JP2903171B2 - 4-wheel drive vehicle with driving force distribution control - Google Patents

Description

Description: BACKGROUND OF THE INVENTION The present invention relates to a four-wheel drive vehicle, and more particularly to a four-wheel drive vehicle that controls driving force distribution so that driving force distribution can be performed stably.

[Conventional technology]

Conventionally, there has been proposed a four-wheel drive vehicle in which the distribution of the transmission torque between the front wheels and the rear wheels is controlled in accordance with the driving state. As a driving force distribution means for such a four-wheel drive vehicle, for example, a differential limiting device such as a VCU (Viscous Coupling Unit) is attached to the center differential to appropriately regulate the difference in the number of revolutions of the center differential. Devices such as those described above and devices capable of adjusting the power transmission state according to control hydraulic pressure by a hydraulic multi-plate clutch or the like have been developed.

[Problems to be solved by the invention]

By the way, in the four-wheel drive vehicle as described above, there is a demand that the control of the distribution of the driving force should be performed stably with an allowance against disturbance in general driving.

On the other hand, Japanese Unexamined Patent Publication No. Sho 63-80226 discloses a method for controlling front and rear wheel driving force distribution of a four-wheel drive vehicle by controlling a differential limiting force so that an actual rotational speed difference between front and rear wheels approaches a target rotational speed difference. Performing techniques are disclosed.

However, Japanese Patent Application Laid-Open No. 63-8026 discloses that the differential limiting force is increased only when the actual rotational speed difference exceeds the target rotational speed difference, and when the actual rotational speed difference is less than the target rotational speed difference, the differential limiting force is increased. Only a technique for reducing the limiting force is disclosed, and in such a technique, the differential limitation is performed even when the actual rotation speed difference is between 0 and the target rotation speed difference.
Although the tight corner braking phenomenon is avoided to some extent, there is a risk that the behavior of the vehicle may be directed to an unstable direction, and the driving force distribution control cannot be performed stably.

The present invention has been made in view of the above-described problems, and in this type of power transmission device, it is possible to perform stable control that is hardly affected by disturbance and to stabilize the behavior of a vehicle. It is an object of the present invention to provide a driving force distribution control type four-wheel drive vehicle.

[Means for solving the problem]

For this reason, the driving force distribution control expression 4 according to the first aspect of the present invention.
The wheel drive vehicle includes differential limiting means capable of selectively limiting the differential between the front wheel drive shaft and the rear wheel drive shaft, and control means for controlling the differential limiting means, wherein the control means A front wheel side rotation speed detecting unit for detecting a rotation speed of the front wheel side drive shaft; a rear wheel side rotation speed detection unit for detecting a rotation speed of the rear wheel side drive shaft; a front wheel side rotation speed detection unit; A rotational speed difference calculating unit that calculates a rotational speed difference between the front wheel side drive shaft and the rear wheel side drive shaft based on each rotational speed detected by the rear wheel side rotational speed detecting unit, and a rotational speed difference generated by turning the vehicle. Target rotation speed difference calculating means for calculating a target rotation speed difference between the front wheel side drive shaft and the rear wheel side drive shaft based on a steering wheel angle, a vehicle body speed, and a vehicle body acceleration of the vehicle; Front-wheel-side drive shaft calculated by means Rotational speed difference between the rear wheel side drive shaft, the drive force distribution control type for controlling so as to approach the target rotational speed difference calculated by the target rotational speed difference calculation means 4
In the wheel drive vehicle, the control means sets a differential limiting force based on a difference between a rotation speed difference between the front wheel drive shaft and the rear wheel drive shaft and the target rotation speed difference, and sets the front wheel side When the rotational speed difference between the drive shaft and the rear wheel side drive shaft is larger than 0 and smaller than the target rotational speed difference, the differential limiting force is not applied, and the front wheel side drive shaft and the rear wheel side When the rotational speed difference from the drive shaft is equal to or less than 0 and when the rotational speed difference is equal to or greater than the target rotational speed difference, the differential limiting force is set to act.

The second driving force distribution control four-wheel drive vehicle in accordance with claim of the present invention, in one of the first aspect, a final reduction of the front and rear wheels of the final reduction ratio [rho _f and the front wheel of the dynamic load radius r _f between the ratio [rho _r and the rear wheel of the dynamic load radius r _r and the transfer ratio [rho _t, is set to satisfy the relationship of _{(ρ f / r f) <} (ρ r · ρ t / r r) It is characterized by having.

(Operation)

In the driving force distribution control type four-wheel drive vehicle according to the first aspect of the present invention, in the control means, each rotation detected by the front wheel side rotation speed detection means and the rear wheel side rotation speed detection means by the rotation speed difference calculation means. The rotation speed difference between the front wheel drive shaft and the rear wheel drive shaft is calculated based on the speed, and the rotation speed difference between the front wheel drive shaft and the rear wheel drive shaft calculated by the rotation speed difference calculation means is equal to the target. The control is performed so that the rotation speed difference approaches the target rotation speed difference calculated based on the steering wheel angle, the vehicle speed, and the vehicle acceleration of the vehicle.

At this time, the control unit sets the differential limiting force based on the difference between the rotational speed difference between the front wheel side drive shaft and the rear wheel side drive shaft and the target rotational speed difference. As a result, the driving force is distributed in accordance with the difference between the rotational speed difference between the front wheel drive shaft and the rear wheel drive shaft and the target rotational speed difference.

Further, when the rotational speed difference between the front wheel drive shaft and the rear wheel drive shaft is larger than 0 and smaller than the target rotational speed difference, the differential limiting force does not act. Conversely, the differential limiting force acts when the rotation speed difference between the front wheel drive shaft and the rear wheel drive shaft is 0 or less and when the rotation speed difference is equal to or more than the target rotation speed difference.

Also, in the driving force distribution control type four-wheel drive vehicle according to the second aspect of the present invention, the front wheel final reduction ratio ρ _f and the front wheel dynamic load radius
between the r dynamic load of the final reduction ratio [rho _r and the rear wheels of _f and the rear wheel radius r _r and the transfer ratio _{ρ t, (ρ f / r} f) <(ρ r · ρ t / r r) relationship Is established, a differential can be generated between the front wheels and the rear wheels even during normal running, and front and rear wheel driving force distribution control by differential limitation can be realized.

〔Example〕

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS A four-wheel drive vehicle with a driving force distribution control system according to one embodiment of the present invention will be described below in detail with reference to the drawings. FIG. 1 is a schematic overall configuration diagram, and FIG. 3 to 6 are graphs showing the characteristics of the driving force distribution control, FIGS. 7 and 8 are diagrams each showing the control oil pressure of the driving force distribution control, and FIG. 9 is the driving force distribution control of the present invention. It is a schematic whole block diagram which shows the modification of a formula four-wheel drive motor vehicle.

In FIG. 1, reference numeral 2 denotes an engine, and the output of the engine 2 is transmitted to an output shaft 8 via a torque converter 4 and an automatic transmission 6. The output of the output shaft 8 is transmitted through an intermediate gear 10 to a planetary gear differential (center differential) 12
It is transmitted to.

On the other hand, the output of the planetary gear type differential device 12 is transmitted through a reduction gear mechanism 19 and a front wheel differential gear device 14 to an axle 17.
L, 17R are transmitted to the left and right front wheels 16, 18, while the axles 25L, 25R are transmitted via a bevel gear mechanism 15, a propeller shaft 20, a bevel gear mechanism 21, and a differential gear unit 22 for rear wheels.
From the left and right rear wheels 24, 26. The planetary gear type differential device 12 includes a sun gear 12a, a planetary gear 12b disposed outside the sun gear 12a, and a ring gear 12c disposed outside the planetary gear 12b, similarly to a conventionally known one. In addition, the output of the output shaft 8 of the automatic transmission 6 is input to the carrier 12d that supports the planetary gear 12b, and the sun gear 12a
Is interlocked with the front-wheel differential gear device 14, and the ring gear 12c is interlocked with the propeller shaft 20.

Further, a hydraulic multi-plate clutch 28 is interposed between the ring gear 12c and the carrier 12d as differential control means whose frictional force changes according to the pressure applied to its own hydraulic chamber.

Therefore, the planetary gear type differential device 12 can control the torque transmitted to the front wheel side and the rear wheel side by appropriately controlling the hydraulic multi-plate clutch 28 from a completely free state to a locked state.

Further, the rotation angle from the code 30 is the neutral position of the steering wheel 32, i.e. a steering sensor for detecting a steering angle theta _S,
Lateral acceleration sensor 34 for detecting acceleration G _Y of the lateral direction acting on the vehicle body, 36 is a longitudinal acceleration sensor for detecting acceleration G _X in the longitudinal direction acting on the vehicle body, 38 detects the throttle opening theta _T of the engine 2 Throttle sensor, 39 is engine 2
The engine key switches 40, 42, 44, and 46 are wheel speed sensors for detecting the rotational speeds of the left front wheel 16, the right front wheel 18, the left rear wheel 26, and the right rear wheel 28, respectively. The output is input to the controller 48.

Reference numeral 50 denotes an anti-lock brake device. When the anti-lock brake device 50 outputs an operation signal of an anti-lock brake, a signal indicating the state is input to the controller 48. Reference numeral 52 denotes a warning light that is turned on based on a control signal from the controller 48.

Reference numeral 54 denotes a hydraulic power source, and 56 denotes a hydraulic power source 54 and a hydraulic multi-plate clutch.
The pressure control valve is interposed between the pressure control valve and the hydraulic chamber 28. The pressure control valve 56 is controlled by a control signal from the controller.

Further, reference numeral 60 front wheel speed sensor (front-wheel-side rotational speed detection means) for detecting a rotational speed (rotational speed) N _F of the front wheel side output shaft of the center differential 12, 62 wheel output shaft after the center differential 12 Is a rear wheel side rotation sensor (rear wheel side rotation speed detecting means) for detecting the rotation speed (rotation speed) N _R of the motor.

By the way, the final reduction ratio ρ _{f of} the front wheel, the dynamic load radius r _{f of} the front wheel,
Rear wheel final reduction gear ratio [rho _r of between dynamic load radius r _r and the transfer ratio [rho _t of the rear wheels, _{generally, (ρ f / r f)} = (ρ r · ρ t / r r) relationship Is established in such a manner that, in all driving states, the relationship of (ρ _f / r _f ) <(ρ _r · ρ _t / r _r ) (1) It is set to hold.

Note that the final reduction ratio ρ _f of the front wheels is related to, for example, the reduction gear mechanism 19, the final reduction ratio ρ _f of the rear wheels is related to the bevel gear mechanism 21, for example, and the transfer ratio ρ _t is, for example, the bevel gear mechanism 15.
It is a value about.

In general, because equal and the dynamic load radius r _r of the rear wheels and the front wheels of the dynamic load radius r _f, equation (1) becomes _{ρ f <ρ r · ρ t} ...... (1) '.

With such a setting, if both the front wheel and the rear wheel are not slipping, the rotation speed on the rear wheel side of the clutch disc of the hydraulic multi-plate clutch 28 is faster than the rotation speed on the front wheel side. Therefore, when the pressure in the hydraulic chamber of the hydraulic multi-plate clutch 28 is increased to bring the clutch into a connected state,
Normally, torque is transmitted from the rear wheel side to the front wheel side based on the rotation difference between the front wheel side clutch disc and the rear wheel side clutch disc. In this case, the transmission torque capacity has a magnitude corresponding to the decrease in the rotation difference between the two clutch disks, that is, a magnitude corresponding to the magnitude of the rotation difference originally generated and the contact pressure between the two clutch disks.

FIG. 3 is a characteristic diagram showing the relationship between the front-rear driving force distribution ratio and the control hydraulic pressure when the clutch 28 transmits torque from the rear wheels to the front wheels, and FIG. FIG. 6 is a characteristic diagram showing a relationship between engine torque and clutch transmission torque when torque is transmitted to the motor. FIG. 5 is a characteristic diagram showing, similarly to FIG. 3, the relationship between the front-rear driving force distribution ratio and the control hydraulic pressure when the clutch 28 transmits torque from the front wheels to the rear wheels, and FIG. FIG. 5 is a characteristic diagram showing the relationship between the engine torque and the clutch transmission torque when torque is transmitted from the front wheels to the rear wheels as in FIG.

When transmitting torque from the rear wheels to the front wheels, as shown in FIG.
When the control hydraulic pressure is zero and completely free (that is, the differential limiting force is 0), the ratio is about 32:68 (this depends on the load balance between the front wheel system and the rear wheel system, etc. It can be seen that the ratio of the driving force distribution toward the front wheels increases as the control oil pressure increases (that is, the differential limiting force is applied). In particular, when the engine torque _TE is small, the ratio of the driving force distribution to the front wheels is greatly increased only by slightly increasing the control oil pressure,
As the engine torque T _E increases, greater control oil pressure to increase the driving force distribution to the front wheels becomes necessary, also, it is also difficult to significantly increase the proportion of the driving force distribution to the front wheel side, the control oil pressure Since the driving force distribution is adjusted little by little with respect to the change, it is easy to reliably adjust to the target driving force distribution.

Also, as shown in FIGS. 4 and 6, it can be seen that a clutch transmission torque proportional to the engine torque is required to obtain a constant distribution of the driving force.

On the other hand, when torque is transmitted from the front wheels to the rear wheels, as shown in FIG. 5, the variable range of the driving force distribution is narrow, and the distribution closer to the front becomes impossible, which is not preferable.

In order to transmit torque from the rear wheels to the front wheels,
The inequality (1) is set to hold.

With such a setting, the torque to be distributed to the front wheels can be significantly larger than that to the rear wheels, and the torque distribution ratio at which the torque to the front wheels is maximized is (ρ _f / r _f ) And the value of (ρ _r · ρ _t / r _r ) can greatly increase the torque distribution to the front wheels. For example, the torque distribution ratio to the front and rear wheels, that is, the front wheels: rear wheels Can be set to 100: 0 or more (e.g., 120: -20). Therefore, for example, the torque distribution ratio (front wheel: rear wheel) can be adjusted to an extremely wide range from 33:67 to 100: 0.

The controller 48 as a control means includes a CPU, a ROM, a RAM, an interface and the like (not shown) necessary for the control described later, and includes a target rotation speed difference calculation section (target rotation speed difference calculation means) 48a, It has a rotation speed difference calculation section (rotation speed difference calculation means) 48b and a torque control section (torque control means) 48c for controlling the hydraulic multi-plate clutch 28 as a transmission torque adjustment mechanism.

In rotational speed difference calculating section 48b, .DELTA.N from the rotational speed N _R of the rotational speed N _F and the rear wheel output shaft of the front wheel side output shaft of the center differential 12 detected by the rotational speed sensor 60, 62 (= N _R - N _F) is calculated.

The target rotational speed difference calculating section 48a calculates the center differential 1
Set the target value (target rotation speed difference) ΔN ₀ of the rotation speed difference (rotation speed difference) between the front wheel side output shaft and the rear wheel side output shaft of 2,
This target rotational speed difference ΔN ₀ is calculated by calculating the steering wheel angle (steering angle) θ _H ,
It can be determined according to the running state of the vehicle such as the vehicle speed V and the vehicle acceleration V '.

That is, the target rotation speed difference ΔN ₀ is calculated based on these angle values θ _H , V,
A function of V ′ ΔN ₀ = f (θ _H , V, V ′) (2)

The reason for setting the target rotation speed difference ΔN ₀ in this manner is as follows.

As described above, in the drive system of this automobile, when the wheels do not slip, a rotational speed difference is generated between the front wheel output shaft and the rear wheel output shaft of the center differential 12. When the clutch 28 is connected, torque is transmitted from the rear wheel side to the front wheel side through the clutch 28, and the torque distribution to the front and rear wheels is adjusted. Depending on the capacity of the transmitted torque through the clutch 28 at this time, The rotation speed difference described above also changes. Therefore, in order to set the torque distribution to the front wheel and the rear wheel to be optimal for the traveling state, paying attention to the rotational speed difference between the front wheel output shaft and the rear wheel output shaft of the center differential 12, It is conceivable to control the rotational speed difference to take a value corresponding to the optimal driving force distribution state.

The optimal driving force distribution state can be determined based on the running state of the vehicle such as the steering wheel angle θ _H , the vehicle speed V, and the vehicle acceleration V ′. In other words, when the vehicle is steered, in general, the oversteer tendency increases when the rear wheel drive power distribution is large, and the unsteering tendency increases when the front wheel drive power distribution is large. Affects the steer characteristics. Therefore, the front and rear wheel driving force distribution state such that the steering characteristic is desirable for the vehicle,
It can be said that "optimum drive power distribution to front and rear wheels". Since this desirable steering characteristic generally changes depending on the vehicle speed V and the vehicle acceleration V ′, when focusing on the steering characteristic, the optimal driving force distribution state is determined by the steering wheel angle θ _H , the vehicle speed V, and the vehicle acceleration V ′. It can be determined according to the running state of the vehicle. In addition, not only when turning, but also when the vehicle is accelerating, the weight sharing of the rear wheels increases, so that if the rear wheel driving force distribution is large, the driving force can be efficiently exhibited, and conversely, the weight sharing of the front wheels when the vehicle decelerates. Since it increases, the braking force can be efficiently exhibited if the front wheel driving force distribution is large. Therefore, from such a viewpoint, "optimum drive force distribution to front and rear wheels" can be determined according to vehicle body acceleration V '.

That is, as described above, the target rotational speed difference ΔN ₀ can be represented as a function of each of these values θ _H , V, V ′ [see equation (2)]. However, since the characteristics described above differ depending on the vehicle type, the function, that is, the correspondence of the target rotation speed difference ΔN _{0 to} each value θ _H , V, V ′ is set based on experimental data for each vehicle type. Will be.

From another point of view, when torque is transmitted through the clutch 28, slip occurs on the front wheels and the rear wheels. When the slip of the wheels is in an optimum state, that is, when the wheels The theoretical ΔN on the center differential that occurs at the rate is determined by the running state of the vehicle. It can be said that such a theoretical ΔN is set as the rotational speed difference ΔN ₀ .

Torque control section 4c based on the information from the target rotational speed difference calculating section 48a and the rotational speed difference calculating unit 48b, the actual rotational speed difference .DELTA.N is as transmitting torque adjusting mechanism such that it converges to the target rotational speed difference .DELTA.N ₀ The hydraulic multi-plate clutch 28 is controlled.

Specifically, as shown in FIG.
The control hydraulic pressure P (which corresponds to the differential limiting force) is set according to the actual rotational speed difference ΔN. That is, when the actual rotation speed difference ΔN becomes larger than the target rotation speed difference ΔN ₀ (> 0), the control oil pressure P is changed to the actual rotation speed difference ΔN and the target rotation speed difference ΔN.
As proportional to the difference D from N ₀ (= ΔN−ΔN ₀ ), (P =
αD), and the actual rotation speed difference ΔN is equal to the target rotation speed difference Δ
If it is smaller than N ₀ and equal to or greater than 0 (range indicated by K ₁ ),
The control oil pressure P is set to 0, and if the actual rotation speed difference ΔN is equal to or less than 0, the control oil pressure P is changed according to the magnitude of ΔN (P = −β
ΔN). The coefficients α and β are positive values.

If the target rotation speed difference ΔN ₀ is negative (ΔN ₀ <0), the control oil pressure P may be set as shown in FIG. 7, but in this case, the reaction to the rear wheel slip is deteriorated. However, it is not preferable for handling stability.

The reason why the differential limitation is not performed with the oil pressure set to 0 in the range indicated by K ₁ and K ₂ is that if the differential limitation is performed in this case, the torque is further distributed to the slip wheels, which is not preferable. .

Thus, in general, the target rotational speed difference .DELTA.N ₀ exactly, that is, set to a state a larger number of revolutions N _R of the rear wheel output shaft than the rotational speed N _F of the front wheel side output shaft of the center differential 12, As shown in FIG. 8, the hydraulic pressure P is controlled.

The torque control section 48c, while adjusting the way hydraulic P, the actual rotational speed difference .DELTA.N controls the connection state of the hydraulic multi-plate clutch 28 so as to converge to the target rotational speed difference .DELTA.N _0.

Next, such control contents will be described with reference to the flowchart shown in FIG.

Each control element is initialized (step S1), and if the engine is not turned off, the steps from step S2 are performed.
Proceed to S3, read steering wheel angle theta _H, the vehicle speed V, vehicle acceleration V ', the shaft rotational speed N _F, signals such as N _R (detection data).

Then, the rotation speed difference detection unit 48b calculates the actual differential rotation speed (actual rotation speed difference) ΔN from the shaft rotation speeds N _F and N _R (step
S4) The target rotational speed difference calculator 48a calculates a target differential rotational speed (target rotational speed difference) ΔN ₀ from the steering wheel angle θ _H , the vehicle speed V, and the vehicle acceleration V ′ (step S5).

Further, in step S6, it is determined whether the actual or rotational speed difference .DELTA.N is 0 or not, unless .DELTA.N is 0 or less, at step S7, whether the actual rotational speed difference .DELTA.N is the target rotational speed difference .DELTA.N ₀ or more (That is, ΔN−ΔN ₀ > 0). If ΔN <0 in these determinations in steps S6 and S7, the process proceeds to step S9, where the control hydraulic pressure P of the hydraulic multiple disc clutch 28 is set to P = −βΔN If 0 <ΔN <ΔN ₀ , the process proceeds to step S8, and the control oil pressure P of the hydraulic multiple disc clutch 28 is set to P =
If ΔN ₀ <ΔN, the process proceeds to step S10, where the control oil pressure P of the hydraulic multiple disc clutch 28 is set to P = α (ΔN−ΔN ₀ ).
Set as

Then, the pressure control valve is set so that these oil pressures are set.
The oil pressure is controlled through 56 etc.

As a result, feedback control is performed so that the actual rotational speed difference ΔN converges to the target rotational speed difference ΔN ₀ , the torque distribution to the front and rear wheels is always appropriately adjusted, and each wheel always maintains the state of the optimal slip ratio. While traveling. As a result, the driving force distribution control is not performed in the unstable direction, and the driving force distribution control can be stably performed.

Further, when controlling the differential limiting force so that the actual rotational speed difference between the front and rear wheels approaches the target rotational speed difference generated along with the turning, the actual rotational speed difference may be from 0 to the target rotational speed difference. By avoiding the differential limiting force in the middle, it is possible to avoid the tight corner braking phenomenon, as well as to suppress the differential limiting that may lead the vehicle behavior to an unstable direction. And the driving force distribution control can be stably performed.

Conversely, when the actual rotational speed difference between the front and rear wheels is equal to or greater than the target rotational speed difference, and when the actual rotational speed difference is equal to or less than 0, the differential limiting force is applied. By applying the differential limiting force, it is possible to bring the actual rotational speed difference between the front and rear wheels closer to the target rotational speed difference, and to guide the behavior of the vehicle in a stable direction, including avoiding the die corner braking phenomenon. Becomes possible.

In the above-described embodiment, the case where the hydraulic differential multi-plate clutch 28 for limiting the differential is combined with the center differential 12 has been described. However, the configuration of the present invention can also be applied to a transfer clutch as shown in FIG. .

In FIG. 9, reference numeral 11 denotes a front wheel clutch for driving the front wheels, and 13 denotes a rear wheel clutch for driving the rear wheels. The control is performed in substantially the same manner as the clutch 28. The other parts are configured in substantially the same manner as in the embodiment (see FIG. 1), and the description is omitted.

That is, in this case, the target value of the difference between the rotation speed of the front wheel output shaft and the rotation speed of the rear wheel output shaft in the transfer portion is set so that the actual rotation speed difference converges to the target rotation speed difference. The control hydraulic pressure of both clutches 11 and 13 is set to control the coupling state of both clutches 11 and 13.

In each of the above-described examples, the control is performed based on the rotation difference in the center differential or the transfer unit. However, since the wheels also rotate in response to the rotation of these parts, the hydraulic pressure is controlled based on the rotation speed of the wheels. It may be controlled. That is,
The actual rotational speed difference between the front and rear wheels is calculated from the detected values of the wheel speed sensors 40 to 46, while the target rotational speed difference between the front and rear wheels is set, and the actual rotational speed difference converges to this target rotational speed difference. That is, the control hydraulic pressure is set to perform the control. This also
The same operation and effect as described above can be obtained.

〔The invention's effect〕

According to the driving force distribution control type four-wheel drive vehicle of the first and second aspects of the present invention, the driving force distribution control is not performed in an unstable direction, and the driving force distribution control is stabilized. In particular, when the rotation speed difference between the front wheel side drive shaft and the rear wheel side drive shaft is larger than 0 and smaller than the target rotation speed difference, the differential limiting force is not applied.
The behavior of the vehicle can be stabilized by suppressing the differential limitation that may lead the behavior of the vehicle to an unstable direction, and the rotational speed of the front wheel drive shaft and the rear wheel drive shaft can be further reduced. When the difference is equal to or less than 0 and when the difference is equal to or greater than the target rotational speed difference, the differential limiting force is applied. By approaching the difference, the behavior of the vehicle can be guided in a stable direction, and there is an advantage that the performance as a four-wheel drive vehicle can be greatly improved.

[Brief description of the drawings]

FIGS. 1 to 9 show a driving force distribution control type four-wheel drive vehicle as an embodiment of the present invention. FIG. 1 is a schematic overall configuration diagram, and FIG. 2 relates to the driving force distribution control. 3 and 6 are graphs showing the characteristics of the driving force distribution control, FIGS. 7 and 8 are diagrams each showing the control oil pressure of the driving force distribution control, and FIG. 9 is the driving force distribution control of the present invention. It is a schematic whole block diagram which shows the modification of a control type four-wheel drive motor vehicle. 2 ... Engine 4 ... Torque converter 6 ... Automatic transmission 8 ... Output shaft 10 ... Intermediate gear 12 ... Planetary gear type differential (center differential) 14 ... For front wheels Differential gearing, 15… Bevel gear mechanism, 16, 18… Front wheel, 17L, 1
7R: axle, 19: reduction gear mechanism, 20: propeller shaft, 21: bevel gear mechanism, 22: differential gear device for rear wheel, 24, 26 ... rear wheel, 25L, 25R: axle , 28 ... hydraulic multi-plate clutch as differential control means, 30 ... steering sensor,
34 ... lateral acceleration sensor, 36 ... longitudinal acceleration sensor, 38 ...
… Throttle sensor, 39 …… Engine key switch, 4
0, 42, 44, 46 ... wheel speed sensor, 48 ... controller as control means, 48a ... target rotation speed difference calculation unit (target rotation speed difference calculation means), 48b ... rotation speed difference calculation unit (rotation Speed difference calculating means), 50 ... anti-lock brake device,
52 warning light, 54 hydraulic pressure source, 56 pressure control valve, 60
... Front wheel side rotation speed sensor (front wheel side rotation speed detecting means), 62
... Rear wheel side rotation speed sensor (rear wheel side rotation speed detecting means).

Claims (2)

(57) [Claims] 1. A vehicle comprising: a differential limiting means for selectively limiting a differential between a front wheel side driving shaft and a rear wheel side driving shaft; and a control means for controlling the differential limiting means. A front wheel side rotation speed detecting means for detecting a rotation speed of the front wheel side drive shaft; a rear wheel side rotation speed detection means for detecting a rotation speed of the rear wheel side drive shaft; a front wheel side rotation speed detection means; Rotation speed difference calculation means for calculating a rotation speed difference between the front wheel side drive shaft and the rear wheel side drive shaft based on each rotation speed detected by the rear wheel side rotation speed detection means; Target rotation speed difference calculation means for calculating a target rotation speed difference between the front wheel side drive shaft and the rear wheel side drive shaft based on a steering wheel angle, a vehicle body speed, and a vehicle body acceleration of the vehicle; Front-wheel-side drive shaft calculated by means A driving force distribution control type four-wheel drive vehicle that controls a rotation speed difference between the motor and the rear wheel side drive shaft to approach a target rotation speed difference calculated by the target rotation speed difference calculation unit. A differential limiting force is set based on a difference between a rotation speed difference between the front wheel drive shaft and the rear wheel drive shaft and the target rotation speed difference, and the front wheel drive shaft and the rear wheel drive shaft are set. When the rotation speed difference between the front wheel side drive shaft and the rear wheel side drive shaft is 0, the differential limiting force is not applied when the rotation speed difference between the front wheel side drive shaft and the rear wheel side drive shaft is smaller than the target rotation speed difference. A driving force distribution control type four-wheel drive vehicle characterized in that the differential limiting force is applied in the following cases and when the difference is equal to or larger than the target rotational speed difference. 2. A front wheel final reduction ratio ρ _f and a front wheel dynamic load radius r _f.
Between the dynamic load radius r _r and the transfer ratio [rho _t of final reduction ratio [rho _r and the rear wheel of the rear wheels, the relationship of _{(ρ f / r f) <} (ρ r · ρ t / r r) Establishment and The driving force distribution control type four-wheel drive vehicle according to claim 1, characterized in that the setting is made to perform the following.