JP6170012B2 - Toroidal continuously variable transmission - Google Patents

Description

  The present invention relates to a toroidal continuously variable transmission that suppresses slippage of a power roller held between an input disk and an output disk by a pressing force generated by a double piston hydraulic loader.

  A hydraulic loader of a toroidal-type continuously variable transmission that changes a gear ratio steplessly by tilting a power roller sandwiched between an input disk and an output disk includes a first oil chamber, a first piston, and a second oil chamber. Patent Document 1 below discloses that an input disk is pressed in the axial direction by supplying the same hydraulic pressure to the first oil chamber and the second oil chamber to suppress slippage of the power roller. It is.

Japanese Patent No. 4696537

  By the way, when the input disk is pressed in the axial direction by the hydraulic loader, if the input disk is deformed, various problems such as reduction in the control accuracy of the transmission ratio occur. It is desirable to keep it to a minimum. In particular, when the hydraulic loader includes a piston that locally presses a specific position on the back surface of the input disk and an oil chamber that uniformly presses a wide area on the back surface of the input disk, the two pressing means are appropriately used. Otherwise, the amount of deformation of the input disk may be increased.

  Therefore, the first piston is easily deformed because it locally presses the radially outer end of the input disk, but the oil pressure of the second oil chamber presses the entire area of the back surface of the input disk evenly and is difficult to deform. Attention is paid to switching between a first control state in which oil pressure is supplied only to the second oil chamber and a second control state in which oil pressure is supplied to both the first oil chamber and the second oil chamber. When the input disk is pressed preferentially and the oil pressure in the second oil chamber alone is insufficient, the input disk is pressed by the first piston to minimize the deformation of the input disk while suppressing the slippage of the power roller. It is conceivable to limit it to the limit.

  However, when such a double piston type hydraulic loader is employed, the hydraulic pressure in the first oil chamber is released while the hydraulic pressure is supplied only to the second oil chamber in the first control state, and the vehicle is suddenly accelerated. For example, when the hydraulic pressure is supplied to the first oil chamber by shifting from the first control state to the second control state, the hydraulic pressure in the first oil chamber is delayed and the pressing force is insufficient, and the power roller can slip. There is sex.

  The present invention has been made in view of the above-described circumstances, and an object thereof is to quickly raise the pressing force of a double piston hydraulic loader of a toroidal continuously variable transmission.

  In order to achieve the above object, according to the first aspect of the present invention, a rotary shaft connected to a drive source, an input disk supported so as not to rotate relative to the rotary shaft, and relative to the rotary shaft are provided. An output disk supported rotatably, a power roller supported tiltably around a trunnion shaft and sandwiched between the input disk and the output disk, and the input disk attached in a direction approaching the output disk A hydraulic loader that urges the input disk, and the hydraulic loader is fitted to an inner peripheral surface of the first cylinder housing so as to be slidable in the axial direction. A first piston abutting on an axial end of a second cylinder housing projecting on the back surface, and an axial sliding self-fixing on an inner peripheral surface of the second cylinder housing fixed to the rotating shaft. A second piston fitted between the side wall of the first cylinder housing and the first oil chamber defined between the first piston, a second oil defined between the rear surface of the input disk and the second piston. And a hydraulic control circuit that controls the hydraulic pressure supplied to the first oil chamber and the second oil chamber, wherein the hydraulic control circuit is connected to the rotary shaft from the drive source. And a second control for supplying the control oil pressure to both the first oil chamber and the second oil chamber based on the input torque input to the second oil chamber. A toroidal continuously variable transmission is proposed in which a lubricating oil pressure lower than a control oil pressure is supplied to the first oil chamber in the first control state.

  According to the second aspect of the present invention, in addition to the configuration of the first aspect, when the input torque increases by a predetermined value or more in the first control state, the second control state is changed from the first control state. A toroidal-type continuously variable transmission characterized by shifting to is proposed.

  According to the third aspect of the present invention, in addition to the configuration of the first or second aspect, the first oil passage for supplying the control oil pressure to the first oil chamber, and the first oil chamber are lubricated. A second oil passage for supplying hydraulic pressure; and a switching valve for selectively connecting the first oil passage and the second oil passage to the first oil chamber, wherein the switching valve is the second oil in an initial state. A toroidal continuously variable transmission is proposed in which a road is connected to the first oil chamber.

  The input shaft 18 of the embodiment corresponds to the rotation shaft of the present invention, and the engine E of the embodiment corresponds to the drive source of the present invention.

  According to the first aspect of the present invention, the toroidal continuously variable transmission is supported by the rotating shaft connected to the drive source, the input disk supported so as not to rotate relative to the rotating shaft, and supported by the rotating shaft so as to be relatively rotatable. An output disk, a power roller supported so as to be tiltable around the trunnion shaft and sandwiched between the input disk and the output disk, and a hydraulic loader for urging the input disk in a direction approaching the output disk. The hydraulic loader includes a first cylinder housing fixed to a rotating shaft, and an axial direction of a second cylinder housing that is fitted to the inner peripheral surface of the first cylinder housing so as to be slidable in the axial direction and protrudes from the back surface of the input disk. A first piston abutting on the end, a second piston fixed to the rotary shaft and fitted to the inner peripheral surface of the second cylinder housing so as to be axially slidable, and between the side wall of the first cylinder housing and the first piston A first oil chamber partitioned into a first oil chamber, a second oil chamber partitioned between the back surface of the input disk and the second piston, and a hydraulic control circuit for controlling the hydraulic pressure supplied to the first oil chamber and the second oil chamber. Since it is provided, the first piston that contacts the second cylinder housing is driven by the hydraulic pressure supplied to the first oil chamber to press the input disk in the axial direction, and is supplied to the second oil chamber defined by the second piston. Oil The back of the input disk by pressing in the axial direction, can be nipped power rollers between an input disk and output disk to prevent the occurrence of slip in.

  Since the first piston locally presses the radially outer end of the input disk (that is, the second cylinder housing), it is easy to deform the input disk, but the oil pressure in the second oil chamber is uniform over the entire back surface of the input disk. To reduce the deformation of the input disk. Therefore, the hydraulic control circuit switches between the first control state in which the hydraulic pressure is supplied only to the second oil chamber and the second control state in which the hydraulic pressure is supplied to both the first oil chamber and the second oil chamber. The input disc is preferentially pressed by the hydraulic pressure of the chamber, and when the hydraulic pressure of the second oil chamber alone is insufficient, the input disc is auxiliaryly pressed by the first piston, thereby suppressing the slip of the power roller and the input disc. To minimize the deformation of the shaft, improve the gear ratio control accuracy, improve the power transmission efficiency, prevent fretting of the toroidal curved surfaces of the input disk and output disk, and prevent the first piston and the second piston from contacting each other. Can do.

  In addition, since the hydraulic pressure control circuit supplies a lubricating hydraulic pressure lower than the control hydraulic pressure to the first oil chamber in the first control state, air is prevented from being mixed into the first oil chamber in the first control state. When the control oil pressure is supplied to the first oil chamber from the control state to the second control state, the control oil pressure can be quickly raised to reliably prevent the power roller from slipping.

  According to the second aspect of the present invention, when the input torque increases to a predetermined value or more in the first control state, the second control state is shifted from the first control state. By supplying the control oil pressure only to the chamber and supplying the control oil pressure to both the first oil chamber and the second oil chamber when the input torque is large, it is possible to prevent the power roller from slipping with the minimum required oil pressure. Thus, the load of the hydraulic supply source can be reduced.

  According to the configuration of claim 3, the first oil passage for supplying the control oil pressure to the first oil chamber, the second oil passage for supplying the lubricating oil pressure to the first oil chamber, the first oil passage and the second oil A switching valve that selectively connects the passage to the first oil chamber, and the switching valve connects the second oil passage to the first oil chamber in an initial state, so that the switching valve is provided in the event that the power supply fails. In the initial state, the lubricating oil pressure is supplied from the second oil passage to the first oil chamber, and it is possible to automatically shift to the first control state suitable for failure.

The longitudinal section which passes along the engine of a toroidal type continuously variable transmission. The longitudinal cross-sectional view which passes along the electric motor of a toroidal type continuously variable transmission. 3 is an enlarged view of part 3 of FIG. The figure which shows the hydraulic control circuit of a hydraulic loader. The flowchart explaining control of a hydraulic loader. The time chart explaining the effect | action of a hydraulic loader. The graph explaining the responsiveness of the torque which can be transmitted with respect to loading oil_pressure | hydraulic. The graph which shows the relationship between the loading hydraulic pressure in the state of a double piston, and a transmittable torque.

  Hereinafter, embodiments of the present invention will be described with reference to FIGS.

  As shown in FIGS. 1 and 2, a transmission case 11 of a toroidal-type continuously variable transmission T for a hybrid vehicle includes a first case 12, a second case 13 coupled to the first case 12, and a second case. And a third case 14 coupled to 13. An input shaft 18 is supported so as to straddle the first case 12, the second case 13, and the third case 14, and the left end portion of the input shaft 18 extending into the first case 12 has a damper 19 and a generator 20. To the crankshaft 21 of the engine E.

  A toroidal transmission mechanism 22 having a well-known structure is provided at the right end portion of the input shaft 18. The single cavity type toroidal speed change mechanism 22 includes a substantially cone-shaped input disk 23 supported on the input shaft 18 so as not to rotate relative to the input shaft and slidable in the axial direction, and rotatable relative to the input shaft 18 and slidable in the axial direction. A pair of crank-shaped pivots having one end rotatably supported by a substantially cone-shaped output disk 24 supported and opposed to the input disk 23, and a pair of trunnions (not shown) arranged so as to sandwich the input shaft 18 therebetween. The shafts 25 and 25, a pair of power rollers 26 and 26 which are rotatably supported by the other ends of the pivot shafts 25 and 25 and abut against the input disk 23 and the output disk 24, and the input disk 23 is hydraulically connected to the output disk 24 side. And a hydraulic loader 27 that suppresses the slippage of the power rollers 26 and 26.

  When the pair of trunnions are driven in the opposite directions along the trunnion shafts 28 and 28 by hydraulic pressure, the power rollers 26 and 26 tilt in one direction around the trunnion shafts 28 and 28, and the contact point with the input disk 23 becomes the contact point. While moving radially outward with respect to the input shaft 18, the point of contact with the output disk 24 moves radially inward with respect to the input shaft 18, so that the rotation of the input disk 23 is accelerated to the output disk 24. The transmission ratio of the toroidal transmission mechanism 22 is continuously reduced.

  On the other hand, when the pair of trunnions are driven along the trunnion shafts 28, 28 in the opposite direction, the power rollers 26, 26 are tilted around the trunnion shafts 28, 28 in the other direction and contact points with the input disk 23. Moves radially inward with respect to the input shaft 18, and the contact point with the output disk 24 moves radially outward with respect to the input shaft 18, so that the rotation of the input disk 23 is decelerated to the output disk 24. The transmission ratio of the toroidal transmission mechanism 22 is continuously increased.

  A first gear 32 is rotatably supported on the outer periphery of the input shaft 18, and the first gear 32 can be coupled to the output disk 24 of the toroidal transmission mechanism 22 via a wet multi-plate hydraulic clutch 33. A drive gear 34 is fixed to the input shaft 18, and two thrust bearings 35 and 36 are disposed between the drive gear 34 and the first gear 32.

  An output shaft 37 parallel to the input shaft 18 is supported on the first case 12 and the second case 13, and a second gear 38 that meshes with the first gear 32 and a final drive gear 39 are supported on the output shaft 37. It is fixed. A differential gear D is disposed inside the first case 12 and the second case 13, and a final driven gear 40 fixed to the casing meshes with the final drive gear 39. A pair of axles 41, 41 extends from the differential gear D to the left and right through the first case 12 and the second case 13.

  Further, a first oil pump 42 is disposed inside the first case 12 and the second case 13, and a drive gear 34 fixed to the input shaft 18 is engaged with a driven gear 43 fixed to the pump shaft. The first oil pump 42 is driven in conjunction with the rotation of the input shaft 18. The first oil pump 42 pumps up oil stored in the bottom of the mission case 11 and supplies it to the hydraulic clutch 33 and the hydraulic loader 27 as hydraulic oil and also supplies it to the toroidal transmission mechanism 22 as lubricating oil.

  An electric motor M is supported on the right side surface of the second case 13, and a motor output gear 45 fixed to the motor shaft 44 meshes with a second gear 38 fixed to the output shaft 37. A second oil pump 46 is disposed inside the first case 12 and the second case 13, and a driven gear 47 fixed to the pump shaft meshes with the final driven gear 40. The second oil pump 46 driven by the differential gear D supplies oil instead of the first oil pump 42 when the vehicle travels with the engine E stopped.

  As shown in FIG. 3, the hydraulic loader 27 includes a first cylinder housing 51 fixed to the input shaft 18, an outer periphery and an inner periphery on the inner peripheral surface of the peripheral wall 51 a of the first cylinder housing 51 and the outer periphery of the input shaft 18. A first piston 52 slidably supported on the surface, a second cylinder housing 23a that protrudes axially from the input disk 23 and abuts against the first piston 52, and an outer peripheral surface that is an inner peripheral surface of the second cylinder housing 23a. A second piston 53 that is slidably supported by the inner shaft and is fixed to the input shaft 18, a first oil chamber 54 defined between the side wall 51b of the first cylinder housing 51 and the first piston 52, And a second oil chamber 55 defined between the rear surface of the input disk 23 and the second piston 53.

  The outer peripheral portion of the input disk 23 is spline-fitted 56 to the inner peripheral surface of the first cylinder housing 51 so as not to rotate relative to the axial direction and to be slidable in the axial direction. In a movable state, it rotates integrally with the input shaft 18. When the input disk 23 is pressed toward the power rollers 26, 26 by the hydraulic loader 27, the radial outer end portion tends to spread outward in the axial direction due to the reaction load that the input disk 23 receives from the power rollers 26, 26. Can be suppressed by the spline fitting 56 of the first cylinder housing 51 and the input disk 23.

  When the hydraulic pressure supplied to the first oil chamber 54 drives the first piston 52 to the left in the drawing with respect to the first cylinder housing 51, the first piston 52 presses the right end of the second cylinder housing 23a to input. The hydraulic pressure supplied to the second oil chamber 55 urges the disc 23 to the left and urges the input disc 23 to the left with respect to the second piston 53. As a result, the power rollers 26 and 26 are pinched between the input disk 23 and the output disk 24, and a loading thrust that suppresses the slip between the input disk 23 and the output disk 24 and the power rollers 26 and 26 can be generated. it can.

  At this time, the load by which the hydraulic loader 27 of the toroidal transmission mechanism 22 presses the input disk 23 to the left in FIG. 1 is the path of the power rollers 26, 26 → the output disk 24 → the hydraulic clutch 33 → the thrust bearing 35 in the second case. It is transmitted to and supported by a support member 13 a fixed to 13. Further, the load that the hydraulic loader 27 of the toroidal transmission mechanism 22 presses the input shaft 18 to the right in FIG. 1 is transmitted to and supported by the support member 13a through the path of drive gear 34 → thrust bearing 36.

  Further, since the hydraulic pressure in the second oil chamber 55 presses the entire back surface of the input disk 23, the deformation amount of the input disk 23 is relatively small, whereas the first piston 52 that operates with the hydraulic pressure in the first oil chamber 54 is used. Presses only the right end portion of the second cylinder housing 23a provided on the radially outer end side of the input disk 23, so that the deformation amount of the input disk 23 becomes relatively large. Therefore, in order to minimize the deformation of the input disk 23 and improve the speed ratio control accuracy and power transmission efficiency, it is desirable to apply hydraulic pressure to the second oil chamber 55 rather than the first oil chamber 54. .

  For this purpose, the hydraulic pressure supply to the first oil chamber 54 and the second oil chamber 55 is controlled by the hydraulic pressure control circuit 57 connected to the first and second oil pumps 42 and 46.

  As shown in FIG. 4, the hydraulic control circuit 57 includes a first oil passage P1 that extends from the first and second oil pumps 42 and 46 and connects to the second oil chamber 55. A hydraulic loader linear solenoid valve 58 is interposed. A switching valve 59 is connected to the first oil passage P1 on the downstream side of the hydraulic loader linear solenoid valve 58 and the second oil passage P2 to which lubricating oil pressure is supplied, and the downstream side of the switching valve 59 is the third oil passage. It is connected to the first oil chamber 54 via the path P3. The switching valve 59 formed of a solenoid valve connects the second oil passage P2 to the third oil passage P3 in the demagnetized state and supplies the lubricating oil pressure to the first oil chamber 54. In the excited state, the switching valve 59 connects the first oil passage P1 to the third oil passage 54. The loading hydraulic pressure is supplied to the first oil chamber 54 by connecting to the oil passage P3.

  Next, the operation of the embodiment of the present invention having the above configuration will be described.

  The vehicle according to the present embodiment includes an engine E and an electric motor M as a driving source for traveling. When the remaining capacity of the battery is sufficient, the vehicle travels with the driving force of the electric motor M without using the engine E. To do. That is, when the electric motor M is driven in a state where the hydraulic clutch 33 is disengaged, the driving force is the axle 41 in the path of the motor output gear 45 → second gear 38 → final drive gear 39 → final driven gear 40 → differential gear D. , 41, the vehicle travels forward or backward depending on the direction of rotation of the electric motor M.

  When the remaining capacity of the battery becomes a predetermined value or less, the traveling by the electric motor M is switched to the traveling by the engine E. That is, when the engine E is driven with the hydraulic clutch 33 engaged, the driving force thereof is the input shaft 18 → toroidal transmission mechanism 22 → hydraulic clutch 33 → first gear 32 → second gear 38 → final drive gear 39 → final. The vehicle is transmitted to the axles 41 and 41 through a path of the driven gear 40 → the differential gear D, and the vehicle travels forward with the driving force of the engine E. At this time, the gear ratio can be arbitrarily changed by driving the trunnion of the toroidal transmission mechanism 22 to change the tilt angle of the power rollers 26 and 26.

Further, the electric motor M can be driven when the vehicle starts, and both the engine E and the electric motor M can be driven when traveling uphill, and the electric power for driving the electric motor M at that time is the electric motor M by the engine E. Can be obtained by driving as a generator,
The hydraulic loader 27 of the present embodiment can generate a loading thrust by the first oil chamber 54 and the first piston 52 and a loading thrust by the second oil chamber 55 and the second piston 53. Since the loading thrust by the oil chamber 54 and the first piston 52 is generated when the first piston 52 presses the second cylinder housing 23a provided at the radially outer end of the input disk 23, the input disk 23 has the second cylinder. The loading thrust is concentrated on the radially outer end provided with the housing 23a, and there is a concern that the deformation of the input disk 23 due to the loading thrust becomes large.

  On the other hand, the loading thrust by the second oil chamber 55 and the second piston 53 is generated by the oil pressure of the second oil chamber 55 being uniformly pressed across the entire radial direction of the back surface of the input disk 23. Deformation of the input disk 23 due to thrust is relatively small.

  Therefore, in the present embodiment, the second oil chamber 55 and the second piston 53 are preferentially used to generate the required loading thrust, and when the loading thrust is insufficient, the first oil chamber 54 and the first piston 52 is used to generate an insufficient loading thrust so that the deformation of the input disk 23 is minimized.

  That is, when the input torque input from the engine E to the input shaft 18 is small, the switching valve 59 is demagnetized and the first oil passage P1 is shut off from the third oil passage P3, and the hydraulic loader linear solenoid valve 58 is used. The regulated loading oil pressure is supplied only from the first oil passage P1 to the second oil chamber 55 of the hydraulic loader 27, and is not supplied to the first oil chamber 54.

  When the input torque is small, the loading thrust required to suppress the slipping of the power rollers 26 and 26 is also small, so that the slipping of the power rollers 26 and 26 is sufficiently achieved only by supplying the loading oil pressure to the second oil chamber 55. Can be suppressed. At this time, if the loading hydraulic pressure is supplied only to the first oil chamber 54, the radially outer portion of the input disk 23 is greatly bent toward the output disk 24, so that the control accuracy of the transmission ratio is reduced or power transmission is performed. Although the efficiency decreases, at this time, the deformation of the input disk 23 is minimized by supplying the loading oil pressure only to the second oil chamber 55 and shutting off the supply of the loading oil pressure to the first oil chamber 54. be able to.

  On the contrary, when the input torque input from the engine E to the input shaft 18 is large, the switching valve 59 is excited to connect the first oil path P3 to the third oil path P3, whereby the hydraulic loader linear solenoid valve 58 is excited. The loading hydraulic pressure adjusted in step S1 is supplied from the first oil passage P1 to both the second oil chamber 55 and the first oil chamber 54 of the hydraulic loader 27.

  When the input torque is large, the loading thrust required to suppress the slip of the power rollers 26 and 26 also becomes large. Therefore, by supplying the loading hydraulic pressure to both the second oil chamber 55 and the first oil chamber 54, a large amount is obtained. A slippage of the power rollers 26 and 26 can be reliably suppressed by generating a loading thrust. At this time, the loading hydraulic pressure supplied to the second oil chamber 55 and the first oil chamber 54 uniformly presses the radially inner portion and the radially outer portion of the input disc 23, so that the deformation of the input disc 23 is minimized. Can be suppressed.

  The above operation will be described with reference to the flowchart of FIG. 5. First, in step S1, the input torque input from the engine E to the toroidal continuously variable transmission T is calculated, and then in step S2, a double piston state or a single piston state. The loading oil pressure corresponding to each is calculated. The loading hydraulic pressure in the state of the double piston is the loading hydraulic pressure that should be generated by the linear solenoid valve 58 for the hydraulic loader when the hydraulic pressure is supplied to both the second oil chamber 55 and the first oil chamber 54. The loading hydraulic pressure is a loading hydraulic pressure that should be generated by the hydraulic loader linear solenoid valve 58 when the hydraulic pressure is supplied only to the second oil chamber 55.

  If the switching between the double piston state and the single piston state is in progress in step S3, the switching process is continued in step S7. If the switching is not in progress, switching determination is made in step S4 based on the magnitude of the input torque. . That is, if the loading oil pressure does not change across the threshold value in step S5, the current double piston state or single piston state loading oil pressure is selected in step S6, and the loading oil pressure is switched from single to double in step S5. If the state changes from a state less than the first threshold value to a state equal to or higher than the first threshold value, switching from the loading hydraulic pressure in the single piston state to the loading hydraulic pressure in the double piston state is executed in step S7. On the other hand, if the input torque changes from the state above the second threshold at the time of switching from double to single at the step S5 to a state below the second threshold at the step S5, the loading hydraulic pressure from the double piston state is changed to the single at the step S7. The piston state is switched to the loading hydraulic pressure.

  In step S8, the loading hydraulic pressure is converted into a control current for the hydraulic loader linear solenoid valve 58. In step S9, the hydraulic loader linear solenoid valve 58 is driven with the control current, thereby supplying the hydraulic pressure to the hydraulic loader 27. .

  An example of control according to the flowchart of FIG. 5 is shown in the time chart of FIG. FIG. 6A shows the input torque of the toroidal type continuously variable transmission T calculated in step S1, and shows a case where the input torque increases stepwise. FIGS. 6B and 6C show changes in the required loading oil pressure corresponding to the single piston state and the double piston state calculated in step 23. The required loading oil pressure in the single piston state is as follows. The required loading hydraulic pressure in the double piston state is larger. As shown in FIG. 6 (B), when the driver suddenly depresses the accelerator pedal and the increase amount of the required loading hydraulic pressure in the single piston state becomes equal to or greater than the threshold value (see step S4), FIG. As shown in FIG. 6D, the single piston state is switched to the double piston state (see step S5), and the target loading hydraulic pressure in the double piston state is selected as shown in FIG. 6E (step S5). (See S6 and Step S7).

  As shown in FIG. 7, in the single piston state, the increase in the transmittable torque with respect to the increase in the loading hydraulic pressure is small, whereas in the double piston state, the increase in the transmittable torque with respect to the increase in the loading hydraulic pressure is large. . Therefore, when the transmittable torque is increased from A to B, the loading hydraulic pressure needs to be increased greatly from p2 to p3 in the single piston state, whereas the loading hydraulic pressure is slightly increased from p1 to p2 in the double piston state. It is only necessary to increase it, and the response of the transmittable torque is improved.

  As described above, it is possible to improve the responsiveness of the transmittable torque by setting the hydraulic loader 27 to the double piston state. However, when the hydraulic loader 27 is in the single piston state, the loading hydraulic pressure is not supplied. When oil in one oil chamber 54 leaks and air enters, the first oil chamber moves to a double piston state and loading oil is supplied to the first oil chamber 54 until the mixed air is discharged. Since the rising of the loading hydraulic pressure 54 is delayed, there is a possibility that the response of the transmittable torque may be reduced.

  For example, as shown in FIG. 8, when the transmittable torque is increased from A to B in the double piston state, it is necessary to increase the loading hydraulic pressure from p1 to p3. However, air enters the first oil chamber 54. If so, the loading hydraulic pressure increases only to p2 lower than the target value p3, and the target value B of the target transmittable torque may not be obtained.

  However, according to the present embodiment, the lubricating oil pressure is supplied from the second oil passage P2 to the first oil chamber 54 of the hydraulic loader 27 by the switching valve 59 shown in FIG. It is possible to prevent air from entering the oil chamber 54 and ensure the responsiveness of the hydraulic loader 27. Since the lubricating oil pressure is much lower than the loading oil pressure, it does not affect the loading thrust generated by the hydraulic loader 27, but is sufficient to prevent air from entering.

  When the power supply fails, the switching valve 59 is in the initial state shown in FIG. 4 by the spring force of the spring, and the lubricating oil pressure is supplied from the second oil passage P2 to the first oil chamber 54 of the hydraulic loader 27. Thus, the hydraulic loader 27 is automatically shifted to the single piston state in the event of failure, thereby preventing an excessive loading thrust from acting.

  The toroidal type continuously variable transmission T has been described above. However, when the hydraulic loader for generating the side pressure of the pulley is a double piston type in the belt type continuously variable transmission, the technique of the present invention is applied. Thus, it is possible to prevent an excessive loading thrust from being generated while preventing the endless belt from slipping.

  The embodiments of the present invention have been described above, but various design changes can be made without departing from the scope of the present invention.

  For example, the toroidal transmission mechanism 22 of the embodiment is of a single cavity type, but may be of a double cavity type.

  The drive source of the present invention is not limited to the engine E of the embodiment, and may be any drive source such as an electric motor.

18 Input shaft (rotary shaft)
23 input disk 23a second cylinder housing 24 output disk 26 power roller 27 hydraulic loader 28 trunnion shaft 51 first cylinder housing 51b side wall 52 first piston 53 second piston 54 first oil chamber 55 second oil chamber 57 hydraulic control circuit 59 Switch valve P1 First oil passage P2 Second oil passage E Engine (drive source)

Claims (3)

A rotary shaft (18) connected to a drive source (E), an input disk (23) supported so as not to rotate relative to the rotary shaft (18), and a rotary shaft supported by the rotary shaft (18). An output disk (24), a power roller (26) supported in a tiltable manner around the trunnion shaft (28) and sandwiched between the input disk (23) and the output disk (24), and the input disk A hydraulic loader (27) for urging (23) in a direction approaching the output disk (24);
The hydraulic loader (27) is fitted to a first cylinder housing (51) fixed to the rotating shaft (18) and an inner peripheral surface of the first cylinder housing (51) so as to be axially slidable. A first piston (52) that abuts on an axial end of a second cylinder housing (23a) projecting from the back surface of the input disk (23), and a second piston housing fixed to the rotating shaft (18). A second piston (53) fitted to the inner peripheral surface of (23a) so as to be axially slidable, and defined between a side wall (51b) of the first cylinder housing (51) and the first piston (52). A first oil chamber (54), a second oil chamber (55) defined between the back surface of the input disk (23) and the second piston (53), the first oil chamber (54) and the Oil supplied to the second oil chamber (55) A toroidal type continuously variable transmission and a hydraulic control circuit (57) for controlling,
The hydraulic control circuit (57) is a first control that supplies a control hydraulic pressure only to the second oil chamber (55) based on an input torque input from the drive source (E) to the rotary shaft (18). And a second control state in which a control hydraulic pressure is supplied to both the first oil chamber (54) and the second oil chamber (55) can be switched, and the first oil chamber in the first control state can be switched. A toroidal continuously variable transmission characterized in that a lubricating oil pressure lower than the control oil pressure is supplied to (54).   2. The toroidal continuously variable transmission according to claim 1, wherein when the input torque increases by a predetermined value or more in the first control state, the first control state shifts to the second control state.   A first oil passage (P1) for supplying control oil pressure to the first oil chamber (54), a second oil passage (P2) for supplying lubricating oil pressure to the first oil chamber (54), and the first oil And a switching valve (59) for selectively connecting the path (P1) and the second oil path (P2) to the first oil chamber (54), and the switching valve (59) is the second valve in an initial state. The toroidal continuously variable transmission according to claim 1 or 2, wherein an oil passage (P2) is connected to the first oil chamber (54).

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