FR2554768A1 - Control for all wheel drive - Google Patents

Abstract

(+18.11.83,30.11.83,27.04.84, 27.04.84,01.05.84, 31.05.84,05.10.84(2) -JP- 217494,225985, 062793, 085225,088087,111507,207949,209055) (1524DB) The single engine drives both the drive shaft for the front wheels and that for the rear wheels. A hydraulic pump is driven through the speed difference of the two drive shafts and provides a servo pressure to change the transmission from four wheel drive to two wheel drive. The pump is driven in either direction to provide the servo pressure and the two wheel drive is applied to either front or rear wheels.

Description

The present invention relates to a power transmission device intended for a vehicle of the type whose front and rear wheels are driven by the same motor.
In a vehicle of the four-wheel drive type, in which the front and rear wheels are driven by the same engine, there is a slight difference between the effective radii of the tires of the front and rear wheels, or there may appear a difference between the tire lengths of the front and rear wheels as the vehicle crosses a curve - As a result, the tires tend to slip during vehicle travel, and excessive force tends to be exerted on the vehicle engine system .It is therefore necessary to specify an effective means to prevent the communication of such an undesirable force to -: - the vehicle's engine system.

For this purpose, a vehicle of the type with constantly driving four wheels, in which the four wheels must all be driven at the same time, has been provided with a third differential unit commonly affected by the term "central differential", so as to allow the driving force of the engine to be transmitted to all the wheels even when there is a difference in rotation between a first rotary shaft transmitting the driving force to the front wheels, and a second rotary shaft transmitting the driving force to the rear wheels. the vehicle of the type with four-wheel drive constantly has disadvantages quarter to its weight, its dimensions and its cost when compared to a vehicle of the type with four-wheel drive intermittently, whose four wheels are not necessarily driven at the same tempo Similarly, due to the ability of the front and rear wheels to have different rotational speeds, the drive of all four wheels rices may not always be achieved when necessary, and, in order to ensure such entrainement of the four-wheel drive, it is necessary to provide in addition a differential lock mechanism. Thus, the vehicle of the type four wheel drive constantly has the disadvantage, among other things, that its power transmission system has a complex structure.

A method relating to such a vehicle of the type with four-wheel drive constantly, is described in Japanese Patent Application NO 58-20521 (1983). According to the described method, the central differential is replaced by a liquid type multi-plate clutch arranged in the clutch side of the drive, and such a clutch is caused to slide while the vehicle crosses a curve, thus absorbing the difference between the speed3 of rotation of the front and rear wheels. However, the proposed method has, inter alia, the disadvantage that said multi-plate clutch of the liquid type tends to be deteriorated by heat in gender due to the effect of said sliding, and that the clutch has limited torque transmission capacity.

Such a central differential does not generally equip vehicles of the four-wheel drive type intermittently. Consequently, the driver of a vehicle of this type must carry out the operations required so that only two of the four wheels are driven when a phenomenon of braking in tight turns occurs, a phenomenon peculiar to four-wheel drive which can occur. occur when the vehicle crosses a curve. Consequently, the vehicle of the intermittently four-wheel drive type has, inter alia, the disadvantage that the driver must carry out complex manipulations in order to drive the vehicle.

In order to eliminate all the drawbacks of vehicles according to the prior art, in which the front and rear wheels are driven by the same motor3 one of the primary aims of the present invention is to provide a power transmission device intended for a vehicle, which has small dimensions and low weight.

The present invention achieves this goal by proposing a power transmission device intended for a vehicle, characterized in that it comprises a first rotary shaft transmitting the driving force to the front wheels, a second rotary shaft transmitting the driving force to the rear wheels , a hydraulic oil pump connecting the first rotary shaft to the second rotary shaft and intended to be driven as a function of the difference between the speeds of rotation of the first and second rotary shafts, thereby delivering a quantity of hydraulic oil corresponding to the difference rotation speeds, the hydraulic oil pump having at least two lights alternately switched between the outlet side and the suction side, according to the direction of relative rotation of the first and second rotary shafts, and a hydraulic control circuit comprising a passage hydraulic oil establishing communication between one of the two lights with the other and a means of comma Hydraulic valve arranged in the hydraulic oil passage to control the pressure of the hydraulic oil delivered by the hydraulic oil pump.

Various embodiments of the power transmission device according to the invention will now be described in more detail, by way of non-limiting examples, with reference to the accompanying drawings in which: - Figure 1 is a diagram showing in schematic form the structure of a vehicle of the four-wheel drive type to which the power transmission device according to the present invention is applied; - Figure 2 is a schematic sectional view showing a first embodiment of the power transmission device according to the present invention; - Figures 3 (a) and 3 (b) illustrate the respective directions of flow of the hydraulic oil in the first embodiment; - Figure 4 is a schematic sectional view showing a second embodiment of the power transmission device according to the present invention; ; - Figure 5 is a schematic sectional view showing a third embodiment of the power transmission device according to the present invention; - Figure 6 is a schematic sectional view showing a fourth embodiment of the power transmission device according to the present invention; - Figure 7 is a graph showing, in the case of the fourth embodiment, the pressure delivered by the vane pump as a function of the difference between the rotational speeds of the front wheels and the rear wheels; - Figure 8 is a schematic sectional view of a part of the hydraulic control circuit of a fifth embodiment of a power transmission device according to the present invention; - Figure 9 is a schematic sectional view of part of the hydraulic control circuit of a sixth embodiment of a power transmission device according to the present invention; - Figure 10 is a schematic sectional view of a seventh embodiment of the power transmission device according to the present invention; - Figure 11 is a diagram showing the structure of another vehicle of the four-wheel drive type, to which the power transmission device according to the present invention is applied; - Figure 12 is an axial sectional view of a gear pump incorporated in a ninth embodiment of the power transmission device according to the present invention; - Figure 13 is a sectional view along line A-A of Figure 12 ;; - Figure 14 is a diagram showing, in schematic form, the structure of the hydraulic control circuit of the ninth embodiment; - Figures 15 to 24 relate to a tenth embodiment of the present invention; - Figure 15 is a vertical sectional view showing part of the hydraulic oil pump; FIG. 16 is a diagram showing, in schematic form, the structure of a vehicle of the four-wheel drive type to which the power transmission device according to the present invention is applied; - Figure 17 is a schematic view showing part of the hydraulic oil pump; - Figures 18 (a) and 18 (b) are diagrams showing the principles of operation of the transmission device according to the invention; - Figure 19 is a sectional view along the line V-V of Figure 15 ;; - Figure 20 is a schematic view of the hydraulic circuit; - Figures 21 (a), 21 (b), 21 (c) and 22 are schematic views respectively showing the conditions of engagement in a key forming part; - Figures 23 (a), 23 (b) and 23 (c) respectively illustrate the engagement conditions in a key-forming part, Figure 23 (a) being a view in vertical section, Figure 23 (b) a plan view of a rotor shaft, and Figure 23 (c) being a schematic view showing the respective engagement conditions; - Figures 24vau, 24 (b) and 24 (c) are schematic views of the hydraulic circuits intended to cause the lifting of the pallets; - Figure 25 is a vertical sectional view of a part showing an example of a variant of a hydraulic oil supply line according to an eleventh embodiment of the power transmission device according to the invention; - Figure 26 is a vertical sectional view of a part showing another example of a variant of a hydraulic oil supply line according to a twelfth embodiment of the power transmission device according to the invention; - Figure 27 is a graph showing the pressure delivered by a vane pump; - Fig 28 is a vertical sectional view showing part of a thirteenth embodiment of the power transmission device according to the invention5 - Figure 29 is a vertical sectional view showing part of a supply line in hydraulic oil, intended to adjust the backlash play in the thirteenth embodiment of the power transmission device according to the invention; - Figure 30 is a schematic side view showing part of a rotor; - Figure 31 is a sectional view along line X-X of Figure 30; and - Figure 32 is a sectional view showing a vane portion of a preferred vane pump according to the present invention.

Referring first to FIG. 1 showing the structure of the four-wheel drive type vehicle to which the present invention is applied, a transmission mechanism 2 is connected to an engine 1 comprising a crankshaft extending in the transverse direction of the vehicle, and an output shaft 3 of the transmission mechanism 2 transmits the driving force to a driving pinion 4. From the driving pinion 4, the driving force is transmitted, via a idler pinion 5, to a shaft of intermediate transmission 8 at the two ends of which pinions 6 and 7 are respectively mounted. From a first pinion or from pinion 7 mounted at a first end of the intermediate transmission shaft 8, the driving force is transmitted to a differential 10 intended driving the front wheels 9. On the other hand, the driving force transmitted to the front wheels 9 is transmitted directly, via a pinion 12, to a first rotary shaft 11. the driving force is then transmitted se, via a power transmission device 13, to a second rotary shaft 14. The driving force is then transmitted, via a gear mechanism 15 which changes the direction of the power transmission , to a differential 17 intended to drive the rear wheels 16.

A first embodiment of the power transmission device 13 according to the present invention will now be described with reference to FIG. 2. With reference to this figure, the power transmission device 13 constituting an embodiment of the present invention, comprises a hydraulic oil pump produced in the form of a vane pump 20, and an associated hydraulic control circuit 21. The vane pump 20 includes a rotor 20a and a cam ring 20b. The rotor 20a is coupled to the first rotary shaft 11 to which the driving force transmitted to the front wheels is directly transmitted. The cam ring 20b is coupled to the second rotary shaft 14 transmitting the driving force to the rear wheels 16. The vane pump 20 delivers a quantity of hydraulic oil which is proportional to its speed of rotation.

More specifically, this vane pump 20 operates as a hydraulic oil pump when it appears a relative rotation between the rotor 20a and the cam ring 20b, that is to say when it appears a relative rotation between the first rotary shaft 11 and the second rotary shaft 14. The operation of this vane pump 20 is such that when the lights acting as exit lights (which are head lights in the direction of relative rotation) are closed, the rotor 20a and the cam ring 20b are made to rotate integrally, constituting a rigid body under the effect of the static pressure of the hydraulic oil. To this end, the cam ring 20b is provided with two pump chambers arranged in diagonal positions, and the cam ring 20b is also provided with four lights 22, 23, 24 and 25 arranged in substantially diagonal positions, so that those of them which are at the tail and at the head in the direction of relative rotation, act respectively as suction and exit lights. The lights 22 and 24 arranged diagonally to perform the same function communicate with each other by means of a first passage of hydraulic oil 26. Similarly, the lights 23 and 25 arranged diagonally to perform the same function communicate with each other by means of a second hydraulic oil passage 27. The first and second hydraulic oil passages 26 and 27 are each connected to the associated ports 22, 23, 24 and 25 by the through a mechanism which allows a supply of hydraulic oil even when the cam ring 20b is being rotated.

The first and second hydraulic oil passages 26 and 27 communicate with a hydraulic oil tank 30 respectively via a first valve unit or first control valve 28 and a second valve unit or second valve 29. The two passages 26 and 27 allow the hydraulic oil to flow exclusively from the hydraulic oil tank 30. In addition, these passages 26 and 27 communicate with each other by means of two control selector valves 31 and 32 which respectively allow the flow of hydraulic oil coming from the associated passages 26 and 27. The intermediate space between these two control valves 31 and 32 serving as flow selector member, communicates with a check valve 33 arranged in a hydraulic oil outlet passage 41 serving as a flow control member. A communication passage 35 extends between an intermediate part of the check valve 33, which comprises a valve element provided with an associated spring 34, and intermediate passage parts between the oil tank 30 and the two valves control 28 and 29. A piston 36 is arranged on the other side of the spring 34 to control the force of the spring 34 which normally exerts pressure in the direction of closing of the check valve 33. The pressure of the hydraulic oil , which is controlled in a servo-controlled manner by the conditions of use in a manner which will be described later, acts on the piston 36. For the purposes of this servo-controlled, hydraulic oil, introduced at constant pressure by a nozzle 37, is controlled by means of a solenoid valve 38. This solenoid valve 38 is electrically connected to a calculation means 39. Signals representative of the speed of rotation of the motor 1, of the speed of rotation of the first rotary shaft 11, from the speed of rotation of the second rotary shaft 14 and the opening of the throttle valve are applied, together with the output signals delivered by a brake actuation detection switch and a steering angle detector, by calculation means 39 which controls the pressure of the hydraulic oil acting on the piston 36. The oil at constant pressure introduced by the nozzle 37 can be constituted by hydraulic oil used for controlling the transmission 2 when the transmission is of the automatic type , or by an oil pump when the transmission 2 is of the manual type. Such a hydraulic oil can also be constituted by the hydraulic oil used for the power-assisted steering, the hydraulic oil supplying the servo-brakes or the hydraulic oil derived from the outlet side of the vane pump 20.

Thanks to such an arrangement of the hydraulic control circuit 21, the pressure of the hydraulic oil delivered by the outlet lights of the vane pump 20 always acts on the active element of the check valve 33 and the oil tank. hydraulic 30 communicates with the suction lights of the vane pump 20 regardless of the direction of relative rotation of the rotor 20a and the cam ring 20b.

The operation of such a power transmission device 13 will now be described with reference to the case in which the force pushing the check valve 33 in the direction of its opening against the force of the spring 34, is constant, that is to say when the force of the spring 34 is the only one to oppose the opening of the check valve 33
When the vehicle is in the usual condition for traveling in rectilinear forward motion, the effective radius of the tires of the front wheels 9 is the same as that of the rear wheels 16, and the slip rate of the tires in rotation is very small. In such a condition, there appears to be no difference between the rotational speeds of the first rotary shaft 11 and the second rotary shaft 14 of the power transmission device 13. Consequently, the vane pump 20 does not deliver any hydraulic oil under pressure , and no driving force is transmitted to the rear wheels 16. Thus, the vehicle is driven only by the front wheels 9 alone, that is to say that the vehicle moves in a front-wheel drive mode.

However, when the forward vehicle is caused to accelerate, for example, a slip of less than about 1% generally occurs on the front wheels 9, although this phenomenon is not appreciable.

When there is a difference in rotational speed between the first and second rotary shafts 11 and 14 under the effect of the sliding of the aforementioned front wheels 9, the vane pump 20 is made active to raise the pressure as a function of said difference in rotational speeds. The rotor 20a and the cam ring 20b rotate in an integral manner, and the motive force corresponding to the accumulated pressure and to the surface of the pallets subjected to said pressure, is transmitted to the rear wheels 16 to establish a mode of operation in four drive wheels. In such a case, the flow of hydraulic oil in the vane pump 20 is that shown in Figure 3 (a). We can see in this Figure 3 (a) that, due to the rotation of the rotor 20a relative to the cam ring 20b, the lights 23 and 25 then act as suction lights for the hydraulic oil, and the latter is sucked from the hydraulic oil tank 30 to penetrate the suction lights 23 and 25 via the control valve 29. On the other hand, the lights 22 and 24 act as outlet lights to close the control valve 28 and the control selector valve 32, and, at the same time l the hydraulic oil delivered is sent, via the control selector valve 31, in the direction of the check valve 33. In FIG. 3 (a) the solid lines and the dashed dashed lines indicate the directions of flow respectively hydraulic oil delivered and hydraulic oil drawn e.

It is now assumed that the speed of rotation of the front wheels 9 becomes very high compared to that of the rear wheels 16, as, for example, when the vehicle is traveling on a snow-covered road or is being accelerated or accelerated. sudden braking which results in a locking of the rear wheels 16. In such a case, the difference between the speeds of rotation of the first and second rotary shafts 11 and 14 included in the power transmission device 13 becomes very large. As a result, high pressure is generated in the vane pump 20, and hydraulic oil then flows in the directions shown in Figure 3 (a). When the hydraulic oil pressure exceeds a predetermined level, the check valve 33 is open against the force of the spring 34, and the pressure of the hydraulic oil delivered is controlled so as to be substantially constant. Thus, a constant driving force corresponding to the regulated pressure of the hydraulic oil delivered, is transmitted to the rear wheels 16 in order to establish the four-wheel drive traction mode. Consequently, the rotational speed of the front wheels 9 decreases while that of the rear wheels 16 increases, so that the difference between the rotational speeds of the front and rear wheels 9 and 16 is decreased. (this operation is the same as that of the differential without sliding).

Thus, when a slip occurs on the front wheels 9, the engine torque intended for the rear wheels 16 is increased in order to prevent them from being unable to roll, while, when the rear wheels 16 tend to lock up , the braking torque for the front wheels 9 is increased so as to prevent the rear wheels 16 from locking.

A case is also assumed in which the speed of rotation of the rear wheels 16 is very high compared to that of the front wheels 9, such as, for example, when the front wheels 9 tend to lock under the effect of an actuation brake. In such a case, a very large difference appears between the rotational speeds of the first and second rotary shafts 11 and 14 of the power transmission device 13, in a direction opposite to the aforementioned direction. Consequently, the flow of hydraulic oil in the vane pump 20 is now that shown in FIG. 3 (b). It can be seen in this FIG. 3 (b) that the lights 22 and 24 then act as lights hydraulic oil suction, and that hydraulic oil is sucked from the reservoir 30 to enter the suction ports 22 and 24 via the control valve 28, while, on the other apart, the lights 23 and 25 act as hydraulic oil outlet lights. Consequently, hydraulic oil flowing through the second passage 27 closes the control valve 29 and the control selector valve 31, and, at the same time, high pressure hydraulic oil introduced from the second passage 27 flows through the selector valve 32 towards the check valve 33. The hydraulic oil pressure is also kept constant by the check valve 33, the corresponding constant driving force is transmitted to the wheels rear 16 to establish the four-wheel drive mode. This results in an increase in the braking torque communicated to the rear wheels 16 in order to prevent the front wheels 9 from locking.

In a usual condition for the vehicle to cross a curve, the speed of rotation of the front wheels 9 is slightly higher than that of the rear wheels 16, and the vehicle thus crosses the curve in the four-wheel drive mode of traction, mode in which the braking torque is communicated to the front wheels 9 while the engine torque is communicated to the rear wheels 16.

In the same way as that described above, the outlet pressure of the hydraulic oil is controlled, by means of the check valve 33 disposed in the power transmission device 13, so as not to exceed a constant value. Consequently, contrary to the case of the prior art in which manipulation is required on the part of the driver to establish the four-wheel drive mode of traction in a vehicle of the four-wheel drive type intermittently, the switching between the four-wheel drive mode and the two-wheel drive mode can be performed automatically, and the four-wheel drive mode is established, according to the present invention, by the driving force corresponding to the difference between the rotational speeds of the front and rear wheels. Similarly, when compared to the central differential which necessarily equips a vehicle of the type with four-wheel drive constantly, the power transmission device 13 according to the present invention is small, has a compact structure and is of light weight and low cost.

As has already been described, the pressure of the hydraulic oil acting on the lower end of the piston 36 is controlled in a controlled manner on the conditions of use in order to regulate the force required for the opening of the valve. restraint 33. By means of such regulation, the pressure of the hydraulic oil leaving the vane pump 20 can be regulated or controlled, and the driving force transmitted to the rear wheels 16 can also be regulated.

Consequently, the way in which such a slave control is carried out can be such that, when an increased load of the motor 1 is detected on the basis of the output signal delivered by the throttle opening sensor, the oil pressure The hydraulic power delivered by the vane pump 20 is increased accordingly, so that the vehicle can operate in the four-wheel drive mode, in which the driving force transmitted to the rear wheels is correspondingly increased.

In addition, the manner in which such a servo control is carried out, may be such that the pressure of the hydraulic oil supplied by the vane pump 20 is increased when the output signal of the brake actuation detection switch detecting the foot brake actuation state indicates that this contact is activated, so that the front and rear wheels 9 and 16 can be locked so as to reduce the vehicle's braking distance and the vehicle can also be braked retaining its stability.

In addition, the way in which such a servo control is carried out, may be such that the pressure of the hydraulic oil supplied by the vane pump 20 is reduced when the steering angle detected by the steering angle sensor increases. , that the vehicle can cross a curve regularly without giving rise to the phenomenon of braking in a tight curve. Furthermore, the vehicle can run stably when the pressure of the hydraulic oil supplied by the vane pump 20 is regulated or controlled as a function of the speed of rotation of the engine 1 and of the linear speed of the vehicle detected on the basis sensor output signals, which are sent to the computing means 39.

A second embodiment of the present invention, in which the hydraulic control circuit 21 is different from that used in the first embodiment, will now be described with reference to FIG. 4. In this FIG. 4, the same are used reference numerals to designate the parts appearing identically in Figure 2. In the power transmission device 13 shown in Figure 4, the structure of the hydraulic oil pump 20 is the same as that described above. Figure 4, the flow selector member or the two control selector valves 31 and 32 shown in Figures 3 (a) and 3b) arranged (es), in the first embodiment, between the first and second passages d hydraulic oil 26 and 27 of the hydraulic control circuit 21 to allow the flow of hydraulic oil from the only respective passages 26 and 27, are replaced by a single selector valve 40 of the drawer type. s hydraulic oil passages defined by the two ends of the selector valve slide 40 communicate with the first or second passages 26 and 27, and also communicate with the hydraulic oil tank 30 respectively via the control valves 28 and 29, and a hydraulic oil outlet passage 41 communicates with the middle part of the hydraulic oil passages switched by means of the slide of the selector valve 40.

The check valve 33 is replaced by a pressure control valve 42 which comprises a drawer 42a having two spaced zones. The hydraulic oil outlet passage 41 communicates with the space defined between the two zones of the drawer 42a, and a passage 43 intended to introduce the hydraulic oil at a regulated pressure, communicates with the oil tank 30 and also with the space defined between the two zones of the drawer 42a. An elastic force of a spring 42b acts on the left end of the slide 42a, and the pressure of the hydraulic oil controlled in a controlled manner by the combination of the nozzle 37 and the solenoid valve 38 also acts on this same end.

Consequently, hydraulic oil is obtained at a pressure regulated as a result of a balance between the elastic force acting on the slide and established by the combination of the force of the spring 42b and the controlled pressure of the hydraulic oil, and the elastic force acting on the slide and corresponding to the difference in area of the two areas of the slide 42a to which the hydraulic oil from the outlet step 41 is applied. The hydraulic oil having a pressure thus regulated is introduced into the passage 43 to be brought back to passage 26 or 27 connected to the suction lights.

It is now assumed that the relative speed of rotation of the first rotary shaft 11 is greater than that of the second rotary shaft 214. It is then assumed that, in such a case, the rotor 20a rotates clockwise, as well as it has been described with reference to Figure 3 (a). In accordance with the arrangement of such a hydraulic control circuit 21, the first and second passages 26 and 27 are then now connected respectively to the outlet and suction ports. Consequently, the pressure of the hydraulic oil delivered acts on the left end face of the drawer of the selector valve 40 to urge this drawer towards its rightmost position, and the first passage 26 communicates with the outlet passage 41. consequently, the hydraulic oil delivered by the first passage 26 is introduced, via the outlet passage 41, towards the pressure control valve 42, and the hydraulic oil at the regulated pressure is circulated towards the lights suction via the control valve 29.

It is then assumed that the relative speed of rotation of the second rotary shaft 14 is greater than that of the first rotary shaft 11. The cam ring 20b then rotates clockwise> as has been described with reference in Figure 3 (b), and the second and first passages 27 and 26 are now connected respectively to the outlet and suction ports.

Consequently, the pressure of the hydraulic oil delivered acts on the right end face of the drawer of the selector valve 40 disposed in the hydraulic control circuit 21 of FIG. 4, and this drawer is biased towards its leftmost position in order to to allow the communication of the second passage 27 with the outlet passage 41. The hydraulic oil coming from the second passage 27 is introduced towards the pressure control valve 42, so that hydraulic oil at the regulated pressure is introduced by through the control valve 28 in order to be circulated to the suction ports.

Thus, whatever the direction of relative rotation of the first and second rotary shafts 11 and 14, hydraulic outlet oil is always introduced into outlet passage 41. Consequently, when the pressure of the hydraulic oil acting on the slide 42a of the pressure control valve 42 is regulated by means of the solenoid valve 38, the pressure of the hydraulic oil supplied by the hydraulic oil pump 20 can be controlled, so that the mode of traction corresponding to the running state of the vehicle at this precise instant can be established in the manner already described.

A third embodiment of the present invention, in which the hydraulic control circuit 21 is different from that used in the first and second embodiments, will now be described with reference to FIG. 5. In this FIG. 5, we use the same reference numerals to designate identical parts appearing in Figures 2 and 4. -
In the embodiment shown in FIG. 5, the selector valve 40 used in the second embodiment is replaced by a valve valve 45 controlled by the opening and closing of a solenoid valve 44. This valve valve 45 comprises a drawer 45a comprising three spaced zones. In FIG. 5, the first and second passages of hydraulic oil 26 and 27 communicate respectively with the spaces defined between the three zones of drawer 45a and also communicate with the reservoir 30 and the passage regulated oil 43 from the pressure control valve 42 via the control valves 28 and 29 respectively. The outlet passage 41 communicates with the valve 45, to be opened or closed by the central zone of the drawer 45a, so that the first and second passages 26 and 27 can also communicate with the pressure control valve 42. A spring 45b is mounted at the left end of the slide valve 45a of the slide valve valve 45 , and the solenoid valve 44 with controlled opening and closing is arranged upstream of a nozzle 46 formed in a passage of hydraulic oil in communication with the right end of the slide valve 45a 45. The calculation means 39 is electrically connected to the solenoid valve 44.

It is assumed that the relative speed of rotation of the first rotary shaft 11 is greater than that of the second rotary shaft 14, and that, in this case, the rotor 20a rotates clockwise. As has been described with reference to FIGS. 3 (a), the first and second passages 26 and 27 are then connected respectively to the outlet and suction ports placed in such a hydraulic control circuit 21.

The calculation means 39 detects the direction of rotation (the direction of relative rotation) of the hydraulic oil pump 20, on the basis of the input signals representative of the speeds of rotation of the first and second rotary shafts 11 and 14. The calculation means 39 then turns the solenoid valve 44 to the open position in order to open the passage upstream of the nozzle 46. The slide valve valve 45a 45 is biased in its rightmost position by force elastic stressing the drawer provided by the combination of the force of the spring 45b and the force corresponding to the difference in surface area of the two zones of the drawer 45a to which the hydraulic oil from the first passage 26 is applied, thus establishing communication between the first passage 26 and the outlet passage 41. The hydraulic oil introduced into this outlet passage 41 is circulated as in the case of FIG. 4.

Thus, by virtue of the use of the slide valve 45 whose movement is controlled by the solenoid valve 44, the valve 45 can be actuated reliably, regardless of the direction of relative rotation of the oil pump hydraulics 20.

A fourth embodiment of the present invention, in which the hydraulic control circuit 21 is different from that used in the first embodiment, will now be described with reference to FIG. 6. In this FIG. 6, the same are used reference numbers to designate identical parts appearing in FIG. 2.

In the embodiment shown in FIG. 6, the pressure by means of which the check valve 33 is biased in its open position is not controlled by the calculation means 39, but is fixed at a predetermined value defined by the spring force 34, so that the vane pump outlet lights 20 may not communicate with the reservoir 30 when the hydraulic oil pressure is not greater than the setting pressure of the check valve 33 .

A first auxiliary hydraulic oil passage 52 comprising a nozzle or flow restrictor 50, and a second auxiliary hydraulic oil passage 53 comprising a nozzle or flow restrictor 51 are connected respectively between the ports 23 and 24, and between the lights 22 and 25, so that the check valve 33 can perform its function of regulating or controlling the pressure. Thus, the check valve 33 constitutes a control member for the flow of hydraulic oil, associated with the nozzles 50 and 51.

More specifically, the flow of hydraulic oil through the auxiliary passages 52 and 53 is generally restricted by the respective nozzles 50 and 51. Consequently, when the difference in the speeds of rotation of the first and second rotary shafts 11 and 14 is small, the amount of hydraulic oil flowing in the auxiliary passages 52 and 53 is also small. As the difference in rotational speeds between the first and second rotary shafts 11 and 14 increases, the pressure of the hydraulic oil supplied by the vane pump 20 becomes higher and higher until it finally overcomes the resistance nozzles 50 and 51 opposing the flow. In such a situation, hydraulic oil flows through the auxiliary passages 52 and 53 before the pressure of the hydraulic oil delivered reaches the control pressure value assigned to the check valve 33, so as to transmit to the second rotary shaft 14 the torque corresponding to the difference in rotation speeds. Thus the auxiliary passages 52 and 53 act to adjust the pressure control function of the check valve 33.

When the difference in the rotational speeds of the front wheels 9 and the rear wheels 16 is small, regardless of the direction of relative rotation of the rotor 20a and the cam ring 20b of the vane pump 20, the hydraulic oil pressure delivered by the vane pump 20 overcomes the flow resistance of the nozzles 50 and 51 in such a hydraulic control circuit 21, and hydraulic oil flows through the auxiliary passages 52 and 53 to penetrate the lights d suction, so that the torque corresponding to the difference in rotational speeds is transmitted to the rear wheels 16. When, on the contrary, the difference in rotational speeds of wheels 9 and 16 is large, the hydraulic oil tank 30 does not not communicate with the suction lights of the vane pump 20 before the pressure of the hydraulic oil delivered reaches the level exceeding the setting of the active spherical element 36 of the check valve 33.

The difference in the rotational speeds of the front and rear wheels 9 and 16 is shown, as a function of the pressure of the hydraulic oil delivered, in FIG. 7, in which the characteristic curve a represents this relationship when the nozzles 50 and 51 are provided (that is to say when the auxiliary passages 52 and 53 are provided), while the characteristic curve b represents this relationship when the nozzles 50 and 51 are not provided. It will be seen in FIG. 7 that the pressure of the hydraulic oil delivered is lower when the device includes the nozzles 50 and 51.

In the power transmission device 13 shown in FIG. 6, the pressure of the hydraulic oil delivered by the vane pump 20 is maintained at a predetermined value thanks to the operation of the check valve 33, and the auxiliary passages 52 and 53, together with the respective nozzles 5O and 51, are provided such that, when the difference in rotational speeds exceeds a predetermined value which may be small, the torque corresponding to this difference in rotational speeds can be transmitted to the rear wheels 16 Consequently, the present invention eliminates the disadvantage of transmitting a high torque to the rear wheels 16, despite the fact that the closing of the hydraulic oil passage results in an accumulation and an increase in pressure so , and that the difference in rotational speeds is not that great. Thus, the torque corresponding precisely to the difference in rotation speeds can be transmitted to the rear wheels 16.

A fifth embodiment of the present invention will now be described with reference to Figure 8. In the embodiment shown in this figure 8, the arrangement of the hydraulic oil pump 20 is the same as that shown in Figure 6 , except that a flow restrictor 55, having the structure shown, is disposed in each of the auxiliary passages 52 and 53 instead of the nozzles 50 and 51. This flow restrictor 55 is disposed in each of the auxiliary passages 52 and 53 establishing a connection between the outlet and suction ports of the pump 20. The flow restrictor 55 comprises a housing 57 disposed between the outlet passage and the suction passage, a needle valve 60 mounted in the housing 57 by means of a spring 58 to openably block a communication lumen formed in the wall of the auxiliary passage 52 (53), a diaphragm 59 fixed to the needle valve 60 to make the expansion and contraction dependent on the engine load (not shown), and a pressure transmission lumen 62 formed in the wall of the bolt case 57 to transmit to the diaphragm 59 the depression prevailing in the intake manifold of the engine.

The higher the engine torque, the higher the vacuum in the engine intake manifold and the smaller the opening of the communication light 61. However, the communication light 61 is generally kept in a slightly open state without being completely closed, and when the motor load decreases, the communication light 61 opens more widely. On the other hand, when the engine load increases, the driving force increases correspondingly. The communication light 61 is made narrower by means of the advancing movement of the needle valve 60 so as to establish the four-wheel drive mode of traction.

 A sixth embodiment of the present invention will now be described with reference to Figure 9. In the embodiment shown in this Figure 9, the arrangement of the hydraulic oil pump 20 is the same as that shown in Figure 6 , except that a flow restrictor 65 having the structure shown, is disposed in each of the auxiliary passages 52 and 53 instead of the nozzles 50 and 51. This flow restrictor 65 comprises a bolier 66 disposed between the outlet passage and the suction passage. A pressure transmission light 67 is provided in the wall of the housing 66 for transmitting the pressure of the hydraulic oil delivered by the oil pump of the power steering mechanism fitted to the steering wheel, and another pressure transmission light 68 is provided in the wall of the bolier 66 to transmit the vacuum prevailing in the intake manifold of the engine. A piston 70 loaded by a spring 69 is t mounted in the boltier 66 between the port 67 for transmitting the pressure at the inlet and the port 68 for transmitting the outlet pressure of the oil pump, so as to make any vertical movement dependent on the outlet pressure of the oil pump and inlet pressure. A needle valve 71 intended to open and close the auxiliary passage 52 (53) is fixed to the lower end of the piston 70.

The larger the steering angle, the higher the oil pressure of the power steering mechanism. In such a case, the needle valve 71 placed in the flow restrictor 65 having the structure described above, performs a withdrawal movement to widely open the auxiliary passage 52 (53), thereby increasing the difference in speed difference of rotation of the front and rear wheels 9 and 16. On the other hand, when a large difference occurs between the speeds of rotation of the front and rear wheels 9 and 16 while the vehicle is traveling in rectilinear forward travel, this tolerance of difference in rotational speeds is canceled, and the required torque is transmitted to the front wheels 9.

In addition, the vacuum prevailing in the intake manifold, which is transmitted by the transmission light 68 to open and close the auxiliary passage 52 (53), can be used to act in mutual locking relationship with the engine torque, so that the mode of traction of the vehicle can between modified suitably between the four-wheel drive or two-wheel drive modes, depending on the relative magnitudes of the engine torque and the steering angle.

it should be noted in this connection that the spring 69 included in the restrictor 65 must have a considerable force, since there is a very large difference between the pressure prevailing in the intake manifold and the oil pressure of the power steering.

It is also possible to use a means for variably restricting the flow of hydraulic oil in the auxiliary passages 52 and 53 according to the operating conditions of the vehicle, in order to achieve a goal similar to that described above. For example, the flow of hydraulic oil through the auxiliary passages 52 and 53 can be suitably restricted as a function of the brake oil pressure or as a function of actuation or non-actuation of the pedal d 'accelerator. In addition, the flow of hydraulic oil through the auxiliary passages 52 and 53 can be suitably restricted, for control purposes, depending on the running speed of the vehicle, the angular turning speed, etc. .

A seventh embodiment of the present invention will now be described with reference to FIG. 10, in which the pressure of the hydraulic oil delivered by the hydraulic oil pump or the vane pump 20 is used to urge the vanes of the pump 20.

With reference to this figure 10, the hydraulic oil pump or vane pump 20 is provided with several or, for example, eight radial slots 80 regularly spaced on the circumference of the outer peripheral surface 20c of the rotor 20a, and, in each of said slots 80, a pallet 81 is inserted so as to come into sliding contact with the inner peripheral surface 20e of the cam ring 20b.

In this embodiment, the vanes 81 of the pump 20 are normally biased towards and resting on the inner peripheral surface 20c of the cam ring 20b by means of a hydraulic oil supply unit M acting as a unit for soliciting pallets. More specifically, the rotor 20a is provided with an annular recess 83 communicating with the bottom (the radially inner end) of the slots 80, and this annular recess 83 communicates, via a passage for hydraulic oil 84, with an accumulator 85 comprising a piston 85a biased by a spring 85b. A control valve comprising an active valve element 86a in the form of a parasol is disposed in the passage 84 which is connected to the outlet passage 41.

According to this arrangement, the pallets 81 come into sliding contact with the inner peripheral surface 20e of the cam ring 20b while being stressed by the high pressure working oil accumulated in the accumulator 85, when the hydraulic oil pressure fall due to a decrease in the difference in the speeds of rotation of the rotor 20a and the cam ring 20b, and the control valve 86 is urged to return to its closed position. Consequently, even when the relative rotation between the rotor 20a and the cam ring 20b ceases following the stopping of the vehicle, the pallets 81 are still biased towards and bearing on the inner peripheral surface 20e of the ring with cam 20b, so that the coupling function of the pump 20 is maintained at a sufficient level.

The power transmission device according to the present invention is not only applicable to vehicles of the four-wheel drive type shown in Figure 1, but it is also applicable to> for example, other vehicles of the four-wheel type drive wheels shown in Figure 11.

Referring to Figure 11 which shows an application of an eighth embodiment of the present invention, a transmission 2 is connected to an engine 1 having a crankshaft extending in the longitudinal direction of the vehicle, and an output shaft 3 of the transmission 2 transmits the driving force to a driving pinion 4. From this driving pinion 4, the driving force is transmitted, by means of a driven pinion 5, to a second rotary shaft 14. From this second rotary shaft 14, the driving force is transmitted to a differential 17 associated with the rear wheels 16, so that the rear wheels 16 are driven directly by the motor 1 via the aforementioned power train. , the driving force transmitted to the second rotary shaft 14 is transmitted, via the power transmission device 13 according to the present invention, to a first rotary shaft 11, and this driving force is transmitted to an associated differential 10 to the front wheels 9, so that the front wheels 9 are driven indirectly, by means of the power transmission device 13, by means of the driving force driving the rear wheels 16. The rotor 20a of the vane pump 20 (not shown) is coupled to the second rotary shaft 14 to which the driving force driving the rear wheels 16 is transmitted without modification, and the cam ring 20b of the vane pump 20 is coupled to the first rotary shaft 11 transmitting the driving force to front wheels 9.

According to such an arrangement, a relatively large driving force is applied to the rear wheels 16, for which the adhesion limit torque becomes high due to the fact that they receive a proportion of the weight of the vehicle which is higher at the time of abrupt acceleration. As a result, engine power can be used effectively to provide excellent acceleration performance. In addition, since only the fraction of the engine torque exceeding the adhesion limit of the rear wheels and giving rise to a slip is transmitted to the front wheels 9 by means of the vane pump 20, the torque transmission capacity of the vane pump 20 need not be very high.

In addition, as the driving force is distributed in a greater proportion on the rear wheels 16 than on the front wheels 9 while being limited by the torque transmission capacity of the vane pump 20, the adhesion force front wheels 9, which are the steered wheels, can be maintained at a level high enough to improve the steering performance of the vehicle.

A ninth embodiment of the present invention, in which the vane pump is replaced by a gear pump, will now be described with reference to Figures 1? to 14.

With reference to these Figures 12 to 14, the power transmission device of this embodiment comprises a gear pump 120 and a hydraulic control circuit 121 associated with the gear pump 120.

The gear pump 120 comprises a boltier 122 provided with cylindrical cavities 122a, 122b and 122c. A first pinion 123, a sun pinion 124 and a second pinion 123 meshing with one another are housed respectively in these cavities 122a, 122b and 122c, and the two pinions 123 and the single sun pinion 124 constitute two pumps. A toothing 125 is formed on the outer periphery of the bolier 122 to transmit power to the first rotary shaft 143 which will be described later.

Lights 126, 127, 128 and 129 communicating with the gear pump 120 are provided in one of the side walls of the boltier 122, and a casing 133 provided with hydraulic oil passages communicating with these lights 126, 127, 128 and 129 is disposed on this same side wall.

A first cover 135 is arranged on the other side wall of the bolt case 122, and a second cover 136 is arranged outside the first cover 135. The casing 133, the first cover 135 and the second cover 136 are all fixed to the housing 122 by means of bolts 140. A key 141 is formed on the inner peripheral surface of the end portion of the second cover 136, and, by means of this key 141, the second cover 136 is coupled to a first rotary shaft 143 transmitting the driving force to the front wheels 9. This first rotary shaft 143 is hollow and has a gear portion 145 which is coupled to a differential 147 associated with the front wheels 9, so that the driving force can be transmitted to the front wheels 9 from the first rotary shaft 143.

A second rotary shaft 149 transmitting the driving force to the rear wheels 16 extends freely over the entire length of the first shaft 143, the first cover 135 and the second cover 136. This second rotary shaft 149 is arranged coaxially with the axis of rotation of the sun gear 124 to be coupled to the latter. Bearings 142 are placed so as to freely rotate the gear pump 120 etc. inside the transmission housing.

The operation of the gear pump 120 is similar to that, already described, of the vane pump 20. When there is a difference in the relative rotational speeds between the first and second rotary shafts 143 and 149, the sun gear 124 and the two pinions 123 operate by forming the gear pump. When those of the lights 126, 127, 128 and 129 which serve as exit lights, are closed, the sun pinion 124 and the pinions 123 rotate integrally, so that the housing 122 can also rotate integrally with the sun pinion 124.

The hydraulic control circuit 121 is shown in FIG. 14.

With reference to this FIG. 14, the hydraulic control circuit 121 comprises a first passage of hydraulic oil 153 ensuring communication between the lights 126 and 128 of the gear pump 120 and communicating with a reservoir of hydraulic oil 152 by the through a first control valve 150, a second hydraulic oil passage 156 ensuring communication between the lights 127 and 129 of the pump 120 and communicating with the reservoir 152 by means of a second control valve 155, a first check valve 157 allowing the flow of hydraulic oil from the first pass 153 to the second pass 156 exclusively, and a second check valve 158 allowing the flow of hydraulic oil from the second pass 156 to the first passage 153 exclusively. Each of the first and second check valves 157 and 158 comprises an active valve element 162 resiliently biased by a compression spring 161 which controls the pressure with which the valve is biased in the open position. These first and second check valves 157 and 158 constitute a flow selector means and a flow control means.

it is obvious that the operation and the effect of the power transmission device comprising the gear pump 120 described above with reference to FIGS. 12 to 14, are equivalent to those of the power transmission device comprising the vane pump 20 .

The embodiments which will now be described aim to provide a compact structure for a hydraulic oil pump. A hydraulic oil pump unit 213, which will be described later, is provided with both a hydraulic oil pump and a hydraulic control circuit.

A tenth embodiment of the present invention will now be described with reference to FIGS. 15 to 24.

With reference to FIGS. 15 and 16, which show the structure of the four-wheel drive type vehicle to which the present invention is applied, a transmission mechanism 202 is connected to an engine 201 comprising a crankshaft extending in the transverse direction of the vehicle , and an output shaft 203 of the transmission mechanism 202 transmits the driving force to a return pinion 204. From this return pinion 204, the driving force is transmitted to a hydraulic oil pump unit 213 via a gear assembly 220a which is disposed on the periphery of the hydraulic oil pump unit 213. From this hydraulic oil pump unit 213, the driving force is transmitted to a second rotary shaft 214.The force drive is then transmitted, via a gear mechanism 215 which reverses the direction of power transmission, to a differential 217 associated with the rear wheels 216 in order to center said rear wheels 216 .

With reference to FIGS. 15 to 17, the hydraulic oil pump unit 213 comprises a hydraulic oil pump produced in the form of a vane pump 200 and an associated hydraulic control circuit 221.

The vane pump 200 comprises a rotor 219 and a cam ring 220.

The cam ring gear 220 is connected, via a first rotary shaft 211 and a differential 210, to the front wheels 209. The rotor 219 is coupled to the second rotary shaft 214 transmitting the driving force to the rear wheels 216.

With reference to FIG. 19, the rotor 219 of the vane pump, the operation of which is that of a hydraulic oil pump, comprises several vane grooves 219b (10 in number in the present embodiment) formed in an area peripheral 219a, the pallet grooves 219b being regularly spaced in the peripheral direction. Each of the grooves 219b is provided with a pallet 218 which is in contact with an internal surface 220d of the cam ring 220. The number of pallets 218, which is 10 in the present embodiment, is determined so as not to not form an integer multiple of the number of pump chambers 236a, 236b and 236c, which will be described later, i.e. it is determined by taking an integer multiple of the number of chambers plus one.

The vane pump 200 delivers a quantity of hydraulic oil proportional to its speed of rotation. More precisely, this vane pump 200 generates hydraulic pressure in the pump chambers 236a, 236b and 236c when there appears a relative rotation between the rotor 219 and the cam ring 220, that is to say when it a relative rotation appears between the first rotary shaft 211 and the second rotary shaft 214. The operation of this vane pump 200 is such that, when the outlet lights (lights 222a, 223a and 224a or lights 222b, 223b and 224b which are head lights in the direction of relative rotation between the cam ring 220 and the vanes 218) of the vane pump 220 are closed, the rotor 219 and the cam ring 220 are made to rotate integrally, forming a rigid body, under the action of the static pressure of hydraulic oil.

For this purpose, the inner peripheral surface of the cam ring 220 is shaped so as to have a triangle shape to form three pump chambers 236a, 236b and 236c near the vertices of a triangle inscribed between the inner peripheral surface 220d of the cam ring and the rotor 219. Six lumens 222a, 222b, 223a, 223b, 224a and 224b are arranged at the two ends of the chambers 236a, 236b and 236c so that those of them which are in the tail in the direction of relative rotation have a role of suction lights and that of them which are at the head in the direction of relative rotation have a role of exit lights. The lights 222a, 223a and 224a, which have the same function, communicate with each other through a second passage of hydraulic oil 227. Likewise, the lights 222b, 223b and 224b, which have the same function, communicate with each other via a first passage of hydraulic oil 226. A check valve 223 is arranged between the first passage 226 and the second passage 226., to allow the flow of the hydraulic oil from the first passage 226 to the second passage 227 when the pressure of the hydraulic oil exceeds a predetermined value. A check valve 231 is also provided to allow the flow of hydraulic oil from the second passage 227 to the first passage 226 when the pressure of the hydraulic oil exceeds a predetermined value. The first and second passages 226 and 227 communicate with a hydraulic oil tank 230 via control valves 228 and 229 which allow the flow of 1 hydraulic oil from the tank 230 exclusively.

Thanks to such an arrangement of the hydraulic control circuit 221, the pressure of the hydraulic oil supplied by the vane pump 200 is always exerted on the active elements of the check valves 231 and 233, and the reservoir 230 communicates with the lights suction of the vane pump 200 regardless of the relative direction of rotation of the rotor 219 and the crown ~ 220.

The hydraulic oil pump unit 213 has the structure shown in FIG. 15 and is arranged below the transmission mechanism 202.

The structure of the vane pump 200 will now be described in detail.

The vane pump 200 comprises the rotor 219 and the cam ring 220, a cover 251 engaging the end surfaces of the rotor 219 and the cam ring 220, a pressure retaining member 241 engaging the other end surfaces of the rotor 219 and of the cam crown 220, and a flange 245 fixed by means of a bolt 248 integrally with the cover 251 and the retaining member 241, to the cam crown 220 for transmitting the driving force of the crown gear 220.

A pump body of the vane pump 200 comprises the cam ring 220, the retaining member 241, the cover 251 and the flange 245. A housing of the pump body comprises the retaining member 241, the cover 251 and the collar 245.

Thus, the rotor 219 is disposed inside a space defined by the cam ring 220, the cover 251 and the retaining member 241, so that the two end surfaces of the rotor 219 are in abutment against the surfaces of the cover 251 and of the retaining member 241.

The rotor 219 is connected, by means of a key 257, to a centering shaft 243 of the rear wheels, which is integrally connected to the second rotary shaft 214, to allow the rotational force of the rotor 219 d 'be transmitted in one direction or the other by the drive shaft 243 of the rear wheels. One end of the centering shaft 243 is inserted into a bore 251a formed in the center of the cover 251.

A bearing 252 is inserted between the drive shaft 243 and the bore 251a. This bearing 252 supports the drive shaft 243, so as to allow its relative rotation relative to the cover 251 and to ensure the liquid-tightness of the interior of the cover 251.

The cover 251 is mounted for rotation on a casing 292a of the transmission mechanism 202 by means of a bearing 259.

The drive shaft 243 of the rear wheels extends externally through a bore 241a formed in the center of the pressure retaining member 241. A bearing 256 is inserted between the bore 241a and the shaft d 'drive 243. The pad 256 supports the retaining member 241 so as to allow its relative rotation relative to the shaft 243 and to ensure the liquid-tightness of the interior of the retaining member 241.

The pressure retaining member 241 is rotatably mounted by means of the flange 245 fixed by means of a bolt 248 and bearings 260a and 260b to the casing 202a.

The cover 251, which comprises a part of the pump body, has an opening close to the center of rotation of the pump body (in the present embodiment, it is aligned with the center of the drive shaft 243 rear wheels) and an oil hole 251b extending in the direction of the center line of the rotary shaft towards the pump body.

Thus the oil hole 251b is constituted by a part, located more externally than the end of the drive shaft 243 of the rear wheels, of the bore 251a formed in the cover 251, and extending in the direction of the drive shaft 243.

A part more external than the end of the centering shaft 243 in the hole 251b is provided with a filter consisting of a mesh of nylon or similar material and a magnet 255 to remove all the foreign materials contained in hydraulic oil entering it from the end of the hole 251 a.

In the figures, the reference 242 represents a pulsation damper, and 246 a pulsation volume which reduces pulsations of the hydraulic oil delivered.

An outer peripheral part of the retaining member 241 extends outside the outer periphery of the rotor 219 and at the end of the pallet 218, up to the outer periphery of the cam ring 220. The outer peripheral part of the retaining member 241 is fixed and made integral with the cam crown 220 fixed by means of the flange 245, the cover 251 and the bolt 248.

When the vane pump 200 is installed inside the transmission mechanism 202, the outside diameter of the pump 200 must necessarily be as small as possible.

In order to compact the vane pump 200 while using, as previously, the bolt 248 fixing the flange 245 and the cam crown 220, it is necessary to offset the bolt 248 in the direction of the inner peripheral surface 220d of the crown to cam 220, as shown in Figure 19. However, this method is not preferable because, as indicated by the dashed lines in Figure 19, the engagement surface of the retainer 241 and the cam ring 220 is small and that the joint section of the flange 245 and the cam ring 220 defines a three-sided joint associated with the retaining member 241, which results in a reduced joint surface.

However, as described above, the joint surface can be increased and the parts in play can be supported stably in order to ensure sufficient lateral pressure, by extending the external peripheral zone of the member. retainer 241 to the outer peripheral zone of the cam ring 220 and by fixing the flange 245 and the cam ring 220 in an integral manner with the retainer 241. Thus, the vane pump 200 is produced under the shape of a compact pump with large capacity.

Referring to FIG. 19, the outer peripheral zone of the cam ring 220, which constitutes a part of the pump body, has a gear zone 220a which engages with a return pinion 204, as shown in FIG. 16.

Thus the rotational force (the driving force of the front wheels or that of the front and rear wheels) is transmitted to the pump body via the return pinion 204 and the gear forming area 220a. The cam crown 22C is made of an abrasion-resistant material, such as case-hardened steel, in order to minimize wear by friction with the pallets 218.

The idler gear 204 engaging with the gear area 220a is also made of the same material, such as cemented steel, and the gear 220a is protected from wear due to its engagement with the idler gear 204, which is an advantageous solution compared to a case in which the joint-forming zone is made of another material.

The gear zone 220a and the cam ring 220 must be made concentrically. This can be easily done by machining the pinions, taking the inner peripheral surface 220d of the cam ring 220 as a base.

Thus, several constituent parts can be eliminated, by using the cam ring 220 associated with the gear forming zone 220a as the joint forming zone.

Although the gear area 220a poses a slight problem in that it must be made of an abrasion resistant material, it can be formed on the outer peripheral area of the cover 251 or the retainer 241 .

The cover 251 and the retaining member 241 are made of gray cast iron or of a sintered material, and pose a problem as to their resistance to abrasion. If the cover 251 and the retaining member 241 are made of case-hardened steel or similar material, there is a problem of seizure resistance on the sliding surface receiving the rotor 219.

However, the above-mentioned problems can be eliminated by making the cover 251 and the retaining member 241 from an abrasion-resistant material such as case-hardened steel, and by treating the sliding surface receiving the rotor by means of a surface treatment intended to improve lubrication, and the cover 251 or the retaining member 241 can be provided with the outer peripheral pinion zone 220a to serve as a zone for transmitting the driving force.

Thus the hydraulic oil pump unit 213 is produced in a simple and compact form.

On the other hand, the rotor 219 is fixed, by means of a key 257, to the drive shaft 243 of the rear wheels, to transmit the motive force received or delivered by the rotor 219 by means of the key 257.

The key 257 has a shape shorter than an axial width of the rotor 219 at the entry point d of the axial direction of the rotor 219, so that the driving force of the key 257 and of the rotor 219 is transmitted without causing tightening of the rotor 219 with the retaining member 241 or the cover 251.

In the case of a vane pump provided with a rotor 219 having a greater width, it will be difficult to fix the rotor 219 in a perpendicular position to be specified on the drive shaft 243 of the rear wheels, and a seizure may occur between the terminal surface of the rotor 219 and the retaining member 241 or the cover 251. This is considered to be due to the fact that, with the rotor 219 being mounted obliquely, a localized force is exerted on part of the terminal surfaces of the rotor 219 and of the retaining member 241 or of the cover 251, which results in a solution of continuity of the oil film.

This phenomenon can be avoided by reducing the width of the key 257.

With reference to FIGS. 21 (a) to 21 (c), it is assumed that the rotor 219 is mounted obliquely on the centering shaft 243 of the rear wheels with a minimum clearance between the rotor 219 and the retaining member 241 and the cover 251 in the direction of the rotary shaft of the rotor 219.

When the rotor 219 and the drive shaft 243 are in rotation, a torque is transmitted to the rotor 219 via the points A and A 'of the key 257. A force exerted on the points A and A' acts as an unbalanced force on the rotor 219 which then exerts a strong thrust on the retaining member 241 and the cover 251 at points B and
B ', which results in a solution of continuity of the oil film and in seizure.

In Fig. 21 (a), the solid lines denote the engagement conditions in a case in which the rotor 219 is mounted obliquely on the drive shaft 243 of the rear wheels, and the dashed lines denote the arrival conditions engaged in a case in which the rotor 219 is not mounted obliquely on the shaft 243. If the mounting is not oblique, there is a linear contact between the rotor 219 and the key 257. However, in the case of an oblique assembly, there will be a point contact at points A and A>
By reducing a longitudinal width "l" of the key 257, the points
A and A 'are brought closer to the center of this width, thereby reducing the imbalance, and seizure is avoided.

With reference to FIG. 22, when the rotor 219 is connected to the key 257 in such a way that the key 257 is offset by a length "a" from the transverse center of the rotary shaft of the rotor 219, an unbalanced force F is applied, this force F then constituting a moment which causes the inclination of the rotor 219 due to the offset of the longitudinal center of the key 257.

This problem can be resolved by reducing the offset "a" of the key 257 to zero to eliminate the moment causing the inclination of the rotor 219, thus preventing the rotor 219 from seizing up with the retaining member 241 and the cover 251.

The aforementioned unbalanced force F is also caused by an imbalance and eccentricity of the rotor 219 itself, or by an unbalanced force applied to the rotor 219 under the effect of a leakage of hydraulic oil.

For the reasons mentioned above, the key 257 is provided with a shorter length and is disposed in the transverse center of the rotary shaft of the rotor 219, thus preventing the rotor 219 from seizing with the retaining member 241 and cover 251.

With reference to FIGS. 23 (a) to 23 (c), the point of engagement of the key 57 and of the rotor 219 can also be brought closer to the transverse center of the rotary shaft of the rotor 219 by giving the rotor 219 a curved shape in the transverse direction of the rotary shaft.

Thus, by giving the key 257 a shape such that part of its longitudinal center is convex, as shown in FIGS. 23 (b) and 23 (c), the points of engagement A and A 'of the key 257 and the rotor 219 are brought closer to the transverse center of the rotary shaft of the rotor 219, thereby reducing the pressure exerted on the retaining member 241 and the cover 251 due to an inclination of the rotor 219 on the drive shaft. rear wheel drive 243.

Likewise, the method described above allows the convex part located at the longitudinal center of the key 257 to engage over its entire length with the rotor 219, thus improving the distribution of the pressure on the surface of the key 257.

The key 257 is machined in the aforementioned form by means of a machine tool bearing the name "ROTEFLOX". The key 257 can also be machined in the aforementioned form by means of a specific milling machine with developing lead screw, by adjusting its advance rate appropriately.

The aforementioned bending of the key 257 can be applied to the key 257, whether or not the latter has been made shorter as described above.

The hydraulic control circuit will now be described in detail.

With reference to FIG. 15, a casing 202a of the transmission 202 leaving opposite the bore 251a is provided with an oil guide 253.

This oil guide 253 is produced in the form of a cube divided into two halves along one of its diagonals, one end of which extends inside the bore 251a, and collects the hydraulic oil s' flowing along an inner wall of the casing 202a to introduce the oil thus collected inside the bore 251a.

The cover 251 is provided with a hydraulic oil supply line 235 extending between the end of the drive shaft 243 of the rear wheels up to the bore 251a, and disposed obliquely in the direction of the outer periphery of the cover 251 on the side of rotor 219, thus establishing communication with the hydraulic control circuit 221.

A zone forming a seal of this oil supply line 235 with the hydraulic control circuit 21 is provided with a communication passage 235a extending between the end of the supply line 235 and the bore 251a and opening at level of a pad 251 in the bore 251a, and is also provided with a spherical active element of valve 229b at the end of the communication passage 235a, to form a control valve 229.

The configuration described above allows hydraulic oil to be introduced, via a mounting zone of the bearing 252 of the bore 251a, from the supply line 235 to the cam ring 220, and, at the same time, prevents the hydraulic oil from flowing back from the crown 220.

There is also a hydraulic oil supply line 235 ', a communication passage 235'a and a control valve 228 at another point in the bore 251 a, the construction being identical to that of the supply line 235 , the communication passage 235a and the control valve 229 (see Figures 17 and 20).

In the cover 251, the hydraulic control circuit 221 is produced in the form shown in part (a) (area of the cover) and part (b) (view in vertical section) of FIG. 20.

A first hydraulic oil passage 226 is provided which connects the lights 222b, 223b and 224b having simultaneously the role of outlet or suction lights as a function of the relative direction of rotation of the rotor 219 and of the cam ring 220, an end zone 233a of the check valve 233 and an end zone 231b of the check valve 231, and which communicates, via the communication passage 235a and the supply line 235, with the bore 251a.

The circuit also comprises the hydraulic oil supply line 235 'and the communication 235'a which communicates with the light 222a and the bore 251a.

On the other hand, the retaining member 241 is provided with the second hydraulic oil passage 227 which connects the ports 222a, 223a and 224a, an end zone 233b of the check valve 233 and an end zone 231a of the check valve 231, as shown in part (c) (area of the retaining member) and part (b) of Figure 20. The second passage 227 communicates, via the light 222a and the pump chamber 236a formed in the cam ring 220, with the communication passage 235'a, the hydraulic oil supply line 235 'and the bore 251 a.

The first and second passages 226 and 227 are connected by means of the retaining valves 231 and 233 which extend from the cover 251 to the retaining member 241 by passing through the cam ring 220.

In the check valve 231, a spherical active element disposed in the end zone 231a is subjected to the force of a spring 232, thus allowing the flow of hydraulic oil from the second passage 227 to the first passage 226 exclusively when the hydraulic oil pressure exceeds a predetermined value.

In the check valve 233, a spherical active element disposed in the end zone 233a is subjected to the force of a spring 234, thus allowing the flow of hydraulic oil from the first passage 226 to the second passage 227 exclusively when the hydraulic oil pressure exceeds a predetermined value.

The configuration described above constitutes a hydraulic circuit as shown in FIG. 17.

Hydraulic oil is introduced into the lower part of the casing 202a of the transmission mechanism 202 until the lower half of the hydraulic oil pump 213 is filled. The hydraulic oil is sprayed by the rotation of the cam crown 220 to lubricate the parts, and is introduced into the vane pump 200 by means of the bore 251a.

Thus, hydraulic oil flows over the surface of a wall of the casing 202a, reaches an oil guide 253 where it is collected, and is introduced into the bore 251a from the end of the oil guide. 253.

Hydraulic oil is also introduced, via the supply lines 235 and 235 'and the communication passages 235a and 235'a, into the first passage 226, the light 222a and the bearing 252.

Thus, the hydraulic oil is collected by the oil guide 253, thus allowing the bore 251a to always be supplied with hydraulic oil even if the oil level is below the bore 251a.

The supply lines 235 and 235 ′ are inclined, to allow the introduction of a sufficient quantity of hydraulic oil by means of a centrifugal force into the first passage of hydraulic oil 226 included in the hydraulic control circuit 221 .

When the hole 251a rotates, hydraulic oil is drawn in under the effect of centrifugal force, thus preventing air from being brought into the central zone of the end of the hole 251a and coming into contact with hydraulic oil.

Then, hydraulic oil is introduced into the first oil passage 226, the lumen 222a and the key 257, its reflux being prevented by the operation of the control valves 228 and 229.

The hydraulic oil is guided over the entire length of the first passage 226 disposed in the cover 251, up to the ports 222b, 223b and 224b and the end zone 233a of the check valve 233, out via the port 222a and from the second passage 227, up to the lumens 223a and 224a and the end zone 231a of the check valve 231.

The hydraulic oil from the first passage 226 or the second passage 227 is pressurized in the pump chambers 236a, 236b and 236c by means of the relative rotation of the rotor 219 and the cam ring 220, and is delivered respectively to the second pass 227 or the first pass 226.

The second hydraulic oil passage 227 comprises an annular groove formed in the external peripheral surface of the retaining member 241 and the internal peripheral surface of the flange 245 engaging with the external periphery of the retaining member 241.

Until now, such a passage of hydraulic oil has been carried out by implementing a difficult machining process, according to which very long holes are drilled in different directions, said holes thus drilled then being connected and plugs then being inserted into these holes. However, the configuration of the present invention as described above makes it easier to carry out the passage of hydraulic oil.

The number of vanes 218 is determined as an integer which is not an integer multiple of the number of pump chambers 236a, 236b and 236c, and results in different phases of variation of the oil pressure hydraulics delivered by each light, as shown in figure 27. (see figures 27 (b), 27 (c) and 27 (d).) Consequently, there is no increase in the pulsation of the oil pressure in the first passage 226 or the second passage 227 where these pressures P1, P2 and P3 of the hydraulic oil delivered are combined.

Thus, with reference to FIG. 27 (a), a variation H2 due to a pulsation of the oil pressure Po in the first passage 226 or in the second passage 227 will be suppressed to a level lower than each variation H2 of the pressures P1, P2 and P3 of the hydraulic oil delivered.

A hydraulic valve lifting circuit intended to slide the pallets 218 will now be described.

With reference to FIGS. 24 (a), 24 (b) and 24 (c), the retaining member 241, the cover 251 and the rotor 219 constitute a hydraulic circuit 250 for lifting the pallet.

The rotor 219 is provided with a paulette lifting oil passage 250a, which passes through the rotor 219 at its two sides at the level of an end zone of a pallet groove 219 in which the pallet 218 slides. By introducing hydraulic oil pressure in the passage 250a of hydraulic lifting oil, the pallet 218 is pushed upwards and the end of the pallet 218 comes to bear on the inner peripheral surface of the cam ring 220, thus ensuring the increase in pressure in the chambers 236a, 236b, and 236c of the pump.

The passage 250a of lifting oil is connected to annular recesses 250b and 250c arranged at the two ends of the rotor 219, and the retaining member 241 is provided with a communication passage 250d opening at an opposite point of the recess ring finger 250b.

The cover 251 is provided with a communication passage 250e opening at an opposite point of the annular recess 250c.

The communication passages 250d and 250e are arranged so as to incline relative to the annular recesses 250b and 250c in the direction of the second and first passages of hydraulic oil 227 and 226, respectively by means of valves. 250f and 250g control.

The control valves 250f and 250g containing spherical valve elements and the communication passages 250d and 250e extending to the outer periphery, each of the valve elements is supported on the corresponding valve seat, by means of a centrifugal force generated by the rotation of the retaining member 241 and the cover 251, and these control valves are normally kept closed.

This ensures the introduction of hydraulic pressure from the first or second passage of hydraulic oil 226 or 227, whichever is higher, into the passage 250a of lifting oil.

Thus one of the two passages 226 or 227 becomes the suction side, while the other becomes the outlet side, depending on the relative direction of rotation of the cam ring 220 and the rotor 219. However, the passage 250a is always supplied by means of a hydraulic oil pressure coming from the first passage 226 or from the second passage 227, and the pallet 218 is always in abutment against the inner peripheral surface 220d of the crown 220, that the fastest rotation that of the front wheels or that of the rear wheels. An elastic member, such as a spring, can be inserted in the end region of the groove 219b in which the pallet 218 is placed.

The operation of the power transmission device thus formed will now be described in detail.

In the usual state of rectilinear forward driving of the vehicle, the effective radius of the tires of the front wheels 209 is the same as that of the rear wheels 216, and the slip rate of the rotating tires is very small. In such a state, there does not appear to be a difference between the speeds of rotation of the first rotary shaft 211 and of the second rotary shaft 214 connected to the hydraulic oil pump unit 213. Consequently the vane pump 200 does not deliver hydraulic oil under pressure and no driving force is transmitted to the rear wheels 216. Thus the vehicle is driven solely by the front wheels 209; that is, the vehicle is running in the front-wheel drive mode.

However, when the vehicle in straight forward direction undergoes, for example, an acceleration, a slip less than about 1% generally occurs on the front wheels 209, although it is not appreciable. A difference then appears between the rotational speeds of the first and second rotary shafts 211 and 214, due to the aforementioned slippage of the front wheels 209.

In such a case, the vane pump 200 is activated in order to accumulate the pressure corresponding to said difference in rotation speeds. The rotor 219 and the cam crown 220 rotate in solidarity with one another, and the driving force corresponding to the accumulated pressure and to the receiving surface of the pallets is then transmitted to the rear wheels 216 to establish the traction mode at four wheel drive.

The flow of hydraulic oil in the vane pump 200 is shown, for such a situation, in Figure 18 (a). It will be seen in this figure 18 (a) that, due to the rotation of the rotor 219 relative to the crown 220, the lights 222b, 223b and 224b then have a role of suction lights, the hydraulic oil is sucked in the reservoir 230 for entering the suction lights 222b, 223b and 224c via the control valve 229. On the other hand, the lights 222a, 223a and 224a have the role of exit lights. Consequently, hydraulic oil is introduced into the first passage 226 and, at the same time supplied into the second hydraulic passage 227 in the direction of the check valve 231.

In FIG. 18 (a), the solid arrowed lines denote the directions of flow of the hydraulic oil delivered, and the dashed dashed lines denote the directions of flow of the hydraulic oil drawn.

It is therefore assumed that the speed of rotation of the front wheels 209 becomes very high compared to that of the rear wheels 216, for example, when the vehicle is traveling on a snow-covered wheel or is subjected to sudden acceleration or sudden braking resulting in blockage of the rear wheels 216. In such a case, the difference between the rotational speeds of the first and second rotary shafts 211 and 214 connected to the hydraulic oil pump unit 213 becomes very large.

As a result, high pressure is generated in the vane pump 200, and hydraulic oil flows in the directions shown in Figure 18 (a). When the hydraulic oil pressure exceeds a predetermined level, the check valve 231 is opened against the force of the spring 234, and the pressure of the delivered oil is controlled so as to remain substantially constant. Thus, a constant driving force corresponding to the regulated pressure of the hydraulic oil delivered, is transmitted to the rear wheels 216 in order to establish the traction mode with four driving wheels.

Consequently, the rotational speed of the front wheels 209 decreases, while that of the rear wheels 216 is increased, so that the difference between the rotational speeds of the front and rear wheels 209 and 216 is decreased. (This function is the same as that of the slip-free differential). Thus, when a slip occurs on the front wheels 209, the engine torque intended for the rear wheels 216 is increased in order to avoid being unable to turn, while, when the rear wheels 216 tend to lock, the braking torque for the front wheels 209 is increased in order to prevent the rear wheels 216 from locking.

It is also supposed to be a case in which the speed of rotation of the rear wheels 216 is very high compared to that of the front wheels 209, such as, for example, when the front wheels 209 tend to lock after actuation brake. In such a case, a very large difference appears between the rotational speeds of the first and second rotary shafts 211 and 214 connected to the hydraulic oil pump unit 213, in the opposite direction to that mentioned previously.

As a result, the flow of hydraulic oil into the vane pump 200 is now in a direction opposite to that shown in Figure 18 (a). It will be seen in FIG. 18 (b) that the lights-222a, 223a and 224a then have a role of suction lights with respect to the hydraulic oil and that hydraulic oil is sucked in from the reservoir 230 to enter the suction lights 222a, 223a and 224a via the control valve 228, while, on the other hand, the lights 222b, 223b and 224b have the role of exit lights.

Consequently, hydraulic oil is introduced into the second passage 227 and, at the same time, delivered by the first passage 226 in the direction of the check valve 233. The pressure of this hydraulic oil is also kept at a constant level by the check valve 233, the corresponding constant driving force is transmitted to the rear wheels 216 in order to establish the four-wheel drive traction mode.

Consequently, the braking torque communicated to the rear wheels 216 is increased in order to prevent a blocking of the front wheels 209.

In the usual case where the vehicle crosses a curve, the speed of rotation of the front wheels 209 is slightly higher than that of the rear wheels 216, and the vehicle crosses the curve in the four-wheel drive mode of traction, in which the torque of braking is communicated to the front wheels 209 while the engine torque is communicated to the rear wheels 216.

According to the process described above, the outlet pressure of the hydraulic oil is controlled by the check valves 231 and 233 of the pump unit 213 so as not to exceed a constant value. Consequently, and contrary to the situation of the prior art in which manipulation is required of the driver to establish the four-wheel drive mode of traction in a vehicle of the four-wheel drive type intermittently, in accordance with the present invention, the switching between four-wheel drive and front-wheel drive modes can be performed automatically, and the four-wheel drive mode is established by means of the driving force corresponding to the difference between the rotational speeds of the front wheels and back.

Likewise, the hydraulic oil pump 213 in accordance with the present invention is small, has a compact structure and is light in weight and of low cost, compared with the central differential necessarily fitted to a vehicle according to the prior art. of the constantly driving four-wheel type.

FIG. 25 shows an eleventh embodiment of the present invention, in which the structure of the hydraulic oil supply lines 235 and 235 ′ is different from that used in the tenth embodiment of the present invention.

According to this eleventh embodiment of the present invention, the supply lines 235 and 235 ′ extend obliquely in the direction of the outer periphery of the rotor 219 in the cover 251 until the first passage of hydraulic oil 226. Consequently , a centrifugal force is generated by the rotation of the cover 251, and a sufficient quantity of hydraulic oil is introduced into the first passage 226 and the lumen 222a. This structure being identical to that of a centrifugal separator, air is collected at the center of the hole 251a, and is not sucked into the supply lines 235 and 235 '
A twelfth embodiment of the present invention, in which the structure of the hydraulic oil supply lines 235 and 235 ′ is different from that used in the tenth embodiment, will now be described with reference to FIG. 26.

In the embodiment shown in FIG. 26, the end zones of the center wheel shaft 243 of the rear wheels is provided with a supply line 235 ", which opens into this end zone and extends up to the location of the first hydraulic oil passage 226, and, at the same time, a communication passage 235 "a which places the supply line 235" in communication with the first passage 226. As a result, a centrifugal force is generated by the rotation of the drive shaft 243 of the rear wheels, and a sufficient quantity of hydraulic oil is introduced into the first passage 226.

The bearing 252 does not extend until the opening of the communication passage 235 "a on the peripheral surface of the drive shaft 243 of the rear wheels, and a hydraulic oil passage having the form of a ring is formed between the key 257 and the end surface of the bearing 252 which allows the introduction of hydraulic oil into the lumen 222a in addition to the first passage 226. The key 257 and the bearing 252 are also supplied with hydraulic oil serving as lubricant .

A thirteenth embodiment of the present invention, in which the structure of the pressure retaining member 241 is different from that used in the tenth embodiment, will now be described with reference to Figures 28 to 31. In these figures 28 to 31, the parts identical to those already described are designated by identical references and are not described again in the following.

The centering shaft 243 of the rear wheels extends through a bore 345b disposed in the center of a flange 345 serving as a retaining plate. A pressure retaining member 341 serving as a retaining plate is disposed between the inner wall surface of the rotor side 219 of the flange 345 and the end surface of the rotor 219. The pressure retaining member 341 includes a large area diameter 341b, the diameter of which is slightly larger than that of the rotor 219, and a small diameter area 341c, the diameter of which is slightly smaller than that of a hole 345b in the collar 345.

The retaining member 341 is arranged on the flange 345 in such a way that the large diameter area 341b is inserted into a recess 345a made in the flange 345, and that the small diameter area 341c is inserted into the bore 345b . An O-ring 358c is inserted between the large diameter area 341b of the retaining member 341 and the recess 345a of the flange 345, and an O-ring 358b is inserted between the small diameter area 341c and the bore 345b.

In addition, the retaining member 341 is disposed on the centering shaft 243 of the rear wheels so that a bearing 256 disposed between the bore 341a and the drive shaft 243 of the rear wheels allows sliding of the retaining member 341 in the axial direction of the drive shaft 243 and a rotation relative to the drive shaft 243.

Thus the retaining member 341 is displaceable in the axial direction of the centering shaft 243 of the rear wheels, relative to the shaft 243 and to the flange 345, and the end surface rotor side 219 of the retaining member 341 can project into the inner peripheral surface of the cam ring 220. Similarly, the retaining member 341 can rotate integrally with the flange 345 which constitutes a part of the pump body, and a hydraulic oil chamber 361 is formed by the inner wall surface of the recess 345a and the end surface of the large diameter area 341b, on the side of the small diameter area 341c, in the retaining member 341.

The internal pressure of the chamber 361 is maintained by means of the O-rings 358c and 358b, and that of the pump chambers 236a, 236b and 236c is maintained by means of the retaining member 341 and the O-ring 358c. As shown in FIG. 29, the retaining member 341 is provided with a hydraulic oil passage 350h intended to regulate the lateral clearance, this passage 350h allowing the passage 250a of oil for lifting pallets to be placed in communication. with hydraulic oil chamber 361.

This allows the introduction into the chamber 361 of the pressure of the hydraulic oil delivered by the chambers 236a, 236b and 236c to the passage 250a of the pallet lifting oil.

The pressure of the hydraulic oil supplied by the vane pump 200 constitutes the pressure necessary to keep the retaining member in abutment against the rotor 219, and the retaining member 341 slides in the axial direction of the shaft. entrainment 243 of the rear wheels in order to automatically adjust the lateral clearance between the end surface of the retaining member 341 and that of the rotor 219. The hydraulic oil chamber 361 is provided with a spring 362 which maintains a necessary lateral clearance even when there is no relative rotation between the rotor 219 and the cam ring 220 in the vane pump 200 and the pressure of the hydraulic oil delivered by the chambers 236a, 236b and 236c is at a low level , thus maintaining the pressure of the hydraulic oil from the chambers 236a, 236b and 236c when the vane pump 200 starts to rotate.

In addition, as shown in FIGS. 30 and 31, a chamfer 219c is formed on the inner radial edge of the paddle groove 219b of the rotor 219. Thus the chamfer 219c prevents it from a solution of continuity of the film. oil on the edge of the vane groove 219b, which would cause seizure between the end surfaces of the rotor 219 and of the retaining member 341. The chamfer 219c extends internally to the internal periphery of the rotor 219 and does not reach not the lights 222a, 222b, 222c, 223a, 223b and 223c defined by the flange 345. This is intended to prevent a leakage of the hydraulic oil pressure coming from the passage 250a of pallet lifting oil towards the lights d suction, lights 222a, 222b and 222c or lights 223a, 223b and 223c.

The vane pump used in the power transmission device according to the present invention being driven by the difference between the rotational speeds of the rotor and the cam ring, its range of rotational speeds is lower than that of vane pumps conventional, and, at the same time, it must have a compact structure and have a high torque transmission capacity. Consequently, the vane pump used in the present invention must necessarily include as many vanes as possible and many lights. However, if for the purpose of using a conventional vane pump in the context of the present invention , the pallet lifting value is simply increased, a bending moment will be applied to the pallets due to a high pressure of hydraulic oil exerted on the lateral surface in the direction of rotation of the pallets, and the pallets can then be tilted, which would result in temporary sticking to the rotor. In such a case, when the rotor turns eccentrically due to a variation in the pres
the hydraulic oil delivered or due to a backlash generated during
from the manufacture or assembly of the pump, a defect appears
of operation consisting of the formation of a gap between the ex-
end of the pallet and the inner peripheral surface of the crown
cam, which results in an oil leak or a drop in pressure
hydraulic oil. This defect is particularly noticeable in
the case of the vane pump used in the present invention, which
works in a lower speed range.

In order to solve the above-mentioned problem, a preferred form of a
vane pump, intended to be used in the power transmission device according to the present invention, has been defined experimentally, this preferred form now being described with reference to FIG. 32 which shows a sectional view in a plane perpendicular to the rotary shaft of a vane pump 200.

It is assumed that a pallet 218 is subjected to a hydraulic oil pressure
lique P delivered by the vane pump 200 in the form of a force of
lifting of the pallet, t designating the thickness of the pallet 218, 21 designating the length of the pallet 218 projecting from a rotor 219 at the time of its maximum lifting, t2 designating the length of the part of the pallet 218 inserted in a groove 219b in the previous situation, etj (Q, + Q2) being the total length of the pallet 218. In addition, in the protruding end of the pallet 218, a part slightly offset towards a surface 1 in the direction of its rotation is the most projecting, and this part is designated by the term of sliding zone
S. The sliding zone S is positioned at a distance t1 (t1 <t2) from surface A.

When the vane pump passes from the suction process to the outlet process, the pallet 218 is applied in contact with the rotor 219 at points a and b with a force Fo of the hydraulic oil delivered, and is then subjected to a moment bending. Even if the rotor turns
eccentric, a relative slip must appear between the pallet 218 and the rotor 219 at points a and b to maintain close contact between
the protruding end of the pallet 218 and a peripheral surface in
220d of a cam crown 220. Such a condition is defined
in the description below.

First of all, a force exerted on the pallet 218 is determined.

The equilibrium of the moment with respect to point a is given by the following equation:
t1 / 2.Fo P2.2 =
The bending moment due to the pressure P of the oil delivered is given by the following equation:
Fo = 21.P --------------------------------------------- ----- (2)
The lifting force of the pallet 218 is:
ql = tP (3)
The equilibrium equation in the direction of rotation is: Fo + F2 = F1 -------------------------------- -------- - (4)
The equilibrium equation in the radial direction is:
t1.P + (F1 + F2) = ql ------------------------------------- (5 ) (5)
We derive equations (1) to (4): F2 = # 1/2. 2.P ---------------------------------------- (6)
F1 =. (L + + l / 2.2) .P ------------------------------------ ( 7)
By substitution of equation (5) in equations (6) and (7), we have :: t1 - t2 = ffi 1 / # 2. #
The condition for the appearance of a slip between the rotor 219 and the pallet 218 being in the left-right direction in equation (5),
t1 - t2> c. # 1 / # 2 ---------------------------------- (9)
(where c is a critical coefficient of friction) and the shape of the pallet 218 can be determined so as to satisfy equation (9) above.

The rotor 219 is generally made of nickel-chromium steel (SNCN) and the pallet 218 is made of high-speed steel (such as SKH9) or the like. The critical coefficient of friction c of these same materials being approximately 0.04 to 0.05, a typical value Xc = 0.045 has been used here, and the shape of the pallet 218 can be determined so as to satisfy the following formula :
ti - t2> 0.045.1 / 2.

In addition, the vane pump used in the present invention preferably being of the heavy lifting type, the shape of the vane 218 can be determined to satisfy the condition 2> 0.9, so that the length 1 of the lifting area is relatively long compared to the total length. (in conventional vane pumps, this length is approximately 2 2 C0.8).

Although a vane or gear pump has been described by way of example of the hydraulic oil pump preferably used in the embodiments of the present invention which have just been described, and whether it is , in the case of the balanced type vane pump, it is obvious that an unbalanced type vane pump comprising two lights acting alternately as outlet and suction lights, can also be used depending on the amount of motive power transmitted, and that a hydraulic oil pump of any other suitable type, such as a axial piston pump or a radial piston pump, can also be used. The only essential requirement is that the pump be '' a type which delivers a quantity of hydraulic oil corresponding to the difference in rotation speeds. The transmission can be either manual or automatic. Likewise, the control mode of the check valve is in no way limited to a slave control, and it may be any other suitable type of mechanical control.

The present invention is in no way limited to its application to a vehicle of the four-wheel drive type, and it can also be implemented for the transmission of power to the front and rear wheels of a vehicle of the six-wheel type driving. The present invention can also be implemented for the transmission of power to the front front wheels and to the rear wheels of a vehicle of the type with two front driving axles, and for the power transmission to the rear front wheels and to the rear wheels of a vehicle of the type with two rear axles powered.

It will be understood from the detailed description which has just been given, which present invention provides a power transmission device intended for a vehicle, said device comprising a first rotary shaft transmitting power to front wheels, a second rotary shaft transmitting power to the rear wheels, and a hydraulic oil pump connected between the first and second rotary shafts to be driven as a function of the difference between the rotational speeds of the first and second rotary shafts, thereby delivering a quantity of oil hydraulic corresponding to said difference between the rotational speeds. The hydraulic oil pump transmits the driving force by means of the static pressure of the hydraulic oil delivered to establish the four-wheel drive traction mode, and it includes lights suction and outlet automatically switched according to the direction of relative rotation of the first and second shafts rotary. Consequently, the four-wheel drive mode of traction can be established without requiring any manipulation on the part of the driver. Thus, the present invention makes it possible to eliminate the drawbacks such as the sudden cornering braking phenomenon which is observed with vehicles of the four-wheel drive type intermittently, while simultaneously eliminating any disturbance of the driver imposed by driving. . The power transmission device according to the present invention has, compared to the conventional central differential fitted to vehicles of the type with four-wheel drive constantly, the advantages of having small dimensions, low weight, simple construction and low cost.

Claims (44)

1. Power transmission device for a vehicle, characterized in that it comprises a first rotary shaft (11; 143; 211) transmitting the driving force to the front wheels (9; 209), a second rotary shaft (14 ; 149; 214) transmitting the driving force to the rear wheels (16; 216), a hydraulic oil pump ~ (20; 120; 200) connecting the first rotary shaft (11; 143; 211) to the second rotary shaft (14 ; 149; 214) and intended to be entrainée according to the difference between the rotational speeds of the first and second rotary shafts (11,14; 143,149; 211,214), thus delivering a quantity of hydraulic oil corresponding to the difference of rotational speeds, the hydraulic oil pump (20; 120; 200) having at least two lights (22-25; 126-129; 222a-224a, 222b-224b) alternately switched between the outlet side and the suction side, in function of the direction of relative rotation of the first and second rotary shafts (11,14; 1439149; 211,214), and a circuit of c hydraulic control (21; 121; 221) comprising a hydraulic oil passage (26,27; 153,156; 226,227) establishing communication between one of said two lights with the other and a hydraulic control means disposed in the hydraulic oil passage (26,27 ; 153,156; 226,227) to control the pressure of the hydraulic oil supplied by the hydraulic oil pump (20; 120; 200). 2. Power transmission device according to claim 1, characterized in that said first rotary shaft (11; 143; 211) is coupled to the engine (1; 201) of the vehicle so that said front wheels (9; 209) can be driven directly by said motor (1; 201). 3. Power transmission device according to claim 1, characterized in that said second rotary shaft (14; 149; 214) is coupled to the engine (1; 201) of the vehicle so that said rear wheels (16; 216) can be driven directly by said motor (1; 201). 4. Power transmission device according to claim 1, characterized in that said hydraulic oil passage disposed in said hydraulic control circuit (21; 121; 221) comprises a first hydraulic oil passage (26; 153; 226 ) communicating with a first of said lights (22,24; 126,128; 222a-224a) and comprising a first valve unit (28; 150; 228) allowing a flow of hydraulic oil inside said oil pump hydraulic (20; 120; 200) exclusively, a second hydraulic oil passage (27; 156; 227) communicating with the second of said lights (23,25; 127,129; 222b224b) and comprising a second valve unit (29; 155 ; 229) allowing a flow of hydraulic oil inside said hydraulic oil pump (20; 120; 200) exclusively, a first communication passage (35) capable of establishing communication between said second oil passage hydraulic (27; 156. 227) and the area of said first hydraulic oil passage (26; 153; 226) located between the first of said lights (22,24; 126,128; 222a-224a) and said first valve unit (28; 150; 228), and a second communication passage capable of establishing communication between said first passage hydraulic oil (26; 153; 226) and the area of said second hydraulic oil passage (27; 156; 227) located between the second of said lights (23.25; 127.129; 222b-224b) and said second valve unit ( 29; 155; 229), and further characterized in that said hydraulic control circuit (21; 121; 221) includes a selector member (31,32) for switching between said first communication passage and said second communication passage communication. 5. Power transmission device according to claim 1, characterized in that said hydraulic oil passage disposed in said hydraulic control circuit comprises a first hydraulic oil passage communicating at a first of its ends with a first of said lights and comprising a first valve unit allowing a flow of hydraulic oil inside said hydraulic oil pump exclusively, a second hydraulic oil passage communicating with the second of said lights and comprising a second valve unit allowing a flow of hydraulic oil inside said hydraulic oil pump exclusively, and a passage so as to establish communication between the second end of said first hydraulic oil passage and the second end of said second hydraulic oil passage, and characterized in in addition that said hydraulic control circuit comprises a selector member to selectively establish communication between the second end of said first hydraulic oil passage and the second end of said second hydraulic oil passage with said outlet passage. 6. Power transmission device according to claim 4 or 5, characterized in that said selector member operates so as to allow the hydraulic oil to be delivered by said hydraulic oil pump. 7. Power transmission device according to claim 4, characterized in that said hydraulic control member is disposed in each of said first and second communication passages. 8. Power transmission device according to claim 5, characterized in that said hydraulic control member is disposed in said outlet passage. 9. Power transmission device according to claim 1, characterized in that said hydraulic control member comprises an active valve element biased by an elastic biasing means. 10. Power transmission device according to claim 1, characterized in that said hydraulic control member is controlled as a function of the vehicle operating parameters. 11. Power transmission device according to claim 10, characterized in that said vehicle operating parameters are constituted by the speed of rotation of the vehicle engine. 12. Power transmission device according to claim 10, characterized in that said vehicle operating parameters are constituted by the speed of rotation of said first rotary shaft. 13. Power transmission device according to claim 10, characterized in that said vehicle operating parameters are constituted by the speed of rotation of said second rotary shaft. 14. Power transmission device according to claim 10, characterized in that said vehicle operating parameters are constituted by the opening of the butterfly. 15. Power transmission device according to claim 10, characterized in that said vehicle operating parameters are constituted by the degree of actuation of the brakes. 16. Power transmission device according to claim 10, characterized in that said vehicle operating parameters are constituted by information representative of the steering rate. 17. Power transmission device according to claim 1, characterized in that said hydraulic control member comprises a flow restriction means. 18. Power transmission device according to claim 17, characterized in that said flow restriction means is a nozzle. 19. Power transmission device according to claim 17, characterized in that said flow restriction means is controlled as a function of the vacuum prevailing in the intake manifold of the vehicle engine. 20. Power transmission device according to claim 17, characterized in that said flow restriction means is controlled as a function of the outlet pressure of an oil pump included in a power steering mechanism. 21. Power transmission device according to claim 17, characterized in that said flow restriction means is controlled both as a function of the vacuum prevailing in the intake manifold of the engine and of the outlet pressure of a oil pump included in a power steering mechanism. 22. Power transmission device according to claim 1, characterized in that said hydraulic control member comprises a valve element urged by an elastic urging means, and a flow restriction means. 23. Power transmission device according to claim 5, characterized in that said hydraulic control member comprises a flow restriction means arranged in an auxiliary passage ensuring communication of the area of said first hydraulic oil passage located between said first valve unit and a first of said lumens with the area of said second hydraulic oil passage located between said second valve unit and the second of said lumens, and a valve member disposed in said outlet passage and biased by means of elastic stress.  24. Power transmission device according to claim 1,  characterized in that said hydraulic oil pump is a  pallets (20; 200).  25. Power transmission device according to claim 24,  characterized in that said vane pump (200) comprises two or more  pump chambers (236a, 236b, 236c).  26. Power transmission device according to claim 25, characterized in that each of said pump chambers (236a, 236b, 236c) comprises at least two lights (222a-224a, 222b-224b). 27. Power transmission device according to claim 24, characterized in that the number of vanes (218) of said vane pump (200) is an integer which is not an integer multiple of the number of said pump chambers ( 236a, 236b, 236c). 28. Power transmission device according to claim 24, characterized in that said vane pump (200) comprises a pump body and a rotor (219) able to rotate inside said pump body. 29. Power transmission device according to claim 28, characterized in that said pump body comprises a cam ring (220) sliding at the same time as said pallet (218), and a housing (241, 251, 245) supporting said cam ring on its two faces. 30. Power transmission device according to claim 28, characterized in that said hydraulic control circuit is arranged inside said pump body. 31. Power transmission device according to claim 28, characterized in that said pallet (218) is produced in a form satisfying a condition such as tî - t2 A # c.} 2.e, in which is a maximum length of projection of said pallet relative to said rotor (219) in which it is slidably mounted, P is a length of the part of said pallet inserted in said rotor,} 1 + t 2) is a total length of said pallet, t is a thickness of said pallet, the position of the end of said sliding pallet bearing on an inner peripheral surface (220d) of said cam ring (220) being defined as a distance t1 from the surface on the direction side rotation of said pallet, and / Le being a critical coefficient of friction between said rotor and said pallet. 32. Power transmission device according to claim 29, characterized in that said rotor (219) is subjected, in the direction of its rotary shaft, to a force intended to maintain an appropriate clearance between said housing and said rotor. 33. Power transmission device according to claim 29, characterized in that said pump chamber (236a, 236b, 236c) is a space defined by said bolt-holder (241, 251, 245), said cam ring (220) and said rotor (219), and in that said hydraulic oil passage establishes communication through said cam ring between a first of said lights (222a-224a) formed in said housing on a first side of said pump chamber, and the other side of said pump chamber (236a, 236b, 236C). 34. Power transmission device according to claim 1, characterized in that said hydraulic oil pump (120) is a gear pump. 35. Power transmission device according to claim 34, characterized in that said gear pump (120) comprises a gear pump housing (122) capable of turning and coupled to said first rotary shaft (143), a sun gear ( 124) rotatably mounted in said housing (122) and coupled to said second rotary shaft (149), and a pinion (123) rotatably mounted in said housing and meshing with said sun gear. 36. Power transmission device according to claim 35, characterized in that said gear pump housing (122) is provided with said hydraulic control circuit (121). 37. Power transmission device according to claim 1, characterized in that said hydraulic oil pump 200) and said hydraulic control circuit (221) form integrally a hydraulic oil pump unit (213). 38. Power transmission device according to claim 37, characterized in that said hydraulic oil pump unit (213) is connected to an output shaft (203) of a transmission mechanism (202) of said engine (201) . 39. Power transmission device according to claim 38, characterized in that said hydraulic oil pump unit (213) is contained in the same casing (202a) as said transmission mechanism (202). 40. Power transmission device according to claim 39, characterized in that said hydraulic oil pump unit (213) is arranged below said transmission mechanism (202) and occupies a lower part of said casing (202a). 41. Power transmission device according to claim 38, characterized in that said transmission mechanism (202) is a manually actuated transmission mechanism, comprising an input shaft driven by said motor, and an output shaft (203 ) arranged approximately parallel to said input shaft. 42. Power transmission device according to claim 38, characterized in that said hydraulic pump (200) comprises a gear zone (220a) engaging by meshing with a pinion (204) disposed on said output shaft (203) said transmission mechanism (202). 43. Power transmission device according to claim 42, characterized in that said gear zone (220a) is arranged on an outer peripheral surface of said cam ring (220) of said vane pump (200). 44. Power transmission device according to claim 42, characterized in that said gear zone (125) is disposed on an outer peripheral surface of said gear pump housing (122) on which a pinion is mounted to rotate. engaging with said sun gear (124) of said gear pump (120).