JP2737406B2 - Differential limit control device for four-wheel drive vehicle - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a differential limiting control device for a four-wheel drive vehicle for controlling the differential limitation of left and right wheels of a four-wheel drive vehicle.

[0002]

2. Description of the Related Art Between the right and left driving wheels of an automobile,
A differential mechanism is provided to allow the differential that occurs when turning, etc., but with this differential mechanism, the load on one of the left and right wheels is in the groove and the friction coefficient with the road surface is extremely small. comes to the other wheel only this one pair of wheels are rotated is longer little rotation, there is the state that can not transmit the driving torque to the road surface occurs.

Therefore, in such a case, a differential limiting mechanism (LSD = Limited Slip Differential) capable of limiting the differential has been developed. Such a left and right wheel differential limiting mechanism includes a type that is proportional to the rotational speed difference between the left and right wheels and a type that is proportional to the input torque. The right and left wheel rotation speed difference proportional type uses VC that utilizes the viscosity of the liquid.
(Viscous coupling) type LSD, etc.
There is an advantage that the running stability of the vehicle can be improved. On the other hand, the input torque proportional type includes a friction type such as a general LOM (lock automatic) type LSD, and a TORSE using a friction resistance of a worm gear.
There is a mechanical type such as an N (Truesen) type LSD, which has an advantage that the turning performance of the vehicle can be improved.

[0004]

There are cases in which a right-left differential limiting mechanism is provided on the rear wheel or the front wheel of a four-wheel drive vehicle. However, in a four-wheel drive vehicle capable of adjusting the torque distribution to the front and rear wheels. In this case, one of the elements that controls the differential limitation of the left and right wheels
First, it is considered that such a torque distribution state is also effective.

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

[0006]

Therefore, the four-wheel drive vehicle differential limiting control device according to the first aspect of the present invention does not apply the driving force from the engine to the front wheel side drive shaft and the rear wheel side drive shaft. Front and rear wheel driving force distribution means for equally distributing, front and rear wheel differential limiting means capable of selectively limiting the differential between the front wheel side driving shaft and rear wheel side driving shaft, and the front and rear wheel differential limiting means A front and rear wheel differential limiting force setting means for setting a differential limiting force to be applied; and a front and rear wheel drive provided on a side of the front wheel and the rear wheel to which the driving force is more distributed by the front and rear wheel driving force distribution means. Left and right wheel drive force distributing means for distributing the driving force input from the force distributing means to the left wheel drive shaft and the right wheel drive shaft, and selectively selecting the differential between the left wheel drive shaft and the right wheel drive shaft. Left and right wheel differential limiting means that can be limited to Left and right wheel differential limiting force setting means for setting a dynamic limiting force, wherein the left and right wheel differential limiting force setting means determines the right and left wheel differential based on the driving force input to the left and right wheel driving force distribution means. The differential limiting force to be applied by the limiting means is set, and the driving force value calculated from the differential limiting force set by the front and rear wheel differential limiting force setting means as the driving force; It is characterized in that the setting is performed by using the larger one of the driving force when the shaft and the rear wheel side drive shaft are directly connected.

[0007] A differential limiting control apparatus for a four-wheel drive vehicle according to the present invention according to claim 2 is the apparatus according to claim 1, wherein
A target rotational speed difference calculating means for calculating a target rotational speed difference between the left wheel drive shaft and the right wheel drive shaft to be generated by turning of the vehicle; and The present invention is characterized in that the differential limiting force is reduced with an increase in the target rotation speed difference calculated by the rotation speed difference calculation means.

[0008]

In the above-described differential limiting control device for a four-wheel drive vehicle according to the first aspect of the present invention, the front-rear wheel drive force distribution means
The drive force from the engine is unequally distributed to the front wheel drive shaft and the rear wheel drive shaft. At this time, the front and rear wheel differential is set according to the differential limiting force set by the front and rear wheel differential limiting force setting means. The motion limiting means selectively limits the differential between the front wheel drive shaft and the rear wheel drive shaft. On the side of the front wheels and the rear wheels where the driving force is distributed more by the front and rear wheel driving force distribution means, the driving force input from the front and rear wheel driving force distribution means is changed by the left and right wheel driving force distribution means to the left wheel side. It is distributed to the drive shaft and the right wheel drive shaft. At this time, based on the differential limiting force set by the left and right wheel differential limiting force setting unit, the left and right wheel differential limiting unit selectively selects the differential between the left wheel drive shaft and the right wheel drive shaft. Restrict to In the left and right wheel differential limiting force setting means,
The differential limiting force of the left and right wheels is set based on the driving force input to the left and right wheel driving force distribution means. In particular, as this driving force, the driving force value calculated from the differential limiting force set by the front and rear wheel differential limiting force setting means and the front wheel side driving shaft and the rear wheel side driving shaft are directly connected. The above setting is made using the larger value of the driving force in the case.

In the differential limiting control device for a four-wheel drive vehicle according to the first aspect of the present invention, the target rotational speed difference calculating means includes:
The target rotation speed difference between the left wheel drive shaft and the right wheel drive shaft to be generated by turning of the vehicle is calculated, and the left and right wheel differential limiting force setting means is calculated by the target rotation speed difference calculation means. The differential limiting force is reduced as the target rotational speed difference increases.

[0010]

BRIEF DESCRIPTION OF THE DRAWINGS FIG. 1 is a block diagram of a differential limiting control apparatus for a four-wheel drive vehicle according to an embodiment of the present invention.
FIG. 1 is a block diagram showing the overall configuration, FIG. 2 is an overall configuration diagram of a drive torque transmission system including the differential control device, FIG. 3 is a cross-sectional view showing a rear differential as the left and right wheel differential, and FIG. (A) and (b) are both A-
FIG. 5 is a configuration diagram illustrating a rotation speed difference corresponding control unit, FIG. 6 is a configuration diagram illustrating a rear wheel torque gain correction unit,
FIG. 7 is a configuration diagram showing a part of the input torque corresponding control unit,
FIG. 8 is a configuration diagram showing a part of the input torque corresponding control unit,
9 is a configuration diagram of a portion from the maximum value selection unit to the control current output unit, FIG. 10 is a configuration diagram showing details of the steering angle detection unit, FIG. 11 is a configuration diagram showing details of the vehicle speed detection unit,
FIG. 12 is a plan view schematically showing a wheel state for explaining the ideal rotational speed difference, FIG. 13 is a diagram showing an ideal rotational speed difference setting map, FIG. 14 is a diagram showing the rotational difference gain setting map, 15 (a) and 15 (b) are diagrams each showing a map for setting the differential corresponding clutch torque, FIG. 16 is a diagram showing an example of the engine torque map, and FIG. 17 is a diagram showing an example of the transmission torque ratio map. , FIG. 18 is a center differential input torque setting map, FIG. 19 is a flowchart showing a control flow of the whole vehicle including the device, FIG. 20 is a flowchart showing a rear differential control flow, and FIG. FIG. 22 is a flowchart showing a flow of setting the corresponding clutch torque, and FIG. 22 is a flowchart showing a flow of setting the corresponding clutch torque corresponding to the input torque.

First, referring to FIG. 2, the difference for the four-wheel drive vehicle will be described.
An overall configuration of a drive system of a vehicle including a motion restriction control device will be described.

In FIG. 2, reference numeral 2 denotes an engine. The output of the engine 2 is transmitted to an output shaft 8 via a torque converter 4 and an automatic transmission 6. The output of the output shaft 8 is transmitted to a planetary gear type differential device 12 serving as front and rear wheel driving force distribution means for distributing the engine torque of the front wheels and the rear wheels to a required state via the intermediate gear 10. Is transmitted.

On the other hand, the output of the planetary gear type differential device 12 is the reduction gear mechanism 19 and the differential gear device 1 for the front wheels.
4 through the axles 17L, 17R from the left and right front wheels 16, 1
8 while the bevel gear mechanism 15, the propeller shaft 20 and the bevel gear mechanism 21, and the left
The power is transmitted from the axles 25L, 25R to the left and right rear wheels 24, 26 via a differential gear device (rear differential) 22 as right driving force distribution means .

The planetary gear type differential device 12 includes a sun gear 121, a planetary gear 122 disposed outside the sun gear 121, and a ring gear disposed outside the planetary gear 122, as in the prior art. 123 and
The output of the output shaft 8 of the automatic transmission 6 is input to the carrier 125 that supports the planetary gear 122, and 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 interlocked with the propeller shaft 20 via the rear wheel output shaft 29 and the bevel gear mechanism 15. .

Further, the planetary gear type differential device 14 restricts (or limits) the differential between the front wheel side output portion and the rear wheel side output portion, thereby reducing the output torque of the engine between the front wheel and the rear wheel. A hydraulic multi-plate clutch 28 is provided as front and rear wheel differential limiting means or differential adjusting means capable of changing the distribution.

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

Therefore, the planetary gear type differential 12 is
By appropriately controlling the hydraulic multi-plate clutch 28 from a completely free state to a locked state, the torque transmitted to the front and rear wheels can be reduced from about 32:68 to 50:50 for the front and rear wheels. It can be controlled between. Front wheel in completely free condition: rear wheel value: approx. 3
2:68 can be defined by setting the tooth ratio of the input gears on the front wheel side and the rear wheel side of the planetary gears. In this case, when the pressure in the hydraulic chamber of the hydraulic multi-plate clutch 28 is zero and completely free, It is set to be about 32:68.

The ratio in the completely free state (about 3)
2:68) changes depending on the load balance between the front wheel system and the rear wheel system and the like, but usually takes such a value. Further, when the pressure in the hydraulic chamber is set to the set pressure (9 kg / cm ² ) and the hydraulic multi-plate clutch 28 is in the locked state and the differential limit becomes substantially zero, the torque distribution between the front wheels and the rear wheels is reduced. , 50: 5
It becomes 0 and it is in a directly connected state. The control of the hydraulic multi-plate clutch 28 is performed by a control means (controller) 4 described later.
8, the controller 48 includes:
A duty is set for hydraulic control of the hydraulic multiple disc clutch 28 (this function is referred to as front and rear wheel differential limiting force setting means), and the hydraulic multiple disc clutch 28 is controlled based on the set duty.

The details of the rear differential 22 will be described later.

Reference numeral 30 denotes a steering wheel angle sensor for detecting the rotation angle from the neutral position of the steering wheel 32, that is, the steering wheel angle θ, and reference numerals 34a and 34b denote lateral accelerations Gyf and Gyr acting on the front and rear portions of the vehicle body, respectively. In this example, the lateral acceleration data is obtained by averaging two pieces of detection data Gyf and Gyr. In this example, only one lateral acceleration sensor 34 is provided near the center of gravity of the vehicle body, and this detection value is obtained. May be used as the lateral acceleration data. 3
6 is a longitudinal acceleration sensor for detecting a longitudinal acceleration Gx acting on the vehicle body, and 38 is a throttle opening θ of the engine 2.
t is a throttle position sensor for detecting t, 39 is an engine key switch of the engine 2, 40, 42, 44, 4
Left front wheel 16 is 6, respectively, the right front wheel 18, left rear wheel 26, a wheel speed sensor for detecting a rotational speed of the right rear wheel 28, this
Among them, 46 and 44 are rear wheel speed detecting means (front wheel speed).
Sensor), in particular, 46 is a left-wheel-side rotational speed / speed detecting means.
The reference numeral 44 denotes a right wheel side rotational speed detection means. The outputs of these switches and each sensor are input to a controller 48 as control means .

Reference numeral 50 denotes an anti-lock brake device, which operates in conjunction with a brake switch 50A. That is, when the brake switch 50A is turned on when the brake pedal 51 is depressed, an antilock brake operation signal is output in conjunction therewith, and the antilock brake device 50 is operated. Further, when the operation signal of the anti-lock brake is output, a signal indicating the state is simultaneously output to the controller 4.
8 is input. Reference numeral 52 denotes a warning lamp that lights based on a control signal from the controller 48.

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

Reference numeral 54 denotes a hydraulic pressure source, and 56 denotes a pressure control valve system (hereinafter referred to as a pressure control valve system) 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 as control valve). Also, this vehicle is provided with an automatic transmission.
Is a shift lever position sensor (shift range detecting means) for detecting a selected shift range of the shift lever 160A of the automatic transmission, and this detection information is also sent to the controller 48.

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

In this example, a traction control system 151 is also provided. That is, the engine 2 includes the main throttle valve 152 whose opening is controlled in accordance with the amount of depression of the accelerator pedal 162, and constitutes an accelerator pedal-based engine output adjusting device together with the accelerator pedal 162 and the connection measures. .

An auxiliary throttle valve 153 as engine output control means controlled independently of the accelerator pedal system engine output adjusting device is provided in series with the main throttle valve 152 in the intake passage of the engine 2. . The auxiliary throttle valve 153 is driven by a motor, and the motor is driven and controlled based on detection results of the rear wheel speed sensors 44 and 46, the front wheel speed sensors 40 and 42, the engine speed sensor 170, the engine load sensor 172, and the like.

The above-described rear differential (rear differential) 22 is provided with a differential limiting device 23, and is configured as shown in FIG.

That is, as shown in FIG.
Are connected to the rear end of the propeller shaft 20, and the drive pinion gear 402 is supported by the input shaft 401 so as to rotate integrally therewith. The input shaft 401 is rotatably supported in the front part of the case 413 via a bearing 412.

A crown gear 403 meshes with the drive pinion gear 402.
The power transmission annular member 404 is
And the first housing 405 are integrally connected.

The rear differential 22 is a planetary gear type differential using a planetary gear mechanism, and includes a power transmission annular member 404 and an annular member 404 formed therein.
A ring gear 407 formed on the inner peripheral surface of the left wheel 24
A sun gear 408 spline-coupled to the axle 25L of
A carrier 409 which is spline-coupled to the axle 25R of the right wheel 26, and shafts 410a, 410b
Planetary gears 411a, 41 attached via
1b.

Therefore, the rotational torque input from the input shaft 401 passes through the drive pinion gear 402 and the crown gear 403, and from the ring gear 407 of the annular member 404 to the planetary gears 411 a and 411 b and the carrier 4.
09 to the axle 25R of the right wheel 26 and the planetary gears 411a, 411b and the sun gear 4
08 to the axle 25L of the left wheel 24.

Further, a second housing 406 is provided on the right side of the carrier 409, and the second housing 406 is connected to the annular support member 4 via a bearing 428.
18 supported.

The rear differential 22 is provided with a differential limiting device 23. The differential limiting device 23
A multi-plate clutch 414 as a differential limiting mechanism; a driving device 417 for driving the multi-plate clutch;
Differential control unit 48 of the controller 48 for controlling the
a.

That is, the multi-plate clutch 414 is provided inside the annular member 404, and the holder 415a for supporting one clutch disk 414a has the shaft 410a,
The clutch disc 414a is coupled to the carrier 409 via the base 410b so that the clutch disc 414a rotates integrally with the carrier 409, and the holder 415b supporting the other clutch disc 414b is formed on the hollow shaft 416 provided with the sun gear 408. The clutch disk 4
14b is configured to rotate integrally with the sun gear 408.

Further, the driving device 417 is connected to the carrier 40.
9 and a second housing 406, and a force direction conversion mechanism 429 and an electromagnetic clutch mechanism 430 that drives the force direction conversion mechanism 429. In addition,
Since the rear differential 22 limits the differential by an electromagnetic clutch mechanism, an electromagnetically controlled differential (EMC)
D: Electro Magnetic Controlled Differentia
l).

As shown in FIG. 3, the force direction changing mechanism 429 includes a ball 421 interposed between the carrier 409 and the second housing 406, and as shown in FIG.
A diamond-shaped (or rectangular) chamber 42 for accommodating the ball 421
5, the chamber 425 includes a groove 422 formed on the carrier 409 side, the carrier 409, and the second housing 40.
6 is formed by a groove 424 formed in the annular member 423 between them. The annular member 423 is interposed between the carrier 409 and the second housing 406, is normally free in the rotational direction with respect to these members, and rotates integrally with the carrier 409 via the ball 421. However, the second housing 406 side (that is,
When receiving the rotation torque of the crown gear 403 and the power transmission annular member 404), a differential rotation is generated with respect to the carrier 409, and the force due to the rotation torque is changed in direction and acts as a pressing force of the clutch 414. It is supposed to.

The second housing 406 is attached to the annular member 423.
An electromagnetic clutch mechanism 430 applies the rotational torque on the side, and the electromagnetic clutch mechanism 430 is connected to the annular member 423 and the second housing 406 side (the crown gear 403 and the power transmission annular member 404 side). The clutch 427 is interposed between the member 426 and an electromagnetic clutch drive system including a magnet 419 and a solenoid (EMCD coil) 420 as a differential limiting mechanism control means.

That is, the clutch 427 is provided inside the second housing 406, and the magnet 419 is provided further inside the second housing 406, while the magnet 419 is provided outside the second housing 406. , A solenoid 420 capable of attracting the magnet 419 is provided. Thus, when the solenoid 420 operates, the magnet 419
Of the second housing 406 so that the clutch 416 is engaged with the second housing 406.

When the clutch 416 is engaged, the annular member 423 receives the rotational torque of the second housing 406 and tries to rotate integrally with the second housing 406. . At this time, the second housing 406 side (therefore, the sun gear 407 side)
If there is a rotational speed difference (differential rotation) between the and the carrier 409, that is, if there is a rotational speed difference between the left and right wheels,
The annular member 423 makes a differential rotation with respect to the carrier 409, and when the annular member 423 makes a differential rotation with respect to the carrier 409, the ball 421 and the grooves 422, 424 as shown in FIG. The force Δr in the differential rotation direction is converted into a force in a direction orthogonal to the direction, that is, a force F parallel to the axial direction of the rear differential and the axle direction through the inclined surface of the carrier 409. Are driven in the axial direction, and the shafts 410a and 410b and the holder 415a
, The multi-plate clutch 414 is pressed and engaged.

The engagement force of the multi-plate clutch 414 corresponds to the rotational speed difference between the left and right wheels and the magnitude of the electromagnetic force generated by the solenoid 420.
By adjusting the current to zero, the engagement force of the multi-plate clutch 414, and thus the differential limiting force, can be controlled.

The controller 48 is provided with a rear differential control unit 48a for controlling the differential limiting force by adjusting the current to the solenoid 420.

Here, the rear differential control section 48a will be described.

As shown in the block diagram of FIG. 1, the rear differential control section 48a includes sensors (wheel speed sensors 40, 42,
44, 46, steering angle sensors 30a, 30b, 30c, lateral acceleration sensor 34, longitudinal acceleration sensor 36, throttle position sensor 38, engine speed sensor 170,
The clutch torque of the multi-plate clutch 414 is set based on the detection information from the transmission speed sensor 180, the shift position sensor 160, and the like, and the clutch torque is set to the electromagnetic clutch mechanism 430 of the driving device 417 so as to obtain the target clutch torque. Is controlled.

The ABS information, wheel speed,
Steering angle, gear position, comprehensive communication between ABS control unit and engine control unit (SCI communication: SCI =
Data such as serial communication interface is digitally input, and analog input is applied to longitudinal acceleration, lateral acceleration, accelerator opening, hydraulic control to a multi-plate clutch, 4WD control unit control, current to an electromagnetic clutch of a rear differential, and the like. .

In this apparatus, the setting of the clutch torque of the multi-plate clutch 414 is set to an ideal differential state by focusing on the differential state of the left and right wheels (the rotational speed difference is also expressed as the rotational speed difference). Clutch torque Trn for controlling the vehicle, and an input that is set in proportion to the torque input to the rear wheels so as to obtain a large road surface transmission torque by suppressing wheel slip at the time of sudden start or the like. One is selected from the torque proportional clutch torque Tra, and each of these clutch torques Trn,
The setting unit of Tra will be described in order.

The clutch torque Trn corresponding to the differential is a clutch torque for improving the control accuracy so as to make the vehicle behave in accordance with the driver's intention at the time of turning and for avoiding the tight corner braking phenomenon. In other words, when turning, there is a geometric difference between the left and right wheels and between the front and rear wheels due to the track difference and the body posture.
We want to control the differential between the left and right wheels and between the front and rear wheels so that this differential can be properly tolerated.

By the way, at the time of turning, as a driver's will,
Available information is the required steering angle (pseudo steering angle) δr of the driver.
ef and the vehicle speed (pseudo vehicle speed) Vref, and the differential corresponding clutch torque Trn is set based on the information δref and Vref.

Therefore, the differential corresponding clutch torque Tv
As shown in FIG. 1, the portion related to the setting of the right and left wheel actual rotation speed difference detection unit 500 as the actual rotation speed difference detection means
And ideal rotation speed difference setting means (target rotation speed difference setting
Left and right wheels ideal rotational speed difference setting portion 510 as a stage), and the left and right wheel actual rotational speed difference ΔVrd left and right wheels ideal rotational speed difference ΔV
hr and a differential corresponding clutch torque setting unit 520 as a differential limiting force setting means for setting a clutch torque (differential limiting force) Trn ′ from the hour and a correction for correcting the clutch torque Trn ′ with a rear wheel torque gain k _3. part (k ₃ correction unit) 54
6 is comprised.

The left and right wheel actual rotation speed difference detection section 500
As shown in (1), it is configured to include filters 202c and 202d and a left and right wheel actual rotation speed difference calculation unit 506. The filters 202c and 202d are used to detect the minute vibration component of the data generated by disturbance or the like from the rotation speed data signals RL and RR of the left rear wheel 24 and the right rear wheel 26 detected by the wheel speed sensors 44 and 46, respectively. Is to get rid of. Further, the left and right wheel actual rotation speed difference calculation unit 50
In No. 6, by subtracting the right wheel rotation speed Vrr of the rear wheel from the left wheel rotation speed Vrl of the rear wheel, the actual rotation speed difference of the left and right wheels [the rotation speed difference of the left and right wheels (the left and right rotation difference: this rotation difference is the rotation in the rear differential) ΔVrd is calculated.)

The right and left wheel ideal rotational speed difference setting section 510 includes a driver required steering angle calculating section (pseudo steering angle calculating section) 212 as steering angle detecting means, and a driver required vehicle speed calculating section as vehicle speed data detecting means. Section (pseudo vehicle speed calculation section) 216
And an ideal rotation speed difference setting unit 518 as an ideal operation state setting unit.

As shown in FIG. 10, the driver-requested steering angle calculating section 212 as the steering angle detecting means includes a steering angle sensor 3
0 (first steering angle sensor 30a, the second steering angle sensor 30b installed in the steering wheel, the neutral position sensor 30c) detects data theta ₁ from, theta _2, based on θn sensor corresponding steering angle .delta.h [= f ( θ ₁ , θ ₂ , θn)]
Sensor corresponding steering angle data setting unit 212a for calculating the value of
A lateral acceleration data calculator 212b that averages the data Gyf and Gyr detected by the lateral acceleration sensors 34a and 34b to calculate lateral acceleration data Gy; a direction of the sensor corresponding steering angle δh and a direction of the lateral acceleration data Gy; And a driver request steering angle setting unit (vehicle speed data setting unit) 212d that sets the driver request steering angle δref according to the comparison result of the comparison unit 212c.

The function δh = f (θ ₁ , θ ₂ , θn) for obtaining the sensor-corresponding steering angle δh depends on the specifications of the steering wheel angle sensor. Further, both the sensor-assisted steering angle δh and the lateral acceleration data Gy are, for example, positive in the right turning direction.

In order to compare the directions of the sensor corresponding steering angle δh and the lateral acceleration data Gy, a function SIG (x) relating to the following direction is set for the detection data x. When x> 0, SIG (x) = 1, when x = 0, SIG (x) = 0, when x <0, SIG (x) = − 1. SIG is compared with the direction of the lateral acceleration data Gy.
(Δh) is compared with SIG (Gy).

When the direction SIG (δh) of the sensor corresponding steering angle δh is equal to the direction SIG (Gy) of the lateral acceleration data Gy, the driver required steering angle setting section 212d sets the sensor corresponding steering angle δh to the driver. The required steering angle (steering angle data) δref is set, and the direction S of the sensor corresponding steering angle δh is set.
IG (δh) and the direction SIG (G
If y) is not equal, 0 is set to the driver's requested steering angle δ.
Set to ref.

The direction SIG (δ) of the steering angle δh corresponding to the sensor
h) and the direction SIG (Gy) of the lateral acceleration data Gy are not equal, the reason why the driver-requested steering angle δref is set to 0 is, for example, when the driver performs a steering operation such as counter-steering, the steering position of the steering wheel. And the actual steering angle (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, if the direction SIG (δh) of the sensor corresponding steering angle δh and the direction SIG (Gy) of the lateral acceleration data Gy are not equal, the driver's required steering angle is set. Is set to 0.

Thus, the driver's required steering angle δref is
The steering angle δh based on the steering wheel angle is a steering angle that reflects the driver's intention more than the steering angle δy based on the lateral acceleration Gy, for example. Are suitable. For example, if the driver wants to make a larger turn than the current situation, | δh |> | δ
y |, and the steering angle value | δh |
It is possible to increase the magnitude of the ideal rotation speed difference (slip target value) as compared with adopting δy |, while if the driver wants to hold down the current turn, | δh | <| δy | By using | δh |, it is possible to reduce the magnitude of the ideal rotation speed difference (slip target value) as compared with the case of using the steering angle value | δy |.

The driver-requested vehicle speed calculating section 216 calculates the vehicle speed from the wheel speed when a wheel slips.
As shown in FIG. 11, the wheel speed sensors 40, 42, 44,
Left front wheel 16, right front wheel 18, left rear wheel detected by 46
24 , the rotation speed data signals FL, FR, R of the right rear wheel 26
A wheel speed selector 216a for selecting the second largest wheel speed data from the bottom (lower one) of L and RR, and a driver request for setting a driver requested vehicle speed from the selected wheel speed data and the like It comprises a vehicle speed calculation unit 216c.

In particular, the driver-requested vehicle speed calculating section 216c detects the wheel speed data SVW obtained by filtering the wheel speed data selected by the wheel speed selecting section 216a through the filter 216b to remove noise components, and the longitudinal acceleration sensor 36. The vehicle speed during the slip can be estimated from both data SVW and Gx at a certain time before the slip based on the longitudinal acceleration data Gx obtained by removing the noise component by applying the filtered longitudinal acceleration to the filter 216d. .
That is, assuming that the wheel speed data SVW at a certain point in time is V ₂ and the longitudinal acceleration data Gx is a, the theoretical body speed Vref after a time t from this point is Vref = V ₂ + a
It can be calculated by t.

Further, instead of the longitudinal acceleration data Gx, the wheel speed data SVW or the driver's required vehicle acceleration V'ref obtained by differentiating the driver's required vehicle speed Vref with time may be adopted.

The rotation speed data signals FL, FR, R
The second largest wheel speed data from L and RR is adopted because each wheel usually slips toward the over-rotation side in many cases, so the wheel speed of the lowest rotation speed is originally adopted. It is desirable to use the second wheel speed from the bottom in consideration of data reliability. Then, in the ideal rotation speed difference setting unit 518, the driver's required steering angle δref calculated by the driver's required steering angle calculation unit 212 is calculated.
An ideal rotation speed difference ΔVhr is set based on the map shown in FIG. 13 based on the driver required vehicle speed Vref calculated by the driver required vehicle speed calculation unit 216.

That is, as for the steering angle, the larger the steering angle, the larger the required rotation difference between the left and right wheels. Therefore, the larger the steering angle data δref, the ideal rotation speed difference ΔVh
The value of r also increases. For example, as the steering angle data δref increases in the right turning direction, the value of the ideal rotation speed difference ΔVhr increases in the forward direction (in the direction in which the left wheel has a higher speed), and the steering angle data As δref increases in the left turning direction, the value of the ideal rotation speed difference ΔVhr increases in the negative direction (in the direction in which the speed of the left wheel is lower). Regarding the vehicle speed, at low vehicle speeds, the value of the ideal rotation speed difference ΔVhr increases as the vehicle speed increases, but at high speeds, the tendency of the ideal rotation speed difference ΔVhr to increase with increasing vehicle speed decreases. That is, at high speed, the ideal rotational speed difference ΔVhr between the left and right wheels is determined mainly according to the steering angle data δref.

Referring to FIG. 12, a description will be given of the rotational speed difference ΔVhr between the left and right wheels due to the difference in the orbit radii of the left and right wheels. The example shown in FIG. 12 is the case of a right turn, in which the right wheel speed on the inside of the turn (turning inner wheel speed) is Vi, the left wheel speed on the outside of the turn (outer wheel speed) is Vo, and the center of the left and right wheels is the center. Where V is the vehicle speed at the portion, R is the turning radius of the vehicle (the turning radius at the center of the left and right wheels), Lt is the interval between the left and right wheels (tread), l is the wheelbase, and the distance between the front wheel center and the center of gravity is If, the distance between the center of the rear wheel and the center of gravity is lr, and the vehicle weight is m. If the vehicle body slip angle β is sufficiently small and cos β ≒ 1 and sin β ≒ β, the rotational speed difference ΔVhr between the left and right wheels is expressed as follows. ΔVhr = Vo−Vi = (Lt / R) · V (1.1) where R = (1 + A · V ² ) · 1 / δ (1.2) where A is a stability factor If the front cornering power is kf and the rear cornering power is kr, A = − (m / 2l ² ) · (lf · kf−lr · kr) / (kf · kr)

As shown in the above equations (1.1) and (1.2),
From the vehicle speed V and the steering angle δ, the ideal rotational speed difference ΔV between the left and right wheels
hr can be calculated. However, the pseudo vehicle body speed Vref and the driver's required steering angle δref are used as the vehicle speed V and the steering angle δ, respectively.

Then, the left and right wheel actual rotational speed difference detecting section 500
Is subtracted by the subtractor 522 from the left and right wheel actual rotational speed difference ΔVrd detected by the subtractor 522 and the left and right wheel ideal rotational speed difference ΔVhr set by the left and right wheel ideal rotational speed difference setting unit 510.
ΔVhr) and the obtained difference ΔVr (= ΔVrd−Δ
Vhr) and the ideal rotation speed difference ΔVhr between the left and right wheels are input as data to the differential corresponding clutch torque setting unit 520.

The differential corresponding clutch torque setting section 520
Left and right wheel actual rotation speed difference ΔVrd and left and right wheel ideal rotation speed difference Δ
The clutch torque Trn ′ is set in accordance with the difference ΔVr from Vhr (= ΔVrd−ΔVhr), and the clutch torque Trn ′ is set depending on whether the left and right wheel ideal rotational speed difference ΔVhr is positive or negative.

(I) When ΔVhr ≧ 0 In this case, it is desired that the speed of the rear wheel be higher than that of the front wheel.
Set.

If .DELTA.Vrd.gtoreq..DELTA.Vhr, the rear wheels are slipping due to overspeed, and it is desired to suppress a rear wheel slip by transferring a part of the engine torque largely distributed to the rear wheels toward the front wheels. . Therefore, the clutch torque T
Trn ′ = a × (ΔVrd−ΔVhr) = a × ΔVr (1.3) so that rn ′ increases in proportion to the magnitude of the difference ΔVr (ΔVrd−ΔVhr) (where a Is a proportional constant).

If ΔVhr>ΔVrd> 0, the front wheels are slipping, so if the clutch torque T
When rn 'is increased, the engine torque distributed to the front wheels increases, and the slip of the front wheels is promoted. For this reason, it is desired to reduce the engine torque distributed to the front wheels by making the differential limitation free. Therefore, in this case, the clutch torque Trn 'is set to 0 to set a so-called dead zone.

If 0.gtoreq..DELTA.Vrd, the front wheels are slipping, and it is desired to reduce the slip of the front wheels by increasing the distribution of engine torque to the front wheels. Therefore, Trn ′ = − a × ΔVrd = −a × (ΔVr + ΔVhr) (1.4) is set so that the clutch torque Trn ′ increases in proportion to the magnitude of ΔVrd (where a is proportional. constant).

FIG. 15A shows a map of the relationship between Trn 'and ΔVr, and the map shows the differential corresponding clutch torque Trn based on the difference ΔVr and the ideal rotational speed difference ΔVhr between the left and right wheels. Can be requested.

When ΔVhr = 0, ΔVhr>
There is no dead zone region where ΔVrd> 0.

(Ii) When ΔVhr <0 In this case, it is desired that the speed of the front wheels be faster than that of the rear wheels.
Set.

If .DELTA.Vrd.gtoreq.0, the rear wheel is over-rotating and slipping, and it is desired to suppress a rear wheel slip by transferring a part of the engine torque largely distributed to the rear wheel toward the front wheel. . Therefore, the clutch torque Trn '
Is set so as to increase in proportion to the magnitude of ΔVrd: Trn ′ = a × ΔVrd = a × (ΔVr + ΔVhr) (1.5) (where a is a proportional constant).

If 0>ΔVrd> ΔVhr, the rear wheel is slipping, so if the clutch torque T
If rn 'is increased, the engine torque distributed to the rear wheels increases, and the slip of the rear wheels is promoted. For this reason, it is desired to reduce the engine torque distributed to the rear wheels by setting the differential limit to free. Therefore, in this case, the clutch torque Trn 'is set to 0 to set a so-called dead zone.

If .DELTA.Vhr.gtoreq..DELTA.Vrd, the front wheels are slipping, and it is desired to reduce the slip of the front wheels by increasing the distribution of the engine torque to the front wheels. Therefore, Trn ′ = − a × (ΔVrd−ΔVhr) = − a × ΔVr (1.6) so that the clutch torque Trn ′ increases in proportion to the magnitude of ΔVr (ΔVrd−ΔVhr). (Where a is a proportional constant). Such Tr
When the relationship between n ′ and ΔVr is mapped, FIG.
According to this map, the differential corresponding clutch torque Trn ′ can be obtained from the difference ΔVr and the ideal rotational speed difference ΔVhr between the left and right wheels.

In this way, the differential clutch torque Trn ′ obtained from ΔVr and ΔVhr by the differential clutch torque setting section 520 with reference to the map [FIGS. 15A and 15B] is obtained. As shown in FIG.
6 is k ₃ corrected, so that the differential corresponding clutch torque Trn obtained.

[0078] The correction unit 546 is subjected to a lateral acceleration correction by multiplying the rear wheel torque gain k ₃ to a differential corresponding clutch torque Trn', but adapted to obtain a differential response clutch torque Trn, this Rear wheel torque gain k
₃ is a rear wheel torque gain setting unit 544, which is set as follows.

That is, to the rear wheel torque gain setting section 544, the rear wheel shared torque Tre as the input torque calculated from the rear wheel shared torque calculation section 560 as the input torque setting means is sent. setting the rear wheel torque gain k ₃ depending on the rear wheel torque distributed Tre from the map shown in the block 544.

The rear wheel torque gain k ₃ is a coefficient to be corrected in accordance with the magnitude of the rear wheel shared torque Tre.
In consideration of the magnitude of e, when the rear wheel shared torque Tre is small, the differential corresponding clutch torque Trn is also reduced, and the differential corresponding clutch torque Trn is also increased according to the increase of the rear wheel shared torque Tre. It is trying to increase.

The rear wheel shared torque Tre is equal to Tre = T
rm, the rear wheel torque gain k ₃ becomes a constant value (k ₃ = 1). Trm is the rear wheel shared torque assuming the maximum acceleration running on a dirt road, and Trm = μ [Wr + ( h / l) · W · Gx] · Rt (1.7) where μ is the road surface friction coefficient, h is the height of the center of gravity, Gx is the longitudinal acceleration of the vehicle, and Rt is the tire radius.

Rear wheel shared torque proportional clutch torque Tra
Is a set torque for preventing an initial slip of the rear wheel when the transmission torque is expected to increase when the vehicle suddenly starts from a stop state, and is set from the rear wheel shared torque Tre. This seeking proportional torque Tra' obtained by k ₄ corrected.

In order to obtain such a rear wheel shared torque proportional clutch torque Tra and thereby control the clutch 414, a rear wheel shared torque calculator 560 and a proportional torque calculator 570 are used as shown in FIG. , and k ₄ correcting unit 572 as a proportional adjustment means, is provided and switch 574a. Note that the proportional torque calculation unit 570 and k
_The differential limiting force setting means is constituted by the _four correction units 572.

As shown in FIG. 7, the rear wheel shared torque calculating section 560 includes an engine torque detecting section 264 for detecting an engine torque Te at a certain moment, and a torque converter torque ratio detecting section 266 for detecting a torque converter torque ratio t at that time. A transmission reduction ratio detection unit 276 for detecting the transmission reduction ratio ρm at that time;
To center differential torque (center differential clutch torque) T
The respective information is sent from the center differential torque setting unit 562 that obtains c, and the rear wheel shared torque Tre is calculated from the engine torque Te, the torque converter torque ratio t, the transmission reduction ratio ρm, and the center differential torque Tc. It is configured to be.

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

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

Transmission reduction ratio detecting section 276
Then, based on the selected shift stage information from the shift position sensor 160, a shift stage-reduction ratio correspondence map (not shown)
To determine the transmission reduction ratio ρm.

The center differential torque setting section 562 calculates the center differential torque Tc from the following equation based on the longitudinal acceleration Gx. Tc = (Z / Zr) · (Rt / ρrd) [(Wf-Wa · Zs / Z) · Gx-h / l · Wa · Gx ² ] (2.1) where Zs is the tooth of the sun gear The number, Zr is the number of teeth of the ring gear, Wf is the load shared by the front wheels, Wa is the vehicle weight, ρrd is the final reduction ratio, and Z is Zs + Zr.

The rear wheel shared torque calculating section 560 calculates the rear wheel shared torque Tre based on the engine torque Te, the torque converter torque ratio t, the transmission reduction ratio ρm, and the center differential torque Tc set as described above. This calculation is based on the calculation results Tre ₁ , T
The larger value of re ₂ is adopted.

The arithmetic expression for calculating Tre ₁ is as follows:
This is a front-rear distribution formula of torque assuming that the clutch 28 slips. Tre ₁ = (Te · t · ρ ₁ · ρm−Tc) · ρrd · Zr / (Zr + Zs) (2.2) where ρrd is the final reduction ratio, and Zs is the tooth of the sun gear 121.
The number Zr is the number of teeth of the ring gear 123.

The calculation formula for calculating Tre ₂ is the following formula, assuming that the clutch 28 does not slip. The rear wheel shared torque Tre ₂ obtained by this is the stationary rear wheel shared torque. It is. Tre ₂ = (Wr / W a ) · Te · t · ρ ₁ · ρm · ρrd (2.3) where Wr is the load shared by the rear wheels, Wa is the vehicle weight, and ρrd is the end.
Reduction ratio. Note that the equations (2.2) and (2.3)
Chi (Te ・ t ・ ρ ₁ ・ ρm) is from transmission
Output torque. Hydraulic pressure as front and rear wheel differential limiting means
If the multi-plate clutch 28 does not slip, the center differential 12 is fixed.
It is constant, since the front and rear wheels to rotate integrally, considered at rest
Then, as shown in equation (2.3),
The output torque (Te · t · ρ ₁ · ρm) is added to the reduction ratio ρ
rd and the ratio of the rear wheel shared load Wr to the vehicle weight Wa (=
Wr / Wa) to obtain the rear wheel shared torque (at rest)
Rear wheel sharing torque) Tre ₂ can be obtained. on the other hand,
The hydraulic multi-plate clutch 28 as the front and rear wheel differential limiting means slips.
Then, the center differential 12 allows the differential between the front wheel side and the rear wheel side
Therefore, as shown in equation (2.2), the transmission
From output torque (Te ・ t ・ ρ ₁ ・ ρm)
The value obtained by subtracting the differential torque Tc is added to the reduction ratio ρrd and the
Gear ratio [Zr / (Zr + Zs)]
To obtain the rear wheel shared torque Tre ₁
it can.

[0092] _{Then, Tre = MAX (Tre 1,} Tre 2) than ... (2.4), to determine the rear wheel sharing torque Tre.

[0093] Incidentally, as the rear wheel torque distributed Tre, to adopt a resting rear wheel torque distributed Tre ₂ as described above, the arithmetic expression of Tre ₁ may the value of Tre ₁ becomes negative, thus When the rear wheel is at rest, torque Tre
₂ is adopted. Further, since the clutch control based on the rear wheel shared torque Tre is aimed at starting, the adoption of the stationary rear wheel shared torque Tre ₂ is suitable for this.

The proportional torque calculating section 570 calculates the clutch torque Tra 'proportional to the rear wheel shared torque Tre, and sets the proportional relationship [inclination m (= Tr
a ′ / Tre) is, for example, 0.8], and from Tre to Tr
a ′ is calculated.

The k ₄ correction section 57 as a proportional relation adjusting means
2, the clutch torque Tra thus obtained is
Is multiplied by the rotation difference gain k ₄ to perform correction (proportional relation adjustment). The rotation difference gain k ₄ is set in the rotation difference gain setting section 576 as follows, as shown in FIG.

That is, the rotation difference gain k ₄ is to avoid the tight corner braking phenomenon, and the ideal rotation speed difference ΔVhr set by the ideal rotation speed difference setting section 510 is used.
Are determined according to a map as shown in FIG. The rotation difference gain k ₄ in this map is the ideal rotation speed difference Δ
The relationship with Vhr is expressed by the following equation. K ₄ = 0.9 × (| ΔVhrmax || ΔVhr |) /|ΔVhrmax|+0.1 (2.5) where ΔVhrmax = MAX | ΔVhr (δ = MAX) | and coefficient 0.9 and constant 0 ..1 sets the lower limit of k ₂ to .0.
This is to make it 1.

As described above, the rotation difference gain k decreases linearly as the ideal rotation speed difference ΔVhr increases.
By correcting by ₄ , the ideal rotation speed difference Δ
When Vhr increases, the clutch torque Tra 'is reduced so that turning performance (performance that can prevent the tight corner braking phenomenon) is prioritized over sudden starting performance.

In order to obtain the clutch torque Tra,
The clutch torque Tra is set so as to have a proportional relationship with the ideal rotation speed difference ΔVhr, and the proportional coefficient is determined by the magnitude of the rear wheel shared torque Tre as the magnitude of the rear wheel shared torque Tre is increased. (= Tra / ΔV
hr)] may be provided.

Further, the clutch 574a is connected to the judgment means 5
At the time of low vehicle speed (in this example, Vref
<20 km / h), the clutch torque Ta
Is output as data, but when the vehicle speed further increases (Vref ≧ 20 km / h), it is turned off, and a value 0 is output as data of the rear wheel shared torque proportional clutch torque Tra. This is because the rear wheel shared torque proportional control is intended to secure the transmission torque to the road surface by preventing the wheels from slipping. According to the rear wheel shared torque proportional clutch torque Tra, the tight corner braking phenomenon occurs. May occur, or other control speeds may be excluded in situations where slippage is required.To avoid these, a condition has been established in which this rear wheel shared torque proportional control is performed only at a constant vehicle speed. -ing

The clutch torques of the differential corresponding clutch torque Trn and the rear wheel shared torque proportional clutch torque Tra are set for each control cycle repeated at an appropriate timing, and sent to the maximum value selection section 580. The maximum value selection unit 580 selects the maximum clutch torque Trn and Tra (this clutch torque is referred to as Tr ′) for each control cycle. However, when the switch 574a is OFF, 0 is sent as the clutch torque Tra, so that the maximum value selection unit 5
At 80, the clutch torque Trn is selected.

[0102] Thus clutch torque Tr' selected in, as shown in FIG. 9, is subjected to k ₅ corrected by k ₅ correcting unit 584. The k ₅ correction, when the traction control is being performed is a correction that reduces the clutch torque Tr ', for example, k ₅ = 0.5, and traction control during traction control (when ON) When it is not (OFF), k ₅ = 1.
It is set to 0.

The clutch torque Tr obtained by the correction in this way is sent to a torque-current conversion section 586, where it is converted into a clutch supply current I such that the set clutch torque Tr is obtained. It has become so. Here, the clutch supply current I is obtained from the clutch torque Tr by using a map (see the block 586 in FIG. 9).

As described above, when the clutch supply current I is obtained, the current limiter (limiter) 588 holds the clutch supply current I at the limit value when the clutch supply current I exceeds the limit value (for example, 3 A). It has become.

Information on the clutch supply current I that has passed through the limiter 588 as described above is taken in by the peak hold filter 590. This peak hold filter 590 is a kind of limiter for preventing an excessive sudden change in the current so that hunting does not occur in the control due to the sudden change in the current. 4 A / s), and a slightly lower limit speed (for example, 5.2 A)
/ S).

When the information of the clutch supply current I is sent such that the speed of the current change exceeds such a limit, the control current according to the limit speed is stopped.

Further, the control current I having passed through the filter 590 passes through the switch 592a and passes through the EMCD coil 42
0. The switch 592
a, if ABS control (anti-lock brake control) is being performed by a signal from the determination means 592 (O
It is turned off (if N state), and turned on if ABS control is not performed. That is, the signal of the control current I is transmitted on condition that the ABS control is not performed. This is because it is necessary to reliably operate the ABS during the ABS control, and at this time, the torque distribution state of the left and right wheels is controlled because it interferes with the ABS control and is not preferable.

The coil 420 generates a magnetic force in accordance with the control current I sent in this way to adjust the connection state of the clutch 414.

Since this 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. 9, first, each control element is initially set (step a1), learning of a neutral position of a steering angle (step a2), and learning of preload of a clutch (step a3) are performed. The front and rear wheel driving force distribution control is performed while controlling the clutch 28 in accordance with the duty (step a4), and further, the rear differential is controlled (step a5).

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

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

Of these, the control of the rear differential (step a)
5) will be described with reference to the flowcharts of FIGS.

As shown in FIG. 20, first, the wheel speeds FR,
FL, RR, RL, steering angle θ₁, Θ_Two , Θn, lateral acceleration G
y, longitudinal acceleration Gx, throttle opening θth, engine rotation
Number of rotations Ne, transmission rotation number Nt, selective shift
Detects various data, such as steps, and calculates the driver
Calculate the required vehicle speed Vref, the driver's required steering angle δref, etc.
Step j1).

Then, an ideal rotational speed difference ΔVhr between the left and right wheels is obtained from the driver required vehicle speed Vref and the driver required steering angle δref according to a map (step j2).
Setting the rotational difference gain k ₄ in accordance map (step j3).

Further, two kinds of rear wheel shared torques Tre1, Tre1, based on the engine torque Te, the torque converter torque ratio t, the transmission reduction ratio ρm, and the center differential torque Tc.
Tre2 is calculated (steps j4 and j5). And
The larger of the two types of rear wheel shared torques Tre1 and Tre2 is adopted as rear wheel shared torque Tre (step j).
6).

Further, a rear wheel torque gain k ₃ is obtained from the rear wheel shared torque Tre according to a map (step j).
7). Then, a differential corresponding clutch torque Trn from left and right wheel actual rotational speed difference ΔVrd left and right wheels ideal rotational speed difference ΔVhr rear wheel torque gain k ₃ Metropolitan (Step j
8).

[0117] the other hand, the rotational difference gain k ₄ Metropolitan and rear wheel torque distributed Tre, determine the rear wheel allotted torque proportional clutch torque Tra (step j9).

Further, at step j10, the larger one of these clutch torques Trn and Tra is set as the set clutch torque Tr '.

Further, at step j11, the clutch torque Tr ′ determined in this way is applied to the engine output state,
That is, k ₅ corrected by whether the traction control to obtain the clutch torque Tr.

Subsequently, the clutch torque Tr thus corrected and obtained is subjected to torque-current conversion to obtain a clutch supply current I (step j12), and the clutch supply current I is set to a limit value (for example, 3A). ) Is determined (step j13). If I exceeds the limit value, I is held at the limit value (step j14).

Thus, the clutch supply current I subjected to the limiter processing is filtered by the peak hold filter 590 (step j15), and it is determined whether or not ABS control (anti-lock brake control) is being performed (step j16). If the ABS control is performed, the clutch supply current I is set to 0 (step j1).
7) R / D control (rear differential control), that is, differential limiting control through the EMCD coil 420 is performed (step j).
18).

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

First, from the right wheel speed Vrl to the left wheel speed V
The difference ΔVrd (= Vrl−Vrr) obtained by subtracting rr is calculated (step k1), and the difference (actual rotational speed difference between the left and right wheels) ΔVrd is calculated as described above (step j).
3) The difference ΔVr (= ΔVrd−ΔVhr) is obtained by subtracting the obtained ideal rotational speed difference ΔVhr between the left and right wheels (step k2).

Then, in step k3, it is determined whether or not the above-described ideal rotational speed difference ΔVhr of the left and right wheels is 0 or more. If ΔVhr is 0 or more, the process proceeds to step k4, and if ΔVhr is less than 0, the process proceeds to step k5.

When the process proceeds to step k4, the map [FIG.
(A) to obtain the clutch torque Trn from ΔVr.
'Is set.

Specifically, if ΔVr ≧ ΔVhr,
The clutch torque Trn ′ is equal to the difference ΔVr (= ΔVrd−ΔV
Trn ′ = a × (ΔVrd−ΔVhr) = a × ΔVr so as to increase in proportion to the magnitude of (hr) (where a is a proportional constant).

If ΔVhr>ΔVrd> 0,
The clutch torque Trn 'is set to 0 to set a so-called dead zone.

If 0 ≧ ΔVrd, Trn ′ = − a × ΔVrd = −a × (ΔVr + ΔVh) so that the clutch torque Trn ′ increases in proportion to the magnitude of ΔVrd.
r) (where a is a proportional constant).

When ΔVhr = 0, ΔVhr> Δ
There is no dead zone region where Vrd> 0.

When the process proceeds to step k5, the map [FIG.
(Refer to (b)) to calculate the clutch torque Trn from ΔVr.
'Is set.

Specifically, if ΔVrd ≧ 0, Trn ′ = a × ΔVrd = a × (ΔVr + ΔVhr) is set such that the clutch torque Trn ′ increases in proportion to the magnitude of ΔVrd (where a Is a proportional constant).

If 0>ΔVrd> ΔVhr,
The clutch torque Trn 'is set to 0 to set a so-called dead zone.

Further, if ΔVhr ≧ ΔVrd, the clutch torque Trn ′ becomes ΔVr (ΔVrd−ΔVhr)
Trn ′ = − a × (ΔVrd−ΔVhr) = − a × ΔVr so as to increase in proportion to the magnitude of (where a is a proportional constant).

As described above, the differential corresponding clutch torque Trn 'obtained in steps k4 and k5 is
6 lateral acceleration corresponding corrected by being integrated with the lateral G gain k ₁ in (step k6), differential corresponding clutch torque Tr
n is obtained.

[0135] Such a differential corresponding clutch torque Trn
With this setting, the magnitude of the clutch torque Tr is appropriately set, and while the differential between the left and right wheels is appropriately allowed, the vehicle can behave in accordance with the driver's will during turning.

In particular, the direction SIG of the sensor corresponding steering angle δh
When (δh) is not equal to the direction SIG (Gy) of the lateral acceleration data Gy, the driver-requested steering angle is set to 0, so that, for example, when the driver operates a steering wheel such as a counter steer, etc. Even if the steering position of the steering wheel is different from the actual steering angle (turning state) of the vehicle, inappropriate data is not adopted, which contributes to improvement in control performance.

Further, as the driver request vehicle speed Vref,
Since the second largest wheel speed data from among the rotation speed data signals FL, FR, RL, RR is adopted,
Data reliability is ensured.

As shown in the map (see FIG. 13), the setting of the ideal rotation speed difference ΔVhr increases as the steering angle increases, and increases at low vehicle speeds as the vehicle speed increases. The rear wheels are likely to slip at high speeds, and the responsiveness of the posture of the vehicle body required at higher speeds is ensured. As for the steering angle, the larger the steering angle, the larger the difference in rotation required for the front and rear wheels, and this is appropriately allowed, and there is an advantage that the tight corner braking phenomenon can be avoided.

On the other hand, the calculation of the above-described rear wheel shared torque proportional clutch torque Tra is performed as shown in FIG.

First, the rear wheel shared torque Tre is converted into a clutch torque Tra 'proportional to this (step m).
1), the clutch torque Tra' by the rotation difference gain k ₄
Is corrected to obtain the clutch torque Tra (step m
2).

Further, when the vehicle speed is low (Vref <20 km / h in this example) by the determining means 574, the clutch obtained in step m2 described above is obtained at low vehicle speed (Vref <20km / h). Although the torque Tra is used as a control signal, if the vehicle speed is higher than this (Vref ≧ 20 km /
h), 0 is set as a control signal for the rear wheel shared torque proportional clutch torque Tra.

When the rear wheel sharing torque Tre and the rear wheel sharing torque proportional clutch torque Tra proportional to the ideal rotational speed difference ΔVhr are applied to the left and right wheels at the time of starting or sudden acceleration from a low speed, for example. Differential can be appropriately restricted, and high torque can be transmitted to the road surface as appropriate, preventing tire slip during starting and sudden acceleration, stabilizing running and improving turning performance, etc. The driving performance can be improved at the same time, and the durability of the driving system can be improved.

This rear differential mechanism can be applied to vehicles other than the four-wheel drive vehicle, and can also be applied to the front differential.

[0144]

As described above in detail, according to the differential limiting control device for a four-wheel drive vehicle according to the first aspect of the present invention, the driving force from the engine is applied to the front wheel drive shaft and the rear wheel drive shaft. Front and rear wheel drive force distribution means for unequally distributing the front and rear wheel drive shaft, front and rear wheel differential limiting means for selectively limiting the differential between the front wheel drive shaft and the rear wheel drive shaft Front and rear wheel differential limiting force setting means for setting a differential limiting force applied by the means,
The front wheel and the rear wheel are provided on the side where the driving force is distributed more by the front and rear wheel driving force distribution means, and the driving force input from the front and rear wheel driving force distribution means is applied to the left wheel driving shaft and the right wheel driving. Left and right wheel driving force distributing means for distributing to the shaft; left and right wheel differential limiting means capable of selectively limiting the differential between the left wheel driving shaft and the right wheel driving shaft; And left and right wheel differential limiting force setting means for setting the differential limiting force to be actuated according to the above, wherein the left and right wheel differential limiting force setting means is configured based on the driving force input to the left and right wheel driving force distribution means. A driving force value calculated from the differential limiting force set by the front and rear wheel differential limiting force setting means, and a differential limiting force to be applied by the left and right wheel differential limiting means; Of the front wheel side drive shaft and the rear wheel side drive shaft The difference between the left and right wheels is appropriately limited when starting or when suddenly accelerating from a low speed, because the setting is made using the larger value of the driving force As a result, it is possible to transmit high torque to the road surface as appropriate, preventing tire slip when starting or suddenly accelerating, improving running performance by simultaneously achieving running stability and improving turning performance In addition, it also contributes to improving the durability of the drive system. That is, as the input torque to the left and right wheel driving force distribution means, the calculated value of the input torque when it is assumed that a slip occurs in the differential limiting means between the front and rear wheels, and the slip in the differential limiting means between the front and rear wheels. Is compared with the input torque calculation value in the case where it is not assumed that the differential limiting means between the front and rear wheels is slipping by using the larger input torque for calculating the differential limiting force. Even if the calculated input torque value assumes a negative value in the calculation, this negative value is not adopted as the input torque, and accurate differential limiting force control can be performed.

Further, according to the differential limiting control device for a four-wheel drive vehicle of the present invention, the above effects can be obtained more reliably.

[Brief description of the drawings]

FIG. 1 is a block diagram showing an overall configuration of a vehicle differential limiting control device as one embodiment of the present invention.

FIG. 2 is an overall configuration diagram of a drive torque transmission system including a differential control device according to the present embodiment.

FIG. 3 is a cross-sectional view showing a rear differential as a left and right wheel differential of the embodiment.

FIGS. 4A and 4B are cross-sectional views taken along the line AA of FIG. 3;

FIG. 5 is a configuration diagram illustrating a rotation speed difference correspondence control unit according to the present embodiment.

FIG. 6 is a configuration diagram illustrating a rear wheel torque gain correction unit according to the embodiment.

FIG. 7 is a configuration diagram illustrating a part of an input torque corresponding control unit according to the present embodiment.

FIG. 8 is a configuration diagram illustrating a part of an input torque corresponding control unit according to the embodiment.

FIG. 9 is a configuration diagram of a portion from a maximum value selection section to a control current output section in the present embodiment.

FIG. 10 is a configuration diagram showing details of a steering angle detection unit of the present embodiment.

FIG. 11 is a configuration diagram illustrating details of a vehicle speed detection unit of the present embodiment.

FIG. 12 is a plan view schematically showing a wheel state for explaining an ideal rotational speed difference of the embodiment.

FIG. 13 is a diagram showing an ideal rotation speed difference setting map according to the present embodiment.

FIG. 14 is a diagram illustrating a rotation difference gain setting map according to the present embodiment.

FIGS. 15 (a) and (b) are diagrams showing differential-corresponding clutch torque setting maps according to the present embodiment.

FIG. 16 is a diagram illustrating an example of an engine torque map according to the present embodiment.

FIG. 17 is a diagram illustrating an example of a transmission torque ratio map according to the present embodiment.

FIG. 18 is a center differential input torque setting map of the present embodiment.

FIG. 19 is a flowchart showing a flow of control of the entire vehicle including the apparatus of the present embodiment.

FIG. 20 is a flowchart illustrating a flow of control of a rear differential according to the present embodiment.

FIG. 21 is a flowchart illustrating a flow of setting a clutch torque corresponding to a rotational speed difference according to the present embodiment.

FIG. 22 is a flowchart illustrating a flow of setting an input torque corresponding clutch torque according to the present embodiment.

[Explanation of symbols]

Reference Signs List 2 engine 4 torque converter 6 automatic transmission 8 output shaft 10 intermediate gear (transfer idler gear) 12 center differential (center differential) 14 differential gear device for front wheel 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 differential (EMCD) as differential gear for rear wheel 23 Differential limiter 24 Left rear wheel 26 Right rear wheel 25L, 25R For rear wheel axle 27 front wheel output shaft 27a hollow shaft member 28 multi-plate clutch 29 rear wheel output shaft 30,30a as differential limiting means, 30b, 30c wheel angle sensor 32 the steering wheel 34, 34a, 34b lateral acceleration sensor 36 Rear acceleration sensor 38 Throttle sensor 39 Engine key switch 40, 42, 44, 46 Wheel speed sensor 48 Controller 48a Rear differential controller 50 Antilock brake device 50A Brake switch 121 Sun gear 122 Planetary pinion (planetary gear) 123 Ring gear 125 Planet carrier 160 Shift lever Position sensor (shift range detecting means) 160A Shift lever of automatic transmission 170 Engine speed sensor 180 Transmission speed sensor 202a-202d Filter 212 Driver required steering angle calculator as steering angle detector (pseudo steering angle calculator) 212a Sensor-based steering angle data setting unit 212b Lateral acceleration data calculation unit 212c Comparison unit (comparing means) 212d Driver-requested steering angle setting unit ( (Vehicle speed data setting unit) 216 driver required vehicle speed calculation unit (pseudo vehicle speed calculation unit) 216a wheel speed selection unit 216c driver required vehicle speed calculation unit 216d filter 218 ideal operating state setting unit Ideal rotation speed difference setting unit 262a, 262b, 262c Filter 264 Engine torque detection unit 266 Torque converter torque ratio detection unit 276 Transmission reduction ratio detection unit 401 Input shaft 402 Drive pinion gear 403 Crown gear 404 Power transmission annular member 405 First housing 406 second housing 407 ring gear 408 sun gear 409 carrier 410a, 410b shaft 411a, the multi-plate clutch 414a as 411b planetary gear 412 bearings 413 case 414 differential limiting means, 41 b clutch disc 415a, 145b holder portion 416 hollow shaft 417 drives 418 the annular support member 419 solenoid as a magnet 420 differential limiting mechanism control means (E
MCD coil) 421 Ball 423 Annular member 424 Groove 425 Chamber 426 Member on second housing side 427 Clutch 428 Bearing 429 Force direction conversion mechanism 430 Electromagnetic clutch mechanism (EMCD) 431 Bolt 500 Left and right wheels as actual rotation speed difference detecting means Actual rotation speed difference detection unit 506 Left and right wheel actual rotation speed difference calculation unit 510 Ideal rotation speed difference setting means (target rotation speed difference setting unit)
Right and left wheel ideal rotation speed difference setting unit 518 as means) Ideal rotation speed difference setting unit 520 as ideal operation state setting unit 520 Differential clutch torque setting unit (differential limiting force setting means) 522 Subtractor 546 Correction unit (k ₃ corrector) after 544 rear wheel torque gain setting unit 560 wheel torque distributed computing unit 562 center differential torque setting unit 570 proportional torque computing section (a part of the differential limiting force setting means) 572 k ₄ correction of the proportional adjustment means part (a part of the differential limiting force setting means) 574 determining unit 574a switches 576 rotation difference gain setting unit 580 maximum value selection unit 584 k ₅ correcting unit 586 torque - current converting unit 588 the current limiting unit (limiter) 590 peak Hol de filter 592 Judging means 592a Switch

Continuation of the front page (72) Inventor Yoshihito Ito 5-33-8 Shiba, Minato-ku, Tokyo Inside Mitsubishi Motors Corporation (56) References JP-A-62-292529 (JP, A) 96938 (JP, U)

Claims (2)

(57) [Claims] 1. A front / rear wheel drive force distribution means for unequally distributing a drive force from an engine to a front wheel drive shaft and a rear wheel drive shaft, and a differential between the front wheel drive shaft and the rear wheel drive shaft. Front and rear wheel differential limiting means for selectively limiting the front and rear wheel differential limiting force setting means for setting a differential limiting force applied by the front and rear wheel differential limiting means; Left and right wheel driving forces provided on the side where the driving force is distributed more by the wheel driving force distribution means and distributing the driving force input from the front and rear wheel driving force distribution means to the left wheel driving shaft and the right wheel driving shaft. Distributing means, left and right wheel differential limiting means capable of selectively limiting the differential between the left wheel drive shaft and right wheel drive shaft, and a differential limiting force applied by the left and right wheel differential limiting means. A left and right wheel differential limiting force setting means for setting the left and right wheel differential limiting force. Setting means sets a differential limiting force applied by the left and right wheel differential limiting means based on the driving force input to the left and right wheel driving force distribution means, and sets the front and rear wheel differential limiting as the driving force. The greater of the driving force value calculated from the differential limiting force set by the force setting means and the driving force when the front wheel-side drive shaft and the rear wheel-side drive shaft are directly connected. A differential limiting control device for a four-wheel drive vehicle, characterized in that the above setting is performed using 2. A left-right wheel differential limiting force setting device comprising: target rotation speed difference calculating means for calculating a target rotation speed difference between the left wheel drive shaft and the right wheel drive shaft to be generated by turning of the vehicle. 2. The four-wheel drive according to claim 1, wherein the means is configured to decrease the differential limiting force with an increase in the target rotation speed difference calculated by the target rotation speed difference calculation means. Differential limit control device for driving vehicles.