JPH04201729A - Differentially adjusting front/rear wheel torque distribution control device - Google Patents

Abstract

(57) [Summary] This bulletin contains application data before electronic filing, so abstract data is not recorded.

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.

For example, in a so-called full-time four-wheel drive vehicle, a center differential (hereinafter abbreviated as center differential) is used to appropriately distribute the driving force from the engine between the front wheels and the rear wheels. It is conceivable to provide a differential limiting mechanism such as a viscous coupling that limits movement and to adjust the differential limiting mechanism to control the torque ratio according to the operating state.

[Problems to be Solved by the Invention] In the conventional torque distribution control device for front and rear wheels, even when the torque is less than a predetermined value, the differential limiting clutch may rotate at a high difference, which may impair the durability of the clutch. Furthermore, simply releasing the clutch from time to time does not allow the vehicle's running attitude to be controlled.

The present invention was devised in view of the above-mentioned problems, and an object of the present invention is to provide a differentially adjustable front and rear wheel torque distribution control device that can protect a differential limiting clutch and control the running attitude of a vehicle. do.

[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 axle side and the rear wheel side, thereby adjusting the front and rear wheels. A differential adjustment mechanism that adjusts the differential state between the front wheels and the rear axle through a clutch in a four-axle drive vehicle with differential adjustment between the front and rear axles that controls torque distribution to the front and rear axles;
and a control means for controlling the differential adjustment mechanism, and the control means controls the clutch of the differential adjustment mechanism when the differential rotation speed difference between the front and rear wheels continues to be equal to or greater than a predetermined value for a predetermined period of time. It is characterized by being configured to be directly connected for a predetermined period of time.

Furthermore, when the differential rotation speed difference between the front and rear wheels continues to be equal to or greater than a predetermined value for a predetermined time or more, the control means directly connects the clutch of the differential adjustment mechanism for a predetermined time, and then gradually reduces the clutch engagement force. It is preferable to be configured so as to decrease the amount.

[Function] In the differentially adjustable front and rear wheel torque distribution control device of the present invention described above, the differential state between the front wheels and the rear wheels is adjusted through the clutch, but the control means does not control the front and rear wheels. If the differential rotational speed difference between the two continues for a predetermined value or more for a predetermined time or more, the clutch of the differential adjustment mechanism is directly connected for a predetermined time, thereby preventing damage to the clutch due to high differential rotation.

When the control means is configured as in the second claim, sudden changes in the torque control state after clutch protection control are prevented, and each control is performed stably and the vehicle attitude is maintained stably. It can be done.

[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. Figure 5 is a schematic circuit diagram of the hydraulic supply system; Figure 6 is a schematic circuit diagram of the hydraulic supply system; Figure 7 is a cross-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, 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, 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 for longitudinal acceleration, FIG. 15 is a map for setting the clutch torque for longitudinal acceleration, and FIG. 16 is a map for setting the clutch torque corresponding to longitudinal acceleration. FIG. 17 is a diagram showing an example of the transmission torque ratio map, FIG. 18 is a block diagram showing a modified example of the engine torque proportional clutch torque setting means, and FIG. 19 is a diagram showing an example of the transmission torque ratio map. Center differential input torque setting map, Fig. 20 is a characteristic diagram of clutch torque for protection control, Fig. 21(a) is a diagram showing duty characteristics related to the first preload learning, and Fig. 21(b) is a diagram showing its first preload learning characteristic. Fig. 22 is a diagram showing the pressure characteristics related to the second preload learning, and Fig. 23 (a) is the duty characteristic related to the third preload learning. Figures 23(b) and 23(c) are diagrams showing the pressure characteristics related to the third preload learning, Figure 24 is a diagram showing the torque distribution state display means, and Figure 25 is a diagram showing the pressure characteristics related to the third preload learning. is a characteristic diagram for explaining torque distribution by the torque distribution state estimating means, FIG. 26 is a flowchart showing the control flow of the entire vehicle including the device, and FIG. 27 is a flowchart of the front and rear wheel torque distribution control. 28 is a flowchart showing the flow of setting the clutch torque corresponding to the differential, FIG. 29 is a flowchart showing the flow of setting the clutch torque corresponding to longitudinal acceleration, and FIG. 30 is the flowchart showing the flow of setting 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, with reference to FIG. 2, the overall configuration of the cantering system of a vehicle equipped with this differentially adjustable front and rear wheel torque distribution control device will be described.

In FIG. 2, reference numeral 2 denotes an engine, and the output of the engine 2 is 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. 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
Left and right rear @24.2 from axle 25L, 25R through 2
6. The planetary gear type differential device 12 is conventionally qh
Sungear 121. similar to q's. A carrier 125 that supports the planetary gear 122 and includes a planetary gear 122 arranged outside the sun gear 121 and a ring gear 123 arranged outside the planetary gear 122.
The output of the output shaft 8 of the automatic transmission 6 is inputted to the output shaft 8 of the automatic transmission 6, and the sun gear 121 is interlocked with the front circumferential differential gear W14 via the front wheel output shaft 27 and the reduction gear mechanism 19, and
23 is interlocked with the propeller shaft 2o via the rear wheel output shaft 29 and the bevel gear mechanism 15.

In addition, the planetary gear type differential device 14 changes the distribution of engine output torque between the front wheels and the rear wheels by restricting (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 differential limiting means or differential adjusting means.

That is, the hydraulic multi-disc clutch 28 is connected to the sun gear 121 (
The sun gear 121 (or the ring gear 1
23) and the carrier 125 is restricted.

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. from about 32:68 to 50:
It can be controlled between 50 and 50. The value of front wheel:rear wheel in a completely free state: approximately 32:68 can be specified by setting the tooth number ratio of the input gears on the front wheel side and the rear wheel side of the planetary gear, and here, the hydraulic multi-disc clutch 28
The ratio is set to approximately 32:68 when the pressure inside the hydraulic chamber is zero and it is completely free. Also,
This ratio in a completely free state (approximately 32 x 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 (
9 kg/cJ) and the hydraulic multi-plate clutch 28 is in the locked state, and the differential restriction 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 + G yr acting 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 the lateral acceleration data, but only one lateral acceleration sensor 34 is provided near the overlapping part 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; 40
．． 42, 44, and 46 are left front wheel 16 and right front wheel 18 respectively.
, a wheel speed sensor that detects the rotational speed of the right rear wheel 26 and the right rear wheel 28, and the outputs of these switches and each sensor are input to the controller 48.

Reference numeral 50 denotes an anti-lock brake device, and this anti-lock brake device 50 operates in conjunction with a brake switch 50A. In other words, 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.
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. Further, 52 is a controller 48
This is a warning light that lights up based on the control signal.

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.
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 detection means) that detects the selected shift range of the shift lever 160A of the automatic transmission, and this detection information is also sent to the controller 48.

Further, 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. In other words, 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 serving as an engine output control means that is controlled independently of the accelerator pedal system 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 driven by a motor, which is driven by the rear wheel speed sensor 44.
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 parts extending 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 connected through serrations.
58 and the spacer member 115b, the 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), and includes an expanded diameter section that houses the planetary gear element and a rear (leftward direction in FIGS. 3 and 4) In FIG. 4, 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 FIG. 3.4), which will be described later. Further, a plurality of grooves 124d are formed on the outer periphery of the opening.

Center differential 12. 1 is a planetary gear type center differential using a planetary gear mechanism, and includes a sun gear 121 and a plurality of planetary pinions (planetary gears) 122 arranged outside the sun gear 121 so as to surround the sun gear 121.
, a ring gear 123 disposed around the planetary pinion 122, and a planet carrier 125 that supports the planetary pinion 122, each gear being 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 can rotate together. Note that the hollow shaft 145 is formed with a piston accommodating portion 145a, which will be described later.

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 force of the ring gear 123 is applied 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 so as to cover the planetary gear mechanism unit 12 after being installed into the case 115.

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 28a 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 arranged alternately 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 long sword side disk plate 28b.

Note that the first piston 14] 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 a pump 58 that sucks oil in the reservoir 6 pumps oil from its discharge port through a check valve 60 and a pressure control valve 62. and is connected to the hydraulic chamber of 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/a#) to release oil to the reservoir of the automatic transmission 6.
An accumulator 66 and a pressure switch 68 are also connected to this passage. A detection signal from the pressure switch 68 is input to the controller 48. In addition, the pump 58
The motor 58a that drives the 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 includes a solenoid 162a and a valve body 162b that is rotated 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 the return spring 162C to close the oil passage 169f. This reduction valve 162 is electronically controlled by a controller (computer) 48 based on information from various sensors.

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 it exceeds the pressure, it will automatically move to eliminate the oil 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
By moving 1b and 161c to the left, oil passage 169C
and 169g are opened, and the line pressure P works is operating hydraulic pressure (
4WD clutch pressure) P4 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, 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.
As shown in the figure, the lower the duty, the lower the 4WD clutch pressure P4, and the higher the duty, the higher the 4WD clutch pressure P4. It is also conceivable to have the opposite setting, that is, a configuration in which the characteristic is a straight line sloping downward to the right, such that the lower the duty, the higher the 4WD clutch pressure P is, and the higher the duty, the lower the 4WD clutch pressure P4.

Next, the components of the controller related to the control (hereinafter referred to as parking 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 shown in FIG. This will be explained with reference to a block diagram.

In this control, each sensor (wheel speed sensor 40°42.4
4,46. Steering angle sensors 30a, 30b.

30c, lateral acceleration sensor 342 longitudinal acceleration sensor 36,
Throttle position sensor 38. Engine speed sensor 170. 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.

Additionally, ABS information, wheel speed, and steering angle are included in the data.

Comprehensive communication (SCI communication: SCI = Se) between the gear stage, ABS control unit and engine control unit
real communication interface
Data such as ace) are input digitally, and hydraulic control is applied to longitudinal acceleration, lateral acceleration, accelerator opening, and multi-disc clutch.

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

In addition, the clutch torque of the hydraulic multi-disc clutch 28 is set to achieve an ideal differential state 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). ■ Clutch torque Tv for differential to perform control in response to the longitudinal acceleration acting on the vehicle; ■ Clutch torque Tx for longitudinal acceleration to perform control in response to the longitudinal acceleration acting on the vehicle; The engine torque proportional clutch torque Ta is set in proportion to the engine torque so as to obtain a large road surface transmission torque, and the protection control clutch torque Tpc is used to protect the clutch portion of the wet multi-disc clutch. One of these clutch torques Tv, T
The x, Ta, and Tpc setting sections will be explained in order.

The differential compatible clutch torque Tv is a clutch torque that allows the vehicle to behave in accordance with the driver's will when turning.In order to control the attitude of the vehicle, the rear wheels are used as the drive base and slip starts from the rear wheels. Since it is effective to set the clutch torque Tv for the differential, 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 wheel actual rotation speed difference detection section 200, a front and rear wheel ideal rotation speed difference setting section 210, and a front and rear wheel actual rotation speed difference detection section 200. Speed difference ΔVcd and ideal rotational speed difference ΔV of front and rear wheels
A differential compatible clutch torque setting unit 220 that sets clutch torque TV' from hc, and this clutch torque Tv
and a correction section 246 that corrects the lateral acceleration.

The front and rear wheel actual rotational speed difference detection unit 200 includes a filter 2028
~202d, and a front axle rotation speed data calculation unit 204a
, a rear wheel rotation speed data calculation section 204b, and a front and rear axle actual rotation speed difference calculation section 206.

The filters 2028-202d are connected to the left front wheel 16. which is detected by the wheel speed sensors 40, 42, 44, and 46, respectively.
Right front axis 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, in the front wheel axle rotation speed data calculation unit 204a,
The front wheel rotation speed Vf is obtained by averaging the front wheel speeds determined from the front wheel rotation speed data signals FL and FR. The rear wheel rotation speed Vr is obtained by averaging the determined rear wheel speeds.

Furthermore, in the front-rear axis actual rotational speed difference calculation unit 206.

By subtracting the front wheel rotation speed Vf from the rear wheel rotation speed Vr, the actual rotation speed difference between the front and rear wheels [rotation speed difference between the front and rear wheels (front and rear rotation difference: this rotation difference is also the rotation difference at the center differential)
ΔVcd is calculated.

The front-rear axis ideal rotation speed difference setting section 210 includes a driver-required steering angle calculation section (II-like steering angle calculation section) 212 as a steering angle data detection means, and a driver-required vehicle body speed calculation section as a vehicle speed data detection means. (pseudo vehicle speed calculation unit) 216 and ideal rotational speed difference setting unit 218 as an ideal operating state setting unit
It is composed of the following.

As shown in FIG. 8, the driver-required steering angle calculation section 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
0b, detection data θ1 from the neutral position sensor 30c). θ2. Based on θn, the sensor-compatible steering angle δh[
= f(θ0.θ2.θn)], a sensor-compatible steering angle data setting unit 212a, and a lateral acceleration sensor 34.
Data G yf , G yr detected at a, 34b
A lateral acceleration data calculator 8 212b that calculates lateral acceleration data Gy by averaging the lateral acceleration data The driver-required steering angle setting section (vehicle speed data setting section) 212d sets the driver-required steering angle δref in accordance with the driver's requested steering angle δref.

Note that the function δh=f(θ
□, θ2. θn) is determined according to the specifications of the steering wheel angle sensor.

In addition, the sensor-compatible steering angle δh and the lateral acceleration data ay 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 SIG (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)
δref, direction 5IG of sensor-compatible steering angle δh
(δh) is not equal to the direction 5IG(Gy) of the lateral acceleration data Gy, 0 is set to the driver requested steering angle δref.
Set to .

When the direction 5IG (δh) of the sensor-compatible steering angle δh and the direction 5IG (Gy) of the lateral acceleration data ay are not equal, O is set as the driver-required steering angle δref. When performing operations, the steering wheel position and the actual steering angle (turning state) of the vehicle may differ, and in such cases, setting the vehicle steering angle from the steering wheel steering position will ensure proper control. difficult to do.

Therefore, in order to eliminate such problems, the direction 5IG (δh) of the steering angle δh corresponding to the sensor and the lateral acceleration data G
If the y direction 5IG (Gy) is not equal, the driver requested steering angle is set to O.

As shown in FIG. 9, the driver-required vehicle body speed calculation unit 216 calculates the left front wheel 16. Right front axis 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 FR, 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.

Particularly in the driver-required vehicle speed calculation unit 216C.

The wheel speed data SvW selected by the wheel speed selection unit 216a is filtered through a filter 216b to remove noise components, and the longitudinal acceleration detected by the longitudinal acceleration sensor 36 is filtered through a filter 216d to remove noise components. Based on the longitudinal acceleration data Gx, the subsequent vehicle speed is estimated from both data SV W and G x at a certain point in time. In other words, if the wheel speed data SvW at a certain point is ■22 and the longitudinal acceleration data Gx is a, then the theoretical vehicle speed V after a time t after this point
ref can be calculated as Vref=V2+at.

Also, instead of 1 longitudinal acceleration data Gx, wheel speed data S
The driver-required vehicle acceleration V' ref obtained by time-differentiating vW or the driver-required vehicle speed Vref may be used.

Note that the rotational speed data signals FL, FR, and 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 wheels (the so-called inner wheel difference) when turning 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 increases relative to the rotational speed Vf of the front wheels. This makes it easy for the wheels to slip. This ensures the responsiveness of the vehicle body attitude required at high speeds. Also.

Regarding the steering angle, the larger the steering angle BL, the larger the required rotation difference between the front and rear wheels, so the steering angle data δre
The larger the magnitude of f 1δref l, 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 axle wheel speed is Vr, the vehicle speed at the center of gravity of the vehicle is V, the rotation radius of the front wheels is Rf, the rotation radius of the rear wheels is Rr, the rotation radius of the vehicle center of gravity is R5, the vehicle body slip angle is β, and the wheel base is 1. , the distance between the front wheel center and the center of gravity is if, 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]A'ref...(1,1) Note.

Rr=(R"+1r"-2R1r-cos(π/2-β
))””Rf=(R”+l f”2 R1f-e
as(π/2÷β))""β=(11/2l-1fll
r-kr-v)/(l+A-v2)・lr/1・δ However, m is the vehicle weight, kr is the cornering power, and A is the stability factor. Also, the front wheel speed Vf and the rear wheel speed Vr Considering this as a theoretical value, Vf:Vr=Rf:Rr, Vf:V
=Rf:11: Furthermore, there is a relationship between the angles βf and βr shown in FIG. function [A Vhc= f c(V .

δ)]. However, ■ in this case corresponds to the theoretical value, that is, the driver-required vehicle speed V ref, and δ
also corresponds to a theoretical value, that is, the driver requested steering angle δref. Such a function [ΔVhc=fc(Vref, δ
ref)] is mapped, it becomes as shown in Fig. 10.

By the way, regarding the steering angle, in addition to the actual steering angle (sensor-compatible steering angle) δh based on the steering wheel angle θ, the lateral acceleration during turning (turning G) G! There is a turning G equivalent steering angle δy determined from /. This turning G equivalent steering angle δy can be calculated using the following equation.

δy=[(1◆A-Vref2)lvref2]・1-
Gy... (1, 2) where A is the stability factor, Vref is the theoretical vehicle speed (vehicle speed requested by the driver), which will be described later, and ■ is the wheel base.

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, 1δh1>1δy1, and by adopting the steering angle value 1δh1, the magnitude of the ideal rotational speed difference (slip target value) will be greater 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 Iδh
By adopting 1, the magnitude of the ideal rotational speed difference (slip target value) can be made smaller than by adopting the steering angle value 1δy1.

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

The differential compatible clutch torque setting unit 220 sets the clutch torque Tv' in response 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. Wheel ideal rotational speed difference Δ
Clutch torque T is divided into cases depending on the positive/negative of Vhc.
v' is set.

(i) When ΔVhc≧0 In this case, it is desired to make the speed of the rear wheels faster than the speed of the front wheels, and the clutch torque Tv1 is set as shown in (■) to (■) below.

■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
Tv' = aX (ΔVcd - ΔVhc) = a
Set c...(1,3) (where a is a proportionality constant).

■If ΔVhc > ΔVcd > 0, the front axle is slipping, so if clutch torque Tv' is increased at this time, the engine torque distributed to the front wheels will increase and front wheel slip will be accelerated. Become. 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 O 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.

Tv'=-aXΔVcd=-aX(ΔVc+ΔVhc)
...(1, 4) (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(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≧0, 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, so that the clutch torque Tv' increases in proportion to the magnitude of ΔVcd, Tv'=aX ΔVcd=aX(ΔVc+ΔV
hc)... (1, 5) (where a is a proportionality constant).

■If O〉ΔVcd)ΔVhc, the rear wheels are slipping, so if the clutch torque Tν' 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 O 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, the map [ΔV with reference to Figures 13(a) and (b)]
The differential clutch torque Tv' obtained from c and ΔVhc is corrected for lateral acceleration by a correction unit 246.

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

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

The value of 1 for this lateral G gain is intended to reflect the state of the friction coefficient μ of the road surface in the control, and it can be determined that the larger the lateral acceleration Gy is, the larger the road surface μ is, and the larger the road surface μ is, 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 (and therefore the magnitude of the lateral acceleration Gy) increases, the lateral G gain is decreased by 1 to perform correction to decrease the set clutch torque Tv.

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 front wheel shared load calculation means 254a.

It is composed of 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, and this front wheel shared load Wf' is determined by the front wheel shared load Wf at rest, the vehicle weight Wa, the height of the center of gravity, and the wheel. It is determined by the following formula from known numerical values such as base 1 and longitudinal acceleration data Gx.

Wf'=Wf-(h/l)・Wa-Gx" (2,
1) The total output torque calculating means 254b calculates the required total output torque Ta (torque considered on the propeller shaft) from the 1 longitudinal acceleration data Gx, but the required total output torque Ta
are vehicle weight Wa 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/ρ''・(2,2) The front axle shared torque calculation means 254c calculates the front wheel shared load Wf' during acceleration calculated by the front wheel shared load calculation means 254a,
From the required total output torque Ta determined by the total output torque calculation means 254b, the front wheel shared torque Tf is determined using 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 one longitudinal acceleration is calculated Tx.
'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.

Tf=[Zs/(Zs+Zr)]・Ta+[Zr/(Z
s+Zr)l"rx'... (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)]Ta/[Zr
/(Zs+Zr)]...(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
, Ta and solve Tx' for Gx. First, by substituting equations (2, 1) and (2, 2) into equation (2, 3), T f = (Rt/ρ)・(Wf-Gx-h/l
-Wa-Gx”)...(2,5) From equations (2,1), (2,4), (2,5), T
x'=-A-C・(Gx-B/2C)"+AB/4C
...(2,6) However, A=[(Z s + Z r)/ Z rl・
(Rt/ρ)B=Wf-[Zs/(Zs+Zr)]・
Wac=(h/1)・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=0
．． 55 (m), l = 2.6 (m), Tx' = -
40,7(GX-0,552)2+12.4, T
x' can be expressed as a quadratic curve with respect to Gx as shown in FIG.

However, when Gx≠0.55, Tx' takes an extremely thick value, and Gx
In the region >0.55, Tx' decreases, but here,
Considering control safety, T even in the region of Gx>0.55.
X' is set to a constant value equal to the maximum value. In addition,
Such a setting can also be applied when the clutch torque calculating means 254d calculates the longitudinal acceleration-corresponding clutch torque Tx'.

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 section 256.

The correction section 256 performs the same correction as the above-mentioned correction section 246, and performs lateral acceleration correction by multiplying the longitudinal acceleration corresponding clutch torque Tx' by the lateral G gain to obtain the longitudinal acceleration corresponding clutch torque Tx. However,
The lateral G gain of 1 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 the 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 Tx that corresponds to longitudinal acceleration is output only when the front slip occurs, and is not output in other cases (in this case, TX = 0, and hereinafter, generally, the switch is turned off and the clutch torque is When the clutch torque is not applied, the value of the clutch torque is set to ○).

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 ρ of the transmission at that time, and a torque converter torque ratio detection section 276 that detects the torque converter torque ratio t at that time. An engine torque proportional torque setting unit 268 that obtains the engine torque proportional torque Ta' from the engine torque Te based on the map, and
a' is the torque converter torque ratio mentioned above, and the reduction ratio ρm of the transmission.

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 km/h) 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 speed data Ne sent from the engine speed sensor 120 and from which disturbance components are removed through a filter 262b, and engine speed data Ne sent from the transmission 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 l.

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 t, the transmission reduction ratio ρ, the final reduction p1, 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(IΔVhcmaxllΔVhcl)/
IΔVhcmaxl◆0.1...(3,1) However, ΔVhcmax=MAXIΔVhc(δ=MA
X) 1 Also, the coefficient 0.9 and the constant 0.1 are used to set the lower limit of k2 to 0.1.

In this way, as the ideal rotational speed difference ΔVhc increases, the rotational difference gain linearly decreases by 2, and by multiplying and correcting this rotational difference gain, the ideal rotational speed difference ΔVhc during turning etc. When this becomes large, the engine torque proportional clutch torque Ta is reduced so as to prioritize turning performance (performance that can prevent tight corner braking) 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 unit 266, and the transmission reduction ratio ρ sent from the transmission reduction ratio detection unit 276, the center differential input torque (transmission output torque ) Calculate Ta.

Ta=t・ρ−・ρ,・Te...(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 t may be determined from the speed ratio i, and the relationship between Ta and Te may be determined based on this.

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

Tc=[(Zs+Zr)/Zr-Wf/Wa-1ko・T
a --- (a, s) where Zs is the number of teeth of the sun gear, Zr is the number of teeth of the ring gear, Wf is the front wheel shared load, W
a is the 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
It may be determined from m using the following equation.

Tc==[(Zs+Zr)/Zr"Wf/Wa-1
]・tt f) m' /) l" Te... (3,4
) Further, the switch 274a is activated by a signal from the determining means 274 to determine when the vehicle speed is low (in this example, V ref (20h/
h), the engine torque proportional clutch torque Ta can be output as data, but if the vehicle speed becomes higher than this (Vref≧20km/h), the O
becomes FF, engine torque proportional clutch torque Ta
Stop outputting as data. 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 rpc for protecting the clutch portion of the wet multi-disc clutch 28 will be explained. This clutch torque rpc is set by the protection control section 230.

In other words, in the wet type multi-disc clutch 28, if the differential rotation between the clutch plates becomes large, there is a risk of damage such as seizure of the clutch facing or increased wear.Of course, if the differential rotation is large, the duration of this state will be long. The larger the size, the more likely it is to cause damage. - On the other hand, in order to avoid such a situation and protect the clutch 28, it is possible to make the clutch free (to release the connection between the clutch plates). If this is done instantaneously, there is a risk that the vehicle's attitude may 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 section 230, the front and rear wheel actual rotational speed difference calculation section 20
In response to the front and rear wheel actual rotational speed difference Vcd calculated in step 6, this front and rear wheel 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 protection control clutch torque Tpc 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 rpc is first set for a short period of time (in this example, for 1 second).
is set to the upper limit value, and then gradually decreased to 0 (natural release). In this example, the relationship between rpc at the time of decrease and time 11 is as shown in the following equation.

rpc= 40-14 tt...(
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 according to the characteristics of each vehicle. is desirable.

Further, if the above-mentioned detection condition is not satisfied, the value of the protection control clutch torque rpc is set to zero.

The above-mentioned differential clutch torque Tν2, clutch torque corresponding to longitudinal acceleration Tx, engine torque proportional clutch torque Ta, and protection control clutch torque rpc are each set for each control cycle that is repeated at an appropriate timing, Each clutch torque Tv+Tx, Ta, and Tpc set in this way is sent to the maximum value selection section 280.

In this maximum value selection section 280, for each control cycle,
The maximum clutch torque (this clutch torque is designated as Tc) is selected from among clutch torques Tv, Tx, Ta, and Tpc. 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 manner is sent to the torque-pressure converter 282, where the latch control pressure P is adjusted so that the set clutch torque Tc is obtained.
c is now set.

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.

Further, 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 by the centrifugal correction pressure setting section 286 from the clutch control pressure Pc.
A map such as that shown in block 286 of the diagram
It is determined from the front wheel vehicle speed Vf calculated in step 04a. this is,
Since the piston chamber rotates in synchronization with the front wheel side shaft, centrifugal hydraulic pressure.

This is because the centrifugal correction pressure Pv is generated in response to the front wheel speed Vf, and the centrifugal correction pressure Pv is set to be proportional to the square of the front wheel speed Vf.

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 various parts and changes over time have a large effect on the state of clearance between the clutch plates.

For this reason, it is necessary to constantly maintain the contact between the clutch plates 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 Pi 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 Pi from the preload learning value), and here, three types of preload learning will be explained.

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. Therefore, the conditions for preload learning are set as follows, for example.

■More than 30 minutes have passed since the ignition key was turned on.

■Shift selector is 1 (1st speed), 2 (2nd speed), D
(drive) or neutral (N). In this example, there is no range for P (parking) and R (reverse).
2. This is because a different hydraulic pressure is output than in the case of D and N.

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

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

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

First, as shown in FIG. 21(a), the multi-disc clutch 28
Apply a duty for 2 seconds equivalent to a pressure [for example, P = 0.4 kgf/3 @] that is greater than the biasing pressure of the return spring of the clutch 28 and smaller than the designed initial engagement pressure of the clutch 28. , after this 5 e.g. 1.5%
/S increasing rate, e.g. P = 3, 0 kgf/c
Sweep slowly until the duty is equivalent to "g".

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, and the pressure P also gradually increases, until the hydraulic piston reaches a certain position. When 141 and 142 move, the clutch plates come into contact and the pressure P
The force of the return spring is also applied to the , 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 increases, 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 P1.

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

Furthermore, once the above-mentioned preload learning is performed when the ignition key is turned on, it will not be performed again unless the ignition key is turned off and then turned on again.

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 we would like to perform this preload learning several times, the conditions for the preload learning described above are slightly relaxed, 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= Oks/h (vehicle speed Vref is O
).

■Tc≦1kgfm [Clutch torque Tc must be less than 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 P1 (= Pl
) for a predetermined period of time (e.g. 2
2 seconds), and then for a predetermined period of time (for example, 1 second), the duty equivalent to P = 8.8 kgf/mouth 2 (habo 1
00% duty).

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

In other words, if the clutch is released at the initial pressure P1,
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-state maximum pressure Pc (known value, here 8°8kgf/a1
peak value (maximum value) P that is larger than a certain level (approximately 12)
■ax occurs.

On the other hand, if the clutch is in contact and dragging at the initial pressure P, 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 stored, and the difference α (=Pax−Pc) between this value Pm+ax and the steady maximum pressure 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 trial at appropriate time intervals (for example, every 5 minutes) while increasing or decreasing the starting pressure P from the initial value P□ as appropriate. .

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 P I NT
G (n) and initial pressure Piii-PINT (n
). Therefore, the previous initial learning value is P
INTG (n-1), initial pressure is PINT (n
-1), and in the n-th learning stage, the previous initial pressure is learned using rINT(n-1).

Then, the difference α (= PIlax - Pc) obtained by a predetermined duty sweep and the threshold value α. The current initial learning 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=PINT
G (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 worth of pressure), and the initial pressure PINT(n) is set by adding β (= 1 bit of pressure) to the previous initial learned value PINTG (n-1).
1 bit of pressure), that is, the current initial learned 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 P I NTG (n-1) by β.
(=1 bit of pressure) is added.

Note that 1 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/am" or 1 bit.
Bit=Q, set to an appropriate value such as 1 kgf/am.

On the other hand, α〈α. When , the initial learning value PINTG
For (n), the previous initial learning value PINT
It is set to G (n-1) plus β (=1 bit), but the initial pressure PINT (n) is set to the previous initial learned value PINTG (n-1).

This is αkuα. 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
INTG (n) is the previous initial learning value PINTG
(n-1) 4Cβ (=1 bit) is added, but the initial pressure PINT(n) is set to the previous initial learned value PINTG (n-1).

This is done by α〈α. From the results alone, clutch 28
It is not possible to determine whether the contact is in the marginal contact state or in the excessive contact state, which may lead to chattering, so in order to avoid this, the results of this learning should not be immediately adopted as the initial pressure Pi. , the previous learning value is used.

Therefore, if there is excessive contact, α will continue for at least two consecutive cycles. It is considered that this state will continue, and the initial pressure Pi will be reduced and approach the appropriate value, albeit 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 O
).

■Tc≦1kgfm [Clutch torque Tc must be less than 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-mentioned conditions (■' to ■) is satisfied at the same time, 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 □ and hold this for a predetermined time (e.g. 2 seconds), and then for a predetermined time (e.g. 1 second) P =
8, 8 kgf/an” equivalent duty (Habo 1
00% duty), P =
A duty corresponding to 8.8 kgf/CM" is maintained for a predetermined time (for example, 2 seconds). This pattern is continuously repeated while changing the initial pressure Pi as appropriate.

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

Then, the time point to (
Or from the point in time (t) when the pressure P starts to rise, the straight line LO
Until the steady maximum pressure Pc (or a constant pressure value close to this) is reached, the area surrounded by this straight line LO and the curve L1 or L2 that depicts the state of change in pressure P (the diagonal line in the figure) Comparing the areas SL and S2 of curve L1 with an overshoot, the area S1 in the case of the curve L1 with an overshoot is clearly larger than the area S2 in the case of the curve L2 without an 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 Pi. 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. This initial learning [
P I NT G (n) and initial pressure PINT
(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<So.

PINTG (n)=PINTG (n-1)-βPI
NT (n)=PINTG (n-1) That is, S
2: In the case of S, α≧α of the second preload learning. Corresponding to the case, when S<S, α<α of the second preload learning. Corresponds to the case of

That is, when S≧80, 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, this time's initial learning value P
INTG(n) is the previous initial learning value PINTG
(n-1) plus β (=1 bit of pressure), and the current initial pressure P'INT(n) is added to the previous initial learned value PINTG (n-1) by β (=1bi).
t of pressure) is applied.

On the other hand, when S<S, 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 pxNT(n-1). Therefore, this time's initial learning value PI
NTG (n) is the previous initial learning value PINTG
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). The reason for doing this is also the aforementioned α〈α. As in the case of S
<S, it is not possible to determine whether the clutch 28 is engaged, close to contact, or excessively in contact based only on the result of <S.
Since there is a risk of causing chattering, in order to avoid this, the previous learning value is used instead of immediately adopting the current learning result as the initial pressure Pi.

Therefore, if there is an excessive contact state, it is thought that the state of S<S will continue for at least two consecutive cycles, and the initial pressure Pi will be reduced and approach the appropriate value, although it will be delayed by one cycle. I'm going to go.

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.

In addition, in this third preload learning, the area 5 (Sl) of the portion 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', S2').

In this case, the calculation of the area S' starts at the time to when the duty sweep starts (or the time tx when the pressure P starts to rise), and the end of the calculation of the area S' is at the steady state shown by the straight line LO. This is the point in time when the maximum pressure Pc (or a constant pressure value close to this) is reached. Then, set the judgment standard value to 8
0', it can be determined that there has been an overshoot when S'≧L/, and it can be determined that there has been no overshoot when s'<s,'.

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/c++”/s), and set a slightly lower limit speed (e.g. 15,7 k
g/ax”/s).

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.

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

Note that the switch 292a is turned off by 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 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 the signal from the 1 determining means 294, and when the set clutch torque TC is low at low speed, This is an attempt to reduce the duty to O. As a low speed condition, for example, Vref≦5km/h,
For example, the condition for clutch torque TC is Tc≦1k.
It can be stipulated that it is gfm. And this 2
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-duty conversion section 298.

The pressure feedback 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. Note that the signal sent from the pressure sensor 304 to the pressure feedback correction section 296 is filtered by a filter 306 to remove noise components due to disturbances and the like.

The pressure-duty converter 298 sets (outy) corresponding to the control pressure P feedback-corrected by the pressure feedback corrector 296, and follows the map shown in the block of the clutch pressure car duty converter 298 in FIG. As shown, the duty increases linearly with respect to the 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.

That is, as shown in FIGS. 1 and 24, a torque distribution display section 312 is provided in the meter cluster to graphically display (or meter display) the state of torque distribution to the front wheels (or rear wheels). The torque distribution state is displayed 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 front wheel output torque in each of these cases. (or rear wheel output torque), a selection means 310b for selecting the smaller front export force torque (or rear export force torque), and these portions 310a,
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, we will consider torque distribution in each of these cases and when the state changes.

First, as a prerequisite, when differential restriction is not performed as in this four-wheel drive system, the rear wheels are set mainly (the torque ratio between the front wheels and the rear wheels is 32:68, for example), and furthermore,
The differential limiting clutch 28 is assumed to always transmit torque from the rear axle side to the front wheel side, and is set as follows for the sake of simplicity.

p f/ r f< p r ・p t/ r r
”・(5,1) However, ρf: Front differential ratio ρr: Rear differential ratio pt: Transfer ratio rf: Front axle tire radius rr: Rear axle tire radius. If the clutch does not slip, direct 4-wheel drive distribution Therefore, the front wheel torque Tf and the rear wheel torque Tr are as follows.

Tf = Wf/Wa(Tm++ k WrIIrf/ p e
(rf p r p t/rr p t-1))...
(5,2) rr = Wr/Wa・(Tm−k W?rr/p ・(rf
p r p t/rr p t-1))...(5,
3) However, Wf: Front axle shared load Wr: Rear axle shared load Wa: Vehicle weight (=Wf+Wr) Tm: Mission output torque (=center differential input torque) k Nislip ratio coefficient ρ: Final reduction ratio [=(ρf+ρr・pt )/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, 5) However, Tc: Clutch transmission torque capacity a: Number of sangya teeth b: Number of ring gear teeth And, if the latch slips as described above, the front and rear torque caused by the weight distribution, differential ratio difference, etc. This means that the clutch allows for the difference.

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)
・If (5, 6) Tf>Tf', the clutch is in a slip 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 Figure 1, when the pressure P is 2 kgf/cn'' (P=2
) and 8 kgf/am” (P==8).

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, when P is 2 kgf/m'', in a region where the torque Te is smaller than Te1, the straight line marked as direct connection is located below the curve of P=2.

The front wheel torque distribution ratio m follows this straight line. Furthermore, in a region where the torque Te is larger than Te1, 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/am'', the direct connection straight line is always lower in the region shown in this graph, so the front wheel torque distribution ratio m follows the direct connection.

Once the front wheel torque distribution ratio m is 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 axle. In this example, the torque distribution to the front wheels is about 32% to 50%, so the torque distribution display section 312 is marked with a scale corresponding to this, and the lamp is lit or the pointer is turned on until the corresponding scale is reached. Display it in an easy-to-understand manner by moving it around.

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.

First, the flow of the entire operation of the drive system is as shown in Fig. 26. First, each control element is initialized (step a1), the steering angle neutral position is learned (step a2), and the clutch preload is set. Learning (step a3) is performed, and then front and rear wheel drive 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 a predetermined time (15m5ec) has elapsed, and when the predetermined time (15msec) has elapsed, a runaway check is performed by the watchdog (step a15).
), return to step a2 above and perform steps a2 to a1.
Repeat the series of controls in step 3.

In other words, the above-mentioned front and rear wheel drive force distribution control, rear differential control, and engine output control are performed at a predetermined period (15 msec).

Of these, the front and rear wheel drive force distribution control will be explained with reference to the flowchart of FIG. 27.

As shown in FIG. 27, first, wheel speeds FR, FL, RR
, RL, rudder angle θ1. θ21 On! Lateral acceleration G Y
Detect and import data such as + longitudinal acceleration Gx, throttle opening θ, engine speed Ne, and transmission speed Nt9 selected shift stage (step b1
), from these data, front wheel speed Vf, rear wheel transport speed Vr, driver requested vehicle speed V ref , driver requested steering angle δ
ref etc. are calculated (step b2).

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

Furthermore, in steps b6 to b9, 1 is added to the actual rotational speed difference ΔV c , the ideal rotational speed difference ΔVhc, and the lateral G gain.
Calculate clutch torque Tv corresponding to the differential (converted from the map in this example), calculate clutch torque Tx corresponding to the longitudinal acceleration from the longitudinal acceleration Gx and lateral G gain (converted from the map), and calculate the throttle opening. θth, 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 value from these clutch torques Tv, Tx, Ta, and Tpc is calculated as the set clutch torque Tc.

Furthermore, in step bll, the clutch torque Tc determined in this manner is converted into a clutch engagement pressure Pc using a map.

Next, preload correction (adding preload Pi) to this pressure Pc
and centrifugal pressure correction (reducing centrifugal pressure Pv) (step b12) to obtain center differential control pressure Pcd.

Furthermore, by connecting the peak hold filter, the pressure P
(step b13)
).

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

The center differential control pressure Pcd set in this way is
In step b16, pressure foot bath correction is performed. In other words, the difference Δ between the value of Pcd 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 one-operation limited clutch.

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

First, calculate the difference AVcd (=Vr - Vf) by subtracting the front wheel speed Vf from the rear wheel speed Vr (step cl).
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 rotation speed difference ΔVhc between the front and rear axes is 0 or more. If ΔVhc is 0 or more, the process proceeds to step C4, and if ΔVhc is less than 0, the process proceeds to step c5.

Proceeding to step c4, clutch torque Tv' is set from ΔVc using a map (see FIG. 13(a)).

Specifically, if ■ΔVcd≧ΔVhc, the clutch torque Tv' is the difference A Vc (A Vcd −ΔVhc)
Tv' = aX (ΔVcd - ΔVhc) = aXΔVc so that Tv' increases in proportion to the magnitude of (where a is a proportionality constant).

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

Furthermore, if ■O≧ΔVcd, clutch torque Tv'
Tv'=
Set -aXΔVcd=-aX (ΔVc+ΔVhc) (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

Specifically, if ■ΔVcd≧0, the clutch torque T
Tv'=aXΔVcd=aX(ΔVc+ΔVhc) is set so that v' increases in proportion to the magnitude of ΔVcd (where a is a proportionality constant).

If 0>ΔVcd>ΔVhc, the clutch torque Tv' is set to 0, and a so-called dead zone region is set.

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

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 wheels, 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 ay are not equal, the driver-required steering angle is set to O. 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.

Further, since the second largest wheel speed data from the bottom among the rotational speed data signals PL, FR, RL, and RR is used as the driver-required vehicle speed Vref, reliability of the data is ensured.

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 a map (FIG. 15) (step di).

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

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, and switch 2
58a, when the front wheel speed Vf is greater than the vehicle body speed Vref, that is, when the front wheels are slipping (
When the front wheel speed Vf is not greater than the vehicle speed Vref, that is, when the front wheels are not slipping, the clutch torque Tx corresponding to the longitudinal acceleration is used as control data. Torque Tx is set to O (step d4).

As a result, when accelerating, such as when front slips, the torque distribution is the same as in direct-coupled 4WD.

More torque than this will be distributed to the base distribution ratio (toward the rear wheels), preventing the vehicle from becoming too underpowered and allowing smooth turning.

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 e1).

Next, the engine torque proportional torque setting section 268 reads the engine torque proportional torque Ta' from the engine torque Te through a 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′ obtained in this manner, the torque converter torque ratio t, the transmission reduction ratio detection unit 276, the transmission reduction ratio ρm, the final reduction ratio ρ1, and the rotation difference gain setting unit 275. From the obtained rotational difference gain and 2, the center differential input torque (transmission output torque) Ta (=t・ρm・
ρ1·Te) is calculated (step e4).

Further, in step e5, when the vehicle speed is low (in this example, Vre
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 (Vref ≧ 201 s/h), , O is set as the engine torque proportional clutch torque Ta (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 calculation of the protection control clutch torque rpc described above is performed as shown in FIG. 31.

First, in step fl, 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, and is set to 0 when overall control is started.

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 rpc is set to O, and further, in step f14, the flag FLG is set to zero.

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 (8,6)on/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 (1sec), 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. Then, in step f14, the flag FLG is set to O.

During several control cycles, if the actual rotational speed difference Vcd of the front and rear wheels 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).
It is determined whether the timer value exceeds the reference time or not, and 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 Tpc 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 TPC: 40 continues and the timer value becomes 2 seconds or more, the process proceeds from step f6 to step f7, and the protection control is performed based on the relationship rpc = 40-14 X (timer value - 2). The value of clutch torque rpc 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.

After finishing the timer count and clearing the timer, in step f12, the protection control clutch torque rp is
The value of c is set to 0, and in step f14, the flag FL
Let G be 0.

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 Tpc is set to the upper limit value for 1 second) and then gradually decreases to 0 (natural release).

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.

The above-mentioned preload correction will now be explained with reference to FIGS. 32 to 34.

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

If all of these conditions are satisfied, the process proceeds to step g5, and if any of these conditions is not satisfied, learning control is not performed 5).

Proceeding to step g5, it is determined whether preload learning was performed after the ignition key was turned on, and if preload learning has already been performed, learning control is not performed and preload learning is performed. If so, 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, 4 kgf/ci2 equivalent duty (d
uty) for 2 seconds, and then, for example, 1.5%/S
For example, P ” 3 , 0 kgf/am
``Slowly sweep it up to a considerable duty.

On the other hand, the hydraulic piston 141 changes as shown in FIG. 21(b). ．． 142 to this pressure P
The maximum value of the second-order differentiated value (difference) P'' with respect to time and the pressure P at this time are stored.

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 (for example, 5 minutes or an appropriate period shorter than this) has elapsed since the shift selector is set to 1 (1st gear), 2 (2nd gear),
Whether any of D (tribun, neutral to N) is selected, ■Vref= Okm/
h, and whether ■Tc≦1kgfm is determined. If both of these conditions are met, the process proceeds to step h6, and if any of these conditions are not met, 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 (duty) corresponding to the preset initial initial pressure P□ is maintained for a predetermined time (
For example, it is held for 2 seconds), and then it is swept for a predetermined time (for example, 1 second) to a duty equivalent to P=8.8 kgf/2 (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 next 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 the difference (Pmax-P
c) This α is the threshold value (α.

), it can be determined that there was an overshoot. Therefore, it can be determined that the clutch 28 is disengaged at the initial initial pressure P□, and if this α is not larger than the threshold value, an overshoot has occurred.

In other words, it can be determined that the clutch 28 is in a marginal contact state or an excessive contact state during the initial initial pressurization.

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 = PINT
G (n) In other words, for the initial learning value PINTG (n), the previous initial learning value PINTG (n-1)
The initial pressure PINT (n) is set by adding β (==1 bit of pressure) to the previous initial learned value PINTG (n-1). In other words, the current initial learning value PINTG (n) is set.

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

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 PINT
Set to G (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 above-mentioned preload learning conditions
As long as ■ is satisfied, it 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. In other words, in steps h1 to h5, ■' Check whether 10 minutes or more have passed since the ignition key was turned on. ■Whether the shift selector is set to 1 (1st gear) or 2 (2nd gear)
, D (drive), N (neutral), ■Vref= Okm/
h, and whether ■Tc≦1kgfa is determined.

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 (duty) corresponding to the preset initial initial pressure P is set 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/cm'' (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), from the time point to when the duty sweep starts (or the time point t when the pressure P starts to rise), the steady maximum pressure Pc as shown by the straight line LO (or a constant pressure value similar to this) , S2).

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 clearly larger than the area S2 in the case of the curve L2 without overshoot, the presence or absence of an overshoot is determined by comparing the area S with the threshold value S0.

Therefore, if the area S is larger than the threshold value S, the process proceeds to step h8.

PINTG (n)=PINTG (n-1)+βPI
NT (n)-PINTG (n-1)+β=PINT
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 (=1b) 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, if the area S is not larger than the threshold S0, 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 PI NTG (n), β(
= 1 bit), but the initial pressure PINT(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 Pi 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.

In particular, in the judgment based on the integral value (area), it is possible to relatively appropriately judge whether the initial pressure Pi is appropriate.
Initial pressure P without relying heavily on pressure sensor ability
i can be set.

Furthermore, 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 provided to graphically display (or meter display) the state of torque distribution to the front @ (or rear 1), 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 driving more enjoyable and greatly improving marketability.

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 wheel torque distribution control device of the present invention, by adjusting the differential state between the front wheels and the rear wheels, A front and rear wheel differential adjustment type 4@Note vehicle that controls torque distribution to the wheels includes a differential adjustment mechanism that adjusts the differential state between the front wheels and the rear wheels through a clutch, and the differential adjustment mechanism described above. and a control means for directly connecting the clutch of the differential adjustment mechanism for a predetermined time when the differential rotation speed difference between the front and rear wheels continues to be equal to or greater than a predetermined value for a predetermined time or more. This structure prevents damage caused by high rotation of the clutch and improves the durability and reliability of the device.

When the control means is configured as in the second claim, a sudden change in the torque control state after clutch protection control is prevented, each control is performed stably, and the stability of the vehicle attitude is maintained. , driving performance can be significantly improved.

[Brief explanation of the drawing]

1 to 34 show a differentially adjustable front and rear wheel 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,
Figure 6 is a circuit diagram of the main part of the oil pressure supply system, Figure 7 is a diagram showing the characteristics of the oil pressure setting duty, Figure 8 is a block diagram showing details of the total steering angle data detection means, and Figure 9. 10 is a block diagram showing the details of the vehicle speed detection means, FIG. 10 is a diagram showing its ideal rotation speed difference setting map, FIG. 11 is a diagram showing its lateral acceleration gain setting map, and FIG.
Both a) and (b) are plan views schematically showing the wheel conditions for explaining the ideal rotational speed difference, and FIG. 13(a),
(b) is a diagram showing a map for setting the clutch torque corresponding to the differential, FIG. 14 is a block diagram showing the clutch torque setting means corresponding to longitudinal acceleration, and FIG. Figure 16 is a diagram showing an example of the engine torque map, Figure 17 is a diagram showing an example of the transmission torque ratio map, and Figure 1 is a diagram showing an example of the transmission torque ratio map.
Fig. 8 is a block diagram showing a modification of the engine torque proportional clutch torque setting means, Fig. 19 is a center differential input torque setting map, Fig. 20 is a characteristic diagram of the clutch torque for protection control, and Fig. 21 (a). is a diagram showing the duty characteristic related to the first preload learning, FIG. 21(b)
is a diagram showing the pressure characteristics applied to the first preload learning, and the second
Figure 2 is a diagram showing the pressure characteristics related to the second preload learning, Figure 23 (a) is a diagram showing the duty characteristics related to the third preload learning, and Figures 23 (b) and (c) are Both are diagrams showing the pressure characteristics applied to the third preload learning, and the second
FIG. 4 shows the torque distribution state display means, FIG. 25 is a characteristic diagram for explaining the torque distribution by the torque distribution state estimation means, and FIG. 26 shows the control flow of the entire vehicle including the device. 27 is a flowchart showing the flow of the front and rear wheel torque distribution control, FIG. 28 is a flowchart showing the setting of the clutch torque corresponding to the differential, and FIG. 29 is the flowchart showing the setting of the clutch torque corresponding to the longitudinal acceleration. Flow chart showing the flow, No. 30
The figure is a flowchart showing the flow of setting the engine torque proportional clutch torque, Figure 31 is a flowchart showing the flow of setting the protection control clutch torque, and Figure 32 is the flowchart of the first preload learning. FIG. 33 is a flowchart showing the flow of the second preload learning, and FIG. 34 is a flowchart showing the flow of the third preload learning. 2-engine, 4-torque controller, 6-automatic transmission, 8-output shaft, 10-intermediate gear (transfer idler gear), 12-center differential (center differential), 14=-differential gear for front shaft Device, 15-bevel gear mechanism, 15A-bevel gear shaft, 15a--bevel gear, 16.18-front shaft, 17L, 17R--front wheel side axle, 19-reduction gear mechanism, 19a-output gear, 2
0- Propeller shaft, 21- Bevel gear mechanism, 22
-Differential gear device for rear wheels, 24.26-・Rear wheels, 25L
, 25R--1 rear wheel axle, 27--front wheel output shaft, 2
7a--Hollow shaft member, 28--Differential limiting mechanism,
28a-・Front export force side disc plate, 28b-
Input end disc plate, 29-output shaft for rear wheels, 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, 40.42.44.
46-wheel speed sensor, 48-controller, 50-anti-lock brake device, 50A--brake switch,
51--brake pedal, 52-warning light, 54-hydraulic source, 56-pressure control valve system (pressure control valve), 58-pump, 60-check valve, 62-pressure control valve, 64-relief valve, 6-low accumulator, 68-pressure switch,
68a-motor, 113-human power gear, 114a-114
f--bearing, 115-transmission case, 11
5a--end cover, 115b--spacer member, 11
6-Supporting member, 117a, 117b-・Oil passage, 121-
- Sun gear, 122...Planetary pinion (planetary gear), 123-Ring gear, 124=Input case, 125...Planet carrier, 125a-Base plate part, 125b-Planetary pinion housing part,
l 25 f-・Clutch disc support part, 126...
- Pinion shaft, 130 - Connection member, 141...-
First piston, 142-/Second piston, 143-/Partition plate 5144a-First oil chamber, 144b = 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), 163-orifice, 164-oil filter, 165-reducing valve, 170-engine speed sensor, 180-
- Transmission rotation speed sensor, 200-= Front and rear wheel actual rotation speed difference detection section, 202 a to 202 d- Filter, 204 a - Front wheel rotation speed data calculation section, 2
04 b - Rear wheel rotation speed data calculation unit, 206.
... Front and rear wheel actual rotational speed difference calculation unit, 210 - Front and rear wheel ideal rotational speed difference setting unit, 212 - Driver requested steering angle calculation unit (pseudo steering angle calculation unit) as steering angle data detection means
, 212a--sensor compatible steering angle data setting section, 21
2 b - Lateral acceleration data calculation section, 212 c - Comparison section, 212 d - Driver requested steering angle setting section (vehicle speed data setting section), 216 - Driver requested vehicle body speed calculation section as vehicle speed data detection means (pseudo vehicle body speed calculation unit), 216a-one wheel speed selection unit, 216c-driver requested vehicle speed calculation unit,
216 d-filter, 218--ideal rotation speed difference setting section as ideal operating state setting section, 22o--differential clutch torque setting section, 222--subtractor, 230--protection control section, 242--filter, 244- - Lateral G gain setting section, 246-correction section, 254-clutch torque setting means corresponding to longitudinal acceleration, 254a-front wheel shared load calculation means, 2
54b--total output torque calculation means, 254cm front wheel shared torque calculation means, 254d--clutch torque calculation means, 256--lateral acceleration corresponding correction section, 258a--.
-Switch, 258...-Judgment means, 264--Engine torque detection section, 266-.)-Lecondreque ratio detection section, 267"-Center differential input torque calculation section, 268.
-Engine torque proportional torque setting unit, 269-・Clutch torque calculation unit, 270-・Engine torque proportional clutch torque calculation unit, 272-・Turning 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 part, 288--initial engagement pressure setting part (pre-pressure setting part), 290-peak holt filter, 292a, 294a-switch, 295-duty setting part, 292° 294--judgment means, 296-.
-Pressure feedback correction section, 298--Pressure-duty conversion section, 300-1302--Duty solenoid, 304-Pressure sensor, 306--Filter, 31
0--Torque estimating means, 310 a-Calculating means, 3
10 b - selection means, 312 - torque distribution display section, A
M - Mechanical part of differentially adjustable front and rear wheel torque distribution control device. Fig. 5 Fig. 7 Fig. 8.8 P (Kgf/cm ) I:in Fig. 9, 216 1'216d Fig. 10 Fig. 11 M VnCmOX Fig. 12 β 132 Fig. 15 Fig. 16 Fig. 17 Fig. 19 Fig. 2 o Fig. 21 Fig. 0 Haruma t (sa) Fig. 22 Suea I%! ! 1st episode Figure 23 CG) Figure 23 (b) (C) Sl' t (excerpt) Figure 24 Figure 25 1e' Tor 7 (Te) Figure 28 Figure 29 Figure 30 Figure 310 Figure 32 3311ffl 'M34 figure

Claims (2)

[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 differential adjustment mechanism that adjusts the differential state between the
and a control means for controlling the differential adjustment mechanism, and the control means controls the clutch of the differential adjustment mechanism when the differential rotation speed difference between the front and rear wheels continues to be equal to or greater than a predetermined value for a predetermined period of time. A differentially adjustable front and rear wheel torque distribution control device, characterized in that it is configured to be directly connected for a predetermined period of time. (2) When the differential rotational speed difference between the front and rear wheels remains at a predetermined value or more for a predetermined period or more, the control means directly connects the clutch of the differential adjustment mechanism for a predetermined period, and then gradually increases the clutch engagement force. 2. The differentially adjustable front and rear wheel torque distribution control device according to claim 1, wherein the differentially adjustable front and rear wheel torque distribution control device is configured to decrease the amount of torque.