JPH0813609B2 - Drive force distribution controller for four-wheel drive vehicle - Google Patents

Description

Description: TECHNICAL FIELD The present invention relates to a drive force distribution control device for a four-wheel drive vehicle used for a drive force distribution clutch of a transfer device for a four-wheel drive vehicle.

(Prior Art) As a conventional drive force distribution control device for a four-wheel drive vehicle, for example, a device described in JP-A-61-157437 is known.

This conventional device calculates the rotational speed difference between the front and rear wheels based on the sensor signals from the front and rear wheel rotational speed sensors, and the larger the difference between the front and rear wheel rotational speeds, that is, the larger the occurrence of the drive wheel slip, the engagement of the transfer clutch. It is intended to increase the force and change the distribution of the driving force to the four-wheel drive side to promptly suppress the driving wheel slip.

(Problems to be Solved by the Invention) However, in such a conventional device, the clutch engagement force control characteristic with respect to the front-rear wheel rotational speed difference is uniquely determined, and therefore the friction of the traveling road surface is reduced. It left a problem that it could not immediately respond to changes in coefficients and running conditions.

For example, if you set the control characteristics (characteristics that do not distribute a large driving force to the front and rear wheels) so that you can easily drive under conditions such as dry dry roads and power drifting, spin on low friction coefficient roads such as ice and snow roads. There is a strong tendency that running stability and stack escapeability will not be sufficient, and conversely, if set to a control characteristic (characteristic that distributes a large amount of driving force to the front wheels) to obtain running stability on a low friction coefficient road In the case of a high friction coefficient road, the torque transmitted to the front wheels is large, and there is a tendency to drift out early.

On the other hand, the present applicant proposes a solution to the above-mentioned problem in the specification of Japanese Patent Application No. 61-288498, that is, the road surface friction coefficient is indirectly detected by the magnitude of the lateral acceleration to determine the lateral acceleration. We proposed a device with a control content that increases the clutch engagement force increase rate relative to the front-rear wheel rotation speed difference when the torque is large, and increases the clutch engagement force increase rate relative to the front-rear wheel rotation speed difference when the lateral acceleration is small.

However, since there is a time delay including the dead time of the control system, especially when the lateral acceleration is small and the increase rate of the clutch engagement force with respect to the front and rear wheel rotation speed difference is large, the transmission torque of the front and rear wheel rotation speed difference ΔN is The control gain is large and torque hunting occurs.

Therefore, in order to secure the stability of the control system, it is conceivable to introduce a time differential element of the front and rear wheel rotation speed difference, which is a phase advance element, and increase / decrease the clutch engagement force according to the value.

However, if the increase / decrease rate of the clutch engaging force is set to a unique rate, the following problems will occur.

When the increase / decrease rate of clutch engagement force is set small,
When turning on a high friction coefficient road, the clutch engagement force increase rate is small and the turning performance is improved, but when traveling straight or on a low friction coefficient road, there is a delay in the clutch engagement force increase, and straight running stability and low The running performance on the friction coefficient road is poor.

When the increase / decrease rate of clutch engagement force is set to a large value,
Although straight running stability and low friction coefficient road running performance are improved, when accelerating on a high friction coefficient road, the clutch engagement force increase rate is too large and the steer characteristic becomes a strong understeer characteristic.

(Means for Solving the Problems) The present invention has been made for the purpose of solving the above problems, and in order to achieve the object, the present invention has the following solution means. did.

The solution means of the present invention will be explained with reference to the conceptual diagram of the claims shown in FIG. 1. The solution torque is provided in the middle of an engine drive system for distributing and transmitting the engine drive force to the front and rear wheels, and the transmission torque can be changed by an external clutch engagement force. A drive force distribution control device for a four-wheel drive vehicle including drive system clutch means 1 and clutch control means 3 for outputting a control signal for increasing / decreasing the clutch engagement force based on a detection signal from a predetermined detection means 2. In the above, as the detecting means 2, front and rear wheel rotational speed difference detecting means 201, front and rear wheel rotational acceleration difference detecting means 202 for detecting a temporal change rate of the front and rear wheel rotational speed difference, and lateral acceleration detecting means 203 for detecting lateral acceleration. The clutch control means 3 increases the clutch engagement force in accordance with the front-rear wheel rotational speed difference and the front-rear wheel rotational acceleration difference, and the increase ratio value increases the lateral acceleration. And wherein the decreasing characteristic is the means obtained As the.

(Operation) For straight running, starting with a small lateral acceleration, turning with a low friction coefficient road, etc., the control characteristic with a large increase rate is obtained, so the clutch engagement force is large, and the driving force distribution tendency is mainly on the four-wheel drive side. Therefore, it is possible to improve the starting performance, improve the straight running stability, and prevent an early spin or drift out during turning.

Here, in terms of control stability, if the control gain is large, the control system easily oscillates. This is because, in the case of control, there is a dead time due to the control cycle in the electronic control circuit, a dead time in the hydraulic circuit, and a time delay of the first-order lag system. In order to guarantee this and prevent the control system from oscillating in a stable direction, a front-rear wheel rotational acceleration difference, that is, a time differential element is included. As a result, the gain can be increased while suppressing the oscillation, the response of the control system is improved, and it is effective when it is desired to move to the four-wheel drive side faster on a low friction coefficient road or the like.

When the vehicle is turning on a high friction coefficient road with a large lateral acceleration, a control characteristic with a small increase rate of the clutch engagement force is obtained, so the clutch engagement force T is small, and a driving force distribution tendency mainly on the two-wheel drive side tends to occur. Good turning performance on a high friction coefficient road can be realized without drifting out.

At this time, since the increase rate value is small, the control system is originally stable and the need for the differential element is small. In contrast,
Since the increase rate value is small, the clutch engagement force due to the front and rear wheel rotational acceleration difference is also small, and the influence of the differential element is suppressed to a small level.

Therefore, the increase rate of the clutch engagement force can be increased for straight running, starting with a small lateral acceleration, and low-friction coefficient road turning, and for increasing the clutch engagement force during a high-friction coefficient road turning with a large lateral acceleration. It is possible to reduce the value, and it is possible to achieve both suitable running performance while suppressing oscillations in straight running, starting, turning with a low friction coefficient road, and improvement of turning performance during turning with a high friction coefficient road.

(Examples) Hereinafter, examples of the present invention will be described in detail with reference to the drawings. In describing this embodiment, a drive force distribution control device for a four-wheel drive vehicle based on rear wheel drive will be taken as an example.

 First, the configuration will be described.

The four-wheel drive vehicle to which the driving force distribution control device D of the first embodiment is applied is, as shown in FIG.
Engine 11, transmission 12, transfer input shaft 13, rear wheel side drive shaft 14, multi-disc friction clutch (drive system clutch means) 15, rear differential 16, rear wheel 17, front differential 18, front wheel 19, gatorane 2
0, front wheel side drive shaft 21 is provided.

The multi-plate friction clutch 15 is the transfer input shaft.
It is interposed between the front wheel drive shaft 21 and the rear wheel drive shaft 14 (directly connected to the rear wheel drive shaft 14).
It is possible to change the transmission torque to the side.

Incidentally, FIG. 3 shows a specific example of the transfer device 10, and the multi-plate friction clutch 15, gears and shafts are housed in a transfer case 22.

In FIG. 3, 15g is a dish plate, 15h is a return spring, 24 is a clutch pressure oil input port, 25 is a clutch pressure oil passage, 26 is a rear wheel side output shaft, 27 is a lubricating oil passage, and 28 is a speedometer. Pinion, 29 is an oil seal, 30 is a bearing, 31 is a needle bearing, 32 is a thrust bearing, and 33 is a joint flange.

Next, the hydraulic control device 40 that supplies a predetermined clutch hydraulic pressure P for engaging the multi-disc friction clutch 15 is
As shown in the figure, the front wheel rotation speed sensor is used as the detection means.
41, a rear wheel rotation speed sensor 42, a lateral acceleration sensor 43,
As a control processing means, a control unit 45 is provided,
As a control actuator, an electromagnetic proportional relief valve 46
It has.

The left and right wheel speed sensors 41, 42 are provided in the left and right front wheel 19 positions and in the middle of the rear wheel side drive shaft 14, respectively, and a sensor rotor fixed to the shaft and a pickup for detecting a magnetic force change due to rotation of the sensor rotor. A rotation sensor or the like is used, and signals (nf) and (nr) corresponding to the shaft rotation are output from these sensors 41 and 42, respectively.

The lateral acceleration sensor 43 directly detects the lateral acceleration of the vehicle and outputs a signal (yg) corresponding to the lateral acceleration Yg.

As the control unit 45, a control circuit centering on a vehicle-mounted microcomputer is used, and as an internal circuit, an input interface 451, RAM452, ROM453, CPU45
4, output interface 455 is provided.

The electromagnetic proportional relief valve 46 outputs the command current signal (i) from the control unit 45 as the command current value I.
^{When *} = 0, the clutch oil pressure P = 0, but when the output of the command current signal (i) is the command current value I ^* > 0, the valve moves in the closing direction and the line pressure from the hydraulic power source 50 is reduced. By controlling the drain oil amount, the clutch hydraulic pressure P is set according to the magnitude of the command current value I ^* (Fig. 5). The relationship between the clutch hydraulic pressure P and the clutch engaging force T is expressed by the following equation (6th
Figure).

P = T / (μ ・ S ・ 2n ・ Rm) where μ: friction coefficient of clutch plate, S: pressure acting area on piston, n: number of friction discs, Rm: effective radius of torque transmission of friction disc, clutch When the hydraulic pressure P is increased, the clutch engagement force T is also increased proportionally.

 Next, the operation will be described.

First, the flow of driving force distribution control operation in the first embodiment will be described.
This will be described with reference to the flowchart shown in FIG.

In the first embodiment, the lateral acceleration Yg is directly detected by the lateral acceleration sensor 43, and the clutch engagement force T is increased in accordance with the increase in the front / rear wheel rotational speed difference ΔN and the front / rear wheel rotational acceleration difference Δ, and its increase rate value. Is an example in which a characteristic that decreases as the lateral acceleration Yg increases is obtained.

In step a, the front wheel rotation speed is detected from each sensor 41, 42, 43.
Nf, rear wheel rotation speed Nr, and lateral acceleration Yg are read.

At step b, the front and rear wheel rotation speed difference ΔN from the front wheel rotation speed Nf and the rear wheel rotation speed Nr read at step a.
(= Nr-Nf) is calculated.

The front-rear wheel rotation speed difference ΔN, which is the time change rate, is calculated by the front-rear wheel rotation speed difference ΔN obtained in the current control cycle and the front-rear wheel rotation speed difference ΔN ₀ obtained in the previous control cycle.
(= ΔN−ΔN ₀ ) is calculated.

At step c, the target proportional coefficient Kt is calculated from the lateral acceleration Yg read at step a.

The arithmetic expression is Kt = A ₁ / Yg, where A ₁ is a constant.

In the next step d to step l, a low-pass filter that restricts the rate of change in the increasing rate of the real proportional coefficient Kx and the changing speed in the increasing direction is realized.

In step d, the difference ΔK between the target proportional coefficient Kt of this time obtained in step c and the actual proportional coefficient Ko one cycle before is calculated.

In step e, it is determined whether ΔT is positive or negative, and it is determined whether the proportional coefficient K is in the increasing direction or the decreasing direction, and the subsequent processing route is changed.

If ΔK is positive in step e, that is, the proportional coefficient K is in the increasing direction, it is determined in step g whether or not the change width is larger than the set value A ₂ , and the set value A ₂ is the proportional coefficient. It becomes the value of the low-pass filter when K increases.

When ΔK is larger than A ₂ in step g, the process proceeds to step j, the target proportional coefficient Kt is filtered, and the actual proportional coefficient Kx is obtained by the arithmetic expression of Ko + A ₂ .

If ΔK is smaller than A ₂ in step g, the process proceeds to step k, and the target proportional coefficient Kt remains the actual proportional coefficient Kx.
Is set as.

On the other hand, if ΔK is negative in step e, that is, the proportional coefficient K is in the decreasing direction, it is determined in step f whether or not the change width is larger than the set value A ₃ , and the set value A ₃ is proportional. It is the value of the low-pass filter when the coefficient K decreases.
The set value A ₃ is larger than the set value A ₂ .

Then, if | KΔ | is larger than A ₃ in step f,
It proceeds to step i, the target proportional coefficient Kt is filtered, the actual proportional coefficient Kx is obtained by the calculation formula of the Ko-A _3.

If | KΔ | is smaller than A ₃ in step f, the process proceeds to step h, and the target proportional coefficient Kt remains the actual proportional coefficient.
Set as Kx.

At step l, the value of the actual proportional coefficient Kx obtained in the current control cycle is stored as Ko for calculating ΔK.

In step m, the target clutch engagement force T is calculated by Kx and ΔN.

 The calculation formula is as follows.

When ΔN <0; T = 0 When ΔN ≧ 0; T = Kx * (ΔN + B * Δ) where B is a constant.

In step n, the command current signal (i) is output according to the command current value I ^* that obtains the clutch pressure P corresponding to the clutch engaging force T obtained in step m.

 Next, the operation during traveling will be described.

With the above-described control operation, as shown in FIG. 7, the larger the lateral acceleration Yg, the smaller the control characteristic of the proportional coefficient Kx, which is the increase rate value of the clutch engaging force T, and the low-pass filtering process of the proportional coefficient Kx. The change speed is fast on the decreasing change side and slow on the increasing change side. As a result, the following operation is achieved.

(A) During low lateral acceleration running When the lateral acceleration Yg is small, the clutch engagement force T is large, mainly for four-wheel drive, because a control characteristic with a large increase rate can be obtained for straight running, starting, and low friction coefficient road turning. The driving force tends to be distributed on the side, improving the starting performance, improving the straight running stability, and preventing early spinning or drifting out during turning.

Here, in terms of control stability, if the control gain is large, the control system easily oscillates. This is because, in the case of control, there is a dead time due to the control cycle in the electronic control circuit, a dead time in the hydraulic circuit, and a time delay of the first-order lag system. In order to guarantee this and prevent the control system from oscillating in a stable direction, a front-rear wheel rotational acceleration difference Δ, that is, a time differential element is included. As a result, the proportional coefficient Kx (gain) can be increased while suppressing oscillation, the response of the control system is improved, and it is effective when it is desired to move faster to the four-wheel drive side on a low friction coefficient road. Become.

In addition, when a large lateral acceleration Yg is generated during a turn, the proportional coefficient Kx changes from a state of a smaller proportional coefficient Kx to a large proportional coefficient Kx with a decrease in the lateral acceleration Yg toward the rise in the latter half of the turning. Since Kx changes slowly in the increasing direction, the proportional coefficient Kx increases in response with a slight delay in the decrease of the lateral acceleration Yg, that is, the driving force distribution gradually shifts from the two-wheel drive side to the four-wheel drive side. As a result, the understeer tendency in the latter half of turning is suppressed.

(B) High lateral acceleration running When the vehicle is turning on a high friction coefficient road with a large lateral acceleration Yg, a control characteristic with a small increase rate value (proportional coefficient Kx) of the clutch engaging force T is obtained, so the clutch engaging force T is small, The driving force tends to be distributed mainly on the two-wheel drive side, and good turning performance on a high friction coefficient road can be realized without early drift-out.

At this time, since the proportional coefficient Kx is small, the control system is originally stable, and the need for the differential element is small. In contrast,
Since the proportional coefficient Kx is small, the clutch engaging force due to the front-rear wheel rotational acceleration difference Δ is also small, and the influence of the differential element is suppressed to a small level.

Also, when a large proportional coefficient Kx is selected when entering a corner from a straight road on a high friction coefficient road, the lateral acceleration Yg is generated because the proportional coefficient Kx changes rapidly in the decreasing direction by low-pass filtering. The proportional coefficient Kx decreases while responding quickly to, that is, the driving force distribution direction on the two-wheel drive side shifts early, the understeer tendency at the beginning of turning is suppressed, and when entering a corner in the coast state. As described above, since the control characteristic having a small gain with respect to the front-rear wheel rotation speed difference ΔN is shifted to early, the occurrence of hunting in the power train system is suppressed even when the tire grip is high.

As described above, in the driving force distribution control device D of the first embodiment, straight driving, starting,
The increase rate of the clutch engaging force T can be increased for low friction coefficient road turning, and the increase rate of the clutch engaging force T can be decreased for high friction coefficient road turning with a large lateral acceleration Yg. It is possible to achieve both good running performance while suppressing oscillations in straight running, starting, turning with a low friction coefficient road, and improvement of turning performance when turning with a high friction coefficient road.

Next, the second embodiment device shown in FIGS. 9 and 10 will be described.

Regarding the configuration of the device of the second embodiment, as shown in FIG. 9, the lateral acceleration sensor in the first embodiment is used as the detecting means.
While 43 is used, in the second embodiment, the left front wheel speed wf ₁ and the right front wheel speed wf ₂ are detected and the respective wheel speed signals (wf ₁ ), (w
f ₂ ) output front left wheel speed sensor 47, right front wheel speed sensor 4
The difference is that 8 is used.
As shown in FIG. 0, the lateral acceleration Yg is determined by using the calculated value of the turning radius R based on the left and right front wheel speeds wf ₁ and wf ₂ , and the changing speed of the calculated value of the turning radius R is restricted. Therefore, the rate of change of the proportional coefficient K, which is the increase rate of the clutch engaging force T, is indirectly regulated.

This flow of the driving force distribution control operation in the second embodiment is
This will be described with reference to the flowchart shown in FIG.

At step a, the left front wheel speed wf ₁ ,
The right front wheel speed wf ₂ and the rear wheel rotation speed Nr are read.

In step b, each data used in the subsequent processing is calculated.

The vehicle speed Vf = r / 2 · (wf ₁ + wf ₂ ) is calculated from the left front wheel speed wf ₁ , the right front wheel speed wf _2, and the tire diameter r read in step a, and the rear wheel rotation speed Nr and the front wheel speed are calculated. Rotation speed Nf ＝
The front and rear wheel rotation speed left ΔN = Nr−Nf is calculated from (wf ₁ + wf ₂ ) / 2, and the left and right wheel rotation speed difference Δn (= | wf ₁ from the left front wheel speed wf ₁ and the right front wheel speed wf ₂ −wf ₂ |) is calculated, and the front-rear wheel rotation speed difference ΔN obtained in the current control cycle and the front-rear wheel rotation speed difference Δ obtained in the previous control cycle are calculated.
The difference in rotational acceleration between the front and rear wheels Δ = ΔN−ΔN ₀ is calculated from N ₀ .

In step c, the turning radius R is calculated based on the data obtained in step b.

 The formula for calculating the turning radius R is as follows.

R = f (Δn, Vf) ≈K · Vf / Δn In the following steps d to l, a low-pass filter is realized which restricts the increasing rate of the turning radius R and the changing speed in the increasing direction.

In step d, the amount of change ΔR per unit time is calculated by the difference between the turning radius R obtained in step c and the turning radius R ₀ one cycle before.

In step e, it is determined whether ΔR is positive or negative, and whether the turning radius R is increasing or decreasing is determined, and the subsequent processing route is changed.

If ΔR is positive in step e, that is, if the turning radius R is in the increasing direction, it is determined in step g whether or not the change width is larger than the set value A ₄ , and the set value A ₄ is the turning radius. It becomes the value of the low-pass filter when R increases.

When ΔR is larger than A ₄ in step g, the process proceeds to step j and filtering is performed, and the turning radius Rx is
It is calculated by the arithmetic expression of R ₀ + A ₄ .

When ΔR is smaller than A ₄ in step g, the process proceeds to step k, and the calculated turning radius R is set as it is as the turning radius Rx.

On the other hand, if ΔR is negative in step e, that is, if the turning radius R is in the decreasing direction, it is determined in step f whether the change width is larger than the set value A ₅ , and this set value A ₅ is turned. It is the value of the low-pass filter when the radius R decreases.
The set value A ₅ is larger than the set value A ₄ .

Then, if | ΔR | is larger than A ₅ in step f,
The process proceeds to step i and is filtered, and the turning radius Rx is
It is calculated by the formula of R ₀ −A ₅ .

When | ΔR | is smaller than A ₅ in step f, the process proceeds to step h, and the calculated turning radius R is set as it is as the turning radius Rx.

In step l, the value of the turning radius Rx obtained in the current control cycle is stored as R ₀ for calculating ΔR.

At step m, the lateral acceleration Yg is calculated by the low-pass filtered turning radius Rx and the vehicle speed Vf.

 The calculation formula is as follows.

Yg = (Vf) ² / Rx In step n, a proportional coefficient (gain) K is calculated by using the lateral acceleration Yg.

 Note that K = A / Yg (A; constant).

In step o, the correction value ΔNx of the front-rear wheel rotational speed difference ΔN obtained in step b is obtained.

When the correction value ΔNx is ΔN <0, it is regarded as a tight corner and ΔNx = 0, and when ΔN ≧ 0, the turning locus is corrected and ΔNx = ΔN−f (Rx, Vf). It is a thing.

At step p, the target clutch engagement force T is obtained by calculation (the following equation).

T = K * (ΔNx + B * Δ) B; constant In step q, a command current signal (i) is output according to a command current value I ^* that gives a clutch pressure P corresponding to the clutch engaging force T found in step p. To be done.

As described above, in this second embodiment, by filtering the value of the turning radius R for obtaining the calculated value of the lateral acceleration Yg, the value of the proportional coefficient K indirectly determined by the lateral acceleration Yg is filtered. Although an example of doing so is shown, this is effective when the lateral acceleration Yg is detected based on the turning radius R and the vehicle speed Vf and the proportional coefficient K is obtained, and the reason will be described below.

In terms of detection accuracy, the vehicle speed detection accuracy is calculated based on a large number of pulse count values in the current mainstream rotation pulse type, so even if there is a slight pulse detection error, it has no significant effect, but the turning radius In the case of detecting R, for example, in the present embodiment, the calculation is performed from the rotation speed difference Δn between the left and right front wheels and the vehicle speed Vf. However, since the detection error of the pulse with respect to the left and right front wheel rotation speed difference Δn is large, the calculated turning radius R It greatly affects the detection accuracy.

Incidentally, when the steering angle is counted by the pulse encoder and the turning radius R is detected, the pulse detection error also greatly affects.

Therefore, it is necessary to apply a filter when detecting the turning radius R, and the lateral acceleration Yg is calculated based on the filtered turning radius R and the vehicle speed Vf detected with relatively high accuracy.
By calculating the proportional coefficient K with this value, the proportional coefficient K is indirectly filtered.

In this case, if the turning radius R can be filtered and calculated relatively accurately, the turning locus difference of the front and rear wheel rotation speed difference ΔN can be similarly corrected with high accuracy.

The embodiment of the present invention has been described in detail above with reference to the drawings.
The specific configuration is not limited to this embodiment, and even if there are design changes and the like within the scope of the present invention, they are included in the present invention.

For example, in the embodiment, an example in which an electromagnetic proportional relief valve is used as the hydraulic control actuator is shown, but other means, for example, a solenoid opening / closing valve structure when a duty control signal is used, may be used.

Further, in the embodiment, the multi-plate friction clutch by hydraulic engagement is shown as the clutch means, but other clutches such as an electromagnetic clutch and a viscous clutch may be used.

(Effects of the Invention) As described above, in the four-wheel drive vehicle driving force distribution control device of the present invention, the clutch control means is configured to change the clutch according to the front and rear wheel rotational speed difference and the front and rear wheel rotational acceleration difference. Since the fastening force is increased and the characteristic that the rate of increase is decreased as the lateral acceleration is increased,
The increase rate of the clutch engagement force can be increased for straight running, starting with a small lateral acceleration, and low-friction coefficient road turning, and for the high friction coefficient road turning with a large lateral acceleration, the clutch engagement force increase rate value can be increased. It is possible to reduce the torque consumption, and it is possible to achieve both good running performance while suppressing oscillations in straight running, starting, low friction coefficient road turning, etc. and improvement of turning performance when turning with a high friction coefficient road. To be

[Brief description of drawings]

FIG. 1 is a conceptual view of claims showing a drive force distribution control device for a four-wheel drive vehicle according to the present invention, and FIG. 2 is a view showing a four-wheel drive vehicle to which a drive system clutch control device according to an embodiment is applied. Is a cross-sectional view showing a transfer device of the embodiment device, FIG. 4 is a block diagram showing a control unit of the embodiment device, FIG. 5 is a characteristic diagram of a relation between clutch hydraulic pressure and clutch engaging force, and FIG. 6 is a command current value. And FIG. 7 is a characteristic diagram of the relationship between the clutch pressure and the clutch pressure. FIG. 7 is a characteristic diagram for controlling clutch engagement force with respect to front-rear wheel rotation speed difference according to the first embodiment device. FIG. FIG. 9 is a flow chart showing the flow of operation, FIG. 9 is a block diagram showing the control unit of the second embodiment device, and FIG. 10 is a drive system clutch control operation in the control unit of the second embodiment device. Is a flowchart showing the flow. 1 ... Drive system clutch means 2 ... Detection means 201 ... Front / rear wheel rotational speed difference detection means 202 ... Front / rear wheel rotational acceleration difference detection means 203 ... Lateral acceleration detection means 3 ... Clutch control means

Claims (1)

[Claims] 1. A drive system clutch means, which is provided in the middle of an engine drive system for distributing and transmitting the engine drive force to the front and rear wheels, and is capable of changing the transmission torque by an external clutch engagement force, and detection from a predetermined detection means. In a driving force distribution control device for a four-wheel drive vehicle, which comprises a clutch control means for outputting a control signal for increasing or decreasing the clutch engagement force based on a signal, the front and rear wheel rotational speed difference detecting means as the detecting means, The clutch control means includes front and rear wheel rotational acceleration difference detecting means for detecting a temporal change rate of the front and rear wheel rotational speed difference, and lateral acceleration detecting means for detecting a lateral acceleration. A four-wheel drive vehicle characterized in that the clutch engagement force is increased in accordance with the difference, and the increase rate is decreased as the lateral acceleration increases. Drive force distribution control device.