JP2006118688A - Belt type continuously variable transmission - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To miniaturize a belt type continuously variable transmission by transmitting driving force to a movable sheave even without a gear group. <P>SOLUTION: A power transmitting means 551 comprising a first power transmission member 551a capable of relatively turning around a pulley shaft 51, and a second power transmission member 551b integrally turning with the pulley shaft 51, and capable of moving the movable sheave 53 near to a fixed sheave 52 or away from it, a clutch mechanism 58 providing relative movement between the first power transmission member 551a and the second power transmission member 551b, and capable of stopping the relative movement, and an oil pressure chamber 57 capable of pressing the movable sheave 53 toward the fixed sheave 52 are provided in an internal space formed in the movable sheave 53. The clutch mechanism 58 is arranged in a space formed by a hydraulic piston member 57a composing a wall surface of the oil pressure chamber 57. <P>COPYRIGHT: (C)2006,JPO&NCIPI

Description

  The present invention relates to a belt-type continuously variable transmission, and in particular, a belt including a movable sheave sliding mechanism that slides a movable sheave in an axial direction and a clutch mechanism that can stop relative rotation of the movable sheave with respect to a fixed sheave. The present invention relates to an improvement of a continuously variable transmission.

  In general, a belt-type continuously variable transmission includes two rotating shafts arranged in parallel, a primary pulley and a secondary pulley separately attached to each of the rotating shafts, and a V-shape of each of the primary pulley and the secondary pulley. And a belt wound around a groove having a shape. Here, each of the primary pulley and the secondary pulley is a vertical fixed sheave fixed to the rotary shaft (primary shaft and secondary shaft), and a vertical movable sheave that slides in the axial direction on the rotary shaft. The V-shaped groove is formed by the inclined portion of the fixed sheave and the inclined portion of the movable sheave facing each other.

  In this type of belt-type continuously variable transmission, the movable sheave is slid in the axial direction of the rotating shaft to change the V-shaped groove width, whereby each of the belt, the primary pulley, and the secondary pulley. The contact radius can be changed steplessly, whereby the gear ratio can be changed steplessly. In other words, since the ratio of the contact radius on the primary pulley side and the contact radius on the secondary pulley side becomes the gear ratio of the belt-type continuously variable transmission, this belt-type continuously variable transmission has a groove width of the primary pulley. By controlling, the gear ratio can be varied steplessly.

  Thus, conventionally, in order to change the gear ratio in a belt-type continuously variable transmission, it is necessary to slide the movable sheave in the direction of the rotation axis. Therefore, in this belt-type continuously variable transmission, the primary pulley is movable. A mechanism for sliding the sheave (movable sheave sliding mechanism) is provided. As an actuator that functions as the movable sheave sliding mechanism, for example, there is an actuator that uses a driving force of a motor such as an electric motor or a hydraulic motor.

  For example, in Patent Documents 1 and 2 below, there is a movable sheave sliding mechanism that transmits the driving force of an electric motor to a ball screw mechanism through a plurality of gear groups, and the movable sheave is slid in the axial direction by the ball screw mechanism. Further, a clutch mechanism that switches transmission of driving force of the electric motor between the electric motor and the movable sheave is also disclosed.

JP 2000-291895 A JP 2000-283253 A

  However, in the movable sheave sliding mechanism of Patent Documents 1 and 2, the electric motor is disposed at a position separated from the movable sheave, and a plurality of gear groups, clutch mechanisms, and ball screws are provided between the electric motor and the movable sheave. Since the mechanism is interposed, there is an inconvenience that these arrangement places must be secured and the transmission becomes large.

  Accordingly, an object of the present invention is to provide a belt type continuously variable transmission that can improve the disadvantages of the conventional example and can be downsized.

  In order to achieve the above object, according to the first aspect of the present invention, in the internal space formed in the movable sheave, a first power transmission member and a pulley shaft that can rotate relative to the pulley shaft with the pulley shaft as a central axis. A power transmission means comprising a second power transmission member rotating integrally and capable of moving the movable sheave closer to or away from the fixed sheave, and relative movement between the first power transmission member and the second power transmission member is possible. On the other hand, a clutch mechanism capable of stopping relative movement therebetween and a hydraulic chamber capable of pressing the movable sheave toward the fixed sheave are provided, and the clutch is provided in a space formed by a hydraulic piston member that constitutes a wall surface of the hydraulic chamber. The mechanism is arranged.

  Thus, in the invention described in claim 1, the power transmission means and the hydraulic chamber for sliding the movable sheave, and the relative movement between the movable sheave and the fixed sheave can be made possible. Since the clutch mechanisms that can stop the movement can be arranged in the inner space of the movable sheave, the transmission itself can be downsized.

  In order to achieve the above object, according to a second aspect of the present invention, in the belt-type continuously variable transmission according to the first aspect, an outer thread portion is provided on the outer peripheral surface of the first power transmission member, while the second An inner screw portion that is screwed with the outer screw portion is provided on the inner peripheral surface of the power transmission member, and a sliding surface between the movable sheave and the hydraulic piston member is provided on the inner diameter side of the outer screw portion of the first power transmission member. Yes.

  According to the second aspect of the present invention, since the power transmission means and the hydraulic chamber can be arranged inwardly of the maximum outer diameter of the movable sheave, the transmission itself can be further reduced in size. Further, since the hydraulic oil leaked from the hydraulic chamber to the sliding surface between the movable sheave and the hydraulic piston member due to the centrifugal force accompanying the rotation of the movable sheave is supplied to the outer screw portion and the inner screw portion, The threaded portion and the inner threaded portion can be lubricated.

  In order to achieve the above object, according to a third aspect of the present invention, in the belt type continuously variable transmission according to the first or second aspect, the pulley shaft rotates relative to the pulley shaft with the pulley shaft as a central axis. A hydraulic motor as a drive source in the axial direction of the movable sheave provided with a motor case and a motor shaft that rotates integrally with the pulley shaft is provided, and a first power transmission member is provided integrally on the outer peripheral surface of the motor case Alternatively, the first power transmission member is configured by the motor case. The motor shaft and the hydraulic piston member are spline-fitted at the inner diameter portion of the motor shaft, and a lubricating oil supply path is provided for supplying the lubricating oil from the gap of the spline groove in the spline fitting portion to the clutch mechanism. ing.

  According to the third aspect of the present invention, it is possible to reduce the size of the transmission itself in the axial direction by spline fitting the motor shaft and the hydraulic piston member at the inner diameter portion of the motor shaft. The lubricating oil supply path through the spline fitting portion enables the lubricating oil to be supplied to the clutch mechanism as well as downsizing.

  According to the belt type continuously variable transmission according to the present invention, the driving force can be transmitted to the movable sheave without the conventional gear group. Further, as described above, various components can be transferred to the inside of the movable sheave. Since it can arrange | position collectively in space, size reduction of transmission itself can be achieved.

  Embodiments of a belt type continuously variable transmission according to the present invention will be described below in detail with reference to the drawings. The present invention is not limited to the embodiments.

  A belt type continuously variable transmission according to a first embodiment of the present invention will be described with reference to FIGS.

  First, an overall configuration of a power transmission device including a belt type continuously variable transmission according to the present invention will be described with reference to FIG.

  The power transmission device includes an internal combustion engine 10 and a transaxle 20 disposed on the output side of the internal combustion engine 10.

  As shown in FIG. 1, the transaxle 20 includes, in order from the output side of the internal combustion engine 10, a transaxle housing 21 attached to the internal combustion engine 10, a transaxle case 22 attached to the transaxle housing 21, A transaxle rear cover 23 attached to the transaxle case 22 is provided, and a housing is constituted by these.

  First, a torque converter (starting device) 30 is accommodated in the transaxle housing 21. The torque converter 30 increases the torque of the internal combustion engine 10 and transmits the torque to a belt-type continuously variable transmission 1 described later. A pump impeller 31, a turbine liner 32, a stator 33, a lock-up clutch 34, and a damper device 35 are provided. Etc.

  An input shaft 38 that is rotatable about the same axis as the crankshaft 11 of the internal combustion engine 10 is provided inside the transaxle housing 21. Here, the turbine liner 32 is attached to the end of the input shaft 38 on the internal combustion engine 10 side, and the lock-up clutch 34 is provided via the damper device 35.

  On the other hand, a front cover 37 of the torque converter 30 is connected to the end of the crankshaft 11 on the transaxle 20 side via a drive plate 12, and the pump impeller 31 is connected to the front cover 37. .

  The pump impeller 31 is disposed opposite to the turbine liner 32, and the stator 33 is disposed inside the pump impeller 31. A hollow shaft 36 is connected to the stator 33 via a one-way clutch 39, and the input shaft 38 is disposed inside the hollow shaft 36.

  Here, hydraulic oil is supplied into a casing (not shown) formed by the front cover 37 and the pump impeller 31 as described above.

  The operation of the torque converter 30 will be described below.

  First, the torque of the internal combustion engine 10 is transmitted from the crankshaft 11 to the front cover 37 via the drive plate 12. Here, when the lockup clutch 34 is released by the damper device 35, the torque transmitted to the front cover 37 is transmitted to the pump impeller 31 and circulates between the pump impeller 31 and the turbine liner 32. Torque is transmitted to the turbine liner 32 via the hydraulic oil. The torque transmitted to the turbine liner 32 is transmitted to the input shaft 38.

  Here, an oil pump (hydraulic pump) 26 shown in FIG. 1 is provided between the torque converter 30 and a forward / reverse switching mechanism 40 described later. The oil pump 26 is connected to the pump impeller 31 by a rotor 27 via a cylindrical hub 28, and a body (housing) 29 is fixed to the transaxle case 22 side. Further, the hub 28 is spline-fitted to the hollow shaft 36. With the configuration as described above, the power of the internal combustion engine 10 is transmitted to the rotor 27 via the pump impeller 31, so that the oil pump 26 can be driven.

  Next, inside the transaxle case 22 and the transaxle rear cover 23, a forward / reverse switching mechanism 40, the belt-type continuously variable transmission 1, and a final speed reducer 70 as a differential device are housed.

  First, the forward / reverse switching mechanism 40 transmits the torque of the internal combustion engine 10 transmitted to the input shaft 38 in the torque converter 30 to the primary pulley 50 of the belt-type continuously variable transmission 1 described later. The mechanism 41, the forward clutch 42, and the reverse brake 43 are comprised.

  The planetary gear mechanism 41 includes a sun gear 44, a pinion (planetary pinion) 45, and a ring gear 46.

  Here, the sun gear 44 is spline-fitted to a connecting member (not shown), and the connecting member is spline-fitted to the primary shaft 51 that is the rotation shaft of the primary pulley 50. With this configuration, the torque transmitted to the sun gear 44 is transmitted to the primary shaft 51.

  A plurality of (for example, three) pinions 45 are arranged around the sun gear 44 and meshed with the sun gear 44. Here, each pinion 45 is supported by a carrier 48 that supports the pinion 45 itself so as to be capable of rotating, and supports the pinion 45 so as to be integrally revolved around the sun gear 44. The carrier 48 is connected to the reverse brake 43 at its outer peripheral end.

  The ring gear 46 is engaged with each pinion 45 held by the carrier 48 and is connected to the input shaft 38 in the torque converter 30 via the forward clutch 42.

  Subsequently, the forward clutch 42 is ON / OFF controlled by the hydraulic oil supplied to the hollow portion of the input shaft 38. Here, a brake piston (not shown) is used for the ON / OFF control. During forward travel, the forward clutch 42 is turned on and the reverse brake 43 is turned off. During reverse travel, the forward clutch 42 is turned off and the reverse brake 43 is turned on.

  Next, a schematic configuration of the belt type continuously variable transmission 1 will be described.

  The belt type continuously variable transmission 1 includes a primary shaft (pulley shaft) 51 disposed concentrically with the input shaft 38, and a secondary shaft disposed in parallel with the primary shaft 51 at a predetermined interval. (Pulley shaft) 61. Here, the primary shaft 51 is rotatably supported by bearings 81 and 82 shown in FIG. 1, and the secondary shaft 61 is rotatably supported by bearings 83 and 84 shown in FIG.

  First, the configuration on the primary shaft 51 side will be described.

  The primary shaft 51 is provided with a primary pulley 50 shown in FIG. As shown in FIG. 2, the primary pulley 50 includes a fixed sheave 52 that is integrally disposed on the outer periphery of the primary shaft 51, and a movable sheave 53 that is slidable in the axial direction of the primary shaft 51. .

  Here, a V-shaped groove 80 a is formed between the opposing surfaces of the fixed sheave 52 and the movable sheave 53. Further, a cylindrical first extending portion 53a extends on the inner diameter side of the back surface of the movable sheave 53 (opposite side of the groove 80a), and on the inner peripheral surface of the first extending portion 53a. The movable sheave 53 is spline-fitted to the outer peripheral surface of the primary shaft 51 via the spline 54A. For this reason, the movable sheave 53 can rotate integrally with the primary shaft 51 and the fixed sheave 52 and can slide on the primary shaft 51 in the axial direction.

  Further, the primary shaft 51 is provided with a movable sheave sliding mechanism 55 that slides the movable sheave 53 in the axial direction of the primary shaft 51 to approach or separate from the fixed sheave 52. Hereinafter, the movable sheave sliding mechanism 55 according to the first embodiment will be described in detail with reference to FIGS.

  As shown in FIG. 2, the movable sheave sliding mechanism 55 includes a hydraulic motor 550 that is a driving source for sliding the movable sheave 53 in the axial direction of the primary shaft 51, and a driving force (rotational direction) of the hydraulic motor 550. Force transmission means 551 for transmitting the force) to the movable sheave 53 and sliding the movable sheave 53 in the sliding direction.

  First, as the hydraulic motor 550 of the first embodiment, a motor having a structure in which the rotation of the outer rotor generated by the relative rotation with the inner rotor is used as a driving force is used. For example, at least two oil chambers are formed by at least two vanes (blades) arranged in a motor case constituting the outer rotor, and each vane is relatively rotated by the hydraulic pressure of the hydraulic oil flowing into the oil chamber. A so-called vane type hydraulic motor that generates a driving force is used.

  In the vane type hydraulic motor 550 of the first embodiment, as shown in FIGS. 2 and 3, a cylindrical shape that is fitted or press-fitted to the outer peripheral surface of the primary shaft 51 and rotates integrally with the primary shaft 51. Motor shaft 550a and a motor case 550b fixed to the motor shaft 550a so as to be rotatable relative to the motor shaft 550a. Thus, the motor case 550b rotates relative to the primary shaft 51 and the motor shaft 550a. It is comprised so that relative rotation centering on can be performed.

  Here, the motor shaft 550a of the first embodiment is spline-fitted to the outer peripheral surface of the primary shaft 51 via the spline 54B on the inner peripheral surface on one end side thereof.

Further, the motor case 550b according to the first embodiment includes a cylindrical portion 550b ₁ arranged with the motor shaft 550a as a central axis, and a screw member B shown in FIG. 2 so as to close the opening of the cylindrical portion 550b ₁ . It is constituted by an annular portion 550b ₂ to be fixed. The motor case 550b is fixed to the motor shaft 550a through a bearing 550c at its annular portion 550b _2.

  Since first and second oil chambers 550g and 550h, which will be described later, are formed in the motor case 550b, the hydraulic oil in the first and second oil chambers 550g and 550h is external to the motor case 550b. An annular seal member 550d shown in FIG. 2 is provided between the motor case 550b and the motor shaft 550a so as not to leak.

Specifically, in the vane-type hydraulic motor 550 of the first embodiment, as shown in FIG. 3, the motor case 550b of the cylindrical portion 550b ₁ of the inner peripheral surface to the two first vane 550e, 550e of the motor shaft 550a periphery The motor case 550b and the first vanes 550e and 550e constitute an outer rotor. On the other hand, the motor shaft 550a of the outer peripheral surface two second vanes 550f to, 550f is integrally erected toward the inner circumferential surface of the cylindrical portion 550b _1, the motor shaft 550a and a second vane 550f , 550f constitute an inner rotor.

  That is, in the vane type hydraulic motor 550 of the first embodiment, the first and second vanes 550e and 550f are arranged in an annular space between the motor shaft 550a and the motor case 550b as shown in FIG. First and second oil chambers 550g and 550h are formed.

Here, in the first embodiment, seal members 550e ₁ and 550e ₁ are provided between the first vanes 550e and 550e and the outer peripheral surface of the motor shaft 550a and the inner wall surface of the motor case 550b, respectively. By providing seal members 550f ₁ and 550f ₁ between the second vanes 550f and 550f and the inner wall surface of the motor case 550b, the first and second oil chambers 550g and 550h together with the annular seal member 550d described above. Airtightness is ensured.

  The vane type hydraulic motor 550 of the first embodiment configured as described above is disposed on the movable sheave 53 on the opposite side of the groove 80a. Specifically, a cylindrical second extending portion 53b is extended from the outer diameter side of the movable sheave 53 toward the opposite side of the groove 80a, and the second extending portion 53b and the first extending portion are extended. A vane hydraulic motor 550 is disposed in a substantially annular inner space formed by the portion 53a and the primary shaft 51.

  On the other hand, the vane hydraulic motor 550 arranged at such a position is attached to the inner peripheral surface of the second extending portion 53b via the power transmission means 551. Next, the power transmission means 551 will be described in detail.

For example, as the power transmission unit 551 of the first embodiment, a so-called motion screw such as a multi-thread screw or a slide screw that converts the rotational force of the outer rotor into a force in the axial direction is used. As shown in FIG. 2, this type of power transmission means 551 includes a cylindrical first power transmission member 551 a integrally provided on the outer peripheral surface of the cylindrical portion 550 b ₁ of the motor case 550 b, and a second of the movable sheave 53. It is comprised by the cylindrical 2nd power transmission member 551b integrally provided in the internal peripheral surface of the extending part 53b.

  In the power transmission means 551, an outer thread portion in the circumferential direction is formed on the outer peripheral surface of the first power transmission member 551a, while the outer thread portion is also screwed onto the inner peripheral surface of the second power transmission member 551b. An inner thread portion in the circumferential direction is formed. Incidentally, the motor case 550b of the vane hydraulic motor 550 rotates in the circumferential direction together with the primary shaft 51 or relative to the primary shaft 51, but does not move in the axial direction thereof. For this reason, the power transmission means 551 can slide the movable sheave 53 in the axial direction by rotating the motor case 550b in the circumferential direction relative to the primary shaft 51.

  Here, the second power transmission member 551b of the first embodiment is spline-fitted to the inner peripheral surface of the second extending portion 53b via the spline 551c, and a holding member (here, the spline fitting) The movement in the axial direction with respect to the second extending portion 53b is restricted by the annular snap ring 551d). As described above, by adopting a structure in which the second power transmission member 551b is spline-fitted to the second extending portion 53b, the assembly is better than the second power transmission member integrated with the movable sheave 53. As a result, a cost reduction effect can be achieved.

  In the first embodiment, the first power transmission member 551a is provided as a separate member in the motor case 550b. However, the first power transmission can be achieved by forming an external screw portion on the outer peripheral surface of the motor case 550b. The member 551a may be integrated with the motor case 550b. In the first embodiment, the second power transmission member 551b is also provided as a separate member on the movable sheave 53, but an internal thread portion is formed on the inner peripheral surface of the second extending portion 53b of the movable sheave 53. By doing so, the second power transmission member 551b may be integrated with the movable sheave 53.

  As described above, in the first embodiment, since the power transmission means 551 such as a motion screw is provided, the reaction force of the large movable sheave thrust can be borne by a relatively small torque by the friction of the screw surface. it can. For this reason, the output (torque) of the vane hydraulic motor 550 can be lowered. In particular, the efficiency can be increased by reducing the holding hydraulic pressure at the position in the axial direction of the movable sheave 53 when the gear ratio is steady, and the vane hydraulic motor 550 can be downsized. (Reducing diameter) is possible.

  Further, since the power transmission means 551 rotates the motor case 550b and the movable sheave 53 integrally in the rotation direction of the primary shaft 51 when the power transmission means 551 does not operate, the vane hydraulic motor 550 and the movable sheave 53 are integrated. It also functions as an integral rotation mechanism that rotates.

  The bearing 550c and the power transmission means 551 described above constitute a relative movement mechanism that enables relative movement between the vane hydraulic motor 550 and the movable sheave 53. For example, when the motor case 550b rotates relative to the motor shaft 550a, this rotational force (torque) becomes the thrust of the vane type hydraulic motor 550 for sliding the movable sheave 53 via the power transmission means 551. . Here, the reaction force against the thrust is applied to the bearing 550c. The bearing 550c is fixed to the primary shaft 51 via the motor shaft 550a, and the reaction force against the thrust is received by the primary shaft 51. The case 550b does not move so much in the direction of the reaction force. As a result, the movable sheave 53 moves relative to the vane hydraulic motor 550 and approaches the fixed sheave 52. Thus, the movable sheave 53 can be slid in the axial direction of the primary shaft 51 by rotating the motor case 550b relative to the motor shaft 550a.

  Further, as described above, the reaction force against the thrust of the vane type hydraulic motor 550 can be received by the primary shaft 51 via the bearing 550c, and the relative rotation between the motor case 550b and the primary shaft 51 is caused by the movable sheave. Limited by 53 sliding strokes. For these reasons, in the first embodiment, the reaction force is not received by the stationary system such as the transaxle case 22 and the transaxle rear cover 23, and the rolling of the bearing 550c hardly occurs. Loss at 550c can be reduced.

  Here, as described above, since the second vanes 550f and 550f of the vane hydraulic motor 550 rotate integrally with the primary shaft 51, the motor case 550b of the vane hydraulic motor 550 rotates the vane hydraulic motor 550. If it is stopped, it rotates at the same rotational speed as that of the primary shaft 51, and if relative rotation occurs between the motor case 550b and the second vanes 550f and 550f, it rotates at a rotational speed different from that of the primary shaft 51.

  Next, the oil path to the first and second oil chambers 550g and 550h in the vane hydraulic motor 550 described above will be described.

First, as shown in FIGS. 2 and 3, while the motor shaft 550a first oil chamber 550g, an oil passage 550a _1, 550a ₁ communicating with 550g are formed, the oil passage 550a ₁ to the primary shaft 51, An oil passage 51a that communicates with 550a ₁ is formed, and supply of hydraulic oil to the first oil chambers 550g and 550g by these oil passages 550a ₁ , 550a ₁ , 51a or from the first oil chambers 550g and 550g, respectively. The hydraulic oil is discharged.

Furthermore, on its motor shaft 550a second oil chamber 550h, while the oil passage 550a _2, 550a ₂ which communicates with the 550h are formed, an oil passage 51b communicating with the oil passage 550a _2, 550a ₂ to the primary shaft 51 The hydraulic oil is supplied to the second oil chambers 550h and 550h or discharged from the second oil chambers 550h and 550h through these oil passages 550a ₂ , 550a ₂ and 51b.

  Here, the oil passages 51a and 51b of the primary shaft 51 communicate with a gear ratio control switching valve 56A as shown in FIG. The transmission ratio control switching valve 56A includes an oil tank OT, an oil pump (O / P) OP, an oil passage 56a, a regulator valve 56C, an oil passage 56b, a clamping pressure regulating valve 56B, and an oil passage 56c shown in FIG. Hydraulic fluid is supplied via

  The transmission ratio control switching valve 56A switches the position of the valve in which a plurality of oil passages are formed, thereby providing an oil chamber (the first oil chamber 550g, 550g or the second oil chamber 550h, which is the supply target of hydraulic oil). 550h) is switched. This switching is performed by adjusting the difference between the repulsive force of the spring disposed inside the cylinder and the pressure of the fluid such as air or hydraulic oil supplied to the inside. This is performed by a control unit (ECU) C.

  For example, the gear ratio control switching valve 56A switches the supply destination of hydraulic oil to the first oil chambers 550g and 550g by switching the valve position as shown in FIG. 5-1, and as shown in FIG. 5-3. By switching, the supply destination of the hydraulic oil is switched to the second oil chambers 550h and 550h.

  Further, the gear ratio control switching valve 56A supplies hydraulic oil of the same pressure to the first oil chambers 550g and 550g and the second oil chambers 550h and 550h by switching the position of the valve as shown in FIG. . As a result, the rotation of the vane hydraulic motor 550 is stopped, so that the gear ratio control switching valve 56A is also used when the gear ratio is fixed.

  Here, in the first embodiment, the rotation direction of the motor case 550b when the hydraulic pressure in the first oil chambers 550g and 550g is increased is referred to as normal rotation, and the movable sheave 53 is moved during the normal rotation. It is defined as approaching the fixed sheave 52 and increasing the speed (upshift) as the gear ratio decreases. The rotation direction of the motor case 550b when the hydraulic pressures in the second oil chambers 550h and 550h are increased is referred to as reverse rotation. During this reverse rotation, the movable sheave 53 is separated from the fixed sheave 52, and the gear ratio is increased. Defined as decelerating (downshifting).

  As described above, in the first embodiment, the vane type hydraulic motor 550 and the movable sheave 53 are integrally arranged on the primary shaft 51 in the internal space by the second extending portion 53b. The vane hydraulic motor 550 and the movable sheave 53 can be gathered together in a compact manner, and the movable sheave sliding mechanism 55 that slides the movable sheave 53 can be downsized. Further, the downsizing of the movable sheave sliding mechanism 55 enables the downsizing of the belt type continuously variable transmission 1 itself. Furthermore, the use of a hydraulic motor (vane hydraulic motor 550) and the provision of the power transmission means 551 described above eliminates the need for a gear group for transmitting the driving force of the motor to the movable sheave 53. Further downsizing of the movable sheave sliding mechanism 55 and the belt type continuously variable transmission 1 can be achieved.

  Further, since the movable sheave 53 is slid using the power transmission means 551 as described above, the driving loss generated by the conventional gear group is eliminated, and the driving loss in the movable sheave sliding mechanism 55 is reduced.

  Subsequently, on the primary shaft 51 side of the first embodiment, the movable sheave 53 of the primary pulley 50 is pressed against the fixed sheave 52 side, and the belt clamping pressure in the axial direction is applied between the fixed sheave 52 and the movable sheave 53. A pressing mechanism is provided for generation.

  2 and 4, the pressing mechanism communicates with a hydraulic chamber 57 formed between the vane type hydraulic motor 550 (motor case 550 b) and the back surface of the movable sheave 53, for example, the hydraulic chamber 57. An oil passage 53c formed in the first extending portion 53a of the movable sheave 53, an oil passage 51c formed in the primary shaft 51 communicating with the oil passage 53c, and a clamping pressure regulating valve communicating with the oil passage 51c 56B.

  In the first embodiment, the hydraulic piston member 57a is disposed between the vane type hydraulic motor 550 (motor case 550b) and the back surface of the movable sheave 53, and the wall surface of the hydraulic piston member 57a is movable. A hydraulic chamber 57 is formed in a space surrounded by the back surface of the sheave 53, the outer peripheral surface of the first extending portion 53a, the inner peripheral surface of the second extending portion 53b, and the outer peripheral surface of the primary shaft 51.

For example, as shown in FIG. 2, the hydraulic piston member 57a includes a first annular portion 57a ₁ disposed at a distance from the back surface of the movable sheave 53, and a vane from the outer diameter side of the first annular portion 57a _1. The first cylindrical portion 57a ₂ extending toward the hydraulic motor 550, the cylindrical portion 57a ₃ extending toward the vane hydraulic motor 550 from the inner diameter side of the first annular portion 57a ₁ , and the cylindrical shape A second annular portion 57a ₄ that closes an annular opening between the extending end of the portion 57a ₃ and the primary shaft 51, and the vane type hydraulic motor 550 from the primary shaft 51 side of the second annular portion 57a _4. and a second cylindrical portion 57a ₅ Metropolitan was extended.

Here, a cylindrical groove 550a ₃ is formed on the inner diameter side of the other end of the motor shaft 550a (the back side of the movable sheave 53), and the other end and the second annular portion 57a ₄ of the hydraulic piston member 57a. The second cylindrical portion 57a ₅ is inserted into the groove 550a ₃ until it comes into contact with the wall surface. Hydraulic piston member 57a of the first embodiment, the outer peripheral surface of the second cylindrical portion 57a ₅ and the wall surface of the groove 550a ₃ are spline-fitted through a spline 54C, which is therefore the motor shaft 550a, the primary It rotates integrally with the shaft 51, the movable sheave 53 and the like.

In the first embodiment, a gap in the axial direction is provided between the second cylindrical portion 57a ₅ in the groove 550a ₃ , and lubricating oil to the clutch mechanism 58 described later is provided from the gap. Supplied.

On the other hand, the first cylindrical portion 57a ₂ is formed such that the outer peripheral surface thereof is in contact with the inner peripheral surface of the second extending portion 53b in the movable sheave 53 or has a minute gap between the inner peripheral surface. . The outer peripheral surface of the first cylindrical portion 57a ₂ and between the inner peripheral surface of the second extending portion 53b, is provided an annular sealing member 57b shown in FIG.

Here, the outer diameter of the first cylindrical portion 57a ₂ (the sliding surface with the movable sheave 53) is larger than the thread inner diameter of the outer screw portion of the first power transmission member 551a in the power transmission means 551 as shown in FIG. preferably also in small, thereby, the hydraulic oil in the hydraulic chamber 57 oozed from the annular sealing member 57b between the outer peripheral surface of the first cylindrical portion 57a ₂ and the inner peripheral surface of the second extending portion 53b Is supplied to the outer screw portion and the inner screw portion of the power transmission means 551 by the centrifugal force accompanying the rotation of the primary shaft 51. Therefore, the lubrication performance is improved and the durability is improved without providing a lubricating oil supply mechanism, a lubricating oil supply path, and the like to the power transmission means 551 separately.

  The pressing mechanism supplies the hydraulic pressure from the clamping pressure regulating valve 56 </ b> B whose hydraulic oil supply pressure has been adjusted by the electronic control unit C to the hydraulic chamber 57, so that the belt is interposed between the fixed sheave 52 and the movable sheave 53. A pinching pressure can be generated to prevent the belt 80 described later from slipping.

  Here, since the clamping pressure regulating valve 56B communicates with the gear ratio control switching valve 56A described above via the oil passage 56c shown in FIG. 4, the hydraulic pressure from the clamping pressure regulating valve 56B is The oil is also supplied to the first oil chambers 550g and 550g and the second oil chambers 550h and 550h in the vane hydraulic motor 550 via the gear ratio control switching valve 56A.

  The hydraulic chamber 57 and the first and second oil chambers 550g and 550h of the vane hydraulic motor 550 include an oil passage 51c, an oil passage 56c, a transmission ratio control switching valve 56A, an oil passage 51a, and an oil passage 51b. Communicated through. Therefore, the hydraulic oil can be exchanged between the hydraulic chamber 57 and the first and second oil chambers 550g and 550h. This is particularly useful during a sudden deceleration downshift, and the hydraulic oil discharged from the hydraulic chamber 57 can be supplied to the second oil chambers 550h and 550h as will be described later. Can improve. Moreover, since the exchange of the hydraulic oil is made possible, the consumption amount of the hydraulic oil supplied from the oil pump OP can be reduced, and thereby the capacity of the oil pump OP can be reduced.

  As described above, on the primary pulley 50 side, there are two actuators for sliding the movable sheave 53 in the axial direction by hydraulic pressure, that is, the first actuator including the hydraulic chamber 57 and the clamping pressure regulating valve 56B, and the vane. A hydraulic actuator 550 and a second actuator composed of power transmission means 551 and the like are provided.

  Although two types of actuators are illustrated here, more types of actuators may be prepared. Moreover, although the actuator by hydraulic pressure was illustrated, it is not necessarily limited to this.

  Furthermore, on the primary shaft 51 side of the first embodiment, relative movement between the first power transmission member 551a and the second power transmission member 551b in the power transmission means 551 can be enabled, but the relative movement therebetween is stopped. A possible clutch mechanism 58 is provided.

  Here, as described above, the power transmission means 551 slides the movable sheave 53 in the axial direction when the motor case 550b rotates relative to the motor shaft 550a. In this case, the first power transmission member 551a is integrally fixed to the motor case 550b, and the second power transmission member 551b is integrally fixed to the movable sheave 53.

  That is, the first power transmission member 551a rotates integrally with the motor case 550b at the same rotational speed, while the second power transmission member 551b includes the movable sheave 53, the motor shaft 550a, the primary shaft 51, the fixed sheave 52, and the like. It rotates at the same number of rotations integrally with the hydraulic piston member 57a and the like.

  For this reason, at least one of the first power transmission member 551a itself or a member rotating at the same rotational speed and at least one of the second power transmission member 551b itself or a member rotating at the same rotational speed is used. The clutch mechanism 58 is provided between two members, and the clutch mechanism 58 is fastened or released to stop the relative movement between the first power transmission member 551a and the second power transmission member 551b, or the relative movement thereof. Can be performed.

  In the first embodiment, the clutch mechanism 58 is provided between the motor case 550b that rotates at the same rotational speed together with the first power transmission member 551a and the hydraulic piston member 57a that rotates at the same rotational speed together with the second power transmission member 551b. An example will be described in the case of providing.

  As shown in FIG. 6, the clutch mechanism 58 of the first embodiment includes a first clutch engaging portion 58a provided integrally with the wall surface of the motor case 550b, and a second clutch engaging provided on the hydraulic piston member 57a. The joint portion 58b and a clutch operating portion 58c for fastening and releasing the first and second clutch engaging portions 58a and 58b are provided.

For example, sliding the first clutch engaging portion 58a has a cylindrical member 58a ₁ which is provided on the wall surface of the motor case 550b to the primary shaft 51 to the center axis, axially provided on the outer peripheral surface of the cylindrical member 58a ₁ At least one possible and a clutch engaging member 58a ₂ annular. In the first embodiment, while forming an axial spline grooves on the outer circumferential surface of the cylindrical member 58a _1, the inner peripheral surface of the clutch engaging member 58a ₂ a groove can slide along the spline groove To form.

The second clutch engaging portion 58b is slid in the axial direction provided on the inner peripheral surface of the cylindrical member that can include the first clutch engaging portion 58a with the primary shaft 51 as the central axis. And at least one annular clutch engaging member 58b ₁ that is movable. In the first embodiment, the first cylindrical portion 57a ₂ of the hydraulic piston member 57a is used as the cylindrical member, and an axial spline groove is formed on the inner peripheral surface of the first cylindrical portion 57a ₂ . a groove can slide along the spline groove formed on the outer peripheral surface of the clutch engaging members 58b _1.

Here, the clutch engaging members 58a ₂ and 58b ₁ of the first and second clutch engaging portions 58a and 58b are alternately arranged with their annular surfaces facing each other. For this reason, a frictional force is generated on each annular surface in accordance with the axial pressing force when the respective clutch engaging members 58a ₂ and 58b ₁ are in contact with each other, and the first and second clutch engagements are generated. The joint portions 58a and 58b are fastened. In the first embodiment, the axial pressing force is generated by the clutch operating portion 58c.

First, in the clutch operating portion 58c of the first embodiment, a clutch hydraulic piston member 58c for pushing and fastening the clutch engaging members 58a ₂ and 58b ₁ of the first and second clutch engaging portions 58a and 58b. ₁ and a clutch hydraulic chamber 58c ₂ for pushing the clutch hydraulic piston member 58c ₁ are provided.

The clutch hydraulic piston member 58c ₁ of the first embodiment is formed so that one end on the outer peripheral side thereof is spline-fitted into the spline groove of the first cylindrical portion 57a ₂ and can slide in the axial direction. The other end on the peripheral side is brought into contact with the outer peripheral surface of the cylindrical portion 57a ₃ or is formed so as to have a minute gap with the outer peripheral surface.

Also provided is an annular seal member 58c ₃ shown in FIG. 6, thereby the hydraulic pressure chamber 58c ₂ is a clutch therebetween is formed between the hydraulic piston member 58c ₁ and the hydraulic piston member 57a for the clutch .

In the first embodiment, while the oil passage 57a ₆ communicating with the clutch hydraulic chamber 58c ₂ is formed in the hydraulic piston member 57a, the oil passage 51d primary shaft 51 communicating with the oil passage 57a ₆ Is formed. In the first embodiment, the oil passage 56d shown in FIG. 4 communicates with the oil passage 51d of the primary shaft 51, while the oil passage 56e shown in FIG. 4 communicates with the regulator valve 56C. A clutch oil pressure switching valve 56D is provided. The clutch hydraulic switching valve 56D is for regulating the hydraulic pressure of the clutch hydraulic chamber 58c _2, the operation of opening or closing is controlled by an electronic control unit (ECU) C.

For example, since the electronic control unit C is the clutch hydraulic oil is supplied to the clutch hydraulic chamber 58c ₂ by opening the clutch oil pressure switching valve 56D, with the hydraulic pressure of the rising of the clutch hydraulic chamber 58c ₂ clutch hydraulic piston member 58c ₁ is use is pushed toward the clutch engaging member 58a _2, 58b ₁ of each, it can be fastened to the clutch engaging member 58a _2, 58b ₁ of the respective.

Further, in the clutch operating portion 58c of the first embodiment, the clutch hydraulic piston member 58c ₁ pushed by the hydraulic pressure of the clutch hydraulic chamber 58c ₂ is pushed back in the opposite direction, and the respective clutch engaging members 58a ₂ , An elastic member 58c ₄ for releasing the fastening state of 58b ₁ is provided. As the elastic member 58c _4, for example, compressed by a hydraulic piston member 58c ₁ clutch is pushed by hydraulic pressure, using a disc spring or the like to push back the hydraulic piston member 58c ₁ clutch by the hydraulic pressure is decreased .

For example, when the electronic control unit C closes the clutch hydraulic pressure switching valve 56D, the clutch hydraulic fluid in the clutch hydraulic chamber 58c ₂ is discharged, and the elastic member 58c ₄ is accompanied with a decrease in the hydraulic pressure in the clutch hydraulic chamber 58c _2. Since the clutch hydraulic piston member 58c ₁ is pushed back, the clutch engaging members 58a ₂ and 58b ₁ can be released.

Further, in the clutch mechanism 58 of the first embodiment, supplying lubricating oil to the sliding portions (first and second clutch engaging portion 58a, 58b and a hydraulic piston member 58c ₁ like sliding portion of the clutch) A lubricating oil supply path is provided.

In the first embodiment, the gap in the groove 550a ₃ of the motor shaft 550a and the gap of the spline groove in the spline 54C communicating with the gap are used as the lubricating oil supply path.

Further, oil as a lubricating oil supply path, as shown in FIG. 6, thereby forming an oil passage 51e communicating with the gap in the groove 550a ₃ of the motor shaft 550a to the primary shaft 51, communicating with the gap of the splines 54C The path 550a ₄ is provided on a contact surface of the motor shaft 550a with the second annular portion 57a ₄ of the hydraulic piston member 57a.

Here, by causing through the oil passage 56f and the communication that indicates the oil passage 51e of the primary shaft 51 in FIG. 4, for example, supplies lubricating oil from Lubrication valve 56E to the gap in the groove 550a _3. Incidentally, separately provided switching valve communicating with and regulator valve 56C communicates with an oil passage 51e of the primary shaft 51, by opening and closing operation of the switching valve may be supplied to the lubricating oil in the gap in the groove 550a _3.

Further, in the vane type hydraulic motor 550 of the first embodiment, as shown in FIG. 6, the end of the motor shaft 550a protrudes from the wall surface of the motor case 550b, whereby the wall surface of the motor case 550b and the hydraulic piston member 57a. a gap is formed between the second annular portion 57a ₄ of the. For this reason, this clearance and the sliding portion of the clutch mechanism 58 and the oil passage 550a ₄ of the motor shaft 550a communicate with each other, and the lubricating oil supplied to the clearance in the groove 550a ₃ moves to the sliding portion of the clutch mechanism 58. Sent.

In this way, by configuring the lubricating oil supply path to the sliding portion of the clutch mechanism 58 using the gap in the groove 550a ₃ and the gap of the spline 54C communicating with the gap, the durability of the clutch mechanism 58 is increased. In addition to the improvement in performance, the belt-type continuously variable transmission 1 can be shortened in the axial direction.

Further, in the clutch mechanism 58 of the first embodiment, the outermost diameter portion (the inner peripheral surface portion of the first cylindrical portion 57a ₂ ) is closer to the inner diameter side than the thread inner diameter of the outer screw portion of the first power transmission member 551a. Has been placed. As a result, the lubricating oil supplied to the sliding portion of the clutch mechanism 58 is supplied to the outer screw portion and the inner screw portion of the power transmission means 551 by the centrifugal force accompanying the rotation of the primary shaft 51. For this reason, the lubricating performance of the power transmission means 551 can be improved in combination with the hydraulic oil in the hydraulic chamber 57 that has oozed out from the seal member 57b described above, and the durability can be further improved.

Further, the clutch mechanism 58 is configured to be disposed in a space (an annular space surrounded by the first cylindrical portion 57a ₂ and the cylindrical portion 57a ₃ ) formed by the hydraulic piston member 57a. Therefore, an increase in the size of the belt type continuously variable transmission 1 due to the provision of the clutch mechanism 58 can be suppressed.

A friction material such as a facing material may be provided on each of the annular surfaces of the clutch engaging members 58a ₂ and 58b ₁ , and in such a case, when the friction material causes slipping by oil, It is preferable not to supply lubricating oil to the friction material.

  Next, the configuration on the secondary shaft 61 side will be described.

  The secondary shaft 61 is provided with a secondary pulley 60 shown in FIG. As shown in FIG. 7, the secondary pulley 60 includes a fixed sheave 62 that is integrally disposed on the outer periphery of the secondary shaft 61, and a movable sheave 63 that can slide in the axial direction of the secondary shaft 61. . A V-shaped groove 80 b is formed between the opposed surfaces of the fixed sheave 62 and the movable sheave 63.

  Here, the movable sheave 63 is splined to the secondary shaft 61 by a spline 64A. For this reason, the movable sheave 63 rotates integrally with the secondary shaft 61 and the fixed sheave 62 and slides on the secondary shaft 61 in the axial direction so as to approach or separate from the fixed sheave 62. . Since the fixed sheave 62 and the movable sheave 63 rotate integrally as described above, even if the following pressing mechanism generates a belt clamping pressure on the belt 80 between the fixed sheave 62 and the movable sheave 63, the metal sheave 62 is made of metal. Belt 80 can be used, and the belt type continuously variable transmission 1 can be increased in torque.

  In addition, the secondary shaft 61 is provided with a pressing mechanism that presses the movable sheave 63 toward the fixed sheave 62 and generates a belt clamping pressure in the axial direction between the fixed sheave 62 and the movable sheave 63. Here, as the pressing mechanism of the first embodiment, two types of torque cam 65 and hydraulic chamber 66 are prepared.

  First, the torque cam 65 will be described in detail.

  The torque cam 65 of the first embodiment includes, for example, a mountain-shaped first engaging portion 65a provided in an annular shape on the movable sheave 63, as shown in FIGS. It is composed of a torque cam main body 65c having a mountain-and-valley-like second engaging portion 65b facing the joint portion 65a, and a plurality of spherical members 65d arranged between the first and second engaging portions 65a and 65b. The

  Here, the torque cam main body 65 c is fixed to the secondary shaft 61 and the movable sheave 63 by a bearing 61 a shown in FIG. 7 fixed to the secondary shaft 61 and a bearing 61 b arranged between the secondary shaft 61. Relative rotation about the rotation axis is possible, and the position in the axial direction is kept constant.

  For example, relative rotation between the torque cam main body 65c and the movable sheave 63 with the approach of the movable sheave 63 to the fixed sheave 62 (in other words, the separation of the first engagement portion 65a from the second engagement portion 65b). The torque cam 65 changes from the state shown in FIG. 8A to the state shown in FIG. 8B while each spherical member 65d rotates. Thereby, a surface pressure is generated between the first engaging portion 65a, the second engaging portion 65b, and the spherical member 65d, and the second engaging portion 65b and the spherical member 65d press the first engaging portion 65a. Therefore, a belt clamping pressure is generated between the fixed sheave 62 and the movable sheave 63, and the belt 80 can be prevented from slipping.

  Further, thrust is generated in the movable sheave 63 when the second engaging portion 65b and the spherical member 65d press the first engaging portion 65a. However, since the torque cam main body 65c and the movable sheave 63 rotate relative to each other, the movable sheave 63 63 and the fixed sheave 62 do not twist each other. For this reason, the durability of the belt 80 can be improved and the speed ratio can be widened, and the relative position between the primary pulley 50 and the secondary pulley 60 can be maintained at the initial set value. It contributes to the improvement.

  Here, the reaction force against the thrust of the torque cam 65 due to the above surface pressure can be received by the secondary shaft 61 via the bearing 61a. Thus, the reaction force is not received in the stationary system as in the case of the primary pulley 50, and the rolling of the bearing 61a hardly occurs, so that the loss in the bearing 61a can be reduced.

  Further, since the operating portion (first and second engaging portions 65a and 65b, spherical member 65d) of the torque cam 65 is disposed on the outer diameter side of the movable sheave 63, the first engaging portion 65a and the second engaging portion 65a It is also possible to reduce the surface pressure between the engaging portion 65b and the spherical member 65d.

  Next, the hydraulic chamber 66 will be described in detail.

  The hydraulic chamber 66 according to the first embodiment is formed of a space portion on the opposite side of the groove 80 b in the movable sheave 63 and a circular member 67 provided on the secondary shaft 61 and having the secondary shaft 61 as a central axis.

  Here, since the hydraulic chamber 66 is arranged on the inner diameter side of the movable sheave 63, the volume of the hydraulic chamber 66 can be reduced, so that the flow rate of the hydraulic chamber 66 at the time of sudden shift or the like can be reduced.

  The hydraulic chamber 66 communicates with, for example, an oil passage 61c shown in FIG. 4 formed in the secondary shaft 61, and further communicates with the clamping pressure regulating valve 56B via the oil passage 51c communicating with the oil passage 61c. is doing.

  Thus, the pressing mechanism of the secondary pulley 60 constituted by the hydraulic chamber 66, the oil passage 61c, and the clamping pressure regulating valve 56B is from the clamping pressure regulating valve 56B in which the supply pressure of the hydraulic oil is adjusted by the electronic control unit C. Is supplied to the hydraulic chamber 66 to generate a belt clamping pressure between the fixed sheave 62 and the movable sheave 63, thereby preventing the belt 80 from slipping.

  Further, even when the torque ratio is changed (when the movable sheave 63 is driven / non-driven in the secondary pulley 60) and the torque is disturbed and the thrust by the torque cam 65 cannot be obtained, the torque cam 65 is separately hydraulically independent. A desired belt clamping pressure can be generated by a pressing mechanism including an operating hydraulic chamber 66 and the like. As a result, the belt 80 can be more reliably prevented from slipping, so that reliability and drivability can be improved.

  Here, the hydraulic chamber 66 according to the first embodiment is provided with an elastic member 68 such as a coil spring having one end fixed to the wall surface of the space portion of the movable sheave 63 and the other end fixed to the circular member 67. ing.

  In the first embodiment, the torque cam 65 is set at a cam angle (for example, a non-linear cam) such that the thrust by the torque cam 65 is lower than the required thrust, and the shortage is pressed by the hydraulic chamber 66 or the like. The mechanism or / and the elastic member 68 are set to compensate. As a result, the belt 80 does not have to be pinched with an unnecessarily large force, so that the durability of the belt 80 can be improved, loss in the belt 80 can be reduced, and power transmission efficiency can be improved. .

  Further, the thrust corresponding to the torque when the internal combustion engine 10 is not driven may be set to be received by the pressing mechanism including the hydraulic chamber 66 or the like and / or the elastic member 68, and this may occur due to the operation of the torque cam 65. It is possible to suppress the movement of the movable sheave 63 (in other words, speed change) and keep the speed ratio constant. Further, the belt clamping pressure can be kept at a required value.

  Further, the pressing mechanism on the secondary pulley 60 side is not necessarily limited to two types as in the first embodiment, and may be one type or three or more types. In order to improve the controllability of the belt clamping pressure between the fixed sheave 62 and the movable sheave 63, it is preferable to provide at least two types of pressing mechanisms. That is, the belt clamping pressure is shared by each pressing mechanism, and at least one of them is a pressing mechanism (hydraulic chamber 66 of the first embodiment) that is operated by hydraulic pressure, thereby improving the controllability of the belt clamping pressure. Can be made.

  As described above, in the first embodiment, the secondary pulley 60 side also includes an actuator that slides the movable sheave 63 in the axial direction by hydraulic pressure, that is, the hydraulic chamber 66, the clamping pressure regulating valve 56B, and the like. An actuator is provided. In addition, although the actuator by hydraulic pressure was illustrated here, it is not necessarily limited to this.

  In the belt type continuously variable transmission 1 shown above, the belt 80 is wound around the V-shaped grooves 80a and 80b of the primary pulley 50 and the secondary pulley 60, respectively. The belt 80 is an endless belt composed of a number of metal pieces and a plurality of steel rings, and the torque of the internal combustion engine 10 transmitted to the primary pulley 50 via the belt 80 is transmitted to the secondary pulley 60. Communicated.

  Here, in the first embodiment, torque is transmitted by the torque cam main body 65c. A counter drive pinion 92 is fixed to the internal combustion engine 10 side of the secondary shaft 61 that rotates together with the torque cam main body 65c as shown in FIG. 1, and bearings 87 of the secondary shaft 61 are provided on both sides of the counter drive pinion 92. , 88 are arranged.

  For this reason, the torque transmitted to the secondary pulley 60 is transmitted to the power transmission path 90 and the final speed reducer 70, which will be described later, via the torque cam main body 65c, the secondary shaft 61, and the counter drive pinion 92. It is transmitted to the drive shaft 101 via the gear group of the machine 70.

  In the first embodiment, as shown in FIG. 1, a parking gear 89 is integrally provided on the torque cam main body 65c. For example, the parking gear 89 is fitted and fixed to the outer peripheral surface of the torque cam main body 65c. Therefore, the axial length of the belt type continuously variable transmission 1 on the secondary pulley 60 side can be shortened. That is, the conventional parking gear is arranged on the secondary shaft 61 between the secondary pulley 60 and the transaxle rear cover 23. In the first embodiment, the arrangement location is separately secured for the parking gear. Since it is not necessary, the axial length on the secondary pulley 60 side can be shortened.

  Next, a power transmission path 90 having an intermediate shaft 91 parallel to the secondary shaft 61 is provided between the counter drive pinion 92 and a final reduction gear 70 described later. The intermediate shaft 91 is rotatably supported by bearings 85 and 86, and includes a counter driven gear 93 and a final drive pinion 94 which are engaged with the counter drive pinion 92 on the shaft.

  Next, the final reduction gear 70 will be described. The final reduction gear 70 includes a differential case 71 having a hollow inside, a pinion shaft 72, pinions 73 and 74, and side gears 75 and 76.

  First, the differential case 71 is rotatably supported by bearings 77 and 78, and a ring gear 79 meshed with the final drive pinion 94 is provided on the outer periphery thereof.

  The pinion shaft 72 is attached to the hollow portion of the differential case 71, and the pinions 73 and 74 are fixed to the pinion shaft 72.

  The side gears 75 and 76 are fixed to a drive shaft 101 (here, a front drive shaft) to which the wheel 100 is attached.

  In the transaxle case 22 configured as described above, the lubricating oil stored in the bottom (oil pan) is scraped up by the rotating ring gear 79 and transmitted to the meshing surfaces of the gears 94, 93, 92. While splattering, each component (for example, each shaft 101, 91, 61, each bearing 83-88 etc.), such as the final reduction gear 70, is lubricated, and it falls on the inner wall surface of the transaxle case 22 and falls. 51 etc. are lubricated.

  Each component including the belt-type continuously variable transmission 1 described above is controlled by an electronic control unit (ECU) C which is a control means of the belt-type continuously variable transmission 1 based on information from various sensors. The electronic control unit C includes a shift control for the belt-type continuously variable transmission 1 in accordance with a driving state based on data for performing shift control of the belt-type continuously variable transmission 1, for example, information such as an accelerator opening degree and a vehicle speed. Data for controlling the ratio is stored in advance. The operation of the movable sheave sliding mechanism 55 and the pressing mechanism (torque cam 65, hydraulic chambers 57, 66) and the operation of the clutch mechanism 58 when controlling the transmission ratio will be described in detail below with reference to the flowchart of FIG.

  First, the electronic control unit C compares the actual speed ratio (hereinafter referred to as “actual speed ratio”) with the target speed ratio, and the actual speed ratio becomes the target speed ratio (that is, the actual speed ratio). = Target gear ratio) is determined (step ST1).

  Here, it is determined whether or not the actual speed ratio and the target speed ratio are completely the same, but within a range that does not affect the power performance or drivability of the vehicle, for example ± 0.1% A width may be given to the target speed ratio, and if the actual speed ratio is within the range of the width, it may be set to determine that “actual speed ratio = target speed ratio”.

  If the actual gear ratio does not reach the target gear ratio in step ST1, the electronic control unit C next determines the magnitude relationship between the actual gear ratio and the target gear ratio. In the first embodiment, first, it is determined whether or not the actual gear ratio is larger than the target gear ratio (step ST2).

  Here, if the actual gear ratio is larger than the target gear ratio, the electronic control unit C puts the clutch mechanism 58 into a released state (step ST3).

Leave Specifically, already if the closed clutch hydraulic switching valve 56D, maintaining the left supply stop state of the clutch hydraulic fluid to the clutch hydraulic chamber 58c ₂ of the condition, to release the clutch mechanism 58 To. Also, if the opened clutch hydraulic switching valve 56D, which by closed to stop the supply of the clutch operating oil to the hydraulic chamber 58c ₂ clutch and to release the clutch mechanism 58.

  In such a state where the clutch mechanism 58 is released, the power transmission means 551 is operated by driving the vane hydraulic motor 550, and the movable sheave 53 of the primary pulley 50 can be slid in the axial direction. it can.

  However, the movable sheave 53 is subjected to a force that separates from the fixed sheave 52 by the belt tension caused by the belt clamping pressure on the secondary pulley 60 side (in other words, the reaction force of the belt clamping pressure in the primary pulley 50). If the position of the movable sheave 53 in the axial direction is constant (that is, the oil pressure in the first oil chambers 550g and 550g and the second oil chambers 550h and 550h in the vane type hydraulic motor 550 is substantially equal). If there is no change in the hydraulic pressure in the hydraulic chamber 57), the first and second power transmission members 551a and 551b of the power transmission means 551 are applied to the outer screw portion and the inner screw portion.

  On the other hand, an axial backlash is provided between the external thread and the internal thread, but the backlash is clogged away from the fixed sheave 52 due to the reaction force of the belt clamping pressure.

  For this reason, in order to operate the power transmission means 551 so that the movable sheave 53 approaches the fixed sheave 52 in a state where the reaction force of the belt clamping pressure acts on the outer screw portion and the inner screw portion, It is necessary to increase the driving force of the vane hydraulic motor 550 (that is, increase the hydraulic pressure of the hydraulic oil), and there is a risk that a shift shock will occur.

  Therefore, the electronic control unit C according to the first embodiment performs the process of step ST3 to release the clutch mechanism 58, and then controls the clamping pressure regulating valve 56B to control the hydraulic pulley as a pressing mechanism for the primary pulley 50. The hydraulic pressure applied to 57 is increased (step ST4).

  As a result, the movable sheave 53 slides toward the fixed sheave 52 within a backlash range between the outer screw portion and the inner screw portion of the power transmission means 551, and the belt clamping pressure applied to the outer screw portion and the inner screw portion. The burden of reaction force is reduced. For this reason, in such a state, the power transmission means 551 can be operated with a small amount of energy (specifically, a small rotational force of the motor case 550b), so that the driving force (hydraulic pressure) of the vane hydraulic motor 550 can be reduced. It becomes possible.

  Here, in the first embodiment, the hydraulic pressure increased in step ST4 depends on the first oil chamber 550g, 550g of the vane hydraulic motor 550 and / or the first oil chamber 550g depending on the valve position of the transmission ratio control switching valve 56A. Two oil chambers 550h and 550h are also hung.

  For this reason, the hydraulic pressure is preferably increased to a predetermined hydraulic pressure that can be changed from the actual gear ratio to the target gear ratio. For example, the electronic control unit C corresponds to the relationship between the hydraulic pressure applied to the first oil chamber 550g, 550g or the second oil chamber 550h, 550h and the moving amount of the movable sheave 53 in the axial direction or the position of the movable sheave 53 in the axial direction. , And map data indicating the correspondence between the position of the movable sheave 53 in the axial direction and the gear ratio are prepared in advance, and the hydraulic pressure of the clamping pressure regulating valve 56B is set based on the respective map data.

  Note that if the oil pressure is increased to a predetermined hydraulic pressure that can be changed from the actual gear ratio to the target gear ratio after the operation of step ST5 described later, in step ST4, for example, the external thread portion in the power transmission means 551 described above is used. The hydraulic sheave 53 is raised to a hydraulic pressure that allows the movable sheave 53 to slide within a backlash range between the inner screw portion and the inner screw portion.

  Subsequently, the electronic control unit C performs pressure control of the working fluid of the gear ratio control switching valve 56A in a state where the predetermined hydraulic pressure is applied as described above, and adjusts the valve position as shown in FIG. The hydraulic oil supply destination is switched to only the upshift side oil chamber (first oil chamber 550g, 550g) of the vane hydraulic motor 550 (step ST5).

  As a result, the hydraulic oil is supplied to the first oil chambers 550g and 550g and the hydraulic oil in the second oil chambers 550h and 550h is discharged while the motor case 550b rotates relative to the motor shaft 550a (primary shaft 51). Then, the movable sheave 53 of the primary pulley 50 approaches the fixed sheave 52 via the power transmission means 551. Accordingly, the movable sheave 63 of the secondary pulley 60 is separated from the fixed sheave 62.

  As described above, the electronic control unit C reduces the gear ratio and performs speed increase control (upshift control).

  Here, after the oil passage switching to the first oil chambers 550g and 550g in step ST5, the movable sheave 63 rotates together with the fixed sheave 62, the secondary shaft 61 and the bearing 61a on the secondary pulley 60 side. Relative rotation occurs between the sheave 63 and the torque cam main body 65c, and the torque cam 65 changes, for example, from the separated state shown in FIG. 8B to the close state shown in FIG. Therefore, a belt clamping pressure is generated between the fixed sheave 52 and the movable sheave 53, and the belt 80 can be prevented from slipping.

  On the primary pulley 50 side, hydraulic oil corresponding to the sliding amount of the movable sheave 53 is supplied to the hydraulic chamber 57 via the oil passage 51c, so that the hydraulic pressure in the hydraulic chamber 57 is movable by the vane hydraulic motor 550. The sliding force of the sheave 53 is assisted. For this reason, even if the vane type hydraulic motor 550 has a low output, the movable sheave 53 can be slid sufficiently, so that a small vane type hydraulic motor 550 having a reduced output can be used.

  On the secondary pulley 60 side, the hydraulic oil in the hydraulic chamber 66 is discharged through the oil passage 61c. Here, since the oil passage 61 c communicates with the oil passage 51 c as shown in FIG. 4, the hydraulic oil discharged from the hydraulic chamber 66 of the secondary pulley 60 is supplied to the hydraulic chamber 57 of the primary pulley 50. Further, the hydraulic oil discharged from the hydraulic chamber 66 is also supplied to the first oil chambers 550g and 550g via the gear ratio control switching valve 56A. As described above, since the discharged hydraulic oil can be circulated and sent to another oil chamber, the consumption amount of the hydraulic oil can be reduced, and the capacity of the oil pump OP can be reduced.

  Subsequently, the electronic control unit C returns to the determination process of step ST1 and executes the above-described process or a process described later according to the determination result.

  For example, if the actual gear ratio has reached the target gear ratio in step ST1, the electronic control unit C engages the clutch mechanism 58 (step ST6).

In the first embodiment, the electronic control unit C performs the opening control of the clutch hydraulic switching valve 56D, increase the hydraulic pressure by supplying the clutch hydraulic oil to the hydraulic chamber 58c ₂ clutch. As a result, the clutch hydraulic piston member 58c ₁ is pushed in the axial direction, and the clutch engaging members 58a ₂ and 58b _{1 of} the first and second clutch engaging portions 58a and 58b are fastened. The relative rotation and the relative movement in the axial direction between the power transmission member 551a and the second power transmission member 551b are stopped, and the power transmission means 551 cannot be operated.

  In this state, the electronic control unit C adjusts the valve position of the transmission ratio control switching valve 56A as shown in FIG. 5-2, and adjusts the pressure between the first oil chambers 550g and 550g and the second oil chambers 550h and 550h. The same oil pressure is applied from the pressure valve 56B (step ST7).

  Accordingly, the relative rotation of the vane hydraulic motor 550 with respect to the primary shaft 51 stops, and the vane hydraulic motor 550 rotates together with the primary shaft 51 and the movable sheave 53. For this reason, since there is no rotational difference between the vane type hydraulic motor 550 and the primary shaft 51 or the movable sheave 53, it is possible to reduce loss due to unnecessary relative rotation, friction, or the like.

  The hydraulic pressure from the clamping pressure regulating valve 56B is also applied to the hydraulic chamber 57 of the primary pulley 50 and the hydraulic chamber 66 of the secondary pulley 60. Therefore, the fixed sheave 52 and the movable sheave 53 in the primary pulley 50 are A belt clamping pressure is generated between the stationary sheave 62 and the movable sheave 63 in the secondary pulley 60 and the belt 80 can be prevented from slipping.

  Thereafter, the electronic control unit C controls the clamping pressure regulating valve 56B to reduce the hydraulic pressure applied to the first and second oil chambers 550g and 550h and the hydraulic chambers 57 and 66 to a predetermined hydraulic pressure (step ST8). Thereby, the power of the oil pump OP can be reduced and the efficiency of the transmission can be improved.

  Here, the predetermined hydraulic pressure in this case is set based on the hydraulic pressure applied to the respective hydraulic chambers 57 and 66 that can generate the belt clamping pressure to such an extent that the belt 80 does not slip.

  In this manner, in the belt type continuously variable transmission 1, the actual speed ratio is fixed to the target speed ratio.

  Here, in the first embodiment, the clutch mechanism 58 stops the relative rotation and the axial movement of the first and second power transmission members 551a and 551b of the power transmission means 551 in step ST6. I am letting.

  Therefore, the axial direction between the fixed sheave 52 and the movable sheave 53 is not affected by the backlash between the outer screw portion and the inner screw portion of the first and second power transmission members 551a and 551b. The position can be kept constant at the target gear ratio.

  Even if a torque (rotational torque) is generated in the power transmission means 551 due to the reaction force or disturbance of the belt clamping pressure, the axial direction position between the fixed sheave 52 and the movable sheave 53 remains the target speed ratio. Can be kept constant.

  Furthermore, it is possible to suppress an increase in oil pressure and oil amount for absorbing the return torque. That is, if the clutch mechanism 58 is not provided, for example, the return torque must be absorbed by driving the vane hydraulic motor 550, but the return torque is absorbed by the clutch mechanism 58 that can be operated with a small amount of hydraulic oil. As a result, an increase in hydraulic pressure or oil amount is suppressed. In particular, the smaller the coefficient of friction between the outer screw portion and the inner screw portion, the more remarkable is the effect of suppressing the increase in oil pressure and oil amount.

  By the way, in the processing operation described above, after the clutch mechanism 58 is released in step ST3, the hydraulic pressure in the hydraulic chamber 57 is increased in step ST4, and then the vane hydraulic motor 550 is driven to upshift. It is carried out. However, if the clutch mechanism 58 is first released when the frictional resistance between the outer screw portion and the inner screw portion of the power transmission means 551 is small, the belt clamping pressure described above at the outer screw portion and the inner screw portion. Therefore, the movable sheave 53 may slide slightly in the separation direction.

  Therefore, when using such a power transmission means 551 having a small frictional resistance between the outer screw portion and the inner screw portion, the pressing force toward the fixed sheave 52 is increased by increasing the hydraulic pressure in the one-end hydraulic chamber 57. It is preferable that the vane type hydraulic motor 550 is driven by putting it on the movable sheave 53 and then releasing the clutch mechanism 58.

  On the other hand, if the actual speed ratio is smaller than the target speed ratio in step ST2, the electronic control unit C determines whether or not the difference Δγ (= target speed ratio−actual speed ratio) is larger than a predetermined threshold value. Determination is made (step ST9). As the threshold value used here, a minimum value (for example, 0.5) of the difference that is determined to be a rapid downshift by the electronic control unit C is set in advance.

  Here, if the difference Δγ is larger than the threshold value (that is, when a rapid downshift is requested), the electronic control unit C releases the clutch mechanism 58 (step ST10), and the clamping pressure regulating valve 56B. To control the hydraulic pressure applied to the hydraulic chamber 57 of the primary pulley 50 (step ST11).

  In such a case, it is preferable to increase the hydraulic pressure to a predetermined hydraulic pressure that is the closest gear ratio to the target gear ratio when the actual gear ratio is brought close to the target gear ratio. For example, based on the respective map data described above. The hydraulic pressure of the clamping pressure regulating valve 56B is set.

  If the predetermined hydraulic pressure is increased after the operation of step ST12 described later, in step ST11, for example, the movable sheave 53 can slide within the range of the play of the power transmission means 551 described above. Increase to hydraulic pressure.

  Thus, by increasing the hydraulic pressure in the hydraulic chamber 57, the burden of reaction force of the belt clamping pressure applied to the outer screw portion and the inner screw portion of the power transmission means 551 is reduced.

  Therefore, in this state, the electronic control unit C controls the gear ratio control switching valve 56A to adjust the valve position as shown in FIG. 5-3, and the supply destination of the hydraulic oil is lowered to the vane hydraulic motor 550. Only the shift side oil chamber (second oil chamber 550h, 550h) is switched (step ST12).

  As a result, hydraulic oil is supplied to the second oil chambers 550h and 550h, and the motor case 550b rotates relative to the motor shaft 550a (primary shaft 51) while the hydraulic oil in the first oil chambers 550g and 550g is discharged. Then, the movable sheave 53 of the primary pulley 50 is separated from the fixed sheave 52 via the power transmission means 551, and the movable sheave 63 of the secondary pulley 60 approaches the fixed sheave 62.

  After the speed ratio closest to the target speed ratio described above is thus obtained, the electronic control unit C controls the clamping pressure regulating valve 56B to reduce the hydraulic pressure applied to the hydraulic chamber 57 of the primary pulley 50 (step). ST13).

  The hydraulic pressure in this case is the hydraulic pressure required to change from the gear ratio closest to the target gear ratio to the target gear ratio, for example, by the power transmission means 551 operated by the reaction force of the hydraulic pressure and the belt clamping pressure. The hydraulic pressure of the clamping pressure regulating valve 56B is set based on the map data indicating the correspondence relationship between the movable sheave 53 in the axial direction and the correspondence relationship between the position of the movable sheave 53 in the axial direction and the gear ratio.

  As a result, the power transmission means 551 having a small frictional resistance between the outer screw portion and the inner screw portion is operated by the reaction force of the belt clamping pressure, and the movable sheave 53 of the primary pulley 50 is fixed to the fixed sheave 52 until the target gear ratio is reached. The movable sheave 63 of the secondary pulley 60 approaches the fixed sheave 62.

  The electronic control unit C increases the gear ratio as described above and performs rapid deceleration control (rapid downshift control).

  Then, the electronic control unit C returns to the determination process of step ST1 and executes the above-described process or the process described later according to the determination result.

  Here, in the processing operation described above, after the clutch mechanism 58 is released in step ST10, the hydraulic pressure in the hydraulic chamber 57 is increased in step ST11, and then the vane hydraulic motor 550 is driven to decrease. Although the shift is performed, the hydraulic pressure in the hydraulic chamber 57 is increased to apply a pressing force toward the fixed sheave 52 to the movable sheave 53, and then the clutch mechanism 58 is released and the vane hydraulic pressure is released. The motor 550 may be driven.

  Subsequently, when the difference Δγ is smaller than the threshold value in step ST9 (that is, when a rapid downshift is not requested), the electronic control unit C puts the clutch mechanism 58 into a released state (step ST14).

  Thereby, the power transmission means 551 having a small frictional resistance between the outer screw portion and the inner screw portion is operated by the reaction force of the belt clamping pressure, and the movable sheave 53 of the primary pulley 50 is separated from the fixed sheave 52. Start sliding.

  Here, when the difference Δγ is smaller than the threshold value (when the rapid downshift is not required), the moving amount of the movable sheave 53 in the axial direction until reaching the target gear ratio is small. Therefore, if the clutch mechanism 58 is released as described above in this case, the movable sheave 53 may suddenly slide in the axial direction due to the reaction force of the belt clamping pressure, and a shift shock may occur. As described above, the movable sheave 53 may slide and not be fixed to the target gear ratio.

  Therefore, in such a case, the electronic control unit C controls the gear ratio control switching valve 56A and the clamping pressure regulating valve 56B, and supplies the oil pressure as the standby pressure to the upshift-side oil chamber (the first oil pressure chamber (first output)) of the vane hydraulic motor 550. Oil chamber 550g, 550g) (step ST15). The hydraulic pressure here is set so that the driving force of the vane hydraulic motor 550 is generated so that the movable sheave 53 does not slide suddenly due to the reaction force of the belt clamping pressure.

  Subsequently, the electronic control unit C reduces the hydraulic pressure applied to the hydraulic chamber 57 of the primary pulley 50 in step ST13 described above, and separates the movable sheave 53 of the primary pulley 50 from the fixed sheave 52 until the target gear ratio is reached. The movable sheave 63 of the secondary pulley 60 is moved closer to the fixed sheave 62.

  After performing the downshift control as described above, the electronic control unit C returns to the determination process of step ST1 and executes the above-described process according to the determination result.

As described above, according to the belt type continuously variable transmission 1 of the first embodiment, a hydraulic motor that does not require a gear group (here, the vane type hydraulic motor 550) is used as the main drive source of the movable sheave 53. A power transmission means 551 as a means for transmitting the driving force to the movable sheave 53 is also arranged in the same space in the internal space on the back side. Further, between the rear and the hydraulic motor of the movable sheave 53, while forming a hydraulic chamber 57 by placing the hydraulic piston member 57a, the space (the first cylindrical portion 57a ₂ and the cylindrical portion of the hydraulic piston member 57a It is arranged the clutch mechanism 58 surrounded by an annular space) by a 57a _3. For this reason, this belt type continuously variable transmission 1 can be reduced in size.

  Further, since the vane type hydraulic motor 550, the power transmission means 551, the hydraulic chamber 57 and the clutch mechanism 58 are arranged on the same axis, not only the above-described downsizing but also the drag resistance and the gear ratio at those sliding parts. The holding oil pressure can be reduced to keep the pressure constant, and the response at the time of shifting can be improved.

  Next, a belt type continuously variable transmission according to a second embodiment of the present invention will be described with reference to FIGS.

  The belt-type continuously variable transmission 1 of the second embodiment is obtained by changing the configuration on the primary pulley 50 side in the above-described first embodiment as follows, and other configurations are the belt-type continuously variable transmission of the first embodiment. Same as machine 1.

  First, in the second embodiment, a drive source in the axial direction of the movable sheave 53 in the movable sheave sliding mechanism 55 is assigned to the hydraulic chamber 57 instead of the vane hydraulic motor 550. That is, the hydraulic chamber 57 of the second embodiment functions as a pressing mechanism similar to that of the first embodiment that generates the belt clamping pressure between the fixed sheave 52 and the movable sheave 53, and the movable sheave 53 is fixed to the fixed sheave 52. It is functioning as a drive source for approaching. In the second embodiment, since the vane hydraulic motor 550 is not provided, the oil passages 51a and 51b and the gear ratio control switching valve 56A associated therewith are not provided.

  Here, in the internal space on the back surface of the movable sheave 53 (substantially annular space surrounded by the first extending portion 53a, the second extending portion 53b and the primary shaft 51), the hydraulic piston is the same as in the first embodiment. The member 57a is disposed, and this embodiment is performed in a space surrounded by the wall surface of the hydraulic piston member 57a, the back surface of the movable sheave 53, the outer peripheral surface of the first extending portion 53a, and the inner peripheral surface of the second extending portion 53b. The hydraulic chamber 57 of Example 2 is formed.

As shown in FIG. 10, the hydraulic piston member 57a of the second embodiment includes a first annular portion 57a ₁ disposed at a distance from the back surface of the movable sheave 53, and an outer diameter of the first annular portion 57a ₁ . The first cylindrical portion 57a ₂ extending in the same direction as the first extending portion 53a from the side, and the cylindrical portion extending in the same direction as the first extending portion 53a from the inner diameter side of the first annular portion 57a ₁ 57a ₃ and a second annular portion 57a ₄ that closes an annular opening between the primary shaft 51 at the extended end of the tubular portion 57a ₃ and the second annular portion 57a _{4 from} the primary shaft 51 side. 1 consists of extending portion 53a and the second cylindrical portion 57a ₅ Metropolitan that extend in the same direction.

In the second embodiment, the inner peripheral surface of the second cylindrical portion 57a _{5 and} the outer peripheral surface of the primary shaft 51 are spline-fitted via the spline 54D. Therefore, the hydraulic piston member 57a The shaft 51, the fixed sheave 52 and the movable sheave 53 can be rotated together.

The second cylindrical portion 57a ₅ has a larger outer diameter on the second annular portion 57a ₄ side, and a bearing 551e of the first power transmission member 551a and a clutch hydraulic piston member 58c ₁ which will be described later at this portion. It is formed so as to hold the elastic member 58c _4.

  Here, as in the first embodiment, the hydraulic chamber 57 of the second embodiment includes an oil passage 53c formed in the first extending portion 53a, an oil passage 51c formed in the primary shaft 51, and a clamping pressure regulating valve. It communicates with 56B. Therefore, by controlling the clamping pressure regulating valve 56 </ b> B to raise the hydraulic pressure of the hydraulic chamber 57, the movable sheave thrust force that brings the movable sheave 53 closer to the fixed sheave 52, or between the fixed sheave 52 and the movable sheave 53. The belt clamping pressure is generated.

  In the second embodiment, the movable sheave thrust and the belt clamping pressure are not directly transmitted from the back surface of the movable sheave 53 in the hydraulic chamber 57 to the movable sheave 53, but the power transmission means 551 similar to that in the first embodiment is provided. Communicate through.

  As shown in FIG. 10, the power transmission means 551 of the second embodiment includes a first power transmission member 551a disposed so as to be able to rotate relative to the primary shaft 51, and an inner periphery of the second extending portion 53b. It is comprised by the 2nd power transmission member 551b similar to Example 1 provided integrally in the surface.

Here, the first power transmission member 551a of the second embodiment is formed of a substantially annular member with the primary shaft 51 as the central axis, and the axial position relative to the primary shaft 51 is constant on the inner peripheral surface thereof. While being arranged via the bearing 551e so that relative rotation can be performed as it is, an outer screw portion similar to that of the first embodiment is formed on the outer peripheral surface of the second power transmission member 551b. . For example, in the present embodiment 2 is held via a bearing 551e in the second cylindrical portion 57a ₅ of the first power transmission member 551a of the hydraulic piston member 57a.

  Here, when the movable sheave 53 is brought close to the fixed sheave 52 (in the case of an upshift), the hydraulic pressure in the hydraulic chamber 57 is increased. At this time, the axial pressing force by the hydraulic pressure is increased by the second extending portion. It hooks on the internal thread portion of the second power transmission member 551b via 53b. On the other hand, in the reverse case (downshift), the hydraulic pressure in the hydraulic chamber 57 is reduced. At this time, the reaction force of the belt clamping pressure in the axial direction is applied to the movable sheave 53, and the reaction force is extended to the second extension. It is transmitted to the inner thread portion of the second power transmission member 551b via the portion 53b.

  Therefore, in the second embodiment, in order to operate the power transmission means 551 in both cases, the axial force applied to the outer screw portion and the inner screw portion (reaction force of the pressing force or the belt clamping pressure). The first power transmission member 551a and the second power transmission member 551b are rotated relative to each other. For example, when the axial force is applied to the outer screw portion and the inner screw portion, the lead of the outer screw portion and the inner screw portion so that the first power transmission member 551a and the second power transmission member 551b can rotate relative to each other. Increase the corner.

By the way, also in the second embodiment, the first cylindrical portion 57a ₂ has the outer peripheral surface abutting on the inner peripheral surface of the second extending portion 53b or the inner peripheral surface, as in the first embodiment. The outer diameter of the first power transmission member 551a in the power transmission means 551 is smaller than the inner diameter of the screw thread. As a result, as in the first embodiment, the hydraulic oil in the hydraulic chamber 57 that has oozed out of the annular seal member 57b is supplied to the outer screw portion and the inner screw portion of the power transmission means 551 by the centrifugal force accompanying the rotation of the primary shaft 51. Therefore, the lubricating performance is improved and the durability is improved without providing a lubricating oil supply mechanism or a lubricating oil supply path to the power transmission means 551 separately.

  Further, even in the second embodiment, the primary shaft 51 side can allow relative movement between the first power transmission member 551a and the second power transmission member 551b in the power transmission means 551, while between them. A clutch mechanism 58 similar to that of the first embodiment that can stop the relative movement is provided.

As shown in FIG. 10, the clutch mechanism 58 of the second embodiment includes a first clutch engaging portion 58a provided integrally with a wall surface of the first power transmission member 551a and a first piston provided on the hydraulic piston member 57a. A two-clutch engaging portion 58b and a clutch operating portion 58c for engaging and disengaging the first and second clutch engaging portions 58a and 58b. It is disposed (enclosed annular space by the first cylindrical portion 57a ₂ and the cylindrical portion 57a ₃₎ in the space of the member 57a.

On the other hand, also in the second embodiment, an oil passage 57a ₆ communicating with the clutch hydraulic chamber 58c ₂ of the clutch operating portion 58c is formed in the hydraulic piston member 57a, and an oil passage communicating with the oil passage 57a ₆ 51d is formed in the primary shaft 51, and further, there is provided a clutch hydraulic pressure switching valve 56D communicating with the oil passage 51d and communicating with the regulator valve 56C via the oil passage 56e shown in FIG.

For this reason, when the electronic control unit C opens the clutch hydraulic pressure switching valve 56D, the respective clutch engaging members 58a ₂ and 58b _{1 of} the first and second clutch engaging portions 58a and 58b are engaged, and the clutch hydraulic pressure is determined. By closing the switching valve 56D, the elastic member 58c ₄ releases the respective clutch engaging members 58a ₂ and 58b ₁ .

Furthermore, even in the second embodiment, the sliding portions (first and second clutch engaging portion 58a, 58b and a hydraulic piston member 58c ₁ like sliding portion of the clutch) of the clutch mechanism 58 lubricating oil to A lubricating oil supply path for supplying is provided.

In the second embodiment, as shown in FIG. 10, an oil passage 57a ₇ communicating with the oil passage 51f of the primary shaft 51 is connected to the bearing 551e portion holding the first power transmission member 551a, and the hydraulic piston member. formed in the second cylindrical portion 57a ₅ of 57a, supplies lubricating oil from the gap of the bearing 551e to sliding portions of the clutch mechanism 58.

Here, the supply of the lubricating oil to the oil passage 57a ₇ is performed by separately providing a switching valve communicating with the oil passage 51f of the primary shaft 51 and communicating with the regulator valve 56C, and opening and closing the switching valve.

Thus configured according to the second embodiment the clutch mechanism 58, from the thread inside diameter of the outermost diameter portion (inner peripheral surface side of the first cylindrical portion 57a ₂₎ is externally threaded portion of the first power transmission member 551a Is also arranged on the inner diameter side. As a result, the lubricating oil supplied to the sliding portion of the clutch mechanism 58 is supplied to the outer screw portion and the inner screw portion of the power transmission means 551 by the centrifugal force accompanying the rotation of the primary shaft 51. For this reason, the lubricating performance of the power transmission means 551 can be improved in combination with the hydraulic oil in the hydraulic chamber 57 that has oozed out from the seal member 57b, and the durability can be further improved.

Further, since the clutch mechanism 58 is arranged in the space of the hydraulic piston member 57a (the annular space surrounded by the first cylindrical portion 57a ₂ and the cylindrical portion 57a ₃ ), the clutch mechanism The increase in size of the belt type continuously variable transmission 1 due to the provision of 58 can be suppressed.

  Here, the operation of the belt type continuously variable transmission 1 of the second embodiment will be described in detail with reference to the flowchart of FIG.

  First, the electronic control unit C determines whether or not the actual transmission gear ratio is the target transmission gear ratio (that is, the actual transmission gear ratio = target transmission gear ratio) as in the case of the first embodiment (step ST21). If the gear ratio has not reached the target gear ratio, it is determined whether the actual gear ratio is larger than the target gear ratio (whether the actual gear ratio is larger than the target gear ratio) (step ST22).

  If the actual gear ratio is larger than the target gear ratio, the electronic control unit C releases the clutch mechanism 58 (step ST23), controls the clamping pressure regulating valve 56B, and controls the hydraulic chamber 57 of the primary pulley 50. Increase hydraulic pressure (step ST24)

  Here, the hydraulic pressure is increased to a predetermined hydraulic pressure that can be changed from the actual gear ratio to the target gear ratio. For example, the electronic control unit C corresponds to the correspondence between the hydraulic pressure applied to the hydraulic chamber 57 and the amount of movement of the movable sheave 53 in the axial direction or the position of the movable sheave 53 in the axial direction, and the position and speed change of the movable sheave 53 in the axial direction. Map data indicating a correspondence relationship with the ratio is prepared in advance, and the hydraulic pressure of the clamping pressure regulating valve 56B is set based on each map data.

  As a result, the axial force accompanying the increase in the hydraulic pressure in the hydraulic chamber 57 is transmitted to the power transmission means 551, so that the movable sheave 53 of the primary pulley 50 approaches the fixed sheave 52 and the secondary pulley 60 moves accordingly. The sheave 63 is separated from the fixed sheave 62.

  As described above, the electronic control unit C reduces the gear ratio and performs speed increase control (upshift control).

  Subsequently, the electronic control unit C returns to the determination process of step ST21 and executes the above-described process or a process described later according to the determination result.

  For example, if the actual gear ratio has reached the target gear ratio in step ST21, the electronic control unit C engages the clutch mechanism 58 as in the case of the first embodiment (step ST25).

  As a result, the relative rotation between the first power transmission member 551a and the second power transmission member 551b is stopped, and the power transmission means 551 cannot be operated. The valve 56B is controlled to reduce the hydraulic pressure applied to the hydraulic chambers 57, 66 to a predetermined hydraulic pressure similar to that in the first embodiment (step ST26). Thereby, the power of the oil pump OP can be reduced and the efficiency of the transmission can be improved.

  In the processing operation described above, the clutch mechanism 58 is released at one end step ST23, and then the hydraulic pressure in the hydraulic chamber 57 is increased at step ST24 to perform an upshift. However, when the clutch mechanism 58 is first released in this manner, the power transmission means 551 is activated by the reaction force of the belt clamping pressure applied to the outer screw portion and the inner screw portion, and the movable sheave 53 moves in the separation direction. There is a possibility of sliding.

  Therefore, when such inconvenience is a concern, the hydraulic pressure of the one-end hydraulic chamber 57 is increased to apply a pressing force toward the fixed sheave 52 to the movable sheave 53, and then the clutch mechanism 58 is released. Thus, it is preferable to operate the power transmission means 551.

  On the other hand, if the actual gear ratio is smaller than the target gear ratio in step ST22 described above, the electronic control unit C releases the clutch mechanism 58 (step ST27) and controls the clamping pressure regulating valve 56B to control the primary pulley. The hydraulic pressure in the 50 hydraulic chambers 57 is reduced (step ST28).

  Here, based on each map data mentioned above, it reduces to the preset oil pressure which can be changed from an actual gear ratio to a target gear ratio.

  As a result, the reaction force of the belt clamping pressure is transmitted to the power transmission means 551, so that the movable sheave 53 of the primary pulley 50 is separated from the fixed sheave 52, and accordingly, the movable sheave 63 of the secondary pulley 60 is fixed to the fixed sheave 62. To approach.

  The electronic control unit C increases the gear ratio as described above and performs deceleration control (downshift control).

  Then, the electronic control unit C returns to the determination process of step ST21 and executes the process described above according to the determination result.

As described above, according to the belt-type continuously variable transmission 1 of the second embodiment, the movable sheave sliding mechanism 55 including the hydraulic chamber 57 and the power transmission means 551 is provided on the rear surface of the movable sheave 53 as in the first embodiment. In the internal space of the hydraulic piston member 57a constituting the hydraulic chamber 57 (the annular space surrounded by the first cylindrical portion 57a ₂ and the cylindrical portion 57a ₃ ). Since the mechanism 58 is disposed, the belt type continuously variable transmission 1 itself can be reduced in size.

  Further, in the belt type continuously variable transmission 1 of the second embodiment, the movable sheave with respect to the movable sheave 53 with respect to the movable sheave 53 by the hydraulic chamber 57 and the power transmission means 551 without using the vane type hydraulic motor 550 as in the first embodiment. Since thrust is generated, it is possible to reduce drive loss by reducing hydraulic oil and hydraulic pressure.

  Next, Embodiment 3 of the belt type continuously variable transmission according to the present invention will be described with reference to FIG.

In the belt type continuously variable transmission 1 of the third embodiment, the first power transmission member 551a held in the second cylindrical portion 57a ₅ of the hydraulic piston member 57a in the second embodiment is directly connected as shown in FIG. The primary shaft 51 is held via a similar bearing 551e, and the other configuration is the same as that of the belt-type continuously variable transmission 1 of the second embodiment except for the following changes.

Here, in order to shorten the axial direction in the belt-type continuously variable transmission 1 by directly holding the first power transmission member 551a on the primary shaft 51, the second of the hydraulic piston member 57a is used. The cylindrical portion 57a ₅ needs to be shortened in the axial direction with respect to the second embodiment.

Therefore, in the third embodiment, the axial length of the second cylindrical portion 57a ₅ is shortened relative to the second embodiment, and the outer diameter and the inner diameter are increased, so that the second cylindrical portion 57a ₅ to extend from the intermediate portion of the second annular portion 57a _4.

As a result, an increase in the size of the belt-type continuously variable transmission 1 is suppressed, and a belt-type continuously variable transmission is provided by disposing a bearing 551e in the internal space of the second cylindrical portion 57a ₅ as shown in FIG. 1 can be shortened in the axial direction.

On the other hand, with the shape change of the second cylindrical portion 57a _5, spline 54E between the inner and outer circumferential surfaces of the first extending portion 53a of the movable sheave 53 of the cylindrical portion 57a ₃ of the hydraulic piston member 57a With this spline 54E, the hydraulic piston member 57a is spline-fitted to the movable sheave 53 with the axial position kept constant. Thereby, the hydraulic piston member 57a of the third embodiment rotates integrally with the primary shaft 51, the fixed sheave 52, and the movable sheave 53 as in the second embodiment.

Further, the first power transmission member 551a is directly held by the primary shaft 51, whereby the oil passage 57a of the hydraulic piston member 57a constituting the lubricating oil supply passage to the sliding portion of the clutch mechanism 58 in the second embodiment. ₇ is abolished and lubricating oil is directly supplied from the oil passage 51f of the primary shaft 51 to the bearing 551e.

  The belt-type continuously variable transmission 1 of the third embodiment configured as described above operates in the same manner as in the second embodiment and has the same effects.

  Here, the clutch mechanism 58 of each of the first to third embodiments described above is a so-called multi-plate clutch using hydraulic pressure as a drive source. However, the clutch mechanism is not necessarily limited to this, for example, an electromagnetic clutch or the like. It may be.

  In each of the first to third embodiments, the movable sheave sliding mechanism 55 and the clutch mechanism 58 are provided on the movable sheave 53 on the primary pulley 50 side, but the present invention is not necessarily limited thereto. For example, the movable sheave sliding mechanism 55 and the clutch mechanism 58 may be provided for the movable sheave 63 on the secondary pulley 60 side, or provided on the movable sheaves 53 and 63 of both the primary pulley 50 and the secondary pulley 60. May be.

  In addition, you may provide the buffer mechanism 69 shown in FIG.13 and FIG.14 in the secondary pulley 60 of each Examples 1-3 mentioned above.

  The buffer mechanism 69 includes a donut-shaped outer case 691 disposed on the circular member 67 and a plate-shaped member 692 erected on the torque cam main body 65c. The outer case 691 has two hollow portions 691 a filled with a viscous fluid (for example, hydraulic oil), and rotates integrally with the circular member 67. The plate-like member 692 has a through hole (orifice) 692a formed on the surface thereof, and rotates integrally with the torque cam main body 65c.

  Here, a plate-like member 692 is disposed in each of the hollow portions 691a. When the outer case 691 and the plate-like member 692 rotate relative to each other, the plate-like member 692 moves in the hollow portion 691a. To do. A gap is provided between the end of the plate-like member 692 and the inner wall surface of the hollow portion 691a.

  As a result, the torque cam 65 is actuated when the speed ratio is changed, whereby the plate-like member 692 moves in the hollow portion 691a. At that time, resistance is generated by the viscous fluid flowing through the orifice 692a and the gap, and the relative movement between the torque cam main body 65c and the movable sheave 63 can be performed gently. For this reason, it is possible to reduce the shock when the backlash of the torque cam 65 is clogged when the gear ratio is changed (when the torque cam 65 is driven / non-driven).

  The magnitude of the resistance is adjusted by the gap between the end of the plate-like member 692 and the inner wall surface of the hollow portion 691a and the diameter of the orifice 692a.

  Further, in this buffer mechanism 69, the middle portion of the hollow portion 691a shown in FIG. 14 is made wider than both end portions thereof so that the degree of buffering (buffer force) can be changed according to the gear ratio. Also good. That is, the gap between the end portion of the plate-like member 692 and the inner wall surface of the hollow portion 691a is large when the plate-like member 692 is located in the middle portion of the hollow portion 691a, and the plate-like member 692 has a hollow portion 691a. A hollow portion 691a having a width changed in the circumferential direction is formed so as to become smaller as it approaches the both end portions.

  As a result, the moving speed of the plate-like member 692 is high when the plate-like member 692 is located in the middle portion of the hollow portion 691a, and becomes slower as the plate-like member 692 approaches both end portions of the hollow portion 691a. Accordingly, the degree of buffering (buffering force) can be changed according to the above, and the shock when the backlash of the torque cam 65 is clogged can be reduced. For example, drivability can be improved by setting the gap so that the buffering force is increased during downshifting.

  Here, since the movable sheave 63 is attached to the secondary shaft 61 via the spline 64A, the movable sheave 63 and the fixed sheave 62 have the same rotational direction and rotational speed. Therefore, the buffer mechanism 69 is not limited to the position between the movable sheave 63 and the torque cam 65 as in the second embodiment, but may be provided on the fixed sheave 62 side. In this case, for example, the buffer mechanism 69 is provided with a rotating member (not shown) that rotates in the same manner as the torque cam main body 65c on the opposite side of the groove 80b in the fixed sheave 62, and the plate member 692 is attached to the rotating member. The outer sheave 691 may be attached to the fixed sheave 62. The rotating member may be independent of the torque cam 65, or may be extended from the torque cam main body 65c, for example.

  As described above, the belt-type continuously variable transmission according to the present invention is useful for those including a plurality of actuators that function as a sliding mechanism and a pressing mechanism of a movable sheave, and in particular, independently controls each actuator. Therefore, it is suitable for a technique for improving the shift controllability and reducing the drive loss.

1 is a skeleton diagram showing an overall configuration of a power transmission device including a belt-type continuously variable transmission according to the present invention. It is a figure which shows the structure of Example 1 by the side of the primary pulley in the belt-type continuously variable transmission which concerns on this invention. It is sectional drawing of the hydraulic motor seen from the XX line | wire shown in FIG. It is explanatory drawing explaining the hydraulic circuit structure in the belt type continuously variable transmission of Example 1. FIG. It is explanatory drawing explaining operation | movement of the switching valve for gear ratio control of Example 1, Comprising: It is a figure which shows the valve position in the case of supplying hydraulic pressure to the 1st oil chamber. It is explanatory drawing explaining operation | movement of the switching valve for gear ratio control of Example 1, Comprising: It is a figure which shows the valve position in the case of supplying hydraulic pressure to the 1st and 2nd oil chamber. It is explanatory drawing explaining operation | movement of the switching valve for gear ratio control of Example 1, Comprising: It is a figure which shows the valve position in the case of supplying hydraulic pressure to a 2nd oil chamber. It is a figure which shows the structure of Example 1 by the side of the primary pulley in the belt-type continuously variable transmission which concerns on this invention, Comprising: It is an enlarged view explaining a clutch mechanism. It is explanatory drawing explaining the structure by the side of the secondary pulley in the belt-type continuously variable transmission of Example 1. FIG. It is explanatory drawing explaining the torque cam of Example 1, Comprising: It is the figure which illustrated the case where the fixed sheave and the movable sheave of a secondary pulley are in the separated state. It is explanatory drawing explaining the torque cam of Example 1, Comprising: It is the figure which illustrated the case where the fixed sheave of a secondary pulley and the movable sheave have approached. 3 is a flowchart illustrating the operation of the belt type continuously variable transmission according to the first embodiment. It is a figure which shows the structure of Example 2 by the side of the primary pulley in the belt-type continuously variable transmission which concerns on this invention. 6 is a flowchart illustrating the operation of the belt type continuously variable transmission according to the second embodiment. It is a figure which shows the structure of Example 3 by the side of the primary pulley in the belt-type continuously variable transmission which concerns on this invention. It is other examples of a secondary pulley, Comprising: It is explanatory drawing explaining a buffer mechanism. It is sectional drawing of the buffer mechanism seen from the YY line shown in FIG.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Belt type continuously variable transmission 50 Primary pulley 51 Primary shaft 52 Fixed sheave 53 Movable sheave 53b 2nd extension part 54C Spline 55 Movable sheave sliding mechanism 57 Hydraulic chamber 57a Hydraulic piston member 58 Clutch mechanism 58a 1st clutch engaging part 58b Second clutch engaging portion 58c Clutch operating portion 58c ₂ clutch hydraulic chamber 58c ₄ elastic member 60 secondary pulley 61 secondary shaft 62 fixed sheave 63 movable sheave 80 belt 80a, 80b V-shaped groove 550 Hydraulic motor (vane hydraulic) motor)
550a Motor shaft 550a ₃ groove 550a ₄ oil passage 550b motor case 550g first oil chamber 550h second oil chamber 551 power transmission means 551a first power transmission member 551b second power transmission member C electronic control unit

Claims (3)

Two pulley shafts arranged in parallel at a predetermined interval, a movable sheave arranged on each pulley shaft and slidable on the pulley shaft in the axial direction, and opposed to each movable sheave. A fixed sheave that is arranged on the pulley shaft and forms a groove with the movable sheave, and a belt that is wound around each groove in the movable sheave and the fixed sheave arranged opposite to each other. In a step transmission,
The internal space formed in the movable sheave includes a first power transmission member that can rotate relative to the pulley shaft with the pulley shaft as a central axis, and a second power transmission member that rotates integrally with the pulley shaft. While enabling the relative movement between the power transmission means that allows the movable sheave to approach or separate from the fixed sheave and the first power transmission member and the second power transmission member, the relative movement therebetween is stopped. A clutch mechanism that can perform, and a hydraulic chamber that can press the movable sheave toward the fixed sheave,
A belt type continuously variable transmission, wherein the clutch mechanism is arranged in a space formed by a hydraulic piston member constituting a wall surface of the hydraulic chamber. While providing an outer screw portion on the outer peripheral surface of the first power transmission member, an inner screw portion that is screwed with the outer screw portion is provided on the inner peripheral surface of the second power transmission member,
The belt-type continuously variable transmission according to claim 1, wherein a sliding surface between the movable sheave and the hydraulic piston member is provided on an inner diameter side with respect to an outer screw portion of the first power transmission member. A motor case capable of rotating relative to the pulley shaft about the pulley shaft as a central axis and a hydraulic motor serving as a drive source in the axial direction of the movable sheave including a motor shaft rotating integrally with the pulley shaft are provided. The first power transmission member is integrally provided on the outer peripheral surface of the motor case, or the first power transmission member is constituted by the motor case.
The motor shaft and the hydraulic piston member are spline-fitted at the inner diameter portion of the motor shaft, and a lubricating oil supply path for supplying the lubricating oil to the clutch mechanism from the gap of the spline groove in the spline fitting portion is provided. The belt-type continuously variable transmission according to claim 1, wherein the belt-type continuously variable transmission is provided.

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