JPH05178110A - Drive force distribution control device for differential adjusted vehicle - Google Patents

Abstract

PURPOSE:To provide a drive force distribution control device for a differential adjusted vehicle which controls the drive force distribution between two wheel systems by adjusting the differential state between the two wheel systems, which device is adapted to control the driving force distribution of wheels while protecting a differential limiting clutch. CONSTITUTION:A drive force distribution control device for a differential adjusted vehicle comprises a differential adjusting mechanism 28 for adjusting the differential state between two wheel systems through a clutch and a control means 48 for controlling the differential adjusting mechanism 28, wherein the control means 48 includes a clutch protection control unit 230 for controlling the clutch 28 to be directly-coupled for protecting the clutch 28 when the quantity of heat generated in the clutch 28 is more than a designated value. The clutch protection control unit 230 is constructed so that after clutch direct- coupling control, the engagement force of the clutch 28 is gradually decreased, and the slip amount of the clutch 28 is more than a designated value in the course of decreasing the engagement force, the clutch direct-coupling control is again conducted.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a differential adjustment type driving force distribution control for a vehicle, which adjusts the differential state between two wheel systems to control the driving force distribution between these wheel systems. More particularly, the present invention relates to a drive force distribution control device for a differential adjustment type vehicle, which is suitable for use as a front and rear wheel drive force distribution control device in a four-wheel drive vehicle.

[0002]

2. Description of the Related Art Power transmission of an automobile configured to control a ratio of a driving force (also referred to as a driving torque or a torque) transmitted to a front wheel side and a torque transmitted to a rear wheel side according to a driving state. Various devices are known. For example, in a so-called full-time four-wheel drive vehicle, a center differential for appropriately distributing the torque from the engine to the front wheel side and the rear wheel side, and a viscous coupling for limiting the differential in this center differential, etc. A differential limiting mechanism may be provided and the differential limiting mechanism may be adjusted to control the torque ratio according to the operating state.

In addition to such driving force distribution control to the front and rear wheels, driving force distribution control to the left and right wheels can be considered.

[0004]

However, in each of the differential limiting mechanisms as described above, the differential control characteristic is determined by the physical properties and the like, and the differential control is not always positively performed and is always always appropriate. Torque distribution control is not possible. Therefore, as a differential limiting mechanism, between two rotating members that generate a differential (for example, between a front wheel side rotating member and a rear wheel side rotating member, or between a right wheel side rotating member and a left wheel side rotating member), For example, a mechanism for limiting the differential by interposing a wet multi-plate clutch or the like and adjusting the engagement state of this clutch is conceivable.

By the way, when such a clutch mechanism is used, the torque is often transmitted while the clutch is frictionally increasing, and the load on the clutch is increased. Therefore, there is a problem in terms of the durability of the clutch. Moreover, it is desired that the control for protecting the clutch does not impair the control of the clutch for improving the originally intended running performance of the vehicle.

The present invention has been devised in view of the above-mentioned problems. It is possible to appropriately control the driving force distribution of the wheels while protecting the differential limiting clutch and ensuring the durability of the device. An object of the present invention is to provide a driving force distribution control device for a dynamic adjustment type vehicle.

[0007]

For this reason, the drive force distribution control device for a differential adjustment type vehicle of the present invention adjusts the differential state between two wheel systems so that these wheel systems are adjusted. In the differential adjustment type vehicle driving force distribution control device for controlling the driving force distribution, the differential adjustment mechanism for adjusting the differential state between the two wheel systems through the clutch, and the control for controlling the differential adjustment mechanism. Means for controlling the clutch to directly connect the clutch in order to protect the clutch when the load on the clutch becomes excessive, and the clutch protection controller controls the clutch protection controller. After performing the direct connection control of the clutch, the clutch engaging force is gradually reduced, and the clutch is directly connected again when the slip amount of the clutch exceeds a predetermined value while the engaging force is decreasing. It is characterized by being configured to perform control.

It is preferable that the predetermined value is a target slip amount of the clutch. The determination that the load on the clutch is excessive is made by the fact that the state in which the rotation difference occurring in the clutch is larger than the reference value has continued for a reference time or more, or the amount of heat generated in the clutch is calculated and this heat generation is calculated. It can be performed when the amount exceeds a predetermined value.

[0009]

In the above-described differential adjustment type driving force distribution control device for a vehicle according to the present invention, the differential state between the two wheel systems is adjusted through the clutch of the differential adjustment mechanism so that the difference between the two wheel systems is adjusted. The distribution of the driving force is controlled. At this time, the clutch protection control unit provided in the control means performs control to directly connect the clutch to protect the clutch when the load on the clutch becomes excessive, and then the above-mentioned The engagement force of the clutch is gradually reduced. As a result, excessive heat generation in the clutch is avoided, the load on the clutch is suppressed to a certain limit, and the transition to other clutch control is smooth. Then, when the slip amount of the clutch becomes equal to or larger than a predetermined value while the engaging force is decreasing, the clutch is directly connected again. This ensures the protection of the clutch, prevents hunting of control, etc., and makes the transition to another clutch control smoother. In particular, by setting the predetermined value of the slip amount that defines the condition for directly connecting the clutch again during the decrease of the engagement force as the target slip amount of the clutch, vehicle performance improvement other than clutch protection control can be achieved. The above-mentioned protection of the clutch and stabilization of the clutch control are realized while not impairing the clutch control.

[0010]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS A differential adjustment type driving force distribution control device for a vehicle according to an embodiment of the present invention will be described below with reference to the drawings. FIG. 1 is a block diagram showing a configuration of a main portion thereof, and FIG. FIG. 3 is an overall configuration diagram of the torque transmission system, FIG. 3 is a cross-sectional view showing a main part of the drive torque transmission system, FIG. 4 is a cross-sectional view of a main part of the front and rear wheel torque distribution mechanism, and FIG. 5 is a schematic circuit of the hydraulic pressure supply system. FIG. 6, FIG. 6 is a circuit diagram of a main part of the hydraulic pressure supply system, FIG. 7 is a diagram showing characteristics of the hydraulic pressure setting duty, FIG. 8 is a block diagram showing details of the steering angle data detecting means, and FIG. FIG. 10 is a block diagram showing details of the speed detecting means.
FIG. 11 is a diagram showing the ideal rotational speed difference setting map, FIG. 11 is a diagram showing the rotational speed difference gain setting map, and FIG. 12 is a plan view schematically showing wheel states for explaining the ideal rotational speed difference. 13 is a diagram showing the differential corresponding clutch torque setting map, FIG. 14 is a block diagram showing the longitudinal acceleration corresponding clutch torque setting means, FIG. 15 is the longitudinal acceleration corresponding clutch torque setting map, and FIG. 16 is its engine torque. FIG. 17 is a diagram showing an example of a map, FIG. 17 is a diagram showing an example of a transmission torque ratio map thereof, FIG. 18 is a block diagram showing a modification of the engine torque proportional clutch torque setting means, and FIG. 19 is a center differential input torque setting map thereof. 20 is a characteristic diagram of the clutch torque for protection control, FIG. 21 is a diagram for explaining the first preload learning, and FIG. The figure which shows the pressure characteristic concerning the 2nd preload learning, FIG. 23 is a figure for demonstrating the 3rd preload learning, FIG. 24 is the figure which shows the torque distribution state display means, FIG. 25 is the torque distribution state. FIG. 26 is a characteristic diagram for explaining the torque distribution by the estimating means. FIG. 26 is a flowchart showing the flow of control of the entire vehicle including the device.
Is a flowchart showing the flow of the front and rear wheel torque distribution control, FIG. 28 is a flow chart showing the flow of setting the differential corresponding clutch torque, and FIG. 29 is a flow chart showing the flow of setting the front and rear acceleration corresponding clutch torque.
0 is a flow chart showing the flow of setting the engine torque proportional clutch torque, FIG. 31 is a flow chart showing the flow of setting the protection control clutch torque, and FIG.
Is a flow chart showing the flow of the first preload learning, FIG. 33 is a flow chart showing the flow of the second preload learning, FIG. 34 is a flow chart showing the flow of the third preload learning, and FIG. 35 is the clutch protection thereof. FIG. 36 is a characteristic diagram of the protection control clutch torque of the modification of the control, and FIG. 36 is a flowchart showing the flow of setting the protection control clutch torque of the modification of the clutch protection control.

First, with reference to FIG. 2, an overall structure of a drive system of a vehicle equipped with the differential adjustment type front and rear wheel driving force distribution control device as the differential adjustment type vehicle driving force distribution control device will be described. In FIG. 2, reference numeral 2 is an engine, and the output of the engine 2 is transmitted to the output shaft 8 via the torque converter 4 and the automatic transmission 6. The output of the output shaft 8 is a center differential (hereinafter abbreviated as center differential) 12 as an operating device that distributes engine torque (= driving torque or driving force) between the front wheels and the rear wheels to a required state via an intermediate gear 10. Be transmitted to.

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

The center differential 12 is provided with a sun gear 121, a planetary gear 122 arranged outside the sun gear 121, and a ring gear 123 arranged outside the planetary gear 122, similarly to the conventionally known one. The automatic transmission 6 is attached to the carrier 125 that supports the planetary gears 122.
The output of the output shaft 8 is input, the sun gear 121 is interlocked with the front wheel differential gear device 14 via the front wheel output shaft 27 and the reduction gear mechanism 19, and the ring gear 123 is the rear wheel output shaft 29 and the bevel gear mechanism. It is interlocked with the propeller shaft 20 via 15.

Further, the center differential 12 can change the distribution of engine output torque between the front wheels and the rear wheels by restraining (or limiting) the differential between the front wheel side output section and the rear wheel side output section. A hydraulic multi-plate clutch 28 serving as a differential limiting mechanism or a differential adjusting mechanism is additionally provided. That is, the hydraulic multi-plate clutch 28 includes the sun gear 121 (or the ring gear 1).
23) is interposed between the carrier 125 and the carrier 125, and the frictional force is changed by the control pressure applied to its own hydraulic chamber, so that the differential between the sun gear 121 (or the ring gear 123) and the carrier 125 is constrained. It has become.

The hydraulic multi-plate clutch 28 may be interposed between the sun gear 121 and the ring gear 123. Therefore, the center differential 12 appropriately controls the hydraulic multi-plate clutch 28 from the completely free state to the locked state, so that the torque transmitted to the front wheel side and the rear wheel side is about 32:68 for the front wheels: the rear wheels. It is possible to control from a degree to a ratio (for example, 60:40) according to the ground load of the front and rear wheels.

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

Reference numeral 30 is a steering wheel 3.
The steering wheel angle sensors (steering wheel angle sensors) 34a and 34b for detecting the rotation angle from the neutral position of 2, ie, the steering wheel angle (steering wheel angle) θ, are lateral accelerations G acting on the front and rear portions of the vehicle body, respectively.
This is a lateral acceleration sensor that detects yf and Gyr. In this example, the two detection data Gyf and Gyr are averaged to obtain lateral acceleration data. However, only one lateral acceleration sensor 34 is provided near the center of gravity of the vehicle body. The detected value may be lateral acceleration data.

Reference numeral 36 is a longitudinal acceleration G acting on the vehicle body.
A longitudinal acceleration sensor that detects x, 38 is a throttle position sensor that detects the throttle opening θt of the engine 2, 39 is an engine key switch of the engine 2, 40,
42, 46, 44 are the left front wheel 16, the right front wheel 18,
These are wheel speed sensors that detect the rotational speeds of the left rear wheel 24 and the right rear wheel 26, and the outputs of these switches and each sensor are input to the controller 48.

Reference numeral 50 is an antilock brake device, and this antilock brake device 50 operates in conjunction with the brake switch 50A. That is, when the brake switch 50A is turned on when the brake pedal 51 is stepped on, an antilock brake operation signal is output in conjunction with this, and the antilock brake device 50 is operated. Further, when the operation signal of the antilock brake is output, at the same time, a signal indicating the state is output from the controller 4
8 is input. Further, reference numeral 52 is an indicator lamp which is turned on based on a control signal of the controller 48.

The controller 48 includes a CPU, a ROM, a RAM, an interface, etc., which are not shown, but are required for the control described later. Reference numeral 54 is a hydraulic pressure source, and 56 is a pressure control valve system (hereinafter, referred to as a pressure control valve) which is interposed between the hydraulic pressure source 54 and the hydraulic chamber of the hydraulic multi-plate clutch 28 and controlled by a control signal from a controller 48. Abbreviated).

Further, this vehicle is provided with an automatic transmission, and the reference numeral 160 is a shift lever 1 of the automatic transmission.
It is a shift lever position sensor (shift range detecting means) that detects the selected shift range of 60A, and this detection information is also sent to the controller 48. Further, the engine speed Ne detected by the engine speed sensor (engine speed sensor) 170 and the transmission speed Nt detected by the transmission speed sensor (transmission speed sensor) 180 are also the controller 4
Sent to 8. The details of the hydraulic system relating to the hydraulic multi-plate clutch 28 will be described later.

In this example, a traction control system 151 is also provided. That is, the engine 2 has the main throttle valve 152 whose opening is controlled according to the amount of depression of the accelerator pedal 53, and constitutes the accelerator pedal system engine output adjusting device together with the accelerator pedal 53 and the coupling measure. .. Then, the sub-throttle valve 153 as an engine output control means that is controlled independently of the accelerator pedal system engine output adjusting device.
However, in the intake passage of the engine 2, the main throttle valve 1
It is provided in series with 52. This sub throttle valve 1
53 is driven by a motor, and this motor is drive-controlled based on the detection results of the rear wheel speed sensors 44, 46, the front wheel speed sensors 40, 42, the engine speed sensor 170, the engine load sensor 172, and the like.

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

The input section comprises an input gear 113 meshing with the intermediate shaft 10A side and an input case 124 to which the input gear 113 is serrated and connected.
The bearing 114 in the end cover 115a and the liner 115b fixed to the transmission case 115.
It is mounted rotatably via a and 114b. The input case 124 has a shape in which the diameter is expanded toward the front (to the left in FIGS. 3 and 4), and the expanded portion including the planetary gear element and the rear of the expanded portion (in FIG. 4). , Right side), and an opening is formed in front of the expanded diameter portion. The rear wheel output shaft 29, which will be described later, can be mounted on the output shaft 29 from the rear side (rightward in FIGS. 3 and 4). Also, on the outer periphery of the opening,
A plurality of grooves 124d are formed.

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

Of these, the sun gear 121 is integrally provided on the hollow shaft member 27a.
The front-wheel output shaft 27 and the front-wheel output shaft 27 are both serrated with the hollow shaft 145, and the hollow shaft member 27a and the front-wheel output shaft 27 can rotate integrally via the hollow shaft 145. A piston accommodating portion 145a described later is formed on the hollow shaft 145.

The ring gear 123 is connected to the connecting member 13.
0, and the connecting member 130 is attached to the rear wheel output shaft 2
It is connected to the rear wheel output portion by serration coupling with 9. As a result, the output of the ring gear 123 is input to the propeller shaft 20 via the connecting member 130, the rear wheel output shaft 29, and the bevel gear mechanism 15.

Further, the planet carrier 125 is formed with a convex portion 125i which can be fitted in each groove 124d of the input case 124 on the outer peripheral portion, and by these fittings, it is connected so as to rotate integrally with the input case 124. Has been done.
The sun gear 121 is connected to the front wheel output part, and the ring gear 123 is connected to the rear wheel output part. Also,
A stopper 134 is provided for fixing each pinion shaft 126.

These sun gear 121 and ring gear 12
A plurality of planetary pinions 122 are provided between the planetary pinion 122 and the planetary pinion 122.
Both of them are attached to the planet carrier 125 via the pinion shaft 126. The planet carrier 125 is coupled to the input case 124 so as to rotate integrally therewith, and includes a collar-shaped base plate portion 125 a and a planetary pinion accommodating portion 12 formed rearward thereof.
5b and a tubular clutch disc support 125f formed in the front.

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

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

Of these, the front wheel output side disc plate 2
8a is driven by the first piston 141 and the second piston 142, and can be joined to the input side disc plate 28b. It should be noted that the first piston 141 and the second piston 142 are housed so as to be axially movable in a piston housing portion 145a formed on the outer periphery of the hollow shaft 145. Piston accommodating portion 145 is provided between 141 and 142.
Partition plate 143 that is fixed to a and does not move in the axial direction
Is installed.

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

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

The output gear 19a is provided with bearings 114c,
The bevel gear shaft 15A and the bevel gear 1 are supported by the transmission case 115 side via 114d.
5a is supported on the transfer case 115c side via bearings 114e and 114f. Also, the output gear 1
9a and the input gear 19b constitute the reduction gear mechanism 19, and the bevel gear 15a and the bevel gear 15b constitute the bevel gear mechanism 15.

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

Further, in FIG. 4, 132a, b and c are plates for axially supporting the respective shafts, and 133 is a circlip. On the other hand, the hydraulic system for the hydraulic multi-plate clutch 28 is configured as shown in FIG. 5 (schematic hydraulic circuit diagram) and FIG. 6 (main part hydraulic circuit diagram). That is, as shown in FIG. 5, the reservoir 6a also serves as that of the automatic transmission 6, and the pump 58 for sucking the oil in the reservoir 6a has its outlet through a check valve 60 and a pressure control valve 62. Is connected to the hydraulic chamber of the hydraulic multi-plate clutch 28.
The pressure control valve 62 can take a first position in which the hydraulic chamber of the hydraulic multi-plate clutch 28 communicates with the pump 58 and a second position in which the pump 58 communicates with the reservoir 6a of the automatic transmission 6, The position between them is controlled by the controller 48.

A reducing valve 165 is provided in the passage between the check valve 60 and the pressure control valve 62 to open the oil at a set pressure (for example, about 9 kg / cm ² ) to allow oil to escape to the reservoir of the automatic transmission 6. An accumulator 66 and a pressure switch 68 are connected to this passage. The detection signal of the pressure switch 68 is input to the controller 48. The motor 58a that drives the pump 58 is controlled by a control signal from the controller 48.

The specific structure of the pressure control valve 62 is shown in FIG. In FIG. 6, reference numeral 161 denotes a 4WD control valve, and this 4WD control valve 161 is a spool valve, and two valve body portions 161 provided on the spool body 161a.
b, 161c. Spool body 161a
Receives the duty pressure (solenoid control pressure) Pd and the reducing pressure Pr at its both ends, respectively, when the duty pressure Pd decreases, it goes to the left side in the figure and becomes an open state, and when the duty pressure Pd rises, it goes to the right side in the figure. Go to and enter the closed state. Note that 161d is a spring that biases the spool body 161a in an appropriate direction so that the spool body 161a can be appropriately moved as described above.

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

163 is an orifice, 164 is an oil filter, 165 is a reducing valve, and the orifice 163 is a reducing valve 165 and 4WD.
Between the control valve 161 and the oil filter 1
64 is an oil passage 16 that flows into the reducing valve 165.
9b, respectively. In the reducing valve 165, the valve body 165a has a return spring 165b.
The valve body 165a is automatically supplied with a hydraulic pressure when the hydraulic pressure is equal to or lower than a preset pressure, and is discharged when the hydraulic pressure is equal to or higher than the preset pressure. To move to.

Therefore, for example, the solenoid 162a
When activated and duty valve 162 opens, 4WD
Oil pressure on the left end side of the control valve 161 (duty
Pressure) Pd decreases and the return spring 161d causes
By moving the valve body portions 161b and 161c to the left,
The line between the oil passages 169c and 161g is opened, and the line pressure P ₁
Is the operating oil pressure (4WD clutch pressure) P_FourAs hydraulic multi-plate
It is supplied to each oil chamber 144a, 144b of the latch 28.
Then, the hydraulic multi-plate clutch 28 is connected.
It is configured.

If the solenoid 162a is not operated and the duty valve 162 is closed, the hydraulic pressure (duty pressure) on the left end side of the 4WD control valve 161 is increased.
Pd rises, the valve bodies 161b, 161c move to the right (to the position shown in FIG. 6), and the oil passages 169c and 16
The 9WD is disconnected and the 4WD clutch pressure P ₄ is released so that the hydraulic multi-plate clutch 28 is separated.

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

Next, referring to the block diagram of FIG. 1, the constituent elements of the controller for controlling the differential of the center differential 12 by the hydraulic multi-plate clutch 28 (hereinafter referred to as driving force distribution control or center differential control) will be described. And explain.
In this control, each sensor (wheel speed sensor 40, 42, 4
4, 46, steering angle sensors 30a, 30b, 30c, lateral acceleration sensor 34, longitudinal acceleration sensor 36, throttle position sensor 38, engine speed sensor 170, transmission speed sensor 180, shift position sensor 160, etc.) Based on the above, the clutch torque of the hydraulic multi-plate clutch 28 is set, and the differential hydraulic pressure of the hydraulic multi-plate clutch 28 is controlled so as to obtain the target clutch torque.

Among the data, ABS information, wheel speed,
Steering angle, gear position, comprehensive communication between ABS control unit and engine control unit (SCI communication: SCI = S
data such as erial communication interface) is digitally input, and longitudinal acceleration, lateral acceleration, accelerator opening,
The hydraulic control for the multi-disc clutch, the 4WD control unit control, the current to the electromagnetic clutch of the differential gear device (rear differential) 22 for the rear wheels, and the like are analog input.

Further, the clutch torque of the hydraulic multi-plate clutch 28 is set to an ideal differential state by paying attention to the differential state between the front wheels and the rear wheels (the difference in rotational speed, which is also referred to as the rotational speed difference). Differential corresponding clutch torque Tv for performing control so that the control is performed so as to correspond to the longitudinal acceleration corresponding clutch torque Tb for performing control corresponding to the longitudinal acceleration acting on the vehicle.
And the engine torque proportional clutch torque Ta that is set in proportion to the engine torque so that a large road surface transmission torque can be obtained in a four-wheel drive state in which the front and rear wheels are directly connected when suddenly starting, and the clutch portion of the wet multi-plate clutch is protected. One of the clutch torques for protection control Tpc for selecting the clutch torques Tv, Tb, Ta, and Tpc is described in order.

The differential-compatible clutch torque Tv is a clutch torque that causes the vehicle to behave in accordance with the driver's will during turning. In order to control the attitude of the vehicle body, the rear wheel is used as the drive base from the rear wheel. Since it is effective to set to slip, the differential compatible clutch torque T
v is set to achieve such a state.

Therefore, the portion (the front and rear wheel rotational speed difference proportional calculation means 20) relating to the setting of the differential corresponding clutch torque Tv.
1), as shown in FIG. 1, front and rear wheel actual rotation speed difference detection unit 200, front and rear wheel ideal rotation speed difference setting unit 210, front and rear wheel actual rotation speed difference ΔVcd, and front and rear wheel ideal rotation speed difference ΔVhc.
And a differential-compatible clutch torque setting unit 220 for setting the clutch torque Tv 'and a correction unit 246 for correcting the lateral acceleration of the clutch torque Tv'.

The front / rear wheel actual rotation speed difference detection unit 200 includes filters 202a to 202d, a front wheel wheel rotation speed data calculation unit 204a, and a rear wheel wheel rotation speed data calculation unit 204.
b and a front and rear wheel actual rotation speed difference calculation unit 206. The filters 202a to 202d include rotational speed data signals FL, FR, RL, RR of the left front wheel 16, the right front wheel 18, the left rear wheel 26, and the right rear wheel 28 detected by the wheel speed sensors 40, 42, 44, 46, respectively. From
This is for removing the minute vibration component of the data generated by disturbance or the like.

Further, the front wheel rotational speed data calculation unit 20.
In 4a, the wheel speeds of the front wheels obtained from the rotation speed data signals FL and FR of the front wheels are averaged to obtain the front wheel rotation speed Vf, and in the rear wheel rotation speed data calculation unit 204b, the rotation speed data signal RL of the rear wheels. , RR, the rear wheel rotational speed Vr is obtained by averaging the wheel speeds of the rear wheels.

Further, the front and rear wheel actual rotation speed difference calculation unit 206
Then, by subtracting the front wheel rotation speed Vf from the rear wheel rotation speed Vr, the actual rotation speed difference between the front and rear wheels [the rotation speed difference between the front and rear wheels (the front and rear rotation difference, which is also the rotation difference at the center differential)] ΔVcd To calculate. The front-rear wheel ideal rotation speed difference setting unit 210 includes a driver-requested steering angle calculation unit (pseudo-steering angle calculation unit) 212 as a steering angle data detection unit, and a driver-requested vehicle speed calculation unit (pseudo steering angle calculation unit) as a vehicle speed data detection unit. An estimated vehicle speed calculation unit or a pseudo vehicle speed calculation unit) 216,
Ideal rotation speed difference setting unit 21 as an ideal operating state setting unit
It is configured with 8 and.

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

The function δh = f (θ ₁ , θ ₂ , θn) for obtaining the sensor-corresponding steering angle δh corresponds to the specifications of the steering wheel angle sensor. In addition, the sensor-compatible steering angle δh
Both of the lateral acceleration data Gy and the lateral acceleration data Gy are positive in the right turning direction, for example. In order to compare the directions of the sensor-corresponding steering angle δh and the lateral acceleration data Gy, with respect to the detection data x, the function SIG (x) relating to the direction is as follows.
To set. When x> 0, SIG (x) = 1 When x = 0, SIG (x) = 0 When x <0, SIG (x) = − 1 Therefore, in the comparison unit 212c, the sensor-corresponding steering angle δh The direction and the direction of lateral acceleration data Gy are compared by SIG
This is done by comparing (δh) with SIG (Gy).

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

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

Therefore, in order to eliminate such a problem, when the direction SIG (δh) of the sensor-corresponding steering angle δh and the direction SIG (Gy) of the lateral acceleration data Gy are not equal, the driver-requested steering angle Is set to 0. As shown in FIG. 9, the estimated vehicle speed calculation unit 216 uses the rotational speed data of the left front wheel 16, the right front wheel 18, the left rear wheel 26, and the right rear wheel 28 detected by the wheel speed sensors 40, 42, 44, 46. A wheel speed selection unit 216a that selects the wheel speed data of the second largest value from the bottom (smallest) of the signals FL, FR, RL, and RR, and an estimated vehicle speed is set from the selected wheel speed data and the like. Estimated vehicle speed calculation unit 216
It consists of c and.

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

The wheel speed data having the second largest value from the bottom among the rotation speed data signals FL, FR, RL, RR is adopted as the wheel speed data SVW because each wheel is normally on the over rotation side. In most cases, the wheel speed at which the wheel rotates at the lowest speed is preferably adopted, but the second wheel speed from the bottom is used in consideration of the reliability of the data.

Then, the ideal rotation speed difference setting unit 218 uses the driver-requested steering angle δref calculated by the driver-requested steering angle calculation unit 212 and the estimated vehicle body speed Vref calculated by the estimated vehicle body speed calculation unit 216. The ideal rotation speed difference ΔVhc is set in correspondence with the map as shown in FIG. That is, regarding the vehicle speed, when the vehicle speed is low, the influence of the difference between the orbital radii of the front and rear wheels during turning (so-called inner ring difference) is large, and the rotation speed Vr of the rear wheels is lower than the rotation speed Vf of the front wheels, but at a high vehicle speed. As the rotational speed Vr of the rear wheels increases with respect to the rotational speed Vf of the front wheels, the rear wheels tend to slip at high speeds. This ensures the responsiveness of the posture of the vehicle body that is required at higher speeds. Regarding the steering angle, the larger the steering angle is, the larger the rotation difference required for the front and rear wheels is. Therefore, the larger the magnitude | δref | of the steering angle data δref is, the larger the value of ΔVhc is.

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

Further, the front wheel speed Vf and the rear wheel speed Vr
Is theoretical, Vf: Vr = Rf: Rr,
Vf: V = Rf: R, and the angles βf and βr shown in FIG. 12 (b) have a relationship of βf−βr = AV ² , and ΔVhc is set to V from these relationships and the above equations. It can be defined as a function of δ [ΔVhc = fc (V, δ)]. However, in this case, V is a theoretical value, that is, the estimated vehicle speed Vref
And δ is also a theoretical value, that is, the driver-requested steering angle δ
is equivalent to ref. Such a function [ΔVhc = fc (Vre
f, δ ref)] is converted into a map as shown in FIG.

By the way, regarding the steering angle, the steering wheel angle θ
In addition to the actual steering angle (sensor-corresponding steering angle) δh, there is a turning G equivalent steering angle δy obtained from the lateral acceleration (turning G) Gy during turning. This turning G equivalent steering angle δy can be calculated by the following equation. δy = [(1 + A · Vref ² ) / Vref ² ] · l · Gy (1.2) where A is the stability factor, Vref is the theoretical vehicle speed (estimated vehicle speed) described later, l is a wheel base.

The turning G equivalent steering angle δy obtained in this way
On the other hand, the above-described actual steering angle (sensor-corresponding steering angle) δh is a steering angle that more reflects the driver's intention. In other words, if the driver wants to make a bigger turn than the current one, | δh |> |
δy |, and by using the steering angle value | δh |, the magnitude of the ideal rotational speed difference (slip target value) can be made larger than by using the steering angle value | δy | If you want to hold down, | δh | <| δy |
By adopting the steering angle value | δh |, the magnitude of the ideal rotational speed difference (slip target value) can be made smaller than adopting the steering angle value | δy |.

As described above, the front and rear wheel actual rotational speed difference ΔVcd detected by the front and rear wheel actual rotational speed difference detecting section 200,
The subtractor 222 subtracts (ΔVhc) from the ideal front-rear wheel rotational speed difference ΔVhc set by the front-rear wheel ideal rotational speed difference setting unit 210.
cd−ΔVhc), and the obtained difference ΔVc (= ΔVcd−Δ
Vhc) and the ideal front-rear wheel rotational speed difference ΔVhc are input to the differential corresponding clutch torque setting unit 220 as data.

The differential compatible clutch torque setting unit 220 is
Front / rear wheel actual rotation speed difference ΔVcd and front / rear wheel ideal rotation speed difference ΔV
The clutch torque Tv 'is set in accordance with the difference ΔVc (= ΔVcd−ΔVhc) from hc. The clutch torque Tv ′ is set depending on whether the front / rear wheel ideal rotational speed difference ΔVhc is positive or negative.
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 Tv ′ is set as in the following (1) to (3).

If ΔVcd ≧ ΔVhc, the rear wheels are over-rotating and slipping. Therefore, it is desirable to suppress a part of the engine torque largely distributed to the rear wheels toward the front wheels to prevent the rear wheels from slipping. .. Therefore, the clutch torque Tv '
Is set in proportion to the magnitude of the difference ΔVc (ΔVcd−ΔVhc), Tv ′ = a × (ΔVcd−ΔVhc) = a × ΔVc (1.3) (where a is proportional to constant).

If ΔVhc>ΔVcd> 0, the actual rotational speed of the rear wheels is higher than that of the front wheels even though the front wheels are slipping. Therefore, at this time, the clutch torque Tv '
If is increased, the engine torque distributed to the front wheels is increased, and the slip of the front wheels is promoted. Therefore, it is desirable to reduce the engine torque distributed to the front wheels by freeing the differential limitation. Therefore, in this case, the clutch torque Tv 'is set to 0 to set a so-called dead zone.

If 0 ≧ ΔVcd, the front wheels are slipping, so it is desired to reduce the distribution of the engine torque to the front wheels to reduce the slip of the front wheels. Therefore, Tv '=-a * [Delta] Vcd = -a * ([Delta] Vc + [Delta] Vhc) (1.4) is set so that the clutch torque Tv' increases in proportion to the magnitude of [Delta] Vcd. Proportional constant).

A map of such a relationship between Tv 'and ΔVc is as shown in FIG. 13 (a). This map shows the differential corresponding clutch torque from the difference ΔVc and the ideal front-rear wheel speed difference ΔVhc. It is possible to obtain Tv.
When ΔVhc = 0, the dead zone region of ΔVhc>ΔVcd> 0 disappears. (ii) When ΔVhc <0, 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 follows.

If ΔVcd ≧ 0, the rear wheels are over-rotating and slipping. Therefore, it is desired to suppress a slip of the rear wheels by transferring a part of the engine torque largely distributed to the rear wheels toward the front wheels. .. Therefore, the clutch torque Tv 'is Δ
Tv '= a * [Delta] Vcd = a * ([Delta] Vc + [Delta] Vhc) (1.5) (where a is a proportional constant) so as to increase in proportion to the magnitude of Vcd.

If 0>ΔVcd> ΔVhc, the actual rotational speed of the rear wheels is higher than that of the front wheels, even though the rear wheels are slipping. Therefore, at this time, the clutch torque Tv '
If it is increased, the engine torque distributed to the rear wheels will increase and the slip of the rear wheels will be promoted. Therefore, it is desirable to reduce the engine torque distributed to the rear wheels by freeing the differential limitation. Therefore, in this case, the clutch torque Tv 'is set to 0 to set a so-called dead zone.

If ΔVhc ≧ ΔVcd, the front wheels are slipping, so it is desirable to reduce the distribution of the engine torque to the front wheels to reduce the slip of the front wheels. Therefore, Tv '=-a * ([Delta] Vcd- [Delta] Vhc) =-a * [Delta] Vc (1.6) is set so that the clutch torque Tv' increases in proportion to the magnitude of [Delta] Vc ([Delta] Vcd- [Delta] Vhc). (However, a is a proportional constant).

A map of such a relationship between Tv 'and ΔVc is shown in FIG. 13 (b), and this map shows the differential corresponding clutch torque from the difference ΔVc and the ideal front-rear wheel speed difference ΔVhc. It is possible to obtain Tv.
In this way, the differential compatible clutch torque setting unit 220
Then, referring to the map [FIGS. 13 (a) and 13 (b)], ΔVc
And the differential compatible clutch torque Tv calculated from ΔVhc
′ Is adapted for lateral acceleration correction.

The correcting unit 246 is adapted to perform lateral acceleration correction by multiplying the differential corresponding clutch torque Tv ′ by the lateral G gain k ₁ to obtain the differential corresponding clutch torque Tv. The G gain k ₁ is set as follows. That is, the detection data Gy from the lateral acceleration sensor 34 is sent to the lateral G gain setting section 244 after the fine vibration component of the data generated by disturbance or the like is removed through the filter 242. The lateral G gain setting unit 244 sets the lateral G gain k ₁ from the lateral acceleration data Gy according to the map shown in the block of the setting unit 244 in FIG.

This lateral G gain k ₁ is the friction coefficient μ of the road surface.
Is to reflect the state of
It can be judged that the road surface μ is larger as y is larger, and as the road surface μ is larger, the engine torque is distributed mainly to the rear wheels and the turning performance of the vehicle body can be prioritized. Therefore,
Size of road surface μ (thus, size of lateral acceleration Gy)
When becomes larger, the lateral G gain k ₁ is decreased and the set clutch torque Tv is decreased. Even if the road surface μ is large, if the turning performance of the vehicle body is not given special priority, it is possible to omit the correction by the lateral G gain k ₁ .

The clutch torque Tb corresponding to the wheel slip is
In the clutch torque control corresponding to the above wheel slip,
Depending on when driving on low μ roads (roads with low road friction coefficient μ),
When all four wheels slip and control hunting is likely to occur, the clutch torque is used to prevent the vehicle from performing a smooth turning operation by preventing deterioration of the steering characteristics such as strong understeering of the vehicle. Therefore, the control is performed according to the slip amount Ev of the wheel.

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

In the slip amount calculating unit 254a, the estimated vehicle body speed Vref calculated by the estimated vehicle body speed calculating unit 216d of the estimated vehicle body speed calculating unit 216 and the wheel speed selecting unit 216a of the estimated vehicle body speed calculating unit 216 select 4 The wheel speed SVW, which is the second lowest among the wheel speeds of the wheels, is input, and the slip amount Ev of the wheel is calculated as the difference (= SVW-Vref) between them.
Is calculated.

That is, if the wheels do not slip, the estimated vehicle speed Vref becomes equal to the second smallest wheel speed SVW that is supposed to be substantially equal to the actual vehicle speed, but if the wheels slip, The estimated vehicle body speed Vref is smaller than the wheel speed SVW. And this wheel speed SV
The difference between W and the estimated vehicle speed Vref is the slip amount Ev of the wheel.
I can think about it.

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

That is, when slippage occurs, it is desirable to immediately connect the clutch directly so that the driving force can be surely transmitted to the road surface. However, if the clutch is directly connected suddenly, the behavior of the vehicle body suddenly changes and a shock or discomfort is felt during running. This may occur, and control hunting may occur in the state on the boundary line of whether or not slip has occurred. Therefore, even if a slip occurs, the clutch torque Tb is gradually increased according to the slip amount Ev, the control smoothly shifts to the wheel slip corresponding clutch torque control, and a shock or a shock occurs when the clutch is directly connected. There is no discomfort and hunting is less likely to occur.

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

When the four-wheel slip continues, a dedicated map for four-wheel slip as shown by the broken line in FIG. 15 is prepared so that the clutch is always directly connected even if the slip amount Ev is small. It is conceivable to set the clutch torque Tb based on this map. Although not shown, the wheel slip-compatible clutch torque Tb thus set is applied to the lateral G in the same manner as the correction unit 246.
The lateral acceleration correction may be performed by multiplying the gain k ₁ to obtain the clutch torque Tb corresponding to the wheel slip.

The lateral G gain k 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 in the above description, and therefore the description thereof is omitted here. In addition, here, the wheel speed data SVW has the second lowest wheel speed as a representative wheel speed.
If at least the second lowest wheel speed is higher than the vehicle body speed Vref, it can be determined that the wheels are slipping, and this slip amount is also at least the second lowest wheel speed S.
The second smallest wheel speed is adopted because there is only a portion corresponding to VW.

On the other hand, as the wheel speed data SVW, it is possible to adopt an average value of three wheel speed data excluding the smallest wheel speed data among the rotation speed data signals FL, FR, RL, RR. To determine the slip state of each wheel, it is originally preferable to perform this on the basis of all the wheel speed data of the four wheels, but small wheel speed data is excluded in consideration of the reliability of the data. is there.

The wheel slip corresponding clutch torque Tb thus set is output as data to the maximum value selecting section 280. Engine torque proportional clutch torque Ta
Is a set torque for setting the direct-drive four-wheel drive state in advance so that the initial slip of the rear wheels can be prevented when the transmission torque is expected to increase when the vehicle suddenly starts from the stopped state. Is.

Therefore, this engine torque proportional clutch
The part that sets the torque Ta (engine torque proportional clutch
Chitorque setting means) 260 is shown in the lower left part of FIG.
Engine that detects engine torque Te at a certain moment
Torque detector 264 and torque converter torque ratio t at that time
And a torque converter torque ratio detection unit 266 for detecting the
A transmission that detects the reduction ratio ρm of the transmission.
And the engine torque Te.
Based on the map set in proportion to the engine torque
Engine torque proportional torque Ta 'from torque Te
The gin torque proportional torque setting unit 268 and the engine torque
The torque proportional torque Ta 'to the torque proportional torque ratio t,
Reduction ratio ρm of lance mission, final deceleration ρ₁And rotation difference
Gain k ₂Multiply by the engine torque proportional clutch
Engine torque proportional clutch torque calculation to obtain Luk Ta
Section 270 and the set engine torque proportional clutch
Luk Ta is only derated at low speeds (eg Vref <20km / h).
And a switch 274a for outputting as a data
There is.

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

In the torque converter torque ratio detector 266, the filter 262b is sent from the engine speed sensor 170.
The engine rotation speed data Ne from which the disturbance component has been removed through the transmission rotation speed data Nt which is transmitted from the transmission rotation speed sensor 180 and from which the disturbance component has been removed via the filter 262c. Through the map, the transmission torque ratio t at that time is obtained.

Transmission reduction ratio detector 276
Then, the speed reduction ratio ρm of the transmission is obtained from the selected shift speed information from the shift lever position sensor 160 with reference to the shift speed-reduction ratio correspondence map as shown in block 276 of FIG. A map used for setting the engine torque proportional torque setting unit 268 (see FIG.
In block 268), the engine torque Te and the engine torque proportional torque Ta 'are proportional constants determined from known constants such as the number of teeth Zs and Zr of the sun gear and the ring gear, the front wheel shared load Wf, and the vehicle weight Wa. It is a linear relationship that follows.

In the engine torque proportional clutch torque calculation unit 270, the engine torque proportional torque Ta 'determined as described above, the torque converter torque ratio t, the transmission reduction ratio ρm, the final deceleration ρ ₁ and the rotation difference gain k.
_The rotation difference gain k ₂ is set by the rotation difference gain setting unit 275 as follows. That is, the rotation difference gain k ₂ is intended to avoid the tight corner braking phenomenon, and the ideal rotation speed difference setting unit 218
It is determined according to the map as shown in FIG. 11 from the ideal rotation speed difference ΔVhc set in. The relationship between the rotation difference gain k ₂ in this map and the ideal rotation speed difference ΔVhc can be expressed by the following equation. K ₂ = 0.9 × (| ΔVhcmax || ΔVhc |) / | ΔVhcmax | +0.1 (3.1) where ΔVhcmax = MAX | ΔVhc (δ = MAX) | Also, the coefficient 0.9 and the constant 0.1 are the same as k ₂ of k ₂ . This is to make the lower limit 0.1.

As described above, the rotation difference gain k ₂ linearly decreases as the ideal rotation speed difference ΔVhc increases, and the rotation difference gain k ₂ is multiplied and corrected to correct the ideal rotation during turning. When the speed difference ΔVhc becomes large, turning performance (performance capable of preventing the tight corner braking phenomenon) is given priority over sudden start performance,
The engine torque proportional clutch torque Ta is reduced.

By the way, as shown in FIG. 18, the above-mentioned engine torque proportional torque setting section 268 and engine torque proportional clutch torque calculating section 270 have a center differential input torque calculating section 268 and a clutch torque calculating section 2 respectively.
70 and the turning correction unit 270a may be changed. That is, in the center differential input torque calculation unit 268, the engine torque Te sent from the engine torque detection unit 264 and the torque converter torque ratio detection unit 26 are transmitted.
From the torque converter torque ratio t sent from No. 6 and the transmission reduction ratio ρm sent from the transmission reduction ratio detection unit 276, the center differential input torque (transmission output torque) Td is calculated by the following equation. _{Td = t · ρm · ρ 1} · Te ··· (3.2) However, ρ ₁ is the final reduction ratio.

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

The clutch torque calculation unit 270 calculates the clutch torque Tc at which the front-rear drive distribution becomes equal to the static load distribution from the following equation. Tc = [(Zs + Zr) /Zr.Wf/Wa-Zs/Zr] .Ta (3.3) where Zs is the number of sun gear teeth, Zr is the number of ring gear teeth, and Wf is the front wheel share load. , Wa is the vehicle weight.

[0097] Then, the turning correction unit 270a, by correcting the clutch torque Tc obtained in this manner by the rotation difference gain k ₂ described above, the engine torque proportional clutch torque Ta 'is obtained. It should be noted that the center differential input torque calculation unit 267 and the clutch torque calculation unit 269 may be integrated so as to be obtained from the following formula from the engine torque Te, the torque converter torque ratio t, and the transmission reduction ratio ρm. Tc = [(Zs + Zr) / Zr · Wf / Wa-Zs / Zr ] _{· t · ρm · ρ 1 ·} Te ··· (3.4) In addition, the switch 274a is the signal from the determination unit 274, At low vehicle speed (Vref <20km / h in this example)
Is turned on so that the engine torque proportional clutch torque Ta can be output as data, but when the vehicle speed becomes higher (Vref ≧ 20 km / h), it is turned off and the data of the engine torque proportional clutch torque Ta is output. To stop the output. This is because engine torque proportional control may cause a tight corner braking phenomenon when turning at a certain speed, or eliminate other control speeds when slip tolerance is required. The condition that the engine torque proportional control is performed only when the vehicle speed is low is set.

Next, the setting of the protection control clutch torque Tpc for protecting the clutch portion of the wet multi-plate clutch 28 will be described. The setting of the clutch torque Tpc is performed by the protection control unit 230. ..
That is, in the wet multi-plate clutch 28, generally, when the differential rotation (= ΔN) between the clutch plates becomes large, there is a fear of causing damage such as seizure of the clutch facing and an increase in the amount of wear. The longer the state is, the more likely it is to damage. The slip between the clutch plates is at a level where there is no problem under general driving conditions,
Under a special condition, for example, when the running resistance becomes extremely large, such as when traveling on sand, when mounting a tire with a different diameter, or when mounting a tire chain, it easily occurs, and the load of the clutch 28 becomes excessive under such a special condition. Prone.

Therefore, under such a severe condition of use, the clutch 28 is detected and the clutch 28 is brought into the direct connection state, and the clutch slip control (clutch protection control) is carried out. In the present device, the maximum torque capacity of the clutch 28 is set so that the clutch 28 can be directly connected even under the severe conditions described above.
By applying the maximum hydraulic pressure to 8, the clutch 28 is directly connected.

Further, when the clutch protection control is terminated and the control is shifted to another control, if the clutch 28 is instantly switched from the connected state to the free state, the posture of the vehicle may suddenly change. Therefore, it is necessary to set the protection control clutch torque Tpc by the protection control unit 230 so as to avoid these phenomena. This protection control unit 230
Then, first, the ΔN occurrence state is estimated from the use surface pressure of the clutch 28 and the running condition from the ΔN proportional gain of the differential limiting clutch 28, the absorbed energy q of the clutch 28 is obtained, and the absorbed energy q is used as a base for the clutch. 28 protection control is performed.

That is, the absorbed energy q (here,
q ₂ ) is a threshold value q ₀ [eg 0.6 (kgfm / c
m ² )], the clutch protection control (that is, the control for bringing the clutch 28 into the direct connection state) is executed. The absorbed energy q of the clutch 28 is the clutch 28 because the clutch 28 is a continuous slip clutch.
It is calculated from the unit absorbed energy q'of.

This unit absorbed energy q'is determined by the clutch 2
Of the surface pressure (clutch torque per unit area) P and the differential rotation speed V [that is, q ′ = f (P ·
V)] can be obtained. Therefore, the unit absorbed energy q ′ can be obtained from the clutch torque Tc, the clutch slip speed Δωc (Δωc is obtained from ΔVcd) and the total area A of the clutch by the following equation. q ′ = Tc · Δωc / A Then, the absorbed energy q ₂ is obtained by time-integrating the unit absorbed energy q ′ as in the following equation. q ₂ = ∫q′dt Note that the integration in this case is performed for a predetermined time (for example, 2 seconds), and, for example, when this calculation time is t ₀ (second unit), time (t ₀ −2) to time t _{It is} a definite integral up to ₀ .

In practice, since such a calculation is carried out in a computer such as a microcomputer as a control means, for example, the following unit absorbed energy q ′ is periodically calculated to calculate the wheel speed difference ΔV. And pressure sensor value (clutch 2
The unit absorbed energy q ′ is periodically obtained from the measured value P of the surface pressure applied to P and the learning value of preload (details will be described later) Pi. q ′ = C · ΔV · (P−Pi) Here, assuming that the sampling period of control is T seconds, for example, the absorbed energy q is calculated by the following equation. q ₂ = TΣq ′ where Σq ′ is the sum of the values of q ′ sampled during a predetermined time (for example, 2 seconds).

Absorbed energy q thus obtained
_{When 2} becomes equal to or larger than the threshold value q ₀ (q ₂ ≧ q ₀ ), the protection control is started, but the protection control clutch torque Tpc at this time is started.
Are set as shown in the graph of FIG. That is,
If the control condition is satisfied, the protection control clutch torque Tpc
Is set to a maximum value (for example, 40 kgfm), the clutch 28 is directly connected, and this is held for a certain time (1 second in this example), and then the clutch torque Tpc is kept at a certain inclination (at a certain ratio). It is set to decrease to the initial pressure.

However, at the time of this decrease, the deviation ΔVc from the target value (target wheel speed difference) Vhc of the actual wheel speed difference Vcd.
While monitoring (= Vcd-Vhc), if the magnitude of the deviation ΔVc is the set value ΔV ₀ (for example, ΔV ₀ = 1.0) or more,
That is, if | ΔVc | ≧ ΔV ₀ , the maximum pressure is set again. This is because as the clutch torque Tpc is reduced, the direct connection of the clutch 28 is released and slippage may occur in the clutch 28. At this time, if the road surface, tires, etc. are in a condition where slippage is likely to occur,
The slip of the clutch 28 increases. In this case, the clutch protection may be required again after that, and the clutch torque Tpc is set to the maximum value again before the clutch protection condition (here, the condition that q ₂ ≧ q ₀ is satisfied). By setting the clutch 28 in the directly connected state,
The hunting of the clutch control is prevented, the clutch is reliably protected, and the stable clutch control can be performed.

Here, the slip of the clutch 28 becomes large, and the wheel speed difference Vcd corresponding to this clutch slip becomes larger than the target wheel speed difference Vhc by the set value (ΔV ₀
= 1.0) or more (that is, | Vc | ≧
When ΔV ₀ is reached), the clutch torque Tpc is again set to the maximum value to bring the clutch 28 into the direct engagement state.

This is because if the traveling environment such as the road surface and the tires is not severe and the wheel speed difference Vcd can be brought close to the target wheel speed difference Vhc, the differential corresponding clutch torque Tv corresponding to the target wheel speed difference Vhc. This is because it is considered that the control is performed to improve the original traveling property of the vehicle. Therefore, if the wheel speed difference Vcd cannot be quickly approached to the target wheel speed difference Vhc after the direct connection of the clutch 28 is released, it is determined that the clutch protection is necessary again, and the clutch protection condition (here, q ₂ ≧ q
_The clutch 28 is brought into the direct engagement state before the condition (requiring ₀ ) is established.

If a small wheel speed difference (constant value) is set, and the wheel speed difference Vcd becomes larger than this set value while the clutch torque Tpc is decreasing, the clutch 2 is restarted.
8 may be directly connected. by the way,
Consider the setting of the absorbed energy threshold q ₀ . If the threshold value q ₀ is too small, the clutch protection control is frequently performed, and the clutch control for improving the original running performance of the vehicle cannot be performed. Therefore, it is desirable to set the threshold value q ₀ as large as possible so that the clutch protection control is performed only when the clutch protection is really required.

For this purpose, various severe driving conditions are used.
(For example, when starting on sandy ground, drifting on low μ roads
In some cases, such as during full-speed acceleration on low μ roads)
I found the relationship between the energy q and the durability (life) of the clutch.
Then, based on this, the above-mentioned threshold value q ₀Can be set
Wear. Note that this threshold q₀Of the harsh driving conditions mentioned above
Think about setting only one of them
Be done.

The above-mentioned differential clutch torque Tv, wheel slip-compatible clutch torque Tb, engine torque proportional clutch torque Ta, and protection control clutch torque Tpc are the following clutch torques that are repeated at appropriate timings. The clutch torques Tv, Tb, Ta, Tpc which are set respectively and are set in this way
Is sent to maximum value selection section 280.

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

The clutch torque Tc selected in this way is sent to the torque-pressure conversion unit 282, where:
The clutch control pressure Pc is set so as to obtain the set clutch torque Tc. Here, the clutch control pressure Pc is obtained from the clutch torque Tc by the map (see the block 282 in FIG. 1), but since the clutch torque Tc and the clutch control pressure Pc are generally in a proportional relationship, the map is also shown. It is a linear one.

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

The reason why the centrifugal pressure is corrected is as follows. That is, since the piston chambers such as the oil chambers 144a and 144b rotate, the hydraulic oil inside receives the centrifugal force due to the rotation and drives the pistons 141 and 142 by the centrifugal force (therefore, the clutch control pressure). Will increase. In order to secure the control accuracy of the clutch control pressure, it is necessary to subtractively correct the hydraulic pressure (that is, the centrifugal pressure) generated according to the centrifugal force.

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

This is because the piston chamber rotates in synchronism with the front-wheel side shaft, so that the centrifugal hydraulic pressure is generated in correspondence with the front-wheel vehicle speed Vf. Further, generally, the centrifugal force is proportional to the square of the rotation speed. Therefore, the centrifugal correction pressure Pv is the front wheel vehicle speed Vf.
It is set to be proportional to the square of. The preload correction is a correction in which the initial engagement pressure (initial pressure) Pi set by the initial engagement pressure setting unit (preload setting unit) 288 is added to the clutch control pressure Pc as a preload.

The purpose of this preload correction is to maintain the contact between the clutch plates of the clutch 28 to the extent that no dragging torque is produced (a very slight contact) and to improve the control response. To do. However,
The clearance between the clutch plates of the clutch varies from product to product due to component errors, assembly errors, and the like, and even the same product changes over time. In particular, since the return spring of the clutch plate is generally strong, the error or aging of each part greatly affects the clearance between the clutch plates.

Therefore, it is necessary to always keep the contact between the clutch plates as close as possible while detecting the clearance condition between the clutch plates at an appropriate timing. Therefore, the preload setting unit 288 sets the initial pressure Pi by trying (herein, referred to as learning) how much preload is required at appropriate time intervals.

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 described. First, the first preload learning method will be described. In order to perform preload learning, the engine is in a steady operating state (which can be seen from the fact that the oil temperature of the engine has reached a stable temperature at a predetermined height) and is kept constant. Line pressure is obtained, and further, it is necessary to carry out under the condition that does not affect the control of the other clutch 28.
Therefore, the preload learning condition is set as follows, for example. At least 30 minutes have passed since the ignition key was turned on. The shift selector is selected to one of 1 (1st speed), 2 (2nd speed), D (drive) and N (neutral). In this example, there is no range for P (parking) and R (reverse).
This is because a large hydraulic pressure different from the case of D and N is output. Vref = 0km / h (vehicle speed Vref is 0). Tc ≤ 1 kgfm [Clutch torque Tc is a small predetermined value
(1 kgfm) or less]

When the above conditions are simultaneously satisfied, preload learning is executed as follows. First, as shown in FIG. 21A, a pressure larger than the biasing pressure of the return spring of the multi-plate clutch 28 and smaller than the designed initial engagement pressure of the clutch 28 [for example, P = 0. 4 kgf /
cm ² ], a duty corresponding to ² cm is given for 2 seconds, and thereafter, at an increasing speed of, for example, 1.5% / s, for example, P = 3.
Slowly sweep to a duty equivalent to 0 kgf / cm ² .

Then, the pressure P applied to the hydraulic pistons 141, 142 changes as shown in FIG. 21 (b). That is, since the clutch plates are initially separated, when the duty gradually rises, the hydraulic piston 2 is correspondingly moved.
8, the pressure P gradually rises, but when the hydraulic pistons 141, 142 move to a certain position, the clutch plates come into contact with each other, and the force of the return spring is also added to the pressure P. As a result, the pressure P suddenly increases. Furthermore, the hydraulic pistons 141,1
As 42 moves, the clutch plates make strong contact and the clutch is completely engaged. This state can be seen from the fact that the increase in the pressure P becomes the upper limit.

Here, a value (difference) P ″ obtained by second-order differentiating the detected pressure P with time and the pressure P with time is 1
The differentiated value (difference) P ′ is calculated from time to time in a short cycle, and when the second differentiated value P ″ becomes maximum, it is determined that the clutch plate starts contacting, and the pressure P at this time is calculated. Is determined as the initial pressure, and when the first-order differential value P ′ is maximized, it is determined that the clutch plate is completely engaged.

Specifically, when learning is started and the pressure P rises, the maximum value of the second-order differential value P ″ and the pressure P at this time are stored. This second-order differential value P ″ is stored. The value of is calculated for each short control cycle and updated accordingly. Then, when the first-order differential value P ′ becomes 0 (that is, when the clutch is completely engaged), the calculation of the second-order differential value P ″ is discontinued, and the second-order differential value P ″ is calculated within the period up to this point. The pressure P at which the maximum value is obtained is stored as the initial pressure Pi.

If any of the above conditions (1) to (3) for preload learning is not satisfied during the execution of such preload learning, the preload learning is immediately interrupted and the normal mode is resumed. Further, the above-mentioned preload learning is performed once the ignition key is turned on and is not executed unless the ignition key is turned off and then turned on.

Next, the second preload learning method by the preload setting unit 288 will be described. This pre-pressure learning is also performed under the condition that the engine is in a stable oil temperature state at a predetermined height, a constant line pressure is obtained, and the control of the other clutches 28 is not affected. Although it is necessary to perform this preload learning many times, it is necessary to loosen the above preload learning conditions a little and set the following preload learning conditions. '10 after the ignition key is turned on
More than a minute has passed. The shift selector is selected to one of 1 (1st speed), 2 (2nd speed), D (drive) and N (neutral). Vref = 0km / h (vehicle speed Vref is 0). Tc ≤ 1 kgfm [Clutch torque Tc is a small predetermined value
(1 kgfm) or less] A predetermined time (for example, about 5 minutes or an appropriate time shorter than this) has elapsed since the previous trial.

When the above conditions are simultaneously satisfied, preload learning is executed as follows. First, a duty corresponding to a preset initial pressure Pi (= P ₁ ) is held for a predetermined time (for example, 2 seconds), and then P = 8.8 kgf / for a predetermined time (for example, 1 second). Sweep to a duty equivalent to cm ² (duty of almost 100%).

As a result, the hydraulic pistons 141, 14
The pressure P applied to 2 changes in two types of patterns as shown by curves L1 and L2 in FIG. That is, when the clutch is separated by the initial pressure P ₁ , the clutch contacts and starts dragging at a certain point when the duty is swept as shown by the curve L 1, so that the hydraulic pistons 141, 142 cause a shock. Upon receiving, the pressure P suddenly increases and overshoots, and then oscillates to almost 100
Settles to the full engagement pressure (steady peak pressure) according to the duty of%.

When the pressure P overshoots,
After that, the steady-state maximum pressure Pc (known value, here, 8.8
A peak value (maximum value) Pmax larger than a certain level than that of (kgf / cm ² ) occurs. On the other hand, when the clutch is in contact with the initial pressure P ₁ and is in the drag state, the pressure P increases almost linearly as the duty is swept as shown by the curve L2, and at a certain point, the engagement is smoothly completed. Settle down to pressure (steady maximum pressure) Pc.

From such characteristics, the peak value P of the pressure P is
If max is stored and the difference α (= Pmax-Pc) between this value Pmax and the steady maximum pressure Pc is larger than the predetermined value α ₀ , it can be determined that the clutch is disengaged at the initial pressure P ₁ . Therefore, while appropriately increasing or decreasing the starting pressure P from the initial value P ₁ , the above-described trial is repeated at appropriate time intervals (for example, every 5 minutes) to obtain an appropriate initial pressure Pi.
Can be detected and set.

That is, it is desirable to carry out this preload learning many times as long as the above condition is satisfied, and the initial learning value and the initial pressure Pi set at a certain time point (nth learning stage) can be generalized and expressed. , The initial learning value is PINTG (n) and the initial pressure Pi is PINT
(N) Therefore, the previous initial learning value is PINTG (n-1), and the initial pressure is PINT (n-1).
−1), and in the n-th learning stage, the previous initial pressure is learned by PINT (n−1). Then, the difference α (= Pmax-Pc) obtained by the predetermined duty sweep is compared with the threshold value α _0, and the current initial learning value PINTG (n) and the initial pressure PINT (n) are set as follows. .. When α ≧ α ₀ , PINTG (n) = PINTG (n−1) + β PINT (n) = PINTG (n−1) + β = PINTG (n) When α <α ₀ , PINTG (n) = PINTG ( n−1) −β PINT (n) = PINTG (n−1) That is, when α ≧ α ₀ , the initial learning value PINT
For G (n), the previous initial learning value PINT
The initial pressure PINT (n) is set to a value obtained by adding β (= 1 bit pressure) to G (n-1), and β (= 1bi is added to the previous initial learning value PINTG (n-1).
It is set by adding only the pressure of t), that is, the initial learning value PINTG (n) of this time.

This is because when α ≧ α ₀ , it can be determined that an overshoot has occurred, so the previous initial pressure PINT
At (n-1), it can be determined that the clutch 28 has not approached to the bare contact state. Therefore, it is assumed that the initial learning value PINTG (n) of this time is added to the previous initial learning value PINTG (n-1) by β (= 1-bit pressure), and the initial pressure PINT (n) of this time is changed to the previous initial value. Β (= learning value PINTG (n-1)
It is assumed that only 1 bit of pressure) is applied.
Note that 1 bit is limited by the resolution of the oil pressure sensor signal that detects the oil pressure applied to the piston.
1bit = 0.05kgf / cm ² or 1bit = 0.1kgf / c
Set an appropriate value such as m ² .

On the other hand, when α <α ₀ , the initial learning value PINTG (n) is set to the previous initial learning value PINTG (n-1) minus β (= 1 bit). The pressure PINT (n) is set to the previous initial learning value PINTG (n-1). This is because when α <α ₀ , there is no overshoot, so the previous initial pressure PINT (n-1)
Then, it can be determined that the clutch 28 is in a barely contact state or an excessive contact state. Therefore, the current initial learning value PINTG (n) is changed to the previous initial learning value PI.
It is assumed that NTG (n-1) is reduced by β (= 1 bit), but the initial pressure PINT (n) is set to the previous initial learning value PINTG (n-1).

This is because α <α₀The result of
Clutch 28 is barely in contact or is in excessive contact
There is a fear that chattering can not be judged because it is in a state
Therefore, in order to avoid this, the learning result of this time is immediately
Instead of adopting the initial pressure Pi, the previous learning value is adopted.
-ing Therefore, there is excessive contact
And α <α for at least 2 consecutive cycles ₀State continues
The initial pressure Pi is delayed by one cycle.
However, it will be reduced and approached to the appropriate one.
Becomes

During execution of such preload learning, if any of the above preload learning conditions'- is not satisfied, the preload learning is immediately interrupted and the normal mode is resumed. Further, the above-mentioned preload learning is continued as long as the above-mentioned preload learning conditions'to are satisfied.

Next, a third preload learning method by the preload setting unit 288 will be described. Similarly to the second preload learning, this preload learning is also set to execute the preload learning when the following preload learning conditions are simultaneously satisfied. '10 after the ignition key is turned on
More than a minute has passed. The shift selector is selected to one of 1 (1st speed), 2 (2nd speed), D (drive) and N (neutral). Vref = 0km / h (vehicle speed Vref is 0). Tc ≤ 1 kgfm [Clutch torque Tc is a small predetermined value
(1 kgfm) or less] A predetermined time (for example, about 5 minutes or an appropriate time shorter than this) has elapsed since the previous trial.

When the above conditions ('-) are simultaneously satisfied, preload learning is executed as follows. First, FIG.
The duty is adjusted so that the pressure pattern shown in FIG. That is, first, the duty is held at 0% for a predetermined time (for example, 1 second), then the duty is set to a value corresponding to the initial initial pressure P ₁ and held for a predetermined time (for example, 2 seconds), and then the predetermined value is set. Sweep up to a duty equivalent to P = 8.8 kgf / cm ² (which is almost 100% duty) in a time period (for example, 1 second), and a duty equivalent to P = 8.8 kgf / cm ² for a predetermined time (eg, 2 seconds) Hold. This pattern is continuously repeated while appropriately changing the initial pressure Pi.

As a result, the hydraulic pistons 141, 14
The pressure P applied to 2 is the same as in the case of the second preload learning.
As shown by curves L1 and L2 in FIGS. 23 (b) and 23 (c),
There are two types of pattern changes. Then, from the time t ₀ when the duty sweep is started (or the time t ₁ when the pressure P starts to rise) until the steady maximum pressure Pc (or a constant pressure value close to this) as shown by the straight line L 0 is reached. During this period, the straight line L0 and the curve L1 that describes the changing state of the pressure P
Or, the area S of the part surrounded by L2 (shaded in the figure)
When comparing S1 and S2, the curve L1 with overshoot
In the case of, the area S1 is a curve L without overshoot.
It is obviously larger than the area S2 in the case of 2.

Therefore, also in the third preload learning, similar to the second preload learning, the trial as described above is repeated at an appropriate time interval (for example, every 5 minutes) to detect an appropriate initial pressure Pi. Can be set. That is, this preload learning is performed as many times as long as the above conditions are satisfied, and the initial learning value and the initial pressure Pi set at a certain time point (nth learning stage) are set to the initial learning value as described above. PINTG (n) and initial pressure Pi
Is generalized to PINT (n).

Therefore, the previous initial learning value is P
INTG (n-1), initial pressure is PINT (n-
1), and at the nth learning stage, the previous initial pressure is learned by PINT (n-1). Then, the area S obtained by the predetermined duty sweep is compared with the threshold value S _0, and the initial learning value PINTG (n) and the initial pressure PINT (n) of this time are compared.
Is set as follows. When S ≧ S ₀ , PINTG (n) = PINTG (n−1) + β PINT (n) = PINTG (n−1) + β = PINTG (n) When S <S ₀ , PINTG (n) = PINTG (PINTG (n) n−1) −β PINT (n) = PINTG (n−1) That is, S ≧ S ₀ corresponds to α ≧ α ₀ of the second preload learning, and S <S ₀ . Second preload learning α
It corresponds to the case of <α ₀ .

That is, when S ≧ S ₀ , it can be determined that an overshoot has occurred, so the previous initial pressure PINT
At (n-1), it can be determined that the clutch 28 has not approached to the bare contact state. Therefore, it is assumed that the initial learning value PINTG (n) of this time is added to the previous initial learning value PINTG (n-1) by β (= 1-bit pressure), and the initial pressure PINT (n) of this time is changed to the previous initial value. Β (= learning value PINTG (n-1)
It is assumed that only 1 bit of pressure) is applied.

On the other hand, when S <S ₀ , since there is no overshoot, the previous initial pressure PINT (n-
In 1), it can be determined that the clutch 28 is in a barely contact state or an excessive contact state. Therefore, the initial learning value PINTG (n) this time is added to the previous initial learning value PINTG (n-1) by β (= 1 bit), but the initial pressure PINT (n) is the same as the previous initial learning value PINTG (n). The value PINTG (n-1) is set as it is. As for the reason for doing this, similarly to the case of α <α ₀ described above, it is not possible to determine whether the clutch 28 is in a barely contact state or an excessive contact state, only by the result of S <S ₀ . Since there is a fear of causing chattering, in order to avoid this, the learning value of this time is not immediately adopted as the initial pressure Pi, but the previous learning value is adopted.

Therefore, if it is in an excessive contact state,
If there is no S for consecutive 2 cycles ₀I think that the state of
Although the initial pressure Pi is delayed by one cycle,
Will be reduced and approach the appropriate one.
It It should be noted that, while performing such preload learning,
If any of the learning conditions' ~ are not met
Immediately, the preload learning is interrupted and the normal mode is resumed.

Further, the above-mentioned preload learning is continued as long as the above-mentioned preload learning conditions' ~ are satisfied. In this third preload learning, instead of the area S (S1, S2) of the portion surrounded by the straight line L0 and the curve L1 or L2, the straight line L3 and the curve L showing a constant pressure of about the initial pressure are obtained.
Area S '(S1', S
It is also possible to judge by referring to 2 ').

In this case, the calculation of the area S'starts with the du
When the sweep of the tee is started ₀(Or pressure P is higher
Time t when the ascending starts₁), And the calculation of the area S ′ is completed.
Is the steady-state maximum pressure Pc (or
It is the time when a constant pressure value of a similar level is reached. And
The judgment reference value is S₀′ As S ′ ≧ S₀When '
It can be judged that there was a bashoot, and S '<S₀When '
It can be judged that there was no overshoot.

As described above, the centrifugal pressure correction is performed by subtracting the centrifugal correction pressure Pv from the clutch control pressure Pc which is the effective hydraulic pressure, and the initial pressure (preload) Pi is added to the preload application correction. The control pressure Pcd (= Pc-Pv + Pi) of the oil chamber supply level subjected to the above is taken into the peak hold filter 290. The peak hold filter 290 is a kind of limiter that prevents an excessive sudden change in the hydraulic pressure so that hunting does not occur in the control due to the sudden change in the hydraulic pressure. 4 kg / cm ²
/ S), and a slightly lower limit speed (for example, 15.7 kg / cm ² / s) is set for falling of the hydraulic pressure.

When the control pressure Pcd is sent such that the speed of change in hydraulic pressure exceeds such a limit, the control pressure is kept at a value corresponding to this limit value. Further, the control pressure Pcd ′ that has passed through the filter 290 becomes the switch 292.
a and 294a, and is sent to the duty setting unit 295. The switch 292a is turned off when the ABS control (antilock brake control) is being performed (ON state) by the signal from the determination means 292, and is turned on when the ABS control is not being performed. ..
That is, on condition that ABS control is not performed,
A signal of the control pressure Pcd 'is sent. This is because it is necessary to operate the ABS reliably during the ABS control, and it is not preferable to control the torque distribution state of the front and rear wheels at this time because it interferes with the ABS control.

Further, the switch 294a serves as the judging means 29.
It is a control switch for protecting the duty solenoid valve by a signal from No. 4 and is intended to reduce the duty to 0 when the clutch torque Tc is set at a low speed at a low speed. It can be specified that the low speed condition is, for example, Vref ≦ 5 km / h, and the clutch torque Tc condition is, for example, Tc ≦ 1 kgfm. When these two conditions are met, the switch 294a is turned off and the control pressure Pcd 'signal is not sent.

The duty setting section 295 includes a pressure feedback correction section 296 and a pressure-duty conversion section 298.
It has Pressure feedback correction unit 296
Receives the detection information from the pressure sensor 304 that detects the actual pressure acting on the piston, and controls the control pressure Pcd ′.
Signal for correcting the characteristic of the hydraulic circuit. The signal sent from the pressure sensor 304 to the pressure feedback correction unit 296 is filtered by a filter 306 to remove noise components due to disturbance or the like.

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

Hydraulic circuit 300 functioning as a control execution unit
Then, the duty solenoid 302 operates according to the duty set in this way to control the center differential hydraulic multi-plate clutch 28. on the other hand,
In parallel with such center differential control, the state of torque distribution to the front and rear wheels is displayed in the meter cluster on the instrument panel of 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 so as to graphically display (or meter display) the torque distribution state to the front wheels (or the rear wheels). The torque estimation means 310 displays the torque distribution state according to the estimated distribution torque magnitude. Thus, the reason why the torque distribution state is estimated by the torque estimating means 310 is that it is difficult to measure the torque distribution state.

The torque estimating means 310 includes a front-wheel output torque (or rear-wheel output torque) in the case where there is a rotation speed difference between the front and rear wheels and a rotation speed difference between the front and rear wheels in the multi-plate clutch 28. And the front wheel output torque (or the rear wheel output torque) in each of these cases.
The selection means 310b for selecting the smaller one of the front wheel output torque (or the rear wheel output torque) is provided, and these portions 310a and 310b are adapted to estimate the torque distribution state as follows. ..

That is, when estimating the torque distribution, the following two cases can be considered. One is a case where it is assumed that the tire and the road surface do not slip and are in a state similar to the meshing of gears, and the center differential clutch always slips. The other case is that there is a case where the center differential clutch does not slip because there is always a slip between the tire and the road surface.

Therefore, the torque distribution in each of these cases and when the state switches will be considered. First, as a precondition, when the differential limitation is not performed as in this four-wheel drive system, the rear wheels are mainly set (the torque ratio between the front wheels and the rear wheels is 32:68, for example). The clutch 28 surely transmits torque from the rear wheel side to the front wheel side, and is set as follows for simplification. ρf / rf <ρr · ρt / rr (5.1) where ρf: front differential ratio ρr: rear differential ratio ρt: transfer ratio rf: front wheel tire radius rr: rear wheel tire radius, then the clutch does not slip Is a direct drive four-wheel drive distribution, so the front wheel torque Tf and the rear wheel torque Tr are as follows. Tf = Wf / Wa {{Tm + kWr ・ rf / ρ ・ (rfρrρt / rrρt-1)} ・ ・ ・ (5.2) Tr = Wr / Wa ・ {Tm-kWf ・ rr / ρ ・ (rfρrρt / rrρt -1)} ・ ・ ・ (5.3) However, Wf: Front wheel weighting Wr: Rear wheel weighting Wa: Vehicle weight (= Wf + Wr) Tm: Mission output torque (= center differential input torque) k: Slip ratio coefficient ρ: End Reduction ratio [= (ρf + ρr · ρt) / 2] When the clutch slips, the front wheel torque Tf ′ and the rear wheel torque Tr ′ are as follows. Tf '= (Tm-Tc) * a / (a + b) + Tc ... (5.4) Tr' = (Tm-Tc) * b / (a + b) ... (5.5) However, , Tc: Clutch transmission torque capacity a: Number of teeth of sun gear b: Number of teeth of ring gear If the clutch slips as described above, it means that the clutch allows the front-rear torque difference caused by load distribution, differential ratio difference, and the like. Is. Since the clutch is now considering transmitting torque from the rear wheel side to the front wheel side, the front wheel torques Tf and Tf ′ are Tf and Tf ′.
The smaller value can be considered as the front wheel torque value.

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

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

In the characteristic graph, the characteristic line with the smaller m of the straight line directly connected and the curve for a certain control pressure P is adopted. For example, when P is 2 kgf / cm ^{2, in} the region where the torque Te is smaller than Te ₁ , the straight line with direct connection is below the curve of P = 2.
The front wheel torque distribution ratio m follows this straight line. Further, in the region where the torque Te is larger than Te ₁ , the curve of P = 2 is lower than the direct connection, so that the front wheel torque distribution ratio m follows the curve of P = 2.

On the other hand, when P is 8 kgf / cm ² , in the region shown in this graph, the direct connection straight line is always lower, so the front wheel torque distribution ratio according to the direct connection straight line m. In this way, the front wheel torque distribution ratio m
Is set, a signal corresponding to this set value is sent to the torque distribution display unit 312, and the torque distribution display unit 312 displays the torque distribution state to the front wheels. In this example, the torque distribution to the front wheels is 32% -50%.
Since it is about%, a scale corresponding to this is attached to the torque distribution display portion 312, and a lamp is turned on or a pointer is moved up to the corresponding scale, so that the display can be easily understood.

The display of the torque distribution state may be the torque distribution state to the rear wheels, or the distribution state to the front and rear wheels may be displayed in a graph or the like in an analog manner.
Since the differential adjustment type front and rear wheel drive force distribution control device as the differential adjustment type vehicle driving force 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 shown in FIG.
As shown in FIG. 6, first, each control element is initially set (step a1), information from various sensors is input (step a2), and the steering angle neutral position is learned (step a1).
3) and clutch preload learning (step a4), and then the clutch 2 is operated according to the set duty.
The front / rear wheel driving force distribution control is performed while controlling 8 (step a5), and the rear differential is further controlled (step a6).

In steps a7 to a11, engine output control (traction control) such as slip control, trace control, torque selection, retard control calculation, and SCI (Serias Communication Interface) communication control is performed, and the torque distribution display lamp is turned on. Then (step a
12) In step a13, failure diagnosis (fail diagnosis) is performed. At step a14, a predetermined time (15 mse
c) Judging whether or not the time has passed, and a predetermined time (15 mse
c) After a lapse of time, a runaway check by a watch dog is performed (step a15), the process returns to step a2 described above, and a series of control of steps a2 to a13 is repeated.

That is, the front and rear wheel driving force distribution control, the rear differential control, and the engine output control described above are performed in a predetermined cycle (15
msec). Of these, the front and rear wheel driving force distribution control will be described with reference to the flowchart in FIG. As shown in FIG. 27, first, the wheel speed F
R, FL, RR, RL, rudder angles θ ₁ , θ ₂ , θn, lateral acceleration Gy, longitudinal acceleration Gx, throttle opening θth, engine speed Ne, transmission speed Nt, selected shift stages, and other data are detected. Then, this is taken in (step b1), and from these data, front wheel speed Vf, rear wheel speed Vr, driver required vehicle speed Vref, driver required steering angle δ.
Ref and the like are calculated (step b2).

Then, the ideal rotational speed difference ΔVhc between the front and rear wheels is obtained from the driver-requested vehicle speed Vref and the driver-requested steering angle δref (step b3), and the lateral G gain k ₁ is set from the lateral acceleration Gy according to the map. Then, (step b4), the rotation difference gain k ₂ is set from the ideal rotation speed difference ΔVhc according to the map (step b5). Furthermore, in step b6 to step b9, the actual rotational speed difference ΔV
c, an ideal rotational speed difference ΔVhc, and a lateral corresponding G gain k ₁ to calculate a differential compatible clutch torque Tv (in this example, it is calculated by converting from a map), and based on wheel speeds FR, FL, RR, RL and longitudinal acceleration Gx. The clutch torque Tb corresponding to the wheel slip is calculated (converted from the map) from the slip amount Ev of the wheel, and the throttle opening θth, the engine speed Ne, the transmission speed Nt, the selected shift stage, and the engine torque proportional clutch from the speed difference gain k _2. The torque Ta is calculated (converted from the map), and the protection control clutch torque Tpc is set according to the signal of the actual rotation speed difference ΔVch.

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

Further, the peak hold filter is passed through so that an excessive change in the pressure P can be suppressed (step b13). Then, whether the ABS is operating (step b14), the protection condition of the solenoid valve (Vref
≦ 5 km / h, Tc ≦ 1 kgfm) is judged (step b15), and if any of these is satisfied, at step b19 the center differential control pressure Pcd
Is reset to 0.

The center differential control pressure Pcd thus set is subjected to pressure feedback correction in step b16. That is, the difference ΔP between the value of Pcd and the actual measurement value of the pressure sensor is calculated, and the integral correction gains ki and ΔP are calculated.
The above-described center differential control pressure Pcd is corrected by the integral correction pressure Pi obtained from the product of (i) and the proportional correction pressure Pp obtained from the product of the proportional correction gain kpΔP to obtain the pressure P.
To get

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

Then, in step c3, it is judged whether or not the above-mentioned ideal rotational speed difference ΔVhc between the front and rear wheels is 0 or more,
If ΔVhc is 0 or more, proceed to step c4, and if ΔVhc is less than 0, proceed to step c5. At step c4, the clutch torque Tv 'is set from .DELTA.Vc using the map [see FIG. 13 (a)].

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

Further, if 0 ≧ ΔVcd, Tv ′ = − a × ΔVcd = −a × (ΔVc + ΔVhc) is set so that the clutch torque Tv ′ increases in proportion to the magnitude of ΔVcd (however, a Is a constant of proportionality). In addition, ΔVhc =
When 0, the dead zone region of ΔVhc>ΔVcd> 0 disappears.

When the operation proceeds to step c5, the map [Fig.
(B)] is used to calculate the clutch torque Tv 'from ΔVc.
To set. Specifically, if ΔVcd ≧ 0, Tv ′ = a × ΔVcd = a × (ΔVc + ΔVhc) is set so that the clutch torque Tv ′ increases in proportion to the magnitude of ΔVcd (where a is proportional to constant).

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

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

That is, the direction SI of the sensor-corresponding steering angle δh
Direction SIG (Gy) of G (δh) and lateral acceleration data Gy
Is not equal to the steering angle required by the driver, the steering position of the steering wheel and the actual steering angle of the vehicle (turning angle) are set, for example, when the driver performs steering operation such as counter steering. Even if the status is different, inappropriate data will not be adopted, which contributes to improvement of control performance.

Furthermore, as the vehicle speed Vref requested by the driver,
Since the wheel speed data having the second largest value from the bottom among the rotation speed data signals FL, FR, RL, RR is adopted,
Data reliability is ensured. When the ideal rotation speed difference ΔVhc is set to a low vehicle speed, the difference in the orbital radii of the front and rear wheels at the time of turning (so-called inner ring difference) has a great influence, and the rotation speed Vr of the rear wheels is smaller than the rotation speed Vf of the front wheels. However, as the vehicle speed increases, the rotation speed Vr of the rear wheels becomes higher than the rotation speed Vf of the front wheels. For this reason, the rear wheels are more likely to slip at high speeds, and the required responsiveness of the vehicle body posture is ensured at higher speeds. Also,
Regarding the steering angle, the larger the steering angle, the larger the rotation difference required for the front and rear wheels, which is appropriately allowed, and there is an advantage that the tight corner braking phenomenon can be avoided.

Further, the above-mentioned clutch slip-corresponding clutch torque Tb is calculated as shown in FIG. First, in step d1, the slip amount calculating unit 254a of the wheel slip corresponding clutch torque setting means 254,
The estimated vehicle speed Vref calculated by the estimated vehicle speed calculator 216d of the estimated vehicle speed calculator 216 and the estimated vehicle speed calculator 2
The wheel speed that is the second lowest among the wheel speeds of the four wheels selected by the sixteen wheel speed selectors 216a (or the average of the three wheel speed data excluding the wheel speed data that is the lowest among the four wheel speeds). Value) SVW and wheel slip amount Ev (=
SVW-Vref) is calculated.

Next, at step d2, it is judged whether or not the slip amount Ev is at least a minute reference slip amount Ev _0. If Ev is at least Ev ₀ , it is determined that the vehicle is in a four-wheel slip state, and the routine proceeds to step d3, where It is determined whether the clutch torque Tb corresponding to the wheel slip is maximum (MAX). Here, if Tb is the maximum, it is determined that Tb is being maximized in the four-wheel slip state, the process proceeds to step d5, and the clutch torque setting unit 254 is performed.
b, the clutch torque Tb corresponding to wheel slip is maximum (M
AX).

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

As a result, it becomes possible to smoothly shift from the other control to the clutch torque control corresponding to the wheel slip, and the clutch is directly connected without causing a shock or a feeling of strangeness at the time of shifting the control or causing hunting of the control. However, it becomes possible to suppress the slip of the wheels and efficiently transmit the driving force to the road surface. In addition, the inclination (Tb / Ev) in the map shown in FIG. 15 can be changed,
After setting the inclination (Tb / Ev), the clutch torque T
b such as setting b or the above steps d2 and d
It is also possible to omit 3 and d5 and always set the clutch torque Tb based on the slip amount Ev.

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

Next, the engine torque proportional torque setting unit 2
At 68, the engine torque proportional torque Ta 'is read from the engine torque Te through the map (step e).
2). Further, the torque converter torque ratio detection unit 266 obtains the transmission torque ratio t at that time from the engine speed data Ne and the transmission speed data t through a transmission torque ratio map as shown in FIG. 13 (step e3).

Then, in the engine torque proportional clutch torque calculation unit 270, the engine torque proportional torque Ta ′ thus obtained, the torque converter torque ratio t, and the transmission reduction ratio detection unit 276, the transmission reduction ratio ρm, From the final reduction ratio ρ ₁ and the rotation difference gain k ₂ obtained by the rotation difference gain setting unit 275, the engine torque proportional clutch torque Ta (= t · ρm · ρ ₁ · T
e) is calculated (step e4).

Further, in step e5, it is judged whether or not the vehicle speed is low (Vref <20 km / h in this example). If the vehicle speed is low, the engine torque proportional clutch torque Ta is directly output as data. When the vehicle speed is further increased (Vref ≧ 20 km / h), 0 is set as the engine torque proportional clutch torque Ta (step e6), and this is output as control data.

With such an engine torque proportional clutch torque Ta, a direct connection 4WD state is appropriately set at the time of starting or during rapid acceleration from a low speed, so that a high torque can be transmitted to the road surface, and at the time of starting. This prevents tire slippage during sudden acceleration and improves running performance, and contributes to improved drive system durability. Further, the above-mentioned protection control clutch torque Tpc is calculated as shown in FIG.

First, at step f1, it is judged if the protection control clutch torque Tpc is zero. If Tpc is 0, the clutch protection control is not performed, so the routine proceeds to step f2, where the clutch 28 from 2 seconds before to the present is
Calculate the absorbed energy q ₂ of Then, in the subsequent step f3, it is determined whether or not the absorbed energy q ₂ is equal to or greater than the threshold value q ₀ (0.6).

If q ₂ is not larger than the threshold value q ₀ , the clutch protection control is unnecessary, so Tpc is set to 0 in step f13, and the control cycle this time is ended. on the other hand,
If q ₂ is greater than or equal to the threshold value q ₀ , the process proceeds to step f4 to reset the timer value t to 0 in order to start the clutch protection control, and in the following step f5, the value of Tpc is set to the maximum (MAX).
Set to. Then, in the subsequent step f6, it is determined whether or not the value t of the timer is 1 (second) or more.

If the timer value t is smaller than 1, the process proceeds to step f7, where the timer value t is set equal to the control period (Δt).
Only, and the process proceeds to step f9 and thereafter. Then, in step f9, the clutch torque value Tc according to another control law
Is larger than Tpc, and if Tc is larger than Tpc, Tpc is set to 0 in step f13 and the clutch protection control is automatically ended in the current control cycle. However, if the value of Tpc is MAX, then T
c cannot be greater than Tpc.

When the process goes to step f10, it is determined whether or not the accelerator (accelerator pedal) is being operated.
If it is judged whether the accelerator switch is in the OFF state or not, and if the accelerator pedal is operated, Tpc is set to 0 in step f13 and the clutch protection control is automatically ended in the current control cycle. .. In addition, when the process proceeds to step f11, the brake (brake pedal)
Is operated depending on whether or not the brake switch is in the ON state. If the brake pedal is operated, Tpc is set to 0 in step f13 and the clutch is set in this control cycle. The protection control ends automatically.

Further, when the process proceeds to step f12, the deviation ΔVc between the wheel speed difference ΔVcd and the target wheel speed difference ΔVhc.
(= ΔVcd−ΔVhc) is the set value ΔV ₀ (= 1.0 km /
For h), ΔVc ≧ 1.0 or ΔVc ≦ −1.0
(That is, | ΔVc | ≧ 1.0) is determined, and if this is satisfied, the process proceeds to step f4 in the next control cycle. If this is not the case, the process proceeds to step f1 in the next control cycle.

If the No route is taken in any of the steps f9 to 12 as described above, Tpc is MAX and 0.
Therefore, in the next control cycle, it jumps from step f1 to step f6, and it is judged whether the timer t is 1 (second) or more. If the timer t is 1 (second) or more, Tpc is set to MAX. Then, Δt is added to the timer t to update it.

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

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

Such protection control clutch torque Tpc
Clutch protection control protects the clutch plate and contributes to improving the durability of the device. And Tpc to M
Since Tpc is gradually decreased after setting to AX, the transition to other control can be performed smoothly, so that the stability of control,
As a result, the stability of vehicle behavior can be ensured. Also, if this increases based on the absorbed energy of the clutch, the protection control will start, so it becomes possible to perform the protection control when clutch protection is really required, and other clutch control for improving the original running performance of the vehicle can be performed. There is an advantage that clutch protection can be achieved without hindering too much.

Further, even if the road surface, tires, and the like are likely to slip even when Tpc is decreasing, if the traveling environment is severe, the slip of the clutch 28 becomes large and the clutch protection becomes necessary again. Again, since the clutch torque Tpc is set to the maximum value and the clutch 28 is held in the direct connection state, hunting of the clutch control is prevented, the clutch is protected surely, and stable clutch control can be performed. ..

Particularly, the wheel speed difference Vcd and the target wheel speed difference V
Since the above-mentioned control is performed based on the difference from hc, the differential corresponding clutch torque Tv corresponding to the target wheel speed difference Vhc.
The clutch protection can be carried out with the control by the above described as little as possible. Here, regarding the preload correction described above,
This will be described with reference to FIG. First, in the first preload learning method, as shown in FIG. 32, steps g1 to g4 are performed.
Then, after the ignition key is turned on, 3
Whether the shift selector is set to 1 for 0 minutes or more
Whether it is selected among (1st speed), 2 (2nd speed), D (drive), N (neutral),
Whether the vehicle speed Vref is 0 km / h (stopped state),
The set value of clutch torque Tc is small (1kg
fm) or less is judged respectively.

When all of these conditions are satisfied, the routine proceeds to step g5, and when neither of these conditions is satisfied, learning control is not performed. Step g5
In step g6, it is determined whether preload learning has been performed after the ignition key is turned on. If preload learning has already been performed, learning control is not performed, and if preload learning is not performed, step g6 Go to.

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

On the other hand, the maximum value of the value (difference) P ″ obtained by second-order differentiating the pressure P from the pressure P to the hydraulic pistons 141, 142 which changes as shown in FIG. The pressure P at the time is stored, and the stored pressure P is set to the initial pressure. Specifically, when the pressure P increases as learning is started, the second-order differential value P ″ The maximum value of P and the pressure P at this time are stored, and the value of the second-order differential value P ″ is calculated for each control cycle and updated appropriately, and when the first-order differential value P ′ becomes maximum. (That is, when the clutch is completely engaged), the calculation of the second-order differential value P ″ is stopped, and within the period up to this point, 2
The pressure P when the maximum value of the step differential value P ″ is taken is stored as the initial pressure Pi.

During execution of such preload learning,
If any of the above preload learning conditions ~ is no longer satisfied, immediately stop preload learning and return to normal mode.If this preload learning is performed once with the ignition key turned on, then It is not executed unless the ignition key is once turned off and then turned on.

Further, in the second preload learning method, as shown in FIG.
As shown in, in steps h1 to h5, it is determined whether 10 minutes or more have elapsed since the ignition key was turned on, and whether a predetermined time (for example, about 5 minutes or an appropriate time shorter than this) has been set since the previous trial. ) Whether the shift has passed, the shift selectors are 1 (1st speed), 2 (2nd speed), D
It is determined whether the drive is selected from among (drive) and N (neutral), whether Vref = 0 km / h, and whether Tc ≦ 1 kgfm.

If all of these conditions are satisfied, the process proceeds to step h6, and if any of these conditions is not satisfied, learning control is not performed. Step h6
When the process goes to step S3, the hydraulic pressure is raised and the overshoot value of the hydraulic pressure is detected. That is, the hydraulic pressure is raised by holding a preset duty (duty) corresponding to the initial initial pressure P ₁ for a predetermined time (for example, 2 seconds), and then for a predetermined time (for example, 1 second), P = 8. Sweep up to a duty equivalent to 8 kgf / cm ² (which is almost 100% duty).

Then, the overshoot value α of the pressure P applied to the hydraulic pistons 141 and 142, which changes in accordance with this,
To detect. Further, in the next step h7, it is determined whether or not this α is larger than the threshold value. That is, the peak value (maximum value) Pmax of the pressure P is detected and the maximum value Pmax
If the difference (Pmax-Pc) between the steady state maximum pressure Pc (here, about 8.8 kgf / cm ² ) is taken as the overshoot value α, and this α is larger than the threshold value (α ₀ ), there was overshoot. Therefore, it can be determined that the clutch 28 is separated at the initial initial pressure P ₁ , and if this α is not larger than the threshold value, there is no overshoot, that is, the clutch 28 is barely in contact or at the initial initial pressure P _1. It can be judged that there is excessive contact.

If α is larger than the threshold value, the process proceeds to step h8, where PINTG (n) = PINTG (n-1) + β PINT (n) = PINTG (n-1) + β = PINTG (n) Regarding the initial learning value PINTG (n), β is added to the previous initial learning value PINTG (n-1).
(= Pressure for 1 bit) is set, and the initial pressure PINT (n) is obtained by adding β (pressure for 1 bit) to the previous initial learning value PINTG (n-1), that is, , This initial learning value PINT
Set to G (n).

On the other hand, if α is not larger than the threshold value, the process proceeds to step h9, where PINTG (n) = PINTG (n-1) -β PINT (n) = PINTG (n-1), that is, the initial learning value PINTG. For (n), β is added to the previous initial learning value PINTG (n-1).
The initial pressure PINT (n) is set to the previous initial learning value PINTG (n-1), although the value is reduced by (= 1 bit).

During execution of such preload learning, if any of the above preload learning conditions'- is no longer satisfied, the preload learning is immediately interrupted and the normal mode is resumed. Further, the above-mentioned preload learning is continued as long as the above-mentioned preload learning conditions'to are satisfied.

In the third preload learning method, the 34th preload learning method is used.
As shown in the figure, it is determined whether the conditions are the same as those in the third preload learning. That is, in steps h1 to h5, it is determined whether 10 minutes or more have passed since the ignition key was turned on, and a predetermined time (for example, about 5 minutes or an appropriate time shorter than this) has elapsed since the previous trial. The shift selectors are 1 (1st speed), 2 (2
Speed), D (drive), N (neutral), Vref = 0km / h
Whether or not Tc ≤ 1 kgfm,
Each is judged.

If all of these conditions are satisfied, the process proceeds to step h16, and if any of these conditions is not satisfied, learning control is not performed. Step h
At step 16, the hydraulic pressure is raised and the difference between the predetermined pressure and the hydraulic pressure value is integrated. That is, the hydraulic pressure is raised by holding a preset duty (duty) corresponding to the initial initial pressure P ₁ for a predetermined time (for example, 2 seconds), and then for a predetermined time (for example, 1 second), P = 8. Sweep up to a duty equivalent to 8 kgf / cm ² (which is almost 100% duty).

Then, the difference between the pressure P applied to the hydraulic pistons 141 and 142, which changes accordingly, and the predetermined pressure (pressure close to the maximum pressure) is integrated. That is, FIGS. 23 (b) and 23 (c)
As shown in, the time t when the duty sweep is started
_{From 0} (or time t ₁ when the pressure P starts to rise) to the steady maximum pressure Pc (or a constant pressure value close to this) as shown by the straight line L0, the straight line L0 and the pressure P The area S (S1, S2) of the portion (shaded in the figure) surrounded by the curve L1 or L2 that draws the changed state is calculated.

Further, in the next step h17, it is determined whether or not the calculated area S is larger than the threshold value S ₀ . In other words, the area S1 in the case of the curve L1 with overshoot is obviously larger than the area S2 in the case of the curve L2 without overshoot, so that the area S is compared with the threshold value S ₀ to determine the overshoot. The presence or absence of is determined.

If the area S is larger than the threshold value S ₀ , the process proceeds to step h8, and PINTG (n) = PINTG (n-1) + β PINT (n) = PINTG (n-1) + β = PINTG (n In other words, for the initial learning value PINTG (n), β is added to the previous initial learning value PINTG (n-1).
(= Pressure for 1 bit) is set, and the initial pressure PINT (n) is obtained by adding β (pressure for 1 bit) to the previous initial learning value PINTG (n-1), that is, , This initial learning value PINT
Set to G (n).

On the other hand, if the area S is not larger than the threshold value S ₀ , the process proceeds to step h9, where PINTG (n) = PINTG (n−1) −β PINT (n) = PINTG (n−1). Regarding the learning value PINTG (n), β is added to the previous initial learning value PINTG (n-1).
The initial pressure PINT (n) is set to the previous initial learning value PINTG (n-1), although the value is reduced by (= 1 bit).

Even during the execution of such third preload learning,
If any of the above-mentioned preload learning conditions ′ to 1) is not satisfied, the preload learning is immediately stopped and the normal mode is resumed. Also in this case, as long as the above-mentioned preload learning conditions'to are satisfied, the preload learning is continued. By such first to third preload learning, an appropriate initial pressure Pi can be set, respectively, which greatly contributes to improvement of control response.

Particularly, the first preload learning has an advantage that the initial pressure Pi can be set by one learning, which is extremely simple. Also, in the second and third preload learning, the initial pressure Pi is set by several times of learning, but it has the advantage that the setting accuracy is high and the response improvement effect is large. In particular, in the determination based on the integrated value (area), it can be relatively appropriately determined whether or not the initial pressure Pi is appropriate, and the initial pressure Pi can be set without largely depending on the ability of the pressure sensor.

Further, the control performed through the switch 294a protects the duty solenoid valve and the clutch plate, and contributes to the improvement of the reliability and durability of the device. Further, since the torque distribution display unit 312 is provided in the meter cluster and the torque distribution state to the front wheels (or the rear wheels) is graphically displayed (or meter display), the driver recognizes the torque distribution state of the vehicle. It has the advantages that it can be driven while being used effectively and can be used more effectively, and that driving can be made even more enjoyable, and that the product's marketability is greatly improved.

Further, the result of the torque distribution estimation performed at this time may be used by feeding back to the control of each part. Moreover, the following may be considered as a modification of the protection control unit 230. Protection control unit 230
Then, receiving the front and rear wheel actual rotational speed difference Vcd calculated by the front and rear wheel actual rotational speed difference calculation unit 206, the front and rear wheel actual rotational speed difference Vcd is more than the reference value (8.6 km / h in this example). When the large state continues for a reference time (1 second in this example) or more, the protection control clutch torque Tpc is set in a pattern as shown in FIG.

That is, when the above detection condition is satisfied, the protection control clutch torque Tpc is first set to the upper limit for a short time (1 second in this example), and then gradually decreased to 0 (natural release). ) Let's do it. In this example, T when decreasing
The relation between pc and time tt is as follows. Tpc = 40-14tt (4.1) The time for setting the upper limit value and the speed at which the clutch torque Tpc is gradually reduced to 0 (corresponding to the slope of the straight line in FIG. 35).
Is preferably set to an optimum value according to the characteristics of each vehicle.

When the above detection condition is not satisfied, the value of the protection control clutch torque Tpc is set to zero. Further, the above-mentioned protection control clutch torque Tpc is calculated as shown in FIG. First, in step f1, it is determined whether the flag FLG is 1. The flag FLG is a control flag which is set to 1 when the protection control is executed, and is set to 0 when the entire control is started.

Therefore, at the start of control, step f
The routine proceeds to 2 and it is determined whether the front-rear wheel actual rotation speed difference Vcd is greater than or equal to a reference value (8.6 km / h in this example). If the front-rear wheel actual rotational speed difference Vcd is not equal to or greater than the reference value (8.6 km / h), 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 protection control clutch torque T
The value of pc is set to 0, and the flag FLG is set to 0 in step f14.

On the other hand, if it is determined in step f2 that the front / rear wheel actual rotational speed difference Vcd is equal to or greater than the reference value (8.6 km / h), the process proceeds to step f3, where it is determined whether or not the timer is counting, If the timer count is not started, the process proceeds to step f4 to start the timer count. When the timer count is started in this way, it is determined in step f5 whether the timer value is equal to or longer than the reference time (1 sec). If the timer value does not reach the reference time or longer, the process proceeds to step f12 to perform protection. The value of the control clutch torque Tpc is set to 0, and the flag F is set in step f14.
LG is set to 0.

During the several control cycles, if the front-rear wheel actual rotational speed difference Vcd is equal to or greater than the reference value (8.6 km / h), the timer count is continued during this period, and at step f5, It becomes possible to judge that the value of the timer reaches or exceeds the reference time, and at this time, the process proceeds to step f6. In step f6, it is determined whether or not the value of the timer is equal to or longer than the reference time (2 sec). If the value of the timer does not reach the reference time or longer, the process proceeds to step f10 to set the value of the clutch torque Tpc for protection control to 40. To do.

Then, in step f13, the flag FLG is set to 1, and the flow advances to step f8 to judge whether Tpc is 0 or more. When the process proceeds from step f10 to step f8, since Tpc is naturally 0 or more, the timer count is continued. Then, when the state of Tpc = 40 continues and the value of the timer becomes 2 sec or more, from step f6,
In step f7, the value of the protection control clutch torque Tpc is gradually reduced in the relationship of Tpc = 40−14 × (timer value−2).

In this way, when the protection control clutch torque Tpc is not more than 0 after a number of control cycles, the process proceeds from step f8 to step f11 to end the timer count and clear the timer. Then, the value of the protection control clutch torque Tpc is set to 0 in step f12, and the flag FLG is set to 0 in step f14.

As a result, the state where the front and rear wheel actual rotation speed difference Vcd is equal to or greater than the reference value (8.6 km / h) is the reference time (1 se).
c) If the condition for clutch protection that continues for more than the above condition is satisfied, the characteristic as shown in FIG. 35 is set, that is, first, the upper limit value is set for a short time (1 second in this example), and then 0 is gradually set. The protection control clutch torque Tpc is set so as to decrease (naturally cancel).

Also in this modification, when | ΔVc | ≧ 1.0 while the torque Tpc is decreasing, Tpc is set to MAX again and the clutch protection control is restarted, thereby obtaining the same effect as that of the present embodiment. be able to. Further, various settings can be considered for the degree of decrease of the torque Tpc, that is, the slope of the straight line of Tpc in FIGS.

Although the differential adjustment type front and rear wheel drive force distribution control device has been described in this embodiment, the drive force distribution control device according to the present invention includes the clutch protection control.
It can be widely applied to the driving force distribution control of the left and right wheels.

[0226]

As described above in detail, according to the drive force distribution control device for a differential adjustment type vehicle of the present invention, the differential state between the two wheel systems is adjusted to adjust these wheel systems. In a differential adjustment type driving force distribution control device for controlling driving force distribution between vehicles, a differential adjustment mechanism for adjusting a differential state between two wheel systems through a clutch, and the differential adjustment mechanism are controlled. And a clutch protection control unit for performing control for directly connecting the clutch to protect the clutch when the load on the clutch becomes excessive, and the clutch protection control unit is provided. After the clutch direct connection control is performed, the engagement force of the clutch is gradually reduced, and when the slip amount of the clutch becomes a predetermined value or more during the decrease of the engagement force, the clutch is directly connected again. By being configured to perform a Do control, so the protection of the differential limiting clutch is reliably performed, it becomes possible to greatly improve the durability of the differential limiting clutch. At the same time, the transition from the clutch protection control to the clutch control for improving other vehicle performance can be smoothly performed, and the clutch control can be made stable. As a result, there is an advantage that the practicability of the driving force distribution control device using such a differential limiting clutch can be greatly improved.

Particularly, by setting the predetermined value of the slip amount that defines the condition for re-engaging the clutch again during the decrease of the engaging force as the target slip amount of the clutch, the differential limiting clutch can be operated. It is possible to improve the practicability of the used driving force distribution control device while ensuring the control performance for improving the drivability of the vehicle.

[Brief description of drawings]

FIG. 1 is a block diagram showing a configuration of a main part of a drive force distribution control device for a differential adjustment type vehicle as an embodiment of the present invention.

FIG. 2 is an overall configuration diagram of a drive torque transmission system of a differential adjustment type drive force distribution control device for a vehicle as an embodiment of the present invention.

FIG. 3 is a cross-sectional view showing a main part of a drive torque transmission system of a differential adjustment type drive force distribution control device for a vehicle as an embodiment of the present invention.

FIG. 4 is a cross-sectional view of a main part of a front / rear wheel torque distribution mechanism of a drive force distribution control device for a differential adjustment type vehicle as an embodiment of the present invention.

FIG. 5 is a schematic circuit diagram of a hydraulic pressure supply system of a drive force distribution control device for a differential adjustment type vehicle as an embodiment of the present invention.

FIG. 6 is a circuit diagram of an essential part of a hydraulic pressure supply system of a drive force distribution control device for a differential adjustment type vehicle as an embodiment of the present invention.

FIG. 7 is a diagram showing characteristics of a hydraulic pressure setting duty of the differential adjustment type vehicle driving force distribution control device as one embodiment of the present invention.

FIG. 8 is a block diagram showing details of steering angle data detecting means of the drive force distribution control device for a differential adjustment type vehicle as one embodiment of the present invention.

FIG. 9 is a block diagram showing details of vehicle body speed detecting means of the drive force distribution control device for a differential adjustment type vehicle as one embodiment of the present invention.

FIG. 10 is a diagram showing an ideal rotational speed difference setting map of the drive force distribution control device for a differential adjustment type vehicle as one embodiment of the present invention.

FIG. 11 is a diagram showing a lateral acceleration gain setting map of the drive force distribution control device for a differential adjustment type vehicle as one embodiment of the present invention.

FIG. 12 is a plan view schematically showing a wheel state for explaining an ideal rotational speed difference of the drive force distribution control device for a differential adjustment type vehicle as one embodiment of the present invention.

FIG. 13 is a diagram showing a differential corresponding clutch torque setting map of the drive force distribution control device for a differential adjustment type vehicle as one embodiment of the present invention.

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

FIG. 15 is a wheel slip corresponding clutch torque setting map of the drive force distribution control device for a differential adjustment type vehicle as one embodiment of the present invention.

FIG. 16 is a diagram showing an example of an engine torque map of the drive force distribution control device for a differential adjustment type vehicle as one embodiment of the present invention.

FIG. 17 is a diagram showing an example of a transmission torque ratio map of the drive force distribution control device for a differential adjustment type vehicle as one embodiment of the present invention.

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

FIG. 19 is a center differential input torque setting map of the drive force distribution control device for a differential adjustment type vehicle as one embodiment of the present invention.

FIG. 20 is a characteristic diagram of a clutch torque for protection control of the drive force distribution control device for a differential adjustment type vehicle as one embodiment of the present invention.

FIG. 21 is a diagram related to the first preload learning of the differential adjustment type vehicle driving force distribution control device as one embodiment of the present invention, in which (a) shows its duty characteristic and (b) shows that. The pressure characteristics are shown.

FIG. 22 is a diagram showing a pressure characteristic related to second preload learning of the drive force distribution control device for a differential adjustment type vehicle as one embodiment of the present invention.

FIG. 23 is a diagram relating to a third preload learning of the drive force distribution control device for a differential adjustment type vehicle as one embodiment of the present invention, (a) showing its duty characteristic, (b),
(C) shows the pressure characteristics.

FIG. 24 is a diagram showing a torque distribution state display means of the differential adjustment type vehicle driving force distribution control device as one embodiment of the present invention.

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

FIG. 26 is a flowchart showing a control flow of the entire vehicle including a differential adjustment type driving force distribution control device for a vehicle as one embodiment of the present invention.

FIG. 27 is a flow chart showing a flow of front and rear wheel torque distribution control of the drive force distribution control device for a differential adjustment type vehicle as one embodiment of the present invention.

FIG. 28 is a flowchart showing a flow of setting a differential compatible clutch torque of the differential adjustment type vehicle driving force distribution control device as one embodiment of the present invention.

FIG. 29 is a flowchart showing a flow of setting a clutch torque corresponding to a wheel slip of the drive force distribution control device for a differential adjustment type vehicle as one embodiment of the present invention.

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

FIG. 31 is a flow chart showing a flow of setting a clutch torque for protection control of the drive force distribution control device for a differential adjustment type vehicle as one embodiment of the present invention.

FIG. 32 is a flow chart showing a flow of a first preload learning of the drive force distribution control device for a differential adjustment type vehicle as one embodiment of the present invention.

FIG. 33 is a flowchart showing a second preload learning flow of the differential adjustment type vehicle driving force distribution control device as one embodiment of the present invention.

FIG. 34 is a flowchart showing a third preload learning flow of the differential adjustment type vehicle driving force distribution control device as one embodiment of the present invention.

FIG. 35 is a characteristic diagram of a clutch torque for protection control of a modified example of the clutch protection control of the drive force distribution control device for a differential adjustment type vehicle as one embodiment of the present invention.

FIG. 36 is a flowchart showing a flow of setting a clutch torque for protection control of a modification of the clutch protection control of the drive force distribution control device for a differential adjustment type vehicle as one embodiment of the present invention.

[Explanation of symbols]

2 Engine 4 Torque Converter 6 Automatic Transmission 8 Output Shaft 10 Intermediate Gear (Transfer Idler Gear) 12 Center Differential (Center Differential) 14 Differential Gear Device for Front Wheel (Front Differential) 15 Bevel Gear Mechanism 15A Bevel Gear Shaft 15a Bevel Gear 16, 18 Front Wheel 17L, 17R Front Wheel Side Axle 19 Reduction Gear Mechanism 19a Output Gear 20 Propeller Shaft 21 Bevel Gear Mechanism 22 Rear Wheel Differential Gear Device (Rear Differential) 24, 26 Rear Wheel 25L, 25R Rear Wheel Axle 27 Front Wheel Output Shaft 27a Hollow Shaft member 28 Hydraulic multi-plate clutch (differential limiting mechanism or differential adjusting mechanism) 28a Front wheel output side disc plate 28b Input side disc plate 29 Rear wheel output shaft 30, 30a, 30b, 30c Steering wheel angle sensor 32 steering wheel 34, 34a, 34b lateral acceleration sensor 36 longitudinal acceleration sensor 38 throttle sensor 39 engine key switch 40, 42, 44, 46 wheel speed sensor 48 controller 50 antilock brake device 50A brake switch 51 brake pedal 52 indicator light 54 hydraulic power source 56 pressure control valve system (pressure control valve) 58 pump 60 check valve 62 pressure control valve 64 reducing valve 66 accumulator 68 pressure switch 68a motor 113 input gear 114a to 114f bearing 115 transmission case 115a end cover 115b retainer 116 Support members 117a, 117b Oil passage 121 Sun gear 122 Planetary pinion (planetary gear) 123 Ring gear 124 Input Case 125 Planet carrier 125a Base plate part 125b Planetary pinion accommodating part 125f Clutch disc support part 126 Pinion shaft 130 Connecting member 141 First piston 142 Second piston 143 Partition plate 144a First oil chamber 144b Second oil chamber 145 Hollow shaft 145a Piston accommodating 160 Shift lever position sensor (shift range detecting means) 160A Shift lever of automatic transmission 161 4WD control valve 162 Duty solenoid valve (duty valve) 163 Orifice 164 Oil filter 165 Reducing valve 170 Engine speed sensor 180 Transmission speed Sensor 200 Front and rear wheel actual rotation speed difference detection unit 202a to 202d Filter 204a Front wheel vehicle Rotational speed data calculation unit 204b Rear wheel wheel rotation speed data calculation unit 206 Front / rear wheel actual rotation speed difference calculation unit 210 Front / rear wheel ideal rotation speed difference setting unit 212 Driver requested steering angle calculation unit (pseudo steering) as steering angle data detection means Angle calculation unit) 212a Sensor-corresponding steering angle data setting unit 212b Lateral acceleration data calculation unit 212c Comparison unit 212d Driver required steering angle setting unit (vehicle speed data setting unit) 216 Estimated vehicle speed calculation unit as vehicle speed data detecting means (pseudo) Vehicle speed calculation unit) 216a Wheel speed selection unit 216c Estimated vehicle speed calculation unit 216d Filter 218 Ideal rotation speed difference setting unit 220 as an ideal operating state setting unit 220 Differential compatible clutch torque setting unit 222 Subtractor 230 Protection control unit 242 Filter 244 Horizontal G gain setting unit 246 Correction unit 254 Wheel slip support Torque setting means 254a slip amount calculating section 254b clutch torque setting section 264 engine torque detecting section 266 torque converter torque ratio detecting section 267 center differential input torque calculating section 268 engine torque proportional torque setting section 269 clutch torque calculating section 270 engine torque proportional clutch torque calculating section 272 Turning correction unit 274a Switch 274 Judgment means 275 Rotation difference gain setting unit 276 Transmission reduction ratio detection unit 280 Maximum value selection unit 282 Torque-pressure conversion unit 285 Centrifugal oil pressure correction unit 286 Centrifugal correction pressure setting unit 288 Initial engagement pressure setting Part (preload setting part) 290 peak hold filter 292a, 294a switch 292, 294 judgment means 295 duty setting part 296 pressure feedback correction part 298 pressure-due Tee conversion section 300 Hydraulic circuit 302 Duty solenoid 304 Pressure sensor 306 Filter 310 Torque estimation means 310a Calculation means 310b Selection means 312 Torque distribution display section AM Mechanical part of differential force type vehicle drive force distribution control device

Claims (4)

[Claims] 1. A differential adjustment type vehicle driving force distribution control device for controlling driving force distribution between these wheel systems by adjusting a differential state between the two wheel systems. A differential adjusting mechanism for adjusting the differential state between the systems through a clutch and a control means for controlling the differential adjusting mechanism are provided, and the control means protects the clutch when the load on the clutch becomes excessive. Therefore, a clutch protection control section for performing control for directly connecting the clutch is provided, and the clutch protection control section gradually reduces the engagement force of the clutch after performing the direct connection control of the clutch and the engagement force. When the slip amount of the clutch becomes equal to or more than a predetermined value while the clutch is decreasing, the clutch is controlled so that the clutch is directly connected again. Power distribution control device. 2. The differential adjustment type drive force distribution control device for a vehicle according to claim 1, wherein the predetermined value is a target slip amount of the clutch. 3. The clutch protection control unit is configured to determine that the load on the clutch has become excessive when a state in which the rotation difference occurring in the clutch is larger than a reference value continues for a reference time or longer. Claim 1 characterized by
A drive force distribution control device for a differential adjustment type vehicle as described. 4. The clutch protection control unit is configured to calculate the amount of heat generated in the clutch and determine that the load on the clutch becomes excessive when the amount of heat generated exceeds a predetermined value. The drive force distribution control device for a differential adjustment type vehicle according to claim 1, wherein: