JPH02290731A - Torque distribution control device of four-wheel drive vehicle - Google Patents

Abstract

PURPOSE:To provide a wide range of torque distribution control of front and rear wheels by forming a center differential device into a composite epicyclic gear type, and building in a hydraulic multiple disc clutch which generates differential limit torque when rear wheels slip. CONSTITUTION:An output shaft 15 from an automatic transmission 30 connected to the crankshaft 11 of an engine 10 connected to a center differential device 50 arranged coaxially with it. The rear end of a front drive shaft 16 is connected to the differential device 50 through a pair of reduction gears 17, 18, and the power outputted to a rear drive shaft 20 from the center differential device 50 is transmitted to a rear differential device 22 through a propeller shaft 21. The center differential device 50 is made in a composite epicyclic gear type, where the first sun gear 51 is connected to the output shaft 15 side, a carrier 55 is connected to the front drive shaft 16 side, and the second sun gear 53 is connected to the rear drive shaft 20 side. A hydraulic multiple disc clutch 60 which permits torque movement and differential lock is provided between the carrier 55 and the second sun gear 53.

Description

[Detailed Description of the Invention] [Field of Industrial Application] The present invention relates to a torque distribution control device for a four-wheel drive vehicle equipped with a composite planetary gear type center differential device. The center differential device of the drive system is a vertical type in which the machine output shaft is arranged in the direction of travel of the vehicle, and the transmission output shaft and the rear drive shaft are arranged coaxially, and the hydraulic multi-plate for limiting the differential. This relates to clutch torque distribution control.

[Conventional technology]

In general, the reference torque distribution between the front and rear wheels of a four-wheel drive vehicle is set proportionally based on dynamic weight distribution that takes into account the movement of the center of gravity during acceleration so that the driving force is maximized. Therefore, on a front engine/front drive (FF) basis, front wheel torque TF and rear wheel torque TIL are TF:TI
Front engine rear drive (FR) base TF: RRT Misaki 4o
: Set to 60.

In addition, when the center differential device is divided equally depending on the state of these standard torque distributions, the bevel gear type is used.
In the case of unequal divisions, a simple braneta-ligia formula or the like is selected.

Here, in a vehicle with center differential i = 14-wheel drive, where the torque distribution is equally divided, the driving performance on rough roads is maximized. However, slipping easily occurs on rough roads such as low μ roads, and when this slipping occurs, adding a differential limiting function to the center differential device will certainly improve driving force, but it will not particularly improve maneuverability. However, since the slip conditions for all four wheels are the same, all four wheels may slip at the same time during high-speed turns, making it difficult to control the vehicle. To ensure steering stability even in such slip conditions,
A simple planetary gear type center differential device is used to set the standard torque distribution for rear wheel bias.

This allows the rear wheels to always slip first, causing the rear wheels to power slide when the driver operates the accelerator, allowing the vehicle to steer while drifting its tail.

Conventionally, regarding a four-wheel drive vehicle equipped with the above-mentioned planetary gear type center differential device, there is a prior art disclosed in, for example, Japanese Patent Laid-Open No. 176728/1983. Here, it has a simple planetary gear center differential device, inputs the gear change output to the carrier, and applies torque from one of the sun gear and ring gear to the front wheels, and from the other to the rear wheels, depending on the difference in pitch circle diameter between the sun gear and ring gear. It will be distributed unevenly according to the amount. Further, it is shown that a differential limiting hydraulic multi-disc clutch is disposed between any two elements of the sun gear, ring gear, and carrier.

[Problem to be solved by the invention]

By the way, in the prior art described above, since it is a simple planetary gear type center differential device, the standard torque distribution ratio is determined only by the ratio of the meshing pitch diameters of the sun gear and the ring gear. There is no freedom.

In addition, in a vertically mounted powertrain where the engine crankshaft and transmission output shaft are arranged in the direction of vehicle travel, if the transmission output shaft and rear drive shaft are applied to a coaxial drive system, the inner side of the sun gear Since the transmission output shaft is inserted through the sun gear, there is a limit to the reduction in the size of the sun gear. On the other hand, increasing the size of the ring gear increases the size of the entire transfer and worsens the comfort of the vehicle interior, which has its limits. Therefore, there is relatively little freedom in setting the drive torque distribution ratio to the rear wheels due to surrounding constraints due to the meshing pitch diameter ratio of the sun gear and ring gear, so it is possible to increase the drive torque distribution to the rear wheels, or Even if you want to set the drive torque distribution to the front and rear wheels to a value close to equal distribution, it is difficult, so it is difficult to set the drive torque distribution to the front and rear wheels to a value close to equal distribution. Because torque distribution is difficult to set,
It is difficult to fully demonstrate its effects.

Furthermore, as mentioned above, it is difficult to set a relatively large standard torque distribution for the rear wheels, so even when torque distribution is controlled using a hydraulic multi-disc clutch, the control range is narrow, making it difficult to properly control running performance and maneuverability. There are inconveniences such as not being able to obtain

The present invention has been made in view of this point, and its purpose is to increase the degree of freedom in determining reference torque distribution by a composite planetary gear type center differential device.

In addition, when applied to a drive system in which the transmission output shaft and the rear drive shaft are coaxial in a vertically mounted lance axle type or FR type, the torque distribution is set to be sufficiently biased to the rear wheels, and the hydraulic multi-disc clutch is used. 4. Possible to control torque distribution over a wide range
Our objective is to provide a torque distribution control device for wheel drive vehicles.

[Failure to solve the problem]

In order to achieve the above object, the torque distribution control device for a four-wheel drive vehicle of the present invention includes a front wheel side transmission element on a transmission output shaft. In a four-wheel drive vehicle with a center differential in which rear wheel side transmission elements are arranged coaxially and a center differential device is arranged between these three, the center differential device is integrated with the first and second sun gears. The first sun gear is connected to the transmission output shaft, the carrier is connected to the front wheel side transmission element, and the second sun gear is connected to the transmission output shaft. is connected to the rear wheel side transmission element, and in the transmission output shaft of the human power element of the center differential device, the carrier of the two output elements, and the second sun gear, between the output elements or between the input element and either one The hydraulic multi-disc clutch that limits the differential movement of the center differential device is interposed between the center differential and the output element, bypassing the hydraulic multi-disc clutch.

[For production]

Based on the above configuration, the center differential device of a four-wheel drive vehicle has a meshing pitch circle radius of four gear elements, a first and second sun gear, and a first and second pinion supported by a carrier and meshing with these sun gears. The torque distribution is determined by this, and it becomes possible to set the reference torque distribution between the front and rear wheels over a relatively wide range with a compact structure.

In addition, the torque distribution between the front and rear wheels of a four-wheel drive vehicle is such that the weight is sufficiently biased to the rear wheels without applying differential limiting torque to the center differential device under normal conditions, allowing for free turning.
It exhibits good maneuverability with a touch of understeer, and runs in four-wheel drive, with the rear wheels always slipping first. When rear wheel slip occurs, a differential limiting torque is generated in the hydraulic multi-disc clutch in response to the slip, and the clutch torque is bypassed and transmitted from the output element to the rear wheels to the output element to the front wheels. Alternatively, by bypassing the input element to the output element side and transmitting differential limiting torque to the front wheels, the weight is shifted to the front wheels, the torque distribution to the front wheels increases, the rear wheel slip disappears, and the driving force is secured. Improves running performance.

〔Example〕

Embodiments of the present invention will be described below based on the drawings.

In FIG. 1, a vertical transaxle type drive system to which the present invention is applied is described. A transmission case 3 is connected to the rear of a torque converter case 1 and a differential case 2, and a transfer case 3 is connected to the rear of a torque converter case 1 and a differential case 2. The case 4 is joined, and the oil van 5 is attached to the lower part of the transmission case 3.

The reference numeral 10 designates an engine, and the crankshaft 11 of this engine 0 is connected to a torque converter 13 equipped with a lock-up clutch 12 inside the torque converter case 1, and the input shaft 4 from the torque converter 13 is input to the automatic transmission 30 of. An output shaft 15 from the automatic transmission 30 is output coaxially with the human power shaft 14, and this output shaft 15 is coaxially connected to a center differential device 50 inside the transfer case 4. Inside the transmission case 3, there are people and output personnel 14. A front drive shaft 16 is arranged parallel to the front drive shaft 15, and the rear end of the front drive shaft 16 is connected to a center differential device 50 with a pair of reduction gears 17. Connected via 18, front drive 1
The front end of 6 is configured to transmit power to the front wheels via a front differential device 9 inside the differential case 2.

On the other hand, an output is output from the center differential device 50 to the rear drive shaft 20, and the rear drive shaft 20 is configured to be transmitted to the rear wheels via the propeller shaft 21, rear differential device 22, etc. The center differential device 50 includes a wet multi-disc hydraulic multi-disc clutch 60.
will be provided.

The automatic transmission 30 has a front planetary gear 31 and a rear planetary gear 32, and for these front planetary gear 31 and rear planetary gear 82, a high clutch 33, a reverse clutch 34, and a brake band 35 are provided.
, a forward clutch 36, an overrunning clutch 37, a low and reverse clutch 38, and a one-way clutch 3940 are provided, and by selectively engaging these, four forward speeds and one reverse speed are obtained. Further, in front of the automatic transmission 30, an oil pump 4■ is provided so as to connect the torque converter's inverter sleeve 13a and the drive shaft 42 so as to constantly drive the same, and the oil pan 5 has a control valve body 43. be accommodated. The control valve body 43 supplies and drains oil to each of the above-mentioned friction elements and selectively engages them.

In FIG. 2(a), the center differential device 50 will be described.

First, the transmission output shaft 15 and the rear drive shaft 20 are coaxially rotatably fitted via the push 23a and the thrust bearing 24, and the reduction gear 17 is also connected to the output shaft l5.
The two are coaxially fitted together via a twenty dollar bearing 23, and a center differential device 50 is coaxially provided between these three. The center differential device 50 is a composite planetary gear type, and the output device 15
It has a first sun gear 51 formed on the rear drive shaft 20, a first pinion 52 that meshes with the sun gear 51, a second sun gear 53 that meshes with the rear drive support 20, and a second pinion 54 that meshes with the sun gear 51. 1st. Second pinion 52. A carrier 5 54 is integrated and is integrally formed with the reduction gal I7.
It is pivotally supported by a bin 56 of No. 5 via a twenty-one dollar bearing 23. And a carrier 5 equipped with a reduction gear 17
The front and rear parts of 5 are supported by ball bearings 25.

In this way, the power of the transmission output shaft 15 is transferred to the first sun gear 5.
1 and the first and second pinions 52. Torque is transmitted to the carrier 55 and the second sun gear 53 via the carrier 54 at a predetermined distribution ratio. When such torque is transmitted, the first and second pinions 52. The rotation and revolution of the carrier 54 absorbs the difference in rotation between the carrier 55 and the second sun gear 53.

Here, the torque distribution of the center differential device 50 will be described in detail using the schematic diagram of FIG.

The input torque of the first sun gear 5l is TI, its engagement pitch radius is "S1, the front side torque of the carrier 55 is TF, the first and second pinion 52. 54's meshing pitch radius is rp1, 'p2, the rear side torque of the second sun gear 53 is Tl{, its meshing pitch radius is rs2
Then, for TI-TF+T (1)rs
1+ rp1− rs2+ rp2 (2) holds true.

Further, the tangential load P acting on the meshing point between the first sun gear 5l and the first pinion 52 is the tangential load P1 acting on the carrier 55 and the meshing point between the second sun gear 53 and the second pinion 54. It is equal to the sum of the tangential load P2 acting on the point.

P − T I/rs1 P1-TF / (rs1 + rp1) P2 =TIL/rs2 T I/rs1= ((TF/ (rs1+rp1)]
+TR/rs2(1). Substituting equation (2) into equation (3) and sorting it out, we get the so-called TF (1-rp111's2/rS1
● 『p2）@TITFL − (rp1Φrs2/rs
1●rp2) ●T! becomes. From this, first,
Second sun gear 51.53 and first and second pinions 52
．． It can be seen that the reference torque distribution of the front side torque TF and the rear side torque Tλ can be freely set by the engagement pitch radius with 54.

Here, rs1 -23.5a+m, rp1 -18
．． 5mm, rp2 -18.8mm, rs2 -
If it is set to 21.2 mm, TF-20/53 ・Ti TR. -33753 ・It becomes TI. Therefore, the torque distribution between the front and rear wheels becomes TF:TR-≠38:82, and the reference torque distribution can be set to sufficiently bias the rear wheels.

Referring to the hydraulic multi-plate clutch 60 in FIG.
7, and the hub 62 is spline-coupled to the rear drive shaft 20 which is integrated with the second sun gear 53. In this way, the hydraulic clutch 60 is connected to the support column 5 of the carrier 55 of the center differential device 50, as shown in FIG. 2(b).
Bypass is provided between 5a and second sun gear 53 to enable torque movement and differential locking.

Then, the piston 64 and the bearing 6 are moved by the hydraulic pressure in the hydraulic chamber 63.
5. The clutch torque is generated by pressing the drive plate 67a, retaining plates 67b, 67c, and plate 67 between the drum 6 and the hub 62 via the pressure plate 66.

Referring to FIG. 4(a), the hydraulic and electrical control systems of the hydraulic multi-disc clutch 60 will be described.

Reference numeral 41 designates an oil bomb driven by the engine, and the discharge pressure of this oil pump 41 is regulated by a regulator valve 80 to generate line pressure, which is used for automatic gear shifting. The line pressure oil passage 81 also communicates with the hydraulic chamber 63 of the hydraulic multi-disc clutch 60 via a clutch control valve 82 and an oil passage 83, and the oil passage 81 is connected to an oil passage 86 having a pilot valve 84 and an orifice 85. It communicates with the control side of the duty solenoid valve 87 and the clutch control valve 82. The duty signal from the control unit 90 is input to the duty solenoid valve 87 to generate duty pressure, and by operating the clutch control valve 82 with this duty pressure, the clutch pressure and clutch torque of the hydraulic multi-disc clutch 60 are variably controlled. It looks like this.

On the other hand, the control unit 90 has at least a front wheel rotation speed sensor 91, a rear wheel rotation speed sensor 92, a steering angle sensor 93, and a shift position sensor 100 as input signals for the control unit 90. As shown in FIG. 2(a), the front wheel rotation speed sensor 9 consists of a pickup 9lb facing a parking gear 91a fixed to the front drive support 16.
2 has a pickup 92b facing the pulse gear 92a of the rear drive shaft 20, and detects the front wheel rotation speed NF and the rear wheel rotation speed Nk. These front wheel rotation speed NF, rear wheel rotation speed NR. is input to the slip rate calculation section 94, but TF
<Since the rear wheels always slip first with the standard torque distribution of Tllj, the slip rate S is S-NF/NFL
(S>0). This slip ratio S and the steering angle ψ of the steering angle sensor 93 are inputted to a clutch pressure setting section 95, and the clutch pressure Pc is searched from a map shown in FIG. 5 in a map setting section 96. Here, in a non-slip state where S≧1, the clutch pressure Pc is set to a low value, and in a slip state where S<1, the clutch pressure Pc is set to a low value as the slip rate S decreases.
c is increased, and when the set value is 51 or less, it is set to P wax.

Furthermore, the clutch pressure Pc is reduced in accordance with the increase in the steering angle ψ to avoid a tight corner braking phenomenon. This clutch pressure Pc is input to a duty conversion section 97 and converted into a duty ratio D corresponding to the clutch pressure Pc, and this duty signal is output.

FIG. 7 shows the flow of processing performed within the control unit 90. As mentioned above, first, the rotation speed of the front and rear wheels is detected, the presence or absence of slip of the front and rear wheels is calculated, and if the slip rate is less than the set value, a preset map based on the throttle opening, vehicle speed, gear position, steering angle, etc. Search for the required differential limiting torque from .

Then, a duty ratio corresponding to the torque transmission capacity of this clutch is selected and outputted to the duty solenoid valve 87. Here, if the slip exceeds a set value, it is determined to be a slip, and a map value at the time of slip occurrence is retrieved and output to the duty solenoid valve 87.

Next, the operation of the four-wheel drive vehicle configured as described above will be described using the characteristic diagram shown in FIG.

First, the power of the engine 10 is input to the automatic transmission 30 via the torque converter 13 and the input shaft 14, and the transmission power automatically shifted by the automatic transmission 30 is transferred to the output shaft 5.
is input to the first sun gear 5l of the center differential device 50. Here, depending on the specifications of each gear element of the center differential device 50, for example, TF:
TR special 38: By setting it to 82, 38% of the torque of the shifting power is integrated with the carrier 55 or is output to the duction gear 17, and 62% of the torque is distributed to the second Zangya 53 and output. Ru.

On the other hand, at this time, the front and rear wheel rotational speeds NF, N, and the steering angle ψ
The signal is input to the control unit 90, and the slip rate calculation section 94 calculates the slip rate S.

Therefore, in a non-slip state with S≧1 on a dry road surface, the clutch pressure setting section 95 sets Pc to a low value, and the duty ratio converting section 97 outputs a signal with a duty ratio of 100%, for example, to the duty solenoid valve 87. For this reason, the duty pressure becomes zero, and the clutch control valve 82 operates to close the oil passage 8l and drain the hydraulic multi-disc clutch 60, thereby releasing the hydraulic multi-disc clutch 60 and reducing the clutch torque. becomes zero, making the center differential device 50 described above free.

Therefore, the 38% torque integrated with the carrier 55 or output to the induction gear 17 is directly transmitted to the front wheels via the reduction gear 18, the front drive φ city 1B, and the front differential device 19. Further, 62% of the torque output to the second sun gear 23 is transmitted to the rear wheels via the rear drive shaft 20, blower shaft 21, and rear differential device 22, resulting in four-wheel drive driving with a biased emphasis on the rear wheels. This torque distribution results in a slight understeer, ensuring good maneuverability. Also, when turning, the center differential device 50 operates the first and second pinion 52.
Rotate 54. It revolves around the planet and completely absorbs the difference in rotational speed, making it possible to turn freely.

Next, when driving on a slippery road, the rear wheels always slip first due to torque distribution biased toward the rear wheels, and the control unit 90
A slip rate calculation unit 94 calculates a slip rate S1 (S<1) according to the slip state. Then, a duty signal of clutch pressure Pc1 corresponding to this slip ratio S1 is output to the duty solenoid valve 87 and operates the clutch control valve 82, so that the hydraulic multi-disc clutch 60 is supplied with hydraulic pressure with the line pressure regulated. Predetermined clutch torque T
produces c. Therefore, the center differential device 5
0, a transmission line passing through the hydraulic multi-disc clutch BO is bypassed between the carriers 55 of the two output elements and the second sun gear 53, and the second sun gear with a large torque distribution is formed. 53, torque is bypassed and transmitted to the carrier 55 by the amount of clutch torque Tc. As a result, the torque distribution is as shown in Figure 6: TI7:Tλ-T
F1: Changes to T1, reduces rear wheel torque and no longer causes slip, increases running performance and maintains good maneuverability.

Then, when the slip rate S mentioned above becomes less than the set value St,
The differential limiting torque becomes maximum together with the oil pressure of the hydraulic multi-disc clutch 60, directly connecting the carrier 55 of the center differential device 50 and the second sun gear 53. Therefore, the center differential device 50 is differentially locked, and the vehicle is driven in direct four-wheel drive with torque distribution corresponding to the weight distribution between the front and rear wheels, maximizing all-terrain performance. In this way, torque distribution is controlled widely between the front and rear wheels in response to slip conditions and to avoid slip conditions.

Furthermore, in the torque distribution control associated with the occurrence of slip as described above, when turning, the differential limiting torque of the pressure multi-disc clutch 60 is corrected to decrease depending on the steering angle ψ. For this reason,
The differential restriction of the center differential device 50 is reduced, making it possible to sufficiently absorb the difference in rotational speed, avoiding a tight corner braking phenomenon, and ensuring good maneuverability.

Note that the characteristics of torque distribution control for reference torque distribution, slip, etc. can be arbitrarily set other than those in the embodiments.

The embodiments of the present invention have been described above, and FIG. 8(a)
, (b), another embodiment of torque distribution control using the hydraulic multi-disc clutch 80 will be described.

In FIG. 8(a), since the front wheel output side friction reduction gear 17 is loosely fitted into the transmission output lever 15 on the input side of the center differential device 50, there is a gap between the transmission output lever 15 and the reduction gear 17. A hydraulic multi-disc clutch 60 is interposed therein.

As a result, the bypass system 101 of the pressure multi-disc clutch 60 is configured separately for the drive system from the transmission output 15 to the reduction gal 17 and the rear drive shaft 20 via the center differential device 50. In this bypass system 101, when the rear wheels slip or break, a differential function is established in the center differential device 50 in which the rear wheel rotation speed Nlt>the rotation speed of the transmission output shaft 15>the front wheel rotation speed NF, and the transmission output shaft Torque is directly transmitted from the first sun gear 51 to the reduction gear 17 according to the clutch torque Tc, and the torque obtained by subtracting the clutch torque Tc flowing to the front wheel side is transmitted manually from the first sun gear 51 to the first pinion 52, Torque is also transmitted to the rear drive shaft 20 via the second pinion 54 and second sun gear 53.
As a result, the front and rear wheel torques TF and TIL are as follows.

TF −0.38 (Ti −Tc)+TcTIt
−0.62 (Tl −”re) Therefore, in the non-slip state, the Kurasochi torque Tc is zero, so TF:
T conversion-38: Torque is distributed to the rear wheels of 62, and when rear wheel slip occurs, clutch 1・lux Tc occurs,
Input torque TI is directly transmitted to the front wheels in accordance with clutch torque Tc. The larger the clutch torque Tc is, the more the input torque TI flows to the front wheels via the bypass system 101, and the front wheel torque is actively controlled to increase.

In the center differential device 50 shown in FIG. 8(b), the transmission output shaft l5 is the same as the one described above.
A hydraulic multi-disc clutch 60 is interposed between the transmission output lever I5 and the second sun gear 53, and the hydraulic multi-disc clutch 60 is extended to the sun gear 53 side via the drum 61 of the hydraulic multi-disc clutch 60. is connected to the rear drive link 20.

Therefore, in this embodiment as well, the transmission output shaft 15, the hydraulic multi-disc clutch 60, the rear drive shaft 20, and the second
A bypass system 1Ol that connects to the reduction gear 17 via the sun gear 53, the second pinion 54, and the carrier 55.
is configured, and when the rear wheel slips, the rear wheel rotation speed Nlt
>Rotational speed of transmission output shaft I5> A differential function of the front wheel rotational speed NF is established, and the front and rear wheel torques TF, Tl' are as follows.

TF -0.38 (TI -+-re)TR -0.
62 (TI + Tc) - Tc Thus, when rear wheel slip occurs, differential limiting torque corresponding to clutch torque Tc is added to input torque Ti, and front wheel torque TF is actively controlled to increase.

Note that the above-mentioned embodiment shows the application of the present invention to a vertically mounted transaxle of a front engine;
It can also be applied to configurations in which sprockets and chains are provided instead of transfer gears (reduction gears) to transmit torque to the front wheels, such as in 4-wheel drive vehicles based on rear engine/rear drive (RR). It goes without saying that the present invention can also be applied to drive vehicles.

〔Effect of the invention〕

As described above, according to the present invention, in a four-wheel drive vehicle with a center differential, the compound planetary gear type center differential device is composed of two sun gears, a pinion, and a carrier, and the meshing pitch radius of the two sun gears and pinions is Since the standard torque distribution is determined and the gear specifications of the first sun gear and first pinion and the gear specifications of the first sun gear and second pinion can be changed, the degree of freedom in determining the torque distribution is increased. increases significantly.

Furthermore, by increasing the degree of freedom in torque distribution by the center differential device, it is possible to set a reference torque distribution that is sufficiently biased toward the rear wheels with a compact structure having a predetermined strength.

In addition, sufficient torque distribution biasing the rear wheels provides good maneuverability, accurate slip detection, and improved acceleration.

Furthermore, since it assumes sufficient torque distribution with a bias towards the rear wheels, it is possible to perform a wide range of torque distribution control using the hydraulic multi-disc clutch depending on the vehicle's driving conditions and road surface conditions, improving maneuverability and cruising performance against slips, etc. It is possible to improve the performance of both by finely and appropriately controlling the characteristics.

When the rear wheels slip, rear wheel torque flows to the front wheels according to the differential limiting torque of the hydraulic multi-disc clutch, resulting in linear characteristics.

Furthermore, a bypass system for the hydraulic Tasaka clutch is provided separately for each, and when the rear wheels slip, the front wheel torque is accurately increased according to the clutch torque, so the response is better than that of the conventional differential limiting device.

[Brief explanation of the drawing]

Fig. 1 is an overall configuration diagram of an embodiment of a four-wheel drive vehicle according to the present invention, Fig. 2 (a) is a sectional view of the main parts, Fig. 2 (b) is a schematic diagram showing the carrier support column, and Fig. 3 is a center diagram. A schematic diagram of the differential device, Figure 4 is a circuit diagram of the hydraulic and electrical control system of the hydraulic multi-disc clutch, Figure 5 is a diagram showing a clutch pressure map, Figure 6 is a diagram showing the action of torque distribution control, FIG. 7 is a flowchart showing arithmetic processing within the control unit, and FIGS. 8(a) and 8(b) are scale diagrams showing another embodiment of torque distribution control using a hydraulic multi-disc clutch.

Claims (4)

[Claims] (1) In a four-wheel drive vehicle with a center differential in which a front wheel side transmission element and a rear wheel side transmission element are disposed coaxially on the transmission output shaft, and a center differential device is disposed between these three, the above-mentioned center differential The device includes first and second sun gears, first and second pinions integrally formed, and a carrier, the first sun gear is connected to the transmission output shaft, and the carrier is connected to the transmission output shaft. the transmission output shaft of the input element of the center differential device, the carrier of the two output elements, and the second sun gear; Torque distribution control for a four-wheel drive vehicle in which a hydraulic multi-disc clutch that limits the differential of the center differential device is interposed between output elements or between an input element and either one of the output elements, bypassing the hydraulic multi-disc clutch. Device. (2) The center differential device is configured such that a front wheel side transmission element directly connected to the carrier is rotatably provided on the transmission output shaft directly connected to the first sun gear, and the front wheel side transmission element is rotatably provided between the transmission output shaft and the front wheel side transmission element. The torque distribution control device for a four-wheel drive vehicle according to claim (1), wherein a hydraulic multi-plate clutch is interposed in the four-wheel drive vehicle. (3) The above-mentioned center differential device extends the transmission output shaft directly connected to the first sun gear toward the second sun gear, and interposes a hydraulic multi-disc clutch between the two to connect the second sun gear to the second sun gear. The torque distribution control device for a four-wheel drive vehicle according to claim 1, wherein the hydraulic multi-disc clutch is connected to the transfer shaft via the drum side. (4) The reference torque distribution by the center differential device described above is set to be biased toward the rear wheels, and the hydraulic pressure applied to the hydraulic multi-disc clutch is controlled according to the vehicle running condition and road surface condition, so that the clutch torque can be adjusted from one of the input elements. Claim (1) in which the torque distribution between the front and rear wheels is variably controlled by moving one output element or between the output elements.
) Torque distribution control device for a four-wheel drive vehicle.