the clutch 2 is re-engaged.
8 may be directly connected. by the way,
Consider the above-described setting of the absorption energy threshold value q ₀ . This threshold q ₀ is too small control of the clutch protection will be performed frequently, not performed clutch control for enhancing the driving performance of the original vehicle. Therefore, as really perform only clutch protection control when the clutch protection is required, to be set as large as possible a threshold q _0.

To this end, various severe driving conditions are required.
(For example, when starting on a sandy ground, drifting on a low μ road
At full throttle on low μ roads).
We found the relationship between Nergie q and clutch durability (life)
Then, based on the threshold q ₀Can be set
Wear. Note that this threshold q₀Of the harsh driving conditions described above
Think about setting one of the most important ones
Can be

The above-described clutch torques of the differential corresponding clutch torque Tv, the wheel slip corresponding clutch torque Tb, the engine torque proportional clutch torque Ta, and the protection control clutch torque Tpc are obtained in each control cycle repeated at an appropriate timing. The respective clutch torques Tv, Tb, Ta, Tpc set in this manner are set.
Is sent to the maximum value selection unit 280.

In the maximum value selecting section 280, the largest one among the clutch torques Tv, Tb, Ta and Tpc (this clutch torque is referred to as Tc) for each control cycle.
Select However, when the switch 274a is OFF, the clutch torque Ta is not sent, so that the maximum value selector 280 selects the maximum value from the clutch torques sent.

The clutch torque Tc selected in this way is sent to the torque-pressure converter 282, where
The clutch control pressure Pc is set such that the set clutch torque Tc is obtained. Here, the clutch control pressure Pc is obtained from the clutch torque Tc by using a map (see the block 282 in FIG. 1). However, since the clutch torque Tc and the clutch control pressure Pc are generally in a proportional relationship, the map is also shown. It is linear.

Further, the clutch control pressure Pc set in this way is added to an adder / subtractor 284 as a preload applying means.
, Centrifugal pressure correction and preload correction are performed. In other words, the centrifugal pressure to be corrected (centrifugal correction pressure)
Is Pv, the preload (ie, the initial engagement pressure balanced with the return spring) is Pi, and the corrected clutch pressure (control pressure) is P
Assuming cd, the correction is performed by the following equation. Pcd = Pc + Pi-Pv The centrifugal pressure correction will be described.
6 and a part of the adder / subtractor 284 (subtraction part) constitute a centrifugal oil pressure correction unit 285.

The centrifugal pressure correction is performed for the following reason. That is, since the piston chambers such as the oil chambers 144a and 144b rotate, the hydraulic oil therein receives the centrifugal force accompanying the rotation and drives the pistons 141 and 142 by the centrifugal force (therefore, the clutch control pressure). Increase. In order to ensure the control accuracy of the clutch control pressure, it is necessary to subtract and correct the hydraulic pressure (that is, the centrifugal pressure) generated according to the centrifugal force.

Then, the centrifugal pressure correction is performed by subtracting the centrifugal correction pressure Pv set by the centrifugal correction pressure setting section 286 from the clutch control pressure Pc. The centrifugal correction pressure setting unit 286 uses a map as shown in a block 286 in FIG.
It is determined from the front wheel vehicle speed Vf calculated in 04a.

Since the piston chamber rotates in synchronization with the front wheel shaft, the centrifugal oil pressure is generated corresponding to the front wheel speed Vf. Generally, the centrifugal force is proportional to the square of the rotation speed. Therefore, the centrifugal correction pressure Pv is equal to the front wheel vehicle speed Vf
Is set in proportion to the square of. The preload correction is a correction in which an initial engagement pressure (initial pressure) Pi set by an initial engagement pressure setting unit (preload setting unit) 288 is added to the clutch control pressure Pc as a preload.

The purpose of this preload correction is to increase the control response by maintaining the state of contact between the clutch plates of the clutch 28 in a marginal state (extremely slight contact state) so that no torque is generated. Is what you do. 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 the same product also 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 keep the state of contact between the clutch plates in a very close state while detecting the clearance state between the clutch plates at an appropriate timing. Therefore, the preload setting unit 288 sets the initial pressure Pi by performing a trial (here, learning) at an appropriate time interval to determine how much preload is required.

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 obtained under such conditions that the control of the other clutches 28 is not affected.
For this reason, the conditions of the preload learning are set as follows, for example. 30 minutes or more 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). In this example, there is no range of P (parking) and R (reverse), in the case of P and R, 1, 2, 2,
This is because a large oil pressure different from the cases of D and N is output. Vref = 0 km / h (vehicle speed Vref is 0). Tc ≦ 1 kgfm [Clutch torque Tc is a small predetermined value
(1kgfm) or less].

When the above conditions are simultaneously satisfied, 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. 4kgf /
cm ² ] equivalent duty is applied for 2 seconds, and thereafter, for example, at an increasing speed of 1.5% / s, for example, P = 3.
Slowly sweep to a duty equivalent to 0 kgf / cm ² .

Then, the pressure P applied to the hydraulic pistons 141 and 142 changes as shown in FIG. That is, since the clutch plates are separated at first, if the duty rises gently, the hydraulic piston 2 will respond accordingly.
8, the pressure P gradually increases, but when the hydraulic pistons 141 and 142 move to a certain position, the clutch plates come into contact with each other, and the pressure P also receives the force of the return spring. As a result, the pressure P suddenly increases. Furthermore, the hydraulic piston 141,1
As 42 moves, the clutch plate comes into strong contact and the clutch is completely engaged. This state is understood from the fact that the increase in the pressure P becomes the upper limit.

Here, a value (difference) P ″ obtained by second-order differentiation of the detected pressure P with respect to time, and the pressure P is calculated as 1
The second-order differential value (difference) P 'is calculated from time to time in a short cycle. When the second-order differential value P "becomes maximum, it is determined that the clutch plate has started to contact, and the pressure P at this time is determined. Is determined as the initial pressure, and when the first-order differential value P ′ becomes maximum is determined as when the clutch plate is fully engaged.

More specifically, when the learning is started and the pressure P increases, the maximum value of the second-order differential value P ″ and the pressure P at this time are stored. This second-order differential value P ″ Is calculated for each short control cycle and is updated as appropriate. When the first-order differential value P ′ becomes 0 (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 The pressure P at the time when the maximum value is obtained is stored as the initial pressure Pi.

If any of the conditions for the preload learning is not satisfied during the execution of the preload learning, the preload learning is immediately interrupted and the operation 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 a condition 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. ′10 after the ignition key is turned on
Minutes have passed. 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 ≦ 1 kgfm [Clutch torque Tc is a small predetermined value
(1kgfm) or less]. 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) equivalent to the initial pressure Pi (= P ₁ ) is held for a predetermined time (for example, 2 seconds), and then P = 8.8 kgf / for a predetermined time (for example, 1 second). Sweep to a duty equivalent to cm ² (almost 100% duty).

Thus, the hydraulic pistons 141, 14
The pressure P applied to 2 changes in 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 some point in the going is swept duty, the clutch starts to drag in contact, hydraulic pistons 141 and 142 shocks The pressure P suddenly increased, overshooted, and oscillated almost 100
Settles to the complete engagement pressure (steady peak pressure) according to the duty of%.

When the pressure P overshoots,
Subsequent steady maximum pressure Pc (known value, here 8.8
A peak value (maximum value) Pmax larger than a certain value (about kgf / cm ² ) is generated. On the other hand, if 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 L 2, and at a certain point in time the smooth full engagement is achieved. Pressure (steady maximum pressure) Pc.

From these characteristics, the peak value P of the pressure P
stores the max, the difference between the value Pmax and a constant maximum pressure Pc α (= Pmax -Pc) is greater than the predetermined value alpha _0, it can be determined that the clutch in the initial pressure P ₁ are separated. Therefore, while increasing or decreasing appropriately the starting pressure P from the initial value P _1, by repeating the trials as described above at appropriate time intervals (e.g., every five minutes), the appropriate initial pressure Pi
Can be detected and set.

That is, this preload learning is desirably performed 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 generalized and expressed. , The initial learning value is PINTG (n) and the initial pressure Pi is PINTG
(N). Therefore, the previous initial learning value is PINTG (n-1), and the initial pressure is PINT (n
-1), and in the n-th learning stage, learning is performed with the previous initial pressure by PINT (n-1). Then, the difference α (= Pmax−Pc) obtained by the predetermined duty sweep is compared with the threshold α _0, and the current initial learning value PINTG (n) and the initial pressure PINT (n) are 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 α ≧ α ₀ , the initial learning value PINT
For G (n), the previous initial learning value PINT
G (n-1) is set to the value obtained by adding β (= 1 pressure for one bit), and the initial pressure PINT (n) is set to β (= 1bi) from the previous initial learning value PINTG (n-1).
(initial learning value PINTG (n)).

When α ≧ α ₀ , it can be determined that an overshoot has occurred.
In (n-1), it can be determined that the clutch 28 is not approaching the barely contact state. Therefore, the current initial learning value PINTG (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 set to the previous initial learning value PINTG (n). The learning value PINTG (n-1) is changed to β (=
(1 bit pressure).
Note that 1 bit is limited by the resolution of the hydraulic pressure sensor signal for detecting the hydraulic pressure applied to the piston.
1bit = 0.05kgf / cm ² or 1bit = 0.1kgf / c
set at an appropriate value m ^2, and the like.

On the other hand, when α <α ₀ , the initial learning value PINTG (n) is set to a value obtained by subtracting β (= 1 bit) from the previous initial learning value PINTG (n−1). The 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, so the previous initial pressure PINT (n−1)
Then, it can be determined that the clutch 28 is in a barely contact state or an excessive contact state. Therefore, the current initial learning value PINTG (n) is replaced with the previous initial learning value PINT (n).
It is assumed that NTG (n−1) is reduced by β (= 1 bit), but the initial pressure PINT (n) is set to the previous initial learning value PINTG (n−1).

This is because α <α₀Just the result of
Clutch 28 is in bare contact or excessive contact
I can't tell if it's in a state and I fear chattering
So, in order to avoid this, immediately
Instead of using the initial pressure Pi, use the previous learning value.
-ing Therefore in excessive contact
And α <α for at least two consecutive cycles ₀Continues
And the initial pressure Pi is delayed by one cycle.
While being reduced and approaching what is appropriate
And

If any of the preload learning conditions (1) to (4) is not satisfied during the execution of the preload learning, the preload learning is immediately interrupted to return to the normal mode. Further, the above-described preload learning is continued as long as the conditions of the preload learning described above are satisfied.

Next, a third preload learning method 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. ′10 after the ignition key is turned on
Minutes have passed. 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 ≦ 1 kgfm [Clutch torque Tc is a small predetermined value
(1kgfm) or less]. 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, FIG.
The duty is adjusted so that a pressure pattern as shown in FIG. 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 In a time (for example, 1 second), the duty is swept to a duty equivalent to P = 8.8 kgf / cm ² (almost 100% duty), and a duty equivalent to P = 8.8 kgf / cm ² is applied for a predetermined time (for example, 2 seconds). Hold. This pattern is continuously repeated while appropriately changing the initial pressure Pi.

As a result, the hydraulic pistons 141, 14
The pressure P applied to 2 is, as in the case of the second preload learning,
As shown by curves L1 and L2 in FIGS. 23 (b) and (c),
Two types of patterns change. Then, from the time point t ₀ when the duty sweep starts (or the time point t ₁ when 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. During this time, the straight line L0 and the curve L1 depicting the changing state of the pressure P
Or the area S of a portion (hatched in the figure) surrounded by L2
1 and S2, a curve L1 with overshoot
The area S1 in the case of
2 is clearly larger than the area S2.

Therefore, in the third preload learning, as in the second preload learning, the above-described trial is repeated at appropriate time intervals (for example, every 5 minutes) to detect an appropriate initial pressure Pi. Can be set. In other words, this preload learning is performed repeatedly as long as the above condition is satisfied, and the initial learning value and the initial pressure Pi set at a certain point in time (the nth learning stage) are changed to the initial learning value in the same manner as described above. With PINTG (n) and initial pressure Pi
Is generalized to PINT (n).

Therefore, the previous initial learning value is P
INTG (n-1), initial pressure is PINT (n-
In the n-th learning stage, the learning is performed by the initial pressure PINT (n-1) at the previous time. Then, by comparing the area S and threshold value S ₀ obtained by a predetermined duty sweep, this initial learning value PINTG (n) and the initial pressure PINT (n)
Is set as follows. When S ≧ S ₀ , PINTG (n) = PINTG (n−1) + β PINT (n) = PINTG (n−1) + β = PINTG (n) When S <S ₀ , PINTG (n) = PINTG ( n−1) −β PINT (n) = PINTG (n−1) That is, when S ≧ S ₀ , corresponding to α ≧ α ₀ in the second preload learning, and when S <S ₀ , Α of the second preload learning
<Α ₀ .

That is, when S ≧ S ₀ , it can be determined that an overshoot has occurred.
In (n-1), it can be determined that the clutch 28 is not approaching the barely contact state. Therefore, the current initial learning value PINTG (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 set to the previous initial learning value PINTG (n). The learning value PINTG (n-1) is changed to β (=
(1 bit pressure).

On the other hand, when S <S ₀ , since there is no overshoot, the previous initial pressure PINT (n−
In 1), it can be determined that the clutch 28 is in a barely contact state or an excessive contact state. Therefore, it is assumed that the current initial learning value PINTG (n) is obtained by adding β (= 1 bit) to the previous initial learning value PINTG (n−1), but the initial pressure PINT (n) is the same as the previous initial learning value PINTG (n). The value is set as PINTG (n-1). The reason for this is that, similarly to the case of α <α ₀ , it is not possible to judge whether the clutch 28 is in the barely contact state or in the excessive contact state only from the result of S <S ₀ , In order to avoid chattering, this learning result is not immediately used as the initial pressure Pi but the previous learning value is used to avoid this.

Therefore, if there is an excessive contact state,
S <S for at least two consecutive cycles ₀It is assumed that
Although the initial pressure Pi is delayed by one cycle,
Will be reduced and approached to what is appropriate.
You. During execution of such preload learning, the above preload
If any of learning conditions' ~ is no longer satisfied
If this is the case, the preload learning is immediately interrupted and the mode returns to the normal mode.

The above-described preload learning is continued as long as the conditions of the preload learning described above are satisfied. In the third preload learning, the area L (S1, S2) of the portion surrounded by the straight line L0 and the curve L1 or L2 is replaced with a straight line L3 indicating a constant pressure about the initial pressure and a curve L3.
1 or L2, the area S '(S1', S1)
The determination may be made with reference to 2 ′).

In this case, the start of the calculation of the area S 'is
At the start of the sweep ₀(Or pressure P
Time t when ascending started₁) And the end of the calculation of the area S ′
Is the steady-state maximum pressure Pc as shown by the straight line L0 (or
(Close to a constant pressure value). And
Set the criterion value to S₀′, S ′ ≧ S₀′
It can be determined that there is a basshoot, and S ′ <S₀′
It can be determined that there was no overshoot.

As described above, the centrifugal pressure correction is performed by subtracting the centrifugal correction pressure Pv from the clutch control pressure Pc, which is the effective hydraulic pressure, and the initial pressure (preload) Pi is added to correct the preload application correction. The control pressure Pcd (= Pc-Pv + Pi) of the oil chamber supply level to which the pressure control is applied is taken into the peak hold filter 290. The peak hold filter 290 is a kind of limiter for preventing an excessive sudden change in the hydraulic pressure so that hunting does not occur in the control due to a sudden change in the hydraulic pressure. 4 kg / 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.

When the control pressure Pcd at which the speed of the oil pressure change exceeds such a limit is sent, the control pressure according to the limit value is maintained. Further, the control pressure Pcd ′ passing through the filter 290 is
a, 294a, and is sent to the duty setting unit 295. The switch 292a is turned off by the signal from the judging 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. .
In other words, on condition that ABS control is not performed,
A signal of the control pressure Pcd 'is sent. 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 connected to the judgment means 29.
This is a control switch for protecting the duty solenoid valve in response to a signal from the control unit 4. When the speed is low and the set clutch torque Tc is small, the duty is set to 0. The low-speed condition can be defined as, for example, Vref ≦ 5 km / h, and the clutch torque Tc can be defined 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 section 295 includes a pressure feedback correction section 296 and a pressure-duty conversion section 298.
With Pressure feedback correction unit 296
Receives the control information from the pressure sensor 304 for detecting the actual pressure acting on the piston and receives the control pressure Pcd ′.
This is for correcting the characteristic of the hydraulic circuit. The signal sent from the pressure sensor 304 to the pressure feedback correction unit 296 has a filter 306 from which noise components due to disturbance or the like are removed.

The pressure-duty converter 298 sets (Duty) corresponding to the control pressure P that has been feedback-corrected by the pressure feedback corrector 296, and is set in the block of the clutch pressure-duty converter 298 in FIG. As shown in the map, the duty increases linearly with 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.

Hydraulic circuit 300 functioning as control execution unit
Thus, the duty solenoid 302 is operated in accordance with the duty set in this way to control the center differential hydraulic multiple disc clutch 28. 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 FIGS. 1 and 24, a torque distribution display section 312 is provided in the meter cluster to graphically display (or meter display) the state of torque distribution to the front wheels (or rear wheels). The torque distribution state is displayed by the torque estimation means 310 in accordance with the magnitude of the distribution torque estimated. As described above, the torque distribution state is estimated by the torque estimation unit 310 because it is difficult to actually measure the torque distribution state.

The torque estimating means 310 detects the difference between the front wheel output torque (or the rear wheel output torque) when there is a difference in rotation speed between the front and rear wheels, and the difference in rotation speed between the front and rear wheels. Calculating means 310a for calculating the front wheel output torque (or rear wheel output torque) when the vehicle is not running, and the front wheel output torque (or rear wheel output torque) for each of these cases.
And a selecting means 310b for selecting the smaller front wheel output torque (or rear wheel output torque), and these portions 310a and 310b estimate the torque distribution state as follows. .

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 a state similar to the engagement of the gears, and the center differential clutch always slips. The other case is that the center differential clutch may not slip because there is actually a slip between the tire and the road surface.

Thus, 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 simplification. ρf / rf <ρr ・ ρt / rr (5.1) where ρf: front differential ratio ρr: rear differential ratio ρt: transfer ratio rf: front wheel tire radius rr: rear wheel tire radius Is a direct-coupled four-wheel drive distribution, so the front wheel torque Tf and the rear wheel torque Tr are as follows. Tf = Wf / Wa · ｛Tm + kWr · rf / ρ · (rfρr ρt / rrρt-1)｝ ... (5.2) Tr = Wr / Wa · ｛Tm-kWf · rr / ρ -1)｝ ... (5.3) where Wf: front wheel sharing weight Wr: rear wheel sharing weight Wa: vehicle weight (= Wf + Wr) Tm: transmission output torque (= center differential input torque) k: slip ratio coefficient ρ: end Reduction 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) , 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 load distribution, a differential ratio difference, or the like. It is. Now, the case is considered in which the clutch transmits torque from the rear wheel side to the front wheel side. Therefore, regarding the front wheel torque Tf, Tf ', Tf, Tf'
The smaller one of the values can be considered as the front wheel torque value.

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

FIG. 25 shows the center differential input torque T
and the front wheel torque distribution ratio m with respect to the input torque. The characteristic of the front wheel torque distribution ratio corresponding to the input torque is a straight line with direct connection when the clutch is in the locked state, and the control pressure when the clutch is in the free state. It becomes curved according to the size of P. Note that the drawing shows a case where the pressure P is 2 kgf / cm ² (P = 2) and a case where the pressure P is 8 kgf / cm ² (P = 8).

In the characteristic graph, the characteristic line with the smaller m of the straight line and the curve for 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 front wheel torque distribution ratio m follows this straight line. 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 this graph, the directly connected straight line is always below, so the front wheel torque distribution ratio according to the straight line marked as directly connected m. Thus, the front wheel torque distribution ratio m
Is set, a signal corresponding to the set value is sent to the torque distribution display unit 312, and the torque distribution display unit 312 displays the state of torque distribution to the front wheels. In this example, the torque distribution to the front wheels is 32% to 50%.
%, The torque distribution display section 312 is provided with a scale corresponding thereto, and the lamp is turned on or the pointer is moved to the corresponding scale, so that the torque distribution display section 312 is displayed in an easily understandable manner.

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 a graph or the like in an analog manner.
Since the differentially adjusted front and rear wheel drive force distribution control device as the hydraulically-adjustable vehicle drive force distribution control device is configured as described above, the differential adjustment is performed as follows.

First, the flow of the entire operation of the drive system is shown in FIG.
As shown in FIG. 6, first, each control element is initially set (step a1), information from various sensors is input (step a2), and learning of the neutral position of the steering angle (step a1) is performed.
3) and learning of the preload of the clutch (step a4), and then the clutch 2 is operated in accordance with the set duty.
8, the front and rear wheel driving force distribution control is performed (step a5), and further, the rear differential is controlled (step a6).

In steps a7 to a11, engine output control (traction control) such as slip control, trace control, torque selection, retard control calculation, and SCI (Serias Communication Interface) communication control is performed, and the torque distribution display lamp is turned on. (Step a
12), a failure diagnosis (failure diagnosis) is performed in step a13. At step a14, a predetermined time (15 mse
c) It is determined whether or not a predetermined time (15 mse
c) After a lapse, a runaway check is performed by a watchdog (step a15), and the process returns to step a2 to repeat a series of controls in steps a2 to a13.

That is, the above-described front / rear wheel drive force distribution control, rear differential control, and engine output control are executed at predetermined intervals (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, the wheel speed F
R, FL, RR, RL, steering angles θ ₁ , θ ₂ , θn, lateral acceleration Gy, longitudinal acceleration Gx, throttle opening θth, engine speed Ne, transmission speed Nt, and selected shift stage are detected. (Step b1), and from these data, the front wheel speed Vf, the rear wheel speed Vr, the driver required vehicle speed Vref, and the driver required steering angle δ
ref and the like are calculated (step b2).

[0163] 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), setting the lateral G gain k ₁ according to the map from the lateral acceleration Gy and (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 ΔV
c, the ideal rotational speed difference ΔVhc, and the lateral G gain k ₁ , calculate the differential corresponding clutch torque Tv (in this example, convert from a map), and calculate the differential clutch torque Tv based on the wheel speeds FR, FL, RR, RL and longitudinal acceleration Gx. calculated from the slip quantity Ev wheel slip corresponding clutch torque Tb of the wheel and (converted from the map), the throttle opening [theta] th, engine speed Ne, transmission rpm Nt, selected shift ratio, the rotation difference gain k ₂ from the engine torque proportional clutch The torque Ta is calculated (converted from the map), and the clutch torque Tpc for protection control is set according to the signal of the actual rotation speed difference ΔVch.

Then, in step b10, the maximum one of these clutch torques Tv, Tb, Ta, Tpc is calculated as the set clutch torque Tc. Further, in step b11, the thus determined clutch torque Tc is converted from the map into the clutch engagement pressure Pc. Subsequently, the pressure Pc is subjected to a preload correction (adding the preload Pi) and a centrifugal pressure correction (decreasing the centrifugal pressure Pv) (step b12) to obtain a center differential control pressure Pcd.

Further, an excessive change in the pressure P is suppressed by passing through a peak hold filter (step b13). Then, whether the ABS is in operation (step b14), the protection condition of the solenoid valve (Vref
.Ltoreq.5 km / h, Tc.ltoreq.1 kgfm) (step b15), and if any of these is satisfied, in step b19, the center differential control pressure Pcd
Is reset to 0.

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

Further, in step b17, the pressure P is converted into a corresponding duty, and the center differential control, that is,
The differential limiting clutch is controlled. The above-described calculation of the differential corresponding clutch torque Tv is performed as shown in FIG. First, a difference ΔVcd (= Vr−Vf) obtained by subtracting the front wheel speed Vf from the rear wheel speed Vr is calculated (step c).
1) And this difference (actual rotational speed difference between front and rear wheels) ΔVcd
Is subtracted from the ideal rotational speed difference ΔVhc of the front and rear wheels obtained as described above (see step b3), and the difference ΔVc (=
ΔVcd−ΔVhc) is obtained (step c2).

Then, in step c3, it is determined whether or not the above-described ideal rotational speed difference ΔVhc of the front and rear wheels is equal to or greater than 0,
If ΔVhc is 0 or more, the process proceeds to step c4, and if ΔVhc is less than 0, the process proceeds to step c5. In step c4, a clutch torque Tv 'is set from .DELTA.Vc using a map (see FIG. 13A).

Specifically, if ΔVcd ≧ ΔVhc, Tv ′ = a × (ΔVcd−ΔVhc) = a × such that the clutch torque Tv ′ increases in proportion to the magnitude of the difference ΔVc (ΔVcd−ΔVhc). ΔVc (where a is a proportional constant). Also, ΔVhc
If>.DELTA.Vcd> 0, the clutch torque Tv 'is set to 0 to set a so-called dead zone.

Furthermore, if 0 ≧ ΔVcd, Tv ′ = − a × ΔVcd = −a × (ΔVc + ΔVhc) is set such that the clutch torque Tv ′ increases in proportion to the magnitude of ΔVcd (where a Is a proportional constant). Note that ΔVhc =
When it is 0, there is no dead zone where ΔVhc>ΔVcd> 0.

At step c5, the map [FIG.
(B) to obtain the clutch torque Tv ′ from ΔVc.
Set. Specifically, if ΔVcd ≧ 0, Tv ′ = a × ΔVcd = a × (ΔVc + ΔVhc) is set such that the clutch torque Tv ′ increases in proportion to the magnitude of ΔVcd (where a is proportional constant).

If 0>ΔVcd> ΔVhc, the clutch torque Tv ′ is set to 0 to set a so-called dead zone. Further, if ΔVhc ≧ ΔVcd, Tv ′ = − a × (ΔVcd−ΔVhc) = − a × ΔVc is set so that the clutch torque Tv ′ increases in proportion to the magnitude of ΔVc (ΔVcd−ΔVhc) ( Where a is a proportional constant.

As described above, in steps c4 and c5, the differential corresponding clutch torque Tv 'is corrected by the correction unit 246.
In the lateral acceleration corresponding corrected by being integrated with the lateral G gain k ₁ (step c6), differential corresponding clutch torque Tv is obtained. By setting the differential corresponding clutch torque Tv, the magnitude of the clutch torque Tv is appropriately set without waste, and the posture control of the vehicle body is appropriately performed while appropriately setting the rear wheel as a drive base so as to slip from the rear wheel. Therefore, the vehicle can behave in accordance with the driver's will when turning.

That is, the direction SI of the sensor corresponding steering angle δh
G (δh) and direction SIG (Gy) of lateral acceleration data Gy
If the steering angle is not equal, the driver-requested steering angle is set to 0. For example, when the driver performs a steering operation such as counter-steering, the steering position of the steering wheel and the actual steering angle of the vehicle (turning Even if the state becomes different, inappropriate data is prevented from being adopted, which contributes to improvement in control performance.

Further, as the driver request vehicle speed Vref,
Since the second largest wheel speed data from among the rotation speed data signals FL, FR, RL, RR is adopted,
Data reliability is ensured. When the ideal rotation speed difference ΔVhc is set at a low vehicle speed, the difference in the orbital 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 becomes higher, the rotation speed Vr of the rear wheel becomes larger than the rotation 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. Also,
Regarding the steering angle, the larger the steering angle, the larger the rotation difference 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 calculation of the clutch torque Tb corresponding to the wheel slip is performed as shown in FIG. First, in step d1, the slip amount calculation unit 254a of the clutch torque setting means 254 for wheel slip corresponds to:
The estimated vehicle speed Vref calculated by the estimated vehicle speed calculation unit 216d of the estimated vehicle speed calculation unit 216 and the estimated vehicle speed calculation unit 2
The second smallest wheel speed among the four wheel speeds selected by the sixteen wheel speed selection units 216a (or the average of three wheel speed data excluding the smallest wheel speed data from the four wheel speeds) Value) SVW from the wheel slip amount Ev (=
SVW-Vref) is calculated.

[0177] Next, in step d2, the slip amount Ev is determined whether fine reference slip amount Ev ₀ or more, as Ev is four-wheel slip state if Ev ₀ or more, the flow advances to step d3, the current It is determined whether or not the clutch torque Tb corresponding to the wheel slip is the maximum (MAX). Here, if Tb is maximum, it is determined that Tb is being maximized in the four-wheel slip state, and the routine proceeds to step d5, where the clutch torque setting unit 254 is set.
b, the wheel slip corresponding clutch torque Tb is maximum (M
AX).

On the other hand, Ev is not greater than Ev ₀ or Tb
If is not the maximum, the process proceeds to step d4, where the clutch torque setting unit 254b uses the map as shown in FIG.
The wheel slip corresponding clutch torque Tb is set based on the slip amount Ev. As a result, as shown in the map of FIG. 15, when a slip occurs, the slip amount Ev
The clutch torque Tb is increased in proportion to the increase in the clutch torque Tb.
b becomes the maximum value (MAX).

As a result, it is possible to smoothly shift from the other control to the clutch torque control corresponding to the wheel slip, and to directly connect the clutch without causing a shock or a sense of incongruity at the time of the control shift, and without causing the hunting of the control. Meanwhile, the driving force can be efficiently transmitted to the road surface by suppressing the slip of the wheels. The inclination (Tb / Ev) in the map shown in FIG.
After setting the inclination (Tb / Ev), the clutch torque T
b, or the steps d2 and d
A configuration is also conceivable in which the clutch torque Tb is always set based on the slip amount Ev, omitting 3 and d5.

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 2
At 68, the engine torque proportional torque Ta 'is read from the engine torque Te through a map (step e).
2). Further, the torque converter torque ratio detecting section 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, in the engine torque proportional clutch torque calculating section 270, the engine torque proportional torque Ta 'thus obtained, the torque converter torque ratio t, and the transmission reduction ratio detecting section 276, 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 section 275, the engine torque proportional clutch torque Ta (= t · ρm · ρ ₁ · T
e) is calculated (step e4).

In step e5, it is determined whether the vehicle speed is low (Vref <20 km / h in this example). If the vehicle speed is low, the above-described engine torque proportional clutch torque Ta is directly output as data. When the vehicle speed further increases (Vref ≧ 20 km / h), 0 is set as the engine torque proportional clutch torque Ta (step e6), and this is output as control data.

By such an engine torque proportional clutch torque Ta, at the time of starting or sudden acceleration from a low speed, etc., a direct connection 4WD state is established so that a high torque can be transmitted to the road surface. In addition, tire slip during sudden acceleration is prevented, driving performance is improved, and driving system durability is also improved. Further, the calculation of the protection control clutch torque Tpc is performed as shown in FIG.

First, at step f1, it is determined whether or not the protection control clutch torque Tpc is zero. If Tpc is 0, the clutch protection control has not been performed, so the process proceeds to step f2 and the clutch 28 from two seconds ago until the present time is executed.
To calculate the absorbed energy q ₂ of. Then, it continues in step f3, whether the absorbed energy q ₂ threshold q ₀ (0.6) or more is judged.

If q ₂ is not larger than the threshold value q ₀ , the clutch protection control is unnecessary, so that Tpc is set to 0 in step f13 and the current control cycle ends. on the other hand,
If q ₂ is the threshold value q ₀ or more, in order to start the clutch protection control, and resets the value t of the timer to 0 the process proceeds to step f4, the following step f5, the maximum value of Tpc (MA
X). Then, in the following step f6, it is determined whether or not the value t of the timer is 1 (second) or more.

If the value t of the timer is smaller than 1, the process proceeds to step f7, where the value t of the timer is increased by the control period (Δt).
And the process proceeds to step f9 and subsequent steps. Then, in step f9, the clutch torque value Tc according to another control law is obtained.
Is determined to be greater than Tpc. If Tc is greater than Tpc, Tpc is set to 0 in step f13, and the clutch protection control is automatically terminated in this control cycle. However, if the value of Tpc is MAX, T
c cannot be greater than Tpc.

When the process proceeds to step f10, it is determined whether or not the accelerator (accelerator pedal) is operated.
It is determined whether or not the accelerator switch is off (OFF). If the accelerator pedal is operated, Tpc is set to 0 in step f13, and the clutch protection control automatically ends in the current control cycle. . If the process proceeds to step f11, brake (brake pedal)
Is determined based on whether the brake switch is on (ON). If the brake pedal is operated, Tpc is set to 0 in step f13, and the clutch is operated in the current control cycle. Protection control automatically ends.

When the process proceeds to step f12, the difference ΔVc between the wheel speed difference ΔVcd and the target wheel speed difference ΔVhc is calculated.
(= ΔVcd−ΔVhc) is equal to the set value ΔV ₀ (= 1.0 km /
h), ΔVc ≧ 1.0 or ΔVc ≦ −1.0
It is determined whether or not | ΔVc | ≧ 1.0 holds. If this holds, the process proceeds to step f4 in the next control cycle. If this does not hold, the process proceeds to step f1 in the next control cycle.

If no route is taken in steps f9 to f12, Tpc is MAX and 0
Therefore, in the next control cycle, the process jumps from step f1 to step f6, and it is determined whether or not the timer t is equal to or longer than 1 (second). If the timer t is equal to or longer than 1 (second), Tpc is set to MAX. Then, Δt is added to the timer t for updating.

Therefore, once the clutch protection control is started, Tpc is held at MAX for one second until the accelerator pedal or the brake pedal is operated. When the accelerator pedal or the brake pedal is operated during this time, the clutch protection control ends. If Tpc is held at MAX for one second after the clutch protection control is started, the timer t becomes 1 or more, so the process proceeds from step f6 to step f7, where Tpc is reduced by a small amount (α).

Then, as shown in FIG. 20, Tpc is linearly reduced to the initial pressure over a predetermined time (here, 10 seconds). Also at this time, if the clutch torque value Tc by another control law becomes larger than Tpc during the decrease, the control automatically shifts smoothly from the clutch protection control to another control law. If the accelerator pedal or the brake pedal is operated during this time, the clutch protection control ends. Further, during this decrease, | ΔVc | ≧ 1.0
Then, the process returns to step 4 again, and the clutch protection control for setting Tpc to MAX is started as shown by the chain line in FIG.

The protection control clutch torque Tpc
, The clutch plate is protected and contributes to the improvement of the durability of the device. And Tpc is M
Since Tpc is gradually reduced after AX, the transition to other control can be performed smoothly, so that control stability and
As a result, stability of the behavior of the vehicle can be ensured. In addition, when this becomes large based on the absorbed energy of the clutch, protection control is started, so that protection control can be performed when clutch protection is really necessary, and other clutch control for improving the running performance of the original vehicle can be performed. There is an advantage that the clutch can be protected without much hindrance.

Further, even when Tpc is decreasing, if the running environment is severe if the road surface, tires, etc. are liable to slip, the slip of the clutch 28 becomes large, and it becomes necessary to protect the clutch again. Again, since the clutch torque Tpc is set to the maximum value and the clutch 28 is maintained in the directly connected state, hunting of the clutch control is prevented, the clutch protection is reliably performed, and stable clutch control can be performed. .

In particular, the wheel speed difference Vcd and the target wheel speed difference V
hc, the above-described control is performed based on the difference between the differential wheel clutch speed Tc and the differential corresponding clutch torque Tv corresponding to the target wheel speed difference Vhc.
The clutch protection can be executed without impairing the control of the clutch as much as possible. Here, regarding the preload correction described above,
This will be described with reference to FIG. First, in the first preload learning method, as shown in FIG. 32, steps g1 to g4
Then, after the ignition key is turned on,
Whether the shift selector is set to 1
(1st gear), 2 (2nd gear), D (drive), or N (neutral)
Whether the vehicle speed Vref is 0 km / h (stop state),
When the set value Tc of the clutch torque is a small predetermined value (1 kg
fm) is determined respectively.

If all of these conditions are satisfied, the process proceeds to step g5. If any of these conditions is not satisfied, the learning control is not performed. Step g5
In step g6, it is determined whether or not preload learning has been performed after the ignition key is turned on. If preload learning has already been performed, learning control is not performed. If preload learning has not been performed, step g6 is performed. Proceed to.

At step g6, the hydraulic pressure is raised, the maximum value (MAX) of the second derivative of the hydraulic pressure is detected, and the hydraulic pressure P at that time is stored. That is, first, as shown in FIG. 21A, a duty equivalent to, for example, P = 0.4 kgf / cm ^{2 is} given for 2 seconds, and thereafter, for example, 1.5% /
At the increasing speed of s, for example, P is slowly swept to a duty corresponding to 3.0 kgf / cm ² .

On the other hand, as shown in FIG. 21 (b), the maximum value of the value (difference) P ″ of the pressure P applied to the hydraulic pistons 141 and 142, which is second-order-differentiated by time, from the pressure P applied to the hydraulic pistons 141 and 142 is calculated. The pressure P at that time is stored, and the stored pressure P is set as an initial pressure.Specifically, when learning starts and the pressure P increases, the second order differential value P ″ Is stored and the pressure P at this time is stored, and the value of the second derivative P ″ is calculated for each control cycle and updated as appropriate. (That is, when the clutch is completely engaged), the calculation of the second derivative P ″ is terminated, and within the period up to this point, 2
The pressure P when the maximum value of the differential value P ″ is obtained is stored as the initial pressure Pi.

Then, during execution of such preload learning,
As soon as any of the above conditions for preload learning is no longer satisfied, preload learning is interrupted and the mode returns to the normal mode.If this preload learning is performed once with the ignition key turned on, Once the ignition key is turned off and then turned on, it is not executed.

In the second preload learning method, FIG.
As shown 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, about 5 minutes or less) ), Whether the shift selector is 1 (first speed), 2 (second speed), D
It is determined whether any one of (drive) and N (neutral) is selected, whether Vref = 0 km / h, and whether Tc ≦ 1 kgfm.

When all of these conditions are satisfied, the process proceeds to step h6, and if any of these conditions is not satisfied, the learning control is not performed. Step h6
Then, the hydraulic pressure is raised and the overshoot value of the hydraulic pressure is detected. In other words, start-up of hydraulic pressure and held for a preset initial initial pressure P ₁ corresponding duty (duty) for a predetermined time (e.g., 2 seconds), P = 8 thereafter a predetermined time (e.g. 1 second). Sweep to a duty equivalent to 8 kgf / cm ² (almost 100% duty).

Then, the overshoot value α of the pressure P applied to the hydraulic pistons 141 and 142 which changes according to the
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, and the maximum value Pmax is detected.
The constant maximum pressure Pc difference (here approximately 8.8 kgf / cm ²⁾ and the (Pmax -Pc) as overshoot value alpha, this alpha is larger than the threshold (alpha _0), there is an overshoot, Therefore, the initial initial pressure can be determined that P ₁ in the clutch 28 is separated, is not greater than the α threshold overshoot was not, i.e., the initial initial pressure P ₁ in the clutch 28 or barely contact It can be determined that there is 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) The initial learning value PINTG (n) is calculated by adding β to the previous initial learning value PINTG (n−1).
(= Pressure for 1 bit), and the initial pressure PINT (n) is the value obtained by adding β (= pressure for 1 bit) to the previous initial learning value PINTG (n-1), that is, , This time initial learning value PINT
Set to G (n).

On the other hand, if α is not larger than the threshold value, 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 For (n), the previous initial learning value PINTG (n-1) is β
(= 1 bit), but the initial pressure PINT (n) is set to the previous initial learning value PINTG (n-1).

If any of the preload learning conditions (1) to (5) is not satisfied during the execution of the preload learning, the preload learning is immediately interrupted to return to the normal mode. Further, the above-described preload learning is continued as long as the conditions of the preload learning described above are satisfied.

In the third preload learning method, the thirty-fourth
As shown in the figure, 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 shorter) has elapsed since the previous trial. Whether the shift selector is 1 (first speed), 2 (2
Speed), D (drive), or N (neutral), Vref = 0 km / h
Whether Tc ≤ 1 kgfm,
Each is judged.

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. Step h
In step 16, the oil pressure is raised and the difference between the predetermined pressure and the oil pressure value is integrated. In other words, start-up of hydraulic pressure and held for a preset initial initial pressure P ₁ corresponding duty (duty) for a predetermined time (e.g., 2 seconds), P = 8 thereafter a predetermined time (e.g. 1 second). Sweep to a duty equivalent to 8 kgf / cm ² (almost 100% duty).

Then, the difference between the pressure P applied to the hydraulic pistons 141 and 142 and the predetermined pressure (the pressure close to the maximum pressure), which changes accordingly, is integrated. That is, FIGS. 23 (b) and 23 (c)
As shown in FIG.
₀ (or the time point 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 thereto) is reached. The area S (S1, S2) of the portion (hatched in the figure) surrounded by the curve L1 or L2 depicting the change state is calculated.

[0209] 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.

[0210] Therefore, if the area S is greater than the threshold value S _0, the process proceeds to step h8, PINTG (n) = PINTG (n-1) + β PINT (n) = PINTG (n-1) + β = PINTG (n That is, for the initial learning value PINTG (n), the previous initial learning value PINTG (n−1) is β
(= Pressure for 1 bit), and the initial pressure PINT (n) is the value obtained by adding β (= pressure for 1 bit) to the previous initial learning value PINTG (n-1), that is, , This time initial learning value PINT
Set to G (n).

[0211] On the other hand, 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, initial The learning value PINTG (n) is obtained by adding β to the previous initial learning value PINTG (n−1).
(= 1 bit), but the initial pressure PINT (n) is set to the previous initial learning value PINTG (n-1).

Even during the execution of the third preload learning,
As soon as any of the preload learning conditions ′ is not satisfied, the preload learning is interrupted and the mode returns to the normal mode. Also, in this case, the process is continued as long as the preload learning condition 'is satisfied. Through such first to third preload learning, an appropriate initial pressure Pi can be set for each, 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 times of learning, but there is an advantage that the setting accuracy is high and a response improving effect is great. In particular, in the determination based on the integral value (area), the determination as to whether the initial pressure Pi is appropriate can be made relatively appropriately, and the initial pressure Pi can be set without largely depending on the capability of the pressure sensor.

Furthermore, the duty solenoid valve and the clutch plate are protected by the control performed through the switch 294a, which contributes to the improvement of the reliability and durability of the device. Further, a torque distribution display unit 312 is provided in the meter cluster to graphically display the torque distribution state to the front wheels (or the rear wheels) (or to display the meters), so that the driver can recognize the torque distribution state of the vehicle. While driving can be performed effectively, the driving can be made more enjoyable, and the merchantability is greatly improved.

Further, it is conceivable that the result of the torque distribution estimation performed at this time is used by feedback to the control of each unit. In this embodiment, the hydraulically-adjustable front and rear wheel driving force distribution control device has been described, but the driving force distribution control device according to the present invention is also applicable to driving force distribution control for left and right wheels, including clutch protection control. It is widely applicable.

[0216]

As described above in detail, according to the hydraulically-adjustable driving force distribution control device for a vehicle of the present invention, the driving force distribution control device for a vehicle that controls the distribution of the driving force of the engine in the vehicle to each driving wheel. , An oil chamber provided in a rotating part, a hydraulic piston built in the oil chamber, driving force distribution adjusting means driven by the hydraulic piston to adjust a driving force distribution state of the engine, and Hydraulic pressure adjusting means for adjusting the hydraulic pressure in the chamber, and hydraulic pressure in the oil chamber are set such that the driving force distribution adjusting means performs a required driving force distribution state, and the hydraulic pressure adjusting means is controlled based on the set hydraulic pressure. A centrifugal oil pressure correction unit that compensates for the centrifugal oil pressure generated in accordance with the rotation speed of the oil chamber when setting the oil pressure. The driving force distribution state can be adjusted with high precision by the oil pressure in the oil chamber provided in the rotating part, and the control of the driving force distribution between the front and rear wheels and the control of the driving force distribution between the left and right wheels can be finely controlled. Become so.
As a result, it is possible to positively control the distribution of the driving force of each wheel in a four-wheel drive vehicle, or to positively control the distribution of the driving force of the left and right driving wheels in a two-wheel drive vehicle, thereby turning the vehicle. It can greatly contribute to the improvement of driving performance and straight running performance.

[Brief description of the drawings]

FIG. 1 is a block diagram showing a configuration of a main part of a hydraulically-adjustable driving force distribution control device for a vehicle as one embodiment of the present invention.

FIG. 2 is an overall configuration diagram of a drive torque transmission system of a hydraulically-adjustable vehicle drive force distribution control device as one embodiment of the present invention.

FIG. 3 is a cross-sectional view illustrating a main part of a drive torque transmission system of the hydraulically-adjustable drive power distribution control device for a vehicle as one embodiment of the present invention.

FIG. 4 is a sectional view of a main part of a front and rear wheel torque distribution mechanism of a hydraulically-adjustable driving force distribution control device for a vehicle as one embodiment of the present invention.

FIG. 5 is a schematic circuit diagram of a hydraulic pressure supply system of a hydraulically-adjustable driving force distribution control device for a vehicle as one embodiment of the present invention.

FIG. 6 is a circuit diagram of a main part of a hydraulic pressure supply system of the hydraulically-adjustable driving force distribution control device for a vehicle as one embodiment of the present invention.

FIG. 7 is a diagram showing a characteristic of a hydraulic pressure setting duty of the hydraulically adjustable vehicle driving force distribution control device as one embodiment of the present invention.

FIG. 8 is a block diagram showing details of a steering angle data detecting means of the hydraulically-adjustable driving force distribution control device for a vehicle as one embodiment of the present invention.

FIG. 9 is a block diagram showing details of a vehicle speed detecting means of the hydraulically-adjustable driving force distribution control device for a vehicle as one embodiment of the present invention.

FIG. 10 is a diagram showing an ideal rotational speed difference setting map of the hydraulically-adjustable driving force distribution control device for a vehicle as one embodiment of the present invention.

FIG. 11 is a diagram showing a lateral acceleration gain setting map of the hydraulically-adjustable vehicle driving force distribution control device as one embodiment of the present invention.

FIG. 12 is a plan view schematically showing wheel states for explaining an ideal rotational speed difference of the hydraulically-adjustable driving force distribution control device for a vehicle as one embodiment of the present invention.

FIG. 13 is a diagram showing a differential-corresponding clutch torque setting map of the hydraulically-adjustable driving force distribution control device for a vehicle as one embodiment of the present invention.

FIG. 14 is a block diagram showing a wheel slip corresponding clutch torque setting means of the hydraulically-adjustable vehicle driving force distribution control device as one embodiment of the present invention.

FIG. 15 is a map for setting a clutch torque corresponding to a wheel slip in the hydraulically-adjustable driving force distribution control device for a vehicle as one embodiment of the present invention.

FIG. 16 is a diagram showing an example of an engine torque map of the hydraulically-adjustable driving force distribution control device for a vehicle as one embodiment of the present invention.

FIG. 17 is a diagram illustrating an example of a transmission torque ratio map of the hydraulically-adjustable vehicle driving force distribution control device as one embodiment of the present invention.

FIG. 18 is a block diagram showing a modified example of the engine torque proportional clutch torque setting means of the hydraulically adjustable vehicle driving force distribution control device as one embodiment of the present invention.

FIG. 19 is a center differential input torque setting map of the hydraulically-adjustable vehicle driving force distribution control device as one embodiment of the present invention.

FIG. 20 is a characteristic diagram of the clutch torque for protection control of the hydraulically-adjustable driving force distribution control device for a vehicle as one embodiment of the present invention.

21A and 21B are diagrams illustrating a first preload learning of a hydraulically-adjustable driving force distribution control device for a vehicle as one embodiment of the present invention, wherein FIG. 21A shows a duty characteristic thereof, and FIG. Show characteristics.

FIG. 22 is a diagram showing a pressure characteristic concerning a second preload learning of a hydraulically-adjustable driving force distribution control device for a vehicle as one embodiment of the present invention.

23A and 23B are diagrams illustrating a third preload learning of the hydraulically-adjustable driving force distribution control device for a vehicle as one embodiment of the present invention, wherein FIG.
(C) shows the pressure characteristics.

FIG. 24 is a diagram showing torque distribution state display means of the hydraulically-adjustable driving force distribution control device for a vehicle as one embodiment of the present invention.

FIG. 25 is a characteristic diagram for explaining torque distribution by torque distribution state estimating means of the hydraulically-adjustable driving force distribution control device for a vehicle as one embodiment of the present invention.

FIG. 26 is a flowchart showing a control flow of the whole vehicle including the hydraulically-adjustable driving force distribution control device for a vehicle as one embodiment of the present invention.

FIG. 27 is a flowchart showing the flow of front and rear wheel torque distribution control of the hydraulically-adjustable driving force distribution control device for a vehicle as one embodiment of the present invention.

FIG. 28 is a flowchart showing a flow of setting a differential corresponding clutch torque in the hydraulically-adjustable driving force distribution control device for a vehicle as one embodiment of the present invention.

FIG. 29 is a flowchart showing a flow of setting a clutch torque corresponding to a wheel slip in the hydraulically-adjustable driving force distribution control device for a vehicle as one embodiment of the present invention.

FIG. 30 is a flowchart showing a flow of setting an engine torque proportional clutch torque in the hydraulically-adjustable driving force distribution control device for a vehicle as one embodiment of the present invention.

FIG. 31 is a flowchart showing a flow of setting a protection control clutch torque of the hydraulically-adjustable driving force distribution control device for a vehicle as one embodiment of the present invention.

FIG. 32 is a flowchart showing a flow of first preload learning of a hydraulically-adjustable driving force distribution control device for a vehicle as one embodiment of the present invention.

FIG. 33 is a flowchart showing a flow of a second preload learning of the hydraulically-adjustable driving force distribution control device for a vehicle as one embodiment of the present invention.

FIG. 34 is a flowchart illustrating a flow of third preload learning of the hydraulically-adjustable driving force distribution control device for a vehicle as one embodiment of the present invention.

[Explanation of symbols]

Reference Signs List 2 engine 4 torque converter 6 automatic transmission 8 output shaft 10 intermediate gear (transfer idler gear) 12 center differential (center differential) 14 front wheel differential gear device (front differential) 15 bevel gear mechanism 15A bevel gear shaft 15a bevel gear 16,18 front wheel 17L, 17R Front wheel axle 19 Reduction gear mechanism 19a Output gear 20 Propeller shaft 21 Bevel gear mechanism 22 Differential gear device for rear wheel (rear differential) 24, 26 Rear wheel 25L, 25R Rear wheel axle 27 Front wheel output shaft 27a Hollow Shaft member 28 Hydraulic multi-plate clutch (differential limiting mechanism or differential adjusting mechanism) 28a Front wheel output side disk plate 28b Input side disk plate 29 Rear wheel output shaft 30, 30a, 30b, 30c Handle angle sensor (steering wheel angle) 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 indicator light 54 Hydraulic source 56 Pressure control valve system (pressure control valve) 58 Pump 60 Check valve 62 Pressure control valve 64 Reducing valve 66 Accumulator 68 Pressure switch 68a Motor 113 Input gear 114a to 114f Bearing 115 Transmission case 115a End cover 115b Retainer 116 Supporting members 117a, 117b Oil passage 121 Sun gear 122 Planetary pinion (planetary gear) 123 Ring gear 124 Input Case 125 Planet carrier 125a Base plate portion 125b Planetary pinion housing portion 125f Clutch disk support portion 126 Pinion shaft 130 Connecting member 141 First piston 142 Second piston 143 Partition plate 144a First oil chamber 144b Second oil chamber 145 Hollow shaft 145a Piston housing Section 160 Shift lever position sensor (shift range detecting means) 160A Shift lever of automatic transmission 161 4WD control valve 162 Duty solenoid valve (duty valve) 163 Orifice 164 Oil filter 165 Reducing valve 170 Engine speed sensor 180 Transmission speed Sensor 200 Front / rear wheel actual rotational speed difference detection unit 202a-202d Filter 204a Front wheel vehicle Rotation speed data calculation unit 204b Rear wheel rotation speed data calculation unit 206 Front and rear wheel actual rotation speed difference calculation unit 210 Front and rear wheel ideal rotation speed difference setting unit 212 Driver required steering angle calculation unit as steering angle data detection means (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 Estimated vehicle speed calculation section (pseudo-vehicle speed data detection means) Vehicle speed calculation unit) 216a Wheel speed selection unit 216c Estimated vehicle speed calculation unit 216d Filter 218 Ideal rotation speed difference setting unit as ideal operation state setting unit 220 Differential clutch torque setting unit 222 Subtractor 230 Protection control unit 242 Filter 244 Horizontal G gain setting unit 246 Correction unit 254 Wheel slip Switch torque setting means 254a slip amount calculating section 254b clutch torque setting section 264 engine torque detecting section 266 torque converter torque ratio detecting section 267 center differential input torque calculating section 268 engine torque proportional torque setting section 269 clutch torque calculating section 270 engine torque proportional clutch torque calculating section 272 Turning correction unit 274a Switch 274 Judging means 275 Rotational difference gain setting unit 276 Transmission reduction ratio detection unit 280 Maximum value selection unit 282 Torque-pressure conversion unit 285 Centrifugal oil pressure correction unit 286 Centrifugal correction pressure setting unit 288 Initial engagement pressure setting Section (preload setting section) 290 Peak hold filter 292a, 294a Switch 292, 294 Judging means 295 Duty setting section 296 Pressure feedback correction section 298 Pressure-duty Tee conversion unit 300 Hydraulic circuit 302 Duty solenoid 304 Pressure sensor 306 Filter 310 Torque estimation unit 310a Calculation unit 310b Selection unit 312 Torque distribution display unit AM Mechanical part of hydraulically adjustable vehicle driving force distribution control device

 ──────────────────────────────────────────────────続 き Continuation of the front page (72) Inventor Masayoshi Ito 5-33-8 Shiba, Minato-ku, Tokyo Inside Mitsubishi Motors Corporation (56) References JP-A-62-286840 (JP, A) JP-A-62-286840 63-235728 (JP, A)

Claims (1)

(57) [Claims] 1. A vehicular driving force distribution control device for controlling distribution of driving force of an engine in a vehicle to respective driving wheels,
An oil chamber provided in a rotating part, a hydraulic piston incorporated in the oil chamber, driving force distribution adjusting means driven by the hydraulic piston to adjust a driving force distribution state of the engine; And a control for setting the oil pressure in the oil chamber so that the driving force distribution adjusting means performs a required driving force distribution state, and controlling the oil pressure adjusting means based on the set oil pressure. And a centrifugal oil pressure correction unit configured to cancel and correct a centrifugal oil pressure component generated in accordance with the rotation speed of the oil chamber when setting the oil pressure. , Hydraulic adjustment type driving force distribution control device for vehicles.