JP3642625B2 - Drive device for four-wheel drive vehicle - Google Patents

Description

[0001]
BACKGROUND OF THE INVENTION
The present invention relates to a drive device for a four-wheel drive vehicle used in a horizontally mounted engine.
[0002]
[Prior art]
Conventionally, there is a prior art disclosed in Japanese Patent Application Laid-Open No. 4-83948 regarding a vehicle drive device in which an engine is horizontally disposed. In this prior art, an engine, a torque converter, a forward / reverse switching device equipped with a double pinion planetary gear, and a belt type continuously variable transmission are provided coaxially in the vehicle body width direction, and output from the secondary shaft of the continuously variable transmission is provided. A transmission arrangement is shown for the differential device.
[0003]
[Problems to be solved by the invention]
However, in the above-mentioned prior art, a horizontally disposed engine is provided with a torque converter, a forward / reverse switching device, and a belt-type continuously variable transmission on the same axis as the engine. A differential device is provided below the secondary shaft of the step transmission, and a transmission case in which these are integrated is joined. Therefore, the overall length of the drive device is long in the vehicle width direction, and the engine room side wall and the drive device in the vehicle-mounted state If you try to secure a sufficient crash stroke at the time of a side collision, the degree of freedom in designing the vehicle body will be limited, and it will be difficult to obtain a working space in the engine room. May be disturbed.
[0004]
Further, in the drive device for a four-wheel drive vehicle based on this drive device, a center differential device is further provided on the secondary shaft side of the belt type continuously variable transmission, so that the structure and the control device for controlling them are complicated. There is a problem such as a cost increase.
[0005]
Also, it is desirable that the engine room structure having the same shape has in-vehicle compatibility with a belt-type continuously variable transmission, a manual transmission (manual transmission, MT), an automatic transmission (automatic transmission, AT), etc. If the overall length, the outer circumference of the transmission case, and the so-called waistline dimensions are made substantially the same as a manual transmission that can be designed relatively compactly, it becomes possible to share a support member and an exhaust system for mounting on a vehicle.
[0006]
Therefore, the object of the present invention is to reduce the width of the drive unit, particularly the transmission case, in the width direction of the vehicle, and can be mounted in a conventional engine room while ensuring a working space for freedom of vehicle body design, crash stroke, and transmission / removal. In addition, it is an object of the present invention to provide a drive device for a four-wheel drive vehicle that can simplify the structure and its control device.
[0007]
[Means for Solving the Problems]
  A drive device for a four-wheel drive vehicle according to the present invention that achieves the above-described object includes a horizontally mounted engine, a transmission to which an output from the engine is input, and one of the engines arranged in parallel to the crankshaft of the engine. First and second drive shafts for transmitting power to the differential device and the other differential device, respectively;A double pinion planetary gear disposed coaxially with the first drive shaft; input switching means for selectively transmitting power from the transmission to the ring gear and sun gear of the planetary gear; and output from the carrier of the planetary gear. Means for transmitting power to the second drive shaft, a third friction engagement element interposed between the first drive shaft and the second drive shaft so as to be able to transmit power, and from the sun gear of the planetary gear. A fourth friction engagement element for selectively transmitting power to the first drive shaft; and a fifth friction engagement element for selectively locking the ring gear rotation of the planetary gear. And the first and second power distribution and forward / reverse switching of the input from the transmission at a predetermined ratio via the planetary gear by selectively actuating each friction engagement element. Wherein the power transmission to the live shaft.
[0009]
DETAILED DESCRIPTION OF THE INVENTION
Hereinafter, embodiments of the present invention will be described with reference to the drawings.
[0010]
In FIG. 1, a drive system of a drive device for a four-wheel drive vehicle with a belt type continuously variable transmission will be described as a drive device for a four-wheel drive vehicle to which the present invention is applied.
[0011]
Reference numeral 10 denotes a horizontal engine, which is a torque converter case 1 that is joined to the engine 10 and accommodates the torque converter 20, a belt type continuously variable transmission 30 that is located on the side of the torque converter case 1 and one front differential. Devices, for example, a differential and converter housing 2 and a side case 3 for accommodating a front differential device 40, a case 4 and an end cover 5 for accommodating a transfer unit 50 in cooperation with the torque converter case 1, and a rear side of the torque converter case 1. An extension case 6 that is positioned and accommodates a power transmission mechanism that transmits the output from the transfer unit 50 to the rear wheels is joined to form a transmission case 7, and an oil pan (not shown) is provided below the transmission case 7. It is.
[0012]
The crankshaft 11 of the horizontally mounted engine 10 is connected to the torque converter 20 inside the torque converter case 1, and the input shaft 21 from the torque converter 20 is connected to the primary shaft 31 of the belt type continuously variable transmission 30 inside the differential and converter housing 2. By connecting, the power from the crankshaft 11 is transmitted to the primary shaft 31 of the continuously variable transmission 30 via the torque converter 20.
[0013]
The power continuously variable by the continuously variable transmission 30 is output to the secondary shaft 32 and input to the transfer unit 50 via the counter shaft 39 and the like, and the transfer unit 50 is configured to transmit to the front wheels via the front differential device 40. On the other hand, it is configured to be transmitted to the rear wheel via the propeller shaft 116 and the other differential device, for example, the rear differential device 117.
[0014]
In the transmission case 7, there is provided an oil pump 8 which is always driven by being connected to an oil pump drive shaft 24a provided in the torque converter 20. The oil pump 8 always generates hydraulic pressure and supplies it to the torque converter 20 and the like. The hydraulic pressure of the continuously variable transmission 30 can be controlled, and the hydraulic pressure is based on signals from the vehicle speed sensor 9a, the throttle sensor 9b, the shift switch 9c, the front wheel speed sensor 9d, the rear wheel speed sensor 9e, the steering angle sensor 9f, and the like. The control unit 9 controls the hydraulic pressure of the transfer unit 50.
[0015]
Next, the torque converter 20, the belt type continuously variable transmission 30, the differential device 40, and the transfer unit 50 will be sequentially described with reference to FIGS.
[0016]
As shown in FIG. 2, the torque converter 20 has an input shaft 21 that is rotatably supported coaxially with respect to the crankshaft 11 on the differential and converter housing 2 and the side cover 3 via a ball bearing 21a. have.
[0017]
The outer periphery of the input shaft 21 is substantially cylindrical, and a flange portion provided at the base end is rotatably surrounded by a stator shaft 22 bolted to the torque converter case 1 with an oil pump housing 8c interposed therebetween. The oil pump drive shaft 24a integrally coupled to the impeller 24 is rotatably fitted.
[0018]
The outer periphery of the impeller 24 is integrally coupled to the outer periphery of the front cover 25 and is coupled to the crankshaft 11 via the drive plate 26 so as to be rotated integrally with the crankshaft 11.
[0019]
A turbine 27 that is fitted to the input shaft 21 is disposed so as to face the impeller 24, and a stator 28 that is supported on the stator shaft 22 via a one-way clutch 28 a is interposed between the impeller 24 and the turbine 27.
[0020]
Further, a lock-up clutch 29 is interposed between the turbine 27 and the front cover 25. An inner gear 8a that is rotationally driven by an oil pump drive shaft 24a at the base end of the stator shaft 22, an outer gear 8b that meshes with the inner gear 8a, and the aforementioned An oil pump 8 having an oil pump housing 8c is provided.
[0021]
When the crankshaft 11 of the engine 10 rotates, the impeller 24 is rotationally driven through the drive plate 26, the front cover 25, etc. coupled to the crankshaft 11.
[0022]
The rotation of the impeller 24 causes the oil in the impeller 24 to be discharged to the outside by centrifugal force. The oil flows from the outside of the turbine 27 and transmits torque in the same direction as the rotation of the impeller 24 to the turbine 27. The input shaft 21 for spline fitting is driven to rotate. Furthermore, the torque of the impeller 24 is increased by inverting the outflow direction of the oil flowing out of the turbine 27 by the stator 28 in a direction that promotes the rotational force of the impeller 24. Further, when the rotational speed of the turbine 27 is high, the flow of oil hits the back surface of the stator 28 and the stator 28 is idled by the one-way clutch 28a.
[0023]
On the other hand, when a certain vehicle speed or number of revolutions is reached, the lockup clutch 29 causes the impeller 24 and the turbine 27 to be directly connected via the front cover 25 to eliminate slippage of the so-called torque converter, thereby reducing the number of revolutions of the engine 10. This saves fuel consumption and improves quietness.
[0024]
The belt-type continuously variable transmission 30 includes a primary pulley 33 and a secondary pulley 34 provided on a primary shaft 31 and a secondary shaft 32 that are arranged in parallel to each other, and a drive belt 35 that is wound between the pulleys 33 and 34. And by changing the pulley groove width of each pulley 33, 34, the ratio of the effective winding diameter of the drive belt 35 to each pulley 33, 34 is changed to change the speed steplessly.
[0025]
Therefore, the primary pulley 33 provided on the primary shaft 31 formed integrally with the input shaft 21 moves in the axial direction with respect to the fixed sheave 33a formed integrally with the primary shaft 31 and the fixed sheave 33a. A movable sheave 33b is provided. The fixed sheave 33a and the movable sheave 33b transmit torque by holding the drive belt 35 with a predetermined clamping force in order to ensure a smooth continuously variable transmission of the transmission, and a pulley formed by the fixed sheave 33a and the movable sheave 33b. Since the groove width needs to be variably controlled, a plurality of ball grooves extending in the axial direction and facing each other are formed in the fitting portion between the primary shaft 31 and the movable sheave 33b. A means for transmitting torque via a ball 33c interposed between the two is employed.
[0026]
A substantially cylindrical piston 37a is fixed to the back surface of the movable sheave 33b opposite to the fixed sheave 33a. The piston 37a cooperates with a bottomed cylindrical cylinder 37b whose center is fixed to the primary shaft 31. Thus, a hydraulic actuator 37 including a spring 37c that forms the hydraulic chamber 37A and biases the movable sheave 33b in a direction of narrowing the pulley groove width is provided.
[0027]
An oil passage 31b communicating with the hydraulic chamber 37A is formed in the primary shaft 31, and the hydraulic actuator 37 is controlled via the oil passage 3a formed in the side cover 3 by being controlled by the hydraulic control circuit 9 based on the throttle opening degree and the like. The pulley groove width is variably controlled by moving the movable sheave 33b along the primary shaft 31 by the hydraulic pressure supplied to and discharged from the hydraulic chamber 37A.
[0028]
On the other hand, the secondary shaft 32 arranged in parallel with the primary shaft 31 is rotatably supported by the differential and converter housing 2 and the side cover 3 via roller bearings 32a and ball bearings 32b, and a secondary pulley 34 provided on the secondary shaft 32. Has a fixed sheave 34a formed integrally with the secondary shaft 32 and a movable sheave 34b that enables movement in the axial direction with respect to the fixed sheave 34a. The fixed sheave 34a and the movable sheave 34b are secondary shafts. 32 and the movable sheave 34b are configured to transmit torque via balls 34c that extend between the plurality of ball grooves that extend in the axial direction and are opposed to each other.
[0029]
A substantially cylindrical cylinder 36a is fixed to the back surface of the movable sheave 34b. The cylinder 36a cooperates with a cylindrical piston 36b whose center is fixed to the secondary shaft 32 to form a hydraulic chamber 36A. A hydraulic actuator 36 having a spring 36c that biases the movable sheave 34b in the direction of narrowing the pulley groove width is provided.
[0030]
An oil passage 32c communicating with the hydraulic chamber 36A is formed in the secondary shaft 32, and is controlled by the hydraulic control circuit 9 based on the throttle opening degree and the like, and the hydraulic actuator 36 via the oil passage 1a formed in the torque converter case 1 is formed. A drive gear 38 is provided at one end of the secondary shaft 32.
[0031]
Here, since the pressure receiving area for receiving the hydraulic pressure of the movable sheave 33b of the primary pulley 33 is larger than that of the movable sheave 34b of the secondary pulley 34, the primary pulley 33 and the secondary pulley 34 are in accordance with the hydraulic pressure supplied to and discharged from the hydraulic chambers 37A and 36A. The pulley groove width is changed to the reverse relationship, and the ratio of the effective winding diameter of the drive belt 35 to each pulley 33, 34 is converted into a stepless manner, and the continuously variable power is output to the secondary shaft 32.
[0032]
The speed change output from the secondary shaft 32 is output from the drive gear 38, decelerated by the counter shaft 39, and transmitted to the transfer unit 50 via the driven gear 54 and the transmission shaft 53 that is bolted to the driven gear 54.
[0033]
The counter shaft 39 includes a shaft 39a whose both ends are fixed to the torque converter case 1 and the differential and converter housing 2, and a relatively large-diameter drive-side gear that is rotatably fitted to the shaft 39a and meshes with the drive gear 38. The inner side of the needle bearing 39e and the roller bearing 39f is formed integrally with the drive gear 39c and the driven gear 39d meshed with the driven gear 54, and both sides are supported by the torque converter case 1 and the differential and converter housing 2. And a gear 39b that is restricted from moving in the axial direction via the race.
[0034]
Next, the front differential device 40 and the transfer unit 50 will be described with reference to FIG. 2 and FIG.
[0035]
The front differential device 40 includes a driven gear 54 in which a cylindrical flange portion 54a is rotatably supported on the differential and converter housing 2 through a ball bearing 54b, and a torque converter case 1 through a ball bearing 53a. It is disposed in a differential housing 41 having an enlarged diameter at a joint with a substantially cylindrical transmission shaft 53 that is pivotally supported.
[0036]
The structure of the front differential device 40 is integrally formed with a front drive shaft 51 serving as a first drive shaft, which will be described later, and is substantially cylindrical and is rotatably fitted in the flange portion 54a of the driven gear 54 and the transmission shaft 53. The differential case 42 has a hollow differential case 42, and a pair of pinions 43b are provided in the differential case 42 by pinion shafts 43a supported at both ends of the differential case 42. The differential gear 43 is obtained by engaging the left and right side gears 43c, 43d of the both pinions 43b. Is configured.
[0037]
The drive shaft 44 connected to one side gear 43c passes through the differential and converter housing 2 from the differential case 42 and transmits power to one front wheel via a constant velocity joint, an axle shaft, etc., and is connected to the other side gear 43d. 45 passes through the differential case 42 and the front drive shaft 51 integrally formed with the differential case 42, protrudes from the end cover 5, and transmits power to the other front wheel via a constant velocity joint, an axle shaft, and the like.
[0038]
The transfer unit 50 serves as a front drive shaft 51 and a second drive shaft serving as a first drive shaft disposed in parallel to the crankshaft 11, the input shaft 21, the primary shaft 31, the secondary shaft 32, and the like of the engine 10. A rear drive shaft 52 is provided.
[0039]
The crankshaft 11, the primary shaft 31, the secondary shaft 32, the front drive shaft 51, the rear drive shaft 52, etc. arranged in parallel with each other are arranged as shown in FIG. 4 showing the arrangement from the direction of arrow A in FIG. The rotating shaft core 11a of the shaft 11 and the primary shaft 31 are coaxially positioned in the vehicle body width direction, and the secondary shaft 32 is arranged parallel to the high position at the rear of the vehicle body with respect to the primary shaft 31, and the secondary pulley with respect to the primary pulley 33. 34 is arranged opposite to the high position behind the vehicle body. The front drive shaft 51 is arranged substantially below the secondary shaft 32, and the rear drive shaft 52 is arranged in parallel to the rear of the front drive shaft 51, thereby reducing the size of the entire drive device in the longitudinal direction. It is designed to improve compatibility with a vehicle body equipped with a manual transmission (MT) and an automatic transmission (AT) by improving the storage property in the room.
[0040]
One end of the front drive shaft 51 configured integrally with the differential case 42 is interposed in the torque converter case 1 via a transmission shaft 53 and a ball bearing 53a that supports the transmission shaft 53, and the other end portion is connected through a needle bearing 51c. Each end cover 5 is rotatably supported.
[0041]
Further, on the outer periphery of the other end of the front drive shaft 51, there are a clutch hub 86 of a third multi-plate clutch 84 serving as a third friction engagement element and a fourth multi-plate clutch 93 serving as a fourth friction engagement element. Splines 51b are formed in which the disks 83 that support the clutch hub 95 are fitted.
[0042]
The front drive shaft 51 is surrounded by a substantially cylindrical fixed shaft 62 integrally formed with the torque converter case 1, and a first friction engagement element is formed between the inner peripheral surface of the fixed shaft 62 and the front drive shaft 51. The first multi-plate clutch 68 is closed by the clutch drum 69 to form an oil chamber 62A. The fixed shaft 62 has an oil passage 62a communicating with the oil chamber 62A, and an oil passage 62b on the outer periphery of the fixed shaft 62. It is formed.
[0043]
A hub 61 is rotatably fitted to the front drive shaft 51. The hub 61 has a cylindrical portion 61a fitted to the front drive shaft 51, and a flange portion 61b formed at the proximal end of the cylindrical portion 61a. A double pinion planetary gear 55 is disposed on the outer periphery of the cylindrical portion 61a in the vicinity of the flange portion 61b. A spline 61c to which the sun gear 56 is fitted is formed at the tip, and a spline 61d to which the clutch drum 94 of the fourth multi-plate clutch 93 to be a fourth friction engagement element is fitted is formed, and the flange portion 61b has a first spline 61d. A clutch hub 79 of the second multi-plate clutch 78 that is the second frictional engagement element is formed.
[0044]
The hub 61 includes a thrust bearing 61f supported by a fixed shaft 62, a thrust bearing 61g supported via a clutch drum 69, clutch hubs 86 of the third multi-plate clutch 84 and the fourth multi-plate clutch 93, and By being held by a thrust bearing 61h supported by the extension case 5 via a disk 83 supporting 95, movement in the axial direction is prevented.
[0045]
The double pinion type planetary gear 55 fitted and coupled to the spline 61c formed on the outer periphery of the hub 61 is engaged with the sun gear 56, the ring gear 57, the sun gear 56, and the ring gear 57 that are spline-fitted to the spline 61c. The first and second pinions 58 and 59 that mesh with each other, and the carrier 60 that rotatably supports the first and second pinions 58 and 59 via a needle bearing 60 a, and the power that is input to the ring gear 57 Is transmitted to the sun gear 56 and the carrier 60 by torque distribution according to the gear specifications of the sun gear 56 and the ring gear 57, and the carrier 60 is moved to the sun gear 56 by the power input to the sun gear 56 by locking the ring gear 57 to the case 4. It has a function of rotating in the reverse direction.
[0046]
On the other hand, a transfer driven gear 52a is attached to one end of a rear drive shaft 52 arranged in parallel with the front drive shaft 51, and a bevel gear 52b meshing with a bevel gear 113a provided at one end of an output shaft 113 described later is attached to the other end. A ball bearing 52c rotatably supports the torque converter case 1 and the end cover 5 of the transmission case 7.
[0047]
Between the transmission shaft 53 and the double pinion type planetary gear 55, the first multi-plate clutch 68 and the second friction engagement element for selectively inputting the output from the transmission shaft 53 to the ring gear 57 or the sun gear 56 are provided. Input switching means 67 having a second multi-plate clutch 78 is provided.
[0048]
The first multi-plate clutch 68 will be described. A clutch drum 69 rotatably supported by the fixed shaft 62 is fitted into a spline 53a formed at the tip of the transmission shaft 53, and the clutch hub 70 is a double pinion planetary gear. It is coupled to 55 ring gears 57. Thus, the first multi-plate clutch 68 is interposed between the transmission shaft 53 and the ring gear 57 so as to be able to transmit power. Then, the hydraulic pressure of the hydraulic chamber 71 presses the retaining plate 73c that contacts the snap ring 73d fixed in the clutch drum 69 via the piston 72 and the drive plate 73a between the driven plate 73b and the clutch hub 70 to transmit power. Configured to do. A retainer 75 a is provided on the opposite side of the piston 72 from the hydraulic chamber 71 via a piston 74, and the pressing force of the return spring 76 is biased on the piston 72 via the piston 74.
[0049]
As for the second multi-plate clutch 78, the clutch drum 69 is shared with the first multi-plate clutch 68, and a clutch hub 79 is formed integrally with the hub 61. Thus, the second multi-plate clutch 78 is interposed between the transmission shaft 53 and the sun gear 56 so as to be able to transmit power. Then, the hydraulic pressure of the hydraulic chamber 80 presses the retaining plate 81c that contacts the snap ring 81d fixed to the piston 72 via the piston 74 and the drive plate 81a between the driven plate 81b and the clutch hub 79 so as to transmit power. Configured. Similar to the above, the piston 74 is biased by the pressing force of the return spring 76.
[0050]
On the side opposite to the input switching means 67 with respect to the double pinion planetary gear 55, a transfer drive gear 82 is rotatably supported by the case 4 of the transmission case 7 via a ball bearing 82a, and a hub via a needle bearing 82b. The transfer drive gear and the transfer driven gear 52a of the rear drive shaft 52 are meshed so as to be able to transmit power.
[0051]
The carrier 60 and the transfer drive gear 82 of the double pinion planetary gear 55 are spline-fitted so that power can be transmitted.
[0052]
In the third multi-plate clutch 84, the clutch drum 85 is coupled to the transfer drive gear 82 via the drum member 85a, and is supported on the end cover 5 so as to be rotatable coaxially with the front drive shaft 51. It is fitted to the spline 51 b of the front drive shaft 51 through 83. Thus, the third multi-plate clutch 84 is interposed between the transfer drive gear 82 and the front drive shaft 51 so as to be able to transmit power. The hydraulic pressure in the hydraulic chamber 87 presses the retaining plate 89c that contacts the snap ring 89d fixed in the clutch drum 85 via the piston 88 and the drive plate 89a between the driven plate 89b and the clutch hub 86 to transmit power. The piston 88 is biased by the pressing force of the return spring 92.
[0053]
Between the front drive shaft 51 and the end portion of the hub 61, a fourth multi-plate clutch 93 serving as a fourth friction engagement element for selectively transmitting power between the front drive shaft 51 and the hub 61 is disposed. The
[0054]
In the fourth multi-plate clutch 93, the clutch drum 94 is splined to the spline 61 d of the hub 61, and the clutch hub 95 is spline-fitted to the front drive shaft 51 via the disk 83 to connect the front drive shaft 51 and the hub 61. It is interposed so that power can be transmitted between them. Then, the hydraulic pressure of the hydraulic chamber 96 presses the retaining plate 98c abutting on the snap ring 98d fixed in the clutch drum 94 via the piston 97 and the drive plate 98a between the driven plate 98b and the clutch hub 95 to transmit power. The piston 97 is biased by the pressing force of the return spring 101.
[0055]
Between the case 4 of the transmission case 7 and the ring gear 57 of the double pinion type planetary gear 55, a fifth multi-fitting element for selectively locking to the case 4 and fixing the ring gear 57 is provided. A plate clutch 102 is provided.
[0056]
The fifth multi-plate clutch 102 is provided with a retaining plate 105 c and a driven rate 105 b that are in contact with a snap ring 105 d fixed in the case 4 via a piston 104 by the hydraulic pressure of the hydraulic chamber 103, and a clutch hub 70 provided on the ring gear 57. The ring gear 57 is locked and fixed to the case 4 by pressing the drive plate 105a therebetween, and the pressing force of the return spring 106 is biased to the piston 104.
[0057]
In the oil pan provided under the transmission case 7, the oil pressure from the oil pump 8 is supplied to the vehicle speed sensor 9a, throttle sensor 9b, shift switch 9c, front wheel speed sensor 9d, rear wheel speed sensor 9e, rudder angle sensor 9f. Are controlled by a hydraulic control circuit 9 based on signals from the first, second, third, fourth, and fifth multi-plate clutches 68, 78, 84, 93, 102. 87, 96, 103 and a control valve for selectively supplying to the continuously variable transmission 30 are provided.
[0058]
In the extension case 6 provided at the rear end of the torque converter case 1, an output shaft 113 is supported by a pair of roller bearings 112 that are supported by the extension case 6 by a retainer 110 and spaced apart from each other by a spacer 111. Yes.
[0059]
A bevel gear 113a that meshes with a bevel gear 52b provided on the rear drive gear 52 is provided at the tip of the output shaft 113, and the other end is configured to transmit power to the rear differential device 117 via a universal joint, a propeller shaft 116, and the like. The
[0060]
Next, the operation of the four-wheel drive vehicle configured as described above will be described with reference to the schematic explanatory views shown in FIGS. 5 to 9 and the first, second, third, fourth, and second driving ranges shown in FIG. The multi-plate clutches 68, 78, 84, 93 and 102 of FIG. In this friction engagement element operation explanatory diagram, ◯ indicates that the corresponding multi-plate clutch is engaged or operated, and (◯) indicates that it is engaged or operated as required later. ing.
[0061]
First, the power of the engine 10 is input from the crankshaft 11 to the primary shaft 31 of the continuously variable transmission 30 via the torque converter 20. Then, the primary shaft 31, the primary pulley 33, the drive belt 35, and the secondary pulley 34 are continuously stepped and output to the secondary shaft 32. The speed change output from the secondary shaft 32 is decelerated by the drive gear 38, the counter shaft 39, and the driven gear 54, and the first multi-plate clutch 68 and the second multi-plate clutch 78 through the transmission shaft 53 and the clutch drum 69. Is input. Here, in the neutral (N) range and the parking (P) range, the first and second multi-plate clutches 68 and 78 are released and the power transmission is cut off, and power transmission thereafter is not performed.
[0062]
In the drive (D) range that is the forward gear, the first multi-plate clutch 68 and the fourth multi-plate clutch 93 are engaged, and the power transmission state is shown by a bold line in FIG. That is, hydraulic pressure is supplied from the control valve to the hydraulic chamber 71, and the retaining plate 73c, the driven plate 73b, and the drive plate 73a that are in contact with the snap ring 73d fixed in the clutch drum 69 via the piston 72 are pressed and engaged. Power is transmitted from the driven gear 54 to the ring gear 57 of the double pinion planetary gear 55 via the transmission shaft 53 by the first multi-plate clutch 68, and the fourth multi-plate is connected to the hydraulic chamber 96 via the piston 97 by hydraulic pressure supplied to the hydraulic chamber 96. The sun gear 56 of the double pinion planetary gear 55 and the front drive shaft 51 are connected so that power can be transmitted by a fourth multi-plate clutch 93 that presses and engages the retaining plate 98c, driven plate 98b, and drive plate 98a of the clutch 93. You .
[0063]
Accordingly, in the double pinion type planetary gear 55, as shown in FIG. 6, the input side ring gear 57 meshes with the first pinion 58, and the second pinion 59 meshed with the first pinion 58 meshes with the sun gear 56 and the sun gear 56 and the carrier 60. Is rotated in the same direction as the ring gear 57, and is rotated differentially while transmitting torque to the sun gear 56 and the carrier 60 at a predetermined distribution ratio. The hub 61 is spline-coupled to the sun gear 56, a fourth multi-plate. The clutch 93, the front drive shaft 51 coupled to the front drive shaft 51 via a spline-fitting disk 83 and the like, and the transfer drive gear 82 spline-fitted to the carrier 60 are rotated in the same direction as the ring gear 57, thereby transferring the transfer drive gear. Transfer-driven meshing with 82 And outputs the Ya 52a to rotate the rear drive shaft 52 in a direction opposite to that of the ring gear 57. Then, it functions as a so-called center differential device that absorbs the rotational difference between the sun gear 56 and the carrier 60 by the rotation and revolution of the first and second pinions 58 and 59 during torque transmission.
[0064]
Here, torque distribution of the double pinion planetary gear 55 will be described with reference to the schematic diagram of FIG.
[0065]
The input torque to the ring gear 57 is Ti, the front side torque output to the front drive shaft 51 via the sun gear 56 and the hub 61 is TF, the rear side torque output to the rear drive shaft 52 by the carrier 60 is TR, and the sun gear. If the number of teeth of 56 is ZS and the number of teeth of the ring gear 57 is ZR,
Ti = TF + TR
TF: TR = ZS: (ZR-ZS)
Is established.
[0066]
From this, it is understood that the reference torque distribution of the front side torque TF and the rear side torque TR can be freely set by appropriately setting the number of teeth ZS of the sun gear 56 and the number of teeth ZR of the ring gear 57.
[0067]
If ZS = 37 and ZR = 82,
TF: TR = 37: (82-37)
become. Therefore, the front and rear wheel torque distribution ratio is
TF: TR≈45: 55
Thus, approximately 45% is allocated to the front wheels and approximately 55% is allocated to the rear wheels, and the reference torque distribution of the rear wheel deviation can be set sufficiently.
[0068]
On the other hand, the third multi-plate clutch 84 is configured to generate the clutch torque Tc by pressing the snap ring 89d, the retaining plate 89c, the driven plate 89b, and the drive plate 89a via the piston 88 by the hydraulic pressure of the hydraulic chamber 87, The clutch torque Tc is variably controlled by the hydraulic pressure from the control valve controlled by the hydraulic pressure control circuit 9.
[0069]
Here, the front wheel rotational speed NF and the rear wheel rotational speed NR detected by the front wheel rotational speed sensor 9d and the rear wheel rotational speed sensor 9e are input to the hydraulic control circuit 9, but after TF <TR when traveling on slippery roads. Since the rear wheel always slips first with the reference torque distribution of wheel deviation, the slip ratio is calculated as S = NF / NR (S> O). The slip ratio S and the steering angle ψ input to the hydraulic control circuit 9 from the steering angle sensor 9 f are used to retrieve the clutch pressure Pc from the map shown in FIG. 7 of the hydraulic control circuit 9. Here, when S ≧ 1 non-slip, the clutch pressure Pc is set to a low value, and in the slip state where S <1, the clutch pressure Pc increases as the slip ratio decreases, and the slip ratio S becomes the set value S.₁ P below_max Stipulated in The line pressure is adjusted to the clutch pressure Pc, and the clutch torque Tc of the third multi-plate latch 84 is variably controlled.
[0070]
Accordingly, the third multi-plate clutch 84 drives the front drive shaft 51 through the third multi-plate latch 84, the transfer drive gear 82, the carrier 60, the sun gear 56, the hub 61, and the fourth multi-plate clutch 93. A bypass system 115 extending to the shaft 51 is configured separately. In this bypass system 115, when the rear wheel slips, a differential function of the rear wheel rotational speed NR> the rotational speed of the ring gear 57> the front wheel rotational speed NF is established in the transfer unit 50, and the front drive shaft according to the clutch torque Tc. 51 is transmitted from the transfer drive gear 82 via the third multi-plate clutch 84 to the front drive shaft 51 with an increased torque by Tc, and further to the transfer driven gear 52a meshed with the transfer drive gear 82, the clutch torque that has flowed to the front wheels. Torque obtained by subtracting Tc is input and transmitted to the rear drive shaft 52. As a result, the front and rear wheel torques TF and TR are as follows.
[0071]
TF = 0.45Ti + Tc
TR = 0.55Ti-Tc
Accordingly, in the non-slip state, the clutch torque Tc is zero, so that the torque is distributed to the rear wheel with a bias of TF: TR = 45: 55. Is larger, the input torque Ti flows to the front wheel side via the bypass system 115, and TF: TR = TF as shown in FIG.₁ : TR₁ As a result, the front wheel torque is positively controlled to increase, and the rear wheel torque is reduced to prevent slipping and improve the running performance. When the slip S is equal to or less than the set value, the differential limiting torque is maximized together with the hydraulic pressure of the third multi-plate clutch 84, and the sun gear 56 and the carrier 60 are directly connected. For this reason, the transfer unit 50 is differentially locked, and becomes a direct-coupled four-wheel drive running with a torque distribution corresponding to the axial weight distribution of the front and rear wheels, so that the running performance is maximized.
[0072]
On the other hand, when the front wheel slips, the differential function of rear wheel rotational speed NR <rotational speed of ring gear 57 <front rotational speed NF is established in transfer unit 50, and transfer drive gear 82 from front drive shaft 51 according to clutch torque Tc. Torque is transmitted to the front wheel 51, and torque obtained by reducing the clutch torque Tc flowing to the rear wheel is transmitted from the front drive shaft 51 to the front wheel. As a result, the front and rear wheel torques TF and TR are as follows.
[0073]
TF = 0.45Ti-Tc
TR = 0.55Ti + Tc
Therefore, in the non-slip state, the clutch torque Tc is zero, so that the torque is distributed to the rear wheel with a bias of TF: TR = 45: 55. When the clutch torque Tc is generated when the front wheel slips, the input torque Ti is The rear wheel torque is actively controlled to increase as it flows to the rear wheel side, and the front wheel torque is reduced to prevent slipping and improve running performance. When the slip ratio is less than the set value, the differential limiting torque is maximized together with the hydraulic pressure of the third multi-plate clutch 84, and the sun gear 56 and the carrier 60 are directly connected. Therefore, torque distribution corresponding to the axial load distribution of the front and rear wheels. This is a direct-coupled four-wheel drive, and the running performance is fully demonstrated. In this way, the torque to the front and rear wheels is controlled widely to avoid the slip state.
[0074]
Further, when the vehicle turns in the torque distribution control accompanying the occurrence of the slip described above, the differential limiting torque of the third multi-plate clutch 84 is corrected to decrease by the steering angle ψ. For this reason, the differential limit of the transfer unit 50 is reduced, and the difference in the rotational speed can be sufficiently absorbed, the tight corner braking phenomenon is avoided, and good maneuverability is ensured.
[0075]
In the reverse (R) range that is the reverse speed, the first multi-plate clutch 68 and the fourth multi-plate clutch 93 are released, and the second multi-plate clutch 78, the third multi-plate clutch 84, and the fifth multi-plate clutch 93 are released. The plate clutch 102 is engaged, and the power transmission state shown in FIG. That is, the hydraulic pressure is supplied to the hydraulic chamber, and the snap ring 81d, the retaining plate 81c, the driven plate 81b, and the drive plate 81a are pressed via the piston 74 to engage the second multi-plate clutch 78 and from the transmission shaft 53. Power is transmitted to the sun gear 56 of the double pinion planetary gear 55 via the hub 61, and the snap ring 105d, the retaining plate 105c, the drive plate 105a, and the driven plate 105b are pressed via the piston 104 by the hydraulic pressure supplied to the hydraulic chamber 103. The ring gear 57 is locked and fixed to the case 4 by the fifth multi-plate clutch 102 engaged therewith. Then, the hydraulic pressure is supplied to the hydraulic chamber 87 and the snap ring 89d, the retaining plate 89c, the driven plate 89b, and the drive plate 89a are pressed via the piston 88, and the third multi-plate clutch 84 drives the front drive from the transfer drive gear 82. Power can be transmitted to the shaft 51.
[0076]
Accordingly, as shown in FIG. 9, the double pinion planetary gear 55 rotates along the ring gear 57 while the first and second pinions 58 and 59 meshed with each other by the rotation of the input-side sun gear 56 are rotated in the opposite directions. The carrier 60 is rotated in the opposite direction to the sun gear 56 to cause the transfer drive gear 82 to rotate in the opposite direction with respect to the input side, and the transfer drive gear 82 is connected to the front drive shaft 51 via the third multi-plate clutch 84. Power is transmitted and the rear drive shaft 52 is rotationally driven in the direction opposite to the front drive shaft 51.
[0077]
Therefore, the input from the driven gear 62 is such that the ring gear 57 of the double pinion type planetary gear 55 is locked to the case 4 by the fifth multi-plate clutch 102 to reverse the front drive shaft 51 and the rear in the direction opposite to the drive (D) range state. The double pinion planetary gear 55 is output to the drive shaft 52 and has a forward / reverse switching function.
[0078]
In this case, the gear ratio output to the front drive shaft 51 and the rear drive 52 with respect to the input of the sun gear 56 is set by the following equation.
[0079]
Gear ratio = [ZS + (− ZR)] / ZS
If ZS = 37 and ZR = 82 as in the above,
Gear ratio = [37 + (− 82)] / 37 = −1.216
Thus, the reduction ratio in the reverse (R) range is appropriately secured.
[0080]
On the other hand, the torque Ti input to the sun gear 56 is transmitted to the front drive shaft 51 according to the clutch Tc, and the torque obtained by subtracting the clutch torque Tc transmitted to the front wheels is input to the rear wheels. As a result, the front and rear wheel torques TF, TR is as follows.
[0081]
Ti = TF + TR
TR = Ti-Tc
TF = Tc
Therefore, by increasing the clutch torque Tc when the rear wheel slip occurs, the input torque Ti is caused to flow to the front wheel side, the front wheel torque is actively increased and controlled, the rear wheel torque is reduced to prevent slipping, and the running performance is improved. In addition, when the front wheel slips, the input torque Ti is caused to flow to the rear wheel side by reducing the clutch torque Tc, and the rear wheel torque is positively increased to reduce the front wheel torque so that slip does not occur and the running performance is improved. When the slip ratio is less than the set value, the differential limiting torque Tc is maximized together with the hydraulic pressure of the third multi-plate clutch 84, and the front drive shaft 51 and the transfer drive gear 82 are directly connected, which corresponds to the axial load distribution of the front and rear wheels. The direct running four-wheel drive with the torque distribution is maximized in running performance. In the case of further turning, the differential limiting torque of the third multi-plate clutch 84 is reduced by the rudder angle ψ, and it becomes possible to sufficiently absorb the rotational speed difference, and the tight corner braking phenomenon is avoided. , The maneuverability will be good.
[0082]
Therefore, in the present embodiment described above, the front drive shaft 51 and the rear drive shaft 52 that transmit power to the front differential device 40 or the rear differential device 117 that are configured to be transmitted to the output side of the belt-type continuously variable transmission 30 are laterally arranged. The continuously variable transmission 30 is provided with a double-pinion planetary gear 55 that is arranged in parallel to the crankshaft 11 of the stationary engine 10 and that has a sun gear 56 coupled to the front drive shaft 51 via a hub 61 and a fourth multi-plate clutch 93. A first multi-plate clutch 68 that transmits the output from the ring gear 57, a second multi-plate clutch 78 that transmits the output to the hub 61, and a third multi-link that connects the front drive shaft 51 and the transfer drive gear 82 so as to transmit power. The fifth multi-plate clutch 1 for locking the multi-plate clutch 84 and the ring gear 57 2 and a drive (D) range which is a forward stage by selectively controlling each of the first, second, third, fourth and fifth multi-plate clutches 68, 78, 84, 93 and 102. In the reverse (R) range, which is the reverse stage, it functions as a center differential device that enables appropriate torque distribution and differential limitation to the front drive shaft 51 and the rear drive shaft 52, and good driving performance is obtained. (D) Functions as a forward / reverse switching device when switching to the range or reverse (R) range.
[0083]
Therefore, a dedicated double-pinion planetary gear that functions independently for each of the center differential device and the forward / reverse switching device has been conventionally required. However, both functions are achieved by a single double-pinion planetary gear, and it is driven while maintaining high performance. Device configuration and control can be simplified and reduced in weight, reducing cost and downsizing, especially the overall length in the vehicle body width direction. It is possible to obtain a degree of freedom in designing the vehicle body while ensuring a crash stroke at the time of a side collision and a work space for assembling and maintenance.
[0084]
Further, instead of the torque converter 20, an electromagnetic clutch or a wet clutch can be used as a starting clutch. In this case, the belt-type continuously variable transmission 30 is connected to the primary shaft 31 in the neutral (N) range and the parking (P) range. The power transmission after the continuously variable transmission 30 is cut off after the input is cut off.
[0085]
Further, in the four-wheel drive apparatus according to the present embodiment, as shown in FIG. 11, the first, third and fourth multi-plate clutches 68, 84 and 93, the transfer drive gear 82, and the rear drive The rear wheel drive unit such as the shaft 52 is eliminated, and the hub 61 is changed to the clutch hub 121 coupled to the carrier 60 so that the output from the second multi-plate clutch can be input to the carrier 60 and the clutch drum 69 is formed. The connecting member 120 directly inputs the output from the transmission shaft 53 to the sun gear 56 of the double pinion planetary gear 55, connects the carrier 60 of the double pinion planetary gear 55 and the front drive shaft 51, and eliminates the case 4. The end cover 5 is changed to a two-wheel drive end cover 122, and the fifth multi-plate clutch is replaced with a double pinion type planetary plate. It can be readily modified for two-wheel-drive vehicle drive unit by arranging between the Rigiya 55 and the end cover 122.
[0086]
In the change to the drive device for a two-wheel drive vehicle, detailed description will be omitted by attaching the same reference numerals to the portions corresponding to FIG. 3 in FIG. 11, but there are many other than the connecting member 120 and the end cover 122 described above. The parts can be shared with the drive device for a four-wheel drive vehicle.
[0087]
The two-wheel drive vehicle drive device formed in this way has the second and fifth multi-plate clutches in the neutral (N) range and the parking (P) range as shown in FIG. The power transmission to the front drive shaft 51 is cut off by setting the power transmissions 78 and 102 to a cut-off state, and the second multi-plate clutch 78 is engaged in the drive (D) range as the forward gear, and the power transmission is shown in FIG. The state is indicated by a bold line. That is, hydraulic pressure is supplied to the hydraulic chamber 80, and the snap ring 81d, the retaining clutch 81c, the driven plate 81b, and the drive plate 81a are pressed via the piston 74 to engage the second multi-plate clutch 78, thereby driving the driven gear. Input from 62 is transmitted to the front drive shaft 51 through the carrier 60 of the double pinion planetary gear 55, and rotationally drives the front drive shaft 51 in the same direction as the driven gear 62.
[0088]
On the other hand, in the reverse (R) range, which is the reverse stage, the engagement of the second multi-plate clutch 78 is released, the hydraulic pressure is supplied to the hydraulic chamber 103, and the ring gear 57 of the double pinion planetary gear 55 is supplied by the fifth multi-plate clutch 102. Is engaged with the transmission case 7 so that the power transmission state is indicated by a bold line in FIG. As a result, the first and second pinions 58 and 59 engaged with each other by the rotation of the sun gear 56 due to the input from the driven gear 62 to the sun gear 56 of the double pinion planetary gear 55 rotate along the ring gear 57 while rotating in reverse. Thus, the carrier 60 is rotated in the opposite direction to the sun gear to drive the front drive shaft 51 in the opposite direction with respect to the input side.
[0089]
In this case, the gear ratio output to the front drive shaft 51 with respect to the input of the sun gear 56 is set by the following equation as described above.
[0090]
Gear ratio = [ZS + (− ZR)] / ZS
Here, ZS = 37 and ZR = 82.
Gear ratio = [37 + (− 82)] / 37 = −1.216
The reduction ratio in the reverse (R) range is appropriately secured.
[0091]
Therefore, a forward / reverse switching device having the double pinion planetary gear 55, the second multi-plate clutch 78, and the fifth multi-plate clutch 102 as main parts is configured.
[0092]
【The invention's effect】
According to the drive device for a four-wheel drive vehicle of the present invention described above, the first and second drive shafts that transmit power to one and the other differential devices are arranged in parallel with respect to the crankshaft of the horizontal engine, Input from the engine side is selectively input to the ring gear and sun gear of the double pinion planetary gear, and this input is selectively distributed to the first and second drive shafts by a single double pinion planetary gear. Power transmission to a single double pinion planetary gear achieves both functions as a center differential device and a forward / reverse switching device, while maintaining high performance, simplifying the configuration and control of the drive, reducing weight, and compactness With the downsizing, the drive unit and the engine compartment side wall The effects of peculiar to the present invention, such as being able to be mounted in a conventional engine room, are obtained, while being able to be mounted in a conventional engine room, while securing a working space for side stroke, assembly, maintenance, etc. Have.
[Brief description of the drawings]
FIG. 1 is a diagram showing a drive system for explaining an outline of an embodiment of a drive device for a four-wheel drive vehicle according to the present invention;
FIG. 2 is a cross-sectional view illustrating a four-wheel drive vehicle drive device.
3 is an enlarged cross-sectional view of the main part of FIG.
4 is an explanatory view showing the arrangement of main parts as seen from the direction of arrow A in FIG. 2; FIG.
FIG. 5 is a schematic explanatory view showing the operation in the same manner.
FIG. 6 is a schematic explanatory view showing the operation in the same manner.
FIG. 7 is a schematic explanatory view showing the operation in the same manner.
FIG. 8 is a schematic explanatory view showing the operation in the same manner.
FIG. 9 is a schematic explanatory view showing the operation in the same manner.
FIG. 10 is an explanatory diagram of the operation of a friction engagement element showing the action in the same manner.
FIG. 11 is a view for explaining diversion of the drive device for a four-wheel drive vehicle of the present invention to a drive device for a two-wheel drive vehicle.
12 is a schematic explanatory view showing the operation of the two-wheel drive vehicle driving device shown in FIG. 11. FIG.
FIG. 13 is a schematic explanatory view showing the operation of the two-wheel drive vehicle drive device.
FIG. 14 is an explanatory diagram of the operation of a friction engagement element, similarly showing the action.
[Explanation of symbols]
10 engine
11 Crankshaft
20 Torque converter
30 belt type continuously variable transmission
31 Primary axis
32 Secondary shaft
33 Primary pulley
34 Secondary pulley
35 Drive belt
40 Front differential device
51 Front drive shaft
52 Rear drive shaft
53 Transmission shaft
55 Double pinion type planetary gear
56 Sungear
57 Ring gear
58 First Pinion
59 Second Pinion
60 Carrier
61 Hub
67 Input switching means
68 First multi-plate clutch
78 Second multi-plate clutch
84 Third multi-plate clutch
93 Fourth multi-plate clutch
102 fifth multi-plate clutch
117 Rear differential device

Claims (7)

A horizontal engine,
A transmission to which the output from the engine is input;
First and second drive shafts that are arranged in parallel to the crankshaft of the engine and respectively transmit power to one differential device and the other differential device;
A double pinion planetary gear arranged coaxially with the first drive shaft;
Input switching means for selectively transmitting the output from the transmission to the ring gear and sun gear of the planetary gear;
Means for transmitting power from the planetary gear carrier to the second drive shaft;
A third friction engagement element interposed between the first drive shaft and the second drive shaft so as to be capable of transmitting power ;
A fourth friction engagement element for selectively transmitting the output from the sun gear of the planetary gear to the first drive shaft;
A fifth friction engagement element for selectively locking the ring gear rotation of the planetary gear,
The input switching means and each friction engagement element are selectively actuated to switch the input from the transmission to a first and second drive shafts by switching power distribution and forward / reverse at a predetermined ratio via the planetary gear. A drive device for a four-wheel drive vehicle characterized by transmitting power. The forward speed is a state in which the input switching means is in a state of transmitting power from the transmission to the ring gear, and is in a ring gear rotation permissible state in which the fifth friction engagement element is released. The fourth friction engagement element is in a power transmission state and the third friction engagement element is in a power transmission state, and the differential between the carrier and the sun gear The drive device for a four-wheel drive vehicle according to claim 1 , which performs restriction. The drive device for a four-wheel drive vehicle according to claim 2 , wherein in the forward gear, the third friction engagement element variably controls the transmission torque based on the traveling state to transmit power. In the reverse gear, the input switching means is in a state of transmitting power from the transmission to the sun gear, the fifth friction engagement element is fastened and the ring gear rotation is locked, and the double-pinion planetary gear receives the shift power. The drive device for a four-wheel drive vehicle according to any one of claims 1 to 3, wherein the third friction engagement element is output to the carrier, the fourth friction engagement element is in a power transmission state, and the fourth friction engagement element is in a release state. The drive device for a four-wheel drive vehicle according to claim 4 , wherein in the reverse gear, the third friction engagement element variably controls the transmission torque based on the running state to transmit power. A first friction engagement element that engages at the forward speed and transmits the output from the transmission to the ring gear and an input switching means that engages at the reverse speed and a power that transmits the output from the transmission to the sun gear. The drive device for a four-wheel drive vehicle according to any one of claims 1 to 5 , having two friction engagement elements. And transmission primary shaft, a secondary shaft that is arranged parallel with the primary shaft, a primary pulley and a secondary pulley, each provided on the primary shaft and the secondary shaft, wound between the primary pulley and the secondary pulley and a drive belt, according to any one of claims 1 to 6 is a belt-type continuously variable transmission which steplessly by changing the ratio of winding diameter to the primary pulley and the secondary pulley of the drive belt The drive device for four-wheel drive vehicles.

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