JP2006275276A - Hydraulic control device of belt type continuously variable transmission - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To miniaturize an oil pump, and to further reduce fuel consumption, by reducing control hydraulic pressure for generating belt pinching force, by controlling cancel force of centrifugal hydraulic pressure in low speed rotation, by surely canceling the centrifugal hydraulic pressure in high speed rotation. <P>SOLUTION: This hydraulic control device has a control hydraulic pressure chamber 45A for supplying the control hydraulic pressure for generating the belt pinching force acting on a belt to a movable sheave 43, a centrifugal hydraulic pressure cancel oil chamber 45B, discharge passages 90F and 43D for discharging a hydraulic fluid of the centrifugal hydraulic pressure cancel oil chamber 45B, and a supply means for supplying the hydraulic fluid to the centrifugal hydraulic pressure cancel oil chamber. The discharge passages 90F and 43D are movably formed in the opening area by the shaft directional relative movement between a member 43C connected to the movable sheave on the outer diameter side of the centrifugal hydraulic pressure cancel oil chamber and constituting a part of the centrifugal hydraulic pressure cancel oil chamber and a member 90D connected to a fixed sheave 42 and constituting the other part of the centrifugal hydraulic pressure cancel oil chamber. <P>COPYRIGHT: (C)2007,JPO&INPIT

Description

  The present invention relates to a hydraulic control device for a belt-type continuously variable transmission, and in particular, transmits power between two variable pulleys by means of a belt, and controls the gear ratio by changing the belt winding radius. The present invention relates to a hydraulic control device for a belt-type continuously variable transmission having a configuration.

  In general, a stepped or continuously variable transmission is provided on the output side of the engine for the purpose of operating the engine under optimum conditions according to the traveling state of the vehicle. An example of such a continuously variable transmission is a belt-type continuously variable transmission. This belt-type continuously variable transmission has two rotating members arranged in parallel, and a primary pulley and a secondary pulley separately attached to each rotating member. Both the primary pulley and the secondary pulley are configured by combining a fixed sheave and a movable sheave, and a V-shaped groove is formed between the fixed sheave and the movable sheave.

  Further, a belt is wound around the groove of the primary pulley and the groove of the secondary pulley, and a hydraulic chamber for generating a holding pressure in the axial direction is separately provided on the movable sheave. By separately controlling the hydraulic pressure in each hydraulic chamber, the groove width of the primary pulley is controlled to change the belt wrapping radius and the gear ratio is changed, while the groove width of the secondary pulley is changed. The belt tension is controlled.

  By the way, in the belt type continuously variable transmission as described above, since the hydraulic chamber rotates around the rotation shaft, the hydraulic pressure generated by centrifugal force, so-called centrifugal hydraulic pressure, acts on the hydraulic chamber, and the hydraulic pressure in the hydraulic chamber is reduced. There is a possibility that the hydraulic pressure is higher than the control target. As a result, it is known that the belt clamping pressure increases more than necessary, which deteriorates belt transmission efficiency and adversely affects belt durability. An example of a belt-type continuously variable transmission that takes measures to eliminate such inconvenience due to centrifugal hydraulic pressure is described in Patent Document 1.

  That is, in Patent Document 1, lubricant oil is controlled and supplied to a balance oil chamber (a cancel oil chamber referred to in the present application) that cancels centrifugal hydraulic pressure generated in a hydraulic cylinder for shifting a secondary pulley provided on a secondary shaft. A balance oil pressure control valve that opens to supply the balance oil chamber when the balance oil chamber is short of lubricating oil, and closes when the lubricant oil is discharged from the balance oil chamber during a downshift. A belt type continuously variable transmission that can reduce the required discharge amount to an oil pump by stopping the supply of lubricating oil to is disclosed. The lubricating oil is discharged from the balance oil chamber by a gap formed between the shaft portion of the fixed pulley (the fixed sheave referred to herein) and the end portion of the balance oil chamber cylinder, which is an opening of the balance oil chamber. It is comprised so that it may be performed from.

JP 2001-248699 A

  In general, in a belt-type continuously variable transmission, it is necessary to reliably cancel the centrifugal hydraulic pressure by filling the cancel oil chamber with hydraulic oil during high-speed rotation so that belt clamping pressure is not generated more than necessary. On the other hand, during low-speed rotation, conversely, the amount of hydraulic oil in the cancellation oil chamber is controlled to a small level to reduce the cancellation force of the centrifugal oil pressure, and the control oil pressure that generates the belt clamping pressure is reduced as much as possible. This eliminates the need for generating unnecessary hydraulic pressure, which is effective for reducing the size of the oil pump and further reducing fuel consumption.

  However, like the belt-type continuously variable transmission described in Patent Document 1, a passage for discharging hydraulic oil from the balance oil chamber (cancellation oil chamber referred to in the present application) is provided with a fixed pulley (fixed sheave described in the present application). )) And the end portion of the balance oil chamber cylinder, the passage is positioned on the inner diameter side of the cancel oil chamber, so that the operating oil is discharged from the cancel oil chamber. Will be limited by the position of the gap. That is, in the configuration of the belt-type continuously variable transmission described in Patent Document 1, it goes without saying that the pulley is rotating, and the hydraulic oil is stored up to the gap below the cancel oil chamber even when the pulley is stopped. Therefore, in the configuration described in Patent Document 1, it is difficult to control the filling amount of the hydraulic oil in the cancellation oil chamber, and hence the oil surface position, and it is also difficult to control the cancellation force of the centrifugal hydraulic pressure. is there.

  The object of the present invention is to make it possible to reliably cancel the centrifugal hydraulic pressure at the time of high-speed rotation and to control the cancellation force of the centrifugal hydraulic pressure at the time of low-speed rotation, and generate belt clamping pressure against the background of the above circumstances. It is an object of the present invention to provide a hydraulic control device for a belt-type continuously variable transmission that can reduce the control hydraulic pressure to reduce the size of an oil pump and further reduce fuel consumption.

  In order to achieve the above object, a hydraulic control device for a belt-type continuously variable transmission according to an aspect of the present invention includes a control hydraulic chamber to which a control hydraulic pressure is generated to generate a belt clamping pressure acting on the belt to the movable sheave, A hydraulic control device for a belt-type continuously variable transmission, comprising: a centrifugal hydraulic pressure cancellation oil chamber; a discharge passage for discharging hydraulic oil in the centrifugal hydraulic pressure cancellation oil chamber; and a supply means for supplying hydraulic oil to the centrifugal hydraulic pressure cancellation oil chamber The discharge passage is on the outer diameter side of the centrifugal hydraulic pressure canceling oil chamber, and is connected to the movable sheave and constituting a part of the centrifugal hydraulic pressure canceling oil chamber, and to a fixed sheave and connected to the centrifugal hydraulic pressure. The opening area is variably formed by axial relative movement with a member constituting another part of the cancel oil chamber.

  Here, it is preferable that the discharge passage is configured so that an opening area thereof increases as the movable sheave moves in a direction in which a gear ratio increases.

  Further, the control oil pressure supplied to the control oil pressure chamber is controlled by decreasing the amount of hydraulic oil in the centrifugal oil pressure canceling oil chamber which decreases as the movable sheave moves in the direction in which the gear ratio increases. It may be.

  According to one aspect of the present invention, a control hydraulic pressure chamber that is supplied with a control hydraulic pressure that generates a belt clamping pressure acting on the belt to the movable sheave, a centrifugal hydraulic pressure cancellation oil chamber, and hydraulic oil in the centrifugal hydraulic pressure cancellation oil chamber are supplied. In a hydraulic control device for a belt-type continuously variable transmission including a discharge passage for discharging and a supply means for supplying hydraulic oil to the centrifugal hydraulic pressure cancellation oil chamber, the discharge passage is connected to a member connected to a movable sheave and a fixed sheave. Since the opening area is variably formed by the axial relative movement with the connected member, the opening area changes due to the movement of the movable sheave accompanying the change of the gear ratio. Accordingly, the amount of hydraulic oil discharged from the centrifugal oil pressure canceling oil chamber changes, and the oil level position is adjusted. As a result, the canceling force of the centrifugal hydraulic pressure can be controlled.

  Here, according to the configuration in which the opening area of the discharge passage increases as the movable sheave moves in the direction in which the gear ratio increases, the centrifugal speed is increased on the speed reduction side where the gear ratio increases. Since the amount of hydraulic oil discharged from the hydraulic pressure canceling oil chamber increases, the canceling force of the centrifugal hydraulic pressure can be reduced. Therefore, the control hydraulic pressure that generates the belt clamping pressure can be reduced by the reduction of the canceling force of the centrifugal hydraulic pressure.

  Further, the control oil pressure supplied to the control oil pressure chamber is controlled by decreasing the amount of hydraulic oil in the centrifugal oil pressure canceling oil chamber that decreases as the movable sheave moves in the direction in which the gear ratio increases. According to this, it is possible to reduce the control hydraulic pressure that generates the belt clamping pressure more accurately, and it is possible to reliably reduce the drive loss of the oil pump.

  Here, an embodiment of a hydraulic control device for a belt type continuously variable transmission according to the present invention will be specifically described with reference to the drawings.

  FIG. 1 is a schematic configuration diagram showing a part of a vehicle to which a belt type continuously variable transmission according to the present invention is applied. A vehicle 1 shown in FIG. 1 is configured as a so-called FF vehicle (front engine front drive: front wheel drive vehicle in front of the engine), and includes an engine 2 as a drive source. As the engine 2, a gasoline engine, a diesel engine, an LPG engine, a hydrogen engine, a bi-fuel engine, or the like can be adopted. Here, a description will be given assuming that a gasoline engine is used as the engine 2.

  As shown in FIG. 1, the vehicle 1 has a transaxle 3 that is disposed on the side of a horizontally placed engine 2 and connected to a crankshaft SC of the engine 2. The transaxle 3 includes a transaxle housing 4, a transaxle case 5, and a transaxle rear cover 6. The housing 4 is disposed on the side of the engine 2, and the case 5 is fixed to the opening end of the housing 4 on the side opposite to the engine 2. The rear cover 6 is fixed to the opening end of the case 5 on the side opposite to the housing 4. A torque converter 7 is arranged inside the transaxle housing 4. Inside the transaxle case 5 and the transaxle rear cover 6, a forward / reverse switching mechanism 8, a belt type continuously variable transmission (CVT) 9. A final reduction gear (differential device) 10 is arranged.

  The torque converter 7 includes a drive plate 11 and a front cover 12 fixed to the crankshaft SC of the engine 2 via the drive plate 11. As shown in FIG. 1, a pump impeller 14 is attached to the front cover 12. The torque converter 7 includes a turbine runner 15 that can rotate while facing the pump impeller 14.

  The turbine runner 15 is fixed to an input shaft SI that extends substantially coaxially with the crankshaft SC. Further, a stator 16 is disposed inside the pump impeller 14 and the turbine runner 15, and the rotation direction of the stator 16 is set only in one direction by the one-way clutch 17. The stator 16 is fixed to the hollow shaft 18 via the one-way clutch 17, and the above-described input shaft SI is inserted into the hollow shaft 18. A lockup clutch 20 is attached to the end of the input shaft SI on the front cover 12 side via a damper mechanism 19. The hollow shaft 18 is fixed to the case 5.

  The pump impeller 14, the turbine runner 15, and the stator 16 described above define a hydraulic oil chamber. The hydraulic oil chamber is operated from an oil pump 21 disposed between the torque converter 7 and the forward / reverse switching mechanism 8. Oil is supplied. Then, when the engine 2 is operated and the front cover 12 and the pump impeller 14 are rotated, the turbine runner 15 starts to be dragged by the flow of hydraulic oil. Further, the stator 16 converts the flow of hydraulic oil into a direction that assists the rotation of the pump impeller 14 when the rotational speed difference between the pump impeller 14 and the turbine runner 15 is large.

  Thus, the torque converter 7 operates as a torque amplifier when the rotational speed difference between the pump impeller 14 and the turbine runner 15 is large, and operates as a fluid coupling when the rotational speed difference between the two becomes small. When the vehicle speed reaches a predetermined speed after the vehicle 1 starts, the lockup clutch 20 is operated so that the power transmitted from the engine 2 to the front cover 12 is mechanically and directly transmitted to the input shaft SI. Become. Further, the fluctuation of the torque transmitted from the front cover 12 to the input shaft SI is absorbed by the damper mechanism 19.

  The oil pump 21 between the torque converter 7 and the forward / reverse switching mechanism 8 has a rotor 22, and the rotor 22 is connected to the pump impeller 14 via a hub 23. The main body 24 of the oil pump 21 is rotatably supported on the transaxle case 5 side. Accordingly, the power of the engine 2 is transmitted to the rotor 22 via the pump impeller 14, thereby driving the oil pump 21.

  The forward / reverse switching mechanism 8 has a planetary gear mechanism 25 of a double pinion type. The planetary gear mechanism 25 includes a sun gear 26 attached to an end of the input shaft SI on the continuously variable transmission 9 side, a ring gear 27 disposed concentrically on the outer peripheral side of the sun gear 26, and a plurality of pinion gears that mesh with the sun gear 26. 28, a plurality of pinion gears 29 that mesh with both the ring gear 27 and the pinion gear 28, and a carrier 30 that holds the pinion gears 28 so as to be capable of rotating, and holds the pinion gears 28 in an integrally revolving state around the sun gear 26. Including.

  The carrier 30 of the forward / reverse switching mechanism 8 is fixed to the primary shaft SP included in the belt-type continuously variable transmission 9, and the power transmission path between the carrier 30 and the input shaft SI is connected using the forward clutch CL. Blocked. The forward / reverse switching mechanism 8 has a reverse brake BR that controls the rotation and fixation of the ring gear 27.

  On the other hand, the belt-type continuously variable transmission 9 includes the above-described primary shaft (driving side rotating shaft) SP extending substantially coaxially with the input shaft SI, and a secondary shaft (driven side rotating) arranged in parallel with the primary shaft SP. Axis) SS. The primary shaft SP is rotatably supported by the bearings 31 and 32, and the secondary shaft SS is rotatably supported by the bearings 33 and 34. The primary shaft SP is equipped with a primary pulley 35, and the secondary shaft SS is equipped with a secondary pulley 36.

  The primary pulley 35 includes a fixed sheave 37 that is integrally formed on the outer periphery of the primary shaft SP, and a movable sheave 38 that is slidably mounted on the outer periphery of the primary shaft SP. The fixed sheave 37 and the movable sheave 38 face each other, and a substantially V-shaped pulley groove 39 is formed between them. Further, the movable sheave 38 is movable in the axial direction of the primary shaft SP with respect to the fixed sheave 37, and the continuously variable transmission 9 moves the movable sheave 38 in the axial direction of the primary shaft SP to move with the movable sheave 38. A hydraulic actuator 40 is provided to approach and separate the fixed sheave 37.

  Similarly, the secondary pulley 36 also includes a fixed sheave 42 that is integrally formed on the outer periphery of the secondary shaft SS, and a movable sheave 43 that is slidably mounted on the outer periphery of the secondary shaft SS. The fixed sheave 42 and the movable sheave 43 face each other, and a substantially V-shaped pulley groove 44 is formed between them. The movable sheave 43 is also movable in the axial direction of the secondary shaft SS with respect to the fixed sheave 42. The continuously variable transmission 9 moves the movable sheave 43 in the axial direction of the secondary shaft SS to A hydraulic actuator 45 is provided to approach and separate the fixed sheave 42.

  Around the pulley groove 39 of the primary pulley 35 and the pulley groove 44 of the secondary pulley 36, a belt B composed of a number of metal pieces and a plurality of steel rings is wound. Then, the hydraulic pressures by the hydraulic actuators 40 and 45 are separately controlled, whereby the groove widths of the primary pulley 35 and the secondary pulley 36 are changed, and the winding radius of the belt B is changed. As a result, the speed ratio of the continuously variable transmission 9 is set to a desired value, and the tension of the belt B is adjusted. The bearing 34 that supports the secondary shaft SS is fixed to the transaxle rear cover 6, and a parking gear PG is provided between the bearing 34 and the secondary pulley 36.

  As shown in FIG. 1, a shaft 48 supported by bearings 46 and 47 is connected to the secondary shaft SS of the belt type continuously variable transmission 9. A drive gear 49 is fixed to the shaft 48, and power is transmitted from the belt-type continuously variable transmission 9 to the final reduction gear 10 via the drive gear 49. The final reduction gear 10 includes an intermediate shaft 50 that is arranged in parallel with the secondary shaft SS. The intermediate shaft 50 is supported by bearings 51 and 52, and a counter driven gear 53 that meshes with the drive gear 49 of the secondary shaft SS and a final drive gear 54 are fixed to the shaft 50.

  Further, the final reduction gear 10 has a hollow differential case 55. The differential case 55 is rotatably supported by bearings 56 and 57, and a ring gear 58 is formed on the outer periphery thereof. The ring gear 58 meshes with the final drive gear 54 of the intermediate shaft 50. Further, the differential case 55 supports a pinion shaft 59 therein, and two pinion gears 60 are fixed to the pinion shaft 59. Each of the pinion gears 60 is engaged with two side gears 61. Each side gear 61 is connected to a front drive shaft 62 separately. A wheel (front wheel) FW is fixed to each front drive shaft 62. ing.

  FIG. 2 is an enlarged cross-sectional view showing the main part of the belt-type continuously variable transmission 9 described above. FIG. 2 shows a configuration related to the primary pulley 35 and the primary shaft SP of the continuously variable transmission 9. Yes. The primary shaft SP is rotatable about an axis, a fixed sheave 37 is integrally formed at one end of the primary shaft SP, and an oil passage SPA is formed inside in the axial direction. The primary shaft SP is rotatably supported by a bearing 31 fixed to the transaxle case 5 outside the fixed sheave 37. An oil passage SPA formed in the axial direction inside the primary shaft SP is communicated with a hydraulic circuit of a hydraulic control device described later. Further, the primary shaft SP is provided with an oil passage SPB extending in the radial direction toward the outer peripheral surface thereof and communicating with the oil passage SPA.

  On the other hand, the movable sheave 38 includes an inner cylindrical portion 38A that slides along the outer peripheral surface of the primary shaft SP, and a radial direction portion 38B that is continuous from the end on the fixed sheave 37 side of the inner cylindrical portion 38A toward the outer peripheral side. The outer cylindrical portion 38C is continuous to the outer peripheral end of the radial direction portion 38B and extended in the axial direction toward the bearing 33. The inner cylinder portion 38A is formed with an oil passage 38D penetrating from the inner peripheral surface to the outer peripheral surface. The oil passage 38D and the oil passage SPB are communicated with each other through a spline portion described later formed on the outer peripheral surface of the primary shaft SP.

  That is, as shown in FIG. 2, a plurality of spline teeth (grooves) 38 </ b> S are formed on the inner peripheral surface of the inner cylindrical portion 38 </ b> A of the movable sheave 38. On the other hand, a plurality of spline grooves (teeth) SPG are formed on the outer peripheral surface of the primary shaft SP that slidably supports the movable sheave 38. The spline teeth 38S and the spline grooves SPG are formed such that the tooth surfaces or groove surfaces form an involute curve, and the primary shaft SP and the movable sheave 38 can be relatively moved relative to each other in the axial direction. And the movable sheave 38 are in a state in which relative movement is impossible in the circumferential direction.

  The radial oil passage SPB is formed on the outer side in the axial direction from the spline groove SPG formed in the primary shaft SP. In this case, the radial oil passage SPB is positioned outside the transmission path of the torque transmitted from the belt B to the primary shaft SP via the movable sheave 38. Stress concentration does not occur, and the strength of the primary shaft SP can be ensured.

  Furthermore, the belt-type continuously variable transmission 9 includes a cylinder member 70 that is an annular partition member. As can be seen from FIG. 2, the cylinder member 70 includes a first radial portion 70A extending in the radial direction of the primary shaft SP, and a first cylindrical portion extending substantially parallel to the axis of the primary shaft SP from the first radial direction portion 70A. 70B, a second radial portion 70C extending in the radial direction of the primary shaft SP along the back surface of the movable sheave 38 from the first cylindrical portion 70B, and an outer cylindrical portion of the movable sheave 38 from the second radial direction portion 70C. It has a second cylindrical portion 70D extending substantially parallel to the axis of the primary shaft SP via a curved portion for escape with respect to 38C.

  A small-diameter portion at the tip of the primary shaft SP is press-fitted into the center hole formed in the first radial direction portion 70 </ b> A of the cylinder member 70, and the cylinder member 70 uses a lock nut 75 to form a step portion of the primary shaft SP. It is fixed between. The first cylindrical portion 70B of the cylinder member 70 is rotatably supported by a bearing 32 fixed to the transaxle rear cover 6 by an annular bearing retainer and a bolt (not shown). Accordingly, the primary shaft SP is rotatably supported by the bearing 32 through the cylinder member 70 (first cylindrical portion 70B) together with the above-described bearing 31.

  A seal member 72 is disposed on the outer edge portion of the outer cylindrical portion 38C of the movable sheave 38 so as to be in sliding contact with the inner peripheral surface of the second cylindrical portion 70D of the cylinder member 70. On the other hand, on the outer peripheral side of the axial end portion of the inner cylindrical portion 38A of the movable sheave 38, a second sliding portion described later that slidably contacts the inner peripheral side of the first cylindrical portion 70B of the cylinder member 70. 38F is formed. Thus, the inner cylinder portion 38A, the radial direction portion 38B, the outer cylinder portion 38C, and the cylinder member 70 of the movable sheave 38 define the first control hydraulic chamber 40A that constitutes the hydraulic actuator 40 described above. On the other hand, the second control that constitutes the hydraulic actuator 40 described above is constituted by the first radial direction portion 70A of the cylinder member 70, the first cylindrical portion 70B, the axial end portion of the inner cylindrical portion 38A of the movable sheave 38 and the primary shaft SP. A hydraulic chamber 40B is defined. By controlling the hydraulic pressure in the first control hydraulic chamber 40A and the second control hydraulic chamber 40B, the movable sheave 38 is moved with respect to the fixed sheave 37, and the winding radius of the belt B is changed. A gear ratio can be obtained.

  The movable sheave 38 is provided with a first sliding portion 38E and the above-described second sliding portion 38F that are spaced apart in the axial direction of the primary shaft SP. Of the two sliding portions of the movable sheave 38, the first sliding portion 38E is provided on the fixed sheave 37 side in the axial direction of the primary shaft SP with respect to the spline 38S and on the inner peripheral surface of the movable sheave 38. In contact with the outer peripheral surface of the primary shaft SP. On the other hand, as described above, the second sliding portion 38F is spaced apart from the first sliding portion 38E in the axial direction, and on the outer peripheral surface of the axial end portion of the inner cylindrical portion 38A of the movable sheave 38. Is provided. And the 2nd sliding part 38F contacts the internal peripheral surface of the 1st cylindrical part 70B of the cylinder member 70 instead of the primary shaft SP, as FIG. 2 shows.

  On the other hand, FIG. 3 is an enlarged half sectional view showing another main part of the belt-type continuously variable transmission 9 according to the present invention described above. This figure shows the secondary pulley 36 and the secondary shaft SS of the continuously variable transmission 9. A related configuration is shown. The secondary shaft SS is rotatable about an axis, a fixed sheave 42 is integrally formed at one end of the secondary shaft SS, and oil passages SSA and SSB are formed in the axial direction inside. As described above, the secondary shaft SS is rotatably supported together with the parking gear PG by the bearing 34 fixed to the transaxle rear cover 6 outside the fixed sheave 42. The oil passages SSA and SSB formed in the axial direction inside the secondary shaft SS are communicated with the hydraulic circuit of the hydraulic control device as will be described later. Further, the secondary shaft SS is provided with an oil passage SSC extending in the radial direction toward the outer peripheral surface thereof and communicating with the oil passage SSA, and an oil passage SSD similarly communicating with the oil passage SSB.

  On the other hand, the movable sheave 43 includes an inner cylindrical portion 43A that slides along the outer peripheral surface of the secondary shaft SS, and a radial portion that is continuous from the end on the fixed sheave 42 side of the inner cylindrical portion 43A toward the outer peripheral side. 43B and an outer cylindrical portion 43C which is continuous with the outer peripheral end of the radial portion 43B and extends in the axial direction toward the bearing 33. That is, a plurality of spline teeth (grooves) are formed on the inner peripheral surface of the inner cylindrical portion 43A of the movable sheave 43, while a plurality of spline teeth (grooves) are formed on the outer peripheral surface of the secondary shaft SS that slidably supports the movable sheave 43. Spline grooves (teeth) are formed. The spline teeth and the spline grooves are formed such that the tooth surfaces or groove surfaces form, for example, an involute curve, and the secondary shaft SS and the movable sheave 43 can be smoothly moved relatively in the axial direction. The movable sheave 43 cannot move relative to the circumferential direction.

  In this embodiment, the movable sheave 43 has a partition member 80 having a substantially bowl shape. The partition member 80 includes a radial portion 80A extending in the radial direction and a cylindrical portion 80B that is continuous with the outer peripheral end of the radial direction portion 80A and extends in the axial direction toward the bearing 34. is doing. The tubular portion 80 </ b> B is fitted to the outer tubular portion 43 </ b> C of the movable sheave 43 and is fixed to the movable sheave 43 by a snap ring 72. A seal member 74 that contacts the inner peripheral surface of the tubular portion 80B of the partition wall member 80 is disposed on the outer tubular portion 43C of the movable sheave 43. Further, the outer cylindrical portion 43C of the movable sheave 43 is formed with a communication hole 43D that constitutes a part of a discharge passage described later.

  Further, the belt type continuously variable transmission 9 includes an annular piston member 90. As can be seen from FIG. 3, the piston member 90 has a first radial base portion 90A extending in the radial direction of the secondary shaft SS, and extends substantially parallel to the axis of the secondary shaft SS from the first radial base portion 90A. An inner cylindrical portion 90B having a second radial direction portion 90C extending in the radial direction of the secondary shaft SS along the back surface of the movable sheave 43 from the inner cylindrical portion 90B, and an outer peripheral end of the second radial direction portion 90C. The outer cylindrical portion 90D is continuous and is extended in the axial direction toward the bearing 33 side. Seal members 92 and 94 are slidably in contact with the inner peripheral surface of the outer cylindrical portion 43C of the movable sheave 43 on the outer peripheral end of the second radial direction portion 90C and the outer peripheral surface of the outer cylindrical portion 90D. Each is arranged. A plurality of oil passage grooves 90E communicating with the above-described oil passage SSD and a centrifugal hydraulic pressure canceling oil chamber described later are formed radially on the end surface of the first radial base portion 90A of the piston member 90. A spring 96 that urges the movable sheave 43 in the direction in which the pulley groove 44 narrows is disposed between the back surface of the radial direction portion 43B of the movable sheave 43 and the step portion of the inner cylindrical portion 90B. Further, the outer cylindrical portion 90D of the piston member 90 is formed with a communication hole 90F that constitutes a part of a discharge passage described later.

  Reference numeral 100 denotes a guide member, which is continuous with the radial direction portion 100A extending in the radial direction, and the outer peripheral end of the radial direction portion 100A, and on the bearing 34 side along the inner cylindrical portion 90B of the piston member 90. And a cylindrical portion 100B extended in the axial direction.

  In the piston member 90 described above, the small-diameter portion at the tip of the secondary shaft SS is press-fitted into the center hole formed in the first radial base portion 90 </ b> A of the piston member 90, and the secondary shaft SS is used using the lock nut 95. The guide member 100 and the bearing 33 are fixed together with the step portion.

  Thus, a control hydraulic chamber 45A constituting the hydraulic actuator 45 described above is defined by the inner cylindrical portion 43A, the radial direction portion 43B, the outer cylindrical portion 43C and the piston member 90 of the movable sheave 43. On the other hand, the inner cylindrical portion 90B of the piston member 90, the second radial direction portion 90C, the outer cylindrical portion 90C, the outer cylindrical portion 43C of the movable sheave 43, the radial direction portion 80A of the partition wall member 80, and the cylinder of the guide member 100 The centrifugal portion 100B defines a centrifugal hydraulic pressure canceling oil chamber 45B that constitutes the hydraulic actuator 45 described above. The hydraulic oil controlled by the hydraulic circuit of the hydraulic control device is supplied to the control hydraulic chamber 45A through the oil passage SSA formed in the axial direction of the secondary shaft SS and the oil passage SSC formed in the radial direction. Supplied. On the other hand, in the centrifugal hydraulic pressure canceling oil chamber 45B, an oil passage SSB formed in the axial direction of the secondary shaft SS, an oil passage SSD similarly formed in the radial direction, an oil passage groove 90E formed in the piston member 90, and a guide The hydraulic oil is supplied through a passage formed between the cylindrical portion 100B of the member 100 and the inner cylindrical portion 90B of the piston member 90.

  Further, in an embodiment according to an aspect of the present invention, as a member on the outer diameter side of the centrifugal hydraulic pressure canceling oil chamber 45B and connected to the movable sheave 43 and constituting a part of the centrifugal hydraulic pressure canceling oil chamber 45B. A communication hole 43D formed in the outer cylindrical portion 43C of the movable sheave 43 and an outer cylindrical portion 90C of the piston member 90 as a member connected to the fixed sheave 42 and constituting another part of the centrifugal oil pressure canceling oil chamber 45B. The communication hole 90 </ b> F formed in the above forms a discharge passage whose opening area is changed by the relative movement of the movable sheave 43 relative to the fixed sheave 42 in the axial direction. In this discharge passage, the amount of overlap between the communication hole 90F and the communication hole 43D changes as the movable sheave 43 moves to the left side on the deceleration side in FIG. The opening area for determining the flow rate of the hydraulic oil passing through the cylinder is increased. As a specific example, the discharge passage is closed when the speed ratio γ is from a minimum speed ratio γmin (for example, about 0.4) on the speed increasing side to a predetermined speed ratio γ0 (for example, about 0.8). The positions and diameters of the communication hole 90F and the communication hole 43D are set so that the opening area of the discharge passage increases from the speed ratio to the maximum speed ratio γmax on the deceleration side.

  The communication holes 90F and the communication holes 43D may be round holes. As another shape, the communication hole 90F is a small-diameter round hole, and the communication hole 43D has a short diameter equal to the diameter of the communication hole 90F and a longer diameter longer in the axial direction than the diameter of the communication hole 90F. It can also be formed as a long hole. As another example of the shape of the discharge passage, as shown in FIG. 4, the communication hole 90F of the piston member 90 on one side is formed as a rectangular long hole 90F ′ extending in the circumferential direction of the outer cylindrical portion 90C. The communication hole 43D of the movable sheave 43 on the other side may be formed as a triangular hole 43D ′ along the outer cylindrical portion 43C. However, in this case, the triangular hole 43D ′ is an isosceles triangle with respect to the axial direction, its apex angle is located on the long hole 90F ′ side, and its base is the long side of the rectangular long hole 90F ′. Have approximately equal lengths.

  Furthermore, the hydraulic circuit 300 in the embodiment of the present invention will be described with reference to FIG. In FIG. 5, reference numeral 21 denotes the above-described oil pump. In the present embodiment, hydraulic oil sucked from an oil tank or an oil pan and discharged from the oil pump 21 is supplied to an oil passage 302, which will be described later as a primary regulator valve. The pressure is adjusted to a predetermined line pressure PL by PRV. The oil passage 302 includes an electromagnetically controlled flow rate control valve 330 as a speed ratio control valve for controlling the speed ratio of the belt type continuously variable transmission 9 and the belt clamping pressure, and an electromagnetically controlled pressure reducing valve as a belt clamping pressure control valve. 340. The oil passage 302 is further connected to a clutch pressure control valve (not shown) that controls the engagement pressure of the forward clutch CL and the reverse brake BR of the forward / reverse switching mechanism 8 described above.

  The primary regulator valve PRV is provided in an oil passage 304 branched from the oil passage 302, and the secondary regulator valve SRV is disposed in series with a drain-type oil passage 306 of the primary regulator valve PRV. An oil passage 308 is branched from the oil passage 306, and the oil passage 308 is connected to an oil passage for lubrication and cooling. On the other hand, the drain oil passage of the secondary regulator valve SRV is a return oil passage 310 to the oil pump 21.

  An oil passage 309 is branched from the oil passage 306 together with the oil passage 308. The oil passage 309 is provided with a flow rate adjusting orifice 314 set to a hole diameter for controlling a required amount of hydraulic oil. ing. The oil passage 309 communicates with the centrifugal hydraulic pressure canceling oil chamber 45B of the secondary pulley 36 described above. Then, as will be described later, in this centrifugal hydraulic pressure cancellation oil chamber 45B, the secondary pressure controlled by the secondary regulator valve SRV described above is used as the supply source hydraulic pressure Pc, and the hydraulic oil at a predetermined flow rate is supplied to the oil passage SSB, the oil passage SSD, and the piston. It is configured to be supplied through an oil passage 90 </ b> E formed in the member 90.

  On the other hand, the hydraulic oil supplied to the oil passage 302 is provided in an oil passage 304 branched from the oil passage 302 and is controlled to a predetermined line pressure by a primary regulator valve PRV that is controlled according to a gear ratio, an input shaft torque, and the like. As described above, the pressure is adjusted to PL. The hydraulic oil drained from the primary regulator valve PRV to the oil passage 306 is then regulated to a predetermined secondary pressure by the secondary regulator valve SRV that is similarly controlled according to the gear ratio, input shaft torque, and the like.

  The hydraulic oil having the line pressure PL is supplied to the first and second control hydraulic chambers 40A and 40B of the primary hydraulic actuator 40 by the flow control valve 330. The flow control valve 330 controls the flow rate into the control hydraulic chambers 40A and 40B by a duty-controlled solenoid valve according to the vehicle speed and the accelerator opening, thereby controlling the gear ratio.

  Further, the hydraulic oil having the line pressure PL is controlled by the pressure reducing valve 340 to be the secondary pulley control hydraulic pressure Pdn, and is supplied to the control hydraulic pressure chamber 45A of the secondary hydraulic actuator 45. The pressure reducing valve 340 is controlled by a solenoid valve whose duty is controlled according to the input shaft torque, and controls the line pressure PL to the secondary pulley control oil pressure Pdn.

  Reference numeral 400 denotes a controller that controls the entire vehicle, and includes an arithmetic processing unit (CPU or MPU), a storage unit (RAM and ROM), and a microcomputer mainly including an input / output interface.

  For this controller 400, various parameters indicating the operating state of the engine 2, for example, information on the engine rotation speed Ne, the accelerator opening Ao indicating the depression amount of the accelerator pedal, and the throttle opening To, It is input as a signal of calculation or detection result from the speed sensor 402, the accelerator opening sensor 404, and the throttle opening sensor 406. Further, various parameters representing the state of the transaxle 3, such as the torque ratio of the torque converter 7, the rotational speed of the primary shaft SP (hereinafter referred to as the input shaft rotational speed) Nin, and the rotational speed of the secondary shaft SS (hereinafter referred to as output). Information such as the Nout, the vehicle speed V, the temperature of the hydraulic oil, that is, the oil temperature Temp, is obtained from the primary shaft rotational speed sensor 408, the secondary shaft rotational speed sensor 410, the vehicle speed sensor 412, and the oil temperature sensor 418. The solenoid valve described above is controlled to obtain a required gear ratio γ (= Nin / Nout) and belt clamping pressure based on a map or the like that is input as a calculation or detection result signal and obtained in advance through experiments or the like. A primary pulley control oil pressure Pdr and a secondary pulley control oil pressure Pdn are formed. Note that detection signals from the primary pressure sensor 414 and the secondary pressure sensor 416 that detect the primary pulley control hydraulic pressure Pdr and the secondary pulley control hydraulic pressure Pdn, respectively, are also input to the controller 400.

  Further, the controller 400 also stores data for performing shift control of the engine 2, the lockup clutch, and the belt type continuously variable transmission 9 based on various signals. For example, data for selecting the optimum operating state of the engine 2 by controlling the gear ratio of the belt-type continuously variable transmission 9 based on the traveling state such as the accelerator opening Ao and the vehicle speed V, A lock-up clutch control map using the accelerator opening Ao and the vehicle speed V as parameters is stored in the controller 400, and the lock-up clutch is controlled to each of engagement / release / slip states based on the lock-up clutch control map. The A control signal is output from the controller 400 based on various signals input to the controller 400 and data stored in the controller 400.

  The input shaft torque Tin to the belt type continuously variable transmission 9 is obtained from a map, for example, based on the throttle opening To and the engine speed Ne, and further from the input / output shaft speed of the torque converter 7. A speed ratio is calculated, a torque ratio is obtained from a map based on the speed ratio, and the engine torque Teng is multiplied by the torque ratio. Note that the output shaft rotational speed Nout is substantially proportional to the vehicle speed V.

  Here, the general operation of the belt type continuously variable transmission 9 according to the above embodiment will be described first. A predetermined discharge hydraulic pressure is generated in the oil passage 302 by the rotation of the oil pump 21 based on the rotation of the engine 2, and the hydraulic pressure is adjusted to the line pressure PL by controlling the primary regulator valve PRV. At the same time, the secondary pressure PS supplied to the torque converter 7 via the oil passage 308 is regulated to a predetermined pressure by the secondary regulator valve SRV arranged in series with the primary regulator valve PRV. Secondary regulator valve SRV adjusts the pressure while letting hydraulic oil escape to oil passage 310. Further, when the engine 2 is rotated, that is, when the input shaft SI is rotated, the various bearings 31, 32, 33, 34, the forward / reverse switching mechanism 8, and the lubrication parts of the final reduction gear 10 are provided via the oil passage 308. In addition, hydraulic oil controlled to a required flow rate by the orifice is supplied. At the same time, in the communication oil passage 309 branched from the oil passage 302, the flow rate of the hydraulic oil controlled to the secondary pressure PS as the predetermined supply oil pressure Pc by the secondary regulator valve SRV is controlled by the orifice 314, and the oil passage SSB, The oil is supplied to the centrifugal oil pressure cancellation oil chamber 45B through the oil passage SSD and the oil passage 90E formed in the piston member 90.

  Therefore, control by the hydraulic control apparatus according to the embodiment of the present invention will be described with reference to the flowchart of FIG. When the control is started, in step S601, the current input shaft rotational speed Nin (t), output shaft rotational speed Nout (t) and input shaft torque Tin of the belt type continuously variable transmission 9 are read. In the next step S602, the moving speed Vm of the movable sheave 43 of the secondary pulley 36 is calculated based on a shift speed Sv described later. That is, the current gear ratio γ (t) (= Nin (t) / Nout (t)) is obtained from the current input shaft rotational speed Nin (t) and output shaft rotational speed Nout (t) read in step S601. Desired. Further, from the speed ratio γ (t−Δt) in the previous routine cycle executed Δt time before the current time and the current speed ratio γ (t), the speed Sv of the movable sheave 43 becomes “Sv = { γ (t) −γ (t−Δt)} / Δt) ”. Therefore, the transmission speed Sv and the specifications of the secondary pulley 36, for example, the sandwiching angle of the substantially V-shaped pulley groove 44 formed between the fixed sheave 42 and the movable sheave 43 (2 of the cone cone inclination angle of the pulley). The moving speed Vm of the movable sheave 43 of the secondary pulley 36 is obtained from the above. The moving speed Vm of the movable sheave 43 is positive when the movable sheave 43 moves to the speed increasing side (right side in FIG. 3), that is, moves to the side where the gear ratio γ decreases.

In the next step S603, it is determined whether or not the current speed ratio γ (t) is greater than the above-described predetermined speed ratio γ0 at which communication of the discharge passage is started due to the overlap of the communication hole 90F and the communication hole 43D. Determined. As a result of the determination, when the current gear ratio γ (t) is larger than the communication start gear ratio γ0, that is, when the hydraulic oil is discharged from the discharge passage, the process proceeds to step S604, where the centrifugal oil pressure canceling oil chamber 45B The hydraulic oil discharge amount Qout is calculated by the following equation (1).
(1) Qout = kc1 × Nout × √ {(R1 ² −r (t−Δt) ² )}

Here, r (t−Δt) is the oil surface diameter at the time of the previous routine cycle executed before Δt time from the present time, and R1 is the rotation axis of the centrifugal oil pressure canceling oil chamber 45B as shown in FIG. Is the outermost radius. (R2 mentioned later is the innermost radius of the centrifugal oil pressure canceling oil chamber 45B.)
On the other hand, as a result of the determination in step S603, when the current speed ratio γ (t) is smaller than the communication start speed ratio γ0, that is, the communication hole 90F and the communication hole 43D do not overlap with each other, the centrifugal oil pressure cancellation oil chamber 45B When the hydraulic oil is not discharged, the process proceeds to step S605, and the hydraulic oil discharge amount Qout from the centrifugal oil pressure canceling oil chamber 45B is set to 0.

After the hydraulic oil discharge amount Qout is obtained in step S604 or step S605 described above, the process proceeds to step S606, where the current oil surface diameter r (t) of the hydraulic oil in the centrifugal hydraulic pressure canceling oil chamber 45B is set. Then, it is calculated and estimated by the following equation (2).
(2) r (t) = r (t−Δt) + Vm × k5 × Δt− (Qin−Qout) × k6 × Δt

  Here, k5 and k6 are coefficients determined by the shape of the secondary pulley 36, respectively. However, if the current oil surface diameter r (t) of the hydraulic oil is smaller than the innermost radius R2 of the centrifugal oil pressure canceling oil chamber 45B as a result of the calculation by the equation (2), r (t) = R2. Further, Qin is the amount of hydraulic oil supplied to the centrifugal hydraulic pressure canceling oil chamber 45B. In this embodiment, the secondary pressure PS is the supply hydraulic pressure Pc, and the orifice 314 is controlled to a constant flow rate. .

In the next step S607, when the hydraulic oil is filled in the centrifugal oil pressure canceling oil chamber 45B, the secondary oil required in the control oil pressure chamber 45A to obtain a predetermined belt clamping pressure in the secondary pulley 36 is obtained. The pulley necessary control hydraulic pressure PdnA is calculated by the following equation (3).
(3) PdnA × k2 + f + Nout ² × k3 = Tin × k1

  Here, k1 and k2 are coefficients determined by the specifications of the secondary pulley 36, k3 is a coefficient determined by the specifications of the pulley and the density of the hydraulic oil, and f is a biasing force by the spring 96.

Further, the process proceeds to step S608, and when the estimated oil surface diameter of the hydraulic oil in the centrifugal oil pressure canceling oil chamber 45B is the above-described r (t), the control oil pressure is obtained in order to obtain a predetermined belt clamping pressure by the secondary pulley 36. The secondary pulley required control hydraulic pressure PdnB required in the chamber 45A is calculated by the following equation (4).
(4) PdnB × k2 + f + Nout ² × k3−Nout ² × k4 × {(R1 ² −R2 ² ) ² − (R1 ² −r (t) ² ) ² } = Tin × k1

  Here, R1 is the outermost radius from the rotation axis of the centrifugal hydraulic pressure canceling oil chamber 45B as described above, and R2 is the innermost radius (see FIG. 3). K4 is a coefficient determined by the density of the hydraulic oil.

  Then, in the next step S609, it is determined whether or not the input shaft torque Tin is greater than or equal to a predetermined value, or whether or not the output shaft rotational speed Nout is less than or equal to a predetermined value. When the input shaft torque Tin is equal to or greater than the predetermined value, or when the output shaft rotation speed Nout is equal to or smaller than the predetermined value, the process proceeds to step S610, and the hydraulic pressure supplied to the control hydraulic pressure chamber 45A of the secondary pulley 36 is determined in step S608 described above. The control value of the electromagnetically controlled pressure reducing valve 340 is set so as to be the required secondary pulley necessary control oil pressure PdnB.

  On the other hand, it is determined in step S609 that the input shaft torque Tin is not less than the predetermined value or the output shaft rotation speed Nout is not less than the predetermined value. In other words, the input shaft torque Tin is less than the predetermined value and the output shaft rotation speed Nout. When the pressure exceeds the predetermined value, the process proceeds to step S611, and the electromagnetically controlled pressure reducing valve is set so that the hydraulic pressure supplied to the control hydraulic chamber 45A of the secondary pulley 36 becomes the secondary pulley required control hydraulic pressure PdnA obtained in step S607 described above. A control value of 340 is set. Note that the predetermined value of the output shaft rotation speed Nout, which is the determination element used in step S609, is set so as to substantially correspond to the above-described communication start speed ratio γ0.

  Thus, when the input shaft torque Tin, in other words, the engine load is equal to or higher than the predetermined value, or the output shaft speed Nout, in other words, when the vehicle speed V is equal to or lower than the predetermined value, for example, at the so-called underdrive, the gear ratio γ Is larger than the communication start speed ratio γ0, the hydraulic oil discharge amount Qout in the centrifugal hydraulic pressure canceling oil chamber 45B increases as compared with the supply amount Qin, and the oil surface diameter r is increased. That is, the hydraulic oil filling amount in the centrifugal hydraulic pressure cancellation oil chamber 45B is reduced. As a result, the required hydraulic pressure to be supplied to the control hydraulic chamber 45A of the secondary pulley 36 can be set to a value that does not generate centrifugal hydraulic pressure in the centrifugal hydraulic pressure canceling oil chamber 45B, that is, the required control hydraulic pressure PdnB. Drive loss can be reduced.

  On the other hand, when the input shaft torque Tin is less than the predetermined value and the output shaft rotational speed Nout exceeds the predetermined value, in other words, for example, when the speed ratio γ is smaller than the communication start speed ratio γ0 during so-called overdrive, The discharge amount Qout becomes 0 while the supply amount Qin of the hydraulic oil to the cancel oil chamber 45B remains constant, and the oil surface diameter r is reduced. That is, the hydraulic oil filling amount in the centrifugal hydraulic pressure canceling oil chamber 45B is increased. Thereby, the centrifugal hydraulic pressure generated in the control hydraulic chamber 45A of the secondary pulley 36 is sufficiently canceled by the centrifugal hydraulic pressure generated in the centrifugal hydraulic pressure canceling oil chamber 45B. At this time, as described above, the required hydraulic pressure to be supplied to the control hydraulic chamber 45A is set to the required control hydraulic pressure PdnA.

  As described above, the hydraulic control device for the belt type continuously variable transmission according to the present invention has been described with respect to the embodiment in which the discharge passage is arranged on the outer diameter side of the centrifugal hydraulic pressure canceling oil chamber provided in the secondary pulley. Of course, in the case of having a centrifugal hydraulic pressure canceling oil chamber, the discharge passage may be arranged in the centrifugal hydraulic pressure canceling oil chamber of the primary pulley.

It is a skeleton figure which shows the transaxle to which the hydraulic control apparatus of the belt-type continuously variable transmission which concerns on this invention is applied. It is sectional drawing which shows an example of a structure of the primary pulley of the hydraulic control apparatus of the belt-type continuously variable transmission which concerns on this invention. 1 is a half sectional view showing a configuration of an embodiment of a secondary pulley of a hydraulic control device for a belt type continuously variable transmission according to the present invention. It is a top view which shows the other structural example of the discharge passage of the hydraulic control apparatus of the belt-type continuously variable transmission which concerns on this invention. It is a block diagram which shows an example of the hydraulic circuit of the hydraulic control apparatus of the belt-type continuously variable transmission which concerns on this invention. It is a flowchart which shows an example of the control routine by the hydraulic control apparatus of embodiment of this invention.

Explanation of symbols

30 Primary shaft 31 Secondary shaft 36 Primary pulley 37 Secondary pulley 38 Primary side fixed sheave 39 Primary side movable sheave 42 Secondary side fixed sheave 43 Secondary side movable sheave 43D, 43D 'Communication hole (discharge passage)
45A Control hydraulic chamber 45B Centrifugal hydraulic cancel oil chamber 90 Piston member 90F, 90F 'Communication hole (discharge passage)
300 Hydraulic circuit 314 Flow rate adjusting orifice 330 Flow rate control valve 340 Pressure reducing valve 400 Controller Nin Input shaft speed Nout Output shaft speed Pc Supply source oil pressure Pdr Primary pulley control oil pressure Pdn Secondary pulley control oil pressure PdnA, PdnB Secondary pulley necessary control oil pressure r ( t) Oil surface diameter Tin Input shaft torque γ Gear ratio (= Nin / Nout)
γmax Maximum gear ratio γmin Minimum gear ratio γ0 Communication start gear ratio

Claims (3)

A control hydraulic pressure chamber that is supplied with a control hydraulic pressure that generates a belt clamping pressure acting on the belt to the movable sheave, a centrifugal hydraulic pressure cancellation oil chamber, a discharge passage that discharges hydraulic oil in the centrifugal hydraulic pressure cancellation oil chamber, and the centrifugal hydraulic pressure In a hydraulic control device for a belt-type continuously variable transmission, comprising a supply means for supplying hydraulic oil to a cancel oil chamber,
The discharge passage is on the outer diameter side of the centrifugal hydraulic pressure canceling oil chamber, and is connected to the movable sheave and constituting a part of the centrifugal hydraulic pressure canceling oil chamber. A hydraulic control device for a belt-type continuously variable transmission, wherein an opening area is variably formed by axial relative movement with a member constituting another part of the chamber.   2. The belt-type continuously variable transmission according to claim 1, wherein the discharge passage is configured such that an opening area thereof increases as the movable sheave moves in a direction in which a speed ratio increases. Hydraulic control device. The control oil pressure supplied to the control oil pressure chamber is controlled by decreasing the amount of hydraulic oil in the centrifugal oil pressure canceling oil chamber that decreases as the movable sheave moves in the direction in which the gear ratio increases. The hydraulic control device for a belt type continuously variable transmission according to claim 2.

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