JP2008101736A - Belt type continuously variable transmission - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To inhibit the increase of drive loss of an oil pump by preventing a working fluid supply discharge valve from opening when hydraulic pressure in a clamping force generation hydraulic pressure chamber is kept constant. <P>SOLUTION: The belt type continuously variable transmission is provided with a primary pulley 50 and a secondary pulley 60 rotating in relation to a static member, a belt 110 wound around the same, a primary hydraulic pressure chamber 55 generating belt clamping force at the primary pulley 50, a working fluid supply discharge valve 70 opening when working fluid is supplied to the primary hydraulic chamber 55 and when the working fluid is discharged from the primary hydraulic chamber 55, and an actuator 80 forcibly opening the working fluid supply discharge valve 70 by sliding a piston 82 from an initial position by hydraulic pressure in a drive hydraulic pressure chamber 81. The actuator 80 is provided with a piston initial position retention means (a centrifugal hydraulic pressure cancel chamber 84, a centrifugal hydraulic pressure cancel member 85) retaining the piston 82 at the initial position by mechanical force with resisting force in valve opening direction acting on the piston 82 by centrifugal hydraulic pressure in the drive hydraulic pressure chamber 81 accompanying rotation of the primary pulley 50. <P>COPYRIGHT: (C)2008,JPO&INPIT

Description

  The present invention relates to a belt type continuously variable transmission.

  In general, in order to transmit a driving force (that is, output torque) output from an internal combustion engine or an electric motor, which is a driving source, to a road surface under an optimum condition according to the traveling state of the vehicle, A transmission is provided on the output side. In general, the transmission is roughly classified into a manual transmission and an automatic transmission. Further, the automatic transmission further includes a continuously variable transmission that variably controls the transmission ratio continuously (continuously), and a transmission ratio. It is distinguished from a stepped transmission that is variably controlled stepwise (discontinuously).

  Here, the continuously variable transmission includes two pulleys (a primary pulley to which the output torque from the drive source is transmitted and a secondary pulley that changes and outputs the output torque transmitted to the primary pulley as main components. ) And a belt that is wound around each of these V-shaped grooves and transmits the output torque transmitted to the primary pulley to the secondary pulley, is known. The primary pulley and the secondary pulley have the same main components, and each pulley shaft (primary pulley shaft, secondary pulley shaft) as a rotating shaft arranged in parallel and its own pulley shaft. Fixed sheave fixed on the top (primary fixed sheave, secondary fixed sheave) and movable on the own pulley shaft facing the fixed sheave in the axial direction and sliding on the pulley shaft in the axial direction A sheave (primary movable sheave, secondary movable sheave) and a clamping pressure generating hydraulic chamber (primary hydraulic chamber, secondary hydraulic chamber) for generating a belt clamping pressure with respect to the belt are configured. For example, this type of belt-type continuously variable transmission is disclosed in Patent Document 1 below.

  In these primary pulleys and secondary pulleys, V-shaped grooves are formed between the fixed sheave and the movable sheave, respectively, and belts are wound around the respective V-shaped grooves. And in these primary pulleys and secondary pulleys, each movable sheave slides in the axial direction on the pulley shaft by adjusting the hydraulic pressure of the respective clamping pressure generating hydraulic chambers, and the respective V-shaped The width of the groove changes. In this belt type continuously variable transmission, the width of each V-shaped groove is variably controlled so that the contact radius with the belt in each V-shaped groove of the primary pulley and the secondary pulley is continuously variable. The transmission ratio is variably controlled steplessly. That is, this belt type continuously variable transmission changes the output torque from the drive source continuously. Further, each of the clamping pressure generating hydraulic chambers presses the movable sheave toward the fixed sheave by the hydraulic pressure to generate belt clamping pressure on the belt.

JP 2001-323978 A

  By the way, in the conventional belt type continuously variable transmission described above, there is a case where the movement of the movable sheave in the axial direction is restricted and the position in the axial direction of the movable sheave with respect to the fixed sheave is kept constant to fix the transmission ratio. . Therefore, in this conventional belt type continuously variable transmission, the hydraulic pressure in the clamping pressure generating hydraulic chamber needs to be maintained at a predetermined hydraulic pressure in order to make the position of the movable sheave relative to the fixed sheave in the axial direction constant. It is necessary to continue supplying hydraulic oil to the clamping pressure generating hydraulic chamber not only when changing the ratio but also when fixing the transmission ratio.

  However, in the conventional belt type continuously variable transmission, in order to keep supplying hydraulic oil to the clamping pressure generating hydraulic chamber at all times, the oil pump of the hydraulic oil supply control device that supplies the hydraulic oil must be continuously driven. In other words, the drive loss of the oil pump increases.

  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 suppress an increase in driving loss of the oil pump.

  In order to achieve the above object, according to the first aspect of the present invention, the two pulleys that rotate with respect to the stationary member and the drive force from the drive source that is wound around each pulley and transmitted to one pulley are obtained. A belt that transmits to the other pulley, a clamping pressure generating hydraulic chamber that is formed in each pulley and generates a belt clamping pressure with respect to the belt by hydraulic pressure, and at least one of these pulleys is integrated with the pulley. A hydraulic oil supply valve that is provided to rotate and opens when hydraulic oil is supplied to the clamping pressure generating hydraulic chamber of the pulley, and the hydraulic oil is discharged from the clamping pressure generating hydraulic chamber to which the hydraulic oil is supplied. The hydraulic oil discharge valve that opens at the same time as the pulley and the hydraulic pressure in the drive hydraulic chamber causes the piston to slide from the initial position to one of the sliding directions relative to the drive hydraulic chamber. Supply And a valve-type continuously variable transmission that forcibly opens the hydraulic oil discharge valve, the valve-opening / closing means is an opening that acts on the piston by the centrifugal hydraulic pressure of the drive hydraulic chamber as the pulley rotates. Piston initial position holding means for holding the piston in the initial position with mechanical force against the force in the valve direction is provided.

  For example, the piston initial position holding means presses the piston in the valve closing direction of the hydraulic oil supply valve and the valve closing direction of the hydraulic oil discharge valve by centrifugal force applied when the pulley rotates, as in the second aspect of the invention. A centrifugal hydraulic canceling member is provided.

  In the belt type continuously variable transmission according to the first and second aspects, when the valve opening / closing means does not need to be operated, the piston is held at the initial position by the piston initial position holding means. The operation is suppressed, and the opening of the hydraulic oil supply valve and the hydraulic oil discharge valve can be prevented.

  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 second aspect, the valve body of the hydraulic oil supply valve and the piston associated with the hydraulic oil supply valve are provided. The piston hydraulic pressure canceling member of the piston initial position holding means, the valve body of the hydraulic oil discharge valve, and the centrifugal hydraulic pressure canceling member of the piston initial position holding means for the piston related to the hydraulic oil discharge valve are respectively integrated.

  In order to achieve the above object, in the invention according to claim 4, in the belt-type continuously variable transmission according to claim 2, the hydraulic oil supply valve and the hydraulic oil discharge valve are provided with a common valve body. The valve body and the centrifugal hydraulic pressure canceling member of the piston initial position holding means are integrated.

  In the belt type continuously variable transmission according to claims 3 and 4, since the function of the piston initial position holding means can be used in combination with the hydraulic oil supply valve and the hydraulic oil discharge valve, the number of parts can be reduced and the size can be reduced. It becomes possible.

  In order to achieve the above object, in the invention according to claim 5, in the belt type continuously variable transmission according to claim 2, 3 or 4, a space is provided on the outer diameter side of the piston and the centrifugal hydraulic pressure canceling member. The piston initial position holding means is configured to be provided.

  In the belt type continuously variable transmission according to the fifth aspect, since the hydraulic oil can be stored in the space, the hydraulic oil can be used as a lubricating oil for a sliding portion of the centrifugal hydraulic pressure canceling member with the piston or the like. . In addition, since foreign matter can be collected in the space, it is possible to prevent the foreign matter from being caught in the sliding portion.

  In order to achieve the above object, according to a sixth aspect of the present invention, in the belt type continuously variable transmission according to the first, second, third, fourth, or fifth aspect, the piston is moved to the initial position using the hydraulic force. The piston initial position holding means is configured so as to be held in the cylinder.

  In the belt type continuously variable transmission according to the sixth aspect, the viscosity of the hydraulic oil is low, such as when cold, and the force against the centrifugal hydraulic pressure cannot be generated only by the centrifugal hydraulic pressure canceling member. However, a force against the centrifugal hydraulic pressure can be generated using the hydraulic pressure.

  In the belt-type continuously variable transmission according to the present invention, when it is not necessary to operate the valve opening / closing means (for example, when it is desired to keep the hydraulic pressure in the clamping pressure generating hydraulic chamber constant), the hydraulic oil supply valve and the hydraulic oil discharge valve Therefore, the hydraulic oil is not discharged from the clamping pressure generating hydraulic chamber at that time, and it is not necessary to supply the hydraulic oil for keeping the hydraulic pressure constant in the clamping pressure generating hydraulic chamber. For this reason, according to this belt type continuously variable transmission, there is no need to drive the oil pump in order to keep the hydraulic pressure in the clamping pressure generating hydraulic chamber constant, so that it is possible to suppress an increase in driving loss of the oil pump. it can. In addition, the piston initial position holding means that uses mechanical force without using hydraulic force does not need to drive the oil pump to generate force against centrifugal hydraulic pressure. An increase in driving loss of the pump can be suppressed.

  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. Here, in the following embodiments, an internal combustion engine (gasoline engine, diesel engine, LPG engine, etc.) is exemplified as a drive source that generates a driving force transmitted to the belt type continuously variable transmission. However, the present invention is not necessarily limited to this. In the following embodiments, one pulley is described as a primary pulley and the other pulley is described as a secondary pulley. However, one pulley may be a secondary pulley and the other pulley may be a primary pulley.

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

  Here, the belt-type continuously variable transmission 1A of the first embodiment will be described together with the overall configuration of the power transmission device.

  As shown in FIG. 1, the power transmission device includes an internal combustion engine 10 as a drive source and a transaxle 20 as a stationary component 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, and a housing is constituted by these.

  A torque converter (starting device) 30 is accommodated in the transaxle housing 21. The torque converter 30 increases or transmits the driving force (that is, the output torque of the internal combustion engine 10) from a driving source to a belt-type continuously variable transmission 1A described later, and includes a pump (pump impeller) 31 and a turbine. (Turbine runner) 32, stator 33, lock-up clutch 34, and damper device 35 are provided at least.

  First, the pump 31 is attached to a hollow shaft 36 that can rotate around the same axis as the crankshaft 11 of the internal combustion engine 10. That is, the pump 31 can rotate about the same axis as the crankshaft 11 together with the hollow shaft 36. The pump 31 is connected to the front cover 37. The front cover 37 is connected to the crankshaft 11 via the drive plate 12 of the internal combustion engine 10.

  Further, the turbine runner 32 is arranged so as to face the pump 31. The turbine runner 32 is disposed inside the hollow shaft 36 and is attached to an input shaft 38 that can rotate around the same axis as the crankshaft 11. That is, the turbine runner 32 can rotate around the same axis as the crankshaft 11 together with the input shaft 38.

  A stator 33 is disposed between the pump 31 and the turbine runner 32 via a one-way clutch 39. The one-way clutch 39 is fixed to the transaxle housing 21. A lockup clutch 34 is disposed between the turbine runner 32 and the front cover 37, and the lockup clutch 34 is connected to an input shaft 38 via a damper device 35. The casing formed by the pump 31 and the front cover 37 is a hydraulic oil supply part, and the hydraulic oil is supplied as the hydraulic fluid from the hydraulic oil supply control device 130 that supplies the hydraulic oil to the hydraulic oil supply part. .

  Here, the operation of the torque converter 30 will be described.

  The output torque of the internal combustion engine 10 is transmitted from the crankshaft 11 to the front cover 37 via the drive plate 12. At this time, when the lockup clutch 34 is released by the damper device 35, the output torque from the internal combustion engine 10 transmitted to the front cover 37 is transmitted to the pump 31, and the pump 31 and the turbine runner 32 are connected to each other. It is transmitted to the turbine runner 32 via hydraulic fluid circulating between them. The output torque from the internal combustion engine 10 transmitted to the turbine runner 32 is transmitted to the input shaft 38. That is, the torque converter 30 increases the output torque from the internal combustion engine 10 via the input shaft 38 and transmits it to the belt type continuously variable transmission 1A. In the above, the stator 33 can change the flow of hydraulic fluid circulating between the pump 31 and the turbine runner 32 to obtain a predetermined torque characteristic.

  On the other hand, when the lock-up clutch 34 is locked (engaged with the front cover 37) by the damper device 35, the output torque from the internal combustion engine 10 transmitted to the front cover 37 is directly not via hydraulic oil. It is transmitted to the input shaft 38. That is, the torque converter 30 transmits the output torque from the internal combustion engine 10 via the input shaft 38 as it is to the belt type continuously variable transmission 1A.

  Next, in the transaxle case 22 and the transaxle rear cover 23, a forward / reverse switching mechanism 40, a belt-type continuously variable transmission 1A, and a final reduction gear 90 as a differential device are housed.

  First, the forward / reverse switching mechanism 40 will be described.

  As shown in FIG. 1, the forward / reverse switching mechanism 40 transmits the output torque from 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 1A described later. And includes at least a planetary gear unit 41, a forward clutch 42, and a reverse brake 43.

  The planetary gear device 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 a primary pulley shaft 51 that is a rotating shaft of the primary pulley 50. Therefore, the torque transmitted to the sun gear 44 is transmitted to the primary pulley 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 held by a switching carrier 47 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 switching carrier 47 is connected to the reverse brake 43 at its outer peripheral end.

  The ring gear 46 is meshed with each pinion 45 held by the switching carrier 47 and is connected to the input shaft 38 of the torque converter 30 via the forward clutch 42.

  Subsequently, the forward clutch 42 constituting the forward / reverse switching mechanism 40 is controlled by ON / OFF control by supplying hydraulic oil from the hydraulic oil supply control device 130 to the hollow portion of the input shaft 38 which is a hydraulic oil supply portion. It is what is done. When the forward clutch 42 is OFF, the output torque from the internal combustion engine 10 transmitted to the input shaft 38 is transmitted to the ring gear 46. On the other hand, when the forward clutch 42 is ON, the output torque from the internal combustion engine 10 transmitted to the input shaft 38 is directly transmitted to the sun gear 44 without the ring gear 46, the sun gear 44, and the pinions 45 rotating relative to each other. .

  Further, the reverse brake 43 constituting the forward / reverse switching mechanism 40 is ON / OFF controlled by supplying hydraulic oil from a hydraulic oil supply control device 130 to a brake piston (not shown) which is a hydraulic oil supply portion. It is. When the reverse brake 43 is ON, the switching carrier 47 is fixed to the transaxle case 22 so that each pinion 45 cannot revolve around the sun gear 44. On the other hand, when the reverse brake 43 is OFF, the switching carrier 47 is released, and each pinion 45 can revolve around the sun gear 44.

  Next, the configuration of the belt type continuously variable transmission 1A according to the first embodiment will be described.

  This belt-type continuously variable transmission 1A is arranged in parallel with a primary pulley 50 having a primary pulley shaft 51 arranged concentrically with the input shaft 38, with a predetermined distance from the primary pulley shaft 51. A secondary pulley 60 having a secondary pulley shaft 61, and a belt 110 wound around the primary pulley 50 and the secondary pulley 60.

  First, the configuration on the primary pulley 50 side will be described.

  The primary pulley 50 is one pulley in the belt-type continuously variable transmission 1 </ b> A that rotates with respect to a stationary component (for example, the transaxle 20), and is transmitted from the internal combustion engine 10 transmitted through the forward / reverse switching mechanism 40. The output torque is transmitted to the secondary pulley 60 by the belt 110. As shown in FIGS. 1 and 2, the primary pulley 50 includes a primary pulley shaft 51, a primary fixed sheave 52, a primary movable sheave 53, a primary partition wall 54, and a primary hydraulic chamber 55. .

  The primary pulley shaft 51 is rotatably supported by pulley bearings 111 and 112 as shown in FIG. The primary pulley shaft 51 has a supply / discharge side main passage 51a that opens only at one end (right end in the figure) of both ends in the axial direction, and the other end (left end in the figure). And a drive side main passage 51b that is open only at the top.

  The supply / discharge-side main passage 51a forms a part of the hydraulic oil supply path and the hydraulic oil discharge path for the primary hydraulic chamber 55, and is formed on the primary fixed sheave 52 side to form the hydraulic oil supply control device. 130 communicates with an oil passage R7 described later. The hydraulic oil supplied from the hydraulic oil supply control device 130 to the primary hydraulic chamber 55 flows into the supply / discharge side main passage 51a, while the hydraulic oil discharged from the primary hydraulic chamber 55 flows into the supply / discharge side main passage 51a. That is, the supply / discharge side main passage 51 a allows the hydraulic oil supplied or discharged between the hydraulic oil supply control device 130 and the primary hydraulic chamber 55 to pass therethrough. The supply / discharge-side main passage 51a includes a shaft-side communication passage 51c formed in the vicinity of the tip thereof, an annular space T1 formed between the primary movable sheave 53 and the primary pulley shaft 51, a primary partition wall 54, and a primary partition. It communicates with the partition-side communication passage 54b of the primary partition 54 via an annular space T2 formed between the movable sheave 53 and the primary pulley shaft 51. The spaces T1 and T2 communicate with each other via a spline 53c of the primary movable sheave 53, which will be described later, and the primary pulley shaft 51. Further, the shaft side communication passage 51c of the first embodiment is formed at a plurality of locations (for example, three locations) at equal intervals on the circumference of the supply / discharge side main passage 51a.

  On the other hand, the drive-side main passage 51b forms part of a hydraulic oil supply path and a hydraulic oil discharge path for the drive hydraulic chamber 81 of the actuator 80 described later, and is formed on the primary movable sheave 53 side. The hydraulic oil supply control device 130 communicates with an oil passage R8 described later. The hydraulic fluid of the actuator 80 supplied from the hydraulic fluid supply control device 130 to the drive hydraulic chamber 81 flows into the drive side main passage 51b, while the hydraulic fluid discharged from the drive hydraulic chamber 81 flows. That is, the drive side main passage 51 b allows the hydraulic oil supplied or discharged between the hydraulic oil supply control device 130 and the drive hydraulic chamber 81 to pass therethrough. The drive-side main passage 51b is formed in a shaft-side communication passage 51d formed in the vicinity of the tip, an annular space T3 formed between the primary partition wall 54 and the primary pulley shaft 51, and the primary partition wall 54. The drive hydraulic chamber 81 communicates with the drive communication passage 54e. The shaft-side communication passage 51d of the first embodiment is formed at a plurality of locations (for example, three locations) at equal intervals on the circumference of the drive-side main passage 51b. In the first embodiment, a plurality of drive communication passages 54e of the primary partition wall 54 are also formed at equal intervals (for example, three locations) on the circumference in accordance with the number of the shaft side communication passages 51d. .

  Further, as shown in FIG. 2, the primary fixed sheave 52 constituting a part of the primary pulley 50 is provided to rotate integrally with the primary pulley shaft 51 at a position facing the primary movable sheave 53. The primary fixed sheave 52 illustrated here is formed as an annular portion that protrudes radially outward from the outer periphery of the primary pulley shaft 51. That is, the primary fixed sheave 52 of the first embodiment is integrally formed on the outer periphery of the primary pulley shaft 51.

  On the other hand, as shown in FIG. 2, the primary movable sheave 53 includes a cylindrical portion 53a and an annular portion 53b. The cylindrical portion 53a is formed around the same rotational axis as the primary pulley shaft 51, and the annular portion 53b is formed to project radially outward from an end portion of the cylindrical portion 53a on the primary fixed sheave 52 side. . The primary movable sheave 53 is connected to the primary pulley shaft 51 by spline-fitting a spline 53c formed on the inner peripheral surface of the cylindrical portion 53a and a spline 51e formed on the outer peripheral surface of the primary pulley shaft 51. It is supported so that it can slide to an axial direction.

  Here, between the primary fixed sheave 52 and the primary movable sheave 53 (that is, between the surfaces facing the primary fixed sheave 52 and the primary movable sheave 53), a V-shaped belt around which the belt 110 is wound is formed. A primary groove 110a is formed.

  Further, as shown in FIG. 2, the primary partition wall 54 forming a part of the primary pulley 50 is an annular member, and is centered on the same rotational axis as the primary pulley shaft 51 and sandwiches the primary movable sheave 53. The primary fixed sheave 52 is disposed in a positional relationship facing the primary fixed sheave 52 in the axial direction. The primary partition wall 54 is provided so as to rotate integrally with the primary pulley shaft 51 by being press-fitted into the primary pulley shaft 51 and tightened with a nut through the pulley shaft 112.

  The primary partition wall 54 is formed with a valve arrangement passage 54a extending radially inward in the axial direction. This valve arrangement passage 54a has one end (right side end in the figure) opened to the primary hydraulic chamber 55 and the other end (left side end in the figure) is closed inside the primary partition 54 and connected to the partition side. It communicates with the passage 54b. An annular valve seat surface 72 that is closed by a valve body 71 (described later) of the hydraulic oil supply / discharge valve 70 is formed in the valve arrangement passage 54a. Here, the valve seat surface 72 is formed at the one end of the valve arrangement passage 54a.

  Accordingly, the partition-side communication passage 54 b of the primary partition 54 functions as a hydraulic oil supply passage that forms part of the hydraulic oil supply path that supplies hydraulic oil to the primary pulley 50 via the hydraulic oil supply / discharge valve 70. Also, it functions as a hydraulic oil discharge passage that forms part of the hydraulic oil discharge path for discharging hydraulic oil from the primary pulley 50 via the hydraulic oil supply / discharge valve 70. The partition-side communication passage 54b communicates one end (diameter outer side end in the figure) with the valve arrangement passage 54a and the other end (diameter inner side in the figure) on the inner peripheral surface of the primary partition wall 54. It opens and communicates with the space T2. The valve arrangement passage 54a and the partition wall side communication passage 54b of the first embodiment are formed at a plurality of locations (for example, three locations) at equal intervals on the circumference. In addition, the hydraulic oil supply / discharge valve 70 is arranged in each valve arrangement passage 54a.

  Further, the primary partition wall 54 is formed with sliding support holes 54c shown in FIG. 2 on the same axis as the valve arrangement passages 54a. Each of the sliding support holes 54c has one end portion (right end portion in the figure) communicated with the valve arrangement passage 54a, and the other end portion (left end portion in the figure) is connected to the piston 82 (piston main body 82a) of the actuator 80. ) In the reciprocating movement direction (axial direction) of FIG.

  The primary partition wall 54 is provided with a drive communication passage 54e for supplying or discharging hydraulic oil between the drive side main passage 51b and the drive hydraulic chamber 81. In the first embodiment, a cover member 83 to be described later is provided between the primary partition wall 54 and the pulley bearing 112. The cover member 83 is provided with an annular portion facing each other in the axial direction. The cover member 83 is fitted to the primary pulley shaft 51, for example. Therefore, in the first embodiment, a groove communicating with the drive hydraulic chamber 81 is formed on the surface of the primary partition wall 54 facing the cover member 83, and this groove is used as the drive communication passage 54e. Thereby, since the thickness of the portion where the drive communication passage 54e is formed in the primary partition wall 54 can be reduced, the axial length on the primary pulley 50 side of the belt type continuously variable transmission 1A can be shortened.

  Subsequently, the primary hydraulic chamber 55 forming a part of the primary pulley 50 is one clamping pressure generating hydraulic chamber in the belt type continuously variable transmission 1A, and the primary movable sheave 53 is fixed to the primary as shown in FIG. By pressing toward the sheave 52 side, a belt clamping pressure is generated with respect to the primary pulley 50 (in other words, the belt 110 wound around the V-shaped primary groove 110a). The primary hydraulic chamber 55 is a space formed by the primary movable sheave 53 and the primary partition wall 54. Here, a seal for a primary hydraulic chamber such as a seal ring is provided between the cylindrical protruding portion 53 d of the primary movable sheave 53 and the primary partition wall 54 and between the cylindrical portion 53 a of the primary movable sheave 53 and the primary partition wall 54. Each member S1 is provided. In other words, the space formed by the primary movable sheave 53 and the primary partition wall 54 constituting the primary hydraulic chamber 55 is sealed by the primary hydraulic chamber seal member S1 to ensure oil tightness.

  The primary hydraulic chamber 55 is supplied with hydraulic oil from the hydraulic oil supply control device 130 that has flowed into the supply / discharge side main passage 51 a of the primary pulley shaft 51. Specifically, the primary hydraulic chamber 55 slides the primary movable sheave 53 in the axial direction by the pressure of the hydraulic oil supplied from the hydraulic oil supply control device 130 (that is, the hydraulic pressure P1 of the primary hydraulic chamber 55). The primary movable sheave 53 is moved toward or away from the primary fixed sheave 52. Accordingly, the primary hydraulic chamber 55 generates a belt clamping pressure with respect to the belt 110 by the hydraulic pressure P1 of the primary hydraulic chamber 55, and changes the position of the primary movable sheave 53 in the axial direction with respect to the primary fixed sheave 52. Accordingly, the hydraulic pressure of the primary hydraulic chamber 55 is adjusted mainly when changing the gear ratio of the belt type continuously variable transmission 1A.

  Here, the hydraulic oil supply / discharge valve 70 will be described.

  The hydraulic oil supply / discharge valve 70 is a hydraulic oil supply valve and a hydraulic oil discharge valve. That is, the hydraulic oil supply / discharge valve 70 is opened when hydraulic oil is supplied to the primary hydraulic chamber 55 that is one clamping pressure generating hydraulic chamber in the belt-type continuously variable transmission 1A. The valve is also opened when the hydraulic oil is discharged from the chamber 55. As shown in FIGS. 2, 7, and 9, the hydraulic oil supply / discharge valve 70 is an operation that is a working fluid from the outside of the primary hydraulic chamber 55 (that is, the outside of the primary pulley 50) to the primary hydraulic chamber 55. The oil is supplied, the hydraulic oil is discharged from the primary hydraulic chamber 55 to the outside of the primary pulley 50, and the hydraulic oil in the primary hydraulic chamber 55 is retained. The hydraulic oil supply / discharge valve 70 of the first embodiment is arranged in each valve arrangement passage 54a formed in the primary partition wall 54 of the primary pulley 50 (that is, the hydraulic oil supply / discharge valve 70 of the first embodiment). Is formed at a plurality of locations (for example, three locations) at equal intervals on the circumference). That is, each hydraulic oil supply / discharge valve 70 of the first embodiment is arranged and configured to rotate integrally with the primary pulley 50.

  For example, each hydraulic oil supply / discharge valve 70 of the first embodiment is a ball check valve, and includes a valve body 71, a valve seat surface 72, a valve body elastic member 73, a pressing force transmission member 74, And a snap ring 75.

  The valve body 71 is a spherical body having a diameter larger than the minimum inner diameter of the valve seat surface 72 and is disposed closer to the primary hydraulic chamber 55 than the valve seat surface 72. The valve seat surface 72 here has a tapered shape such as a truncated cone having a diameter that increases toward the primary fixed sheave 52 side (from the other end of the valve arrangement passage 54a to the one end). It arrange | positions on the press part 82b of piston 82, and a concentric circle. Therefore, the hydraulic oil supply / discharge valve 70 brings the valve element 71 into contact with the valve seat surface 72 to block communication between the valve disposition passage 54 a and the primary hydraulic chamber 55, while the valve element 71 is connected to the valve seat surface 72. The valve arrangement passage 54a and the primary hydraulic chamber 55 are communicated with each other by being separated from each other. That is, each hydraulic oil supply / discharge valve 70 of the first embodiment performs a valve opening operation by moving the valve body 71 toward the primary fixed sheave 52 side and moving away from the valve seat surface 72, and the valve body 71 is different from that. The valve is closed by moving in the opposite direction and contacting the valve seat surface 72.

  The valve body elastic member 73 is a valve body valve closing direction pressing force generation means. The valve body elastic member 73 is arranged in a state of being biased in a direction of pressing the valve body 71 against the valve seat surface 72 via the pressing force transmission member 74. For example, here, an annular disc spring arranged on the same axis as the primary pulley shaft 51 is used as the valve body elastic member 73, and the valve body elasticity is used on the same axis as the primary pulley shaft 51 as the pressing force transmission member 74. An annular member disposed closer to the valve body 71 than the member 73 is used. Here, the valve body elastic member 73 is disposed so that only the outer side of the hydraulic oil supply / discharge valve 70 in the closed state generates the urging force in the valve closing direction on the pressing force transmission member 74. For this reason, the inner peripheral surface side of the valve body elastic member 73 is pressed by the snap ring 75 so as not to move in the valve opening direction of the hydraulic oil supply / discharge valve 70 with respect to the primary partition wall 54. The pressing force transmission member 74 here is arranged so as to be able to slide in the axial direction on the outer peripheral surface of the cylindrical portion covering the cylindrical portion 53a of the primary movable sheave 53 in the primary partition 54. The snap ring 75 holds the valve body elastic member 73 on the outer peripheral surface of the cylindrical portion of the primary partition wall 54. Accordingly, the valve body elastic member 73 generates a valve closing urging force in a direction in which the valve body 71 is brought into contact with the valve seat surface 72, so that the valve closing urging force is applied in the valve closing direction of the hydraulic oil supply / discharge valve 70. The valve body 71 acts as a valve body closing direction pressing force, which is an elastic member pressing force, via the pressing force transmitting member 74. Accordingly, the valve body 71 is pressed against the valve seat surface 72, and each hydraulic oil supply / discharge valve 70 functions as a check valve.

  Here, the primary partition wall 54 is formed with an annular recess 54 f having a size equal to or greater than that of the pressing force transmission member 74 on the surface facing the pressing force transmission member 74, via the pressing force transmission member 74. The valve body closing direction pressing force can be applied to the valve body 71. The pressing force transmission member 74 is preferably formed with a concave portion where the contact end of the valve body 71 falls in a contact portion with the valve body 71, so that the valve body 71 is opened or closed. Since rolling on the pressing force transmission member 74 during operation can be avoided, dispersion of the pressing force to the valve body 71 when the valve is opened or closed is prevented, and hydraulic oil is supplied and discharged with good responsiveness. Valve 70 can be actuated.

  Next, the actuator 80 will be described.

  The actuator 80 is a valve opening / closing means for each hydraulic oil supply / discharge valve 70 and forcibly opens each hydraulic oil supply / discharge valve 70. The actuator 80 includes a drive hydraulic chamber 81 and a piston 82, and is disposed on the opposite side of the hydraulic oil supply / discharge valve 70 with the primary partition wall 54 interposed therebetween.

  The drive hydraulic chamber 81 is supplied with hydraulic oil, and the opening / closing operation of each hydraulic oil supply / discharge valve 70 is controlled by the pressure of the supplied hydraulic oil (that is, the hydraulic pressure P2 of the drive hydraulic chamber 81). Is. The drive hydraulic chamber 81 is disposed on the other side (the left side) of the piston 82 in the reciprocating direction (axial direction) of the piston 82, and the side and the cover member 83 shown in FIG. It is comprised by the cyclic | annular space part formed between these. The cover member 83 includes the above-described annular portion that is sandwiched between the primary partition wall 54 and the pulley bearing 112, and the axial portion of the annular piston main body 82 a of the piston 82 that extends in the axial direction from the outer peripheral end of the annular portion. A cylindrical portion covering the outer peripheral surface. Here, between the outer peripheral surface of the piston main body 82a of the piston 82 and the cylindrical portion of the cover member 83 and between the inner peripheral surface of the piston main body 82a and the primary partition wall 54, for example, a drive hydraulic chamber such as a seal ring. Seal members S2 and S3 are provided. That is, the space formed by the piston main body 82a of the piston 82 constituting the drive hydraulic chamber 81 and the cover member 83 is sealed by the drive hydraulic chamber seal member S2 to ensure oil tightness.

  The piston 82 is an annular piston main body 82a that slides in the axial direction by the action of the hydraulic pressure P2 of the drive hydraulic chamber 81, and the hydraulic oil supply / discharge valve 70 is driven by the piston main body 82a. The cylinder 71 includes a plurality of columnar or cylindrical pressing portions 82b that apply a pressing force to the body 71, and is supported slidably in the axial direction with respect to the drive hydraulic chamber 81.

  The piston main body 82a of the piston 82 is axially directed toward the valve element 71 by the piston valve opening direction pressing force acting by the hydraulic pressure P2 of the drive hydraulic chamber 81 (that is, the valve opening direction of each hydraulic oil supply / discharge valve 70). To slide. On the other hand, each pressing portion 82b of the piston 82 is inserted into each sliding support hole 54c of the primary partition wall 54, and is supported to be slidable in the axial direction in each sliding support hole 54c. For this reason, each pressing portion 82b rotates integrally with the primary pulley 50. Further, each of the pressing portions 82b has one end portion (right end portion in FIG. 2) in contact with the valve body 71 of each hydraulic oil supply / discharge valve 70, and the other end portion (left end portion in the figure). The piston main body 82a can be contacted. Therefore, each pressing portion 82b slides in the axial direction with respect to the primary pulley 50 while being in contact with each valve body 71 and the piston main body 82a, so that the actuator 80 and each hydraulic oil supply / discharge valve 70 are in contact with each other. In the meantime, an axial force (ie, a piston valve opening direction pressing force acting on the piston main body 82a of the piston 82) is transmitted to each valve body 71. As a result, the piston 82 can open each hydraulic oil supply / discharge valve 70. Each pressing portion 82b may be a separate body from the piston main body 82a, or may be provided integrally with the piston main body 82a. In the present Example 1, the former separate structure is illustrated.

  Here, the valve body 71 of each hydraulic oil supply / discharge valve 70 described above has a valve body opening direction pressing force in a direction away from the valve seat surface 72 (that is, the valve opening direction of the hydraulic oil supply / discharge valve 70). The valve body closing direction pressing force in the direction of contacting the valve seat surface 72 (that is, the closing direction of the hydraulic oil supply / discharge valve 70) acts. The valve body closing direction pressing force includes an elastic member pressing force that acts on the valve body 71 by the valve closing biasing force of the valve body elastic member 73 and an operation that acts on the valve body 71 by the hydraulic pressure P1 of the primary hydraulic chamber 55. Oil closing valve direction pressing force is included. Therefore, each hydraulic oil supply / discharge valve 70 opens the valve element 71 away from the valve seat surface 72 when the valve element valve opening direction pressing force exceeds the valve element closing direction pressing force. On the other hand, each of the hydraulic oil supply / discharge valves 70 closes when the valve body 71 comes into contact with the valve seat surface 72 when the valve body closing direction pressing force exceeds the valve body opening direction pressing force. That is, in the first embodiment, each hydraulic oil supply / discharge valve 70 is forcibly opened by increasing the valve opening direction pressing force by the piston valve opening direction pressing force of the piston 82. Therefore, in the first embodiment, when supplying hydraulic oil to the primary hydraulic chamber 55 and when discharging hydraulic oil from the primary hydraulic chamber 55, the actuator 80 is operated to force each hydraulic oil supply / discharge valve 70. To open and forcibly supply and discharge hydraulic oil.

  The hydraulic oil valve closing direction pressing force acting on the valve body 71 by the hydraulic pressure P1 of the primary hydraulic chamber 55 acts on the valve body 71 as the valve closing direction pressing force as described above. Even if P1 rises, the valve body 71 does not leave the valve seat surface 72. Accordingly, the closed state of each hydraulic oil supply / discharge valve 70 is maintained unless the valve body opening direction pressing force acting on the valve body 71 exceeds the valve body opening direction pressing force. The hydraulic oil in the primary hydraulic chamber 55, which is a hydraulic chamber, is reliably held in the primary hydraulic chamber 55, and the hydraulic pressure in the primary hydraulic chamber 55 can be kept constant.

  Therefore, as in the conventional belt-type continuously variable transmission, in order to maintain the position of the primary movable sheave 53 in the axial direction relative to the primary fixed sheave 52, one hydraulic pressure generating hydraulic chamber is supplied from the hydraulic oil supply control device 130. When the hydraulic oil is continuously supplied to the primary hydraulic chamber 55, hydraulic oil having a predetermined pressure exists in the hydraulic oil supply path from the hydraulic oil supply control device 130 to the primary hydraulic chamber 55. The conventional hydraulic oil supply path includes a plurality of sliding portions between the stationary member and the movable member, and hydraulic oil of a predetermined pressure is supplied from the sliding portion to the hydraulic oil supply path when the speed ratio is fixed. There was a risk of leaking outside.

  Here, the stationary member is a member that does not rotate, slide, or the like among the members constituting the belt type continuously variable transmission 1A. For example, a transaxle housing 21, a transaxle case 22, and a transaxle rear cover 23 of the transaxle 20. On the other hand, the movable member is a member that rotates, slides, and the like in the members constituting the belt type continuously variable transmission 1A. For example, the primary pulley shaft 51 or the like. Therefore, the sliding portion includes, for example, a portion where the primary pulley shaft 51 rotates with respect to the transaxle housing 21, the transaxle case 22, the transaxle rear cover 23, and the like of the transaxle 20.

  In contrast to the conventional belt-type continuously variable transmission, in the belt-type continuously variable transmission 1A of the first embodiment, each hydraulic oil supply / discharge valve 70 is disposed between the primary hydraulic chamber 55 and the sliding portion. ing. That is, when each hydraulic oil supply / discharge valve 70 is kept closed and the hydraulic oil is held in the primary hydraulic chamber 55, between the primary hydraulic chamber 55 and each hydraulic oil supply / discharge valve 70. In addition, there is no sliding portion between the fixed member and the movable member as in the prior art. As a result, in the belt type continuously variable transmission 1A of the first embodiment, it becomes possible to suppress the leakage of hydraulic oil from the sliding portion, so that an oil pump, which will be described later, in the hydraulic oil supply control device 130 An increase in driving loss 132 can be suppressed.

  In addition, each hydraulic oil supply / discharge valve 70 of the first embodiment is a valve that is shared with the hydraulic oil supply valve and the hydraulic oil discharge valve, and therefore is opened (forcibly opened in the first embodiment). Thus, the operation of supplying the hydraulic oil to the primary hydraulic chamber 55 and the operation of discharging the hydraulic oil from the primary hydraulic chamber 55 are closed, and the holding operation of the hydraulic oil in the primary hydraulic chamber 55 is performed by closing the valve. Can be done with a valve. In the first embodiment, the hydraulic pressure P2 of the drive hydraulic chamber 81 is used as a drive source of the actuator 80 for forcibly opening each hydraulic oil supply / discharge valve 70. However, the drive source is not necessarily limited to this. For example, rotational force such as a motor or electromagnetic force may be used.

  By the way, the shaft side communication passage 51 d of the primary pulley shaft 51 and the drive communication passage 54 e of the primary partition wall 54 rotate together with the rotation of the primary pulley 50. For this reason, the hydraulic oil in the shaft side communication passage 51d and the drive communication passage 54e flows toward the drive hydraulic chamber 81 due to the centrifugal force accompanying the rotation of the primary pulley 50, and increases the hydraulic pressure P2 of the drive hydraulic chamber 81. there is a possibility. That is, although it is a case where it is desired to maintain the closed state of each hydraulic oil supply / discharge valve 70 (that is, when it is not necessary to supply hydraulic oil to the drive hydraulic chamber 81 as in the case of a fixed gear ratio), There is a possibility that the actuator 80 (the piston main body 82a of the piston 82) is operated by the centrifugal hydraulic pressure generated in the drive hydraulic chamber 81, and the hydraulic oil supply / discharge valves 70 are opened.

  Therefore, the actuator 80 according to the first embodiment is provided with a closed state holding means for avoiding the operation of the actuator 80 due to the centrifugal hydraulic pressure and holding each hydraulic oil supply / discharge valve 70 in the closed state.

  Specifically, if the piston 82 is held at the initial position (here, the position when each hydraulic oil supply / discharge valve 70 shown in FIG. 2 is in the closed position is the initial position), each hydraulic oil supply / discharge is performed. Since the valve 70 is kept in the closed state, as the closed state holding means of the first embodiment, a mechanical pressing force against the centrifugal hydraulic pressure is applied to the piston 82, whereby the piston 82 is initially set. The piston initial position holding means for holding the position will be exemplified.

  As shown in FIGS. 2 to 4-2, the piston initial position holding means includes a centrifugal hydraulic pressure cancellation chamber 84 formed on the opposite side of the drive hydraulic pressure chamber 81 around the piston 82, and the centrifugal hydraulic pressure cancellation chamber 84. And at least a centrifugal hydraulic pressure canceling member 85 that applies a pressing force against the centrifugal hydraulic pressure of the drive hydraulic chamber 81 to the piston 82.

As shown in FIGS. 3 and 4-1, the centrifugal hydraulic pressure canceling chamber 84 is a surface opposite to the driving hydraulic pressure chamber 81 across the piston main body 82a of the piston 82 (hereinafter referred to as “centrifugal hydraulic drag generating surface”). ) 82a and a space portion provided on _one side, is surrounded by the recess and the cover member 83 which is formed in the contact surface between the centrifugal hydraulic pressure drag generating surface 82a ₁ of the centrifugal hydraulic drag generating surface 82a ₁ and the primary partition wall 54 It refers to a space. Further, the centrifugal hydraulic pressure canceling member 85 generates a pressing force radially outward in the centrifugal hydraulic pressure canceling chamber 84 by a centrifugal force accompanying the rotation of the primary pulley 50. The portion is applied to the centrifugal hydraulic drag generating surface 82a ₁ of the piston main body 82a according to the shape of the centrifugal hydraulic pressure canceling chamber 84.

Here, the centrifugal hydraulic pressure canceling member 85 is free to roll in the centrifugal hydraulic pressure canceling chamber 84 like a spherical shape or a cylindrical shape, and the piston main body 82a of the piston 82 is pressed radially outward by centrifugal force. A weight having a mass that can be held at the initial position against the centrifugal hydraulic pressure is used. In the first embodiment, a cylindrical weight (hereinafter referred to as “roller”) is exemplified. Further, in the centrifugal hydraulic pressure canceling chamber 84 of the first embodiment, an inclined surface 84a is formed in which the interval between the piston main body 82a and the centrifugal hydraulic drag generating surface 82a ₁ becomes narrower toward the radially outer side.

Accordingly, since the centrifugal force acts on the roller 85 in a state where the roller 85 is sandwiched between the centrifugal hydraulic drag generating surface 82a ₁ and the inclined surface 84a, a part of the pressing force radially outward due to the centrifugal force is in the axial direction. It acts on the centrifugal hydraulic drag generation surface 82a ₁ as the centrifugal hydraulic drag. For this reason, even if centrifugal hydraulic pressure is generated in the drive hydraulic chamber 81 to open each hydraulic oil supply / discharge valve 70, at least the centrifugal hydraulic drag that resists this is caused by the centrifugal hydraulic cancel chamber 84 and the roller 85. Since it is generated in the piston main body 82a, it is possible to avoid the unnecessary opening of each hydraulic oil supply / discharge valve 70 while the piston 82 maintains the initial position. As a result, when it is desired to maintain the closed state of each hydraulic oil supply / discharge valve 70, the hydraulic pressure in the primary hydraulic chamber 55 is kept constant until the actuator 80 is desired to operate. And the gear ratio can be kept fixed. On the other hand, when the actuator 80 is operated, as shown in FIG. 4B, the roller 85 is pushed by the piston main body 82a and rolls inward in the radial direction of the inclined surface 84a.

  Further, by providing such a mechanical piston initial position holding means, the piston 82 can be reduced in size and weight. That is, when hydraulic pressure is used as the source of centrifugal hydraulic drag of the piston initial position holding means, it is necessary to increase the contact area with the hydraulic pressure in the piston main body 82a in order to increase the pressing force by the hydraulic pressure. Will become larger. Further, since the piston 82 is more susceptible to the centrifugal hydraulic pressure of the drive hydraulic chamber 81 as the mass is smaller, if such a piston initial position holding means is not provided, the mass is increased to reduce the influence of the centrifugal hydraulic pressure. There was a need to do. However, in the first embodiment, since the influence of the centrifugal hydraulic pressure can be canceled by the piston initial position holding means using this mechanical force, the piston 82 can be reduced in size and weight. Along with this, the elasticity of the valve body elastic member 73 of the hydraulic oil supply / discharge valve 70 can be reduced. Therefore, the piston initial position holding means according to the first embodiment can easily reciprocate the piston 82, so that the response when the actuator 80 is activated can be improved.

  Further, if the piston initial position holding means is not provided, for example, the hydraulic oil supply / discharge valve 70 is opened by the centrifugal hydraulic pressure when the transmission gear ratio is fixed. Therefore, in order to keep the hydraulic pressure in the primary hydraulic chamber 55 constant, Although it is necessary to drive the pump 132 and supply hydraulic oil to the primary hydraulic chamber 55, the drive of the oil pump 132 for that purpose is unnecessary here, so that an increase in drive loss of the oil pump 132 can be suppressed. . Furthermore, the piston initial position holding means of the first embodiment uses a spontaneous mechanical force called centrifugal force generated by the roller 85 as a source of centrifugal hydraulic drag, and does not use hydraulic pressure. For example, it is not necessary to drive the oil pump 132 of the hydraulic oil supply control device 130 in order to generate the hydraulic drag, so that an increase in driving loss of the oil pump 132 can be suppressed, and when the centrifugal hydraulic drag is generated. Controllability and reliability can be improved.

Here, in order to prevent the influence of the centrifugal hydraulic pressure at the start of operation (in other words, when the primary pulley 50 starts to rotate), the roller 85 is subjected to centrifugal hydraulic drag at least immediately when the primary pulley 50 starts to rotate. It is preferable that the generating surface 82a ₁ and the inclined surface 84a are brought into contact with each other, and a centrifugal hydraulic drag is applied to the centrifugal hydraulic drag generating surface 82a ₁ . Here, for this purpose, as shown in FIG. 4A, the shortest distance L1 between the radially inner surface 84b of the centrifugal hydraulic pressure cancellation chamber 84 and the inner peripheral surface of the piston main body 82a, its surface 84b and the inclined surface 84a. The shortest distance L2 is defined between the two flat portions and the shortest distances L1 and L2 are set to be larger than the radius r of the roller 85 (same radius r in the case of a spherical body) (r> L1). , L2). Thereby, in the first embodiment, the roller 85 rolls smoothly without being caught in the centrifugal hydraulic pressure cancellation chamber 84 or the piston main body 82a at the start of operation, so that the centrifugal hydraulic drag generating surface 82a ₁ and the inclined surface 84a The operation of the actuator 80 due to the centrifugal hydraulic pressure at the start of operation can be avoided, and highly reliable gear ratio control can be executed with high accuracy.

In this piston initial position holding means, for example, hydraulic fluid that has flowed into the piston main body 82a side from between the sliding support hole 54c and the pressing portion 82b of the piston 82 has a minute amount between the primary partition wall 54 and the piston main body 82a. It is transmitted to the centrifugal hydraulic pressure cancellation chamber 84 from the gap. For this reason, the hydraulic oil becomes lubricating oil and prevents sliding resistance and wear between the roller 85 and the centrifugal hydraulic drag generation surface 82a ₁ and between the roller 85 and the inclined surface 84a. Here, since the hydraulic oil flows through all the sliding portions of the belt-type continuously variable transmission 1A, there is a possibility that the hydraulic oil contains foreign matters such as wear powder and dust although it is minute. Therefore, in the first embodiment, a space 84c shown in FIG. 4A is provided radially outside the roller 85 in the centrifugal hydraulic pressure cancellation chamber 84, and foreign matter is collected in the space 84c. As a result, the foreign matter can be confined radially outward from the roller 85, and the situation where the roller 85 bites the foreign matter at the sliding portion is avoided, so that the smooth movement of the roller 85 is ensured. In addition, wear and scratches on the roller 85 and its sliding surface are prevented, and the reliability of the piston initial position holding means can be improved.

  On the other hand, if the hydraulic oil in the centrifugal hydraulic pressure cancellation chamber 84 increases too much, this hydraulic oil becomes a resistance and prevents smooth movement of the roller 85. For this reason, in the present Example 1, the oil drain hole 86 shown in FIG. 4-1 which can drain the excess hydraulic fluid in the centrifugal hydraulic pressure cancellation chamber 84 is provided. Here, the opening position of the oil drain hole 86 in the centrifugal hydraulic pressure cancellation chamber 84 is defined using the three diameters (R1, R2, R3) shown in FIG. The hydraulic oil is allowed to spread over each sliding surface. “R1” is a diameter from the rotation axis center to the opening position of the oil drain hole 86, “R2” is a diameter from the rotation axis center to the sliding surface of the roller 85 and the piston main body 82a, and “R3” is from the rotation axis center. The diameter to the sliding surface of the roller 85 and the inclined surface 84a is represented, and the diameter R1 to the opening position of the oil drain hole 86 is set to be smaller than the diameters R2 and R3 to the respective roller sliding surfaces. (R1 <R2, R3). That is, the opening position of the oil drain hole 86 is disposed radially inward from the sliding surfaces of the rollers 85. Thereby, since hydraulic oil spreads to each sliding surface of the roller 85 and wear of the roller 85 and the sliding surface can be prevented, durability of the piston initial position holding means is improved.

  The combination of the centrifugal hydraulic pressure canceling chamber 84 and the centrifugal hydraulic pressure canceling member (roller) 85 includes a plurality of sets (here, as shown in FIG. 3) so that a balanced pressing force is applied to the entire piston main body 82a of the piston 82. 3 sets), and it is preferable to arrange each of them equally about the axis.

  Next, the configuration on the secondary pulley 60 side will be described.

  This secondary pulley 60 is the other pulley in the belt-type continuously variable transmission 1 </ b> A that rotates with respect to a stationary part, for example, the transaxle 20, and is transmitted from the internal combustion engine 10 transmitted to the primary pulley 50 via the belt 110. The output torque is transmitted to the final reduction gear 90 of the belt type continuously variable transmission 1A. As shown in FIG. 1, the secondary pulley 60 includes a secondary pulley shaft 61, a secondary fixed sheave 62, a secondary movable sheave 63, a secondary hydraulic chamber 64, a secondary partition wall 65, and a torque cam 66. Reference numeral 69 indicates a parking brake gear.

  The secondary pulley shaft 61 is rotatably supported by pulley bearings 113 and 114. The secondary pulley shaft 61 has a hydraulic oil passage (not shown) therein, and hydraulic oil supplied from the hydraulic oil supply control device 130 to the secondary hydraulic chamber 64 flows into the hydraulic oil passage.

  The secondary fixed sheave 62 is provided so as to rotate integrally with the secondary pulley shaft 61 at a position facing the secondary movable sheave 63. The secondary fixed sheave 62 illustrated here is formed as an annular portion that protrudes radially outward from the outer periphery of the secondary pulley shaft 61. That is, the secondary fixed sheave 62 of the first embodiment is integrally formed on the outer periphery of the secondary pulley shaft 61.

  On the other hand, the secondary movable sheave 63 has a spline (not shown) formed on the inner peripheral surface of the secondary movable sheave 63 and a spline (not shown) formed on the outer peripheral surface of the secondary pulley shaft 61. On the other hand, it is supported so that it can slide to an axial direction.

  Here, between the secondary fixed sheave 62 and the secondary movable sheave 63 (that is, between the surfaces facing the secondary fixed sheave 62 and the secondary movable sheave 63), the belt 110 is wound around. A secondary groove 110b is formed.

  The secondary hydraulic chamber 64 is the other clamping pressure generating hydraulic chamber in the belt-type continuously variable transmission 1A, and as shown in FIG. 1, the secondary movable sheave 63 is pressed toward the secondary fixed sheave 62 to A belt clamping pressure is generated with respect to the pulley 60 (in other words, the belt 110 wound around the V-shaped secondary groove 110b). The secondary hydraulic chamber 64 is a space formed by a secondary pulley shaft 61, a secondary movable sheave 63, and a disk-shaped secondary partition wall 65 fixed to the secondary pulley shaft 61.

  Here, the secondary movable sheave 63 is formed with an annular protruding portion 63a that protrudes in one axial direction (that is, protrudes toward the final reduction gear 90). On the other hand, the secondary partition wall 65 is formed with an annular protrusion 65a that protrudes in the other direction in the axial direction (that is, protrudes toward the secondary movable sheave 63). A seal member (not shown) such as a seal ring is provided between the respective protrusions 63a and 65a. In other words, the space formed by the secondary movable sheave 63 and the secondary partition wall 65 constituting the secondary hydraulic chamber 64 is sealed by a secondary hydraulic chamber seal member (not shown) to ensure oil tightness.

  The hydraulic oil from the hydraulic oil supply control device 130 that has flowed into the hydraulic oil passage (not shown) of the secondary pulley shaft 61 is supplied to the secondary hydraulic chamber 64 via a hydraulic fluid supply hole (not shown). That is, the secondary hydraulic chamber 64 slides the secondary movable sheave 63 in the axial direction by the pressure of the hydraulic oil supplied from the hydraulic oil supply control device 130 (that is, the internal pressure of the secondary hydraulic chamber 64). 63 is moved toward or away from the secondary fixed sheave 62. Thus, the secondary hydraulic chamber 64 generates a belt clamping pressure with respect to the belt 110 by the internal pressure of the secondary hydraulic chamber 64, and the contact radius of the belt 110 with respect to the primary pulley 50 and the secondary pulley 60 is made constant. Can be maintained.

  Further, as shown in FIG. 5A, the torque cam 66 includes a mountain-shaped first engagement portion 63 b provided in an annular shape on the secondary movable sheave 63 of the secondary pulley 60, and the first engagement portion 63 b and the secondary pulley. A second engagement portion 67a formed on an intermediate member 67, which will be described later, facing each other in the axial direction of the shaft 61, and a disc-shaped member disposed between the first engagement portion 63b and the second engagement portion 67a. And a plurality of transmission members 68.

  The intermediate member 67 is formed integrally with the secondary partition wall 65 or fixed to the secondary partition wall 65, and can be rotated relative to the secondary pulley shaft 61 and the secondary movable sheave 63 on the secondary pulley shaft 61 by pulley bearings 113 and 115. It is supported. The intermediate member 67 is fixed to the input shaft 101 of the power transmission path 100 by, for example, spline fitting. That is, the output torque from the internal combustion engine 10 transmitted to the secondary pulley 60 is transmitted to the power transmission path 100 via the intermediate member 67.

  Here, the operation of the torque cam 66 will be described. When the output torque from the internal combustion engine 10 is transmitted to the primary pulley 50 and the primary pulley 50 rotates, the secondary pulley 60 rotates via the belt 110. At this time, since the secondary movable sheave 63 of the secondary pulley 60 rotates together with the secondary fixed sheave 62, the secondary pulley shaft 61, and the pulley bearing 113, relative rotation occurs between the secondary movable sheave 63 and the intermediate member 67. . Then, from the state in which the first engagement portion 63b and the second engagement portion 67a shown in FIG. 5A approach each other, the plurality of transmission members 68 cause the first engagement portion 63b and the first engagement portion 63b as shown in FIG. The second engaging portion 67a changes to a separated state. As a result, the torque cam 66 generates a belt clamping pressure with respect to the belt 110 in the secondary pulley 60.

  That is, the secondary pulley 60 is provided with a torque cam 66 as a means for generating a belt clamping pressure with respect to the belt 110 in addition to the secondary hydraulic chamber 64 which is a clamping pressure generating hydraulic chamber. Here, the torque cam 66 mainly generates a belt clamping pressure, and the secondary hydraulic chamber 64 generates a shortage of the belt clamping pressure generated by the torque cam 66. Note that the secondary hydraulic chamber 64 alone may be used as the means for generating the belt clamping pressure with respect to the belt 110 in the secondary pulley 60.

  The belt 110 of the belt-type continuously variable transmission 1 </ b> A transmits the output torque from the internal combustion engine 10 transmitted through the primary pulley 50 to the secondary pulley 60. As shown in FIG. 1, the belt 110 is wound between the primary groove 110 a of the primary pulley 50 and the secondary groove 110 b of the secondary pulley 60. The belt 110 according to the first embodiment is an endless belt including, for example, a large number of metal pieces and a plurality of steel rings.

  Next, the final reduction gear 90 housed in the transaxle case 22 and the transaxle rear cover 23 will be described.

  The final speed reducer 90 transmits the output torque from the internal combustion engine 10 transmitted through the power transmission path 100 from the wheels 120 and 120 to the road surface. The final reduction gear 90 includes a differential case 91 having a hollow portion, a pinion shaft 92, differential pinions 93 and 94, and side gears 95 and 96.

  The differential case 91 is rotatably supported by bearings 97 and 98. A ring gear 99 is provided on the outer periphery of the differential case 91, and the ring gear 99 is engaged with the final drive pinion 105. The pinion shaft 92 is attached to the hollow portion of the differential case 91. The differential pinions 93 and 94 are rotatably attached to the pinion shaft 92. The side gears 95 and 96 are meshed with both of the differential pinions 93 and 94. The side gears 95 and 96 are fixed to one end portions of the drive shafts 121 and 122, respectively. Wheels 120 and 120 are attached to the other ends of the drive shafts 121 and 122, respectively.

  Here, the power transmission path 100 described above is disposed between the secondary pulley 60 and the final reduction gear 90, as shown in FIG. The power transmission path 100 includes an input shaft 101 on the same axis as the secondary pulley shaft 61, an intermediate shaft 102 parallel to the input shaft 101, a counter drive pinion 103, a counter driven gear 104, and a final drive pinion 105. It is comprised by. The input shaft 101 and the counter drive pinion 103 fixed to the input shaft 101 are rotatably held by bearings 118 and 119. Further, the intermediate shaft 102 is rotatably supported by bearings 116 and 117. The counter driven gear 104 is fixed to the intermediate shaft 102 and meshed with the counter drive pinion 103. The final drive pinion 105 is fixed to the intermediate shaft 102.

  Next, the hydraulic oil supply control device 130 will be described.

  The hydraulic oil supply control device 130 supplies hydraulic oil to a hydraulic oil supply portion that requires supply of hydraulic oil in a vehicle on which the belt type continuously variable transmission 1A and the internal combustion engine 10 are mounted. As shown in FIG. 6, the hydraulic oil supply control device 130 supplies hydraulic oil to the primary hydraulic chamber 55, the secondary hydraulic chamber 64, the drive hydraulic chamber 81, etc., and the hydraulic pressure, hydraulic oil supply flow rate, hydraulic oil The transmission ratio of the belt-type continuously variable transmission 1A is also controlled by controlling the discharge flow rate. In the figure, the hydraulic oil supply portion (excluding the above-described hydraulic oil supply portion and the hydraulic oil supply portion of the internal combustion engine 10 (for example, movable parts) is excluded from the primary hydraulic chamber 55, the secondary hydraulic chamber 64, and the drive hydraulic chamber 81. The illustration of a stationary part having a sliding part in between, a movable part or a movable part having a sliding part between stationary parts, a heated part or a driving device driven by oil) is omitted.

  As shown in FIG. 6, the hydraulic oil supply control device 130 includes an oil pan 131, an oil pump 132, a line pressure control device 133, a constant pressure control device 134, a primary hydraulic chamber control device 135, and a drive The hydraulic chamber control device 136 and the secondary hydraulic chamber control device 137 are configured.

  The oil pump 132 sucks, pressurizes and discharges the hydraulic oil stored in the oil pan 131. The oil pump 132 is connected to the line pressure control device 133 via the oil passage R1. Accordingly, the hydraulic oil pressurized and discharged by the oil pump 132 is supplied to the line pressure control device 133. That is, the discharge pressure Pout of the oil pump 132 is introduced into the line pressure control device 133. The oil pump 132 is disposed between the torque converter 30 and the forward / reverse switching mechanism 40 as shown in FIG. The oil pump 132 includes a rotor 132a, a hub 132b, and a body 132c. The oil pump 132 is connected to the pump 31 by a rotor 132a through a cylindrical hub 132b. The hub 132b is spline-fitted to the hollow shaft 36. The body 132c is fixed to the transaxle case 22. Therefore, the oil pump 132 can be driven by the output torque from the internal combustion engine 10 being transmitted to the rotor 132a via the pump 31. That is, the oil pump 132 increases the discharge amount of the hydraulic oil discharged according to the increase in the rotation speed of the internal combustion engine 10, that is, the discharge pressure Pout increases.

  Further, the line pressure control device 133 adjusts the discharge pressure Pout of the oil pump 132 so that it becomes a predetermined line pressure PL. That is, the line pressure control device 133 adjusts the discharge pressure Pout of the oil passage R1, which is an input hydraulic pressure, to the line pressure PL, and outputs the line pressure PL as an output hydraulic pressure to the oil passage R2. The line pressure control device 133 is connected to a second port 135l of a supply-side flow rate control valve 135c, which will be described later, of the primary hydraulic chamber control device 135 through the oil passage R2, and through the oil passage R2 and the branch oil passage R21. Are connected to the constant pressure control device 134, and are connected to the secondary hydraulic chamber control device 137 via the oil passage R2 and the branch oil passage R22. Accordingly, the line pressure PL adjusted by the line pressure control device 133 is introduced into the second port 135l of the supply side flow rate control valve 135c, the constant pressure control device 134, and the secondary hydraulic chamber control device 137. The line pressure control device 133 adjusts the line pressure PL according to the output torque of the internal combustion engine 10. The line pressure control device 133 is provided with an electromagnetic valve (not shown) such as a linear solenoid valve that regulates the pressure of the hydraulic oil discharged from the oil pump 132. The line pressure control device 133 is electrically connected to an ECU (Engine Control Unit) 140, and the valve opening degree of the linear solenoid valve is controlled by a control signal from the ECU 140, thereby adjusting the line pressure PL. be able to.

  The constant pressure control device 134 adjusts the line pressure PL output from the line pressure control device 133 so that the line pressure PL always becomes a constant pressure. That is, the constant pressure control device 134 adjusts the oil pressure PL of the oil passage R2, which is the input oil pressure, and sets the output oil pressure to the constant pressure PS. The constant pressure control device 134 is connected to a first port 135e of a supply-side control valve 135a, which will be described later, of the primary hydraulic chamber control device 135 through an oil passage R3, and is primary through an oil passage R3 and a branch oil passage R31. The hydraulic chamber controller 135 is connected to a first port 135h of a discharge side control valve 135b, which will be described later, and is connected to a drive hydraulic chamber controller 136 via an oil passage R3 and a branch oil passage R32. Accordingly, the constant pressure PS regulated by the constant pressure control device 134 is introduced into the first port 135e of the supply side control valve 135a, the first port 135h of the discharge side control valve 135b, and the drive hydraulic chamber control device 136. Is done.

  The primary hydraulic chamber control device 135 controls supply of hydraulic oil to the primary hydraulic chamber 55 or discharge of hydraulic fluid from the primary hydraulic chamber 55. The primary hydraulic chamber control device 135 according to the first embodiment controls the supply flow rate of the hydraulic oil supplied to the primary hydraulic chamber 55 and the discharge flow rate of the hydraulic oil discharged from the primary hydraulic chamber 55. A control valve 135a, a discharge-side control valve 135b, a supply-side flow control valve 135c, and a discharge-side flow control valve 135d are configured.

  The supply side control valve 135a controls the supply flow rate of the hydraulic oil supplied to the primary hydraulic chamber 55 by the supply side flow rate control valve 135c. The supply-side control valve 135a switches communication between three ports (that is, the first port 135e, the second port 135f, and the third port 135g) by ON / OFF. The first port 135e is connected to the constant pressure control device 134 as described above. The second port 135f is connected to a first port 135k, which will be described later, of the supply-side flow rate control valve 135c via an oil passage R4. The second port 135f is further connected to a later-described fourth port 135u of the discharge-side flow rate control valve 135d via an oil passage R4 and a branch oil passage R41. The third port 135g is connected to the oil pan 131 via the merging oil passage R51 and the oil passage R5. That is, the third port 135g is released to atmospheric pressure.

  When the supply-side control valve 135a is turned on, the first port 135e and the second port 135f communicate with each other. Accordingly, the constant pressure PS introduced into the supply side control valve 135a is introduced into the first port 135k of the supply side flow control valve 135c (see FIG. 8). That is, the constant pressure PS introduced into the supply side control valve 135a is introduced into a control hydraulic chamber 135o, which will be described later, of the supply side flow control valve 135c communicating with the first port 135k. Further, the constant pressure PS introduced into the supply side control valve 135a is introduced into the fourth port 135u of the discharge side flow control valve 135d (see the same figure). On the other hand, when the supply-side control valve 135a is turned off, the second port 135f and the third port 135g communicate with each other. Accordingly, the first port 135k of the supply side flow control valve 135c is released to the atmospheric pressure via the supply side control valve 135a (see FIG. 10). That is, the control hydraulic pressure chamber 135o is released to atmospheric pressure through the first port 135k of the supply side flow control valve 135c. Further, the fourth port 135u of the discharge side flow control valve 135d is released to the atmospheric pressure via the supply side control valve 135a (see FIG. 10). Here, as shown in FIG. 6, the supply-side control valve 135 a is electrically connected to the ECU 140 and is duty-controlled by a control signal from the ECU 140. Therefore, the supply-side control valve 135a can regulate the control hydraulic chamber 135o of the supply-side flow rate control valve 135c from a constant pressure PS to atmospheric pressure by a control signal from the ECU 140.

  Next, the discharge side control valve 135b constituting the primary hydraulic chamber control device 135 performs discharge flow control of the hydraulic oil discharged from the primary hydraulic chamber 55 by the discharge side flow control valve 135d. The discharge side control valve 135b switches communication of three ports (that is, the first port 135h, the second port 135i, and the third port 135j) by ON / OFF. The first port 135h is connected to the constant pressure control device 134 as described above. The second port 135i is connected to a later-described first port 135r of the discharge-side flow rate control valve 135d via the oil passage R6. The second port 135i is connected to a later-described fourth port 135n of the supply-side flow rate control valve 135c via the oil passage R6 and the branch oil passage R61. The third port 135j is connected to the oil pan 131 via the oil path R5. That is, the third port 135j is released to atmospheric pressure.

  When the discharge side control valve 135b is turned on, the first port 135h and the second port 135i communicate with each other. Accordingly, the constant pressure PS introduced into the discharge side control valve 135b is introduced into the first port 135r of the discharge side flow control valve 135d (see FIG. 10). That is, the constant pressure PS introduced into the discharge side control valve 135b is introduced into a later-described control hydraulic chamber 135v of the discharge side flow control valve 135d communicating with the first port 135r. Further, the constant pressure PS introduced into the discharge side control valve 135b is introduced into the fourth port 135n of the supply side flow control valve 135c (see the figure). On the other hand, when the discharge side control valve 135b is turned off, the second port 135i and the third port 135j communicate with each other. Therefore, the first port 135r of the discharge side flow control valve 135d is released to atmospheric pressure via the discharge side control valve 135b (see FIG. 8). That is, the control hydraulic chamber 135v is released to the atmospheric pressure via the first port 135r of the discharge side flow control valve 135d. Further, the fourth port 135n of the supply-side flow rate control valve 135c is released to atmospheric pressure via the discharge-side control valve 135b (see the same figure). Here, as shown in FIG. 6, the discharge-side control valve 135 b is electrically connected to the ECU 140 and is duty-controlled by a control signal from the ECU 140. Accordingly, the discharge-side control valve 135b can regulate the control hydraulic chamber 135v of the discharge-side flow control valve 135d from a constant pressure PS to atmospheric pressure by a control signal from the ECU 140.

  Next, the supply-side flow rate control valve 135 c constituting the primary hydraulic chamber control device 135 controls the supply flow rate of the hydraulic oil supplied to the primary hydraulic chamber 55. The supply-side flow control valve 135c includes a first port 135k, a second port 135l, a third port 135m, a fourth port 135n, a control hydraulic chamber 135o, a spool 135p, and a spool elastic member 135q. It is configured. As described above, the first port 135k is connected to the second port 135f of the supply side control valve 135a. The second port 135l is connected to the line pressure control device 133 as described above. The third port 135m is connected to the primary hydraulic chamber 55 via an oil passage R7. The third port 135m of the first embodiment has a primary hydraulic chamber via the oil passage R7, the supply / discharge side main passage 51a, the shaft side communication passage 51c, the spaces T1 and T2, the partition wall side communication passage 54b, and the valve arrangement passage 54a. 55 is connected. The fourth port 135n is connected to the second port 135i of the discharge side control valve 135b as described above. In addition, as shown in the figure, between the second port 135f of the supply-side control valve 135a and the first port 135k of the supply-side flow control valve 135c, the line pressure control device 133 and the second port of the supply-side flow control valve 135c An orifice, that is, a throttle, is provided between the port 135l and the hydraulic oil flowing from the supply-side control valve 135a to the supply-side flow control valve 135c and the hydraulic oil flowing from the line pressure control device 133 to the supply-side flow control valve 135c. The pressure or flow rate may be adjusted.

  The control hydraulic chamber 135o communicates with the first port 135k, and a spool valve opening direction push that presses the spool 135p in one direction (upward in the figure) of the direction in which the spool 135p moves by the hydraulic pressure. The pressure is applied to the spool 135p. The spool 135p is movably supported in the primary hydraulic chamber control device 135, and moves in one of the moving directions to communicate the second port 135l and the third port 135m. By moving in the other direction of the moving direction, the communication between the second port 135l and the third port 135m is blocked. The spool elastic member 135q is arranged in a state of being biased between the spool 135p and a member stationary with respect to the spool 135p. Accordingly, the spool elastic member 135q generates a spool urging force, and the spool urging force presses the spool 135p in the other direction (downward in the figure) in the direction in which the spool 135p moves. The valve direction pressing force is applied to the spool 135p.

  In the supply-side flow rate control valve 135c, the spool valve opening direction pressing force acting on the spool 135p exceeds the spool closing direction pressing force, so that the spool 135p moves in one of the moving directions. Here, the supply-side flow rate control valve 135c has a degree of communication between the second port 135l and the third port 135m, that is, the second port as the movement amount of the spool 135p increases in one direction. The flow path cross-sectional area of the flow path that connects 135l and the third port 135m increases. In other words, the supply-side flow rate control valve 135c has two ports (ie, the second port 135l and the third port) as the spool 135p moves by the hydraulic pressure in the control hydraulic chamber 135o regulated by the supply-side control valve 135a. 135m), and the supply flow rate is controlled.

  The discharge side flow control valve 135d controls the discharge flow rate of the hydraulic oil discharged from the primary hydraulic chamber 55. The discharge side flow control valve 135d includes a first port 135r, a second port 135s, a third port 135t, a fourth port 135u, a control hydraulic chamber 135v, a spool 135w, and a spool elastic member 135x. It is configured. The first port 135r is connected to the second port 135i of the discharge side control valve 135b as described above. The second port 135s is connected to the oil pan 131 via the merging oil path R52, the merging oil path R51, and the oil path R5. That is, the second port 135s is released to atmospheric pressure. The third port 135t is connected to the primary hydraulic chamber 55 via the branch oil passage R71 and the oil passage R7. The third port 135t of the first embodiment includes a branch oil passage R71, an oil passage R7, a supply / discharge side main passage 51a, a shaft side communication passage 51c, spaces T1 and T2, a partition wall side communication passage 54b, and a valve arrangement passage 54a. Via the primary hydraulic chamber 55. As described above, the fourth port 135u is connected to the second port 135f of the supply side control valve 135a. In addition, as shown in the figure, an orifice, that is, a throttle is provided between the second port 135i of the discharge side control valve 135b and the first port 135r of the discharge side flow control valve 135d, and discharged from the discharge side control valve 135b. The pressure or flow rate of the hydraulic oil flowing into the side flow rate control valve 135d may be adjusted.

  The control hydraulic chamber 135v communicates with the first port 135r, and a spool valve opening direction push that pushes the spool 135w in one direction (upward in the figure) of the direction in which the spool 135w moves by the hydraulic pressure. The pressure is applied to the spool 135w. The spool 135w is movably supported in the primary hydraulic chamber control device 135, and moves in one of the moving directions so as to communicate the second port 135s and the third port 135t. By moving in the other direction, the communication between the second port 135s and the third port 135t is blocked. The spool elastic member 135x is arranged in a state of being biased between the spool 135w and a member stationary with respect to the spool 135w. Therefore, the spool elastic member 135x generates a spool urging force, and the spool urging force presses the spool 135w in the other direction (downward in the figure) of the direction in which the spool 135w moves. The valve direction pressing force is applied to the spool 135w.

  The discharge-side flow rate control valve 135d moves the spool 135w in one of the moving directions when the spool valve opening direction pressing force acting on the spool 135w exceeds the spool closing direction pressing force. Here, the discharge-side flow rate control valve 135d has a degree of communication between the second port 135s and the third port 135t, that is, the second port 135s as the moving amount of the spool 135w in one direction increases. And the channel cross-sectional area of the channel communicating with the third port 135t increases. That is, the discharge-side flow rate control valve 135d has two ports (ie, the second port 135s and the third port) as the spool 135w moves by the hydraulic pressure of the control hydraulic chamber 135v regulated by the discharge-side control valve 135b. 135t), and the supply flow rate is controlled.

  The drive hydraulic chamber control device 136 adjusts the hydraulic pressure P2 of the drive hydraulic chamber 81, and is connected to the drive hydraulic chamber 81 via an oil passage R8. As described above, the constant pressure PS is introduced into the drive hydraulic chamber control device 136 from the constant pressure control device 134. The drive hydraulic chamber control device 136 according to the first embodiment is connected to the drive hydraulic chamber 81 via an oil passage R8, a drive side main passage 51b, a shaft side communication passage 51d, and a drive communication passage 54e. The drive hydraulic chamber control device 136 includes a switching valve (not shown). The drive hydraulic chamber control device 136 is electrically connected to the ECU 140 and controls ON / OFF of the ON / OFF valve by a control signal from the ECU 140. When the drive hydraulic chamber control device 136 is ON-controlled (that is, the switching valve is turned on), the branch oil passage R32 and the oil passage R8 communicate with each other and are introduced into the drive hydraulic chamber control device 136. The constant pressure PS is introduced into the drive hydraulic chamber 81, and the hydraulic pressure P2 in the drive hydraulic chamber 81 becomes the constant pressure PS. On the other hand, when the drive hydraulic chamber control device 136 is controlled to be OFF (that is, the switching valve is turned OFF), the communication between the branch oil path R32 and the oil path R8 is cut off and the oil path R8 is externally connected. The hydraulic pressure P2 in the drive hydraulic chamber 81 becomes atmospheric pressure. Here, the constant pressure PS at that time is that each hydraulic oil supply / discharge valve 70 is opened by the hydraulic pressure P2 of the drive hydraulic chamber 81 at least when the hydraulic pressure P2 of the drive hydraulic chamber 81 becomes the constant pressure PS. It is a pressure that exceeds the hydraulic pressure that can be generated.

  The secondary hydraulic chamber control device 137 controls the supply of hydraulic oil to the secondary hydraulic chamber 64 or the discharge of hydraulic oil from the secondary hydraulic chamber 64, and is connected to the secondary hydraulic chamber 64 via an oil passage R9. It is connected. As described above, the line pressure PL is introduced into the secondary hydraulic chamber control device 137 from the line pressure control device 133. The secondary hydraulic chamber control device 137 of the first embodiment is connected to the secondary hydraulic chamber 64 via an oil passage R9, a hydraulic oil passage (not shown) of the secondary pulley shaft 61, and a hydraulic fluid supply hole (not shown). The secondary hydraulic chamber control device 137 includes a flow rate control valve (not shown). The secondary hydraulic chamber control device 137 is electrically connected to the ECU 140, and regulates the introduced line pressure PL controlled by a control signal from the ECU 140.

  As described above, the hydraulic oil supply control device 130 is connected to at least the ECU 140 that controls the operation of the internal combustion engine 10. Therefore, the hydraulic oil supply control device 130 controls the primary hydraulic chamber control device 135, the drive hydraulic chamber control device 136, and the secondary hydraulic chamber control device 137 based on the control signal from the ECU 140. At least the gear ratio of the belt type continuously variable transmission 1A is controlled.

  The operation of the belt type continuously variable transmission 1A according to the first embodiment will be described below.

  General vehicle forward and reverse will be described. When the driver selects a forward position using a shift position device (not shown) provided on the vehicle, the ECU 140 turns on the forward clutch 42 and turns off the reverse brake 43 by the hydraulic oil supplied from the hydraulic oil supply control device 130. Thus, the forward / reverse switching mechanism 40 is controlled. As a result, the input shaft 38 and the primary pulley shaft 51 are directly connected. That is, the sun gear 44 and the ring gear 46 of the planetary gear device 41 are directly connected, the primary pulley shaft 51 is rotated in the same direction as the rotation direction of the crankshaft 11 of the internal combustion engine 10, and the output torque from the internal combustion engine 10 is converted to the primary pulley. 50. The output torque from the internal combustion engine 10 transmitted to the primary pulley 50 is transmitted to the secondary pulley 60 via the belt 110 and rotates the secondary pulley shaft 61 of the secondary pulley 60.

  The output torque of the internal combustion engine 10 transmitted to the secondary pulley 60 is transmitted from the intermediate member 67 to the intermediate shaft 102 via the input shaft 101 of the power transmission path 100, the counter drive pinion 103, and the counter driven gear 104. The shaft 102 is rotated. The output torque transmitted to the intermediate shaft 102 is transmitted to the differential case 91 of the final reduction gear 90 via the final drive pinion 105 and the ring gear 99, and the differential case 91 is rotated. The output torque from the internal combustion engine 10 transmitted to the differential case 91 is transmitted to the drive shafts 121 and 122 via the differential pinions 93 and 94 and the side gears 95 and 96, and the wheels 120 and 120 attached to the ends thereof. Is transmitted to. As a result, the wheels 120 and 120 are rotated with respect to a road surface (not shown), so that the vehicle moves forward.

  On the other hand, when the driver selects a reverse drive position using a shift position device (not shown) provided in the vehicle, the ECU 140 turns off the forward clutch 42 with the hydraulic oil supplied from the hydraulic oil supply control device 130, and reverse brakes. 43 is turned ON, and the forward / reverse switching mechanism 40 is controlled. As a result, the switching carrier 47 of the planetary gear unit 41 is fixed to the transaxle case 22 and is held by the switching carrier 47 so that each pinion 45 only rotates. Accordingly, the ring gear 46 rotates in the same direction as the input shaft 38, and each pinion 45 meshed with the ring gear 46 also rotates in the same direction as the input shaft 38, and the sun gear 44 meshed with each pinion 45 becomes the input. It rotates in the opposite direction to the shaft 38. That is, the primary pulley shaft 51 connected to the sun gear 44 rotates in the direction opposite to the input shaft 38. As a result, the secondary pulley shaft 61, the input shaft 101, the intermediate shaft 102, the differential case 91, the drive shafts 121, 122, and the like of the secondary pulley 60 rotate in the opposite direction to the case where the driver selects the forward position. Goes backwards.

  Here, the ECU 140 includes a map (for example, an optimal fuel consumption curve based on the engine speed and the throttle opening of the throttle valve, etc.) stored in the storage unit of the ECU 140 and conditions such as the vehicle speed and the accelerator opening of the driver. ), The gear ratio of the belt-type continuously variable transmission 1A is controlled via the hydraulic oil supply control device 130 so that the operating state of the internal combustion engine 10 is optimized. The control of the gear ratio of the belt type continuously variable transmission 1A includes changing the gear ratio and fixing the gear shift (gear ratio γ steady). The change of the gear ratio and the fixing of the gear ratio are performed by controlling the primary hydraulic chamber control device 135, the drive hydraulic chamber control device 136, and the secondary hydraulic chamber control device 137.

  Regarding the change of the gear ratio, the supply of hydraulic oil from the hydraulic oil supply control device 130 to the primary hydraulic chamber 55 or the primary hydraulic chamber 55 via the hydraulic oil supply control device 130 to the outside of the primary pulley 50 is mainly used. This is done by discharging hydraulic oil. That is, when the hydraulic oil is supplied or discharged, the primary movable sheave 53 slides in the axial direction of the primary pulley shaft 51 according to the supply or discharge, and the primary movable sheave 53 and the primary fixed sheave are fixed. The distance between the sheave 52, that is, the width of the primary groove 110a is adjusted. As a result, the contact radius of the belt 110 in the primary pulley 50 changes, and the speed ratio, which is the ratio between the rotation speed of the primary pulley 50 and the rotation speed of the secondary pulley 60, is controlled steplessly (continuously). On the other hand, the gear ratio is fixed mainly by holding hydraulic oil in the primary hydraulic chamber 55.

  In the secondary pulley 60, the ECU 140 controls the secondary hydraulic chamber control device 137 so that the hydraulic pressure in the secondary hydraulic chamber 64 is regulated, and the belt 110 is sandwiched between the secondary fixed sheave 62 and the secondary movable sheave 63. The belt clamping pressure to be applied is adjusted. Thereby, the belt tension of the belt 110 wound around the primary pulley 50 and the secondary pulley 60 is controlled.

  Here, the change in the gear ratio includes an upshift (that is, a gear ratio decrease change that decreases the gear ratio) and a downshift (that is, a gear ratio increase change that increases the gear ratio). Each change operation will be described below.

  First, the gear ratio reduction changing operation will be described with reference to FIGS.

  The gear ratio reduction change is performed by supplying hydraulic oil from the hydraulic oil supply control device 130 to the primary hydraulic chamber 55 and sliding (moving) the primary movable sheave 53 toward the primary fixed sheave 52 side. Therefore, when changing the gear ratio reduction, as shown in FIG. 7, each hydraulic oil supply / discharge valve 70 is forcibly opened by the actuator 80, and is operated from the hydraulic oil supply control device 130 to the primary hydraulic chamber 55. Supply oil. At that time, the ECU 140 calculates a reduction gear ratio and a transmission speed, and outputs a control signal for the transmission ratio based on these to the hydraulic oil supply control device 130.

  Specifically, the drive hydraulic chamber control device 136 is ON-controlled by the ECU 140. Accordingly, the constant pressure PS introduced into the drive hydraulic chamber control device 136 is introduced into the drive hydraulic chamber 81, and the hydraulic pressure P2 of the drive hydraulic chamber 81 becomes the constant pressure PS. As a result, the actuator 80 causes the valve opening direction pressing force acting on the piston 82 by the hydraulic pressure P2 of the drive hydraulic chamber 81 to act on the valve bodies 71 of the respective hydraulic oil supply / discharge valves 70 as the valve body opening direction pressing force. . Here, as described above, the actuator 80 can open each hydraulic oil supply / discharge valve 70 by the hydraulic pressure P2 when the hydraulic pressure P2 of the drive hydraulic chamber 81 becomes a constant pressure PS. The valve direction pressing force exceeds the valve closing direction pressing force in the opposite direction. Accordingly, in each hydraulic oil supply / discharge valve 70, as shown in FIG. 7, the valve element 71 moves in a direction (opening direction) away from the valve seat surface 72 by the action of the actuator 80 to open the valve. I do.

  The supply-side control valve 135a of the primary hydraulic chamber control device 135 is duty-controlled by the ECU 140 to control the supply flow rate of hydraulic oil to the primary hydraulic chamber 55 by the supply-side flow rate control valve 135c. When duty control is performed by the ECU 140, the supply-side control valve 135a repeats ON and OFF, as shown in FIG. 8, and adjusts the control hydraulic pressure in the control hydraulic chamber 135o of the supply-side flow rate control valve 135c to a predetermined pressure during supply. And a predetermined pressure at the time of supply is introduced into the fourth port 135u of the discharge side flow control valve 135d. Here, the predetermined pressure at the time of supply refers to a pressure relating to a supply flow rate controlled by controlling the communication state between the second port 135l and the third port 135m by the spool valve opening direction pressing force acting on the spool 135p. This is the pressure at which the supply flow rate can be set to the supply flow rate based on the reduction gear ratio and the transmission speed. Therefore, in the supply-side flow rate control valve 135c, the spool valve opening direction pressing force based on the control hydraulic pressure of the control hydraulic chamber 135o (that is, the above-mentioned predetermined pressure at the time of supply) exceeds the spool valve closing direction pressing force. As shown by the arrow A, the spool 135p moves in a direction in which the second port 135l and the third port 135m communicate with each other. As a result, the supply-side flow rate control valve 135c is opened, and the supply flow rate of hydraulic oil to the primary hydraulic chamber 55 becomes a supply flow rate based on the reduction gear ratio and the shift speed.

  On the other hand, the discharge side control valve 135b of the primary hydraulic chamber control device 135 is duty-controlled by the ECU 140 to control the discharge flow rate of the hydraulic oil from the primary hydraulic chamber 55 by the discharge side flow rate control valve 135d. When the exhaust control valve 135b is duty-controlled by the ECU 140, as shown in FIG. 8, the discharge side control valve 135b is kept OFF, and the fourth port 135n of the supply side flow control valve 135c and the control hydraulic chamber of the discharge side flow control valve 135d. Release 135v to atmospheric pressure. Accordingly, in the discharge-side flow control valve 135d, only the spool closing direction pressing force acts on the spool 135w, so that the spool 135w is held at a position that maintains the shut-off state between the second port 135s and the third port 135t. The As a result, the discharge-side flow rate control valve 135d remains closed, and the hydraulic oil discharge flow rate from the primary hydraulic chamber 55 becomes zero.

  At this time, as described above, each hydraulic oil supply / discharge valve 70 is forcibly opened by the actuator 80. Accordingly, the hydraulic oil introduced into the supply-side flow control valve 135c at the line pressure PL (when an orifice is provided between the line pressure control device 133 and the second port 135l of the supply-side flow control valve 135c, the line pressure The hydraulic fluid inserted with the pressure adjusted from PL) is controlled by the supply-side flow rate control valve 135c to the supply flow rate based on the reduction gear ratio and the shift speed, and as shown by the arrow U in FIG. It flows into the supply / discharge side main passage 51a through the path R7. The hydraulic oil flowing into the supply / discharge side main passage 51a enters the primary hydraulic chamber 55 from the supply / discharge side main passage 51a through the shaft side communication passage 51c, the spaces T1 and T2, the partition wall side communication passage 54b, and the valve arrangement passage 54a. Supplied. In other words, in the belt type continuously variable transmission 1A of the first embodiment, when supplying hydraulic oil to the primary hydraulic chamber 55 that is one clamping pressure generating hydraulic chamber, the hydraulic oil discharge valve, the hydraulic oil supply valve, It is not necessary to open each hydraulic oil supply / discharge valve 70 designed to be shared by the supply pressure of the hydraulic oil supplied to the primary hydraulic chamber 55. Therefore, in the first embodiment, when the supply pressure of the hydraulic oil supplied to the primary hydraulic chamber 55 is increased, the line pressure PL introduced into the supply-side flow rate control valve 135c by the line pressure control device 133 is increased. Can be suppressed.

  In the belt-type continuously variable transmission 1A at this time, the hydraulic oil P1 in the primary hydraulic chamber 55 is raised by the hydraulic oil supplied as the hydraulic oil supply / discharge valves 70 are opened, and the primary movable sheave 53 is moved to the primary fixed sheave 53. Since the pressing force to move toward the 52 side increases, the primary movable sheave 53 slides in the axial direction toward the primary fixed sheave 52 side. As a result, the contact radius of the belt 110 in the primary pulley 50 is increased, the contact radius of the belt 110 in the secondary pulley 60 is decreased, the transmission ratio is decreased, and the reduced transmission ratio is obtained.

  Subsequently, the gear ratio increase changing operation will be described with reference to FIGS. 9 and 10.

  The change in speed ratio is changed by discharging hydraulic oil from the primary hydraulic chamber 55 via the hydraulic oil supply control device 130 and sliding (moving) the primary movable sheave 53 to the side opposite to the primary fixed sheave 52 side. Is done by. Therefore, when changing the speed ratio, as shown in FIG. 9, each hydraulic oil supply / discharge valve 70 is forcibly opened by the actuator 80 and the hydraulic oil is discharged from the primary hydraulic chamber 55. At that time, the ECU 140 calculates the increased gear ratio and the gear speed, and outputs a gear ratio control signal based on these to the hydraulic oil supply control device 130.

  Specifically, the drive hydraulic chamber control device 136 is ON-controlled by the ECU 140 in the same manner as when the gear ratio decrease is changed. Accordingly, in each hydraulic oil supply / discharge valve 70, as shown in FIG. 9, the valve element 71 is separated from the valve seat surface 72 by the action of the actuator 80 in the same manner as in the case of the gear ratio reduction change described above ( Moves in the valve opening direction) and opens the valve.

  The supply-side control valve 135a of the primary hydraulic chamber control device 135 is duty-controlled by the ECU 140 to control the supply flow rate of hydraulic oil to the primary hydraulic chamber 55 by the supply-side flow rate control valve 135c. When duty control is performed by the ECU 140, the supply side control valve 135a maintains OFF as shown in FIG. 10, and the control hydraulic chamber 135o of the supply side flow rate control valve 135c and the fourth port 135u of the discharge side flow rate control valve 135d. To atmospheric pressure. Therefore, in the supply-side flow rate control valve 135c, only the spool closing direction pressing force acts on the spool 135p, so that the spool 135p is held at a position that maintains the cutoff state between the second port 135l and the third port 135m. The As a result, the supply-side flow rate control valve 135c remains closed, and the supply flow rate of hydraulic oil to the primary hydraulic chamber 55 becomes zero.

  On the other hand, the discharge side control valve 135b of the primary hydraulic chamber control device 135 is duty-controlled by the ECU 140 to control the discharge flow rate of the hydraulic oil from the primary hydraulic chamber 55 by the discharge side flow rate control valve 135d. When duty control is performed by the ECU 140, the discharge-side control valve 135b repeats ON and OFF, as shown in FIG. 10, and introduces a predetermined pressure during discharge into the fourth port 135n of the supply-side flow rate control valve 135c. The control hydraulic pressure of the control hydraulic chamber 135v of the side flow rate control valve 135d is adjusted to a predetermined pressure at the time of discharge. Here, the predetermined pressure at the time of discharge is a pressure relating to the discharge flow rate controlled by controlling the communication state between the second port 135s and the third port 135t by the spool valve opening direction pressing force acting on the spool 135w. It is the pressure that can make the discharge flow rate a discharge flow rate based on the increase gear ratio and the shift speed. Accordingly, in the discharge side flow control valve 135d, the spool valve opening direction pressing force based on the control hydraulic pressure of the control hydraulic chamber 135v (that is, the above-described predetermined pressure at the time of discharging) exceeds the spool closing direction pressing force. As shown by the arrow C, the spool 135w moves in a direction in which the second port 135s and the third port 135t are communicated with each other. As a result, the discharge-side flow rate control valve 135d is opened, and the hydraulic oil discharge flow rate from the primary hydraulic chamber 55 becomes a discharge flow rate based on the increased gear ratio and the shift speed.

  At this time, as described above, each hydraulic oil supply / discharge valve 70 is forcibly opened by the actuator 80. Accordingly, the hydraulic oil in the primary hydraulic chamber 55 passes from the primary hydraulic chamber 55 to the valve arrangement passage 54a, the partition wall side communication passage 54b, the space portions T1 and T2, and the shaft side communication passage 51c as shown by an arrow D in FIG. To the supply / discharge side main passage 51a. The hydraulic oil in the primary hydraulic chamber 55 that has flowed into the supply / discharge-side main passage 51a flows into the discharge-side flow rate control valve 135d via the oil passage R7 and the branch oil passage R71, and is increased by the discharge-side flow rate control valve 135d. It is controlled to a discharge flow rate based on the transmission gear ratio and the transmission speed, and is discharged to the outside of the oil pan 131, that is, the primary hydraulic chamber 55 through the merged oil passages R 52 and R 51 and the oil passage R 5.

  In the belt-type continuously variable transmission 1A at this time, the hydraulic oil is discharged from the primary hydraulic chamber 55 via each hydraulic oil supply / discharge valve 70, whereby the hydraulic pressure P1 of the primary hydraulic chamber 55 decreases, and the primary hydraulic chamber 55 Since the pressing force for moving the movable sheave 53 to the primary fixed sheave 52 side decreases, the primary movable sheave 53 slides in the axial direction opposite to the primary fixed sheave 52 side. Thereby, the contact radius of the belt 110 in the primary pulley 50 decreases, the contact radius of the belt 110 in the secondary pulley 60 increases, the transmission ratio is increased, and the increased transmission ratio is obtained.

  Next, the speed ratio fixing operation will be described with reference to FIGS.

  The transmission gear ratio is fixed (that is, the transmission gear ratio is made steady) without supplying the hydraulic oil to the primary hydraulic chamber 55 and discharging the hydraulic oil from the primary hydraulic chamber 55 without fixing the primary movable sheave 53. This is done by making the position in the axial direction relative to the sheave 52 constant. That is, when the transmission ratio is fixed, the movement of the primary movable sheave 53 in the axial direction relative to the primary fixed sheave 52 is restricted. Note that the gear ratio is fixed when the ECU 140 determines that there is no need to change the gear ratio significantly, such as when the running state of the vehicle is stable.

  When the transmission gear ratio is fixed, as shown in FIG. 2, each hydraulic oil supply / discharge valve 70 is closed, the supply of hydraulic oil to the primary hydraulic chamber 55 via each hydraulic oil supply / discharge valve 70, and each hydraulic oil Discharging hydraulic oil from the primary hydraulic chamber 55 via the supply / discharge valve 70 is prohibited. At that time, ECU 140 outputs a control signal based on the fixed gear ratio to hydraulic oil supply control device 130.

  Specifically, the drive hydraulic chamber control device 136 is OFF-controlled by the ECU 140. For this reason, the drive hydraulic chamber 81 is released to atmospheric pressure, and the hydraulic pressure P2 of the drive hydraulic chamber 81 becomes atmospheric pressure. Therefore, the piston 82 is subjected to a piston closing direction pressing force by the valve body elastic member 73 via the pressing force transmitting member 74, and further, the roller of the piston initial position holding means is caused by the centrifugal force accompanying the rotation of the primary pulley 50. The centrifugal hydraulic drag from 85 acts. For this reason, the piston 82 slides in the valve closing direction. As a result, no valve body opening direction pressing force acts on the valve body 71 of each hydraulic oil supply / discharge valve 70, and only the valve body closing direction pressing force acts. Each valve element 71 moves in the valve closing direction and contacts the valve seat surface 72 to close the valve 70.

  The supply-side control valve 135a of the primary hydraulic chamber control device 135 is duty-controlled by the ECU 140 to control the supply flow rate of hydraulic oil to the primary hydraulic chamber 55 by the supply-side flow rate control valve 135c. When duty control is performed by the ECU 140, the supply-side control valve 135a maintains OFF as shown in FIG. 6, and the control hydraulic chamber 135o of the supply-side flow rate control valve 135c and the fourth port 135u of the discharge-side flow rate control valve 135d. To atmospheric pressure. Therefore, in the supply-side flow rate control valve 135c, only the spool closing direction pressing force acts on the spool 135p, so that the spool 135p is held at a position that maintains the cutoff state between the second port 135l and the third port 135m. The As a result, the supply-side flow rate control valve 135c is kept closed, and the supply flow rate of hydraulic oil to the primary hydraulic chamber 55 becomes zero, so that the supply to the primary hydraulic chamber 55 via each hydraulic oil supply / discharge valve 70 is performed. Supply of hydraulic oil is prohibited.

  On the other hand, the discharge side control valve 135b of the primary hydraulic chamber control device 135 is duty-controlled by the ECU 140 to control the discharge flow rate of the hydraulic oil from the primary hydraulic chamber 55 by the discharge side flow rate control valve 135d. When duty control is performed by the ECU 140, the discharge side control valve 135b maintains OFF as shown in FIG. 6, and the fourth port 135n of the supply side flow rate control valve 135c and the control hydraulic chamber 135v of the discharge side flow rate control valve 135d. To atmospheric pressure. Accordingly, in the discharge-side flow control valve 135d, only the spool closing direction pressing force acts on the spool 135w, so that the spool 135w is held at a position that maintains the shut-off state between the second port 135s and the third port 135t. The As a result, the discharge-side flow rate control valve 135d is kept closed, and the discharge flow rate of the hydraulic oil from the primary hydraulic chamber 55 becomes 0. Therefore, the discharge flow rate control valve 135d from the primary hydraulic chamber 55 via each hydraulic oil supply / discharge valve 70 Hydraulic oil discharge is prohibited.

  By the way, since the belt tension of the belt 110 changes even when the gear ratio is fixed, the contact radius of the belt 110 in the primary pulley 50 tends to change at that time, and the position of the primary movable sheave 53 in the axial direction with respect to the primary fixed sheave 52 is changed. May be changed. However, in the belt-type continuously variable transmission 1A according to the first embodiment, the supply of hydraulic oil to the primary hydraulic chamber 55 and the discharge of hydraulic oil from the primary hydraulic chamber 55 are prohibited, so that the primary hydraulic chamber 55 The hydraulic oil in the hydraulic chamber 55 is held. Furthermore, in the first embodiment, the pressing force in the valve opening direction of each hydraulic oil supply / discharge valve 70 due to the centrifugal hydraulic pressure of the drive hydraulic chamber 81 can be canceled by the mechanical pressing force of the piston initial position holding means. As a result, the hydraulic oil in the primary hydraulic chamber 55 is reliably held. Therefore, in this belt type continuously variable transmission 1A, when the position of the primary movable sheave 53 in the axial direction relative to the primary fixed sheave 52 is changed, the hydraulic pressure P1 of the primary hydraulic chamber 55 changes, but this primary hydraulic chamber 55 Therefore, the position of the primary movable sheave 53 in the axial direction with respect to the primary fixed sheave 52 is kept constant. For this reason, in the belt type continuously variable transmission 1A of the first embodiment, hydraulic oil is supplied to the primary hydraulic chamber 55 in order to keep the position of the primary movable sheave 53 in the axial direction relative to the primary fixed sheave 52 constant. Thus, it is not necessary to increase the hydraulic pressure P1 of the primary hydraulic chamber 55. As a result, when the transmission gear ratio is fixed, it is not necessary to drive the oil pump 132 for supplying hydraulic oil to the primary hydraulic chamber 55, so that an increase in driving loss of the oil pump 132 can be suppressed.

Here, the piston 82 of the actuator 80 of the first embodiment described above may be changed from the piston main body 82a to the piston main body 182a shown in FIG. The piston main body 182a shown in FIG. 11 is in contact with the contact surface 182a ₁ with the pressing portion 82b of the piston 82 and the roller 85 of the piston initial position holding means on the surface facing the primary partition wall 54 in the piston main body 82a described above. The shape is such that a recess 182a ₃ is provided except for the surface 182a ₂ . As a result, the hydraulic fluid leaking from the space between the sliding support hole 54c and the pressing portion 82b of the piston 82 toward the piston main body 182a easily flows into the centrifugal hydraulic pressure canceling chamber 84, and further, the primary partition wall in the concave portion 182a ₃ The hydraulic oil is retained in the closed space between the cover 54 and the cover member 83, and it is possible to avoid running out of the hydraulic oil in the centrifugal hydraulic pressure cancellation chamber 84. Therefore, by using such a piston main body 182a, the smooth movement of the roller 85 and the reliability of the piston initial position holding means can be significantly improved as compared with the above-described piston main body 82a. Further, since the piston main body 182a is lightened by the concave portion 182a ₃ , the responsiveness when the actuator 80 is operated can be further improved.

Further, the piston 82 of the first embodiment may be changed to that shown in FIG. The piston 282 shown in FIG. 12 is extended according to the piston main body 282a in which a cylindrical concave portion 282a ₁ is formed on the contact surface of the piston main body 82a with the pressing portion 82b, and the depth of the concave portion 282a ₁ . And a pressing portion 282b. The pressing portion 282b may be fitted and fixed to the concave portion 282a ₁ , for example, or may be formed to have an outer diameter that can be detached from the concave portion 282a ₁ . Here, there is no problem in the former case, but in the latter case, the depth of the concave portion 282a ₁ and the length of the extension of the pressing portion 282b are separated from each other by at least reciprocating movement of the piston main body 282a in the axial direction. Set to not. Accordingly, the pressing portion 282b functions as a rotation stopper for the piston main body 282a. Therefore, according to this piston 282, unnecessary relative rotation of the piston main body 282a with respect to the primary pulley 50 can be prevented, so that wear of the drive hydraulic chamber seal member S2 can be reduced, and durability and oil in the drive hydraulic chamber 81 can be reduced. Reliability can be improved by improving the density. When the shape of the piston main body 182a of FIG. 11 described above is applied to the piston 282, the pressing portion 282b falls into the concave portion 182a _{3 of} FIG. 11 as the piston main body 282a rotates, and the hydraulic oil supply / discharge valve The inconvenience that 70 cannot be opened can be eliminated. In the latter case, in order to improve the valve opening response of the hydraulic oil supply / discharge valve 70, it is preferable that both ends of the pressing portion 282b are in contact with the bottom surface of the valve body 71 and the recess 282a _{1 at} the initial position. .

Further, the rotation preventing means of the piston main body 282a similar to the piston 282 may be configured as follows. This is configured by replacing the piston main body 82a with the piston main body 382a shown in FIG. Piston main 382a shown in FIG. 13, which has a recess 382a ₂ on the contact surface of the roller 85 is obtained by so as to roll the rollers 85 in the recess 382a within _2. Here, the piston main body 182a shown in FIG. 11 is illustrated with its recess 382a ₂ formed. That is, the piston main body 382a shown in FIG. 13 has a concave portion 382a ₂ formed in the contact surface 182a ₂ with the roller 85 in the piston main body 182a of FIG. The roller 85 uses the piston main body 382a to contact the bottom surface of the recess 382a ₂ to generate centrifugal hydraulic drag, and one end of the roller 85 contacts the side surface of the recess 382a ₂ to rotate the piston main body 382a. Lock. Even if such a piston main body 382a is used, the same effect as the piston 282 of FIG. The concave portion 382a ₂ can also be applied to the piston main body 282a of the piston 282.

  In the first embodiment, each hydraulic oil supply / discharge valve 70 is provided in the primary partition wall 54. However, each hydraulic oil supply / discharge valve 70 only needs to be able to rotate integrally with the primary pulley 50, and is not necessarily limited to this mode. Is not to be done. For example, each hydraulic oil supply / discharge valve 70 may be provided in the primary pulley shaft 51 and the primary fixed sheave 52.

  In the first embodiment, the hydraulic oil supply valve and the hydraulic oil discharge valve are made the same by each hydraulic oil supply / discharge valve 70, but the hydraulic oil supply valve and the hydraulic oil discharge valve may be provided separately. .

  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 1B of the second embodiment is a modification of the valve-closing state holding means (piston initial position holding means) of the belt-type continuously variable transmission 1A exemplified in the first embodiment. Other than this, the configuration is the same as that of the belt type continuously variable transmission 1A of the first embodiment. In the second embodiment, on the basis of the belt type continuously variable transmission 1A having a piston initial position holding means including a centrifugal hydraulic pressure canceling chamber 84 and a centrifugal hydraulic pressure canceling member (roller) 85 shown in FIG. Illustrate.

Generally, when long term belt-type continuously variable transmission 1B has not been operated, since the hydraulic oil of the centrifugal hydraulic cancel chamber 84 will cut off, and the roller 85 and the centrifugal hydraulic drag generating surface 82a ₁ and the inclined surface 84a The sliding resistance between the two increases. For this reason, in such a case, the roller 85 cannot roll smoothly at the start of operation, so that an accurate centrifugal hydraulic drag cannot be applied to the piston main body 82a of the piston 82, and the hydraulic oil supply / discharge valve 70 is opened. There is a risk of letting you speak. Further, in such a case, the frictional resistance between the roller 85 and the respective sliding surfaces is increased, so that wear may occur between them and the durability may be deteriorated. At the start of operation in this case, since the viscosity of the hydraulic oil is high when the belt-type continuously variable transmission 1B is cold, the hydraulic oil is not always introduced to the centrifugal hydraulic pressure cancellation chamber 84, and such operation is started. The above inconvenience is repeated each time. Further, even if hydraulic oil is stored in the centrifugal hydraulic pressure canceling chamber 84, when cold such as at the start of operation, the hydraulic fluid with high viscosity becomes a resistance and the roller 85 does not move smoothly, so that an accurate centrifugal hydraulic pressure is obtained. It is also conceivable that the drag force cannot be applied to the piston main body 82a.

  Therefore, the piston initial position holding means of the actuator 480 according to the second embodiment is configured so that centrifugal hydraulic drag is applied to the piston main body 82a using the hydraulic force. Specifically, the piston initial position holding means exemplified here is the piston initial position holding means of the first embodiment provided with the centrifugal hydraulic pressure cancellation chamber 84 and the roller 85, and supplies hydraulic oil to the centrifugal hydraulic pressure cancellation chamber 84. Centrifugal hydraulic cancellation hydraulic oil supply path is provided.

  For example, the centrifugal hydraulic pressure canceling oil supply path is configured by a centrifugal hydraulic pressure canceling oil path 487 shown in FIG. 14 that connects the centrifugal hydraulic pressure canceling chamber 84 and the space T2. As a result, the hydraulic oil that has flowed into the supply / discharge side main passage 51a via the oil passage R7 flows into the centrifugal hydraulic pressure cancellation oil passage 487 via the shaft side communication passage 51c and the spaces T1 and T2, and the centrifugal hydraulic pressure cancellation chamber. 84 can be supplied. In this case, in the hydraulic oil supply control device 130, when the ECU 140 determines that the movement of the roller 85 is bad (for example, when it is cold with reference to the hydraulic oil temperature, the water temperature or the oil temperature of the internal combustion engine), If it is necessary to hold the piston 82 at the initial position, the oil passage R7 can be operated so that an appropriate centrifugal hydraulic drag according to the operating oil temperature, the rotational speed of the primary pulley 50, etc. can be applied to the piston main body 82a. Regulate hydraulic pressure. The reason for using the hydraulic oil temperature information during the pressure adjustment is that the hydraulic oil viscosity changes according to the hydraulic oil temperature, so that an appropriate hydraulic pressure corresponding to the viscosity is generated in the centrifugal hydraulic pressure cancellation chamber 84. . In addition, the information on the rotational speed of the primary pulley 50 is used at the time of the pressure adjustment because the centrifugal hydraulic pressure varies depending on the rotational speed. Therefore, an appropriate hydraulic pressure corresponding to the centrifugal hydraulic pressure is canceled. This is because it is generated in the chamber 84.

  Therefore, in the second embodiment, the centrifugal hydraulic pressure to the piston main body 82a is obtained by the hydraulic pressure of the hydraulic oil supplied from the centrifugal hydraulic pressure cancellation oil passage 487 even if the roller 85 does not move smoothly in the cold time such as the start of operation. Since the deficiency of the drag can be compensated for, it is possible to prevent unnecessary opening of the hydraulic oil supply / discharge valve 70. Further, since the hydraulic oil is supplied from the centrifugal hydraulic pressure canceling oil passage 487 to the centrifugal hydraulic pressure canceling chamber 84 in the piston initial position holding means of the second embodiment, the wear between the roller 85 and each sliding surface is reduced. Reduced and improved durability.

  Here, the diameter (that is, the cross-sectional area So) of the centrifugal hydraulic pressure canceling oil passage 487 is set to a small value such as a so-called orifice, and is further made smaller than the diameter (that is, the cross-sectional area Sd) of the oil drain hole 86 (So). <Sd) is preferred. As a result, the hydraulic pressure in the centrifugal hydraulic pressure cancellation chamber 84 does not increase unnecessarily, and the hydraulic pressure does not have to be an obstacle when the actuator 80 is operated (that is, the hydraulic oil supply / discharge valve 70 is opened). . For this reason, the belt-type continuously variable transmission 1B according to the second embodiment can prevent an increase in the hydraulic pressure P2 of the drive hydraulic chamber 81 and can suppress an increase in drive loss of the oil pump 132. Further, since the hydraulic pressure in the centrifugal hydraulic pressure cancellation chamber 84 does not rise unnecessarily even when the actuator 80 is operated, it is possible to prevent an increase in the hydraulic pressure in the oil passage R7 and to reduce the drive loss of the oil pump 132. Increase can be suppressed.

  Further, as shown in FIG. 15, such a centrifugal hydraulic pressure canceling oil passage 487 is a space portion formed by the piston main body 82a of the piston 82 facing each other and the facing surface 54d of the primary partition wall 54 (that is, the piston 82). In this case, the oil pressure of the oil passage R7 when generating the centrifugal hydraulic drag can be reduced, so that the oil pump 132 is also provided. The increase in driving loss can be suppressed.

  By the way, although the centrifugal hydraulic pressure cancellation oil passage 487 as described above is provided here, the hydraulic pressure for holding the piston main body 82a at the initial position is, for example, the hydraulic pressure of the hydraulic oil supplied from the centrifugal hydraulic pressure cancellation oil passage 487. Instead of or together with this hydraulic pressure, the hydraulic pressure of the hydraulic oil that leaks out from between the sliding support hole 54c and the pressing portion 82b of the piston 82 to the piston main body 82a side and is sent to the centrifugal hydraulic pressure cancellation chamber 84, and the drive hydraulic pressure chamber The hydraulic pressure of the hydraulic fluid that leaks from 81 through the drive hydraulic pressure chamber seal member S2 and is sent to the centrifugal hydraulic pressure cancellation chamber 84 may be used. Thereby, since the burden of the hydraulic pressure from the supply / discharge side main passage 51a can be eliminated or reduced, an increase in driving loss of the oil pump 132 can be suppressed.

Also, here is exemplified the piston initial position holding unit which comprises a piston main body 82a of FIG. 2, for example, a piston main body 182a shown in FIGS. 11 and 13, by replacing 382a, its recess 182a ₃ centrifuged It becomes easy to supply the hydraulic oil to the hydraulic cancellation chamber 84 and to easily generate a desired centrifugal hydraulic drag.

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

  The belt type continuously variable transmission 1C of the third embodiment operates the function of the valve-closing state holding means (piston initial position holding means) in the belt type continuously variable transmissions 1A and 1B variously exemplified in the first and second embodiments. The oil supply / discharge valve 70 is provided. In the third embodiment, a belt type continuously variable transmission 1A according to the first embodiment shown in FIG.

  Specifically, the belt-type continuously variable transmission 1C according to the third embodiment is the same as the belt-type continuously variable transmission 1A according to the first embodiment except that the hydraulic oil supply / discharge valve 70 and the actuator 80 are supplied with hydraulic oil shown in FIG. Instead of the discharge valve 170 and the actuator 580, the valve body 171 of the hydraulic oil supply / discharge valve 170 is configured to be used as the centrifugal hydraulic pressure canceling member 85 of the piston initial position holding means.

First, the actuator 580 of the third embodiment will be described. As shown in FIG. 16, the actuator 580 includes a piston 582 having a piston main body 582a similar to the piston main body 82a described above and a pressing portion 582b similar to the pressing portion 82b described above. In Embodiment 3, the pressing portion 582b is inclined so that the distal end surface on the valve body 171 side faces radially inward. That is, the pressing portion 82b of the third embodiment keeps the inclined surface 582b ₁ shown in FIG. 17 in contact with the valve body 171, and when the pressing portion 582b moves in the axial direction, the inclined surface 582b ₁ The body 171 is separated from the valve seat surface 172. Therefore, the actuator 580 of the third embodiment, when the inclined surface 582b ₁ and the valve body 171 of the pressing portion 582b can not be contacted, so can not be the same functions as the actuator 80 described above, these positional relationships In order to prevent displacement, the piston main body 82a and the pressing portion 82b are integrated by fitting or the like.

  Subsequently, the hydraulic oil supply / discharge valve 170 of the third embodiment is provided in the primary hydraulic chamber 55 which is one clamping pressure generating hydraulic chamber in the belt-type continuously variable transmission 1A, similarly to the hydraulic oil supply / discharge valve 70 described above. The valve is opened when the hydraulic oil is supplied, and is also opened when the hydraulic oil is discharged from the primary hydraulic chamber 55. Accordingly, the hydraulic oil supply / discharge valve 170 is a working fluid from the outside of the primary hydraulic chamber 55 (that is, outside the primary pulley 50) to the primary hydraulic chamber 55, as shown in FIGS. The hydraulic oil is supplied, the hydraulic oil is discharged from the primary hydraulic chamber 55 to the outside of the primary pulley 50, and the hydraulic oil in the primary hydraulic chamber 55 is held. The hydraulic oil supply / discharge valve 170 of the third embodiment is disposed in each valve arrangement passage 54a formed in the primary partition wall 154 of the primary pulley 50 (that is, the hydraulic oil supply / discharge valve 170 of the third embodiment is also provided). , A plurality of locations (for example, three locations) are formed at equal intervals on the circumference) and rotate integrally with the primary pulley 50.

  For example, each hydraulic oil supply / discharge valve 170 of the third embodiment is a ball type check valve, and as shown in FIG. 17, a valve body 171, a valve seat surface 172, a valve body elastic member 173, The pressure transmission member 174 and the snap ring 175 are configured.

  The valve body 171 is a spherical body having a diameter larger than the minimum inner diameter of the valve seat surface 172 and is disposed closer to the primary hydraulic chamber 55 than the valve seat surface 172. Here, the valve seat surface 172 has a tapered shape such as a truncated cone having a diameter that increases toward the inner side in the radial direction on the primary hydraulic chamber 55 side, and with respect to the axial direction of the pressing portion 582b of the piston 582 described later. And tilt it inward in the radial direction. The hydraulic oil supply / discharge valve 170 contacts the valve seat 71 with the valve seat surface 72 to block communication between the valve disposition passage 54 a and the primary hydraulic chamber 55, while separating the valve disc 71 from the valve seat surface 72. Thus, the valve arrangement passage 54a and the primary hydraulic chamber 55 are communicated with each other. That is, in each hydraulic oil supply / discharge valve 170 of the third embodiment, the valve body 171 moves to the primary hydraulic chamber 55 side and moves away from the valve seat surface 172 in the same manner as the hydraulic oil supply / discharge valve 70 described above. The valve opening operation is performed, and the valve body 171 moves in the opposite direction to contact the valve seat surface 172 to perform the valve closing operation.

  The valve body elastic member 173 is a valve body closing direction pressing force generation means similar to the valve body elastic member 73 of the hydraulic oil supply / discharge valve 70 described above, and the valve body 171 is moved to the valve seat via the pressing force transmission member 174. It is arranged in a state of being urged to press against the surface 172. For example, also here, an annular disc spring disposed on the same axis as the primary pulley shaft 51 is used as the valve body elastic member 173, and the valve body elastic is used as the pressing force transmission member 174 on the same axis as the primary pulley shaft 51. An annular member disposed closer to the valve body 171 than the member 173 is used. Here, the valve body elastic member 173 is formed such that only the inner diameter side thereof is formed so that the pressing force transmission member 174 generates an urging force in the valve closing direction. For this reason, the outer side of the valve body elastic member 173 is pressed by the snap ring 175 so as not to move in the valve opening direction of the hydraulic oil supply / discharge valve 170 with respect to the primary partition 154. Further, the pressing force transmission member 174 of the third embodiment slides in the axial direction on the inner peripheral surface of the cylindrical portion on the same axis as the primary pulley shaft 51 that protrudes from the primary partition 154 to the primary hydraulic chamber 55 side. Arrange for movement. The snap ring 175 holds the valve body elastic member 173 on the inner peripheral surface of the cylindrical portion of the primary partition 154. Accordingly, the valve body elastic member 173 generates a valve closing urging force in a direction in which the valve body 171 is brought into contact with the valve seat surface 172. Therefore, the valve closing urging force is applied in the valve closing direction of the hydraulic oil supply / discharge valve 170. The valve body 171 acts as a valve body closing direction pressing force, which is an elastic member pressing force, via the pressing force transmitting member 174. Thus, when the transmission gear ratio is fixed, the valve body 171 is pressed against the valve seat surface 172 as shown in FIG. 16, and each hydraulic oil supply / discharge valve 170 functions as a check valve.

  Here, the inner diameter of the pressing force transmission member 174 is set so as to leave a gap between the inner peripheral surface and the outer peripheral surface of the cylindrical portion 53a of the primary movable sheave 53, and as a result, FIGS. When the valve body 171 is separated from the valve seat surface 172 as shown in FIG. 5, if the gear ratio is reduced, the hydraulic oil flows into the primary hydraulic chamber 55 from the valve arrangement passage 54a through the path indicated by the arrow U in FIG. If the ratio is increased, the hydraulic oil is discharged from the primary hydraulic chamber 55 via the arrow D in FIG. Note that the operations of the ECU 140 and the hydraulic oil supply control device 130 at that time are performed in the same manner as described in the first embodiment.

  The pressing force transmission member 174 includes a cylindrical portion that protrudes toward the valve body 171 from the inner diameter side radially inward of the center of the valve body 171, and the valve body at the protruding end of the cylindrical portion. 171 is held to prevent the valve seat surface 172 from falling. An end surface of the protruding end of the cylindrical portion is an inclined surface 174 a that is inclined radially inward as it goes to the valve body 171. The pressing force transmission member 174 holds the valve body 171 on the inclined surface 174a and presses it against the valve seat surface 172 with the valve closing biasing force of the pressing force transmission member 174. In this case, the inclination angle θ2 of the inclined surface 174a of the pressing force transmitting member 174 is made larger than the inclination angle θ1 of the valve seat surface 172 on the piston 582 side (θ1 <θ2). Thereby, since the minimum space | interval between the inclined surface 174a and the valve seat surface 172 and the primary partition 154 in the vicinity becomes smaller than the diameter of the valve body 171, the fall of the valve body 171 can be prevented reliably.

Further, the valve body 171 of the third embodiment is in contact with the inclined surface 582b ₁ on the radially inner side of the inclined surface 582b ₁ of the pressing portion 582b of the piston 582 described above. For this reason, when a centrifugal force is generated in the valve body 171 with the rotation of the primary pulley 50, a pressing force acts on the inclined surface 582b ₁ of the pressing portion 582b by this centrifugal force, and this pressing force becomes the above-described centrifugal hydraulic drag. It acts on the pressing part 582b. Therefore, since the centrifugal hydraulic drag acts as a force that resists at least the centrifugal hydraulic pressure of the drive hydraulic chamber 81, the piston 582 can be held at the initial position when the operation of the actuator 580 is unnecessary. That is, in the third embodiment, the piston initial position holding means is mainly constructed by the valve body 171 and the inclined surface 582b ₁ of the pressing portion 582b.

  As described above, in the third embodiment, the hydraulic oil supply / discharge valve 170 is configured and arranged so as to use the function of the closed state holding means (piston initial position holding means). , 2 as well as suppressing the increase in the drive loss of the oil pump 132, and no special parts for the valve-closing state holding means are required. The number of parts can be reduced. Therefore, the belt-type continuously variable transmission 1C according to the third embodiment can be reduced in cost and size, and further improved in reliability.

In this case, since hydraulic oil is filled between the valve body 171 and the inclined surface 582b ₁ of the pressing portion 582b and between the valve body 171 and the inclined surface 174a of the pressing force transmission member 174, the valve Ensuring smooth operation of the body 171 and avoiding wear between them can be realized without providing a separate lubricating structure. For this reason, the belt-type continuously variable transmission 1C of the third embodiment can further reduce the cost, reduce the size, and improve the reliability.

  As described above, the belt type continuously variable transmission according to the present invention is suitable for a technique for suppressing an increase in driving loss of the oil pump.

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 principal part sectional drawing which shows the structure of the primary pulley in the belt-type continuously variable transmission of Example 1, Comprising: It is a figure which shows the state at the time of gear ratio fixed. FIG. 3 is a cross-sectional view of a main part for explaining the piston initial position holding means of the first embodiment viewed from the line XX shown in FIG. 2. It is principal part sectional drawing which shows the structure of the piston initial position holding | maintenance means of Example 1. FIG. It is operation | movement explanatory drawing of the piston initial position holding | maintenance means of Example 1. FIG. It is a figure which shows the torque cam of a secondary pulley. It is operation | movement explanatory drawing of a torque cam. It is a figure which shows the structural example of a hydraulic-oil supply control apparatus, Comprising: It is a figure explaining the operation | movement at the time of gear ratio fixation. It is operation | movement explanatory drawing of the primary pulley at the time of gear ratio reduction. It is operation | movement explanatory drawing of the hydraulic fluid supply control apparatus at the time of gear ratio reduction. It is operation | movement explanatory drawing of the primary pulley at the time of gear ratio increase. It is operation | movement explanatory drawing of the hydraulic-oil supply control apparatus at the time of gear ratio increase. It is a figure which shows the other structural example of a piston main body. It is a figure which shows the other structural example of a piston. It is a figure which shows the other structural example of a piston main body. It is principal part sectional drawing which shows the structure of the piston initial position holding | maintenance means of Example 2. FIG. It is principal part sectional drawing which shows the other structural example of the piston initial position holding | maintenance means of Example 2. FIG. It is principal part sectional drawing which shows the structure of the primary pulley in the belt-type continuously variable transmission of Example 3, Comprising: It is a figure which shows the state at the time of gear ratio fixation. It is principal part sectional drawing explaining the structure of the piston initial position holding means of Example 3. FIG. It is operation | movement explanatory drawing of the primary pulley of Example 3 at the time of gear ratio reduction. It is operation | movement explanatory drawing of the primary pulley of Example 3 at the time of gear ratio increase.

Explanation of symbols

1A, 1B, 1C Belt type continuously variable transmission 10 Internal combustion engine (drive source)
DESCRIPTION OF SYMBOLS 20 Transaxle 30 Torque converter 40 Forward / reverse switching mechanism 50 Primary pulley 51 Primary pulley shaft 52 Primary fixed sheave 53 Primary movable sheave 54 Primary partition 54a Valve arrangement passage 55 Primary hydraulic chamber (one clamping pressure generating hydraulic chamber)
60 Secondary pulley 61 Secondary pulley shaft 62 Secondary fixed sheave 63 Secondary movable sheave 64 Secondary hydraulic chamber (the other clamping pressure generating hydraulic chamber)
70 Hydraulic oil supply / discharge valve (hydraulic oil supply valve, hydraulic oil discharge valve)
71 Valve body 72 Valve seat surface 73 Valve body elastic member 74 Pressing force transmission member 75 Snap ring 80 Actuator (valve opening / closing means)
81 Driving hydraulic chamber 82 Piston 82a Piston main body 82a ₁ Centrifugal hydraulic drag generating surface 82b Pressing portion 83 Cover member 84 Centrifugal hydraulic canceling chamber 84c Space 84a Inclined surface 85 Centrifugal hydraulic canceling member (roller)
86 Oil drain hole 90 Final reduction gear 100 Power transmission path 110 Belt 130 Hydraulic oil supply control device 132 Oil pump 133 Line pressure control device 134 Constant pressure control device 135 Primary hydraulic chamber control device 136 Drive hydraulic chamber control device 137 Secondary hydraulic pressure Room control device 154 Primary partition wall 170 Hydraulic oil supply / discharge valve (hydraulic oil supply valve, hydraulic oil discharge valve)
171 Valve body 172 Valve seat surface 173 Valve body elastic member 174 Pressing force transmission member 174a Inclined surface 175 Snap ring 182a Main body of piston 182a Contact surface with ₁ pressing portion 182a ₂ Contact surface with roller 182a ₃ Recessed portion 282 Piston 282a Main body of piston 282a ₁ concave portion 282 b pressing portion 382 a piston main body 382 a ₂ concave portion 480 actuator (valve opening / closing means)
487 Centrifugal hydraulic canceling oil passage 580 Actuator (valve opening / closing means)
582 Piston 582a Piston main body 582b Pressing portion 582b ₁ Inclined surface S1 Primary hydraulic chamber seal member S2, S3 Drive hydraulic chamber seal member

Claims (6)

Two pulleys rotating with respect to the stationary member;
A belt that is wound around each pulley and transmits a driving force from a driving source transmitted to one pulley to the other pulley;
A clamping pressure generating hydraulic chamber formed on each pulley and generating a belt clamping pressure against the belt by hydraulic pressure;
A hydraulic oil supply valve that is provided at least one of the pulleys to rotate integrally with the pulley and opens when hydraulic oil is supplied to a clamping pressure generating hydraulic chamber of the pulley;
A hydraulic oil discharge valve that opens when the hydraulic oil is discharged from the clamping pressure generating hydraulic chamber supplied with the hydraulic oil, and rotates integrally with the pulley;
Valve opening / closing forcibly opening the hydraulic oil supply valve and the hydraulic oil discharge valve by sliding the piston from the initial position to one of the sliding directions with respect to the driving hydraulic chamber by the hydraulic pressure of the driving hydraulic chamber Means,
A belt type continuously variable transmission comprising:
The valve opening / closing means is a piston initial position that holds the piston in the initial position with a mechanical force against a force in a valve opening direction applied to the piston by centrifugal hydraulic pressure of the driving hydraulic chamber accompanying rotation of a pulley. A belt-type continuously variable transmission comprising a holding means.   The piston initial position holding means includes a centrifugal hydraulic pressure canceling member that presses the piston in a valve closing direction of the hydraulic oil supply valve and a valve closing direction of the hydraulic oil discharge valve by a centrifugal force applied when the pulley rotates. The belt-type continuously variable transmission according to claim 1.   A centrifugal hydraulic pressure canceling member of the piston initial position holding means for the piston related to the hydraulic oil supply valve and the piston related to the hydraulic oil supply valve, a valve body of the hydraulic oil discharge valve, and the hydraulic oil discharge valve 3. The belt type continuously variable transmission according to claim 2, wherein a centrifugal hydraulic pressure canceling member of the piston initial position holding means for the piston is integrated with each other.   3. The belt according to claim 2, wherein the hydraulic oil supply valve and the hydraulic oil discharge valve have a common valve body, and the valve body and a centrifugal hydraulic pressure canceling member of the piston initial position holding means are integrated. Type continuously variable transmission.   The belt-type continuously variable transmission according to claim 2, 3 or 4, wherein the piston initial position holding means has a space on the outer diameter side of the piston and the centrifugal hydraulic pressure canceling member.   6. The belt-type continuously variable transmission according to claim 1, wherein the piston initial position holding means is configured to hold the piston at the initial position using a hydraulic force. Machine.

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