JPH04201735A - Differentially adjusting front/rear wheel torque distribution control device - Google Patents

Abstract

PURPOSE:To obtain a turn smooth and quick matched with a will of a driver by setting an ideal rotational speed difference between front/rear wheels based on steering angle/car speed data and friction torque of a front/rear wheel differentially adjusting mechanism in accordance with a differential rotational speed deviation. CONSTITUTION:A planet gear type differential gear 12, to which an output of an output shaft 8 of an automatic transmission 6 is transmitted, has a hydraulic multidisk clutch for changing distribution of output torque of an engine between front and rear wheels. Here in a controller 48, an ideal rotational speed difference between the front/rear wheels is set based on a steering angle data and a car speed data. In accordance with a difference between an actual rotational speed difference and an ideal rotational speed difference so that the actual rotational speed difference between the front/rear wheels approaches the ideal rotational speed difference, a value of friction torque of the clutch 28 is set such that the larger increased is this difference the larger the value of the friction torque is increased. A pressure control valve 56, that is, the multidisk clutch 28 is controlled so as to obtain this set friction torque.

Description

[Detailed Description of the Invention] [Industrial Application Field] The present invention relates to a four-wheel drive vehicle, and in particular, the present invention relates to a four-wheel drive vehicle, and in particular, it is possible to adjust the differential state between the front wheels and the rear wheels. The present invention relates to a differentially adjustable front and rear wheel torque distribution control device that controls torque distribution.

[Prior Art] Various power transmission devices for automobiles are known that are configured to control the ratio of torque transmitted to the front wheels and torque transmitted to the rear wheels depending on driving conditions.

In particular, as a control during four-wheel spin, a method has been proposed for controlling the torque distribution between the front and rear wheels based on the lateral acceleration acting on the vehicle body and the vehicle speed.

[Problems to be Solved by the Invention] By the way, controlling the torque distribution between the front and rear wheels using lateral acceleration and vehicle speed as described above does not allow for optimal torque distribution in which the driver's will is appropriately reflected in the control. Condition is difficult to obtain.

The present invention was devised in view of the above-mentioned problems, and is a differentially adjustable front and rear wheel torque system that can appropriately control torque distribution between the front and rear wheels so as to realize driving conditions that reflect the driver's will. The purpose of the present invention is to provide a distribution control device.

[Means for Solving the Problems] Therefore, the differentially adjustable front and rear wheel torque distribution control device of the present invention adjusts the differential state between the front wheels and the rear wheels to tl! In front and rear differential adjustable four-wheel drive vehicles, the difference between the front and rear wheels is reduced by applying frictional force between the front and rear wheels. a friction type front and rear axle differential adjustment mechanism for adjusting the dynamic state; a steering angle data detection means for detecting steering angle data; a vehicle speed data detection means for detecting vehicle speed data; and a steering angle detected by the steering angle data detection means. ideal rotational speed difference setting means for setting an ideal rotational speed difference between the front wheels and the rear wheels based on the angle data and the vehicle speed data detected by the vehicle speed data detection means;
Actual rotational speed difference detection means for detecting an actual rotational speed difference between the front wheels and the rear wheels.

The friction torque of the friction type front and rear wheel differential adjustment mechanism is adjusted according to the difference between the actual rotation speed difference and the ideal rotation speed difference so that the actual rotation speed difference between the front and rear wheels approaches the ideal rotation speed difference. and a control means for controlling a friction type front and rear wheel differential adjustment mechanism so as to obtain the friction torque set by the friction torque setting means, and the friction torque setting means includes: The present invention is characterized in that the friction torque value is set to a larger value as the difference between the actual rotational speed difference and the ideal rotational speed difference increases.

[Function] In the differentially adjustable front and rear wheel torque distribution control device of the present invention described above, the ideal rotational speed difference setting means uses the steering angle data detected by the steering angle data detection means and the vehicle speed data detected by the vehicle speed detection means. The ideal rotational speed difference between the front wheels and the rear wheels is set based on the vehicle speed data, and the friction torque setting means sets the actual rotational speed difference between the front and rear wheels detected by the actual rotational speed difference detection means to the ideal rotational speed difference between the front and rear wheels. According to the difference between the actual rotation speed difference and the ideal rotation speed difference, the friction type front and rear wheel differential adjustment mechanism approaches the rotation speed difference as the difference between the actual rotation speed difference and the ideal rotation speed difference increases. Set the friction torque value to a large value. The control means controls the friction type front and rear wheel differential adjustment mechanism so that the set friction torque is obtained.

[Embodiment] Hereinafter, a differentially adjustable front and rear wheel torque distribution control device as an embodiment of the present invention will be explained with reference to the drawings.
Figure 2 is a block diagram showing the configuration of the main parts, Figure 2 is an overall configuration diagram of the drive torque transmission system, Figure 3 is a sectional view showing the main parts of the drive torque transmission system, and Figure 4 is the front and rear wheels. 5 is a schematic circuit diagram of its hydraulic supply system; FIG. 6 is a circuit diagram of its hydraulic supply system; FIG. 7 is a sectional view of the main parts of the torque distribution mechanism.
FIG. 8 is a diagram showing the characteristics of the oil pressure setting duty, FIG. 8 is a block diagram showing details of the steering angle data detection means,
Fig. 9 is a block diagram showing details of the vehicle speed detection means, Fig. 1o is a diagram showing a map for setting the ideal rotation speed difference,
FIG. 11 is a diagram showing the lateral acceleration gain setting map, and FIG. 12(a).

13(b) is a plan view schematically showing the wheel condition to explain the ideal rotational speed difference, and FIGS. 13(a) and 13(b)
14 is a block diagram showing the clutch torque setting means corresponding to the longitudinal acceleration, FIG. 15 is a map for setting the flannel torque corresponding to the longitudinal acceleration, and FIG. 16 shows the map for setting the clutch torque corresponding to the differential. 17 is a diagram showing an example of the engine torque map, FIG. 17 is a diagram showing an example of the transmission torque ratio map, FIG. 18 is a block diagram showing a modification of the engine torque proportional clutch torque setting means, and FIG. 19 is a diagram showing an example of the transmission torque ratio map. The center differential input torque setting map, Fig. 20 is a characteristic diagram of the clutch torque for protection control, Fig. 21 (a) is a diagram showing the duty characteristic related to the first preload learning, and Fig. 21 (b) is the characteristic diagram of the clutch torque for protection control. Figure 22 is a diagram showing the pressure characteristics related to the first preload learning, Figure 23 (a) is the duty characteristic related to the third preload learning. Figures 23(b) and 23(c) are diagrams showing the pressure characteristics applied to the third preload learning, Figure 24 is a diagram showing the torque distribution state display means, and Figure 25 is the diagram showing the pressure characteristics related to the third preload learning. A characteristic diagram for explaining torque distribution by the torque distribution state estimating means, FIG. 26 is a flowchart showing the flow of control of the entire vehicle including the device, and FIG. 27 is a flowchart showing the flow of front and rear wheel torque distribution control. , Fig. 28 is a flow chart showing the flow of setting the clutch torque corresponding to the differential, Fig. 29 is a flow chart showing the setting flow of the clutch torque corresponding to longitudinal acceleration, and Fig. 30 is a flow chart showing the setting flow of the clutch torque corresponding to the longitudinal acceleration. FIG. 31 is a flowchart showing the flow of setting the clutch torque for protection control; FIG. 32 is a flowchart showing the flow of the first preload learning;
FIG. 3 is a flowchart showing the flow of the second preload learning, and FIG. 34 is a flowchart showing the flow of the third preload learning.

First, the overall configuration of the drive system of a vehicle equipped with this differentially adjustable front and rear wheel torque distribution control device will be described with reference to FIG.

In FIG. 2, reference numeral 2 denotes an engine, and 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 transmitted via an intermediate gear 10 to a planetary gear type differential device 12, which serves as an actuating device that distributes engine torque between front wheels and rear wheels as required.

The output of this planetary gear type differential 12 is transmitted on the one hand to a reduction gear mechanism 19. The transmission is transmitted from the axles 17L, 17R to the left and right front wheels 16.18 via the front wheel differential gear mechanism 14, and the bevel gear mechanism 15. Propeller shaft 20 and bevel gear mechanism 21. Differential gear device 2 for rear wheels
2 from the axles 25L and 25R to the left and right rear wheels 24.2
6. The planetary gear type differential device 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, as in the conventionally known one. The output of the output shaft 8 of the automatic transmission 6 is input to the carrier 125 supporting the planetary gear 122, and the sun gear 1
21 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 12
3 is interlocked with a propeller shaft 20 via a rear wheel output shaft 29 and a bevel gear mechanism 15.

Furthermore, the planetary gear type differential device 14 controls the distribution of engine output torque between the front wheels and the rear wheels by restraining (or limiting) the differential between the front wheel side output section and the rear wheel side output section. A hydraulic multi-disc clutch 28 is provided as a changeable differential limiting or adjusting means.・In other words, the hydraulic multi-disc clutch 28 is
121 (or ring gear 123) and carrier 125, and the frictional force changes depending on the control pressure applied to its own hydraulic chamber, and the difference between sun gear 121 (or ring gear 123) and carrier 12S It is designed to restrict movement.

Therefore, the planetary gear type differential device 12 appropriately controls the hydraulic multi-disc clutch 28 from a completely free state to a locked state, thereby transmitting torque to the front wheel side and the rear wheel side. is about 32: 68 to 5
It can be controlled between 0 and 50. Front wheel: rear wheel value in completely free condition: approx. 32:6
8 can be defined by setting the tooth number ratio of the input gears on the front wheel side and the rear shaft side of the planetary gear, and here, when the pressure in the hydraulic chamber of the hydraulic multi-disc clutch 28 is zero and it is in a completely free state, it is about 32 :68. Also,
This ratio in a completely free state (approximately 32=68) changes depending on the load balance between the front wheel system and the rear wheel system, etc., but usually takes such a value. Also, the pressure in the hydraulic chamber is set pressure (
9kg/cd) and the hydraulic multi-disc clutch 28 is in the locked state, and the differential limit becomes substantially zero, the torque distribution between the front wheels and the rear wheels becomes 50:50, resulting in a direct connection state. .

Further, reference numeral 30 is a steering wheel angle sensor that detects the rotation angle of the steering wheel 32 from the neutral position, that is, the steering wheel angle θ, and 34a and 34b are lateral accelerations G yf and G yr that act on the front and rear parts of the vehicle body, respectively.
In this example, the two detection data G yf and G yr are averaged to obtain lateral acceleration data, but only one lateral acceleration sensor 34 is provided near the center of gravity of the vehicle body, This detected value may be used as lateral acceleration data. 36 is a longitudinal acceleration sensor that detects the longitudinal acceleration Gx acting on the vehicle body; 38 is a throttle position sensor that detects the throttle opening θt of the engine 2; 39 is an engine key switch of the engine 2;
0.42, 44, and 46 are respectively 16 left front wheels and 1 right front wheel.
8. A wheel speed sensor that detects the rotational speed of the left rear wheel 26 and right rear wheel 28, and the outputs of these switches and each sensor are input to the controller 48. #- 50 is an anti-lock brake device. , this anti-lock brake device 50 is a brake switch 50A.
It operates in conjunction with. That is, when the brake switch 50A is turned on when the brake pedal 51 is depressed, an anti-lock brake activation signal is output in conjunction with this, and the anti-lock brake device 50 is activated. Further, when the anti-lock brake activation signal is output, a signal indicating the state thereof is simultaneously input to the controller 48. In addition, 52 is the controller 4
This is a warning light that lights up based on the control signal No. 8.

Note that the controller 48 includes a CPU, ROM, RAM, interface, etc., which are not shown but are necessary for control described later.

Reference numeral 54 indicates a hydraulic power source, and 56 indicates a controller 48 which is interposed between the hydraulic power source 54 and the hydraulic chamber of the hydraulic multi-disc clutch 28.
A pressure control valve system (hereinafter referred to as
(abbreviated as pressure control valve).

Additionally, this car is equipped with an automatic transmission.
Reference numeral 160 is a shift lever position sensor (shift range detecting means) for detecting the selected shift range of the shift lever 160A of the automatic transmission.

This detection information is also sent to the controller 48.

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

Note that details of the hydraulic system related to the hydraulic multi-disc clutch 28 will be described later.

In this example, a traction control system 151 is also provided. That is, the engine 2 is equipped with a main throttle valve 152 whose opening degree is controlled according to the amount of depression of the accelerator pedal 162, and together with the accelerator pedal 162 and the connection mechanism, constitutes an accelerator pedal-based engine output adjustment device. There is. A sub-throttle valve 153 as an engine output control means that is controlled independently of the accelerator pedal-based engine output adjustment device is provided in series with the main throttle valve 152 in the intake passage of the engine 2. This sub-throttle valve 153 is moved by a motor, and this motor is driven by the rear wheel speed sensor 4-4.
．． 46, front wheel speed sensor 40.degree. 42, engine rotation speed sensor 170, engine load sensor 172, etc. The drive is controlled based on the detection results.

Furthermore, as a sensor, the piston 1 of the clutch 28
Oil pressure sensor 304 that detects the oil pressure applied to 41 and 142
is provided at a predetermined location.

To explain in more detail the mechanical part AM of the differentially adjustable front and rear wheel torque distribution control device, this part is an input part to which the driving force of the engine is input through the automatic transmission 6, as shown in FIGS. 3 and 4. , a center differential (center differential) 12, a differential limiting mechanism 28, and output sections to the front and rear wheels.

The input section consists of an input gear 113 that meshes with the intermediate shaft 10A side, and an input case 124 to which this input gear 113 is coupled with serrations.
5a and the spacer member 115b, bearings 114a, 11
It is rotatably mounted via 4b. In addition, the input case 124 has a shape whose diameter increases toward the front (toward the left in FIGS. 3 and 4), including an expanded diameter section that houses the planetary gear element and a fourth diameter section behind this expanded diameter section. In the figure, a reduced diameter part is formed on the right side), and an opening is formed in front of the enlarged diameter part. It can be attached to the rear wheel output shaft 29 from behind (rightward in FIGS. 3 and 4), which will be described later. Further, a plurality of grooves 124d are formed on the outer periphery of the opening.

The center differential 12 is a planetary gear type center differential using a planetary gear mechanism, and includes a sun gear 121 and a sun gear 1.
21 includes a plurality of planetary pinions (planetary gears) 122 arranged outwardly to surround the sun gear 121, a ring gear 123 arranged around the planetary pinions 122, and a planetary carrier 125 supporting the planetary pinions 122. Each gear is composed of a straight tooth gear.

Of these, the sun gear 121 is integrally provided with a hollow shaft member 27a, and both the hollow shaft member 27a and the front wheel output shaft 27 are connected to a hollow shaft 145 through serrations. , hollow shaft member 27a
The front wheel output shaft 27 and the front wheel output shaft 27 are configured to rotate together. Note that a piston accommodating portion 145a, which will be described later, is formed in the hollow shaft 145.

Further, the ring gear 123 is fixed to a connecting member 130, and the connecting member 130 is connected to the rear wheel output shaft 29 through serrations, so that the ring gear 123 is connected to the rear wheel output section. As a result, the output of the ring gear 123 is transmitted to the connecting member 130. Rear wheel output shaft 29. The signal is input to the propeller shaft 20 via the bevel gear mechanism 15.

The planet carrier 125 has a convex portion 1251 formed on its outer circumference that can fit into each groove 124d of the input case 124, and by these fittings, the input case 12
4 and are connected to rotate together. Further, the sun gear 121 is connected to the front wheel output section, and the ring gear 123
is connected to the rear wheel output section.

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

A plurality of planetary pinions 122 are provided between the sun gear 121 and the ring gear 123, and each of these planetary pinions 122 is connected to the planet carrier 1 via a pinion shaft 126.
It is installed on 25.

The planet carrier 125 is coupled to the input case 124 so as to rotate together with the input case 124, and includes a brim-shaped base plate portion 125a, a planetary pinion accommodating portion 125b formed in front of this, and a cylindrical pinion accommodating portion 125b formed in the rear. A clutch disk support portion 125f is provided.

And each of these members 121, 122, 123.12
5 can be assembled individually as a planetary gear mechanism unit 12 in advance, and after being formed into a subassembly in this way, the planetary gear mechanism unit 12 can be installed in the transmission case 115.

Moreover, the above-described input case 124 is installed in the case 115 so as to cover the planetary gear mechanism unit 12 .

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

Among these, the front export force side disk plate 28a is the first
It is driven by a piston 141 and a second piston 142, and can be joined to the input side disk plate 28b.

Note that the first piston 141 and the second piston 142 are
Piston housing portion 145 formed on the outer periphery of hollow shaft 145
A partition plate 143 is interposed between these two pistons 141 and 142 so as to be movable in the axial direction. .

The first piston 141 and the piston accommodating portion 145a
A first oil chamber 144a is provided between the second piston 142 and the partition plate 143, and a second oil chamber 144a is provided between the second piston 142 and the partition plate 143.
44b is provided, and these oil chambers 144a, 14
4b, appropriate hydraulic pressure is supplied from a hydraulic pressure supply system (not shown) through an oil passage 117a bored in the support member 116 fixed to the transmission case 115 side and an oil passage 117b bored in the hollow shaft 145. It is now possible to do so.

Each of these members 28a, 28b, 141, 142.14
3,145,146 are also preliminarily equipped with the differential limiting mechanism unit 28.
After being made into a subassembly in this way, the differential limiting mechanism unit 28 can be mounted inside the transmission case 115.

The output section includes a front wheel output section and a rear wheel output section. and an output gear 19a that meshes with an input gear 19b of a differential gear device (diff) 14 for the rear wheels.
A rear wheel output shaft 29 is provided to pass through the front wheel output shaft 27, a bevel gear shaft 15A is connected to the tip of the rear wheel output shaft 29, and a propeller is attached to the bevel gear shaft 15A. It is composed of a bevel gear 15b at the tip of the shaft 20 and a bevel gear 15a that meshes with the bevel gear 15b.

The output gear 19a is supported on the transmission case 115 side via bearings 114c and 114d, and the bevel gear shaft 15A and the bevel gear 15a are supported by the bearing 1
Transmission case 1 via 14e and 114f
It is supported on the 15 side. Further, the output gear 19a and the input gear 19b constitute a reduction gear mechanism 19, and the bevel gear 15a and the bevel gear 15b constitute the bevel gear mechanism 1.
5 are configured.

In addition, in FIG. 3, 101 is a converter housing, and 10
2 is the oil pump, 103 is the front clutch, 104
is the kickdown brake, 105 is the rear clutch, 10
6 is a low reverse brake, 107 is a planetary gear set, 108 is a transfer drive gear, 109 is a rear cover, and 112 is an end clutch.

Further, in FIG. 4, 131 and 132 are plates that support each shaft in the axial direction, and 133 is an O-ring.

On the other hand, the hydraulic system related to the hydraulic multi-disc clutch 28 is configured as shown in FIG. 5 (schematic hydraulic circuit diagram) and FIG. 6 (principal hydraulic circuit diagram).

That is, as shown in FIG. 5, the reservoir is also used as that of the automatic transmission 6, and the pump 58 that sucks the oil in the reservoir 6 operates the check valve 60 and the pressure control valve 62 from its spout. It is connected to the hydraulic chamber of the hydraulic multi-disc clutch 28 via the hydraulic multi-disc clutch 28. The pressure control valve 62 is connected to the hydraulic multi-disc clutch 2
A first position in which the hydraulic chamber of No. 8 communicates with the pump 58 and a second position in which the hydraulic chamber communicates with the reservoir of the automatic transmission 6 can be taken.

A relief valve 64 is provided in the passage between the check valve 60 and the pressure control valve 62 and opens at a set pressure (for example, about 9 kg/d) to release oil to the reservoir of the automatic transmission 6. Accumulator 66 and pressure switch 68
is connected. The calibration signal of the pressure switch 68 is input to the controller 48. Note that the motor 58a that drives the pump 58 is controlled by a control signal from the controller 48.

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

In this FIG. 6, 161 is a 4WD control valve, and this 4WD control valve 161 is
The spool valve has two valves provided on the spool body 161a.
It has two valve body parts 161b and 161c. The spool main body 161a receives duty pressure (solenoid control pressure) Pd and reducing pressure Pr at both ends thereof, and when the duty pressure Pd decreases, it advances to the left in the figure and becomes open, and when the duty pressure Pd increases, it moves to the left in the figure. Proceed to the center right and it will become closed. In addition, 161d is
This is a spring that biases the spool body 161a in an appropriate direction so that the spool body 161a can move appropriately as described above.

162 is a duty solenoid valve (duty valve), and this duty valve 162 is provided with a solenoid 162a and a valve body 162b driven by this solenoid 162a and a return spring 162c.

When the solenoid 162a is activated, it moves backward to open the oil passage 169f, and when the solenoid 162a is not activated, it moves forward by a return spring 162c to close the oil passage 169f. This reduction valve 162 is operated by a controller (based on information from various sensors).
It is electronically controlled by a computer) 48.

Also, 163 is an orifice, 164 is an oil filter,
165 is a reducing valve, and the orifice 16
3 is provided between the reducing valve 165 and the 4WD control valve 161, and the oil filter 164 is provided in an oil passage 169b flowing into the reducing valve 165.

In the reducing valve 165, a valve body 165a is biased with a predetermined pressure by a return spring 165b, and this biasing force causes the valve body 165a to be supplied with hydraulic pressure when the hydraulic pressure becomes lower than the set pressure. If the pressure exceeds the pressure, it will automatically move to drain the hydraulic pressure.

Therefore, for example, when the solenoid 162a operates and the duty valve 162 opens, the oil pressure (duty pressure) Pd on the left end side of the 4WD control valve 161 decreases, and the return spring 161d causes the valve body portion 16 to
1b and 161c move to the left, oil passage 169c
and 169g are opened, and the line pressure P works is operating hydraulic pressure (
4WD clutch pressure) P is supplied to each oil chamber 144a, 144b of the hydraulic multi-disc clutch 28,
A hydraulic multi-disc clutch 28 is configured to be connected.

Further, when the solenoid 162a is not operated and the duty valve 162 is closed, the oil pressure (duty pressure) Pd on the left end side of the 4WD control valve 161 increases, and the valve body parts 161b and 161c are moved to the right (sixth (position shown in the figure), oil passages 169c and 169g are disconnected, 4WD clutch pressure P4 is released, and hydraulic multi-disc clutch 28 is separated.

The relationship between the duty (Du, ty), which is a control index of the duty valve 162, and the 4WD clutch pressure P4 (=control oil pressure P) is, for example, as shown in FIG. When the 4WD clutch pressure P4 is low, the 4WD clutch pressure P4 becomes low, and the higher the duty, the higher the 4WD clutch pressure P4 becomes. In addition, the reverse setting,
In other words.

If the characteristic becomes a straight line downward to the right and the duty is low, it will be 4.
A configuration is also conceivable in which the WD clutch pressure P4 becomes higher and the higher the duty, the lower the 4WD clutch pressure P4 becomes.

Next, the components of the controller related to the control (hereinafter referred to as driving force distribution control or center differential control) for restraining the differential of the planetary gear type differential device 12 by the hydraulic multi-disc clutch 28 are as follows. This will be explained with reference to the block diagram in FIG.

In this control, each sensor (wheel speed sensor 40°42.4
4,46. Steering angle sensor 30 a + 30 b +3
0c, lateral acceleration sensor 342, longitudinal acceleration sensor 36, throttle position sensor 38. Engine speed sensor 170. ) - Lance transmission rotation speed sensor 180. The clutch torque of the hydraulic multi-disc clutch 28 is set based on the detection information from the shift position sensor 160, etc.), and the differential oil pressure of the hydraulic multi-disc clutch 28 is controlled so as to obtain the target clutch torque. ing.

Furthermore, among the data, ABS information, wheel speed, and steering angle.

Comprehensive communication between the gear stage, ABS control unit and engine control unit (SC, I communication: SCI)
=Serial Communication Int
Data such as erface) are input digitally, and hydraulic control is applied to longitudinal acceleration, lateral acceleration, accelerator opening, and multi-disc clutch.

4WD control unit control, current to the differential electromagnetic clutch, etc. are input in analog form.

In addition, the setting of the clutch torque of the hydraulic multi-disc clutch 28 is determined by focusing on the differential state between the front axle and the rear wheels (the difference in rotational speed, also expressed as the difference in rotational speed), and set the ideal differential state. Differential compatible clutch torque Tv to control the
■Clutch and Lux Tx for longitudinal acceleration to perform control corresponding to the longitudinal acceleration acting on the vehicle;■
Latch torque Ta, which is set in proportion to engine torque, is set in proportion to engine torque so that a large road surface transmission torque can be obtained in a four-wheel drive state where the front and rear axles are directly connected when starting suddenly.
■Preserves the clap part of the wet multi-disc clutch.

One of the protection control clutch torques T and pc for protection is selected, and the settings and running parts of each of these clutch torques Tv, Tx, Ta, Tpc will be explained do.

The differential compatible clutch torque Tv is the clutch torque that makes the vehicle behave in accordance with the will of the wheel driver when turning.In order to control the attitude of the vehicle, the rear wheels are used as the drive base and the slip from the rear wheels is used. Since it is effective to set the differential clutch torque Tv to achieve such a state, the differential clutch torque Tv is set to realize such a state.

Therefore, as shown in FIG. 1, the part involved in setting the clutch torque Tv for the differential is a front and rear axle actual rotation speed difference detection section 200, a front and rear axle ideal rotation speed difference setting section 210, and a front and rear axle ideal rotation speed difference setting section 210. Speed difference ΔV, cd and ideal rotational speed difference Δ of the front and rear axes
The differential clutch torque setting section 220 and 1 is configured to set clutch torque Tv' from Vhc, and a compensating section 246 is configured to correct acceleration of this clutch torque Tv'.

The front and rear axis actual rotational speed difference detection unit 2qO includes a filter 202a.
~202d, and front wheel rotational speed data calculation unit 204a
1, a rear axle rotation speed data calculation section 204b, and a front and rear wheel actual rotational speed difference calculation section 206.

Filter 202. ;~202d are wheel speed sensors 4.0. , 42, 44. The left front wheel detected by 46. Right front wheel 18. Left rear wheel 26. This is to remove micro-vibration components of the data caused by disturbances etc. from the rotational speed data signals FL, FR, RL, and RR of the right rear wheel 28.

Further, the front wheel rotational speed data calculation unit 204a averages the respective wheel speeds of the front wheels determined from the front wheel rotational speed data signals PL, FR to obtain the front wheel rotational speed Vf, and the rear wheel rotational speed data calculation unit 204b Here, the rear wheel rotation speed Vr is obtained by averaging the respective wheel speeds of the rear wheels determined from the rear wheel rotation speed data signals RL and RR.

Further, the front and rear wheel actual rotational speed difference calculation unit 206 subtracts the front wheel rotational speed Vf from the rear wheel rotational speed Vr to calculate the actual rotational speed difference between the front and rear wheels [rotational speed difference between the front and rear wheels (front and rear rotational speed difference: this rotational difference is It is also the rotation difference at the center differential)] ΔV
Calculate cd.

The front and rear wheel ideal rotational speed difference setting section 210 includes a driver-required steering angle calculation section (pseudo steering angle calculation section) 212 as a steering angle data detection means, and a driver-required vehicle body speed calculation section (a pseudo steering angle calculation section) as a vehicle speed data detection means. It is configured to include a pseudo vehicle body speed calculation section) 216 and an ideal rotational speed difference setting section 218 as an ideal operating state setting section.

As shown in FIG. 8, the driver-required steering angle calculation unit 212 as driver-required steering angle data setting means includes a steering angle sensor 30 (first steering angle sensor 30a).

Second steering angle sensor 3 installed on the steering wheel
Based on the detection data 01+ 02+ fin from the neutral position sensor 3(, i (: ), the sensor corresponding steering angle δh, [=f (θ1.+Oz+ O
transverse acceleration data G by averaging the data G yf and G yr detected by the sensor-compatible steering angle data setting unit 212a and the lateral acceleration sensors 34a and 34b.

A lateral acceleration data calculation unit 212b that calculates the steering angle, a comparison unit 212G that compares the direction of the sensor-compatible steering angle δh with the direction of the lateral acceleration data Gy, and a driver-required steering angle according to the comparison result of the comparison unit 212o. The driver-required steering angle setting section (vehicle speed data setting section) 212d sets δref.

Note that the function δh=f (
θ0. θ2. θn) corresponds to the specifications of the steering wheel angle sensor.

Furthermore, the sensor-compatible steering angle δh and the lateral acceleration data Gy both assume, for example, the right turning direction as positive.

Steering angle δh and lateral acceleration data Gy corresponding to these sensors
To compare the directions, the following direction-related function 5IG(x) is set for the detection data X.

When x > O, SIG (x) = 1 When X = O, SIG (x) = O When x < O, 5IG (x) = -1 Therefore, the comparator 212c calculates the direction of the sensor-compatible steering angle δh. A comparison between the direction of the lateral acceleration data Gy and the direction of the lateral acceleration data Gy is made using
This is done by comparing with IG(Gy).

Then, when the direction 5IG (δh) of the sensor-compatible steering angle δh and the direction 5IG (Gy) of the lateral acceleration data Gy are equal, the driver-requested steering angle setting unit 212d requests the driver to set the sensor-compatible steering angle δh. Steering angle (steering angle data) δ
ref, and the direction 5IG of the sensor-compatible steering angle δh (
δh) and the direction 5IG(Gy) of the lateral acceleration data Gy are not equal, O is set to the driver requested steering angle δref.

If the direction SIL, 1 (δh) of the corresponding steering angle δh and the direction 5IG ((koy>) of the lateral acceleration data Gy are not equal, setting O as the driver requested steering angle δref is, for example, When performing a steering wheel operation such as Calacita Steering, the steering wheel position and the actual steering angle (turning state) of the vehicle may differ, and in such cases, if you set the steering angle of the vehicle from the steering wheel position. Difficult to perform appropriate control.

Therefore, in order to eliminate such a problem, if the direction SIG (δh) of the sensor corresponding steering angle δh and the direction 5IG (Gy) of the lateral acceleration data Gy are not equal, the driver requested steering angle is set to O. It is set.

As shown in FIG. 9, the driver-required vehicle body speed calculation unit 216 calculates the left front wheel 16. Right front wheel 18. Left rear wheel 26. Rotational speed data signal FL of the right rear wheel 28.

Wheel speed selection unit 216 that selects the second largest wheel speed data from the bottom (from the smallest) among PR, RL, and RR.
a, and a driver-required vehicle speed calculation unit 216c that sets the driver-required vehicle speed from the selected wheel speed data and the like.

In particular, the driver-required vehicle speed calculation unit 216c filters the wheel speed data selected by the wheel speed selection unit 216a through the filter 216.
Wheel speed data S obtained by removing noise components by
Based on vW and longitudinal acceleration data Gx obtained by filtering the longitudinal acceleration detected by the longitudinal acceleration sensor 36 to remove noise components, the subsequent vehicle speed is calculated from both data SV W and G x at a certain point in time. It is supposed to be estimated. In other words, if the wheel speed data SvW at a certain point is 2, and the longitudinal acceleration data Gx is a, then the theoretical vehicle speed Vref after a time t after this point is:
It can be calculated as Vref=V2+at.

Also, instead of longitudinal acceleration data Gx, wheel speed data S
A driver-required vehicle body acceleration V'ref obtained by time-differentiating VW or driver-required vehicle body speed Vref may be employed.

Note that the rotational speed data signals FL, FR, RL,.

The reason why we use the second largest wheel speed data from the bottom of the RR is because each wheel is usually slipping toward the over-rotation side, so normally it would be better to use the wheel speed that rotates at the lowest speed. Although it is desirable, the second wheel speed from the bottom is used in consideration of the reliability of the data.

The ideal rotational speed difference setting unit 218 then calculates the driver requested steering angle δr calculated by the driver requested steering angle calculation unit 212.
ef and the driver-required vehicle speed V ref calculated by the driver-required vehicle speed calculating section 216, an ideal rotational speed difference ΔVhc is set in accordance with a map as shown in FIG. In other words, when the vehicle speed is low, the difference in track radius between the front and rear axles during turning (the so-called inner ring difference) has a large effect.
The rotational speed Vr of the rear wheels is smaller than the rotational speed Vf of the front wheels, but as the vehicle speed increases, the rotational speed Vr of the rear wheels becomes larger than the rotational speed Vf of the front axle. This makes it easier for the rear wheels to slip. This ensures the responsiveness of the vehicle body posture required at high speeds. Also.

Regarding the steering angle, the larger the steering angle, the larger the rotation difference required between the front and rear wheels, so the larger the magnitude 1δref l of the steering angle data δref, the larger the value of ΔVhc.

The rotational speed difference ΔVhc between the front and rear axles due to the difference in orbit radius between the front and rear wheels will be explained with reference to FIGS. 12(a) and 12(b). Note that Fig. 12(a) is a schematic diagram of a two-wheeled vehicle consisting of one front wheel and one rear wheel, and Fig. 12(b) is a further schematic diagram of Fig. 12(a). be. As shown in FIGS. 12(a) and (b), the front wheel speed is set to Vf
, the rear wheel speed is Vr, the vehicle speed at the center of gravity of the vehicle is V, the rotation radius of the front axle is Rf, the rotation radius of the rear wheel is Rr, the rotation radius of the vehicle center of gravity is R1, the vehicle body slip angle is β, the wheel base is 1. If the distance between the front axle center and the center of gravity is 1f, and the distance between the rear wheel center and the center of gravity is lr, then the rotational speed difference ΔVhc between the front and rear axles is expressed as follows.

ΔVhc = Vr-Vf = [(Rr-Rf)/
R]・vref...(1,1) In addition, Rr=(R2+l r"-2Rl r-cos(π
/2-β))””Rf = (R”+l f”-2Rl
f-cos(7C/2+β))""β"(1-m/2
1・1f/1r-k r-V)/(1+A-V2)・l
r/l-δ where m is the vehicle weight, kr is the cornering power, and A is the stability factor.

Also, considering that the front wheel speed Vf and the rear wheel speed Vr are theoretical, Vf:Vr=Rf:Rr, Vf:V=R
f: R, and furthermore, the angle βf shown in FIG. 12(b)
, βr have the relationship βf−βr=AV”, and from these relationships and the above formulas, ΔVhc can be expressed as a function of V and δ [Δ
Vhc=fc(V.

δ)]. However, in this case, ■ corresponds to a theoretical value, that is, the driver-required vehicle speed Vref, and δ also corresponds to a theoretical value, that is, the driver-required steering angle δref. Such a function [ΔVhc=fc(V ref e
When δref ) is mapped, it becomes as shown in FIG.

By the way, regarding the steering angle, in addition to the actual steering angle (sensor-corresponding steering angle) δh based on the steering wheel angle θ, there is a turning G-equivalent steering angle δy obtained from the lateral acceleration (turning G) Gy during turning.

This turning G equivalent steering angle δy can be calculated using the following equation.

δy=[(1+A-Vref2)/Vref2]・l
- Gy... (1, 2) where A is a stability factor, Vref is a theoretical vehicle speed (vehicle speed requested by the driver), which will be described later, and ■ is a wheel height.

In contrast to the turning G-equivalent steering angle δy determined in this manner, the above-mentioned actual steering angle (sensor-corresponding steering angle) δh is a steering angle that more closely reflects the driver's intention. In other words, if the driver wants to make a larger turn than the current one, Iδh1>1δy1, and by adopting the steering angle value 1δh], the ideal rotational speed difference (slip target value) will be larger than by adopting the steering angle value 1δy1. On the other hand, if the driver wants to suppress the current turning, 1δh1<1δy1, and the steering angle value 1δh
By adopting I, the magnitude of the ideal rotational speed difference (slip target value) can be made smaller than by adopting the steering angle value 1δyl.

What is the front and rear wheel actual rotational speed difference ΔVc+1 detected by the front and rear wheel actual rotational speed difference detection unit 200 as described above and the front and rear wheel ideal rotational speed difference ΔVhc set by the front and rear wheel ideal rotational speed difference setting unit 210? , subtracted by the subtractor 222 (ΔVcd-
ΔVhc) and the obtained difference ΔVc (=ΔVcd−
ΔVhc) and the ideal rotational speed difference ΔVhc between the front and rear wheels are input to the differential compatible clutch torque setting unit 220 as data.

The clutch torque setting unit 220 for differential sets the clutch torque Tv' corresponding to the difference ΔVc (=ΔVcd−ΔVhc) between the actual rotational speed difference ΔVcd of the front and rear wheels and the ideal rotational speed difference ΔVhc of the front and rear wheels. Ideal shaft rotation speed difference ΔV
Clutch torque Tv is divided into cases depending on the sign of hc.
' is set.

(i) When ΔVhc≧O In this case, we want to make the speed of the rear axle faster than the speed of the front wheels, and as shown in (1 to ■) below, the clutch torque TV'
Set.

■If ΔVcd≧ΔVhc, the rear wheels are over-rotating and slipping, so it is desired to suppress rear wheel slipping by transferring part of the engine torque that was largely distributed toward the rear axle to the front wheels. Therefore, the clutch torque Tv' is the difference ΔV
so that it increases in proportion to the size of c(ΔVcd−ΔVhc).

Tv'=aX(ΔVcd-ΔVhc)=aXΔVc...
・Set as (1, 3) (where a is a constant of proportionality).

■If ΔVhc > ΔVcd > O, the front wheels are slipping, so if the clutch torque Tν' is increased at this time, the engine torque distributed to the front wheels will increase, promoting front wheel slipping. . For this reason, it is desirable to set the differential limit free and reduce the engine torque distributed to the front wheels.

Therefore, in this case, the clutch torque Tv' is set to 0 to set a so-called dead zone region.

■If O≧ΔVcd, the front wheels are slipping, so
I want to increase the distribution of engine torque to the front wheels to reduce front wheel slip. Therefore, the clutch torque Tv' increases in proportion to the magnitude of ΔVcd.

T v ′= −a XΔVcd= −a X (Δ
Set as VC+ΔVhc) (1, 4) (where a is a proportionality constant).

When we map the relationship between Tν' and ΔVc, we get
The map is as shown in FIG. 13(a), and the differential clutch torque Tv can be determined from the difference ΔVc and the ideal rotational speed difference ΔvhC between the front and rear wheels.

Note that when ΔVhc=O, there is no dead zone region where ΔVhc>ΔVcd>0.

(ii) When ΔVhc<O In this case, it is desired to make the speed of the front wheels faster than the speed of the rear wheels, and the clutch torque Tv' is set as shown in (1) to (2) below.

■If ΔVcd≧O, the rear wheels are over-rotating and slipping, so it is desired to suppress rear wheel slipping by transferring part of the engine torque that was largely distributed toward the rear axle to the front wheels. Therefore, in order to increase in proportion to the magnitude of clutch torque Tv' or ΔVcd, Tv'=aXΔVcd=aX(ΔVc+ΔvhC)...
・Set as (1, 5) (where a is a proportionality constant).

■If 0〉ΔVcd＞ΔVhc, the rear wheels are slipping, so if clutch torque Tv' is increased at this time, the engine torque distributed to the rear wheels will increase, promoting rear wheel slipping. It turns out. For this reason, it is desirable to set the differential limit free and reduce the engine torque distributed to the rear wheels.

Therefore, in this case, the clutch torque Tv' is set to 0 to set a so-called dead zone region.

■If ΔVhc≧ΔVcd, the front wheels are slipping, so it is desired to increase the distribution of engine torque to the front wheels to reduce the front wheel slip. Therefore, the clutch torque T
Tv'=-aX(ΔVcd-ΔVhc)=-a×ΔVc so that v' increases in proportion to the magnitude of ΔVc(ΔVcd-ΔVhc).
...Set as (1, 6) (where a is a proportionality constant).

When we map the relationship between Tv' and ΔVc, we get
The map is as shown in FIG. 13(b), and the differential clutch torque Tv can be determined from the difference ΔVc and the ideal rotational speed difference ΔvhC between the front and rear wheels.

In this way, the differential compatible clutch torque setting section 220
Then, with reference to the map [Fig. 13 (a), (b)], ΔV
The differential compatible clutch torque Tv' obtained from c and ΔVhc is corrected for lateral acceleration in a correction section 246.

The correction unit 246 performs lateral acceleration correction on the differential compatible clutch torque TV' by multiplying the lateral G gain by the
The clutch torque Tv for the differential is obtained, and this lateral G gain is set to 1 as follows.

In other words, the detection data ay from the lateral acceleration sensor 34 is
After removing micro-vibration components from the data caused by disturbance etc. through the filter 242, the lateral G gain setting section 244
It is now sent to This lateral G gain setting section 2
At step 44, the lateral G gain is set from the lateral acceleration data Gy according to the map shown in the block of the setting section 244 in FIG.

This lateral G gain k is intended to reflect the state of the friction coefficient μ of the road surface in the control, and it can be judged that the larger the lateral acceleration Gy, the larger the road surface μ, and the larger the road surface μ, the greater the engine torque. I would like to be able to prioritize the turning performance of the car body by mainly distributing it to the rear wheels. Therefore, when the magnitude of the road surface μ (therefore, the magnitude of the lateral acceleration ay) increases, a correction is made to reduce the lateral G gain and reduce the set clutch torque Tν.

Note that even if the road surface μ is large, if the turning performance of the vehicle body is not particularly prioritized, it may be possible to omit the correction by 1 to the lateral G gain.

The longitudinal acceleration responsive clutch torque Tx is a clutch torque that prevents strong understeer of the vehicle and allows the vehicle to perform smooth turning operations, and is controlled in response to the longitudinal acceleration Gx acting on the vehicle. It has become.

Setting of this longitudinal acceleration corresponding clutch torque Tx is performed by longitudinal acceleration corresponding clutch torque setting means 254.
After the detection data Gx from the longitudinal acceleration sensor 36 passes through the filter 252 and removes minute vibration components caused by disturbances, etc., the clutch torque setting means 254
It is now sent to

As shown in FIG. 14, the clutch torque setting means 254 includes a front wheel shared load calculation means 254a, a total output torque calculation means 254b, a front wheel shared torque calculation means 254c, and a clutch torque calculation means 254d. .

The front wheel shared load calculation means 254a calculates the front wheel shared load Wf' during acceleration from the longitudinal acceleration data Gx.
It is determined by the following formula from known values such as vehicle weight W a , center of gravity height, wheel base 1, etc., and longitudinal acceleration data Gx.

Wf'=Wf-(h/1)・Wa-Gx=7 (2,
1) The total output torque calculation means 254b calculates the required total output torque Ta (torque considered on the propeller shaft) from the longitudinal acceleration data GK, but the required total output torque Ta
are vehicle weight W a and tire radius Rt.

It is determined by the following formula from the final reduction ratio (average value at rear differential and front differential) ρ and longitudinal acceleration data Gx.

Ta=Wa Gx-Rt/p ”・(2,2
) The front wheel shared torque calculation means 254C calculates the front wheel shared load Wf' during acceleration obtained by the front wheel shared load calculation means 254a.
and the required total output torque Ta determined by the total output torque calculation means 254b, the front wheel shared torque Tf is determined by the following formula.

Tf= (Wf'/Wa) ・Ta...
(2, 3) The clutch torque calculation means 254d uses the required total output torque T calculated by the total output torque calculation means 254b.
From a and the front wheel shared torque Tf calculated by the front wheel shared torque calculating means 254C, the clutch torque corresponding to longitudinal acceleration Tx is calculated.
'Calculate.

In other words, the front wheel torque distribution factor f due to the center differential 12 and the differential limiting clutch 28 is set as follows assuming that rear slip occurs first.

TC[Zs/(Zs+Zr)]・Ta+[Zr/(Zs
+Zr)]・Tx'... (2, 4) However, Zs is the number of teeth of the sun gear 12a, and Zr is the number of teeth of the ring gear 12c.

Equations (2, 4) can be transformed as follows.

Tx'=Tf-[Zs/(Zs+Zr)iTa/[Z
r/(Zs+Zr)co...(2,4') Therefore, the required total output torque Ta and the front wheel shared torque T
From f, the clutch torque Tx' corresponding to longitudinal acceleration can be determined.

On the other hand, from equations (2,1) to (2,4), Wf', Tf,
Eliminating Ta and solving Tx' for Gx, first,
Substituting equations (2, 1) and (2, 2) into equation (2, 3), T f = (Rt/p) ・(Wf - Gx - h/
1-Wa-Gx2)...(2,5) From formulas (2,1), (2,4), (2,5), T
X'=-A-c(Gx-B/2C)2+AB/4C...
・(2,6) However, A=[(Zs+Zr)/Zr](Rt/ρ)
B=Wf EZ s/(Z s +Z r
) ko・WaC=”(h/l)・Wa Here, each constant related to constants A, B, and C is Zs=28.
Zr=60. Rt=0.296(m), p=3.6. W
f=880 (kg), Wa=1595 (kg), h=o
. 55 (m), 1 = 2.6 (m), Tx'=
-40,7(GX-0,552)2+12.4,
Tx' can be expressed as a quadratic curve with respect to Gx as shown in FIG.

However, Tx' reaches its maximum value at GxL: 0.55, and G
Tx' decreases in the region of x>0.55, but in consideration of control safety, Tx' is set to a constant equal to the maximum value even in the region of Gx>0.55. . Note that such settings are performed by the clutch torque calculation means 254d.
It can also be applied when calculating the clutch torque Tx' corresponding to the longitudinal acceleration.

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

Clutch torque T corresponding to longitudinal acceleration set in this way
x′ is corrected by the lateral acceleration corresponding correction unit 256.

The correction unit 256 performs the same correction as the above-mentioned correction unit 246, and performs lateral acceleration correction by multiplying the longitudinal acceleration corresponding clutch torque Tx′ by 1 on the lateral G cane to obtain the longitudinal acceleration corresponding clutch torque Tx. However,
The lateral G gain □ has been described above, and its purpose is to reflect the state of the friction coefficient μ of the road surface in the control, as described above, so its explanation will be omitted here.

The clutch torque Tx corresponding to the longitudinal acceleration corrected in this manner is outputted as data in accordance with whether the switch 258a is turned on or off. This switch 258a is turned on in response to a signal from the determining means 258 when the front wheel speed f is greater than the vehicle speed Vref, that is, when the front wheels are slipping (front slip), and is turned ON in other cases. is OF
It becomes F. Therefore, the clutch torque TI that corresponds to longitudinal acceleration is output only when the front slip is set, and is not output in other cases (in this case, it is referred to as TS20, and hereinafter, generally, the clutch torque TI is set when the switch is turned off and the clutch torque TI is not output). When the clutch torque is not output, the value of the clutch torque is set to 0).

The engine torque proportional clutch torque Ta is set in advance to the direct four-wheel drive state in order to prevent the initial slip of the rear wheels when the transmitted torque is expected to increase, such as when suddenly starting from a stopped state. This is the setting torque for setting.

Therefore, the part that sets this engine torque proportional clutch torque Ta (engine torque proportional clutch torque setting means) is an engine torque detection part 2 that detects the engine torque Te at a certain moment, as shown in the lower left part of FIG.
64, a torque converter torque ratio detection section 266 that detects the torque converter torque ratio t at that time, a transmission reduction ratio detection section 276 that detects the reduction ratio ρm of the transmission at that time, and a map set in a proportional relationship with the engine torque Te. An engine torque proportional torque setting section 268 that obtains the engine torque proportional torque Ta' from the engine torque Te based on the engine torque proportional torque T;
a' is the above-mentioned torque converter torque ratio tt and transmission reduction ratio ρm.

An engine torque proportional clutch torque calculation unit 270 multiplies the final deceleration ρ□ and the rotational difference gain by 2 to obtain the engine torque proportional clutch torque Ta, and calculates the set engine torque proportional clutch torque Ta at low speeds (for example, Vre
f<20 kn/h) and a switch 274a that outputs only data as data.

In the engine torque detection unit 264, throttle opening data θth sent from the throttle position sensor 38 and filtered through a filter 262a to remove minute vibration components caused by disturbances, etc., and throttle opening data θth sent from the engine rotation speed sensor 170 and filtered through a filter 262b. The engine torque Te at that time is determined from the engine rotational speed data Ne obtained by removing micro-vibration components caused by disturbances and the like through an engine torque map as shown in FIG. 16, for example.

The torque converter torque ratio detection unit 266 uses engine rotation speed data Ne sent from the engine rotation speed sensor 120 and from which disturbance components are removed through a filter 262b, and engine rotation speed data Ne sent from the transmission rotation speed sensor 130 and from which disturbance components are removed through a filter 262c. From the transmission rotational speed data Nt, for example, through a transmission torque ratio map as shown in FIG.
The transmission torque ratio t at that time is determined.

The reduction ratio detection unit 276 of the transmission detects the selected shift stage information from the shift position sensor 110.
The reduction ratio ρ of the transmission is determined by referring to the shift stage-reduction ratio correspondence map shown in block 276 of FIG.
It is designed to find m.

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

In the engine torque proportional clutch torque calculation section 270,
Engine torque proportional torque T determined as described above
Calculation is performed from a', the torque converter torque ratio, the reduction ratio ρm of the transmission, the final reduction ρ, and the rotation difference gain of 2. However, the rotation difference gain of 2 is determined by the rotation difference gain setting section 275 as follows. is set to

In other words, the rotational speed difference gain of 2 is intended to avoid the tight corner braking phenomenon, and the ideal rotational speed difference setting section 2
It is determined from the ideal rotational speed difference ΔVhc set in step 18 according to a map as shown in FIG. The relationship between the rotational difference gain 2 and the ideal rotational speed difference ΔVhc in this map is given by the following equation.

K, = 0.9X (lΔV hcmax l lΔVhc
l)/lΔVhcmaxl+0.1...(3,1) However, ΔV hcmax=MAX lΔVhc(δ=
MAX) 1 Also, the coefficient 0.9 and constant 0.1 are k2
This is to set the lower limit of 0.1.

In this way, as the ideal rotational speed difference ΔVhc increases, the rotational difference gain linearly decreases by 2, and by multiplying this rotational difference gain by 2 and correcting it, the ideal rotational speed difference ΔVhc when turning etc. When this becomes large, the engine torque proportional clutch torque Ta is reduced so that 1M cycle performance (performance that can prevent the tight corner braking phenomenon) is prioritized over sudden start performance.

By the way, the above-mentioned engine torque proportional torque setting section 26
8 and the engine torque proportional clutch torque calculating section 270, as shown in FIG. can also be considered.

That is, the center differential input torque calculation section 267 uses the engine torque T sent from the engine torque detection section 264.
e, the torque converter torque ratio t sent from the torque converter torque ratio detection section 266, and the transmission reduction ratio ρm sent from the transmission reduction ratio detection section 276, the center differential input torque (transmission output torque) is calculated by the following formula. Calculate Ta.

Ta=t8ρm+ρ□1Te 1°(3,2
) However, ρ is the final reduction ratio.

Note that the relationship between the center differential input torque Ta and the engine torque Te becomes a proportional relationship for each setting shift. For example, if the torque converter torque ratio t is set to 1.5, the 19th
The result will be as shown in the figure. However, in reality, this relationship varies greatly depending on the magnitude of the torque converter torque ratio t, so the torque converter torque ratio may be determined from the speed ratio 1, and the relationship between Ta and Te may be determined based on this.

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

Tc=[(Zs+Zr)/Zr-Wf/Wa-11・T
a...-(3,3) However, Zs is the number of teeth on the sun gear, Zr is the number of teeth on the ring gear, Wf is the front wheel shared load, Nl
/a is vehicle weight.

Then, the turning correction unit 272 corrects the clutch torque Tc obtained in this manner by the above-mentioned rotational difference gain by 2, thereby obtaining the engine torque proportional clutch torque Ta.

Note that the center differential input torque calculation section 267 and the clutch torque calculation section 269 are integrated to calculate the engine torque Te.
, torque converter torque ratio t, and transmission reduction ratio ρ
(2) Alternatively, it may be calculated using the following formula.

Tc=[(Zs+Zr)/Zr-Wf/Wa-1]
・t・p m・p 1+ Te... (3, 4) Furthermore, the switch 274a is activated by a signal from the determining means 274 when the vehicle speed is low (in this example, V raf<20km
/h), it is turned ON and the engine torque proportional clutch torque Ta can be output as data, but if the vehicle speed becomes higher than this (V ref ≧ 20km/h)
), and the output is stopped as data of the engine torque proportional clutch torque Ta. This is because engine torque proportional control may cause tight corner braking when turning at a certain speed, or may exclude other control units in situations where slip tolerance is required. The condition is that this engine torque proportional control is performed only when the vehicle speed is constant.

Next, the setting of the protection control clutch torque TρC for protecting the clutch portion of the wet multi-disc clutch 28 will be explained. The setting of this clutch torque Tpc is performed by the protection control section 230.

In other words, in the wet multi-disc clutch 28, if the differential rotation between the clutch plates becomes large, there is a risk of damage such as seizing of the clutch facing or increased wear. Naturally, the differential rotation is large and the duration of this state is The larger the value, the more likely it is to cause damage. - On the other hand, in order to avoid such a situation and protect the clutch 28, make the clutch free (
It may be possible to release the connection between the clutch plates, but
If the clutch 28 is instantly switched from the connected state to the free state, there is a risk that the attitude of the vehicle will suddenly change. Therefore, in order to avoid any of these phenomena, the protection control clutch torque rpc is set by the protection control unit 230.

In the protection control unit 230, the front and rear wheel actual rotational speed difference calculation unit 20
In response to the front and rear wheel actual rotational speed difference Vcd calculated in step 6, this front and rear axle actual rotational speed difference Vcd is set to the reference value (in this example, 8
．． 6 km/h) continues for a reference time (1 second in this example), the clutch torque Tpc for protection control is set in a pattern as shown in FIG.

In other words, when the above-mentioned detection conditions are met, the protection control clutch torque Tpc is first set for a short period of time (in this example, for 1 second).
is set to the upper limit value, and thereafter it is gradually decreased (naturally released) to O. In this example, the relationship between Tpc at the time of decrease and time 11 is as shown in the following equation.

rpc = 40-14tt "'"
(4,1) The time to set the upper limit value and the speed at which the clutch torque rpc is gradually reduced to O (corresponding to the slope in Fig. 20) should be set to the optimum value as appropriate depending on the characteristics of each vehicle. is desirable.

Moreover, when the above-mentioned detection conditions are not satisfied.

The value of the protection control clutch torque TρC is set to zero.

Each clutch torque of the above-mentioned differential clutch torque Tv2, clutch torque corresponding to longitudinal acceleration Tx, engine torque proportional clutch torque Ta, and protection control clutch torque Tpc is set at an appropriate timing.

Each clutch torque Tv, Tx+ is set for each repeated control cycle, and is set in this way.
Ta and Tpc are sent to the maximum value selection section 280.

In this maximum value selection section 280, for each control cycle,
Clutch torque Tv, Tx, Ta,
The largest one among T pc (this clutch torque is T
C). However, when the switch 258a or 274a is OFF, the clutch torque Tx or Ta is not sent, so the maximum value selection section 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

The latch control pressure Pc is set so that the set clutch torque Tc can be obtained.

Here, the map (see block 282 in Figure 1)
Clutch control pressure P is calculated from clutch torque Tc by
However, since the clutch torque Tc and the clutch control pressure Pc are generally in a proportional relationship, the map is also linear as shown. □Furthermore, the clutch control pressure Pc changed in this way is subjected to centrifugal pressure correction and preload correction in an adder/subtractor 284 serving as a preload applying means.

- The centrifugal pressure correction is performed by subtracting the centrifugal correction pressure PV set in the centrifugal correction pressure setting section 286 from the clutch control pressure Pc. The front wheel speed Vf calculated in 2'04a is determined using a map as shown in the figure. This is because the piston chamber rotates in synchronization with the front wheel side shaft, so the centrifugal oil pressure is generated in response to the front wheel speed Vf, and the centrifugal correction pressure Pν is proportional to the square of the front wheel speed V, f. It is set as follows.

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

The purpose of this preload correction is to create a contact state between the clutch plates of the clutch 28 that is at the limit where no torque is produced.
The aim is to maintain a very slight contact state) in order to increase the control response.

However, the clearance between the clutch plates of the clutch is
Variations occur from the manufacturing stage to each product due to parts errors, assembly errors, etc., and even the same product changes over time. In particular, since the return springs of the clutch plates are generally strong, errors in each part and changes over time have a large effect on the clearance between the clutch plates.

For this reason, it is necessary to always keep the clutch plates in a close contact state while detecting the clearance state between the clutch plates at an appropriate timing.

For this reason, the preload setting unit 288 sets the initial pressure P1 by testing (herein referred to as learning) at appropriate time intervals how much preload is required.

There are various methods for this preload learning (setting of the initial pressure P1 from the preload learning value), and here, three types of preload learning will be explained next.

First, to explain the first preload learning method, in order to perform preload learning, the engine is in a steady operating state (the oil temperature of the engine has reached a stable temperature state at a predetermined level). It is necessary to perform this under conditions such that a constant line pressure can be obtained and furthermore, control regarding other clutches 28 will not be affected. For this reason, the preload learning conditions are set as follows, for example: 3 [30 minutes or more have passed since the ignition key was turned on.

■Shift selector is 1 (1st speed), 2 (2nd speed), F) (
(drive) or (key/neutral) is selected. P (parking) R (reverse)
In this example, when P and R, there is no range of 1.
,2. This is because a different hydraulic pressure is output than in the case of D and N. 1■Vref=Okm/h (
The vehicle speed V ref is O).

■Tc≦1kgfm [Clutch torque Tc must be less than a small predetermined value (1kgfm).

When each of the above conditions is satisfied at the same time, preload learning is executed as follows.

First, as shown in FIG. 21(a), the multi-disc clutch 28
A pressure greater than the biasing pressure of the return spring and smaller than the designed initial engagement pressure of the clutch 28 [e.g. p=o, duty equivalent to 5 ￥ kgf/■2 for 2 seconds] After this, for example 1.
1 for example P=3. with an increasing rate of 5%/S. Okgf/■2
Sweep slowly until Aiichi's duty.

Then, the pressure P applied to the hydraulic pistons 141, 142 changes as shown in FIG. 21(b).

In other words, the clutch plates are initially separated, so when the duty gradually increases, the hydraulic piston 28 moves accordingly, so the pressure P also gradually increases (but up to a certain position, the hydraulic piston 28 moves). When the pistons 141 and 142 move, the clutch plates come into contact and the pressure P
The force of the return spring is also applied, and the pressure P rapidly increases. Furthermore, the hydraulic piston 141
, 142 move, the clutch plates come into strong contact and the clutch becomes completely engaged. This state occurs because the increase in pressure P reaches its limit.

Here, a value (difference) P'' obtained by second-order differentiation of the detected pressure P with respect to time and a value (difference) P′ obtained by first-order differentiation of pressure P with respect to time are calculated at short intervals, When the second-order differential value P'' reaches the maximum, it is determined that the clutch plate starts contacting, and the pressure P at this time is determined to be the initial pressure, and when the first-order differential value P' reaches the maximum. is determined to be when the clutch plate is fully engaged.

Specifically, when learning is started and the pressure P starts to rise, the maximum value of the second-order differential value P'' and the pressure P at this time are memorized.The value of this second-order differential value P'' is It is calculated every short control cycle and updated as appropriate.

Then, when the first-order differential value P' reaches the maximum (that is, when the clutch is fully engaged), the calculation of the second-order differential value P'' is stopped, and within the period up to this point, the second-order differential value P'' is The pressure P when the maximum value is taken is stored as the initial pressure Pi. .

If any of the preload learning conditions (1) to (2) described in L are no longer satisfied during the execution of such preload learning, the preload learning is immediately interrupted and the mode returns to the normal mode.

Further, the above-mentioned preload learning is performed once when the ignition key is turned on, and then the preload learning is performed once after the ignition key is turned on.

It is configured so that it will not be executed unless the ignition key is turned on after being turned off.

Next, a second preload learning method by the preload setting section 288 will be explained.

This preload learning is also performed under conditions such that the engine is at a predetermined height, the oil temperature is stable, a constant line pressure is obtained, and the control of other clutches 28 is not affected. However, since it is necessary to perform this preload learning by trying many times, the conditions for the preload learning described above are relaxed a little, and the following preload learning conditions are set, for example.

■'10 minutes after the ignition key is turned on
More than a minute has passed.

- The shift selector must be selected from 1 (1st speed), 2 (2nd speed), D (drive), or N (neutral).

■Vref= O)m/h (vehicle speed Vref is O)
).

■Tc≦1kgfm [clutch torque Tc is less than or equal to a small predetermined value (1kgfm)].

- A predetermined period of time (for example, about 5 minutes or an appropriate period shorter than this) has elapsed since the previous trial.

When each of the above conditions is satisfied at the same time, preload learning is performed as follows.

First, a preset initial pressure Pi (=P,)
A corresponding duty is maintained for a predetermined period of time (e.g. 2 seconds), and then a predetermined period of time (e.g. 1 second) is maintained.
P = 8, sweep the duty equivalent to 8 kgf/an2 (almost 100% duty).

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

In other words, if the clutch is released during the initial pressurization,
As shown by curve L1, as the duty is swept, at a certain point, the clutch contacts and begins to drag, so the hydraulic pistons 141 and 142 receive a shock, and the pressure P rapidly increases, overshoots, and then vibrates. However, it settles to a complete engagement pressure (steady peak pressure) corresponding to approximately 100% duty.

When the pressure P overshoots, the subsequent steady maximum pressure Pc (known value, here 8°8kgf/a+
Peak value (maximum value) that is more than - room larger than +”
Pmax occurs.

On the other hand, if the clutch is in contact and dragging during the initial pressure P operation, as the duty is swept as shown by curve L2, the pressure P will increase almost linearly, and at a certain point it will smoothly reach full engagement. The pressure settles down to Pc (steady maximum pressure).

Based on these characteristics, the peak value Pwax of the pressure P is memorized, and the difference α between this value Pax and the steady maximum pressure Pc is
(=Pmax - Pc) is the predetermined value α. If it is larger than , it can be determined that the clutch is disengaged at the initial pressure P1.

Therefore, it is possible to detect and set an appropriate initial pressure Pi by repeating the above-described trial at appropriate time intervals (for example, every 5 minutes) while appropriately increasing or decreasing the starting pressure P from the initial value P1.

In other words, it is desirable to perform this preload learning many times as long as the above conditions are met, and the initial learning value and initial pressure Pi set at a certain point (nth learning stage) are
To generalize and express, the initial learning value is PINTG
(n) and the initial pressure Pi are set as PINT(n). Therefore, the previous initial learning value is PINTG
(n-1), the initial pressure can be expressed as PINT (n-1), and at the n-th learning stage, the previous initial pressure is PI
Learning will be performed using NT(n-1).

Then, the difference α (= Pmax - Pc) obtained by a predetermined duty sweep and the threshold value α. Compare with
The current initial learned value P I NTG (n) and initial pressure PINT (n) are set as follows.

■α≧α. When , PINTG (n) = PINTG (n-1) + βPI
NT (n)-PINTG (n-1)+β=P I
NTG (n) ■α〈α. When , PINTG (n) = PINTG (n-1) - βPI
NT (n)=PINTG (n-1) That is, α≧α
. When , the initial learning value P I NTG (n)
For, the previous initial learning value PINTG (n
−1) plus β (=1 bit of pressure), and the initial pressure PINT(n) is set to be β (=1 bit pressure) added to the previous initial learned value PTNTG (n-1).
In other words, the current initial learning value PINTG (n) is set to the current initial learning value PINTG (n).

This means α≧α. At this time, it can be determined that overshoot has occurred, and therefore, it can be determined that the clutch 28 has not approached the contact state at the previous initial pressure PINT (n-1). Therefore, the current initial learning value PINTG (n) is changed to the previous initial learning value PINTG (n).
Assume that β (=1 bit of pressure) is added to G (n-1), and the current initial pressure PINT (n) is added to the previous initial learned value PINTG (n-1) by β (=1b
It is assumed that the pressure corresponding to

Note that l bit is limited by the resolution of the oil pressure sensor signal that detects the oil pressure applied to the piston, but for example, 1 bit = 0, 05 kgf/■2 or 1 b
Set it to an appropriate value such as 0, 1 kgf/a++2, etc.

On the other hand, α〈α. When , the initial learning value PINTG
For (n), the previous initial learning value P I
Set NTG (n - 1,) plus B (=lbit), or set the initial pressure PINT (n) to the previous initial learning value PINTG (n 1
).

This is αkuα. At the time of , since there is no overshoot, it is determined that the clutch 28 is in a marginal contact state or an excessive contact state at the previous initial pressure PINT(n-1). So - this time's initial learning value P
INTG (n) is the initial learning value PINTG of previous H
It is assumed that β (=1 bit) is added to (n-1), but the initial pressure PINT(n) is set to the previous initial learned value PINTG(n-1,). This is done in order to avoid this because it is not possible to determine whether the clutch 28 is in a marginal contact state or an excessively contact state based only on the result of α<α0, which may lead to chattering. Immediately apply the learning results to the initial pressure P.
1, but the previous learned value is used.

Therefore, if there is excessive contact, α will continue for at least two consecutive cycles. It is considered that the state continues, and the initial pressure P1 is reduced and approaches the appropriate value, although with a delay of one cycle.

Furthermore, if any of the above preload learning conditions ■' to ■ are no longer satisfied during the execution of such preload learning,
Immediately interrupt preload learning and return to normal mode.

In addition, the above-mentioned preload learning is performed under the above-mentioned preload learning conditions
As long as ■ is satisfied, it will continue.

Next, a third preload learning method by the preload setting section 288 will be explained.

Similar to the second preload learning, this preload learning is also set to be executed when the following preload learning conditions are simultaneously satisfied.

■'10 minutes after the ignition key is turned on
More than a minute has passed.

- The shift selector must be selected from 1 (1st speed), 2 (2nd speed), D (drive), or N (neutral).

■Vref= Okm/h (Vehicle speed Vref is ○)
To be.

(l T c≦1 kgfm [clutch torque Tc is less than or equal to a small predetermined value (1 kgfm)].

- A predetermined period of time (for example, about 5 minutes or a shorter appropriate time) has elapsed since the previous attempt.

When each of the above-mentioned conditions ((j.'~.6)) is determined simultaneously, preload learning is executed as follows.

First, the duty is adjusted so that the pressure cover turns as shown in FIG. 23(a). In other words, first, the duty is set to 0% for a predetermined period of time (for example, 1 second).
, then set the duty to a value equivalent to the initial initial pressure P□, hold this for a predetermined time (for example, 2 seconds), and then maintain it for a predetermined time (for example, 1 second) until P=8.
Sweep to a duty equivalent to 8 kgf/■2 (almost 100% duty), P = 8, 8
kgf/an2 equivalent duty for a predetermined time (e.g. 2
hold for seconds). This pattern is continuously repeated while changing the initial pressure P1 as appropriate.

As a result, the pressure P applied to the hydraulic pistons 141 and 142 is the same as in the case of the second preload learning.

As shown by curves Ll and L2 in FIGS. 23(b) and 23(c), two types of pattern changes are made.

Then, from the time t when the duty sweep starts (or the time t when the pressure P starts to rise) until reaching the steady maximum pressure Pc (or a constant pressure value close to this) as shown by the straight line LO. Comparing the areas Sl and S2 of the part (hatched in the figure) surrounded by the straight line LO and the curve L1 or L2 that depicts the change state of the pressure P, the area S1 in the case of the curve L1 with overshoot. is clearly larger than the area S2 in the case of the curve L2 without overshoot.

Therefore, in this third preload learning, as in the second preload learning, the above-mentioned trial is repeated at appropriate time intervals (for example, every 5 minutes) to detect and set an appropriate initial pressure PL. I can do it.

In other words, this preload learning is repeated as many times as long as the above conditions are met, and the initial learning value and initial pressure Pi set at a certain point (nth learning stage) are changed to the initial learning value as described above. is expressed as PINTG (n) and the initial pressure P1 is expressed as PINT(n).

Therefore, the previous initial learning value is PINTG(n
-1), the initial pressure can be expressed as PINT (n-1), and n
In the second learning stage, the previous initial pressure is PINT.
(n-1), learning will be performed.

Then, the area S and the threshold value S obtained by a predetermined duty sweep. The current initial learned value P I NTG (n) and the initial pressure P I N T
(n) is set as follows.

■When S≧80, PINTG (n) - PINTG (n-1) + βPI
NT (n)=PINTG (n-1)+β=PINT
G (n) ■When S<S, PINTG (n)=PINTG (n-1)-βP
I N T (n) = P I N T G
(n −1) In other words, when S≧SO, α≧α of the second preload learning. Corresponding to the case, when S<So, α of the second preload learning is α. Corresponds to the case of

That is, when S≧80, it can be determined that overshoot has occurred, so the previous initial pressure P I NT (n-1)
Therefore, it can be determined that the clutch 28 is not approaching the contact state. Therefore, the current initial learning value P I NTG (n) is changed to the previous initial learning value P
Assume that β (=1 bit worth of pressure) is added to INTG (n-1), and the current initial pressure P I NT (
n) as the previous initial learning value PINTG (n-1)
It is assumed that β (=1 bit worth of pressure) is added to .

On the other hand, SkuS. At the time, there is no overshoot, so it can be determined that the clutch 28 is in a marginal contact state or an excessive contact state at the previous initial pressure PINT(n-1). Therefore, this time's initial learning value P
I NTG (n) is the previous initial learning value PINT
It is assumed that β (=1 bit) is added to G (n-1), but the initial pressure PINT (n) is set to the previous initial learned value PINTG (n-1).

The reason for doing this is also the same as mentioned above. As in the case of , it is not possible to determine whether the clutch 28 is in a marginal contact state or an excessive contact state based only on the result of S<So, which may lead to chattering, so in order to avoid this, Instead of immediately adopting the current learning result as the initial pressure Pi, the previous learning value is used.

Therefore, if there is excessive contact, it is thought that the state of S<So will continue for at least two consecutive cycles, and the initial pressure Pi will decrease and approach the appropriate value, albeit with a delay of one cycle. It turns out.

Furthermore, if any of the above preload learning conditions ■' to ■ are no longer satisfied during the execution of such preload learning,
Immediately interrupt preload learning and return to normal mode.

In addition, the above-mentioned preload learning is performed under the above-mentioned preload learning condition (I′
As long as the conditions (5) to (5) are satisfied, the process continues.

In addition, in this third preload learning, the area 5 (SL) of the part surrounded by the straight line LO and the curve L1 or L2.

S2), the area S' of the part surrounded by the straight line L3 indicating a constant pressure similar to the initial pressure and the curve L1 or L2
It is also conceivable to make the determination with reference to (Sl', 52').

In this case, calculation of the area S' starts at the time when the duty sweep is started. (or the time t when the pressure P starts to rise), and the calculation of the area S' ends when the steady maximum pressure Pc (or a constant pressure value close to this) is reached as shown by the straight line LO. . Then, set the judgment reference value to S
. ', when S'≧So', it can be determined that there has been an overshoot, and when S'<So', it can be determined that there has been no overshoot.

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

This peak-hold filter 290 is a kind of limiter that prevents excessive sudden changes in oil pressure so that hunting does not occur in control due to sudden changes in oil pressure. 4
kg/an2/s), and set a slightly lower limit speed (for example, 15.7 kg/c) in response to a drop in oil pressure.
rn2/s) is set.

If a control pressure Pcd is sent that causes the speed of oil pressure change to exceed such a limit, the control pressure is kept at a level corresponding to this limit value.

Furthermore, the control pressure Pcd' that has passed through the filter 290 is controlled by the switches 292a, 294 . a, and is sent to the duty setting section 295.

Note that the switch 292a is turned OFF according to a signal from the determining means 292 if ABS control (anti-lock brake control) is being performed (if it is in the ON state);
If BS control is not performed, it is turned ON. In other words, the control pressure PCd' signal is sent on the condition that ABS control is not being performed. This is because during ABS control, it is necessary for ABS to work reliably, and at this time, it is ABS that controls the torque distribution state between the front and rear wheels.
This is because it may interfere with BS control, which is undesirable.

Further, the switch 294a is a control switch for protecting the duty solenoid valve and the clutch plate according to a signal from the determining means 294, and when the set clutch torque TC is small at low speed, the duty is turned to O. It is something that we try to do. The low speed condition is, for example, V ref≦5km/h, and the clutch torque TC condition is, for example, Tc≦1kg.
fm, etc. When these two conditions are met, the switch 294a is turned off and the control pressure Pcd' signal is no longer sent.

The duty setting section 295 includes a pressure feedback correction section 296 and a pressure-tute conversion section 298.

The pressure feed pack correction section 296 corrects the signal of the control pressure Pcd' in response to the detection from the pressure sensor 304 that detects the actual pressure acting on the piston, and corrects the characteristics of the hydraulic circuit. It is for the purpose of Note that the signal sent from the pressure sensor 304 to the pressure correction unit 296 is filtered by a filter 306 to remove noise components due to disturbances and the like.

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

In the hydraulic circuit 300 functioning as a control execution section, a duty solenoid 302 is operated in accordance with the duty set in this manner to control the differential limiting clutch 28 of the center differential.

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

In other words, as shown in FIGS. 1 and 24, a torque distribution display section 312 is provided within the 5-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 according to the estimated magnitude of the distributed torque by the torque estimating means '310.

The reason why the torque distribution state is estimated by the torque estimating means 310 in this way is that it is difficult to actually measure the torque distribution state.

This torque estimating means 310 calculates the front export force torque (
or rear export force torque) and front export force torque (or rear export force torque) when there is no rotational speed difference between the front and rear wheels, and a calculation means 310a that calculates the front export force torque (or rear export force torque) in the case where there is no rotational speed difference between the front and rear wheels, and the front @roka torque in each of these cases. (or rear export force torque), a selection means 310b for selecting the smaller front export force torque (or rear export force torque);
310b estimates the torque distribution state as follows.

In other words, when estimating torque distribution, the following two cases are possible. One is the case where it is assumed that the tires and the road surface do not slip and are in a state similar to meshing gears, and that the center differential always slips. The other case is that there is always slippage between the tires and the road surface, so the center differential may not slip.

Therefore, let us consider the torque distribution in each of these cases and when the state changes.

First, as a prerequisite, if differential restriction is not performed as in this four-wheel drive system, the rear wheels are set to normal (the torque ratio between the front wheels and the rear wheels is, for example, 32:68), and furthermore,
Assuming that the differential limiting clutch 28 always transmits torque from the rear wheel side to the front wheel side, for simplification, the following 1.
Set this.

ρf/r4<ρr・,ot/rr”(5,1
) However, ρf: Front differential ratio ρr: Rear differential ratio ρt: Transfer ratio rf: Front tire radius rr: Rear tire radius.If the clutch does not slip, the distribution will be for direct four-wheel drive, so the front wheel torque Tf and rear wheel torque Tr are as follows.

Tf = Wf/Wa・(Tm+k Wr+rf/pe(
rf p r p t/rr p t-1))...(
5, 2) rr = Wr/Wa・(Tm−kWf−rr/, o・(rf
ρrρt/rrρt-1))...(5, 3) Where, Wf: Front axle shared load Wr: Rear wheel shared load Wa: Vehicle weight (=Wf+Wr) Tm: Mission output torque (=center differential input torque) knislip Ratio coefficient ρ: Final reduction ratio [=(ρf+ρr・ρt)/2] Also, if the clutch slips, the front wheel torque Tf' and the rear wheel torque T
r' is as follows.

T f '=(Tm-T c)・a/(a+b)+T
c...(5,4)Tr'=(Tm-Tc)・
b/(a+b) , -...(5,s) However, Tc
:Clutch transmission torque capacity a: Number of Sanghya teeth b = Number of ring gear teeth And, if the latch slips as described above, it means that the clutch is allowing the difference in torque between the front and rear caused by the weight distribution, differential ratio difference, etc. .

Now, we are considering the case where the clutch transmits torque from the rear wheel side to the front wheel side, so the front wheel torque Tf.

Regarding Tf', the smaller value of Tf and Tf' can be considered as the front wheel torque value.

That is, if Tf<Tf', the clutch is in the locked state,
The front wheel torque distribution ratio m is m=T f / (T f + T r)
...(5,6) If Tf>Tf', the clutch is in a slipping state, and the front wheel torque distribution ratio m is m = Tf'/ (Tf'+Tr')"
・It can be estimated as (5, 7).

Note that FIG. 25 shows the front wheel torque distribution ratio m with respect to the center differential input torque Tm, and the characteristic of the front wheel torque distribution ratio corresponding to the input torque is a linear shape with direct connection when the clutch is in the locked state. When the clutch is in a free state, the curve becomes curved depending on the magnitude of the control pressure P. In addition, in the figure, when the pressure P is 2 kgf/cm2 (P=2
) and the case of 8 kgf/an'' (P=8) are shown.

In the characteristic graph, the characteristic line with the smaller m of the straight line marked with direct connection and the curve for a certain control pressure P is adopted.

For example, if P is 2 kgf/a++2, the torque T
In the region where e is smaller than Te1, the straight line marked as direct connection is located below the curve of P=2, so the front wheel torque distribution ratio m follows the second straight line. Also, the torque Te is T
In a region larger than e1, the curve of P=2 is located below the direct connection, so the front wheel torque distribution ratio m follows the curve of P=2.

On the other hand, when P is 8 kgf/an2, in the region shown in this graph, the direct connection straight line is always lower, so the @-axis torque distribution ratio m follows the direct connection.

Once the front wheel torque distribution ratio m has been set in this way,
A signal corresponding to this set value is sent to the torque distribution display section 312, and the torque distribution display section 312 displays the state of torque distribution to the front wheels. In this example, the torque distribution to the front wheels is about 32% to 50%, so the torque distribution display section 312 is provided with a scale corresponding to this, and the lamps are lit up to the corresponding scale. By moving the opening, it is displayed in an easy-to-understand manner.

Note that this display of the torque distribution state may be a state of torque distribution to the rear wheels, or may be an analog display of the state of torque distribution to the front and rear wheels using a graph or the like.

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

As shown in Fig. 26, the overall operation flow of the shame system is as follows: First, each control element is initialized (step a1), the steering angle neutral position is learned (step a2), and the clutch is activated. Preload learning (step a3) is performed, and then front and rear @ driving force distribution control is performed while controlling the clutch 28 according to the set duty (step a4).
), and further controls the rear differential (step a5).
.

Then, in steps 87 to all, slip control, trace control, torque selection, retard control calculation, SCI
(Serias Communication In
engine output control (interface) communication control
Traction control) is performed, a torque distribution display lamp is turned on (step a12), and a failure diagnosis is performed in step a13. step a14
Then, it is determined whether or not a predetermined time (15m5et) has elapsed, and when the predetermined time (15m5ec) has elapsed, a runaway check is performed by Woychitosog (step a15).
), return to the above steps ``2'' and perform steps a2-a.
13 series of controls are repeated.

In other words, the above-mentioned front/rear @active power distribution control, rear differential control, and engine output control are performed at a predetermined period (15 msec).

Of these, the front and rear contour power distribution control will be explained with reference to the flowchart in FIG. 27.

As shown in FIG. 27, first, wheel speeds FR, FL, RR
, RL, steering angle θ1tLlon, lateral acceleration Gyr, longitudinal acceleration Gx, throttle opening Oth, engine speed Ne, transmission speed Nt9, selected shift stage, etc. are detected and imported (step b).
1) From these data, front wheel speed Vf, rear wheel speed Vr, driver requested vehicle speed Vref, driver requested steering angle δr
ef etc. are calculated (step b2).

Then, the ideal rotational speed difference ΔVhc between the front and rear wheels is determined according to the map from the driver-required vehicle speed V ref + the driver-required steering angle δref (step b3), and the lateral gain is set to 6 according to the map from the lateral acceleration Gy. (Step b4), and the rotational difference gain is set to 2 according to the map from the ideal rotational speed difference Δ■hc (Step b5).

Furthermore, in steps b6 to b9, the differential compatible clutch torque Tv is calculated from the actual rotational speed difference ΔVc, the ideal rotational speed difference ΔVhc, and the lateral 6 gain (in this example, calculated by converting from the map), and the longitudinal acceleration Calculate the clutch torque Tx corresponding to the longitudinal acceleration from Gx, lateral G gain, and the lateral G gain (converted from the map), and calculate the throttle opening θth and the engine rotation speed Ne.

Transmission rotation speed Nt, selected shift stage.

Calculate the engine torque proportional clutch torque Ta from the rotation difference gain 2 (converted from the map), and calculate the ideal rotation speed difference Δ
Clutch torque for protection control rpc according to the Vhc signal
Set.

Then, in step blo, the maximum clutch torque Tv, Tx, Ta, and Tpc is calculated as the set clutch torque Tc.

Furthermore, in step bll, the thus determined clutch torque Tc is converted from the map into clutch engagement pressure Pc.

Next, preload correction (adding preload Pi) to this pressure Pc
Then, centrifugal pressure correction (elimination of centrifugal pressure Pv) is performed (step b12) to obtain center differential control pressure Pcd.

Furthermore, optimize the peak hold filter.

It is possible to suppress excessive changes in pressure P (step b13).

Then, whether ABS is in operation (step b14),
Solenoid valve protection conditions (V ref≦5km/h
, Tc5kg f m ) is satisfied (
After the judgment in step b15), if any of these conditions apply, in step b19, the center differential control pressure Pcd is
Reset to O.

The center differential control pressure Pcd set in this way is
At step b16, pressure feedback correction is performed. In other words, the difference Δ between the value of Pad and the actual value of the pressure sensor
P is calculated, and the integral correction pressure Pi obtained from the product of the integral correction gain ki and ΔP (i) and the proportional correction gain kp are calculated.
The above-mentioned center differential control pressure Pcd is corrected by the proportional correction pressure Pp found from the product of ΔP and the pressure P is obtained.

Furthermore, in step b17, the pressure P is converted into a corresponding duty to perform center differential control, that is, control of the operation limiting clutch.

The above-mentioned calculation of the differential clutch torque Tv is performed as shown in FIG. 28.

First, calculate the difference ΔVcd (=Vr-Vf) by subtracting the front wheel speed Vf from the rear wheel speed Vr.Step c1)
Then, from this difference (actual rotational speed difference between the front and rear wheels) ΔVcd, the ideal rotational speed difference ΔVhc between the front and rear wheels obtained as described above (see step b3) is subtracted, and the difference ΔVc (=
ΔVcd−ΔVhc) is determined (step c2).

Then, in step c3, it is determined whether the above-mentioned ideal rotational speed difference ΔVhc between the front and rear wheels is greater than or equal to 0. If ΔVhc is greater than or equal to 0, the process proceeds to step c4, and if ΔVhc is less than 0, the process proceeds to step c5.

Proceeding to step c4, the map [see Figure 13(a)]
Clutch torque Tv' is set from ΔVc using

Specifically, if ■ΔVcd≧ΔVhc, then Tv'=aX(ΔVcd-ΔVhc')=a
Set as Vc (where a is a proportionality constant).

If ∆Vhc>∆Vcd>O, the clutch torque Tv' is set to 0, and a so-called dead band region is set.

Furthermore, if ■0≧ΔVcd, clutch torque Tv'
Tv' increases in proportion to the magnitude of ΔVcd.
= −a XΔVed= −a X (ΔVc+ΔV
hc) (where a is a proportionality constant).

Note that when ΔVhc=O, there is no dead zone region where ΔVhc>ΔVcd>0.

Proceeding to step c5, the map [see Figure 13(b)]
Clutch torque Tv' is set from ΔVc using .DELTA.Vc.

Specifically, (if DΔVcd≧0, then Tv'=aXΔVcd=aX(ΔVc+ΔVhc) is set so that clutch torque Tv' increases in proportion to the magnitude of ΔVcd (where a is a proportionality constant).

If (2) 0>ΔVcd)ΔVhc, the clutch torque Tν' is set to O, and a so-called dead zone region is set.

Furthermore, if ■ΔVhc≧ΔVcd, Tv'=-aX(ΔVcd-ΔVhc)=-aX so that the clutch torque Tv' increases in proportion to the magnitude of ΔVc(ΔVcd-ΔVhc).
Set as ΔVc (where a is a proportionality constant).

Thus, step c4. In step c5, the differential clutch torque Tv' obtained is corrected for the lateral acceleration by adding 1 to the lateral G gain in the correction unit 246 (step c6), and the differential clutch torque Tv is obtained.

By setting the differential compatible clutch torque Tv like this,
The magnitude of the clutch torque Tv is set appropriately without wasting any waste, and the attitude control of the vehicle body can be adjusted appropriately while setting the rear wheels as the drive base and slipping from the rear axle, and the driver's will when turning. This makes it possible to make the vehicle behave in accordance with the following.

In other words, if the direction 5IG (δh) of the sensor-compatible steering angle δh and the direction 5IG (Gy) of the lateral acceleration data Gy are not equal, the driver-required steering angle is set to Q, so the driver can, for example, Even when the steering wheel position and the actual steering angle (turning state) of the vehicle differ when performing steering operations such as steering, inappropriate data will not be adopted, contributing to improved control performance. do.

Furthermore, since the wheel speed data with the second largest size from the bottom among the rotational speed data signals FL, FR, RL, and RR is used as the driver-required vehicle speed V r e f, the reliability of the data is ensured. ing.

When the ideal rotational speed difference ΔVhc is set, at low vehicle speeds, the influence of the difference in track radius of the front and rear wheels (so-called inner ring difference) during turning is large, and the rotational speed Vr of the rear wheels is determined by the rotational speed Vf of the front wheels.
However, as the vehicle speed increases, the rotational speed Vr of the rear wheels becomes larger than the rotational speed Vf of the front wheels.

Therefore, the rear wheels tend to slip at high speeds, and the responsiveness of the vehicle body attitude required at high speeds is ensured.

Regarding the steering angle, the greater the steering angle, the greater the rotation difference required between the front and rear wheels, which has the advantage of being appropriately tolerated and avoiding tight corner braking.

Further, the calculation of the clutch torque TX corresponding to the longitudinal acceleration described above is performed as shown in FIG. 29.

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

Then, lateral acceleration correction is applied to this longitudinal acceleration-corresponding clutch torque TX' by applying a lateral G gain (step d2) to obtain longitudinal acceleration-corresponding clutch torque Tx.

Further, in step d2, the front wheel speed Vf is changed to the vehicle body speed V.
It is determined whether it is greater than ref.

Through the switch 258a, the front wheel speed Vf changes to the vehicle body speed V.
When it is larger than ref, that is, when the front wheels are slipping (at the time of front slip), the above-mentioned longitudinal acceleration corresponding clutch torque Tx is directly adopted as control data, and the front wheel speed Vf is larger than the vehicle body speed Vref. If no, that is, the front wheels are not slipping, the clutch torque Tx corresponding to the longitudinal acceleration is set to O (step d4).

As a result, when accelerating, such as when front slips, the torque distribution is the same as that of direct-coupled 4WD, but any torque beyond that is distributed to the base distribution ratio (closer to the rear wheels), preventing strong under-distribution. , you will be able to make smooth turns.

Further, calculation of the engine torque proportional clutch torque Ta is performed as shown in FIG. 30.

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

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

Furthermore, the torque converter torque ratio detection unit 266 detects the engine rotation speed data Ne and the transmission rotation speed data t.
From this, the transmission torque ratio t at that time is determined using a transmission torque ratio map as shown in FIG. 13 (step e3).

Then, the engine torque proportional clutch torque calculation section 27
0, the engine torque proportional torque Ta' and the torque converter torque ratio t obtained in this way, the transmission reduction ratio ρm, the final reduction ratio ρ, and the rotation difference gain setting section in the transmission reduction ratio detection section 276. From the rotation difference gain obtained in 275 and 2, center differential input torque (transmission output torque) Ta (=t・9m
・ρ□・Te) is calculated (step e4).

Further, in step e5, when the vehicle speed is low (in this example, Vrs
f < 20 km/h), and if the vehicle speed is low, the engine torque proportional clutch torque Ta described above is output as data as is, but if the vehicle speed becomes higher than this (V ref ≧ 20 kII/h). Then, the engine torque proportional clutch torque Ta is set to 0 (step e6), and this is output as control data.

With this engine torque proportional clutch torque Ta, when starting or suddenly accelerating from low speed, etc., the direct 4WD state is set as appropriate, and high torque can be transmitted to the road surface. This prevents the tires from slipping at times, improving driving performance and contributing to the durability of the drive system.

Further, the above-mentioned calculation of the protection control clutch torque rpc is performed as shown in FIG. 31.

First, in step f1, it is determined whether the flag FLG is 1 or not. This flag FLG is a control flag that is set to 1 when -protection control is executed.

It is set to O at the start of overall control.

Therefore, at the start of control, the process proceeds to step f2, and the actual rotational speed difference Vcd of the front and rear wheels is set to the reference value (in this example, 8.6
km/h) or more.

Actual rotational speed difference between front and rear wheels Vcd is standard value (8.6 km/h)
If not, the process proceeds to step f9, and if the timer is counting, the counting is ended and the timer is cleared. Then, in step f12, the value of the protection control clutch torque Tpc is set to O, and further, in step f14, the flag FLGt-0 is set.

On the other hand, if it is determined in step f2 that the front and rear wheel actual rotational speed difference Vcd is greater than or equal to the reference value (8,61ai/h),
The process proceeds to step f3, where it is determined whether or not the timer count has been started. If the timer count has not been started, the process proceeds to step f4, where the timer count is started.

When the timer count is started in this way, step f
In step 5, it is determined whether the timer value is equal to or greater than the reference time (lsec), and if the timer value does not reach the reference time or greater, the process proceeds to step f12, where the value of the protection control clutch torque rpc is set to O. At step f14, the flag FL
G@:O.

During several control cycles, if the front and rear wheel actual rotational speed difference Ved is equal to or higher than the reference value (8.6 km/h), the timer continues to count during this period, and in step f5,
It can be determined that the timer value reaches the reference time or more, and in this case, the process advances to step f6.

In step f6, the timer value is set to the reference time (2 seconds).
If the timer value does not reach the reference time or more, the process proceeds to step flo, where the value of the protection control clutch torque rpc is set to 40.

Then, in step f13, the flag FLG is set to 1, and the process proceeds to step f8, where it is determined whether rpc is 0 or more. When proceeding from step flo to step f8, naturally r
Since pc is greater than or equal to 0, timer counting continues.

Then, when this state of rpc=40 continues and the timer value becomes 2 seconds or more, the process proceeds from step f6 to step f7, where protection control is performed based on the relationship rpc=40-14 x (timer value - 2). The value of clutch torque Tpc is gradually decreased.

In this way, after several control cycles, when the protection control clutch torque rpc is no longer 0 or more, the process advances from step f8 to step fil, where the timer count is finished, the timer is cleared, and step f1
In step f14, the value of the protection control clutch torque Tpc is set to O, and in step f14, the flag FLG is set to O.

As a result, the actual rotational speed difference Vcd of the front and rear wheels is set to the reference value (8
, 6km/h) or more for a reference time (isec) or more is established, the characteristics shown in Fig.
In this example, the protection control clutch torque rpc is set so that it is set at the upper limit value for 1 second) and then gradually decreases to O (naturally released).

This protection control clutch torque rpc protects the clutch plate, contributes to improving the durability of the device, and has the effect of helping to prevent the vehicle from spinning.

Here, the above-mentioned preload correction will be explained with reference to FIGS. 32 to 34.

First, in the first preload learning method, as shown in FIG. (1st speed), 2 (2nd speed),
D (drive), N (neutral) is selected, ■Vehicle speed Vref is Ok/km/
h (stopped state) and (1) whether the set value Tc of the clutch torque is less than a small predetermined value (1 kgfm).

If both of these conditions are met, the process proceeds to step g5, and if any of these conditions are not met, no learning control is performed.

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

In step g6, the oil pressure is increased, the maximum value (MAX) of the second-order differential value of the oil pressure is detected, and the oil pressure P at that time is memorized.

That is, first, as shown in FIG. 21(a), for example, P
= 0, duty equivalent to 4 kgf/cm2 (d
uty) for 2 seconds, and then, for example, 1.5%/S
For example, P = 3, 0 kgf/an
Sweep slowly until the duty is equivalent to 2.

On the other hand, as shown in FIG. 21(b), the pressure P on the hydraulic pistons 141, 142 changes, and the maximum value of the value (difference) P'', which is obtained by second-order differentiation of this pressure P with respect to time, and The pressure P is memorized.

Then, the memorized pressure P is set as the initial pressure.

Specifically, when 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 memorized, and the value of this second-order differential value P'' is stored. is calculated every control cycle and updated as appropriate, and when the first-order differential value P' reaches the maximum (that is, when the clutch is fully engaged), the calculation of the second-order differential value P'' is stopped, and at this point The pressure P when the second-order differential value P'' takes the maximum value within the period up to this point is stored as the initial pressure Pi.

During the execution of such preload learning, if any of the conditions for preload learning described above are no longer satisfied,
Immediately interrupts preload learning and returns to normal mode, and once this preload learning is performed once the ignition key is turned on, it will not be performed again unless the ignition key is turned on again after being turned off. .

In addition, in the second preload learning method, as shown in FIG. ■Whether a predetermined period of time (e.g., about 5 minutes or a shorter appropriate time) has elapsed since It is determined whether one of the following is selected, (1) whether Vref=Okrm/h, and (2) whether Tc≦1kgfm. And if both of these conditions are met,
Proceeding to step h6, if any of these conditions is not satisfied, no learning control is performed.Proceeding to step h6, the oil pressure is increased and an overshoot value of the oil pressure is detected.

In other words, to start up the hydraulic pressure, the duty corresponding to the preset initial initial pressure P2 is maintained for a predetermined time (
P = 8, 8 kgf/aI for a predetermined period of time (for example, 1 second full).
Sweep to a duty equivalent to "+" (which is almost 100% duty).

The hydraulic piston 141.1 changes accordingly.
The overshoot value α of the pressure P applied to 42 is detected.

Furthermore, in the first step h7, it is determined whether this α is larger than a threshold value.

That is, the peak value (maximum value) Pmax of the pressure P is detected, and this maximum value Pmax and the steady maximum pressure Pc (here, 8.
8kgf/cm") is set as the overshoot value α. If this α is larger than the threshold value (α6), there is an overshoot. Therefore, at the initial initial pressure P□, the clutch 28 is If it can be determined that they are separated and α is not larger than the threshold value, it can be determined that there was no overshoot, that is, the clutch 28 is in a marginal contact state or an excessive contact state at the initial initial pressure P1.

Therefore, if α is larger than the threshold value, proceed to step h8 and calculate PINTG (n) = PINTG (n-1) + βPI
NT (n)=PINTG (n-1)+β=PI
NTG (n) In other words, for the initial learning value PINTG(n), β is added to the previous initial learning value PINT, G(n-1).
(= 1 bit worth of pressure), and the initial pressure PINT(n) is set to the previous initial learned value PINTG (n-1) plus β (=1 bit worth of pressure).
Only added.

In other words, this initial learning value P I NTG (n
).

On the other hand, if α is not larger than the threshold, the process proceeds to step h9.

PINTG (n)=PINTG (n-1)-βPI
NT (n) = PINTG (n-1) In other words, for the initial learning value PINTG (n), β (= 1bi
t minutes), but the initial pressure PI
NT(n) is the previous initial learning value P I
Set to NTG (n −1).

Furthermore, if any of the above preload learning conditions ■' to ■ are no longer satisfied during the execution of such preload learning,
Immediately interrupt preload learning and return to normal mode.

In addition, the above-mentioned preload learning is performed under the condition JI′ of the above-mentioned preload learning.
As long as ~■ is satisfied, the process will continue.

Furthermore, in the third preload learning method, as shown in FIG. 34, it is determined whether the conditions are the same as those in the third preload learning. That is, in steps h1 to h5.

■'10 minutes after the ignition key is turned on
■Whether a predetermined amount of time has passed since the previous attempt (e.g., about 5 minutes or an appropriate time shorter than this); ■Whether the shift selector is set to 1 (1st gear)
, 2 (2nd speed), D (drive), N (neutral), ■Vref=
It is determined whether Tc is Okm/h and whether ■Tc≦1kgfm.

If both of these conditions are met, the process proceeds to step h16, and if any of these conditions are not met, no learning control is performed.

Proceeding to step h16, the oil pressure is increased and the difference between the predetermined pressure and the oil pressure value is integrated.

In other words, to start up the hydraulic pressure, the duty corresponding to the preset initial initial pressure P1 is maintained for a predetermined time (
For example, it is held for only 2 seconds), and then it is swept for a predetermined time (for example, 1 second) to a duty equivalent to P = 8, 8 kgf/an2 (which is almost 100% duty).

The hydraulic piston 141.1 changes accordingly.
The difference between the pressure P applied to 42 and a predetermined pressure (a pressure close to the maximum pressure) is integrated. That is, as shown in FIGS. 23(b) and 23(C), the time t is when the duty sweep is started. (or the time t when the pressure P starts to rise) until reaching the steady maximum pressure Pc (or a constant pressure value close to this) as shown by the straight line LO, the change in the pressure P with respect to the straight line LO. The area S (Sl, S2) of the portion (hatched in the figure) surrounded by the curve L1 or L2 that depicts the state is calculated.

Furthermore, in the next step h17, the calculated area S is set to the threshold value S. Determine whether it is greater than. In other words, the area S1 in the case of the curve L1 with overshoot is
Since the area S2 is obviously larger than the area S2 in the case of the curve L2 without overshoot, the area S is set to the threshold value S. The presence or absence of overshoot is determined by comparing with .

Therefore, the area S is the threshold value S. If it is larger than .

Proceed to step h8, PINTG (n) = PINTG (n-1) + βPI
NT (n)=PINTG (n-1)+β= P I
N T G (n) In other words, for the initial learned value P I NTG (n), the previous initial learned value P I NTG (n-
1) plus β (= 1 bit worth of pressure), and the initial pressure P I NT (n) is set by adding β (= lb) to the previous initial learned value PINTG (n-1).
The current initial learning value PINTG(n) is set to the current initial learning value PINTG(n).

On the other hand, the area S is the threshold value S. If it is not larger than , proceed to step h9 and calculate PINTG (n) = PINTG (n-1) - βPI
NT (n) = PINTG (n-1) In other words, for the initial learned value PINTG (n), β (=1
However, the initial pressure P I NT (n) is set to the previous initial learned value PINTG (n-1).

Even during execution of the third preload learning, if any of the above-mentioned preload learning conditions (1) to (2) are no longer satisfied, the preload learning is immediately interrupted and the mode returns to the normal mode.

Also in this case, as long as the above-mentioned preload learning conditions (■' to ()) are satisfied, the process is continued.

Through such first to third preload learning, appropriate initial pressures P1 can be set, respectively, and this greatly contributes to improving control response.

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

Further, in the second and third preload learning, the initial pressure Pi is set by learning several times, which has the advantage of high setting accuracy and a large effect of improving response.

Especially in judgments based on integral value (area).

It is possible to relatively appropriately determine whether the initial pressure Pi is appropriate, and the initial pressure P1 can be set without relying heavily on the ability of the pressure sensor.

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

Furthermore, there is a torque distribution display section 312 in the meter cluster.
is installed to graphically display (or meter display) the state of torque distribution to the front wheels (or rear wheels), so the driver can drive while recognizing the torque distribution state of the vehicle, and use it effectively for driving. This has the advantage of making it more enjoyable and greatly improving product quality.

Furthermore, it is also conceivable that the results of the torque distribution estimation performed at this time may be used as feedback to the control of each part.

[Effects of the Invention] As detailed above, according to the differentially adjustable front and rear axle torque distribution control device of the present invention, by adjusting the differential state between the front wheels and the rear axle, A friction type that adjusts the differential state of the front and rear axles by applying frictional force between the front and rear wheels of a four-wheel drive vehicle. A front and rear wheel differential adjustment mechanism, a steering angle data detection means for detecting steering angle data, a vehicle speed data detection means for detecting vehicle speed data, and the steering angle data detected by the steering angle data detection means and the vehicle speed data. ideal rotational speed difference setting means for setting an ideal rotational speed difference between the front wheels and the rear wheels based on the vehicle speed data detected by the detection means; and an actual rotational speed difference between the front wheels and the rear wheels. an actual rotational speed difference detection means for detecting the difference between the actual rotational speed and the ideal rotational speed so that the actual rotational speed difference between the front and rear axes approaches the ideal rotational speed difference. a friction torque setting means for setting the friction torque of the friction type front and rear wheel differential adjustment mechanism;
and a control means for controlling the friction type front and rear wheel differential adjustment mechanism so as to obtain a friction torque, and as the difference between the actual rotational speed difference and the ideal rotational speed difference becomes larger, the friction torque setting means By setting the above-mentioned friction torque value to a large value, the driver's steering wheel operations are reflected in the torque distribution control, making them smoother and more responsive to the driver's will. Allows for quick turns. Therefore, it is possible to improve the turning performance in the early stage of the turn and to stabilize the vehicle body attitude in the latter half of the turn. Furthermore, since torque distribution control can be performed more accurately, it becomes easier to avoid tight corner braking phenomena.

Further, in the above friction torque setting means, by providing a dead zone in which the above friction torque value is set to O in a region where the difference between the above actual rotation speed difference and the ideal rotation speed difference is smaller than a predetermined value, it is possible to torque control becomes possible.

[Brief explanation of the drawing]

1 to 34 show a differentially adjustable front and rear axle torque distribution control device as an embodiment of the present invention. FIG. 1 is a block diagram showing the configuration of its main parts, and FIG. The overall configuration of the transmission system, Figure 3 is a sectional view showing the main parts of the drive torque transmission system, Figure 4 is a sectional view of the main parts of the front and rear wheel torque distribution mechanism, and Figure 5 is a schematic diagram of the hydraulic pressure supply system. circuit diagram,
FIG. 6 is a circuit diagram of the main part of the oil pressure supply system, FIG. 7 is a diagram showing the characteristics of the oil pressure setting duty, and FIG. 8 is a block diagram showing details of the total steering angle data detection means. FIG. 9 is a block diagram showing details of the vehicle speed detection means, and FIG. 10 is a diagram showing a map for setting the ideal rotation speed difference.
FIG. 11 is a diagram showing the lateral acceleration gain setting map, FIGS. 12(a) and (b) are plan views schematically showing the wheel conditions to explain the ideal rotational speed difference, and FIG. 13 is a diagram showing the lateral acceleration gain setting map. (a) and (b) are diagrams respectively showing maps for setting the clutch torque corresponding to the differential, FIG. 14 is a block diagram showing the clutch torque setting means corresponding to the longitudinal acceleration, and FIG. 15 is a diagram showing the clutch torque setting corresponding to the longitudinal acceleration. Map for use, 1st
Figure 6 is a diagram showing an example of the engine torque map.
Figure 18 is a block diagram showing an example of the transmission torque ratio map, Figure 18 is a block diagram showing a modification of the engine torque proportional clutch torque setting means, Figure 19 is the center differential input torque setting map, and Figure 20 is its protection. A characteristic diagram of the control clutch torque, FIG. 21(a) is a diagram showing the duty characteristic related to the first preload learning, FIG.
Figure (b) shows the pressure characteristics applied to the first preload learning, Figure 22 shows the pressure characteristics applied to the second preload learning, and Figure 23 (a) shows the pressure characteristics applied to the third preload learning. Diagrams showing duty characteristics required for learning, Fig. 23 (b) and (c)
24 is a diagram showing the pressure characteristics applied to the third preload learning, and FIG. 24 is a diagram showing the torque distribution state display means.
FIG. 25 is a characteristic diagram for explaining torque distribution by the torque distribution state estimating means, FIG. 26 is a flowchart showing the flow of control of the entire vehicle including the device, and FIG.
7 is a flowchart showing the flow of the front and rear wheel torque distribution control, FIG. 28 is a flowchart showing the flow of setting the clutch torque corresponding to the differential, and FIG. 29 is a flowchart showing the flow of setting the clutch torque corresponding to the longitudinal acceleration. , Fig. 30 is a flowchart showing the flow of setting the engine torque proportional clutch torque, Fig. 31 is a flowchart showing the flow of setting the protection control clutch torque, and Fig. 32 shows the flow of the first preload learning. 33 is a flowchart showing the flow of the second preload learning, and FIG. 34 is a flowchart showing the flow of the third preload learning. 2- Engine, 4 Torque converter, 6--
Automatic transmission, 8.-output shaft, lO-.intermediate gear (transfer idler gear), 12.-center differential (center differential), 14.-differential gear device for front wheels;
15--Bevel gear mechanism, 15A--Bevel gear shaft, 15a--Bevel gear, 16.18--Front shaft, 1
7L, 17R...-Front wheel side axle, 19-Reduction gear mechanism 1
．． 19a-output gear, 20-propeller shaft, 21.
- bevel gear mechanism, 22 - differential gearing for the rear wheels, 24
．． 26-Rear wheel, 25L, 25R-・Axle for rear wheel 27-
Output shaft for front wheels, 27a...Hollow shaft member, 28...Differential limiting mechanism, 28a...Front export force side disc plate, 2
8b--Input end disc plate, 29-Output shaft for rear wheel, 30.30a. 30b, 30cm-handle angle sensor, 32-, steering wheel, 34, 34a, 34b-
Lateral acceleration sensor, 36...-Longitudinal acceleration sensor, 38-
Throttle sensor, 39-・Engine key switch, 4
o, 42.44.46-Wheel speed sensor, 48,-Controller, 50...-Anti-lock brake device, 50
A-brake switch, 51--flake pedal, 5
2-warning light, 54-hydraulic source, 56-pressure control valve system (pressure control valve), 58-pump, 60... check valve, 62
- Pressure control valve, 64 - Relief valve, 66 - Accumulator, 68 - Pressure switch, 68a - Motor, 113 - Input gear, 114a to 114 f - Bearing, 115 - Transmission case, 115 a
- End cover, 115 b- Spacer member, 116-
- Support member, 117a, 117b - Oil passage, 121 - Sun gear, 122 - Planetary pinion (planetary gear), 12:3 - Ring gear, 124 - Input case, 125 - Planet carrier, 125a - Base plate part, 125b...Planetary pinion housing part, 125f-clutch disc support part, 126...
Pinion shaft, 130 - connection member, 141 - first piston, 142 - second piston, 143 - partition plate, 144 a - first oil chamber, 144 b - second
Oil chamber, l 45-Hollow shaft, 145a...Piston housing part, 160...-Shift lever position sensor (shift range detection means), 160A--Shift lever of automatic transmission
1161--4WD control valve, 162,-
Duty solenoid valve (duty valve), 1
63...orifice, 164...-oil filter,
165 - Reducing valve, 170 - Engine rotation speed sensor, 180 - Transmission rotation speed sensor, 200 - Front and rear wheel actual rotation speed difference detection unit, 2028 ~
202d...-Filter, 204 a--Front wheel rotation speed data calculation unit, 204 b-Rear wheel rotation speed data calculation unit, 206, - Front and rear wheel actual rotation speed difference calculation unit, 2
10. - Front and rear wheel ideal rotational speed difference setting unit, 212 - Driver requested steering angle calculation unit as steering angle data detection means (
pseudo steering angle calculation unit), 212a - sensor compatible steering angle data setting unit, 212b, - lateral acceleration data calculation unit,
212c-comparison unit, 212d--driver requested steering angle setting unit (vehicle speed data setting unit), 216...driver requested vehicle body speed calculation unit (pseudo vehicle body speed calculation unit) as vehicle speed data detection means; 216a--Wheel speed selection section, 216c-
... Driver requested vehicle body speed calculation section, 216d - Filter, 218 ... Ideal rotational speed difference setting section as an ideal operating state setting section, 22o - Differential compatible clutch torque setting section, 222 - Subtractor, 230・-Protection control unit, 242-・
- Filter, 244 - - Lateral G gain setting section, 246, -
Correction unit, 254...Clutch torque setting means corresponding to longitudinal acceleration, 254 a-1 front shaft shared load calculation means, 254
b - Total output torque calculation means, 254 c - Front wheel shared torque calculation means, 254d... Clutch torque calculation means, 256... - Lateral acceleration corresponding correction unit, 258a...
switch, 258, -judgment means, 264-engine torque detection section, 266--torque converter torque ratio detection section, 267
- Center differential input torque calculation unit, 268 - Engine torque proportional torque setting unit, 269 - Clutch torque calculation unit, 270 - Engine torque proportional clutch torque calculation unit, 272 - Turn correction unit, 274a, -. Switch, 274--determination means, 275--rotation difference gain setting section, 276--transmission reduction ratio detection section, 280--maximum value selection section, 282--torque-pressure conversion section, 286-. Centrifugal correction pressure setting section, 288...-
Initial engagement pressure setting section (preload setting section), 290--peak holt filter, 292a, 294a--switch,
29゛5...Duty setting section, 292゜294-m
- Determination means, 296--Pressure feedback correction section,
298--Pressure-duty converter, 300--,
302-Duty solenoid, 304--Pressure sensor, 306--Filter, 310-Torque estimation means, 3
10 a--Calculating means, 310 b-- Selection means, 31
2...-Torque distribution display unit, AM...-Mechanical part of differentially adjustable front and rear wheel torque distribution control device. Fig. 5 Fig. 7 Knife Fig. 9 and 216 10th Prisoner 1I Fig. 8 Vhcmax Fig. 12 β Neshi [To, Surua member (b) 130 (b) Fig. 15 Fig. 16 e 1Y Figure 19 Figure 20 Figure 21 O day fT IBl ↑ (〃) Figure 22 Seia Intertalk Figure 23 Figure 23 (bl (C) Figure 24 Figure ?-5・1e1 Torque (Te ) Figure 28 Figure 29 Figure 30 Figure 310 Figure 32 Figure 330 Figure 34

Claims (5)

[Claims] (1) In a front and rear differential adjustable four-wheel drive vehicle that controls torque distribution to the front and rear wheels by adjusting the differential state between the front and rear wheels, a friction-type front and rear wheel differential adjustment mechanism that adjusts the differential state of the front and rear wheels by applying a frictional force between the wheels; a steering angle data detection means that detects steering angle data; and a steering angle data detection means that detects vehicle speed data. A vehicle speed data detection means, and an ideal rotational speed difference between the front wheels and the rear wheels is determined based on the steering angle data detected by the steering angle data detection means and the vehicle speed data detected by the vehicle speed data detection means. An ideal rotational speed difference setting means to set, an actual rotational speed difference detection means to detect the actual rotational speed difference between the front wheels and the rear wheels, and an actual rotational speed difference between the front and rear wheels to determine the ideal rotational speed. friction torque setting means for setting the friction torque of the friction type front and rear wheel differential adjustment mechanism according to the difference between the actual rotational speed difference and the ideal rotational speed difference so that the difference approaches the difference; and the friction torque setting means sets the friction torque. and a control means for controlling the friction type front and rear wheel differential adjustment mechanism so as to obtain a friction torque according to the above-described method. I see. A differentially adjustable front and rear wheel torque distribution control device is characterized in that it is configured to set the above friction torque value to a large value. (2) In the friction torque setting means, a dead zone is provided in which the friction torque value is set to 0 in a region where the difference between the actual rotational speed difference and the ideal rotational speed difference is smaller than a predetermined value. The differentially adjustable front and rear wheel torque distribution control device according to claim 1. (3) The above actual rotational speed difference and ideal rotational speed difference are calculated by subtracting the front wheel rotational speed from the rear wheel rotational speed, respectively.
If the difference between the above actual rotational speed difference and the ideal rotational speed difference is the difference between the above actual rotational speed difference and the above ideal rotational speed difference, then the above dead zone is The difference between the actual rotation speed difference and the ideal rotation speed difference is set between 0 and a negative predetermined value, and when the ideal rotation speed difference is negative, the difference between the actual rotation speed difference and the ideal rotation speed difference is set between 0 and a negative predetermined value. The differentially adjustable front and rear wheel torque distribution control device according to claim 2, wherein the difference between the front and rear wheels is between 0 and a positive predetermined value. (4) The differential adjustment type according to claim 3, wherein the negative predetermined value and the positive predetermined value are numbers with opposite signs of the ideal rotational speed difference in each case. Front and rear wheel torque distribution control device. (5) The ideal rotational speed difference is defined as the rear wheel rotational speed minus the front wheel rotational speed, and when this ideal rotational speed difference is 0, the dead zone is omitted. A differentially adjustable front and rear wheel torque distribution control device according to claim 1.