GB2154522A - Power transmission apparatus for vehicles - Google Patents

Abstract

In a power transmission apparatus for use in a vehicle of the type in which its front and rear wheels are driven by the same engine (1), a first rotary shaft (11) drivingly connected to the front wheels (9) and a second rotary shaft (14) drivingly connected to the rear wheels (16) are coupled to each other through a hydraulic oil pump (13) which is driven according to the rotation speed difference between the first and second rotary shafts (11, 14) and delivers hydraulic oil of an amount corresponding to the rotation speed difference. The pressure of hydraulic oil delivered from the pump is controlled by hydraulic control means (21) so as to automatically change over between the two-wheel drive mode in which the front or rear wheels only are driven and the four-wheel drive mode in which both the front and rear wheels are driven. A relief value (33) controls the relative speeds and driving forces of the front and rear wheels. A computer (39) controls the relief valve (33), and receives inputs responsive to engine speed, speeds of shafts (11, 14), engine throttle opening, brake actuation, vehicle speed, and steering wheel angle. The hydraulic pump may be a vane pump or a gear pump. <IMAGE>

Description

SPECIFICATION Power transmission apparatus for vehicles Technical field ofthe invention This invention relates to a power transmission apparatus for a vehicle of the kind in which its front and rear wheels are driven by the same engine.

Technical background of the invention In a vehicle of 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 occurs a difference between the paths of rolling movement of the tires of the front and rear wheels during running along a curve. As a result, the tires tend to slip during running of the vehicle, and an excessive force tends to act upon the drive system of the vehicle. It is therefore necessary to provide effective means for preventing impartation of such an undesirable force to the drive system of the vehicle.

For this purpose, in a vehicle of full-time four-wheel drive type in which all of the four wheels are necessarily driven at the same time, there has been provided a third differential unit commonly called a center differential unit so as to permit transmission of the drive force of the engine to all the wheels even when a rotation speed difference occurs between a first rotary shaft transmitting the drive force to the front wheels and a second rotary shaft transmitting the drive force to the rear wheels. However, the vehicle of full-time four-wheel drive type is disadvantageous from the aspects of weight, size and cost when compared with a vehicle of part-time four-wheel drive type in which all of the four wheels are not necessarily driven at the same time.Also, owing to the capability of the differential rotation of the front and rear wheels, the four-wheel drive may not always be achieved when the four-wheel drive is required, and, in order to ensure the four-wheel drive, a differential locking mechanism is additionally required. Thus, the vehicle of full-time four-wheel drive type has been defective among others in that its power transmission system is complex in structure.

In connection with such a vehicle of full-time four-wheel drive type, a method is disclosed in Japanese Patent Application Laid-open No. 58-20521(1983). According to the disclosed method, the center differential unit is replaced by a wet type multi-disk clutch disposed in the drive coupling portion, and such a clutch is caused to slide during running of the vehicle along a curve thereby absorbing the rotation speed difference between the front and rear wheels. However, the proposed method has been defective among others in that the wet type multi-disk clutch tends to be damaged by the heat generated due to slippage and that the clutch has a limited torque transmission capacity.

The center differential unit is not usually provided in the vehicle of part-time four-wheel drive type.

Accordingly, the driver of the vehicle of this type must make necessary manipulation to drive the two wheels only among the four wheels when a tight corner braking phenomenon peculiar to the four-wheel drive may occur during running of the vehicle along a curve. Therefore, the vehicle of part-time four-wheel drive type has been defective in that the driver must make complicated manipulation for driving the vehicle.

Summary of the invention With a view to obviate or reduce the prior art defects of vehicles in which the front and rear wheels are driven by the same engine, it is a primary object of the present invention to provide a power transmission apparatus for a vehicle, which is small in size and light in weight.

In accordance with the present invention which attains the above object, there is provided a power transmission apparatus for a vehicle comprising a first rotary shaft transmitting drive force to front wheels, a second rotary shaft transmitting drive force to rear wheels, a hydraulic oil pump connected between the first rotary shaft and the second rotary shaft to be driven according to the difference between the rotation speeds of the first and second rotary shafts thereby delivering hydraulic oil of an amount corresponding to the rotation speed difference, the hydraulic oil pump having at least two ports alternately changed over between the delivery side and the suction side depending on the direction of relative rotation of the first and second rotary shafts, and a hydraulic control circuit including a hydraulic oil passage providing communication of one of the ports with the other and hydraulic control means disposed in the hydraulic oil passage for controlling the pressure of hydraulic oil delivered from the hydraulic oil pump.

Brief description of the drawings Figure 1 is a diagrammatic view showing schematically the structure of a vehicle of four-wheel drive type to which a power transmission apparatus of the present invention is applied.

Figure 2 is a schematic sectional view of a first embodiment of the power transmission apparatus according to the present invention.

Figure 3(a) and 3(b) illustrate the flowing directions of hydraulic oil in the first embodiment respectively.

Figure 4 is a schematic sectional view of a second embodiment of the power transmission apparatus according to the present invention.

Figure 5 is a schematic sectional view of a third embodiment of the power transmission apparatus according to the present invention.

Figure 6 is a schematic sectional view of a fourth embodiment of the power transmission apparatus according to the present invention.

Figure 7 is a graph showing the delivery pressure of the vane pump relative to the rotation speed difference between the front and rear wheels in the fourth embodiment.

Figure 8 is a schematic sectional view of part of the hydraulic control circuit in a fifth embodiment of the power transmission apparatus according to the present invention.

Figure 9 is a schematic sectional view of part of the hydraulic controly circuit in a sixth embodiment of the power transmission apparatus according to the present invention.

Figure 10 is a schematic sectional view of a seventh embodiment of the power transmission apparatus according to the present invention.

Figure 11 is a diagrammatic view showing the structure of another vehicle of four-wheel drive type to which a power transmission apparatus of 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 apparatus according to the present invention Figure 13 is a sectional view taken along the line A-A in Figure 12.

Figure 14 is a diagrammatic view showing schematically the structure of the hydraulic control circuit in the ninth embodiment.

Figures 15 through 24 relate to a tenth embodiment of the present invention.

Figure 15 is a vertical sectional view showing part of the hydraulic oil pump.

Figure 16 is a diagrammatic view showing schematically the structure of a vehicle of four-wheel drive type to which a power transmission apparatus of the present invention is applied.

Figure 17 is a schematic view showing part of the hydraulic oil pump.

Figures 18(a) and 18{by are schematics showing operation principles.

Figure 19 is a sectional view taken along the line V-V in Figure 15.

Figure 20 is a schematic view of the hydraulic circuit.

Figures 21(a), (b) and (c) and Figure 22 are schematic views showing engagement conditions in a spline section respectively.

Figures 23/at, (b) and (c) illustrate engagement conditions in the spline section respectively, Figure 23(a) is a vertical sectional view, Figure 23(b) is a plane view of a rotor shaft, and Figure 23(c) is a schematic view showing engagement conditions respectively.

Figures 24(at, (b) and (c) are schematic views of hydraulic circuits to push up valves.

Figure 25 is a vertical sectional view of part showing a modification example of hydraulic oil path in a eleventh embodiment of the power transmission apparatus according to the present invention.

Figure 26 is a vertical sectional view of part showing another modification example of a hydraulic oil path in a twelfth embodiment of the power transmission apparatus according to the present invention.

Figure 27 is a graph showing delivery pressure of a vane pump.

Figure 28 is a vertical sectional view showing part of a thirteenth embodiment of the power transmission apparatus ofthe present invention.

Figure 29 is a vertical sectional view showing part of a hydraulic oil path for adjusting side clearance in the thirteenth embodiment of the power transmission apparatus of the present invention.

Figure 30 is a schematic side view showing part of a rotor.

Figure 31 is a sectional view of taken along the line X-X in Figure 30.

Figure 32 is a sectional view showing a vane portion of a preferred vane pump according to the present invention.

Detailed description of the preferred embodiments Preferred embodiments of the present invention when applied to a vehicle of four-wheel drive type will now be described in detail.

Referring first to Figure 1 showing the structure of the vehicle of four-wheel drive type to which the present invention is applied, a transmission mechanism 2 is connected to an engine 1 having a crankshaft extending in the widthwise direction of the vehicle, and an output shaft 3 of the transmission mechanism 2 transmits the drive force to a drive gear 4. From the drive gear 4, the drive force is transmitted through an idle gear 5 to an intermediate transmission shaft 8 having gears 6 and 7 mounted on its both ends respectively. From one of the gears or the gear 7 mounted on one end of the intermediate transmission shaft 8, the drive force is transmitted to a differential unit 10 for front wheels 9 to drive the front wheels 9. On the other hand, the drive force transmitted to the front wheels 9 is directly transmitted through a gear 12 to a first rotary shaft 11.The drive force is then transmitted through a power transmission apparatus 13 to a second rotary shaft 14. The drive force is then transmitted through a gear mechanism 15, which changes the direction of power transmission, to a differential unit 17 for rear wheels 16 to drive the rear wheels 16.

A first embodiment of the power transmission apparatus 13 of the present invention will be described with reference to Figure 2. Referring to Figure 2, the power transmission apparatus 13 embodying the present invention comprises a hydraulic oil pump 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 drive force transmitted to the front wheels 9 is directly transmitted. The cam ring 20b is coupled to the second rotary shaft 14 transmitting the drive force to the rear wheels 16. The vane pump 20 delivers hydraulic oil of an amount proportional to the rotation speed thereof.More precisely, this vane pump 20 functions as a hydraulic oil pump when there occurs relative rotation between the rotor 20a and the cam ring 20b, that is, when there occurs 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 ports acting as delivery ports (which are leading ports in the direction of relative rotation) are closed, the rotor 20a and the cam ring 20b are rotated integrally as a rigid body by the action of the static pressure of hydraulic oil.For this purpose, the cam ring 20b is formed with two pump chambers disposed at diagonal positions, and the cam ring 20b is also formed with four ports 22,23, 24 and 25 diposed at substantially diagonal positions so that trailing and leading ones of them in the direction of relative rotation act as suction ports and delivery ports respectively.

The ports 22 and 24 disposed in diagonal relation to perform the same function communicate with each other through a first hydraulic oil passage 26. Also, the ports 23 and 25 disposed in diagonal relation to perform the same function communicate with each other through 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 through a mechanism which permits supply of hydraulic oil even when the cam ring 20b is under rotation.

The first and second hydraulic oil passages 26 and 27 communicate with a hydraulic oil reservoir 30 through a first valve unit or a first check valve 28 and a second valve unit or a second check valve 29 respectively. The two hydraulic oil passages 26 and 27 permit flow of hydraulic oil from the hydraulic oil reservoir 30 only. Further, these hydraulic oil passages 26 and 27 communicate with each other through a pair of opposing selector check valves 31 and 32 which permit flow-out of hydraulic oil from the associated hydraulic oil passages 26 and 27 respectively. The space intermediate between these two check valves 31 and 32 acting as flow selector means communicate with a relief valve 33 disposed in a hydraulic oil delivery passage 41 to act as flow control means.A communication passage 35 extends between an intermediate portion of the relief valve 33 including a valve member having a spring 34 associated therewith and intermediate passage portions between the oil reservoir 30 and the two check valves 28 and 29. A piston 36 is disposed on the other side of the spring 34 for controlling the force of the spring 34 normally applying a pressure in the closing direction of the relief valve 33. The pressure of hydraulic oil which is duty-controlled in a manner as described later acts on the piston 36. For the purpose of the duty control, hydraulic oil at a constant pressure supplied through an orifice 37 is controlled by a solenoid valve 38. This solenoid valve 38 is electrically connected to a computer 39.Signals indicative of the rotation speed of the engine 1, the rotation speed of the first rotary shaft 11, the rotation speed of the second rotary shaft 14, and the throttle opening are applied together with the output signals of a brake actuation detection switch and a steering angle detector to the computer 39 which controls the pressure of hydraulic oil acting on the piston 36. The hydraulic oil at the constant pressure supplied through the orifice 37 may be provided by hydraulic oil used for the control of the transmission 2 when the transmission is of the automatic type or provided by an oil pump when the transmission 2 is of the manual type. Such hydraulic oil may also be provided by hydraulic oil used for power steering, hydraulic oil supplied to the brake booster or hydraulic oil derived from the delivery side of the vane pump 20.

By virtue of such an arrangement of the hydraulic control circuit 21, the pressure of hydraulic oil delivered from the delivery ports of the vane pump 20 acts always on the valve member of the relief valve 33 and the hydraulic oil reservoir 30 communicates with the suction ports of the vane pump 20 regardless of the direction of relative rotation of the rotor 20a and the cam ring 20b.

The function of such a power transmission apparatus 13 will be described with reference to the case where the force urging the relief valve 33 in the open direction against the force of the spring 34 is constant, that is, when the force of the spring 34 only resists the opening movement of the relief valve 33.

In the usual straightforward running state of the vehicle, the effective radius of the tires of the front wheels 9 is the same as that of the rear wheels 16, and the rate of slip of the tires under rotation is quite small. In such a state, there occurs no rotation speed difference between the first rotary shaft 11 and the second rotary shaft 14 of the power transmission apparatus 13. Accordingly, no hydraulic oil under pressure is delivered from the vane pump 20, and no drive force is transmitted to the rear wheels 16. Thus, the vehicle is driven by the front wheels 9 only. That is, the vehicle runs in the two-front-wheel drive mode.

However, when the vehicle running straightforward is, for example, accelerated, a slip of less than about 1 % occurs usually on the front wheels 9 although it is not so appreciable. When there occurs a rotation speed difference between the first and second rotary shafts 11 and 14 due to the above slip of the front wheels 9, the vane pump 20 is energized to build up the pressure corresponding to the above rotation speed difference. The rotor 20a and the cam ring 20b rotate integrally, and the drive force corresponding to the built-up pressure and the pressure receiving area of the vanes is transmitted to the rear wheels 16 to establish the four-wheel drive mode. The flow of hydraulic oil in the vane pump 20 in this case is as shown in Figure 3(a).It will be seen in Figure 3(a) that, due to the rotation of the rotor 20a relative to the cam ring 20b, the ports 23 and 25 act as the hydraulic oil suction ports, and hydraulic oil is sucked from the hydraulic oil reservoir 30 into the suction ports 23 and 25 through the check valve 29. On the other hand, the ports 22 and 24 act as the delivery ports to close the check valve 28 and selector check valve 32, and, at the same time, delivered hydraulic oil is fed through the selector check valve 31 toward the relief valve 33. In Figure 3(a), the solid and broken lines with arrows indicate the flowing directions of delivered and sucked hydraulic oil respectively.

Suppose then that the rotation speed of the front wheels 9 becomes very high compared with that of the rear wheels 16 as, for example, when the vehicle runs on a snow-laden road or is abruptly accelerated or abruptly braked resulting in locking of the rear wheels 16. In such a case, the rotation speed difference between the first and second rotary shafts 11 and 14 in the power transmission apparatus 13 becomes very large. Consequently, a high pressure is generated in the vane pump 20, and such hydraulic oil flows in the directions shown in Figure 3(a). When the pressure of hydraulic oil exceeds a predetermined level, the relief valve 33 is opened against the force of the spring 34, and the pressure of delivered hydraulic oil is controlled to be substantially constant.Thus, a constant drive force corresponding to the regulated pressure of delivered hydraulic oil is transmitted to the rear wheels 16 to establish the four-wheel drive mode.

Consequently, the rotation speed of the front wheels 9 is decreased, while that of the rear wheels 16 is increased, so that the rotation speed difference between the front and rear wheels 9 and 16 is reduced. (This function is the same as that of the non-slip differential unit.) Thus, when a slip occurs on the front wheels 9, the driving torque for the rear wheels 16 is increased to avoid the impossibility of running, while, when the rear wheels 16 tend to be locked, the braking torque for the front wheels 9 is increased to prevent locking of the rear wheels 16.

Suppose further the case where the rotation speed of the rear wheels 16 is very high compared with that of the front wheels 9 as, for example, when the front wheels 9 tend to be locked due to application of the brakes.

In such a case, a very large difference appears between the rotation speeds of the first and second rotary shafts 11 and 14 of the power transmission apparatus 13 in a direction opposite to the aforementioned direction. Consequently, the flow of hydraulic oil in the vane pump 20 is now as shown in Figure 3(b). It will be seen in Figure 3(b) that the ports 22 and 24 act as hydraulic oil suction ports, and hydraulic oil is sucked from the hydraulic oil reservoir 30 into the suction ports 22 and 24 through the check valve 28, while, on the other hand, the ports 23 and 25 act as hydraulic oil delivery ports.Accordingly, hydraulic oil flowing through the second hydraulic oil passage 27 closes the check valve 29 and selector check valve 31, and, at the same time, hydraulic oil at a high pressure fed from the second hydraulic oil passage 27 flows through the selector check valve 32 toward the reliefvalve 33. Since the pressure of such hydraulic oil is also maintained constant by the relief valve 33, the corresponding constant drive force is transmitted to the rear wheels 16 to establish the four-wheel drive mode. As a result, the braking torque imparted to the rear wheels 16 is increased to prevent locking of the front wheels 9.

In a usual running state of the vehicle running along a curve, the rotation speed of the front wheels 9 is slightly higher than that of the rear wheels 16, and the vehicle runs along the curve in the four-wheel drive mode in which the braking torque is imparted to the front wheels 9, while the driving torque is imparted to the rear wheels 16.

In the manner described above, the delivery pressure of hydraulic oil is controlled so as not exceed a constant value by the relief valve 33 in the power transmission apparatus 13. Therefore, in contrast to the prior art case where the driver's manipulation is required to establish the four-wheel drive mode in a vehicle of the part-time four-wheel drive type, the change-over between the four-wheel drive mode and the two-wheel drive mode can be automatically effected, and the four-wheel drive mode is established by the drive force corresponding to the rotation speed difference between the front and rear wheels, according to the present invention.Also, when compared with the center differential unit necessarily provided in a vehicle of the full-time four-wheel drive type, the power transmission apparatus 13 according to the present invention is small in size, compact in structure, light in weight and low in cost.

As described already, the pressure of hydraulic oil acting on the lower and of the piston 36 is duty-controlled to regulate the force required for opening the relief valve 33. By so regulating, the pressure of hydraulic oil delivered from the vane pump 20 can be regulated or controlled, and the drive force transmitted to the rear wheels 16 can be regulated.

Therefore, the manner of duty control may be such that, when a heavier load of the engine 1 is sensed on the basis the output signal of the throttle opening sensor, the pressure of hydraulic oil delivered from the vane pump 20 is correspondingly increased, so that the vehicle can run in the four-wheel drive mode in which the drive force transmitted to the rear wheels 16 is increased correspondingly. Further, the manner of duty control may be such that the pressure of hydraulic oil delivered from the vane pump 20 is increased when the output signal of the brake actuation detection switch sensing the state of actuation of the foot brake indicates turn-on of this switch, locking of the front and rear wheels 9 and 16 can be prevented to shorten the braking distance of the vehicle, and also the vehicle can be stably braked.

Further, the manner of duty control may be such that the pressure of hydraulic oil delivered from the vane pump 20 is decreased with the increase in the steering angle sensed by the steering angle sensor, the vehicle can smoothly run along a curve without giving rise to the tight corner braking phenomenon. Further, the vehicle can stably run when the pressure of hydraulic oil delivered from the vane pump 20 is regulated or controlled depending on the rotation speed of the engine 1 and the velocity of the vehicle sensed on the basis of the sensor output signals applied to the computer 39.

A second embodiment of the present invention in which its hydraulic control circuit 21 differs from that employed in the first embodiment will be described with reference to Figure 4. In Figure 4, the same reference numerals are used to designate the same parts appearing in Figure 2. In the power transmission apparatus 13 shown in Figure 4, the structure of the hydraulic oil pump 20 is the same as that described already. In Figure 4, the flow selector means or the two selector check valves 31 and 32 [shown in Figures 3(a) and 3(b)1 disposed between the first and second hydraulic oil passages 26 and 27 of the hydraulic control circuit 21 to permit flow-out of hydraulic oil from the respective hydraulic oil passages 26 and 27 only in the first embodiment are replaced by a single selector valve 40 of spool type.The hydraulic oil passages defined by the both ends of the spool of the selector valve 40 communicate with the first or second hydraulic oil passage 26 or 27 and communicate also with the hydraulic oil reservoir 30 through the check valves 28 and 29 respectively, and a hydraulic oil delivery passage 41 communicates with the intermediate position of the hydraulic oil passages changed over by the spool of the selector valve 40.

The relief valve 33 is replaced by a pressure control valve 42 which includes a spool 42a having spaced two lands. The hydraulic oil delivery passage 41 communicates with the space between the two lands of the spool 42a, and a passage 43 for feeding hydraulic oil at a regulated pressure communicates with the oil reservoir 30 and also with the space between the two lands of the spool 42a. A biasing force of a spring 42b acts on the left-hand end of the spool 42a, and the pressure of hydraulic oil duty-controlled by the combination of the orifice 37 and the solenoid valve 38 acts also on that end.Accordingly, hydraulic oil at a regulated pressure is obtained as a result of a balance between the spool biasing force provided by the combination of the force of the spring 42b and the duty-controlled pressure of hydraulic oil and the spool biasing force corresponding to the differential area of the two lands of the spool 42a to which hydraulic oil is fed from the delivery passage 41. The hydraulic oil at such a regulated pressure is fed into the hydraulic oil passage 43 to be returned to the hydraulic oil passage 26 or 27 connected to the suction ports.

Suppose that the relative rotation speed of the first rotary shaft 11 is higher than that of the second rotary shaft 14. Suppose then, that the rotor 20a rotates clockwise in such a case, as described with reference to Figure 3(a). Then, according to the arrangement of such a hydraulic control circuit 21, the first and second hydraulic oil passages 26 and 27 are now connected to the delivery and suction ports respectively. As a result, the pressure of delivered hydraulic oil acts on the left-hand end face of the spool of the selector valve 40 to urge this spool to its extreme rightward position, and the first hydraulic oil passage 26 communicates with the delivery passage 41.Therefore, hydraulic oil delivered through the first hydraulic oil passage 26 is fed through the delivery passage 41 toward the pressure control valve 42, and hydraulic oil at the regulated pressure is circulated through the check valve 29 to the suction ports.

Suppose then that the relative rotation speed of the second rotary shaft 14 is higher than that of the first rotary shaft 11. Then, the cam ring 20b rotates clockwise as described with reference to Figure 3(b), and the second and first hydraulic oil passages 27 and 26 are now connected to the delivery and section ports respectively.

As a result, the pressure of delivered hydraulic oil acts on the right-hand end face of the spool of the selector valve 40 in the hydraulic control circuit 21 in Figure 4, and this spool is urged to its extreme leftward position to permit communication of the second hydraulic oil passage 27 with the delivery passage 41.

Hydraulic oil from the second hydraulic oil passage 27 is fed toward the pressure control valve 42, so that hydraulic oil at the regulated pressure is fed through the check valve 28 to be circulated to the suction ports.

Thus, regardless of the directions of relative rotation of the first and second rotary shafts 11 and 14, delivered hydraulic oil is always fed into the delivery passage 41. Therefore, when the pressure of hydraulic oil acting on the spool 42a of pressure control valve 42 is regulated by the duty-controlled solenoid valve 38, the pressure of hydraulic oil delivered from the hydraulic oil pump 20 can be controlled, so that the drive mode corresponding to the present running state of the vehicle can be established as described already.

A third embodiment of the present invention in which its hydraulic control circuit 21 differs from that employed in the first and second embodiments will be described with reference to Figure 5. In Figure 5, the same reference numerals are used to designate the same parts appearing in Figures 2 and 4.

In the embodiment shown in Figure 5, the selector valve 40 employed in the second embodiment is replaced by a spool valve 45 controlled by on-off of a solenoid valve 44. This spool valve 45 includes a spool 45a having three spaced lands. In Figure 5, the first and second hydraulic oil passages 26 and 27 communicate with the spaces between the three lands of the spool 45a respectively and communicate also with the hydraulic oil reservoir 30 and the regulated hydraulic oil passage 43 from the pressure control valve 42 through the check valves 28 and 29 respectively. The delivery passage 41 communicates with the spool valve 45 to be opened and closed by the central land of the spool 45a, so that the first and second hydraulic oil passages 26 and 27 can also communicate with the pressure control valve 42.A spring 45b is mounted on the left-hand end of the spool 45a of the spool valve 45, and the on-off controlled solenoid valve 44 is disposed upstream of an orifice 46 formed in a hydraulic oil passage communicating with the right-hand end of the spool 45a of the spool valve 45. The computer 39 is electrically connected to the solenoid valve 44.

Suppose that the relative rotation speed of the first rotary shaft 11 is higher than that of the second rotary shaft 14, and the rotor 20a rotates clockwise in such a case. Then, as described with reference to Figure 3(a), the first and second hydraulic oil passages 26 and 27 are connected to the delivery and suction ports respectively in such a hydraulic control circuit 21.

The computer 39 detects the direction of rotation (the direction of relative rotation) of the hydraulic oil pump 20 on the basis of the input signals indicative of the rotation speeds of the first and second rotary shafts 11 and 14. Then, the computer 39 turns on the solenoid valve 44 to open the passage upstream of the orifice 46. The spool 45a of the spool valve 45 is urged to its extreme rightward position by the spool biasing force provided by the combination of the force of the spring 45b and the force corresponding to the differential area of the two lands of the spool 45a to which hydraulic oil is fed from the first hydraulic oil passage 26, thereby establishing communication between the first passage 26 and the delivery passage 41.

Hydraulic oil fed into this delivery passage 41 is circulated as in the case of Figure 4.

Suppose, on the contrary, that the relative rotation speed of the second rotary shaft 14 is higher than that of the first rotary shaft 11, and the cam ring 20b rotates clockwise in such a case. Then, as described with reference to Figure 3(b), the second and first hydraulic oil passages 27 and 26 are connected to the delivery and suction ports respectively. In such a case, the computer 39 turns off the solenoid valve 44. The pressure of hydraulic oil acts on the right-hand end face of the spool 45a, together with the force corresponding to the differential area of the two lands of the spool 45a to which hydraulic oil is fed from the second hydraulic oil passage 27, thereby urging the spool 45a to its extreme Ieftward position against the force of the spring 45b, to establish communication between the second passage 27 and the delivery passage 41.Hydraulic oil fed into this delivery passage 41 is circulated as in the case of Figure 4.

Thus, by virtue of the provision of the spool valve 45 which is changed over under control of the solenoid valve 44, the spool valve 45 can be reliably operated regardless of the direction of relative rotation of the hydraulic oil pump 20.

A fourth embodiment of the present invention in which its hydraulic control circuit 21 differs from that employed in the first embodiment will be described with reference to Figure 6. In Figure 6, the same reference numerals are used to designate the same parts appearing in Figure 2.

In the embodiment shown in Figure 6, the pressure with which the relief valve 33 is urged to its open position is not controlled by the computer 39 but is set a predetermined value determined by the force of the spring 34, so that the delivery ports of the vane pump 20 may not communicate with the hydraulic oil reservoir 30 when the pressure of hydraulic oil does not exceed the pressure setting of the relief valve 33. A first auxiliary hydraulic oil passage 52 having a flow restrictor or orifice 50 and a second auxiliary hydraulic oil passage 53 having a flow restrictor or orifice 51 are connected between the ports 23, 24 and between the ports 22, 25 respectively so that the relief valve 33 can achieve the pressure control or regulating function.

Thus, the relief valve 33 constitutes hydraulic oil flow control means together with the orifices 50 and 51.

More precisely, flow of hydraulic oil through the auxiliary passages 52 and 53 is usually restricted by the respective orifices 50 and 51. Therefore, when the rotation speed difference between the first and second rotary shafts 11 and 14 is small, the amount of hydraulic oil flowing into the auxiliary passages 52 and 53 is small. With the increase in the rotation speed difference between the first and second rotary shafts 11 and 14, the pressure of hydraulic oil delivered from the vane pump 20 becomes higher until finally it overcomes the resistance of the orifices 50 and 51 against flow. In such a situation, hydraulic oil flows through the auxiliary passages 52 and 53 before the pressure of delivered hydraulic oil attains the control pressure setting of the relief valve 33, thereby to transmit the torque corresponding to the rotation speed difference to the second rotary shaft 14.Thus, the auxiliary passages 52 and 53 act to adjust the pressure control function of the relief valve 33.

When the rotation speed difference between the front wheels 9 and the rear wheels 16 is small irrespective of the direction of relative rotation of the rotor 20a and cam ring 20b of the vane pump 20, the pressure of hydraulic oil delivered from the vane pump 20 overcomes the flow resistance of the orifices 50 and 51 in such a hydraulic control circuit 21, and hydraulic oil flows through the auxiliary passages 52 and 53 into the suction ports, so that the torque corresponding to the rotation speed difference is transmitted to the rear wheels 16. When on the contrary, the rotation speed difference between the wheels 9 and 16 is large, the hydraulic oil reservoir 30 would not communicate with the suction ports of the vane pump 20 before the pressure of delivered hydraulic oil attains the level exceeding the setting of the spherical valve member 36 of the relief valve 33.The rotation speed difference between the front and rear wheels 9 and 16 relative to the pressure of delivered hydraulic oil is shown in Figure 7 in which the characteristic curve a represents the relation when the orifices 50 and 51 are provided (that is, when the auxiliary passages 52 and 53 are provided), while the characteristic curve b represents the relation when the orifices 50 and 51 are not provided. It will be seen in Figure 7 that the pressure of delivered hydraulic oil is lower when the orifices 50 and 51 are provided.

In the power transmission apparatus 13 shown in Figure 6, the pressure of hydraulic oil delivered from the vane pump 20 is maintained at a predetermined value by the function of the relief valve 33, and the auxilary passages 52 and 53 having the respective orifices 50 and 51 therein are provided, so that, when the rotation speed difference exceeds a predetermined value which may be small, the torque corresponding to the rotation speed difference can be transmitted to the rear wheels 16. Therefore, the present invention obviates the defect that a large torque is transmitted to the rear wheels 16, in spite of the fact that closure of the hydraulic oil passage results in build-up and increase of the delivery pressure and that the rotation speed difference is not so large. Thus, the torque exactly corresponding to the rotation speed difference can be transmitted to the rear wheels 16.

A fifth embodiment of the present invention will be described with reference to Figure 8. In the embodiment shown in 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 a structure as shown is disposed in each of the auxiliary passages 52 and 53 in lieu of the orifices 50 and 51. This flow restrictor 55 is disposed in each of the auxiliary passages 52 and 53 connecting between the delivery and section ports of the hydraulic oil pump 20.The flow restrictor 55 includes a casing 57 disposed between the delivery passage and the suction passage, a needle valve 60 provided in the casing 57 through a spring 58 for openably closing a communication port 61 bored in the wall of the auxiliary passage 52(53), a diaphragm 59 fixed to the needle valve 60 for making expansion and contraction depending on the load of the engine (not shown), and a pressure transmission port 62 bored in the wall of the casing 57 fortransmitting the negative pressure in the engine intake manifold to the diaphragm 59.

The larger the torque of the engine, the lower is the negative pressure in the engine intake manifold, and the opening the communication port 61 becomes smaller. However, the communication port 61 is usually maintained in a slightly open condition without being completely closed, and with the decrease in the load of the engine, the communication port 61 is opened wider. On the other hand, with the increase in the load of the engine, the drive force increases correspondingly. The communication port 61 is narrowed by the advancing movement of the needle valve 60 so as to establish the four-wheel drive mode.

A sixth embodiment of the present invention will be described with reference to Figure 9. In the embodiment shown in 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 a structure as shown is disposed in each of the auxiliary passages 52 and 53 in lieu of the orifices 50 and 51. This flow restrictor 65 is disposed in each of the auxiliary passages 52 and 53 connecting between the delivery and suction ports of the hydraulic oil pump 20. The flow restrictor 65 includes a casing 66 disposed between the delivery passage and the suction passage.A pressure transmission port 67 is bored in the wall of the casing 66 for transmitting the pressure of hydraulic oil delivered from the oil pump of the power steering mechanism provided for the handle, and another pressure transmission port 68 is bored in the wall of the casing 66 for transmitting the negative pressure in the engine intake manifold. A piston 70 loaded with a spring 69 is mounted between the manifold pressure transmission port 67 and the oil pump delivery pressure transmission port 68 in the casing 66 so as to make vertical movement under influence of the oil pump delivery pressure and the manifold pressure. A needle valve 71 for opening and closing the auxiliary passage 52(53) is fixed to the lower end of the piston 70.

The larger the steering angle, the oil pressure of the power steering mechanism becomes higher. In such a case, the needle valve 71 in the flow restrictor 65 of the above structure makes retracting movement to widely open the auxiliary passage 52(53) thereby increasing the allowance for the rotation speed difference between the front and rear wheels 9 and 16. On the other hand, when there occurs a large rotation speed difference between the front and rear wheels 9 and 16 in the straightforward running state of the vehicle, this allowance for the rotation speed difference is cancelled, and the required torque is transmitted to the front wheels 9.

Further, the negative pressure in the intake manifold which is transmitted through the pressure transmission port 68 to open and close the auxiliary passage 52(53) may be arranged, to act in interlocking relation with the engine torque, so that the drive mode of the vehicle can be suitably changed over to the four-wheel drive mode depending on the relative magnitudes of the engine torque and steering angle.

It is to be noted herein that the spring 69 incorporated in the flow restrictor 65 should be considerably strong since there is a very great difference between the intake manifold pressure and the power steering oil pressure.

Means for variably restricting the flow of hydraulic oil through the auxiliary passages 52 and 53 depending the running conditions of the vehicle may also be employed for attaining the object similar to that described above. For example, the flow of hydraulic oil through the auxiliary passages 52 and 53 may be suitably restricted depending on the brake oil pressure or depending on whether or not the accelerator pedal is released. Further, the flow of hydraulic oil through the auxiliary passages 52 and 53 may be suitably restricted for the purpose of control depending on the running speed of the vehicle, the angular velocity of steering, etc.

A seventh embodiment of the present invention will be described with reference to Figure 10 in which the pressure of hydraulic oil delivered from the hydraulic oil pump or vane pump 20 is utilized for biasing the vanes of the vane pump 20.

Referring to Figure 10, the hydraulic oil pump or vane pump 20 is formed with a plurality of, or, for example, eight radial slots 80 in equally circumferentially spaced apart relation along the outer peripheral surface 20c of the rotor 20a, and, in each of these slots 80, a vane 81 is inserted so as to make sliding engagement with the inner peripheral surface 20e of the cam ring 20b.

In this embodiment, the vanes 81 of the vane pump 20 are normally biased toward and onto the inner peripheral surface 20c of the cam ring 20b by a hydraulic oil supply unit M acting as a vane biasing unit. More precisely, the rotor 20a is formed with an annular recess 83 communicating with the bottom (the radially inner end) of the slots 80, and this annular recess 83 communicates through a hydraulic oil passage 84 to an accumulator 85 including a piston 85a urged by a spring 85b. A check valve 86 having a parasol-shaped valve member 86a is disposed in the passage 84 which is connected to the hydraulic oil delivery passage 41.

According to the above arrangement, the vanes 81 make sliding engagement with the inner peripheral surface 20e of the cam ring 20b by being urged by the working oil of high pressure accumulated in the accumulator 85, when the pressure of hydraulic oil drops due to a decrease in the rotation speed difference between the rotor 20a and the cam ring 20b, and the check valve 86 is urged to its closed position. Therefore, even when the relative rotation between the rotor 20a and the cam ring 20b is ceased as a result of the stoppage of the vehicle, the vanes 81 are still urged toward and onto the inner peripheral surface 20e of the cam ring 20b, so that the coupling function of the vane pump 20 is sufficiently maintained.

The power transmission apparatus according to the present invention is not only applicable to the vehicle of four-wheel drive type shown in Figure 1, but also applicable to, for example, another vehicle of four-wheel drive type as shown in Figure 11.

Referring to Figure 11 showing 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 drive force to a drive gear 4. From the drive gear 4,the drive force is transmitted to a second rotary shaft 14 through a driven gear 5. From the second rotary shaft 14, the drive force is transmitted to a differential unit 17 for rear wheels 16 that the rear wheels 16 are driven directly from the engine 1 through the aforementioned power train.On the other hand, the drive force transmitted to the second rotary shaft 14 is transmitted through the power transmission apparatus 13 of the present invention to a first rotary shaft 11, and this drive force is transmitted to a differential unit 10 for front wheels 9, so that the front wheels 9 are indirectly driven through the power transmission apparatus 13 by the drive 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 drive force driving the rear wheels 16 is transmitted intact, and the cam ring 20b of the vane pump 20 is coupled to the first rotary shaft 11 transmitting the drive force to the front wheels 9.

According to such an arrangement, a relatively large drive force is applied to the rear wheels 16 where the grip limit torque becomes large due to the distribution of a greater proportion of the vehicle weight during an abrupt acceleration. Therefore, the power of the engine can be effectively utilized to exhibit an exceilent acceleration performance. Further, since only the engine torque portion exceeding the grip limit of the rear wheels giving rise to a slip is transmitted to the front wheels 9 by the vane pump 20, the torque transmission capacity of the vane pump 20 need not be large.

Further, since the drive force is distributed to the rear wheels 16 in a greater proportion than the front wheels 9 by being limited by the torque transmission capacity of the vane pump 20, the force gripping the front wheels 9 which are the steering wheels can be maintained to be high enough for improving the steering performance.

A ninth embodiment of the present invention in which the vane pump is replaced by a gear pump will be described with reference to Figures 12 to 14.

Referring to Figures 12 to 14, the power transmission apparatus in this embodiment comprises a gear pump 120 and a hydraulic control circuit 121 associated with the gear pump 120.

The gear pump 120 includes a housing 122 formed with cylindrical cavities 122a, 122b and 122c. A first pinion gear 123, a sun gear 124, and a second pinion gear 123 making meshing engagement with each other are received in these cavities 122a, 122b and 1 22c respectively, and the two pinion gears 123 and the single sun gear 124 constitute two pumps. A gear 125 is formed on the outer periphery of the housing 122 to transmit power to a first rotary shaft 143 described later.

Ports 126, 127, 128 and 129 communicating with the gear pump 120 are formed in one of the side walls of the housing 122, and a casing 133 formed with hydraulic oil passages communicating with these ports 126, 127, 128 and 129 is provided on that sidewall. A first cover 135 is provided on the other side wall of the housing 122, and a second cover 136 is provided outside the first cover 135. Ail of the casing 133, first cover 135 and the second cover 136 are fastened to the housing 122 by bolts 140. A spline 141 is formed on the inner peripheral surface of the end portion of the second cover 136, and, by the spline 141, the second cover 136 is coupled to a first rotary shaft 143 transmitting the drive 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 unit 147 for the front wheels 9, so that the drive force can be transmitted to the front wheels 9 from the first rotary shaft 143.

A second rotary shaft 149 transmitting the drive force to the rear wheels 16 extends loosely through the first rotary shaft 143, first cover 135 and the second cover 136. This second rotary shaft 149 is positioned in a relation coaxial with the rotation axis of the sun gear 124to be coupled to the sun gear 124. Bearings 142 are provided so as to freely rotatably support the gear pump 120, etc. within the transmission casing.

The operation of the gear pump 120 is similar to that of the aforementioned vane pump 20. When a relative rotation speed difference occurs between the first and second rotary shafts 143 and 149, the sun gear 124 and two pinion gears 123 function as the gear pump. When the ports acting as delivery ports among the ports 126, 127, 128 and 129 are closed, the sun gear 124 and pinion gears 123 rotate unitarily so that the housing 122 can rotate in unitary relation with the sun gear 124.

The hydraulic control circuit 121 is shown in Figure 14.

Referring to Figure 14, the hydraulic control circuit 121 includes a first hydraulic oil passage 153 providing communication between the ports 126 and 128 of the gear pump 120 and communicating with a hydraulic oil reservoir 152 through a first check valve 150, a second hydraulic oil passage 156 providing communication between the ports 127 and 129 of the gear pump 120 and communicating with the reservoir 152 through a second check valve 155, a first relief valve 157 permitting flow of hydraulic oil from the first passage 153 to the second passage 156 only, and a second relief valve 158 permitting flow of hydraulic oil from the second passage 156 to the first passage 153 only. Each of the first and second relief valves 157 and 158 includes a valve member 162 biased by a compression spring 161 which controls the pressure with which the valve is urged to open position. These first and second relief valves 157 and 158 constitute flow selector means and flow rate control means.

It is apparent that the operation and effect of the power transmission apparatus including the gear pump 120 described above with reference to Figures 12 to 14 are equivalent to those of the power transmission apparatus including the vane pump 20.

Embodiments described below are to provide a compact layout of hydraulic oil pump. A hydraulic oil pump unit 213 described later is provided with both a hydraulic oil pump and a hydraulic control circuit.

A tenth embodiment of the present invention is described referring to Figures 15 through 24.

Referring to Figures 15 and 16 showing the structure of the vehicle of four-wheel drive type to which the present invention is applied, a transmission mechanism 202 is connected to an engine 201 having a crankshaft extending in the widthwise direction of the vehicle, and, and output shaft 203 of the transmission mechanism 202 transmits the drive force to a counter gear 204. From the counter gear 204, the drive force is transmitted to a hydraulic oil pump unit 213 through a gear unit 220a which is formed on the peripheral of the hydraulic oil pump unit 213. From the hydraulic oil pump unit 213, the drive force is transmitted to a second rotary shaft 214. The drive force is then transmitted through a gear mechanism 215, which changes the direction of power transmission, to a differential unit 217 for rear wheels 216 to drive the rear wheels 216.

Referring to Figures 15 through 17, the hydraulic oil pump unit 213 comprises a hydraulic oil pump in the form of a vane pump 200 and an associated hydraulic control circuit 221. The vane pump 200 includes a rotor 219 and a cam ring 220. The cam ring 220 is connected through a first rotary shaft 11 and a differential unit 210 to front wheels 209. The rotor 219 is coupled to the second rotary shaft 214 transmitting the drive force to the rear wheels 216.

Referring to Figure 19, the rotor 219 of the vane pump which functions as a hydraulic oil pump have a number of vane grooves 219b (10 vane grooves in this embodiment) on a peripheral section 219a, the vane grooves 219b being formed in peripheral direction with equal spacings. Each of the vane grooves 219b has a vane 218 which is in contact with an inner surface 220d of the cam ring 220. The number of the vanes 218, which is 10 in this embodiment, is determined so that it is not an integer multiple of the number of pump chambers 236a, 236b and 236c as described hereinafter; that is set to an integer multiple of the number of the pump chambers plus one.

The vane pump 200 delivers hydraulic oil of an amount proportional to the rotation speed thereof. More precisely, this vane pump 200 generates a hydraulic pressure in the pump chambers 236a, 236b and 236c when there occurs relative rotation between the rotor 219 and cam ring 220, that is, when there occurs relative rotation between the first rotary shaft 211 and the second rotary shaft 214. The operation of this vane pump 200 is such that, when delivery ports (ports 222a, 223a and 224a or ports 222b, 223b and 224b which are leading ports in the direction of relative rotation between the cam ring 220 and vane 218) of the vane pump 200 are closed, the rotor 219 and the cam ring 220 are rotated integrally as a rigid body by the action of the static pressure of hydraulic oil.

For this purpose, the inner peripheral surface of the cam ring 220 is formed to a triangle-like shape so that three pump chambers 236a, 236b and 236c are formed near the vertexes of a triangle between the inner cam ring peripheral surface 220d and the rotor 219. Six ports 222a, 222b, 223a, 223b, 224a and 224b are disposed on both ends of individual pump chambers 236a, 236b and 236c so that trailing ones of them in the direction of relative rotation act as suction ports and leading ones act as delivery ports. The ports 222a, 223a and 224a which act to perform the same function communicate with each other through a second hydraulic oil passage 227.Also, the ports 222b, 223b and 224b which act to perform the same function communicate with each other through a first hydraulic oil passage 226. There is provided a relief valve 233 between the first hydraulic oil passage 226 and the second hydraulic oil passage 227 which permits flow of hydraulic oil from the first hydraulic oil passage 226 to the second hydraulic oil passage 227 when the pressure of hydraulic oil exceeds a predetermined value. There is also provided a relief valve 231 which permits flow of hydraulic oil from the second hydraulic oil passage 227 to the first hydraulic oil passage 226 when the pressure of hydraulic oil exceeds a predetermined value.The first and second hydraulic oil passages 226 and 227 communicate with a hydraulic oil reservoir 230 through check valves 228 and 229 which permit flow of hydraulic oil from the hydraulic oil reservoir 230 only.

By virtue of such an arrangement of the hydraulic control circuit 221, the pressure of hydraulic oil delivered from the vane pump 200 acts always on the valve members of the relief valves 231 and 233 and the hydraulic oil reservoir 230 communicates with the suction ports of the vane pump 200 regardless of the direction of relative rotation of the rotor 219 and the cam ring 220.

The hydraulic oil pump unit 213 is formed as shown in Figure 15 and disposed under 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 cam ring 220, a cover 251 engaging with end surfaces of the rotor 219 and the cam ring 220, a pressure retainer 241 engaging with the other end surfaces of the rotor 219 and the cam ring 220, and a flange 245 fastened with a bolt 248 together with the cover 251 and the pressure retainer 241 to the cam ring 220 to transmit the drive force of the cam ring 220.

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

Thus, the rotor 219 is disposed within a space formed by the cam ring 220, the cover 251 and the pressure retainer 241 so that both end surfaces of the rotor 219 are in contact against the end surfaces of the cover 251 and the pressure retainer 241.

The rotor 219 is connected through a spline 257 to a rear wheel drive shaft 243 which is integrally connected with the second rotary shaft 214, permitting rotation force of the rotor 219 to be input or output through the rear wheel drive shaft 243. An end of the rear wheel drive shaft 243 is inserted into a through-hole 251 a formed at the center of the cover 251. A bushing 252 is inserted between the rear wheel drive shaft 243 and the through-hole 251 a. The bushing 252 supports the rear wheel drive shaft 243 permitting its relative rotation to the cover 251 and keeps the inside of the cover 251 liquid-tight.

The cover 251 is rotatably attached through a bearing 259 to a casing 202a of the transmission mechanism 202.

The rear wheel drive shaft 243 extends externally through a through-hole 241 a formed at the center of the pressure retainer 241. A bushing 256 is inserted between the through-hole 241 a and the rear wheel drive shaft 243. The bushing 256 supports the pressure retainer 241 permitting its relative rotation to the rear wheel drive shaft 243 and keeps the inside of the pressure retainer 241 liquid-tight.

The pressure retainer 241 is rotatably attached through the flange 245 fastened with a bolt 248 and bearings 260a and 260b to the casing 202a.

The cover 251 comprising part of the pump body has an opening near the center of rotation of the pump body (in this embodiment, it is in line with the center of the rear wheel drive shaft 243) and an oil hole 251 b extending in the direction of the center line of the rotary shaft toward the pump body.

Thus, the oil hole 251 b is a part, outer than the end ofthe rear wheel drive shaft 243, of the through-hole 251 a formed in the cover 251, extending in the direction of the rear wheel drive shaft 243.

An outer part than the end of the rear wheel drive shaft 243 in the oil hole 251 b is provided with a filter comprising a Nylon net or the like and a magnet 255 to remove any foreign materiais in hydraulic oil flowing in from the end of the through-hole 251a.

In the Figures, 242 is a pulsation damper and 246 is a pulsation volume which reduce pulsation of delivered hydraulic oil.

An outer peripheral portion of the pressure retainer 241 extends out of the outer peripheral of the rotor 219 and the end of the vane 218, up to the outer peripheral of the cam ring 220. The outer peripheral portion of the pressure retainer 241 is fixed and fastened together with the cam ring 220 fastened with the flange 245, the cover 251 and the bolt 248.

When the vane pump 200 is installed within the transmission mechanism 202, the outer diameter of the vane pump 200 is necessary to be as small as possible.

To make the vane pump 200 compact while the bolt 248 fastening the flange 245 and the cam ring 220 is used as in before, it is necessary to shift the bolt 248 toward the inner peripheral surface 220d of the cam ring 220 as shown in Figure 19. However, this method is not preferable because, as shown by broken lines in Figure 19, an engagement area between the pressure retainer 241 and the cam ring 220 is small and the joint section of the flange 245 and the cam ring 220 is a 3-face joint together with the pressure retainer 241, resulting in reduced joint area.

However, as described above, by extending the outer peripheral section of the pressure retainer 241 up to the outer peripheral section of the cam ring 220 and by fastening the flange 245 and the cam ring 220 together with the pressure retainer 241, the joint area can be increased and the parts can be stably supported in position to assure a sufficient side pressure. Thus, the vane pump 200 is formed as a compact, large-capacity pump.

Referring to Figure 19, the outer peripheral section of the cam ring 220 as part of the pump body has a gear section 220a which is engaged with a counter gear 204 as shown in Figure 16.

Thus, rotation force (drive force of the front wheels or that of the front and rear wheels) to the pump body is transmitted through the counter gear 204 and the gear section 220a. The cam ring 220 is made of an abrasion-resistant material such as a cemented steel in order to minimize wearing by friction with the vane 218.

The counter gear 204 engaging with the gear section 220a is also made of the same material such as a cemented steel and the gear 220a is protected from wearing by engagement with the counter gear 204, which is advantageous as compared with a case where the joint section is made of another material.

The gear section 220a and the cam ring 220 must be formed concentrically. This can be easily accomplished by machining the gears on the basis of the inner peripheral surface 220d of the cam ring 220.

Thus, a number of parts can be eliminated by using the cam ring 220 with the gear section 220a as a joint section.

Though the gear section 220a has a slight problem that it must be made of an abrasion-resistant material, it can be formed on the outer peripheral portion of the cover 51 or the pressure retainer 241.

The cover 251 and the pressure retainer 241 are made of an gray cast iron material or a sintered product and have a problem in resistance to abrasion. If the cover 251 and the pressure retainer 241 are made of a cemented steel or the like, there is a problem in resistance to seizure on a surface sliding with the rotor 219.

However, the above problems can be eliminated by making the cover 251 and the pressure retainer 241 from an abrasion-resistant material such as a cemented steel and treating the sliding surface with the rotor 219 with surface treatment for improved lubrication, and the cover 251 or the pressure retainer 241 can be provided with the outer peripheral gear section 220a as a drive force transmission section.

Thus, the hydraulic oil pump unit 213 is formed in a simple and compact shape.

On the other hand, the rotor 219 is attached through a spline 257 to the rear wheel drive shaft 243 to transmit the input and output drive force of the rotor 219 by the spline 257.

The spline 257 is formed shorter than an axial width of the rotor 219 at the enter of the axial direction of the rotor 219 so that the drive force of the spline 257 and the rotor 219 is transmitted without causing jam of the rotor 219 with the pressure retainer 241 or the cover 251.

With a vane pump with the rotor 219 of a larger width, it will be difficult to attach the rotor 219 in an exact perpendicular position to the rear wheel drive shaft 243 and a seizure may occur between the end surface of the rotor 219 and the pressure retainer 241 or the cover 251. This is considered as due to the fact that, because the rotor 219 is attached slantly, a local force is applied to part of the end surfaces of the rotor 219 and the pressure retainer 241 or the cover 251, resulting in discontinued oil film.

This phenomenon can be eliminated by reducing the width of the spline 257.

Referring to Figures 21(a) through (c), suppose that the rotor 219 is attached slantlyto the rear wheel drive shaft 243 with a minimum clearance between the rotor 219 and the pressure retainer 241 and the cover 251 in the direction of the rotary shaft of the rotor 219.

When the rotor 219 and the rear wheel drive shaft 243 rotate, a torque is transmitted through points A and A' of the spline 257 to the rotor 219. Aforce applied to points A and A' acts as an unbalanced force to the rotor 219 which strongly pushes the pressure retainer and the cover 251 at points B and B', resulting in discontinued oil film and a seizure.

In Figure 21 (a), the solid lines indicate the engagement conditions in a case where the rotor 219 is attached slantly to the rear wheel drive shaft 243 and the broken lines indicate the engagement conditions in a case where the rotor 219 is not slanted to the rear wheel drive shaft 243. If not slanted, there is a line-contact between the rotor 219 and the spline 257. However, if slanted, there will be a point-contact at points A and A'.

By reducing a longitudinal width "1" of the spline 257, points A and A' are brought closer to the center of this width with reduced unbalance and seizure is prevented.

Referring to Figure 22, when the rotor 219 is attached to the spline 257 so that the spline 257 is dislocated by a length "a" from the transversal center of the rotary shaft of the rotor 219, then an unbalanced force F is applied, the unbalanced force F acts as a moment to slant the rotor 219 because the longitudinal center of the spline 257 is dislocated.

This problem can be solved by reducing the longitudinal dislocation "a" of the spline 257 to zero to eliminate the moment slanting the rotor 219, thereby preventing seizure of the rotor 219 with the pressure retainer 241 and the cover 251.

The above unbalanced force F is also caused by an unbalance and eccentricity of rotor 219 itself or by an unbalanced force to the rotor 219 due to leakage of hydraulic oil.

For the above described reasons, the spline 257 is made short and disposed at the transversal center of the rotary shaft of the rotor 219, thereby preventing seizure of the rotor 219 with the pressure retainer 241 and the cover 251.

Referring to Figures 23(a) through (c), the engagement point between the spline 57 and the rotor 219 can be further brought closer to the transversal center of the rotary shaft of the rotor 219 by crowning the rotor 219 in the transversal direction of the rotary shaft.

Thus, by forming the spline 257 so that a portion of its longitudinal center is convexed as shown in Figures 23(b) and (c), the engagement points A and A' between the spline 257 and the rotor 219 are brought closer to the transversal center of the rotary shaft of the rotor 219, thereby reducing pressure applied to the pressure retainer 241 and the cover 251 due to a slant of the rotor 219 to the rear wheel drive shaft 243.

Also, the above mentioned permits the convexed portion at the longitudinal center of the spline 257 to engage over the entire portion with the rotor 219, thereby improving a surface pressure distribution of the spline 257.

The spline 257 is machined to the above shape by a machine tool called "ROTEFLOW". The spline 257 can also be machined to the above shape by a specific hob machine while adjusting its feed rate.

The above crowning of the spline 257 can be applied both to the spline 257 which is not made short as described above and to that which is made short as described above.

The hydraulic control circuit will be described in detail below.

Referring to Figure 15, a casing 202a of the transmission 202 opposing the through-hole 251 a is provided with an oil guide 253.

The oil guide 253 is formed to a cube divided into two halves along its diagonal, the end of which extends into the through-hole 251 a, which collects hydraulic oil falling along an inner wall of the casing 202a and supplies collected hydraulic oil into the through-hole 251 a.

The cover 251 is provided with a hydraulic oil supply path 235 from the end of the rear wheel drive shaft 243 to the through-hole 251a and disposed slantly up toward the outer peripheral of the cover 251 at the side of the rotor 219, thereby communicating with the hydraulic control circuit 221.

In a joint section of the hydraulic oil supply path 235 with the hydraulic control circuit 221, there is formed a communication passage 235a extending from the end of the hydraulic oil supply path 235 toward the through-hole 251 a and opening at a bushing 252 in the through-hole 251 a, and there is provided a spherical valve member 229a at the end of the communication passage 235a, forming a check valve 229.

The above configuration permits hydraulic oil to be supplied from the hydraulic oil supply path 235 through a mount section of the bushing 252 of the through-hole 251a to the cam ring 220, and, at the same time, prevents hydraulic oil from flowing back from the cam ring 220.

There are also provided a hydraulic oil supply path 235', a communication passage 235'a and a check valve 228 at another position in the through-hole 251 a with the same construction as the hydraulic oil supply path 235, the communication passage 235a and the check valve 229. (Refer to Figures 17 and 20).

In the cover 251, the hydraulic control circuit 221 is formed as shown in section (a) (cover portion) and section (b) (vertical sectional view) of Figure 20.

There is provided the first hydraulic oil passage 226 which connects the ports 222b, 223b and 224b acting simultaneously as delivery or suction ports depending on the direction of relative rotation between the rotor 219 and the cam ring 220, and end section 233a of the relief valve 233 and an end section 231b of the relief valve 231, and communicates through the communication passage 235a and the hydraulic oil supply path 235 with the through-hole 251 a.

There are also formed the hydraulic oil supply path 235' and the communication 235'a which communicate with the port 222a and the through-hole 251 a.

On the other hand, in the pressure retainer 241, there is provided the second hydraulic oil passage 227 which connects the ports 222a, 223a and 224a, an end section 233b of the relief valve 233 and an end section 231 a of the relief valve 231, as shown in section (c) (pressure retainer section) and section (b) of Figure 20.

The second hydraulic oil passage 227 communicates through the port 222a and the pump chamber 236a in the cam ring 220 with the communication passage 235'a, the hydraulic oil supply path 235' and the through-hole 251 a.

The first and second hydraulic oil passages 226 and 227 are connected by the relief valves 231 and 233 formed from the cover 251 through the cam ring 220 to the pressure retainer 241.

In the relief valve 231, a spherical valve member disposed at the end section 231 a is applied with a force of a spring 232, thereby permitting flow of hydraulic oil from the second hydraulic oil passage 227 to the first hydraulic oil passage 226 only when the pressure of hydraulic oil exceeds a predetermined value.

In the relief valve 233, a spherical valve member disposed at the end section 233a is applied with a force of a spring 234, thereby permitting flow of hydraulic oil from the first hydraulic oil passage 226 to the second hydraulic oil passage 227 only when the pressure of hydraulic oil exceeds a predetermined value.

The above configuration forms a hydraulic circuit as shown in Figure 17.

Hydraulic oil is fed into the lower part of the casing 202a of the transmission mechanism 202 to the extent that the lower half of the hydraulic oil pump 213 is immersed. Hydraulic oil is slung by the rotation of the cam ring 220 for lubricating parts and supplied through the through-hole 251 a into the vane pump 200.

Thus, hydraulic oil flows on a wall surface of the casing 202a, reaches an oil guide 253 where it is collected, and is supplied from the end of the oil guide 253 into the through-hole 251 a.

Hydraulic oil is also supplied through the hydraulic oil supply paths 235 and 235' and the communication passages 235a and 235'a to the first hydraulic oil passage 226, the port 222a and the bushing 252.

Thus, hydraulic oil is collected by the oil guide 253, thereby permitting the through-hole 251 a to be always supplied with hydraulic oil even if the oil level is below the through-hoie 251 a.

The hydraulic oil supply paths 235 and 235' are slanted, permitting a sufficient amount of hydraulic oil to be supplied by a centrifugal force into the first hydraulic oil passage 226 comprising the hydraulic control circuit 221.

When the through-hole 251a rotates, hydraulic oil is sucked by the function of centrifugal force, thereby preventing air from being brought into the center area of the end of the through-hole 251a and coming into hydraulic oil.

Then, hydraulic oil is supplied to the first hydraulic oil passage 226, port 222a and the spline 257, being prevented from back flow by the function of the check valves 228 and 229.

Hydraulic oil is guided through the first hydraulic oil passage 226 disposed in the cover 251 to the ports 222b, 223b and 224b and the end section 233a of the relief valve 233, or through the port 222a and the second hydraulic oil passage 227 to the ports 223a and 224a and the end section 231 a of the relief valve 231.

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

The second hydraulic oil passage 227 comprises a ring groove formed on the outer peripheral surface of the pressure retainer 241 and the inner peripheral surface of the flange 245 engaging with the outer peripheral of the pressure retainer 241.

Heretofore, such a hydraulic oil passage has been formed by using such a difficult machining procedure that long holes are drilled in various directions, these drilled holes are connected, and blank plugs are put into the holes. However, the configuration in the present invention described above facilitates manufacture of the hydraulic oil passage.

The number of the vanes 218 is determined as an integer that is not an integer multiple of the number of the pump chambers 236a, 236b and 236c, resulting in different phases of change in pressure of hydraulic oil delivered from individual delivery ports as shown in Figure 27. (Referto Figures 27(b), (c) and (d).) As a result, there is no increase of pulsation of hydraulic oil pressure in the first hydraulic oil passage 226 or the second hydraulic oil passage 227 where these pressures P1, P2 and P3 of delivered hydraulic oil are combined.

Thus, referring to Figure 27(a), a variation H2 due to pulsation of a hydraulic oil pressure Po in the first hydraulic oil passage 226 or the second hydraulic oil passage is suppressed to below each variation H2 of the pressures P1, P2 and P3 of delivered hydraulic oil.

Then, a valve lifting hydraulic circuit to lift the vane 218 will now be described below.

Referring to Figures 24(a), (b) and (c), the pressure retainer 241, the cover 251 and the rotor 219 form a valve lifting hydraulic circuit 250.

In the rotor 219, there is provided a valve lifting hydraulic oil passage 250a through the rotor 219 to its both sides at an end section of a vane groove 219 engaging with the vane 218. By supplying a hydraulic oil pressure into the valve lifting hydraulic oil passage 250a, the vane 218 is pushed up and the end of the vane 218 comes in positive contact with the inner peripheral surface of the cam ring 220, thereby ensuring pressurization in the pump chambers 236a, 236b and 236c.

The valve lifting hydraulic oil passage 250a is connected to ring recesses 250b and 250c disposed at the both ends of the rotor 219, and the pressure retainer 241 is provided with a communication passage 250d opening at opposing position of the ring recess 250b.

The cover 251 is provided with a communication passage 250e opening at opposing position of the ring recess 250c.

The communication passages 250d and 250e are disposed extending slantly out of the ring recesses 250b and 250c toward the second and first hydraulic oil passages 227 and 226, through check valves 250f and 250g respectively.

Since the check valves 250f and 250g contain spherical valve members and the communication passages 250d and 250e extend out to the outer peripheral, each valve member is pressed against each valve seat by a centrifugal force produced by the rotation of the pressure retainer 241 and the cover 251 and these check valves are normally kept closed.

This ensures the valve lifting hydraulic oil passage 250a to be supplied with a hydraulic pressure from the first or second hydraulic oil passage 226 or 227 whichever higher.

Thus, either one of the first hydraulic oil passage 226 or the second hydraulic oil passage 227 becomes the suction side and the other becomes delivery side depending on the direction of relative rotation of the cam ring 220 and the rotor 219. However, the valve lifting hydraulic oil passage 250a is always supplied with a hydraulic oil pressure from the first hydraulic oil passage 226 or the second hydraulic oil passage 227, and the vane 218 is always pressed against the inner peripheral surface 220d of the cam ring 220, no matter which rotation is faster, the front wheels or the rear wheels. An elastic member such as spring may be inserted at the end section of the vane groove 219b into which the vane 218 is inserted.

The operation of thus formed power transmission apparatus will be described below.

In the usual straightforward running state 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 rate of slip of the tires under rotation is quite small. In such a state, there occurs no difference in rotation speed between the first rotary shaft 211 and the second rotary shaft 214 connected to the hydraulic oil pump unit 213. Accordingly, no hydraulic oil under pressure is delivered from the vane pump 200, and no drive force is transmitted to the rear wheels 216. Thus, the vehicle is driven by the front wheels 209 only. That is, the vehicle runs in the two-front-wheel drive mode.

However, when the vehicle running straightforward is, for example, accelerated, a slip of less than about 1% occurs usually on the front wheels 209 although it is not appreciable. There occurs a difference in rotation speed between the first and second rotary shafts 211 and 214 due to the above slip of the front wheels 209.

In such a case, the vane pump 200 is energized to build up the pressure corresponding to the above rotation speed difference. The rotor 219 and the cam ring 220 rotate integrally, and the drive force corresponding to the built-up pressure and the pressure receiving area of the vanes is transmitted to the rear wheels 216 to establish the four-wheel drive mode.

The flow of hydraulic oil in the vane pump 200 in this case is as shown in Figure 18(a). It will be seen in Figure 18(a) that, due to the rotation of the rotor 219 relative to the cam ring 220, the ports 222b, 223b and 224b act as the suction ports, the hydraulic oil is sucked from the hydraulic oil reservoir 230 into the suction ports 222b, 223b and 224b through the check valve 229. On the other hand, the ports 222a, 223a and 224a act as the delivery ports. Accordingly, hydraulic oil is fed through the first hydraulic oil passage 226 and, at the same time, delivered through the second hydraulic oil passage 227 toward the relief valve 231.

In Figure 18(a), the solid lines with arrows indicate the flowing directions of delivered hydraulic oil and the broken lines with arrows indicate the flowing directions of sucked hydraulic oil.

Suppose then that the rotation speed of the front wheels 209 becomes very high compared with that of the rear wheels 216, for example, when the vehicle runs on a snow-laden road or is abruptly accelerated or abruptly braked resulting in locking of the rear wheels 216. In such a case, the difference in rotation speed between the first and second rotary shafts 211 and 214 connected to the hydraulic oil pump unit 213 becomes very large. Consequently, a high pressure is generated in the vane pump 200 and such hydraulic oil flows in the directions shown in Figure 18(a). When the pressure of hydraulic oil exceeds a predetermined level, the relief valve 231 is opened against the force of the spring 234, and the pressure of delivered hydraulic oil is controlled to be substantially constant. Thus, a constant drive force corresponding to the regulated pressure of delivered hydraulic oil is transmitted to the rear wheels 216 to establish the four-wheel drive mode.

Consequently, the rotation speed of the front wheels 209 is decreased, while that of the rear wheels 216 is increased, so that the difference in rotation speed between the front and rear wheels 209 and 216 is reduced.

(This function is the same as that of the non-slip differential unit.) Thus, when a slip occurs on the front wheels 209, the driving torque for the rear wheels 216 is increased to avoid the impossibility of running, while, when the rear wheels 216 tend to be locked, the braking torque for the front wheels 209 is increased to prevent locking of the rear wheels 216.

Suppose further the case where the rotation speed of the rear wheels 216 is very high compared with that of the front wheels 209 as, for example, when the front wheels 209 tend to be locked due to be locked due to application of the brakes. In such a case, a very large difference appears between the rotation speeds of the first and second rotary shafts 211 and 214 connected to the hydraulic oil pump unit 213 in a direction opposite to the aforementioned direction.

Consequently, the flow of hydraulic oil in the vane pump 200 is now in a direction opposite to that shown in Figure 18(a). It will be seen in Figure 18(b) that the ports 222a, 223a and 224a act as suction ports for hydraulic oil, and hydraulic oil is sucked from the hydraulic oil reservoir 230 into the suction ports 222a, 223a and 224a through the check valve 228, while, on the other hand, the ports 222b, 223b and 224b act as delivery ports. Accordingly, hydraulic oil is fed through the second hydraulic oil passage 227 and, at the same time, delivered through the first hydraulic oil passage 226 toward the relief valve 233.Since the pressure of such hydraulic oil is also maintained constant by the relief valve 233, the corresponding constant drive force is transmitted to the rear wheels 216 to establish the four-wheel drive mode. As a result, the braking torque imparted to the rear wheels 216 is increased to prevent locking of the front wheels 209.

In a usual running state of the vehicle running along a curve, the rotation speed of the front wheels 209 is slightly higher than that of the rear wheels 216, and the vehicle runs along the curve in the four-wheel drive mode in which the braking torque is imparted to the front wheels 209, while the driving torque is imparted to the rear wheels 216.

In the manner described above, the delivery pressure of hydraulic oil is controlled so as not to exceed a constant value by the relief valves 231 and 233 in the hydraulic oil pump unit 213. Therefore, in contrast to the prior art case where the driver's manipulation is required to establish the four-wheel drive mode in a vehicle of the part-time four-wheel drive type, the change-over between the four-wheel drive mode and the two-wheel drive mode can be automatically effected, and the four-wheel drive mode is established by the drive force corresponding to the difference in rotation speed between the front and rear wheels, according to the present invention.

Also, when compared with the center differential unit necessarily provided in a prior art vehicle of the full-time four-wheel drive type, the hydraulic oil pump 213 according to the present invention is small in size, compact in structure, light in weight and low in cost.

Figure 25 shows a eleventh embodiment of the present invention in which the structure of hydraulic oil supply paths 235 and 235' differs from that employed in the tenth embodiment of the present invention.

In the eleventh embodiment of the present invention, the hydraulic oil supply paths 235 and 235' extend slantly out toward the outer peripheral of the rotor 219 in the cover 251 up to the first hydraulic oil passage 226. Consequently, a centrifugal force is generated by the rotation of the cover 251, and a sufficient amount of hydraulic oil is fed into the first hydraulic oil passage 226 and the port 222a. Since this structure is the same as of a centrifugal separator, air is collected to the center of the through-hole 251 a and is not sucked into the hydraulic oil supply paths 235 and 235'.

Atwelfth embodiment of the present invention in which the structure of the hydraulic oil supply paths 235 and 235' differs from that employed in the tenth embodiment will be described with reference to Figure 26.

In the embodiment shown in Figure 26, the end section of the rear wheel drive shaft 243 is provided with a hydraulic oil supply path 235" which opens at the end section and extends to the postion of the first hydraulic oil passage 226, and, at the same time, a communication passage 235" which causes the hydraulic oil supply path 235" to communicate with the first hydraulic oil passage 226. Consequently, a centrifugal force is generated by the rotation of the rear wheel drive shaft 243, and a sufficient amount of hydraulic oil is fed into the first hydraulic oil passage 226.

The bushing 252 does not extend to the opening of the communication passage 235" on the peripheral surface of the rear wheel drive shaft 243, and a ring-form hydraulic oil passage is formed between the spline 257 and the end surface of the bushing 252 which permits hydraulic oil to be fed to the port 222a in addition to the first hydraulic oil passage 226. The spline 257 and the bushing 252 are also supplied with hydraulic oil as a lubricant.

A thirteenth embodiment of the present invention in which the structure of the pressure retainer 241 differs from that employed in the tenth embodiment will be described with reference to Figures 28 through 217. In Figures 28 through 217, the same parts already described are referred to the same numerals and are not described here.

The rear wheel drive shaft 243 extends out through a through-hole 345b disposed at the center of a flange 345 as a plate retainer. A pressure retainer 341 as a pressure plate is disposed between the inner wall surface of the rotor 219 side of the flange 345 and the end surface of the rotor 219. The pressure retainer 341 comprises a large-diameter section 341 b of which the diameter is slightly larger than that of the rotor 219 and a small-diameter section 341c of which the diameter is slightly smaller than that of a through-hole 345b of the flange 345.

The pressure retainer 341 is disposed on the flange 345 so that the large-diameter section 341 b is inserted into a recess 345a formed on the flange 345 and the small-diameter section 341 c is inserted into the through-hole 345b. An O-ring 358c is inserted between the large-diameter section 341 b of the pressure retainer 341 and the recess 345a of the flange 345, and an O-ring 358b is inserted between the small-diameter section 341c and the through-hole 345b.

Further, the pressure retainer 341 is disposed on the rear wheel drive shaft 243 so that a bushing 256 disposed between the through-hole 341 a and the rear wheel drive shaft 243 permits the pressure retainer 341 sliding in the axial direction of the rear wheel drive shaft 243 and rotating relative to the rear wheel drive shaft 243.

Thus, the pressure retainer 341 is movable in the axial direction of the rear wheel drive shaft 243, relative to the rear wheel drive shaft 243 and the flange 345, and the rotor 219 side end surface of the pressure retainer 341 can protrude into the inner peripheral surface of the cam ring 220. Also, the presure retainer can rotate together with the flange 345 which is 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 section 341 b, to the side of the small-diameter section 341c, in the pressure retainer 341.

The internal pressure of the hydraulic oil chamber 361 is maintained with the O-rings 358c and 358b, and that of the pump chambers 236a, 236b and 236c is maintained with the pressure retainer 341 and the O-ring 358c. As shown in Figure 29, the pressure retainer 341 is provided with a side clearance adjusting hydraulic oil passage 350h which permits the vane lifting hydraulic oil passage 250a to communicate with the hydraulic oil chamber 361. This allows the pressure of hydraulic oil delivered from the pump chambers 236a, 236b and 236c to the vane lifting hydraulic oil passage 250a to be fed into the hydraulic oil chamber 361.

The pressure of hydraulic oil delivered from the vane pump 200 acts as a pressure required to press the pressure retainer against the rotor 219, and the pressure retainer 341 slides in the axial direction of the rear wheel drive shaft 243 to automatically adjust the side clearance between the end surface of the pressure retainer 341 and that of the rotor 219. The hydraulic oil chamber 361 is provided with a spring 362 which maintains a required side 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 hydraulic oil delivered from the pump chambers 236a, 236b and 236c is low, maintaining the delivery pressure of hydraulic oil from the pump chambers 236a, 236b and 236c when the vane pump 200 starts rotating.

Further, as shown in Figures 30 and 31, a chamfer 219c is provided on the inner radial edge line of the vane groove 21 9b of the rotor 219. Thus, the chamfer 219c prevents discontinuation of oil film at the edge line of the vane groove 219b which causses seizure between the end surfaces of the rotor 219 and the pressure retainer 341. The chamfer 219c extends within the inner peripheral of the rotor 219 and does not reach the ports 222a, 222b, 222c, 223a, 223b and 223c formed by the flange 345. This will prevent leakage of hydraulic oil pressure from the vane lifting hydraulic oil passage 250a to the suction ports, the ports 222a, 222b and 222c or the ports 223a, 223b and 223c.

Since the vane pump employed in the power transmission apparatus according to the present invention is driven by the difference in the rotation speeds beween the rotor and the cam ring, its rotation speed range is lower compared with conventional vane pumps and, at the same time, it is required to be compact in structure and to have a large torque-transmitting capacity. Therefore, the vane pump employed in the present invention is required to have a lifting amount of vane as large as possible and many ports. However, if, in order to apply a conventional vane pump to the present invention, the lifting amount of vane is simply increased, a bending moment will be applied to the vane due to a high hydraulic oil pressure applied to the lateral surface in the rotating direction of the vane, and the vane may be inclined resulting in temporary sticking to the rotor.In such a case, when the rotor rotates eccentrically due to variation in the pressure of delivered hydraulic oil or a play produced in manufacture or assembly, there occurs such a defect that a gap is formed between the end of the vane and the inner peripheral surface of the cam ring, resulting in oil leakage or a pressure drop of hydraulic oil. This defect is especially remarkable with the vane pump employed in the present invention which is operated in a lower range of rotation speed.

In order to solve the above problems, a preferred shape of vane pump to be employed in the power transmission apparatus according to the present invention has been experimentally determined, which will now be described with reference to Figure 32 showing a sectional view taken on a plane perpendicular to the rotary shaft of a vane pump 200.

Suppose that a vane 218 is applied with a hydraulic oil pressure P delivered from the vane pump 200 as a vane lifting force, where t is the thickness of the vane 218, e1 is the projection length of the vane 218 from a rotor 219 at the time of the maximum lifting, t2 is the length of part of the vane 218 in a vane groove 219b in the above state, and 4 (= t1 + t2) is the total length of the vane 218.Further, in the projected end of the vane 218, a part slightly shifted to a surface A in the direction of its rotation is the most projected, and this part is referred to a sliding section. The sliding section is positioned at a distance (tl < t2) from the surface A.

When the vane pump changes over from the suction process to the delivery process, the vane 218 is applied with a force Fo of delivered hydraulic oil and comes in contact with the rotor 219 at points a and b, thereby being subjected with a bending moment. Even if the rotor rotates eccentrically, a relative sliding must occur between the vane 218 and the rotor 219 at the points a and b to maintain a close contact between the projected end of the vane 218 and an inner peripheral surface 220d of a cam ring 220. Such a condition is determined as described below.

First, a force applied to the vane 218 is determined.

The moment balance with respect to the point a is given by the following equation.

e1l2 Fo F2 t2 (1) The bending moment due to the delivery hydraulic oil pressure P is given by the following equation.

Fo = fl3tP ---------------------------------------- (2) The lifting force of the vane 218 is qi = t.p ---------------------------------------- (3) The balance equation in the rotating direction is Fo + F2 = Fl ---------------------------------------- (4) The balance equation in the radial direction is t1P+F1+F2)=q1 F2) From the equations (1) through (4) F1 = t12/2-f2-P ---------------------------------------- (6) F2 = t1-(1 + t1/2t2)-P ---------------------------------------- (7) Substituting the equation (5) for the equations (6) and (7), tl - t2 = #c.# (1/#2.1-------------------------------- (8) Since the condition where sliding occurs between the rotor 219 and the vane 218 is left side right side in the equation (5), tl - t2 > Ac-f11f2-1 ---------------------------------------- (9) (where c is a critical friction coefficient.) and the shape of the vane 218 can be determined so that the equation (9) above is satisfied.

The rotor 219 is generally made of a nickel-chromium steel (SNCN) and the vane 218 is made of a high-speed steel (such as SKH9) or the like. Since the critical friction coefficient #c of the same materials is approximately 0.04 to 0.05, #c = 0.045 is employed here as a typical value, and the shape of the 218 can be determined to satisfy the following formula.

tl - t2 > 0.045-f1/f2-t Further, since the vane pump employed in the present invention is preferably of a high-lift type, the shape of the vane 218 can be determined to satisfy 1l2 > 0.9 so that the length 11 of the lifting section is relatively long with respect to the total length 1. (In conventional vane pumps, it is approximately 1lf2 < 0.8.) Although a vane pump or a gear pump has been illustrated as an example of the hydraulic oil pump preferably employed in the aforementioned embodiments of the present invention, and, in the case of the vane pump, that of balanced type it, is apparent that a vane pump of unbalanced type having two ports alternately acting as delivery and suction ports may also be employed depending on the quantity of transmitted drive force, and a hydraulic oil pump of any other suitable type such as a trochoid pump, a hypocycloid pump, an axial plunger pump or a radial plunger pump may be also employed. It is the only essential requirement that the pump is of the type which delivers an amount of hydraulic oil corresponding to the rotation speed difference.The transmission may be any one of the manual type and the automatic type. Also, the manner of control of the relief valve is in no way limited to the duty control and may be any other suitable type of mechanical control.

The present invention is in no way limited to its application to a vehicle of four-wheel drive type and can also be utilized for power transmission to front and rear wheels of a vehicle of six-wheel drive type. The present invention can also be utilized for power transmission to front-side front and rear wheels of a vehicle of front two-axle drive type and for power transmission to rear-side front and rear wheels of a vehicle of rear two-axle drive type.

It will be understood from the foregoing detailed description that the present invention provides a power transmission apparatus for a vehicle, which comprises a first rotary shaft transmitting power to front wheels, a second rotary shaft transmitting power to rear wheels, and a hydraulic oil pump connected between the first rotary shaft and the second rotary shaft to be driven according to the difference between the rotation speed of the first and second rotary shafts thereby delivering hydraulic oil of an amount corresponding to the rotation speed difference. The hydraulic oil pump transmits the drive force with the static pressure of delivered hydraulic oil to establish the four-wheel drive mode and has delivery ports and suction ports automatically changed over depending on the direction of relative rotation of the first and second rotary shafts. Therefore, the four-wheel drive mode can be established without requiring any manipulation by the driver. Thus, the present invention can obviate the trouble such as the tight corner braking phenomenon encountered with the vehicle of the part-time four-wheel drive type while, at the same time, eliminating the troublesome driver's manipulation for driving. The power transmission apparatus of the present invention has the advantages of small size, light weight, simple construction and low cost over the conventional center differential unit equipped in the vehicle of full-time four-wheel drive type.

Claims (51)

1. A power transmission apparatus for a vehicle comprising a first rotary shaft transmitting drive force to front wheels, a second rotary shaft transmitting drive force to rear wheels, a hydraulic oil pump connected between said first rotary shaft and said second rotary shaft to be driven according to the difference between the rotation speeds of said first and second rotary shafts thereby delivering hydraulic oil of an amount corresponding to the rotation speed difference, said hydraulic oil pump having at least two ports alternately changed over between the delivery side and the suction side depending on the direction of relative rotation of said first and second rotary shafts, and a hydraulic control circuit including a hydraulic oil passage providing communication of one of said ports with the other and hydraulic control means dipsoed in said hydraulic oil passage for controlling the pressure of hydraulic oil delivered from said hydraulic oil pump.

2. A power transmission apparatus as claimed in Claim 1,wherein said first rotary shaft is coupled to the vehicle's engine so that said front wheels can be directly driven by said engine.

3. A power transmission apparatus as claimed in Claim 1, wherein said second rotary shaft is coupled to the vehicle's engine so that said rear wheels can be directly driven by said engine.

4. A power transmission apparatus as claimed in any preceding claim wherein said hydraulic oil passage in said hydraulic control circuit includes a first hydraulic oil passage communicating with one of said ports and having a first valve unit permitting flow of hydraulic oil into said hydraulic oil pump only, a second hydraulic oil passage communicating with the other of said ports and having a second valve unit permitting flow of hydraulic oil into said hydraulic oil pump only, a first communication passage capable of establishing communication of said second hydraulic oil passage with the portion of said first hydraulic passage between one of said ports and said first valve unit, and a second communication passage capable of establishing communication of said first hydraulic oil passage with the portion of said second hydraulic oil passage between the other of said ports and said second valve unit, and wherein said hydraulic control circuit includes selector means for changing over between said first communication passage and said second communication passage.

5. A power transmission apparatus as claimed in Claim 1, wherein said hydraulic oil passage in said hydraulic control circuit includes a first hydraulic oil passage communicating at one end thereof with one of said ports and having a first valve unit permitting flow of hydraulic oil into said hydraulic oil pump only, a second hydraulic oil passage communicating at one end thereof with the other of said ports and having a second valve unit permitting flow of hydraulic oil into said hydraulic oil pump only, and a delivery passage capable of providing communication between the other end of said first hydraulic oil passage and the other end of said second hydraulic oil passage, and wherein said hydraulic control circuit includes selector means for selectively establishing communication of the other end of said first hydraulic oil passage and the other end of said second hydraulic oil passage with said delivery passage.

6. A power transmission apparatus as claimed in Claim 4 or 5, where said selector means operates to permit delivery of hydraulic oil from said hydraulic oil pump.

7. A power transmission apparatus as claimed in Claim 4, wherein said hydraulic control means is disposed in each of said first communication passage and said second communication passage.

8. A power transmission apparatus as claimed in Claim 5, wherein said hydraulic control means is disposed in said delivery passage.

9. A power transmission apparatus as claimed in any preceding claim, wherein said hydraulic control means includes a valve member biased by biasing means.

10. A power transmission apparatus as claimed in any preceding claim, wherein said hydraulic control means is controlled depending on one or more operating parameters of the vehicle.

11. A power transmission apparatus as claimed in Claim 10, wherein one operating parameter is the rotation speed of the engine of the vehicle.

12. A power transmission apparatus as claimed in Claim 10 or 11, wherein one operating parameter is the rotation speed of said first rotary shaft.

13. A power transmission apparatus as claimed in any of Claims 10 to 12, wherein one operating parameter is the rotation speed of said second rotary shaft.

14. A power transmission apparatus as claimed in any of Claims 10 to 13, wherein one operating parameter is the throttle opening.

15. A power transmission apparatus as claimed in any of Claims 10 to 14, wherein one operating parameter is the degree of brake actuation.

16. A power transmission apparatus as claimed in any of Claims 10 to 15, wherein one operating parameter is information indicative of the rate of steering.

17. A power transmission apparatus as claimed in any preceding claim, wherein said hydraulic control means includes flow restricting means.

18. A power transmission apparatus as claimed in Claim 17, wherein said flow restricting means is an orifice.

19. A power transmission apparatus as claimed in Claim 17 or 18, wherein said flow restricting means is controlled depending on the negative pressure in the intake manifold of the vehicle's engine.

20. A power transmission apparatus as claimed in any of Claims 17 to 19, wherein said flow restricting means is controlled depending on the delivery pressure of an oil pump in a power steering mechanism.

21. A power transmission apparatus as claimed in any of Claims 17 to 20, wherein said flow restricting means is controlled depending on both the negative pressure in the intake manifold of the engine and the delivery pressure of an oil pump in a power steering mechanism.

22. A power transmission apparatus as claimed in any preceding claim, wherein said hydraulic control means includes a valve member biased by biasing means, and flow restricting means.

23. A power transmission apparatus as claimed in Claim 5, wherein said hydraulic control means includes flow restricting means disposed in an auxiliary passage providing communication of the portion of said first hydraulic oil passage between said first valve unit and one of said ports with the portion of said second hydraulic oil passage between said second valve unit and the other of said ports, and a valve member disposed in said delivery passage and biased by biasing means.

24. A power transmission apparatus as claimed in any preceding claim, wherein said hydraulic oil pump is a vane pump.

25. A power transmission apparatus as claimed in Claim 24, wherein said vane pump includes two or more pump chambers.

26. A power transmission apparatus as claimed in Claim 25, wherein each of said pump chambers includes at least two ports.

27. A power transmission apparatus as claimed in any of Claims 24 to 26, wherein the number of vanes of said vane pump is an integer that is not an integer multiple of the number of said pump chambers.

28. A power transmission apparatus as claimed in any of Claims 24 to 27, wherein said vane pump comprises a pump body and a rotor that can rotate in said pump body.

29. A power transmission apparatus as claimed in Claim 28, wherein said pump body comprises a cam ring sliding with said vane and a housing supporting said cam ring from its both sides.

30. A power transmission apparatus as claimed in Claim 28 or 29, wherein said hydraulic control circuit is disposed in said pump body.

31. A power transmission apparatus as claimed in any of Claims 28 to 30, wherein said vane is formed into a shape satisfying a condition of t1 - t2 > ac-f1/f2-t where t 1 is a maximum projection length of said vane from said rotor attached slidingly to said rotor, C2 is a length of part of said vane inserted in said rotor, e (= t1 + t2) is a total length of said vane, t is a thickness of said vane, the position of the end of vane sliding against an inner peripheral surface of said cam ring is set at a distance of tl from the surface of rotation direction side of said vane, and ilc is a critical friction coefficient between said rotor and said vane.

32. A power transmission apparatus as claimed in Claim 29, wherein said rotor is applied with a force in the direction of its rotary shaft to maintain an adequate clearance between said housing and said rotor.

33. A power transmission apparatus as claimed in Claim 29, wherein said pump chamber is a space formed by said housing, said cam ring and said rotor, and said hydraulic oil passage provides communication through said cam ring of one of said ports formed in said housing on one side of said pump chamber with the other port formed in said housing on the other side of said pump chamber.

34. A power transmission apparatus as claimed in any of Claims 1 to 23, wherein said hydraulic oil pump is a gear pump.

35. A power transmission apparatus as claimed in Claim 34, wherein said gear pump includes a rotatable gear pump housing coupled to said first rotary shaft, a sun gear rotatably supported in said housing and coupled to said second rotary shaft, and a pinion gear rotatably supported in said housing and making meshing engagement with said sun gear.

36. A power transmission apparatus as claimed in Claim 35, wherein said gear pump housing is provided with said hydraulic control circuit.

37. A power transmission apparatus as claimed in any preceding claim, wherein said hydraulic oil pump and said hydraulic control circuit integrally comprises a hydraulic oil pump unit.

38. A power transmission apparatus as claimed in Claim 37, wherein said hydraulic oil pump unit is connected to an output shaft of a transmission mechanism of said engine.

39. A power transmission apparatus as claimed in Claim 38, wherein said hydraulic oil pump unit is contained in a same casing together with said transmission mechanism.

40. A power transmission apparatus as claimed in Claim 39, wherein said hydraulic oil pump unit is disposed below said transmission mechanism and in a lower part of said casing.

41. A power transmission apparatus as claimed in Claim 38, wherein said transmission mechanism is a transmission mechanism for manual manipulation having an input shaft driven by said engine and an output shaft disposed approximately parallelly to said input shaft.

42. A power transmission apparatus as claimed in Claim 38, wherein said hydraulic pump includes a gear section making meshing engagement with a gear provided on said output shaft of said transmission mechanism.

43. A power transmission apparatus as claimed in Claim 42, wherein said gear section is disposed on an outer peripheral surface of said cam ring of said vane pump.

44. A power transmission apparatus as claimed in Claim 42, wherein said gear section is disposed on an outer peripheral surface of said gear pump housing rotatably supporting a pinion gear making meshing engagement with said sun gear of said gear pump.

45. A power transmission apparatus for a vehicle comprising a front wheel drive path to transmit drive force to front wheels of the vehicle, a rear wheel drive path to transmit drive force to rear wheels of the vehicle, a reversible hydraulic oil pump connected between first and second portions of the front wheel or rear wheel drive path to be driven according to the speed difference between the first and second drive path portions so that the pump delivery depends of said speed difference, said hydraulic oil pump forming part of a closed hydraulic circuit including a hydraulic control means arranged to control the delivery of the pump.

46. A power transmission apparatus as claimed in Claim 1, wherein the pump is disposed in the rear wheel drive path.

47. A power transmission apparatus as claimed in Claim 1, wherein the pump is disposed in the front wheel drive path.

48. A power transmission apparatus as claimed in any preceding claim, wherein the pump forms part of two hydraulic circuits, one of which circuits includes a one-way valve to permit flow in one direction only through the pump, and the other of which circuits includes a one-way valve to permit flow in the other direction only through the pump, and the hydraulic control means including a selector operable to select one or the other of said circuits.

49. A power transmission apparatus as claimed in any of Claims 45 to 48 and having the additional features recited in any of Claims 2 to 44.

50. A power transmission apparatus substantially as hereinbefore described with reference to and as illustrated in the accompanying drawings.

51. A four or more wheel drive motor land vehicle having a power transmission apparatus as claimed in any preceding claim.