JP2012237370A - Toroidal continuously variable transmission - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a toroidal continuously variable transmission that can suppress both slip of a power roller and cost increase.SOLUTION: The toroidal continuously variable transmission includes: a plurality of power rollers 22 that transmit power between an input disk 20b and an output disk 21b and include a first power roller and a second power roller; a first pressure output mechanism 101 that outputs a first pressure P1 according to a slant angle of the first power roller; a second pressure output mechanism 102 that outputs a second pressure P2 according to a slant angle of the second power roller, and a signal pressure output mechanism 103 that outputs a signal pressure Ps for controlling the slip of the power roller according to a differential pressure between the first pressure and the second pressure.

Description

  The present invention relates to a toroidal continuously variable transmission.

  Conventionally, a technique for calculating a slip ratio at a contact portion between a disk and a power roller in a toroidal-type continuously variable transmission has been disclosed. For example, in Patent Document 1, an average slip ratio is calculated based on the input / output disk rotation speed and the roller tilt angle, and when the average slip ratio is equal to or greater than a predetermined value, the input / output disk and the roller are A technique for a toroidal CVT control device that determines that a traction coefficient is in the vicinity of a maximum value is disclosed. According to the toroidal CVT control device of Patent Document 1, it is possible to accurately predict the macro slip between the disk and the roller and prevent the slip.

JP 2005-344867 A

  There is still room for improvement in suppressing slip in a toroidal continuously variable transmission. For example, if slip can be suppressed without requiring a sensor for detecting the tilt angle, an increase in cost for realizing slip suppression can be suppressed. In a toroidal-type continuously variable transmission, it is desired that both suppression of slip and suppression of increase in cost can be achieved.

  An object of the present invention is to provide a toroidal-type continuously variable transmission that can both suppress slippage of a power roller and suppress cost increase.

  The toroidal continuously variable transmission of the present invention transmits power between an input disk and an output disk, and includes a plurality of power rollers including a first power roller and a second power roller, and a tilt of the first power roller. A first pressure output mechanism that outputs a first pressure according to a turning angle; a second pressure output mechanism that outputs a second pressure according to a tilt angle of the second power roller; the first pressure and the first pressure And a signal pressure output mechanism that outputs a signal pressure that suppresses slippage of the power roller in accordance with a differential pressure between the two pressures.

  The toroidal continuously variable transmission further includes a clamping pressure correction mechanism that corrects a clamping pressure in which the input disk and the output disk pinch the plurality of power rollers, and the clamping pressure correction mechanism is configured to adjust the signal pressure. Accordingly, it is preferable to increase the clamping pressure.

  In the toroidal continuously variable transmission, it is preferable that the signal pressure output mechanism outputs the signal pressure when a differential pressure between the first pressure and the second pressure is greater than or equal to a threshold value.

  In the toroidal continuously variable transmission, it is preferable that the output of a prime mover that outputs power to the input disk is further reduced based on a differential pressure between the first pressure and the second pressure.

  A toroidal continuously variable transmission according to the present invention transmits power between an input disk and an output disk, and includes a plurality of power rollers including a first power roller and a second power roller, and an inclination of the first power roller. A first pressure output mechanism that outputs a first pressure according to the turning angle, a second pressure output mechanism that outputs a second pressure according to the tilt angle of the second power roller, a first pressure and a second pressure, A signal pressure output mechanism that outputs a signal pressure that suppresses slipping of the power roller in accordance with the differential pressure. According to the toroidal continuously variable transmission according to the present invention, when the power roller slips, each mechanism automatically generates a signal pressure to suppress the slip, thereby suppressing the power roller slip and increasing the cost. There is an effect that it is possible to achieve both suppression.

FIG. 1 is a schematic cross-sectional view of a toroidal continuously variable transmission according to an embodiment. FIG. 2 is a schematic cross-sectional view around the pair of power rollers of the toroidal continuously variable transmission according to the embodiment. FIG. 3 is a schematic configuration diagram of a hydraulic circuit according to the hydraulic control apparatus of the embodiment. FIG. 4 is a schematic configuration diagram of the clamping pressure correction system according to the embodiment. FIG. 5 is an operation explanatory diagram of the first pressure output mechanism. FIG. 6 is a schematic configuration diagram of the clamping pressure correction system according to the first modification. FIG. 7 is a flowchart of control according to the first modification. FIG. 8 is a schematic configuration diagram of the clamping pressure correction system according to the second modification.

  Hereinafter, a toroidal continuously variable transmission according to an embodiment of the present invention will be described in detail with reference to the drawings. In addition, this invention is not limited by this embodiment. In addition, constituent elements in the following embodiments include those that can be easily assumed by those skilled in the art or those that are substantially the same.

[Embodiment]
The embodiment will be described with reference to FIGS. 1 to 5. The present embodiment relates to a toroidal continuously variable transmission. FIG. 1 is a schematic cross-sectional view of a toroidal continuously variable transmission according to the embodiment, FIG. 2 is a schematic cross-sectional view around a pair of power rollers of the toroidal continuously variable transmission according to the embodiment, and FIG. It is a schematic block diagram of the hydraulic circuit which concerns on this hydraulic control apparatus.

  The toroidal continuously variable transmission according to the present embodiment includes a mechanism that detects one slip among a plurality of power rollers and performs fail-safe. Specifically, a first pressure output mechanism (see reference numeral 101 in FIG. 4) that outputs a first pressure P1 corresponding to a tilt angle (speed ratio) of one power roller, and a tilt of the other power roller. A second pressure output mechanism (see reference numeral 102 in FIG. 4) that outputs a second pressure P2 corresponding to the rotation angle (speed ratio), and a power roller according to a differential pressure between the first pressure P1 and the second pressure P2. A signal pressure output mechanism (reference numeral 103 in FIG. 4) that outputs a signal pressure Ps that suppresses slipping is provided. Thereby, according to the toroidal-type continuously variable transmission of this embodiment, it is possible to achieve both slip suppression of the power roller and cost reduction.

  As shown in FIG. 1, the vehicle 1 includes an engine 5 serving as a drive source, a torque converter 6 coupled to the engine 5, a forward / reverse switching mechanism 7 coupled to the torque converter 6, and a forward / reverse switching mechanism 7. And a toroidal continuously variable transmission 8 connected thereto. A reduction gear 9 is connected to the toroidal continuously variable transmission 8, a differential device 10 is connected to the reduction gear 9, and a drive wheel 11 is connected to the differential device 10.

  For example, a gasoline engine or a diesel engine is used as the engine 5, and the piston reciprocates in the central axis direction of the cylinder formed in a cylindrical shape, and the crankshaft 15 converts the reciprocating motion of the piston into a rotational motion. Outputs driving force. The engine 5 is a prime mover that outputs power to the input disks 20 a and 20 b of the toroidal continuously variable transmission 8. In addition, it is not restricted to said structure, Instead of the engine 5 as a motor | power_engine, electric motors, such as a motor, may be used, and an engine and an electric motor may be used together.

  The torque converter 6 is a kind of fluid clutch, and transmits the driving force output from the engine 5 to the forward / reverse switching mechanism 7 via hydraulic oil. The torque converter 6 has a lockup mechanism, for example, and increases the output torque from the engine 5 or transmits it to the forward / reverse switching mechanism 7 with the output torque as it is.

  The toroidal continuously variable transmission 8 changes the rotational speed of the driving force input from the forward / reverse switching mechanism 7 to a desired rotational speed according to the driving state of the vehicle, and outputs it. The detailed description of the toroidal continuously variable transmission 8 will be described later.

  The reduction gear 9 reduces the rotational speed of the driving force input from the toroidal continuously variable transmission 8 and transmits the driving force to the differential device 10. The differential device 10 absorbs the speed difference between the center driving side, that is, the inner driving wheel 11 and the outer driving wheel 11 generated when the vehicle 1 turns.

  Therefore, when the engine 5 is driven in the vehicle 1, the driving force output from the engine 5 is transmitted to the torque converter 6 via the crankshaft 15. The driving force obtained by amplifying the output torque by the torque converter 6 is transmitted to the forward / reverse switching mechanism 7 and is changed to a desired rotation direction by the forward / backward switching mechanism 7. The driving force in the desired rotational direction is input to the toroidal continuously variable transmission 8, and the rotational speed of the input driving force is changed according to a predetermined speed ratio of the toroidal continuously variable transmission 8. The The driving force whose rotational speed has been changed by the toroidal-type continuously variable transmission 8 is input to the reduction gear 9, and the input driving force is decelerated by the reduction gear 9 and then output to the differential device 10. Then, when the differential device 10 transmits the input driving force to the driving wheels 11, the driving wheels 11 rotate, and thereby the vehicle 1 travels.

  Next, the toroidal continuously variable transmission 8 will be described. The toroidal continuously variable transmission 8 converts the driving force (output rotational speed and output torque) from the engine 5 into an appropriate driving force according to the traveling state of the vehicle 1 and transmits it to the drive wheels 11. There is a so-called CVT in which the gear ratio can be controlled steplessly (continuously).

  As shown in FIGS. 1 and 2, the toroidal continuously variable transmission 8 includes a pair of input disks 20a and 20b to which driving force is input via an input shaft 28, and a pair of input disks 20a and 20b. And a pair of output disks 21a and 21b disposed to face the input disks 20a and 20b, and between the input disks 20a and 20b and the output disks 21a and 21b, respectively. And two pairs of power rollers 22 as a pair. The plurality of power rollers 22 transmit power between the input disk 20a and the output disk 21a and between the input disk 20b and the output disk 21b.

  Each power roller 22 (22a, 22b, 22c, 22d) is rotatably supported by each trunnion 23. The toroidal-type continuously variable transmission 8 also includes a disk pressing mechanism 24 that causes the input disks 20a and 20b and the output disks 21a and 21b to approach each other and clamps the power rollers 22, and each trunnion 23 with a swing shaft 61a. , 61b (see FIG. 2) and a hydraulic servo mechanism 25 that moves each power roller 22 between the neutral position and the shift position, and the disk pressing mechanism 24 and the hydraulic servo mechanism 25 are It is controlled by the hydraulic control device 26.

  In the toroidal continuously variable transmission 8, each power roller 22 is sandwiched between each input disk 20 a, 20 b and each output disk 21 a, 21 b by the disk pressing mechanism 24, and in this state, each trunnion 23 is driven by the hydraulic servo mechanism 25. Can be moved along the rocking shafts 61a and 61b. As a result, each power roller 22 moves from the neutral position to the speed change position, whereby each power roller 22 tilts, and the speed ratio, which is the rotational speed ratio between the input disks 20a, 20b and the output disks 21a, 21b, is changed. Is done.

  The pair of input disks 20 a and 20 b are pivotally supported on a variator shaft 30 connected to the input shaft 28 of the toroidal continuously variable transmission 8. The pair of input disks 20a and 20b rotate integrally with the variator shaft 30 about the axis X1. The pair of input disks 20a and 20b is provided with a front side input disk 20a on the front side (engine 5 side) with respect to the variator shaft 30, and the rear side of the pair of input disks 20a and 20b is spaced from the front side input disk 20a by a predetermined distance. The rear side input disk 20b is provided on the side (drive wheel 11 side).

  Each of the input disks 20a and 20b is formed in a disk shape protruding outward in the radial direction of the variator shaft 30, and an opening through which the variator shaft 30 is inserted is formed at the axis. And the toroidal surface 33a which each contacts with each power roller 22 is formed in the surface facing each output disk 21a, 21b of each input disk 20a, 20b.

  The front-side input disk 20a is supported by the variator shaft 30 via a ball spline 35. For this reason, the front-side input disk 20 a is restricted from being displaced in the circumferential direction with respect to the variator shaft 30, but is allowed to be displaced in the axial direction with respect to the variator shaft 30. On the other hand, the rear side input disk 20b is supported by the variator shaft 30 through the spline 36, and movement in the axial direction is restricted by a snap ring 37 provided at the rear side end of the variator shaft 30. . For this reason, the rear-side input disk 20b is restricted from being displaced in the circumferential direction with respect to the variator shaft 30, while being restricted from being displaced in the axial rear direction with respect to the variator shaft 30, Displacement is allowed. Although the details will be described later, the front side input disk 20a is pressed axially rearward with respect to the variator shaft 30 by the disk pressing mechanism 24, and the rear side input disk 20b is applied to the variator shaft 30 by the disk pressing mechanism 24. On the other hand, it is pressed to the front side in the axial direction.

  The pair of output disks 21a and 21b are rotatably supported with respect to the variator shaft 30, and rotate about the axis X1. The pair of output disks 21a and 21b are disposed between the pair of input disks 20a and 20b, and the front-side output disk 21a is provided on the front side with respect to the variator shaft 30, and the front-side output disk A rear output disk 21b is provided on the rear side at a predetermined interval with respect to 21a.

  Similarly to the input disks 20a and 20b, each of the output disks 21a and 21b is formed in a disk shape protruding outward in the radial direction of the variator shaft 30, and an opening through which the variator shaft 30 is loosely inserted is formed at the axis. Is formed. And the toroidal surface 33b which each contacts with each power roller 22 is formed in the surface facing each input disk 20a, 20b of each output disk 21a, 21b.

  The front output disk 21a and the rear output disk 21b are supported by the variator shaft 30 via bearings. An output gear 40 is connected between the pair of output disks 21a and 21b, and the output gear 40 rotates integrally with the pair of output disks 21a and 21b. A counter gear 41 is engaged with the output gear 40, and an output shaft 42 is connected to the counter gear 41. Accordingly, the output gears 40 are rotated by the rotation of the output disks 21a and 21b, and the output shaft 42 is rotated by transmitting the rotation from the output gear 40 to the counter gear 41. The output shaft 42 is connected to the drive wheel 11 via the speed reducer 9, the differential device 10, and the like.

  Here, around the input shaft 28, a disk pressing mechanism 24 is disposed to bring the input disks 20a and 20b closer to the output disks 21a and 21b. That is, the disk pressing mechanism 24 is configured to sandwich the pair of power rollers 22 between the input disks 20a and 20b and the output disks 21a and 21b, thereby applying a clamping pressure. The disk pressing mechanism 24 includes a clamping pressure generating hydraulic chamber 45 and a clamping pressure piston 46.

  The clamping pressure generating hydraulic chamber 45 is provided on the front side input disk 20a side, and is disposed between the input shaft 28 and the front side input disk 20a. Although details will be described later, hydraulic oil is supplied to the clamping pressure generating hydraulic chamber 45 from the disk pressing hydraulic system 83 of the hydraulic control device 26 (see FIG. 3).

  The clamping pressure piston 46 is formed in a disc shape, and is provided at one end on the front side of the variator shaft 30 so that the center thereof substantially coincides with the axis X1. Specifically, a flange portion 47 is formed at the front side end portion of the variator shaft 30, and the center portion of the front side end surface of the pressing pressure piston 46 is brought into contact with the flange portion 47, and the pressing pressure force An input shaft 28 is coupled to the outer peripheral portion of the front end surface of the piston 46. That is, the flange portion 47 of the variator shaft 30 is disposed between the input shaft 28 and the clamping pressure piston 46. The clamping pressure piston 46 is arranged with a predetermined gap between the flange portion 47 of the variator shaft 30 and the front input disk 20a in the X1 axis direction. As a result, the above-described clamping pressure generating hydraulic chamber 45 is formed between the clamping pressure piston 46 and the front input disk 20a.

  The clamping pressure piston 46 is provided so as to be rotatable about the axis X1 together with the variator shaft 30 and to be movable in a direction along the axis X1. As a result, the rear side input disk 20b, the variator shaft 30 and the clamping pressure piston 46 are configured integrally with the input shaft 28 so as to be rotatable about the axis X1, and are movable in the direction along the axis X1. Composed. The front-side input disk 20a can be rotated together with the rear-side input disk 20b, the variator shaft 30 and the clamping pressure piston 46 around the axis X1. The input disk 20b, the variator shaft 30 and the clamping pressure piston 46 can move relative to each other in the direction along the axis X1.

  Further, the front side input disk 20a has a front side end surface serving as a first hydraulic working surface 50 for moving the front side input disk 20a, and the clamping pressure piston 46 has a rear side end surface thereof. The second working surface 51 is a hydraulic pressure for moving the rear input disk 20b. That is, a part of the clamping pressure generating hydraulic chamber 45 is partitioned by the first working surface 50 and the second working surface 51.

  Accordingly, the disk pressing mechanism 24 moves the front side input disk 20a to the rear side in the axial direction of the axis X1 by the hydraulic pressure of the hydraulic oil supplied into the clamping pressure generating hydraulic chamber 45, and the clamping pressure piston 46 is moved. Move to the front side in the axial direction of the axis X1. As a result, the pair of power rollers 22a and 22b disposed between the front input disk 20a and the front output disk 21a are sandwiched between the front input disk 20a and the front output disk 21a. On the other hand, the pair of power rollers 22c and 22d disposed between the rear side input disk 20b and the rear side output disk 21b is moved by the pressing pressure piston 46 moving toward the axial front side of the axis X1. Since the rear side input disk 20b moves to the front side, it is sandwiched between the rear side input disk 20b and the rear side output disk 21b. As a result, the toroidal continuously variable transmission 8 can transmit power between the input disks 20a, 20b and the output disks 21a, 21b via the power rollers 22.

  Next, the two pairs of power rollers 22 are disposed between the input disks 20a and 20b and the output disks 21a and 21b, respectively. Specifically, one pair of power rollers 22a and 22b is disposed between the front side input disk 20a and the front side output disk 21a, and the other pair of power rollers 22c and 22d is a rear side input disk. It is arranged between 20b and the rear output disk 21b. Each power roller 22 rolls while being sandwiched between the input disks 20a and 20b and the output disks 21a and 21b by the disk pressing mechanism 24, so that the output disks 21a and 20b are transferred from the input disks 20a and 20b. Driving force is transmitted to 21b, or driven force is transmitted from each output disk 21a, 21b to each input disk 20a, 20b. At this time, each power roller 22 is connected between each power roller 22 and each input disk 20a, 20b and between each power roller 22 and each output disk 21a, 21b by traction oil supplied to the toroidal-type continuously variable transmission 8. The driving force or driven force is transmitted using the shear force of the oil film formed between the two.

  More specifically, each power roller 22 includes a power roller body 55 and an outer ring 56. The outer peripheral surface of the power roller body 55 is formed on a curved contact surface 57 corresponding to the toroidal surfaces 33a and 33b of the input disks 20a and 20b and the output disks 21a and 21b. As shown in FIG. 2, the power roller main body 55 is rotatably supported by a rotation shaft 58 formed on the outer ring 56 via a radial bearing RB. Further, the power roller main body 55 is rotatably supported on the surface of the outer ring 56 facing the power roller main body 55 via a thrust bearing SB. Accordingly, the power roller main body 55 can rotate around the axis X2 of the rotation shaft 58 of the outer ring 56 as a rotation center.

  An eccentric shaft 59 is formed on the outer ring 56 together with the rotating shaft 58 described above. The eccentric shaft 59 is formed such that its axis X3 is shifted from the axis X2 of the rotation shaft 58. The eccentric shaft 59 is rotatably supported by a trunnion 23 described later via a radial bearing RB. Accordingly, the outer ring 56 can rotate around the axis X3 of the eccentric shaft 59. That is, each power roller 22 can revolve about the axis X3 and can rotate about the axis X2.

  Each trunnion 23 that rotatably supports each power roller 22 moves each power roller 22 between a neutral position and a shift position with respect to each input disk 20a, 20b and each output disk 21a, 21b. . Each trunnion 23 has a main body 60 and swing shafts 61a and 61b.

  A space portion in which the power roller 22 is disposed is formed in the main body portion 60. The main body 60 is integrally formed with swing shafts 61 a and 61 b at both ends in the moving direction of the power roller 22. The swing shafts 61a and 61b are formed in a columnar shape, and pivotally support the main body portion 60 about the axis X4 as a rotation center. One swinging shaft 61 a extending downward in the figure is supported by a casing (not shown) via a lower link 62, and the other swinging shaft 61 b extending upward in the drawing is supported by a casing (not shown) via an upper link 63. (Shown). Further, the swing shaft 61a is supported by the lower link 62 via the radial bearing RB so as to be rotatable and movable in the axial direction, and the swing shaft 61b is rotatably supported by the upper link 63 via the radial bearing RB. It is supported movably in the direction. Therefore, the trunnion 23 is supported so that the main body 60 can rotate about the axis X4 together with the swing shafts 61a and 61b, and can move in the direction along the axis X4.

  Here, a hydraulic servo mechanism 25 that moves each trunnion 23 in the X4 axis direction is disposed around the swing shafts 61a and 61b of each trunnion 23. That is, the hydraulic servo mechanism 25 moves each power roller 22 from the neutral position to the shift position via each trunnion 23. The hydraulic servo mechanism 25 includes a transmission control hydraulic chamber 65 and a transmission control piston 66. The hydraulic control mechanism 66 receives the hydraulic pressure of the hydraulic oil introduced into the transmission control hydraulic chamber 65 by the transmission control piston 66. The trunnion 23 is moved in two directions (A1 direction and A2 direction) along the axis X4.

  The speed change control piston 66 includes a piston base 70 and a flange portion 71. The piston base 70 is formed in a cylindrical shape and is inserted into and fixed to one swinging shaft 61a. The flange portion 71 protrudes from the piston base 70 to the outer side in the radial direction of the swing shaft 61 a and is formed in a disk shape and is fixedly provided on the piston base 70.

  The shift control hydraulic chamber 65 is formed in the cylinder body 75, and the cylinder body 75 is configured by a body main body 76 and a lower cover 77 disposed at the bottom of the body main body 76. The body main body 76 is formed with a cylindrical hollow recess serving as a transmission control hydraulic chamber 65 at the bottom. The lower cover 77 is fixed to the bottom of the body main body 76 so as to close the opening of the concave portion of the body main body 76. As a result, the transmission control hydraulic chamber 65 is partitioned by the body main body 76 and the lower cover 77 into a columnar shape (cylinder shape) centered on the axis X4.

  The flange portion 71 of the transmission control piston 66 is housed in the transmission control hydraulic chamber 65 and divides the transmission control hydraulic chamber 65 into a first hydraulic chamber OP1 and a second hydraulic chamber OP2. At this time, the first hydraulic chamber OP1 is formed on the A2 direction side along the axis line X4, and the second hydraulic chamber OP2 is formed on the A1 direction side along the axis line X4.

  Accordingly, when the hydraulic oil is supplied into the first hydraulic chamber OP1, the flange portion 71 is moved in the A1 direction along the axis X4 by the hydraulic pressure, while the hydraulic oil is supplied into the second hydraulic chamber OP2. Then, the flange portion 71 is moved in the A2 direction along the axis X4 by hydraulic pressure. Thereby, the hydraulic servo mechanism 25 shifts each power roller 22 from the neutral position with respect to each input disk 20a, 20b and each output disk 21a, 21b by moving each trunnion 23 in the direction along the axis X4. Each power roller 22 can be tilted with respect to each of the input disks 20a, 20b and each of the output disks 21a, 21b by moving to a shift position corresponding to the ratio. Thereafter, the change of the gear ratio is completed by returning each power roller 22 to the neutral position again.

  An annular seal member S is provided at the distal end of the flange portion 71 in the radial direction, and the hydraulic oil inside the first hydraulic chamber OP1 and the second hydraulic chamber OP2 leaks from the annular seal member S to each other. It is sealed so that there is no. In the pair of trunnions 23 that support the pair of power rollers 22, the first hydraulic chamber OP <b> 1 and the second hydraulic chamber OP <b> 2 formed around the swing shaft 61 a of one trunnion 23 are swung by the other trunnion 23. The positional relationship is reversed with respect to the first hydraulic chamber OP1 and the second hydraulic chamber OP2 formed around the shaft 61a.

  The disk pressing mechanism 24 and the hydraulic servo mechanism 25 are controlled by a hydraulic control device 26. As shown in FIG. 3, the hydraulic control device 26 includes an oil tank 80 that stores hydraulic oil, an oil pump 81 that is driven by the engine 5, and a line pressure that supplies hydraulic oil pumped up by the oil pump 81 toward each part. A system 82, a disk pressing hydraulic system 83 connected to the line pressure system 82, and a servo hydraulic system 84 connected to the line pressure system 82 are provided. For example, the servo hydraulic system 84 is connected to the upstream side of the line pressure system 82 as compared to the disk pressing hydraulic system 83.

  The oil pump 81 operates in conjunction with the operation of the engine 5, for example, the rotation of the crankshaft 15, and sucks, pressurizes, and discharges the hydraulic oil stored in the oil tank 80. The pressurized hydraulic fluid discharged is supplied to the disk pressing hydraulic system 83, the servo hydraulic system 84, and the like via the line pressure system 82.

  The line pressure system 82 supplies hydraulic oil discharged from the oil pump 81 to each part, for example, the forward / reverse switching mechanism 7, the disk pressing mechanism 24, the hydraulic servo mechanism 25, and the like. The hydraulic pressure in the line pressure system 82 is maintained at a predetermined hydraulic pressure (line pressure) by supplying hydraulic oil from the oil pump 81 during operation of the engine 5.

  The disc pressing hydraulic system 83 is a drive source (hydraulic source) of the disc pressing mechanism 24 and connects the line pressure system 82 and the pinching pressure generating hydraulic chamber 45. The disc pressing hydraulic system 83 is provided with a disc pressing control valve 88 that adjusts the flow rate of hydraulic oil supplied to the clamping pressure generating hydraulic chamber 45.

  The disk pressing control valve 88 is controlled by the ECU 90. The ECU 90 switches the connection state between the line pressure system 82 and the pinching pressure generating hydraulic chamber 45 by controlling the disk pressing control valve 88, thereby supplying hydraulic oil to the pinching pressure generating hydraulic chamber 45. When supplying hydraulic oil to the clamping pressure generating hydraulic chamber 45, the ECU 90 opens the disk pressing control valve 88 based on a valve control amount that provides a predetermined clamping pressure. The disk pressing control valve 88 is, for example, a pressure control valve, and can control the pressure of hydraulic oil supplied to the clamping pressure generating hydraulic chamber 45. The disk pressing control valve 88 can be, for example, a spool valve. The disk pressing control valve 88 has a spool valve element urged by an urging member, an input port communicating with the oil pump 81, an output port communicating with the clamping pressure generating hydraulic chamber 45, and a drain port. The disk pressing control valve 88 controls the hydraulic pressure supplied to the clamping pressure generating hydraulic chamber 45 when the spool valve element driven by hydraulic pressure or electromagnetic force moves to the pressure increasing side or the pressure reducing side.

  For example, when the spool valve element is driven to the pressure increasing side, the cross-sectional area of the flow path connecting the input port and the output port increases, or the flow path disconnection of the flow path connecting the output port and the drain port is increased. The area is reduced. As a result, the hydraulic pressure in the clamping pressure generating hydraulic chamber 45 increases. On the other hand, when the spool valve element is driven to the pressure reducing side, the cross-sectional area of the flow path connecting the input port and the output port decreases, or the flow path disconnection of the flow path connecting the output port and the drain port is reduced. Increases area. As a result, the hydraulic pressure in the clamping pressure generating hydraulic chamber 45 decreases. Further, the disk pressing control valve 88 can maintain the hydraulic pressure of the clamping pressure generating hydraulic chamber 45 with none of the input port, the output port, and the drain port communicating with the other ports.

  The ECU 90 can control the hydraulic pressure of the clamping pressure generating hydraulic chamber 45 by adjusting a signal pressure that is a hydraulic pressure for driving the spool valve element of the disk pressing control valve 88, for example. Alternatively, the ECU 90 can control the hydraulic pressure of the clamping pressure generating hydraulic chamber 45 by adjusting an electromagnetic force such as a solenoid that drives the spool valve element of the disk pressing control valve 88.

  The servo hydraulic system 84 is a drive source (hydraulic power source) of the hydraulic servo mechanism 25, and connects the line pressure system 82 to the first hydraulic chamber OP1 and the second hydraulic chamber OP2. The servo hydraulic system 84 is provided with a first shift control valve 92 that adjusts the flow rate of hydraulic oil supplied from the line pressure system 82 to the first hydraulic chamber OP1, and is supplied from the line pressure system 82 to the second hydraulic chamber OP2. A second shift control valve 93 is provided for adjusting the flow rate of the operating oil.

  The first shift control valve 92 is controlled by the ECU 90, and by controlling the first shift control valve 92 by the ECU 90, the connection state between the line pressure system 82 and the first hydraulic chamber OP1 is switched. The hydraulic fluid is supplied from the pressure system 82 to the first hydraulic chamber OP1 and is discharged from the first hydraulic chamber OP1 to the oil pan 94. The second shift control valve 93 is also configured in the same manner as the first shift control valve 92 and is controlled by the ECU 90. By controlling the second shift control valve 93 by the ECU 90, the line pressure system 82 and the second hydraulic chamber are controlled. The connection state with OP2 is switched, whereby hydraulic oil is supplied from the line pressure system 82 to the second hydraulic chamber OP2, and hydraulic oil is discharged from the second hydraulic chamber OP2 to the oil pan 94.

  Therefore, when the flange portion 71 is moved in the A1 direction along the axis X4, the ECU 90 controls the first shift control valve 92 to supply hydraulic oil from the line pressure system 82 to the first hydraulic chamber OP1, while the ECU 90 Thus, the second shift control valve 93 is controlled to discharge the hydraulic oil from the second hydraulic chamber OP2 to the oil pan 94. On the other hand, when the flange portion 71 is moved in the A2 direction along the axis X4, the ECU 90 controls the first shift control valve 92 to discharge the hydraulic oil from the first hydraulic chamber OP1 to the oil pan 94, while the ECU 90 The second transmission control valve 93 is controlled to supply hydraulic oil from the line pressure system 82 to the second hydraulic chamber OP2.

  The ECU 90 determines a target speed ratio of the toroidal continuously variable transmission 8 based on, for example, the vehicle speed and the accelerator opening, and determines a target tilt angle of each power roller 22 based on the target speed ratio. The ECU 90 controls the movement of each trunnion 23 in the direction of the axis X4 by the hydraulic servo mechanism 25 so as to realize the target tilt angle. Further, the ECU 90 determines a target value of the clamping pressure for clamping each power roller 22 based on the output torque of the engine 5, the gear ratio of the toroidal continuously variable transmission 8, and the like. The ECU 90 controls the hydraulic pressure of the clamping pressure generating hydraulic chamber 45 by the disk pressing control valve 88 so as to realize the determined target clamping pressure.

  Here, slip may occur between the power rollers 22 and the toroidal surfaces 33a of the input disks 20a and 20b, or between the power rollers 22 and the toroidal surfaces 33b of the output disks 21a and 21b. It is desirable that the occurrence of slip between the power roller 22 and the toroidal surfaces 33a and 33b can be appropriately suppressed.

  The toroidal continuously variable transmission 8 according to the present embodiment includes a clamping pressure correction system 100. FIG. 4 is a schematic configuration diagram of the clamping pressure correction system 100. The clamping pressure correction system 100 has a function of correcting the clamping pressure for clamping each power roller 22 so as to suppress the slip between each power roller 22 and the toroidal surfaces 33a and 33b. The clamping pressure correction system 100 includes a disk pressing control valve 88, a first pressure output mechanism 101, a second pressure output mechanism 102, and a signal pressure output mechanism 103.

  The first pressure output mechanism 101 is driven in conjunction with the tilt of the first power roller, which is one of the plurality of power rollers 22a, 22b, 22c, and 22d, and the line pressure is adjusted by the driving to adjust the first pressure. This is a first valve mechanism that outputs P1. That is, the first pressure output mechanism 101 can output the first pressure P1 corresponding to the tilt angle of the first power roller. The second pressure output mechanism 102 is driven in conjunction with the tilt of the second power roller, which is another one of the plurality of power rollers 22a, 22b, 22c, and 22d, and adjusts the line pressure by the driving. The second valve mechanism that outputs the second pressure P2. The second pressure output mechanism 102 can output the second pressure P2 corresponding to the tilt angle of the second power roller.

  Here, the first power roller is also described as one of the pair of power rollers 22a and 22b disposed between the rear side input disk 20b and the rear side output disk 21b (hereinafter also referred to as “first power roller 22a”). ) And the second power roller is the other 22b of the pair of power rollers 22a and 22b (hereinafter also referred to as “second power roller 22b”).

  The first pressure output mechanism 101 and the second pressure output mechanism 102 can be, for example, spool valves. The first pressure output mechanism 101 includes a valve body 101a, a spool valve element 101b, a spring 101c, and a piston 101d. The valve body 101a extends in a direction along the axis X1. The valve body 101a includes an input port 101e, an output port 101f, a feedback port 101g, and a drain port 101h. The feedback port 101g is formed at one end of the valve body 101a, and an input port 101e, an output port 101f, and a drain port 101h are arranged in this order toward the other end of the valve body 101a.

  A source pressure, for example, a line pressure is supplied to the input port 101e. The output port 101 f is connected to the first pressure port 103 h of the signal pressure output mechanism 103. The output port 101f is further connected to the feedback port 101g via the aperture 106. The drain port 101h is connected to the oil pan 94, for example.

  The spool valve element 101b is disposed in the valve body 101a, and is hydraulically output via the output port 101f (first hydraulic pressure P1) by moving in the axial direction relative to the input port 101e and the output port 101f. Adjust the pressure. The spool valve element 101b has a first valve portion 101i and a second valve portion 101j. The first valve portion 101i opens and closes the input port 101e to control the flow area of the lubricating oil between the input port 101e and the output port 101f. The second valve portion 101j opens and closes the drain port 101h and controls the flow area of the lubricating oil between the output port 101f and the drain port 101h. The feedback port 101g is located on the opposite side of the output port 101f with the first valve portion 101i interposed therebetween. Therefore, an axial force corresponding to the differential pressure between the output port 101f and the feedback port 101g acts on the first valve portion 101i.

  The end of the spool valve element 101b opposite to the feedback port 101g side is connected to the piston 101d via a spring 101c. The second valve portion 101j receives the spring force Fsp by the spring 101c and the pressure of the output port 101f.

  The piston 101d moves in the axial direction of the valve body 101a according to the tilt angle of the first power roller 22a. The piston 101d is connected to the first power roller 22a via a pin 104 and a rod 105. The rod 105 rotates around the axis X4 in conjunction with the tilt of the first power roller 22a. The rod 105 is connected to, for example, a trunnion 23 that supports the first power roller 22a, and protrudes from the first power roller 22a toward the outside in the radial direction with respect to the axis X1. A pin 104 is connected to the radially outer end of the rod 105. The pin 104 is supported by the piston 101d. In other words, the rod 105 and the piston 101d are connected by the pin 104 and are movable in the axial direction of the valve body 101a, and are relatively rotatable around the pin 104 as a rotation center.

  FIG. 5 is an operation explanatory diagram of the first pressure output mechanism 101. When the tilt angle of the first power roller 22a changes, the rod 105 rotates together with the first power roller 22a about the axis X4 as shown in FIG. As a result, the pin 104 moves in the axial direction of the valve body 101a and drives the piston 101d in the axial direction. For example, when the piston 101d is driven in the B1 direction (the direction from the drain port 101h toward the feedback port 101g in the axial direction of the valve main body 101a) in conjunction with the change in the tilt angle, the piston 101d compresses the spring 101c. The spring force Fsp of the spring 101c increases. As a result, the spool valve element 101b moves in the B1 direction to open the input port 101e, and the first pressure P1 output to the output port 101f increases.

  Since the output port 101f and the feedback port 101g communicate with each other, when the first pressure P1 acting on the spool valve element 101b via the feedback port 101g increases to a pressure that balances the pressing force by the spring force Fsp, the spool valve element 101b returns to the neutral position and closes the input port 101e. The restrictor 106 suppresses the vibration of the supply pressure when the first pressure P1 is adjusted according to the spring force Fsp.

  On the other hand, when the piston 101d is driven in the B2 direction in conjunction with the change in the tilt angle, the piston 101d moves in the direction of extending the spring 101c, and the spring force Fsp decreases. As a result, the spool valve element 101b moves in the B2 direction to open the drain port 101h, and the supply pressure supplied from the output port 101f decreases. When the supply pressure decreases until it balances with the pressing force by the spring force Fsp, the spool valve element 101b returns to the neutral position and closes the drain port 101h.

Thus, the first pressure output mechanism 101 can output the first pressure P1 corresponding to the tilt angle of the first power roller 22a. The relationship between the first pressure P1 and the spring force Fsp is expressed by the following formula (1), where A is the valve area of the spool valve element 101b (for example, the cross-sectional area of the valve portion 101i).
P1 × A = Fsp (1)

  The second pressure output mechanism 102 is provided in the second power roller 22b, and outputs the second pressure P2 corresponding to the tilt angle of the second power roller 22b by the same method as the first pressure output mechanism 101. be able to. The second pressure output mechanism 102 is supplied with the line pressure, and outputs the second pressure P2 adjusted according to the tilt angle of the second power roller 22b to the output port. The first pressure output mechanism 101 and the second pressure output mechanism 102 can output the same pressure with respect to the same tilt angle. That is, when the tilt angle of the first power roller 22a and the tilt angle of the second power roller 22b are equal, the first pressure P1 and the second pressure P2 are the same pressure.

  The signal pressure output mechanism 103 is a correction valve that outputs the signal pressure Ps so that the slip of the power roller 22 is suppressed when the differential pressure between the first pressure P1 and the second pressure P2 exceeds a predetermined threshold. Mechanism. The signal pressure output mechanism 103 can output the signal pressure Ps that suppresses the slip of the power roller 22 in accordance with the differential pressure between the first pressure P1 and the second pressure P2.

  The signal pressure output mechanism 103 includes a valve body 103a, a spool valve element 103b, a first spring 103c, and a second spring 103d. The valve body 103a includes a first input port 103e, a second input port 103f, an output port 103g, a first pressure port 103h, and a second pressure port 103i. The first pressure port 103h is formed at one end of the valve body 103a, and the second pressure port 103i is formed at the other end of the valve body 103a. The first input port 103e and the second input port 103f are formed between the first pressure port 103h and the second pressure port 103i, and the first input port 103e has a first pressure higher than that of the second input port 103f. It is arranged on the port 103h side. The output port 103g is formed between the first input port 103e and the second input port 103f.

  The first pressure P1 output by the first pressure output mechanism 101 is supplied to the first pressure port 103h. The second pressure P2 output by the second pressure output mechanism 102 is supplied to the second pressure port 103i. The first input port 103e and the second input port 103f are supplied with a source pressure from a common pressure source, for example, a line pressure.

  The spool valve element 103b is disposed in the valve body 103a, and opens or closes the first input port 103e or the second input port 103f by moving in the axial direction of the valve body 103a. The spool valve element 103b has a first valve portion 103j and a second valve portion 103k. The first valve portion 103j opens and closes the first input port 103e, and the second valve portion 103k opens and closes the second input port 103f.

  The first spring 103c is disposed on the first pressure port 103h side with respect to the spool valve element 103b, and biases the spool valve element 103b toward the second pressure port 103i. On the other hand, the second spring 103d is disposed on the second pressure port 103i side with respect to the spool valve element 103b, and urges the spool valve element 103b toward the first pressure port 103h. The spool valve element 103b closes the first input port 103e and the second input port 103f in the neutral position. For example, when the first pressure P1 is equal to the second pressure P2, the force toward the first pressure port 103h acting on the spool valve element 103b and the force toward the second pressure port 103i are balanced in the neutral position. To do. As a result, the spool valve element 103b closes the input ports 103e and 103f, and no signal pressure is output.

  When the tilt angle of the first power roller 22a and the tilt angle of the second power roller 22b are equal, such as when neither the first power roller 22a nor the second power roller 22b is slipping, the first pressure P1 It is equal to the second pressure P2. In this case, the signal pressure output mechanism 103 does not output a signal pressure. If slip occurs in the first power roller 22a or the second power roller 22b and a deviation occurs between the target tilt angle and the actual tilt angle, a differential pressure between the first pressure P1 and the second pressure P2 is generated. . For example, when slip occurs in the first power roller 22a, the differential pressure between the first pressure P1 and the second pressure P2 increases as the displacement of the tilt angle of the first power roller 22a increases.

  The signal pressure output mechanism 103 outputs a signal pressure corresponding to a differential pressure between the first pressure P1 and the second pressure P2 (hereinafter simply referred to as “differential pressure ΔP”) to the output port 103g. For example, when the first pressure P1 becomes higher than the second pressure P2, the spool valve element 103b moves toward the C1 direction (from the first pressure port 103h to the second pressure port 103i in the axial direction of the valve body 103a) by the differential pressure ΔP. Direction). Thereby, the second input port 103f is opened. As the magnitude of the differential pressure ΔP increases, the opening of the second input port 103f increases and the signal pressure Ps increases. Thus, when the first pressure P1 increases due to slip, the signal pressure output mechanism 103 outputs a signal pressure Ps having a magnitude corresponding to the degree of increase in the first pressure P1.

  On the other hand, when the first pressure P1 becomes lower than the second pressure P2, the spool valve element 103b moves in the direction C2 (from the second pressure port 103i to the first pressure port 103h in the axial direction of the valve body 103a) by the differential pressure ΔP. Direction). As a result, the first input port 103e is opened. As the magnitude of the differential pressure ΔP increases, the opening of the first input port 103e increases and the signal pressure Ps increases. As described above, when the first pressure P1 decreases due to the slip, the signal pressure output mechanism 103 outputs the signal pressure Ps having a magnitude corresponding to the degree of decrease of the first pressure P1.

  The signal pressure output mechanism 103 can be configured to output a signal pressure when the differential pressure ΔP between the first pressure P1 and the second pressure P2 is equal to or greater than a predetermined threshold. This threshold value can be adjusted by the valve diameter (valve area A) and the spring coefficient of the springs 103c and 103d. For example, the signal pressure output mechanism 103 may be set not to output the signal pressure Ps while allowing the differential pressure ΔP between the first pressure P1 and the second pressure P2 to the extent that is caused by variations in individual components. Further, the signal pressure output mechanism 103 may be set so as to allow the differential pressure ΔP due to a very small degree of slip and not output the signal pressure Ps.

  The same applies to the case where the second pressure P2 fluctuates due to slip, and the signal pressure output mechanism 103 outputs a signal pressure Ps corresponding to the degree to which the second pressure P2 increases and the degree to which the second pressure P2 decreases.

  The signal pressure Ps output by the signal pressure output mechanism 103 is input to the disk pressing control valve 88. The signal pressure Ps increases the clamping pressure generated by the disk pressing control valve 88. The valve control amount for the disc pressing control valve 88 by the ECU 90 and the valve control amount for the disc pressing control valve 88 by the signal pressure Ps act in a superimposed manner to increase the opening of the disc pressing control valve 88. For example, when the disk pressing control valve 88 is a spool valve, the driving force for driving the spool valve element of the disk pressing control valve 88 by hydraulic pressure or the like according to a command from the ECU 90 and the signal pressure Ps are the spool valve elements of the disk pressing control valve 88. The spool valve element is moved in a direction in which the pressure in the clamping pressure generating hydraulic chamber 45 is increased in a superimposed manner.

  In other words, the disk pressing control valve 88 of the present embodiment corrects the clamping pressure with which the input disks 20a, 20b and the output disks 21a, 21b sandwich the power roller 22, and increases the clamping pressure according to the signal pressure Ps. It functions as a pressure correction mechanism.

  The disc pressing control valve 88 increases the hydraulic pressure in the clamping pressure generating hydraulic chamber 45 according to the signal pressure Ps, whereby the clamping pressure for clamping each power roller 22 increases. By increasing the pinching pressure, it is possible to suppress slipping of the power roller 22 due to a tilt angle synchronization shift or the like. In the toroidal-type continuously variable transmission 8 of the present embodiment, slipping of the power roller 22 can be mechanically suppressed by the clamping pressure correction system 100 without providing a sensor or the like for detecting a tilt angle. Therefore, both suppression of slip of the power roller 22 and suppression of cost increase can be achieved.

  Further, the signal pressure Ps increases as the differential pressure ΔP increases, and the correction amount of the clamping pressure with respect to the power roller 22, that is, the increment of the clamping pressure increases. Therefore, when the slip ratio is large, the pinching pressure increases as the angle difference between the tilt angle of the first power roller 22a and the tilt angle of the second power roller 22b increases, and the slip can be appropriately suppressed. .

  According to the present embodiment, it is possible to perform fail-safe operation by performing fail determination (fail detection) with high accuracy when slip occurs. For example, in the case of a conventional method for determining a failure based on a gear ratio estimated from the position of an actuator, there is a possibility that the failure cannot be correctly determined when a slip occurs. When slip occurs, the plurality of power rollers may not slip in the same way. As an example, out of four power rollers, three power rollers do not slip, and if only one power roller slips, the change in the gear ratio with the case of not slipping is small, and fail determination cannot be made correctly. there is a possibility.

  On the other hand, according to the toroidal-type continuously variable transmission 8 of the present embodiment, even when only a part of the power rollers 22 slips, such as when only one power roller 22 slips, the hydraulic pressure A slip can be accurately detected by the circuit and a fail-safe operation can be performed. In the toroidal continuously variable transmission 8 of the present embodiment, slip is automatically suppressed by the hydraulic circuit without the ECU 90 being interposed. Therefore, the load on the ECU 90 can be reduced.

  In addition, the pressure output mechanism which outputs the pressure according to the tilt angle of the power roller 22 is not limited to what was illustrated in this embodiment.

  In the present embodiment, a case where one of the pair of power rollers 22 disposed between the rear side input disk 20b and the rear side output disk 21b is a first power roller and the other 22d is a second power roller will be described. However, it is not limited to this. The first power roller and the second power roller are optional. For example, one of the pair of power rollers 22 disposed between the front side input disk 20a and the front side output disk 21a may be a first power roller, and the other 22b may be a second power roller.

  Alternatively, one of the pair of power rollers 22a and 22b disposed on the front side may be a first power roller, and one of the pair of power rollers 22c and 22d disposed on the rear side may be a second power roller. Further, in addition to the signal pressure output mechanism 103 that outputs the signal pressure Ps corresponding to the difference in tilt angle between the pair of power rollers 22 on the front side or the rear side, the tilt angle of the front side power roller 22 and the rear side. A signal pressure output mechanism 103 that outputs a signal pressure Ps corresponding to a difference from the tilt angle of the power roller 22 may be provided.

  Note that the output of the engine 5 may be reduced based on the signal pressure Ps. For example, a sensor that detects the signal pressure Ps may be provided, and the ECU 90 may perform control to reduce the output of the engine 5 based on the detection result of the sensor. For example, in addition to the correction of the clamping pressure by the clamping pressure correction system 100, by reducing the output of the engine 5, the slip of the power roller 22 can be suppressed at an earlier timing or the slip rate of the power roller 22 can be further reduced. It becomes possible to make it.

  Alternatively, a mechanism for mechanically reducing the output of the engine 5 by the signal pressure Ps may be provided. For example, a mechanism for reducing the opening degree of the throttle valve by the signal pressure Ps or a mechanism for reducing the fuel injection amount may be provided. In addition to these, a slip suppression mechanism that suppresses the slip of the power roller 22 can be operated according to the signal pressure Ps. Note that the slip suppression mechanism including the mechanism for reducing the output of the engine 5 operates based on the differential pressure between the first pressure P1 and the second pressure P2 instead of being driven by the signal pressure Ps. May be.

  The hydraulic circuit that detects a failure based on the mismatch between the first pressure P1 and the second pressure P2 and the hydraulic circuit that performs a fail operation based on the detected failure are not limited to those exemplified in this embodiment.

  In the present embodiment, the disc pressing control valve 88 has a function as a clamping pressure correction mechanism that increases the clamping pressure in accordance with the signal pressure Ps. However, the present invention is not limited to this. For example, a clamping pressure correction mechanism that increases the clamping pressure according to the signal pressure Ps may be provided independently of the disk pressing control valve 88.

[First Modification]
A first modification will be described. Instead of correcting the clamping pressure by the signal pressure of the signal pressure output mechanism 103, the clamping pressure may be corrected by detecting the first pressure P1 or the like output from the first pressure output mechanism 101. FIG. 6 is a schematic configuration diagram of the clamping pressure correction system 110 according to the first modification, and FIG. 7 is a flowchart of control according to the first modification.

  As shown in FIG. 6, the first pressure output mechanism 101 is provided on one side 22c of the pair of power rollers 22 arranged on the rear side. Further, a hydraulic pressure sensor 107 that detects the first pressure P1 output by the first pressure output mechanism 101 is provided. Similarly, a second pressure output mechanism 102 is provided on the other 22d (see FIG. 1) of the pair of power rollers 22 arranged on the rear side. Further, a hydraulic pressure sensor 108 that detects the second pressure P2 output by the second pressure output mechanism 102 is provided. The hydraulic sensors 107 and 108 are connected to the ECU 90, and signals indicating the detection results of the hydraulic sensors 107 and 108 are output to the ECU 90.

  Furthermore, a pair of power rollers 22a and 22b (see FIG. 1) arranged on the front side are provided with a pressure output mechanism and a hydraulic sensor (not shown), respectively. Each pressure output mechanism outputs a pressure corresponding to the tilt angle of the corresponding power roller 22, and the hydraulic sensor detects these pressures and outputs a detection result signal to the ECU 90. The ECU 90 performs a fail determination and a fail safe operation based on the pressure corresponding to the inclination angle of each power roller 22a, 22b, 22c, 22d.

  With reference to FIG. 7, the control according to the present modification will be described. The control flow shown in FIG. 7 is executed while the vehicle 1 is traveling, for example.

  First, in step S1, the ECU 90 determines whether the difference between the maximum and minimum of the plurality of angle values is greater than a predetermined value. In step S <b> 1, it is determined whether or not there is an angle difference exceeding a predetermined value in the tilt angle of each power roller 22. When the difference between the maximum and minimum of the plurality of angle values exceeds a predetermined value, it can be determined that the power roller 22 is slipping. ECU90 acquires the tilt angle of each power roller 22 based on the detection result of the hydraulic sensor arrange | positioned at each power roller 22, and performs determination of step S1 based on the acquired tilt angle. If it is determined in step S1 that the difference between the maximum and minimum of the plurality of angle values is greater than the predetermined value (step S1-Y), the process proceeds to step S2, and otherwise (step S1- In N), the control flow ends.

  In step S2, fail safe operation is executed by the ECU 90. The ECU 90 increases the hydraulic pressure generated by the disk pressing control valve 88 and increases the clamping pressure on the power roller 22. Further, the ECU 90 reduces the output of the engine 5 and reduces the torque input to the toroidal continuously variable transmission 8. By implementing at least one of the increase of the clamping pressure and the reduction of the engine torque, the slip of the power roller 22 is suppressed. When step S2 is executed, this control flow ends.

  In the present modification, the fail-safe operation can be performed based on a comparison result between the maximum value and the minimum value of a plurality of angle values (pressure corresponding to the angle value). Further, it is possible to monitor each angle value and the fluctuation of the angle value at an arbitrary timing. For example, it is possible to perform the slip determination when any of the power rollers 22 has a pressure variation different from that of the other power rollers 22. As an example, it is possible to determine that a slip has occurred when the pressure fluctuation related to one of the power rollers 22 is larger than the pressure fluctuation related to the other power roller 22 in the same period.

  Instead of or in addition to the fail determination based on the difference between the maximum and minimum angle values, the fail determination is performed based on the mismatch between the transmission ratio calculated from the input / output rotation speed and the transmission ratio based on the hydraulic pressure. May be. For example, when the transmission ratio calculated based on the input rotation speed and the output rotation speed of the toroidal continuously variable transmission 8 does not match the tilt angle (transmission ratio) based on the first pressure P1, the power roller 22c slips. It is possible to determine that

[Second Modification]
A second modification will be described. Instead of a pressure output mechanism that outputs a pressure corresponding to the tilt angle of the power roller 22, an angle sensor that detects the tilt angle of the power roller 22 may be provided. FIG. 8 is a schematic configuration diagram of the clamping pressure correction system 120 according to the second modification.

  As shown in FIG. 8, the tilt angle sensor 109 is disposed on the swing shaft 61 b of the trunnion 23. The tilt angle sensor 109 detects the tilt angle of the power roller 22 based on the rotation angle of the swing shaft 61b around the axis line X4. The tilt angle sensor 109 is arranged for each of the power rollers 22a, 22b, 22c and 22d. The tilt angle sensor 109 is connected to the ECU 90, and a signal indicating the detection result of each tilt angle sensor 109 is output to the ECU 90. The tilt angle sensor 109 is not limited to the swing shaft 61b, and may be arranged so as to detect the tilt angle of the power roller 22 or an angle or other physical quantity that changes in conjunction with the tilt of the power roller 22. That's fine.

  The ECU 90 performs a fail-safe operation based on the detection result of each tilt angle sensor 109. For example, the ECU 90 may perform the fail determination and the fail safe operation according to the same flowchart as that of the first modified example (FIG. 7). In this case, in step S1, the angle value detected by the tilt angle sensor 109 may be referred to instead of the angle value based on the hydraulic sensor of the first modification. The ECU 90 performs the fail-safe operation when the difference between the maximum value and the minimum value of the plurality of angle values detected by the tilt angle sensor 109 exceeds a predetermined value. The content of the fail-safe operation can be the same as that of the first modified example.

  Instead of or in addition to the fail determination based on the difference between the maximum and minimum angle values, based on the mismatch between the speed ratio calculated from the input / output rotation speed and the speed ratio based on the detection result of the tilt angle sensor 109. Fail determination may be performed.

  The contents disclosed in the above embodiments can be executed in appropriate combination.

DESCRIPTION OF SYMBOLS 1 Vehicle 8 Toroidal type continuously variable transmission 20a Front side input disk 20b Rear side input disk 21a Front side output disk 21b Rear side output disk 22a, 22b, 22c, 22d Power roller 23 Trunnion 88 Disk pressure control valve (clamping pressure correction mechanism) )
90 ECU
DESCRIPTION OF SYMBOLS 100 Nipping pressure correction system 101 1st pressure output mechanism 102 2nd pressure output mechanism 103 Signal pressure output mechanism P1 1st pressure P2 2nd pressure Ps Signal pressure

Claims (4)

A plurality of power rollers for transmitting power between the input disk and the output disk and including a first power roller and a second power roller;
A first pressure output mechanism that outputs a first pressure according to the tilt angle of the first power roller;
A second pressure output mechanism that outputs a second pressure according to the tilt angle of the second power roller;
A signal pressure output mechanism that outputs a signal pressure that suppresses slippage of the power roller in accordance with a differential pressure between the first pressure and the second pressure;
A toroidal-type continuously variable transmission comprising: Furthermore, a clamping pressure correction mechanism that corrects a clamping pressure with which the input disk and the output disk sandwich the plurality of power rollers is provided,
The toroidal continuously variable transmission according to claim 1, wherein the clamping pressure correction mechanism increases the clamping pressure according to the signal pressure. The toroidal continuously variable transmission according to claim 1, wherein the signal pressure output mechanism outputs the signal pressure when a differential pressure between the first pressure and the second pressure is equal to or greater than a threshold value. The toroidal type according to any one of claims 1 to 3, further comprising: reducing an output of a prime mover that outputs power to the input disk based on a differential pressure between the first pressure and the second pressure. Continuously variable transmission.

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