JP2007271060A - Toroidal type continuously variable transmission - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To immediately push back a power roller into a predetermined movement range when the same exceeds the predetermined movement range. <P>SOLUTION: A hydraulic control device A is provided with end stop levers 48A, 48B rotated when a rotation center d of the power roller 4 exceeds the predetermined movement range, and a first end stop valve 115 and a second end stop valve 116 pressed and changed over to directly supply line pressure PL to hydraulic servos 40A1, 40B1 or hydraulic servo 40A2, 40B2 with bypassing a first or a second reaction pressure control valve 112, 115 for regulating working pressure when the end stop lever 48A or the end stop lever 48B is rotated. Consequently, the power roller 4 can be immediately pushed back to the predetermined movement range. <P>COPYRIGHT: (C)2008,JPO&INPIT

Description

  The present invention relates to a toroidal continuously variable transmission mounted on, for example, an automobile, a working vehicle, and the like, and more specifically, a toroidal continuously variable whose rotation center of a power roller can move with respect to the surface direction of an input disk and an output disk. It relates to a transmission.

  In recent years, a toroidal continuously variable transmission has been proposed as a transmission for a vehicle, which has an input disk, an output disk, and a power roller sandwiched between the two disks, and enables shifting by position control of the power roller. (See Patent Document 1). In this system, the supply pressure of the oil pump (oil supply pump 71) is adjusted to the line pressure by the pressure adjusting valve (72), and the line pressure is further adjusted by the pressure adjusting valve (control valve 75), and the hydraulic servo (shifting) is performed. By supplying the control hydraulic cylinders 76, 76) as operating pressure, the position of the power roller is pressed and controlled. The power roller moved by being pressed and driven with respect to the surface direction of both discs, depending on the rotation direction of both discs, one of the clamping (contact) portions of both discs is on the inner peripheral side and the other is on the outer peripheral side. Since the tilt angle is changed by dragging, the speed ratio is changed by autonomously changing the diameter of the contact portion, and thereby the speed change is performed.

JP 11-82658 A

  However, when the relative rotation of both disks suddenly changes, for example, when the output disk suddenly decelerates due to sudden braking by the foot brake, the force on the surface direction of both disks suddenly changes in the power roller, The hydraulic servo pressing drive control may be delayed. In such a case, there is a possibility that the power roller is moved greatly and exceeds the set moving range in which the gear can be normally shifted. In this case, the power roller is immediately pushed back into the moving range. It is desired to immediately return the shift control of the power roller to the control by the press drive control of the hydraulic servo.

  Accordingly, the present invention provides a toroidal continuously variable transmission that can rapidly increase the operating pressure of a hydraulic servo when detecting that a power roller has exceeded a predetermined movement range, thereby solving the above-described problems. It is for the purpose.

The present invention according to claim 1 (see, for example, FIGS. 1 to 8), an input disk (2A, 2B), an output disk (3), and a variator device having a power roller (4) sandwiched between these disks ( 5) and
A hydraulic servo (40A1, 40A2, 40B1, 40B2) and a link mechanism (41, 42, 45, 46) for supporting the power roller (4) are provided, and the hydraulic servo (40A1, 40A2, 40B1, 40B2) A roller driving device (B) that can move and drive the rotational center (d) of the power roller (4) with respect to the surface direction of the two disks via the link mechanism (41, 42, 45, 46) based on the pressing drive. )When,
A source pressure regulating valve (111) for regulating the hydraulic pressure of the hydraulic pressure generation source (110) as a source pressure (PL);
A toroidal stepless variable pressure control valve (112, 113) that regulates the original pressure (PL) as an operating pressure (PS1, PS2) to be supplied to the hydraulic servo (40A1, 40A2, 40B1, 40B2). In the transmission (1),
Out-of-range detection means (48A, 48B) for detecting that the rotation center (d) of the power roller exceeds a predetermined movement range;
When the out-of-range detection means (48A, 48B) detects that the rotation center (d) of the power roller (4) has exceeded a predetermined movement range, it bypasses the operating pressure regulating valve (112, 113). And a bypass communication valve (115, 116) for supplying the original pressure (PL) as an operating pressure (PS1, PS2) of the hydraulic servo (40A1, 40A2, 40B1, 40B2),
The toroidal-type continuously variable transmission (1) is characterized in that.

The present invention according to claim 2 (see, for example, FIG. 4 and FIG. 8) includes a plurality of the operating pressure regulating valves (112, 113),
The hydraulic servos (40A1, 40A2, 40B1, 40B2) input the different operating pressures (PS1, PS2) to two oil chambers (40Aa, 40Ab, 40Ba, 40Bb) facing the pistons (40Ap, 40Bp), A power roller capable of driving the rotational center (d) of the power roller (4) against the surface direction of the disks (2A, 2B, 3) via the link mechanism (41, 42, 45, 46). Hydraulic servos for control (40A1, 40A2, 40B1, 40B2),
The bypass communication valve (115, 116) acts on the piston (40Ap, 40Bp) with respect to the direction in which the original pressure (PL) is driven to press the power roller (4) within the predetermined movement range. Supplied to the oil chamber (40Aa, 40Ab, 40Ba, 40Bb)
A toroidal continuously variable transmission (1) according to claim 1.

According to a third aspect of the present invention (see, for example, FIGS. 4 and 8), the link mechanism (41, 42, 45, 46) is configured such that both the hydraulic servos (40A1, 40A2, 40B1, 40B2) are driven by pressing. A lever member (45A, 45B) that can be driven to rotate on a plane parallel to the disks (2A, 2B, 3), and a swingable and rotatable connection to the lever member, and the power roller (4) A roller carriage (41) that rotatably supports the center of rotation (d),
The out-of-range detection means (48A, 48B) is rotated by contacting and pressing the lever member (45A, 45B) when the lever member (45A, 45B) rotates more than a predetermined angle. And an end stop lever (48A, 48B) for detecting that the rotation center (d) of the power roller (4) exceeds a predetermined movement range,
When the end stop lever (48A, 48B) is rotated, the bypass communication valve (115, 116) is spooled against the spring (115g, 116g) by the end stop lever (48A, 48B). 115f, 116f) are pressed and driven, and the input ports (115b, 116b) for inputting the original pressure (PL) and the output ports (115a, 116a) directly communicating with the hydraulic servos (40A1, 40A2, 40B1, 40B2). Are connected to each other, thereby bypassing the operating pressure regulating valves (112, 113) and supplying the original pressure (PL) as the operating pressure of the hydraulic servo (40A1, 40A2, 40B1, 40B2). (115, 116)
It exists in the toroidal type continuously variable transmission (1) of Claim 1 or 2.

According to a fourth aspect of the present invention (see, for example, FIGS. 1 to 4 and 8), the variator device (5) is a full toroidal type.
A toroidal continuously variable transmission (1) according to any one of claims 1 to 3.

  In addition, although the code | symbol in the said parenthesis is for contrast with drawing, this is for convenience for making an understanding of invention easy, and has no influence on the structure of a claim. It is not a thing.

  According to the first aspect of the present invention, when the out-of-range detecting means detects that the rotation center of the power roller exceeds the predetermined moving range, the bypass communication valve bypasses the operating pressure regulating valve and the original pressure Is supplied as the operating pressure of the hydraulic servo, so that the operating pressure of the hydraulic servo can be rapidly increased when the rotation center of the power roller exceeds a predetermined movement range. As a result, when the power roller exceeds the predetermined moving range, the power roller can be immediately pushed back into the predetermined moving range, and the shift control of the power roller is hydraulically controlled. It is possible to return immediately under the control by the servo pressing drive control.

  According to the second aspect of the present invention, the toroidal continuously variable transmission includes a plurality of operating pressure regulating valves, and the hydraulic servo inputs different operating pressures to the two oil chambers facing the piston, via the link mechanism. Since this is a hydraulic servo for power roller control that can drive the rotation center of the power roller against the surface direction of both disks, the piston is applied to the original pressure in the direction of pressing and driving the power roller within a predetermined movement range. It can be supplied to the oil chamber acting on. Thus, while the power roller can be immediately pushed back into the predetermined movement range, the shift control of the power roller can be immediately returned under the control by the hydraulic servo pressing drive control.

  According to the third aspect of the present invention, the out-of-range detection means comprises an end stop lever, and the bypass communication valve supplies the original pressure as the hydraulic servo operating pressure when the end stop lever is rotated. Because it consists of an end stop valve, it detects that the power roller has exceeded the predetermined movement range and detects the operation until the hydraulic servo operating pressure is suddenly increased, for example, by a sensor, without electrically increasing the hydraulic pressure. It can be performed mechanically, and can be made inexpensive and compact.

  According to the fourth aspect of the present invention, since the variator device is of a full toroidal type, it is possible to shift the power roller autonomously by moving and driving the power roller on a plane parallel to the two disks. Can do.

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

  As shown in FIG. 1, the continuously variable transmission (IVT) 1 includes a toroidal continuously variable transmission (hereinafter simply referred to as a variator) 5, a planetary gear mechanism 6, a reverse gear mechanism 7, and a low / high switching mechanism 10. And a planetary gear unit U. The variator 5 is a full toroidal continuously variable transmission, and includes two input disks 2A and 2B connected to the input shaft 12, one output disk 3 connected to the hollow shaft 16, and a gap between the two disks. Power rollers 4 and 4 to be sandwiched. The input disks 2A and 2B and the output disk 3 have arc-shaped concave grooves 2a and 3a that form part of a circular shape so as to face each other, and double cavities 31A and 31B sandwiching two rows of power rollers. To cancel the thrust force between the input disks. The power rollers 4 and 4 are inclined by shifting in a direction perpendicular to the shaft, and are continuously variable by changing the contact radius between the input disks 2A and 2B and the output disk 3. The variator 5 has a speed ratio (output speed / input speed) of −0.4 to −2.5. Since the output disk 3 is inverted with respect to the input disks 2A and 2B, the speed ratio is-(minus).

  The planetary gear mechanism 6 includes a front carrier C having two pinions P1 and P2 and a low-plane simple planetary gear 11. The low-mode simple planetary gear 11 includes a dual planetary gear 14 of the reversing gear mechanism 7 and It has a common rear carrier C0. The two pinions P1 and P2 have an integral structure that is rotatably supported by a common pinion shaft, and the front carrier C that supports these pinions is connected to the first ring gear R3 of the low-mode simple planetary gear 11. It is connected. The front carrier C is connected to the input shaft 12 and is connected to the input disk 2B on the rear side, so that constant speed rotation is transmitted.

  The first sun gear S1 connected to the output disk 3 of the variator 5 is meshed with the first pinion P1, and the first sun gear S1 serves as an input gear for inputting variable speed rotation by the variator 5. . The second pinion P2 meshes with the second sun gear S2 serving as the high mode output gear, and is integrally coupled to the third sun gear S3 of the low mode simple planetary gear 11. The second sun gear S2 is connected to the output shaft 13 via the high clutch H of the low / high switching mechanism 10 to constitute a high mode output gear.

  The reverse gear mechanism 7 includes a dual planetary gear 14 having two pinions P4 and P5 meshing with each other, and the rear carrier C0 is integrally formed with the carrier of the low mode simple planetary gear 11 as described above. The ring gear R0 is fixed to the case 20, and the fourth sun gear S0 is connected to the output shaft 13 via the low clutch L.

  The continuously variable transmission 1 is mainly used in an automobile or the like, and an output shaft 13 is connected to a drive wheel via a differential device (not shown), and the output shaft 13 is inverted by a variator 5 or the like. The rotation is configured to be reversed again by the differential device.

  As shown in FIG. 3, a continuously variable transmission (IVT) 1 is housed in a mission case 20, and the case 20 has a cylindrical main case 20a and a housing fixed to the front side of the main case 20a. 20b and an end case 20c fixed to the rear side of the main case 20a. The front end of the housing 20b is coupled to the engine, and a damper device (not shown) is accommodated. That is, since the continuously variable transmission 1 has a gear neutral (GN), which will be described in detail later, the start of a torque converter or the like required for the conventional automatic transmission (AT) and the continuously variable transmission (CVT). A device is not required, and therefore, only a damper device that absorbs vibration and pulsation of the engine is sufficient in the housing 20b.

  The main case 20a is provided with the variator 5, the planetary gear mechanism 6 and the gear mechanism of the reversing gear mechanism 7. The end case 20c is provided with the low / high switching mechanism 10, and the rear end surface and end of the main case 20a are arranged. The front end face of the case 20c is connected and fixed to form an integral storage portion. A lower part of the main case 20 a is opened, and the opening is closed by an oil pan 21. In the oil pan 21 portion, the pump / valve block 22 in which the oil pump and the valve body are integrated and the fulcrum block 23 are housed, and these blocks 22 and 23 are fixed together. The main case 20a is fixed to the opening.

  A front partition plate 25 is fixed so as to be sandwiched between the housing 20b and the main case 20a, and a front portion of the input shaft 12 is rotated to a central boss portion 25a of the partition plate 25 via a needle bearing (not shown). It is supported freely. The distal end portion 12a of the input shaft 12 extends into the housing 20b and is connected to the engine output shaft via a damper device in the housing 20b. A dish-like support plate 27 is integrally fixed to the input shaft 12 and is supported adjacent to the support plate 27 on the front input disk 2A. The input disk 2A is spline-engaged with the support plate 27 at the outer periphery thereof, and a pressing device 29 is disposed between the back surface of the input disk 2A and the support plate 27, and the input disk 2A Is rotated integrally with the input shaft 12 and a necessary pressing force is applied to the variator 5.

  An output disk 3 is rotatably supported on the input shaft 12 via a needle bearing, and a sleeve shaft 16 is integrally connected to a boss portion 3 b of the output disk 3, and the sleeve shaft 16 is connected to the input shaft 12. The first sun gear S1 is formed at the rear end of the input shaft 12 so as to be fitted to the input shaft 12 and extending rearward. A rear input disk 2B is disposed so as to fit on the sleeve shaft 16. The rear input disk 2B is supported by a fulcrum block 23 as will be described later, and is not in contact with or without contact with the sleeve shaft 16. The load is maintained.

  A plurality of (three) power rollers 4 are sandwiched between an annular arc surface 2a formed on the front input disk 2A and an annular arc surface 3a formed in front of the output disk 3, respectively. A plurality of (three) power rollers 4 are sandwiched between an annular arc surface 2a formed on the rear input disk 2B and an annular arc surface 3a formed on the rear side of the output disk 3. Two cavities 31A and 31B having a circular cross section are formed by the arc surfaces 2a and 3a facing each other, and the center of the power roller 4 is at the center of each of the circular cavities 31A and 31B. It rotates about the passing center axis d. That is, the variator 5 is made of a full toroidal double cavity type, and thus has a double cavity (2 rows), so that the torque capacity is doubled compared to that of a single row, and the input disks and output disks are Thus, the thrust force is canceled out, the bearing load is not applied, and the two contact points of the power roller 4 are opposite to each other on the center point of the cavity 31, so that the thrust force is hardly applied to the power roller 4. Does not work.

  The fulcrum block 23 has an L-shaped front sectional view, and includes a variator control block 23a at the bottom and a support block 23b that rises from the rear end side of the bottom. As shown in FIG. 4, the control block 23 a is accurately positioned and fixed to the opening 20 e of the main case 20 a by a plurality of knock pins 33 and a large number of bolts 35. As schematically shown in FIG. 4, the control block 23 a includes hydraulic actuators 40 connected to the three rollers 4 located in the cavities 31 </ b> A and 31 </ b> B via links. Specifically, the control block 23a of the fulcrum block is formed with the same number of hydraulic actuators 40 as the number of rollers 4 and a hydraulic circuit communicating with each actuator. Each actuator 40 is connected to the pump / valve block 22 ( The controlled predetermined hydraulic pressure from (see FIG. 3) acts.

  The pressing device 29 (see FIG. 3) for the support plate 27 has a spring (plate spring) that acts on a predetermined preload and a hydraulic actuator on which the hydraulic pressure from the pump / valve block 22 acts. Each of the power rollers 4 exerts a pressing force that presses the arc surfaces 2a and 3a of the input disks 2A and 2B and the output disk 3 with a predetermined pressure. Traction force acts between the disc and the disc. In other words, the variator 5 transmits power between the input disks 2A and 2B and the output disk 3 by a traction drive.

  As shown in FIG. 4, each of the power rollers 4 is supported by a carriage 41 so as to be rotatable about an axis d, and the carriage 41 is supported by a spherical bearing 42 at one end thereof. Each spherical bearing 42 is provided at the tip of a lever 45 that is swingably supported by a support shaft 43, and each lever 45 is connected to a piston rod 46 of the hydraulic actuator 40. As a result, as will be described in detail later, the inclination of the power roller 4 is autonomously controlled based on the relationship between the traction force acting on the contact surface between the roller and the disk and the pressing force (reaction force) of the hydraulic actuator 40. Is done. That is, the variator 5 is controlled to be shifted by the hydraulic actuator 40.

  On the other hand, the rising support block portion 23b of the volkram block 23 is provided with a cylindrical support hole 50 as shown in detail in FIG. An annular protrusion 54 is integrally formed on the rear surface of the rear input disk 2B, and a boss 55 is integrally formed so as to protrude from the central portion of the protrusion. A ball bearing 51 and a one-way clutch 52 are disposed adjacent to the support hole 50 in the axial direction.

  In the full toroidal variator 5, no thrust force acts on the power roller 4, and shifting can be performed with a small force by shifting the power roller 4 in the direction orthogonal to the axis as described above. The clamping force of the power roller 4 is loosened. For this reason, even if reverse torque is applied to the variator 5 by reverse rotation of the engine or reverse drive from the wheels, the one-way clutch 52 prevents the variator 5 from rotating backward. 3 is a sprocket fixed to the boss portion 58 of the carrier C integral with the boss portion 55 of the input disk. The oil pump (hydraulic pressure generating source) of the pump / valve block 22 is connected via the chain 57. ) 110 (see FIG. 8) is driven.

  As shown in FIG. 3, the front carrier C of the planetary gear mechanism 6 includes a carrier body 61 and a carrier cover 62 fixed to the body. The carrier body 61 is splined on the rear end side of the input shaft 12. The nut 63 is screwed together and the carrier C integrated with the input shaft 12 is fixed. The carrier body 62 has a boss portion 58 whose front end is a dock m, and the front (rear) of the rear boss portion 55 of the rear input disk 2B is also a dock n. 58 and 55 are coupled together in the rotational direction by the engagement of the docks m and n, and the carrier body 61 presses the ball bearing 51 against the back surface k of the input disk 2B via the sprocket 56. .

  A pinion shaft 67 is supported across the carrier body 61 and the carrier cover 62, and a first pinion P1 and a second pinion P2 are supported on the pinion shaft 67 in parallel in the axial direction. Yes. The first pinion P1 and the second pinion P2 are integrally formed and may have the same number of teeth. However, in the present embodiment, the number of teeth is slightly different, and these common pinions P1 and P2 are needle bearings (or The pinion shaft 67 is rotatably supported via a bush. The first sun gear S1 is meshed with the first pinion P1, and the second sun gear S2 is meshed with the second pinion P2.

  The first sun gear S1 is formed at the distal end portion of the hollow shaft 16, and the hollow shaft 16 is rotatably supported by being fitted to the input shaft 12 via a needle bearing. As described above, the end portion is connected to the output disk 3 of the variator 5. The second sun gear S2 is formed at the base end portion of the intermediate shaft 70, and the intermediate shaft 70 is fitted on the input shaft 12 (the front carrier C integral with the input shaft 12) via a needle bearing. While being rotatably supported, the tip (rear) side is connected to the clutch hub 72 of the high clutch H of the low / high switching mechanism 10 by spline engagement. A third sun gear S3 is spline-engaged with the intermediate shaft 70 and coupled with a snap ring so as to be removed, and a first ring gear R3 is integrally engaged with the carrier cover 62 of the front carrier C by a snap ring. Are combined. The simple planetary gear 11 of the planetary gear mechanism 6 is configured by the rear carrier C0 integrated with the third sun gear S3, the first ring gear R3, and the reverse gear mechanism 7.

  The output (common) carrier C0 includes a carrier body 76 located at the center and front and rear carrier covers 77 and 78 constituting left and right side plates. The carrier body 76 is rotatably supported by the intermediate shaft 70, and a pinion shaft 80 is non-rotatably supported between the carrier body 76 and the front carrier cover 77. A pinion P3 is rotatably supported. The third pinion P3 meshes with the third sun gear S3 and the first ring gear R3 to constitute the simple planetary gear 11.

  A pinion shaft 81 and a pinion shaft 82 are rotatably supported between the carrier body 76 and the rear carrier cover 78, and a fourth pinion P 4 is rotatably supported on the pinion shaft 81. The fifth pinion P5 is rotatably supported. Both the pinions P4 and P5 mesh with each other, and one fourth pinion P4 meshes with the fourth sun gear S0, and the other fifth pinion P5 meshes with the second ring gear R0 and reverses. A dual planetary gear 14 for the gear mechanism 7 is configured. The second ring gear R0 is connected to the transmission case 20a by spline engagement and is fixed in the rotational direction. A low clutch L hub 83 is integrally formed on the fourth sun gear S0 by welding or the like. ing. The clutch hubs 72 and 83 constitute a low / high switching mechanism 10 by interposing a low clutch L and a high clutch H, each of which is a wet multi-plate clutch, between the output shaft 13 and a drum 85 integral with the output shaft 13. Yes.

  A boss member 95 is integrally fixed to the end plate 20d of the end case 20c by bolts, and the output shaft 13 is supported by the integrated end case 20c via a ball bearing 96 and a needle bearing. The output shaft 13 is integrally connected to the sleeve shaft 99 fitted to the boss 95 at the tip (front) end portion, and the sleeve shaft 99 and the clutch drum 85 are integrally connected. ing. The clutch drum 85 and the sleeve shaft 99 are used as cylinders, and an oil chamber formed between the piston 100 and the piston 100 is fitted so as to be oil-tight. The low clutch hydraulic servo 101 is configured by contacting the low clutch L through the formed hole. The cylinder plate 102 is fixed to the sleeve shaft 99, and the oil chamber formed between the cylinder plate 102 and the sleeve shaft 99 as a cylinder and the piston 103 is fitted in an oil-tight manner. The piston 103 is in contact with the high clutch H to form a high clutch hydraulic servo 105.

  Next, the operation of the continuously variable transmission (IVT) 1 will be described with reference to FIGS.

  For example, when the vehicle equipped with the continuously variable transmission 1 starts or reverses, the low / high switching mechanism 10 is controlled based on hydraulic control by a shift lever (not shown) or the hydraulic control device A, and the high clutch H is released. At the same time, the low clutch L is engaged, and the continuously variable transmission 1 is brought into the low mode. Then, as shown in FIGS. 1 and 2, the rotation of the input shaft 12 connected to the engine output shaft causes the input disks 2A and 2B of the variator 5, the front carrier C of the planetary gear mechanism 6, and the simple planetary gear for low mode. 11 to the first ring gear R3. Of these, the rotation of the input shaft 12 input to the input disks 2A and 2B is shifted by the variator 5, and the variator output rotation Vout is output from the output disk 3 and input to the first sun gear S1.

  When the variator output rotation Vout is input to the first sun gear S1, the rotation of the input shaft 12 and the variator output rotation Vout are synthesized by (primary) torque circulation in the planetary gear mechanism 6, and based on the gear ratio, It is output from the second sun gear S2 as a slightly increased rotational speed from the variator output rotation Vout, and is input to the third sun gear S3. Then, in the low-mode simple planetary gear 11, the rotation of the input shaft 12 inputted to the first ring gear R3 and the above-mentioned speed-up rotation of the third sun gear S3 are combined by (secondary) torque circulation. And output from the rear carrier C0. The output rotation of the rear carrier C0 is an output rotation that is shifted to a width from the reverse rotation of the deceleration to the normal rotation of the deceleration via the neutral position (GN point) according to the width of the gear ratio of the variator 5.

  The output rotation of the rear carrier C0 is input to the dual planetary gear 14 of the reversing gear mechanism 7, is reversed via the second ring gear R0 fixed to the case 20, and is output from the fourth sun gear S0. The output rotation OutL of the fourth sun gear S0 is output to the output shaft 13 via the low clutch L as the output rotation in the low mode state.

  In the low mode in which the transmission path as described above is formed, when the variator output rotation Vout (gear ratio of the variator 5) is in the gear neutral state GN indicated by the one-dot chain line in FIG. The rotation is in the neutral state, and the rotation reversed in the reversing gear mechanism 7, that is, the output rotation OutL in the low mode is in the neutral state. As described above, in this state, the engine speed (the rotation of the input shaft 12) and the rotation of the output shaft 13 are irrelevant. For example, when switching to the travel range, the gear ratio of the variator 5 is changed to the gear neutral state GN. By engaging the low clutch L after adjusting to the above, it is not necessary to absorb the rotational speed difference, and there is no need to provide a device for absorbing the rotational speed difference such as a torque converter.

  Here, for example, a shift lever (not shown) is in the reverse (R) range, and when the gear ratio of the variator 5 is increased in accordance with the vehicle speed or the accelerator opening, for example, from this gear neutral state GN (in FIG. 2). When the variator output rotation Vout is shifted downward), the output rotation OutL of the output shaft 13 is accelerated to the forward rotation side, reversed by the differential device, that is, accelerated to the reverse side. Go.

  On the other hand, for example, when a shift lever (not shown) is in the drive (D) range and the gear ratio of the variator 5 is made smaller than the gear neutral state GN according to, for example, the vehicle speed and the accelerator opening (in FIG. 2). When the variator output rotation Vout is shifted upward), the output rotation OutL of the output shaft 13 is accelerated toward the reverse rotation side, reversed by the differential device, that is, accelerated toward the forward side. .

  Subsequently, in the low mode state described above, the output rotation OutL of the output shaft 13 is increased (the transmission ratio of the variator 5 is decreased), and reaches the transmission ratio of the sync change SC indicated by the broken line in FIG. For example, when a shift determination is made according to the vehicle speed or the accelerator opening, the low / high switching mechanism 10 is controlled based on the hydraulic control by a hydraulic control device (not shown), the low clutch L is released and the high clutch H is released. Are engaged, and the continuously variable transmission 1 is brought into a high mode state.

  Then, as shown in FIGS. 1 and 2, similarly in the high mode state, the rotation of the input shaft 12 is transmitted to the input disks 2A and 2B of the variator 5 and the front carrier C of the planetary gear mechanism 6 to output the output disk. 3, the variator output rotation Vout is output and input to the first sun gear S1. When the variator output rotation Vout is input to the first sun gear S1, the planetary gear mechanism 6 combines the rotation of the input shaft 12 and the variator output rotation Vout by torque circulation, and is slightly smaller than the variator output rotation Vout based on the gear ratio. Is output from the second sun gear S2 as a speed increasing rotation. The output rotation OutH of the second sun gear S2 is output to the output shaft 13 via the high clutch H as the output rotation in the high mode state.

  In switching between the low mode state and the high mode state at the time of the sync change SC, the gear ratios of the respective gears are set so that switching is performed at the same gear ratio at which the gear ratio of the variator 5 (variator output rotation Vout) is minimized. Is set. In other words, in the low mode state, when the gear ratio of the variator 5 is reduced, the output rotation OutL is increased, and in the high mode state, on the contrary, the gear ratio of the variator 5 is increased with the gear change. As it is done, the output rotation OutH is increased.

  In the high mode, the output rotation OutH of the second sun gear S2 is input to the third sun gear S3, and the rotation of the input shaft 12 is input to the first ring gear R3. The rotation is performed in the same manner as in the low mode, but the fourth sun gear S0 rotates idly because the low clutch L is released. Therefore, torque transmission is not performed in the low planetary simple planetary gear 11 and the dual planetary gear 14 of the reversing gear mechanism 7.

In the continuously variable transmission 1, the number of teeth of the first and second pinions P1 and P2 is set to Z _P1 and Z _P2 , and the number of teeth of the first and second sun gears S1 and S2 is set to Z _S1 and Z _S2. When the gear ratio (Z _S1 / Z _P1 ) × (Z _P2 / Z _S2 ) is 1 or more, the output rotation OutH in the high mode (that is, the rotation of the second sun gear S2) and the variator output rotation Vout ( That is, the rotation of the first sun gear S1 is different from that of the first sun gear S1, and the variator output rotation Vout is increased to become the output rotation OutH (see FIG. 2). Furthermore, the above-mentioned ratio of the number of teeth (Z _S1 / Z _P1 ) × (Z _P2 / Z _S2 ) is configured to be the same so that the output rotation and the variator output rotation are the same. May be.

  Next, autonomous shift control in the variator 5 will be described with reference to FIGS.

  For example, in a vehicle equipped with a continuously variable transmission 1, as shown in FIG. 5, rotation from an engine (not shown) is transmitted, and the input disk 2A (2B) rotates at the variator input rotation Vin. When the input disk 2A (2B) is rotated by the variator input rotation Vin, the surface velocity vector at the contact portion Ei based on the contact radius of the power roller 4 becomes Vdi, and at the same time, the surface velocity vector of the power roller 4 at the contact portion Ei is Vri. And rotation is transmitted.

  Further, the surface velocity of the power roller 4 at the contact portion Eo between the power roller 4 and the output disk 3 based on the contact radius is the surface velocity vector Vro having the same magnitude as the opposite direction to the surface velocity vector Vri. The surface velocity vector of the output disk 3 at the portion Eo is Vdo. That is, the output disk 3 rotates at the variator output rotation Vout according to the surface velocity vector Vdo at the contact portion Eo based on the contact radius.

  On the other hand, in the input disk 2A (2B), a force Tdi based on the engine torque TE transmitted from the engine (not shown) to the input disk 2A (2B) and the contact radius at the contact portion Ei acts. At this time, a traction force Ftri is applied to the power roller 4 according to the force Tdi when torque is transmitted from the input disk 2A (2B) to the power roller 4.

  Similarly, when torque is transmitted from the power roller 4 to the output disk 3, the force Tdo acts on the output disk 3 in accordance with the traction force Ftri at the contact portion Eo. Thus, the variator output torque Tout is transmitted to the output disk 3 as the rotational torque based on the contact radius according to the force Tdo acting on the contact portion Eo.

  When the force Tdo acts on the output disk 3 from the power roller 4, the traction force Ftro acts on the power roller 4 as a reaction force of the force Tdo. The combined force of the traction forces Ftri and Ftro acting on the power roller 4 is connected to the hydraulic actuator 40 while rotatably supporting the power roller 4 as described above, particularly when the power roller 4 is not moved. The reaction force Fp acting on the carriage 41 is balanced.

  Here, for example, when the reaction force Fp is increased by the control of the hydraulic actuator 40, the reaction force Fp and the combined force of the traction forces Ftri and Ftro are not balanced, and the power roller 4 moves in the Y1 direction. Ei and Eo also move. That is, as shown in FIG. 6, the power roller 4 moves from the position x in the drawing to the position y. Then, the surface velocity vector Vdi at the contact portion Ei was parallel in the same direction as the surface velocity vector Vri at the position x in the drawing, but at the position y in the drawing, the surface velocity vector Vdi is the input disk at the contact portion Ei. The tangential direction of 2A (2B) is not parallel to the surface velocity vector Vri, and the difference ΔV between the surface velocity vectors Vdi and Vri is also not parallel.

  Furthermore, since the traction force acting on the power roller 4 is the same as the direction of the difference ΔV between the surface velocity vectors Vdi and Vri, the traction force acting on the power roller 4 in the state indicated by the y position in FIG. Become. As shown in FIG. 7, the traction force Fti has a force Ftti having a direction component perpendicular to the direction component of the traction force Ftri in addition to the direction component of the traction force Ftri.

  On the other hand, at the contact portion Eo between the power roller 4 and the output disk 3, the power roller 4 has a velocity vector in the direction opposite to the surface velocity vector Vdi and is transmitted (pressed) from the power roller 4 (in this case, the outer periphery). The traction force Fto acts as a reaction force of a force including a directional component pressing toward the side), that is, the traction force Fto is a force Ftto having a direction component perpendicular to the direction component of the traction force Ftro and the traction force Ftro. have. By the action of these forces Ftti and Ftto, the center axis d of the power roller 4 is tilted, that is, the gear ratio (contact radius) between the input disk 2A (2B) and the output disk 3 changes autonomously.

  Then, due to the effect of the caster angle β (see FIG. 5), the power roller 4 moves while returning autonomously in the direction along the rotation direction as indicated by the position z in FIG. 6, and the surface velocity vectors Vdi, Vr Become parallel again, and the direction component of the traction force is in the same direction as Ftri and Ftro, so that the power roller 4 is stabilized.

  Thus, in the variator 5, the traction force Ftri + Ftro generated at the contact portions Ei and Eo between the input disk 2 </ b> A (2 </ b> B) and the output disk 3 and the power roller 4, and the carriage 41 that rotatably supports the power roller 4. By changing the reaction force Fp acting in this manner, a speed difference ΔV between the input disk 2A (2B) and the output disk 3 and the power roller 4 is generated, and traction force components Ftti, Ftto generated by the speed difference ΔV are generated. Thus, the shift is performed autonomously.

  Next, the hydraulic control device A of the continuously variable transmission 1 will be described with reference to FIG. FIG. 8 is a schematic diagram showing a part of the variator 5 and a part of the hydraulic control device A of the continuously variable transmission 1 according to the present invention. FIG. 8 is a schematic diagram showing the variator 5 and the hydraulic control device, and shows only necessary elements for explaining the present invention, and the actual hydraulic circuit is more complicated and has many elements. It is.

  The hydraulic control apparatus A according to the present invention includes the oil pump (hydraulic pressure generating source) 110, a primary regulator valve (original pressure regulating valve) 111, the carriage 41, the spherical bearing 42, the lever 45, the piston rod 46, the hydraulic actuator 40A, 40B, a roller driving device B that drives the power roller 4, a first reaction pressure control valve (operating pressure regulating valve) 112 that regulates a servo pressure (operating pressure) PS1 supplied to the hydraulic actuators 40A and 40B, A second reaction pressure control valve (operating pressure regulating valve) 113 for regulating the servo pressure PS2 supplied to the hydraulic actuators 40A, 40B, a first end stop valve (detour communication valve) 115 for the servo pressure PS1, a servo pressure PS2 Second end stop valve for Road communication valve) 116, clutch pressure control valve 120 for low clutch, clutch pressure control valve 121 for high clutch, pilot shuttle check valves 131, 132, 133, ball shuttle valves 138, 139, and bypass check valves 135, 136 And so on. The first end stop valve 115 and the second end stop valve 116 are configured to include end stop levers (out-of-range detection means) 48A, 48B for mechanically controlling the first end stop valve 115 and the second end stop valve 116.

  The primary regulator valve 111 includes a spool 111e, a spring 111d that urges the spool 111e in parallel with a signal pressure PS input to the oil chamber 111c via an oil passage f7, which will be described in detail later, and a pressure of the port 111a. And an oil chamber 111f that urges the spool 111e to face the signal pressure PS. In the primary regulator valve 111, the position of the spool 111e is adjusted by the biasing force of the spring 111d, the signal pressure PS input to the oil chamber 111c via the oil passage f7, and the oil pressure of the oil passage a7 input to the oil chamber 111f. By discharging a part of the hydraulic pressure from the oil pump 110 input to the port 111a via the oil passage a7 from the port 111b, the oil passage a7 to the oil passages a1, a2, a3, a4, a5 The oil pressure of a6 is adjusted as the line pressure (original pressure). The hydraulic pressure discharged from the port 111b is returned to the oil pump 110 via the oil passage a8 (for example, via a secondary regulator valve not shown), and energy loss in the oil pump 110 is reduced.

  The hydraulic control device A includes a secondary regulator valve (not shown) that inputs, for example, the pressure discharged from the port 111b of the primary regulator valve 111 and regulates the pressure to a substantially constant secondary pressure. The secondary pressure is supplied as a lubricating pressure Plube to a lubricating oil passage for supplying lubricating oil to the transmission mechanism U, the variator 5 and the like not shown. Further, the hydraulic control apparatus A is provided with a modulator valve (not shown) that inputs the line pressure PL and adjusts the pressure to a substantially constant modulator pressure. A linear solenoid for adjusting the servo pressure PS1 (not shown). The original pressure is supplied to the valve, the linear solenoid valve for adjusting servo pressure PS2, the linear solenoid valve for adjusting low clutch engagement pressure PSCL, and the linear solenoid valve for adjusting high clutch engagement pressure PSCH. These linear solenoid valves (not shown) linearly adjust the modulator pressure on the basis of an electric signal from a control unit (ECU) (not shown), and signal pressure PSS1, signal pressure PSS2, signal pressure PSCL, and signal pressure PSCH, respectively. Is output.

  The low clutch clutch pressure control valve 120 has a spool 120f and a spring (not shown) that urges the spool 120f in one direction, and the line pressure PL is input via an oil passage a5. Port 120c, a port 120a connected to the hydraulic servo 101 of the low clutch L through the oil passage g1, a drain port 120b, and a signal pressure PSCL input through the oil passage h3 are applied to the spool 120f. There is an oil chamber 120d that acts in a direction that opposes the biasing force of the spring, and a feedback oil chamber 120e in which the hydraulic pressure output from the port 120a acts as a feedback pressure in the biasing direction of the spring via the oil path i3. Configured.

  The clutch pressure control valve 120 for the low clutch resists the urging force of the spring while being feedback controlled by the feedback pressure input to the oil chamber 120e as the signal pressure PSCL input to the oil chamber 120d increases. Thus, the position of the spool 120f is driven to move toward the right half position. Accordingly, the opening amount of the port 120c and the port 120a is increased and the opening amount of the drain port 120b is decreased (or closed), that is, the port 120a is based on the line pressure PL input to the port 120c. Pressure adjustment output so that the clutch pressure PCL for low clutch output to the oil passage g1 increases. Further, when the signal pressure PSCL is not output (or reduced), the position of the spool 120f is driven to move toward the left half position by the biasing force of the spring. As a result, the opening amount of the port 120c and the port 120a is reduced (or closed) and the opening amount of the drain port 120b is increased, that is, the low clutch clutch pressure PCL is discharged from the drain port 120b.

  Similarly, the clutch pressure control valve 121 for the high clutch has a spool 121f and a spring (not shown) that urges the spool 121f in one direction, and the line pressure via the oil passage a6. The port 121c to which the PL is input, the port 121a connected to the hydraulic servo 105 of the high clutch H through the oil passage g2, the drain port 121b, and the signal pressure PSCH input through the oil passage h4 are connected to the spool. An oil chamber 121d acting in a direction opposite to the biasing force of the spring with respect to 121f, and a feedback oil chamber in which the hydraulic pressure output from the port 121a acts as a feedback pressure in the biasing direction of the spring via the oil path i4 121e.

  The clutch pressure control valve 121 for the high clutch resists the urging force of the spring while being feedback-controlled by the feedback pressure input to the oil chamber 121e as the signal pressure PSCH input to the oil chamber 121d increases. Thus, the position of the spool 121f is driven to move toward the right half position. As a result, the opening amount of the port 121c and the port 121a is increased and the opening amount of the drain port 121b is decreased (or closed), that is, the port 121a is based on the line pressure PL input to the port 121c. Pressure adjustment output so that the clutch pressure PCH for the high clutch output to the oil passage g2 increases. When the signal pressure PSCH is not output (or reduced), the position of the spool 121f is driven to move toward the left half position by the biasing force of the spring. As a result, the opening amount of the port 121c and the port 121a is reduced (or closed) and the opening amount of the drain port 121b is increased, that is, the high clutch clutch pressure PCH is discharged from the drain port 121b.

  On the other hand, the first reaction pressure control valve 112 has a spool 112f and a spring (not shown) that urges the spool 112f in one direction, and the line pressure PL is input via an oil passage a3. Port 112c, a port 112a connected to the ball shuttle valve 139 via an oil passage b3, a drain port 112b, and a signal pressure PSS1 inputted via an oil passage h1 is applied to the spring 112f against the spool 112f. An oil chamber 112d that acts in a direction opposite to the urging force, and a feedback oil chamber 112e in which the hydraulic pressure output from the port 112a acts as a feedback pressure in the urging direction of the spring via the oil path i1. It is configured.

  As the signal pressure PSS1 input to the oil chamber 112d increases, the first reaction pressure control valve 112 is feedback-controlled by the feedback pressure input to the oil chamber 112e and resists the biasing force of the spring. The position of the spool 112f is driven to move toward the right half position. Thereby, the opening amount of the port 112c and the port 112a is increased and the opening amount of the drain port 112b is decreased (or closed), that is, the port 112a is based on the line pressure PL input to the port 112c. Pressure is adjusted so as to increase the servo pressure PS1 output to the oil passage b3. When the signal pressure PSS1 is not output (or reduced), the position of the spool 112f is driven to move toward the left half position by the biasing force of the spring. As a result, the opening amount of the port 112c and the port 112a is reduced (or closed) and the opening amount of the drain port 112b is increased, that is, the servo pressure PS1 is discharged from the drain port 112b to the lubricating oil passage.

  Similarly, the second reaction pressure control valve 113 includes a spool 113f and a spring (not shown) that urges the spool 113f in one direction, and the line pressure PL via an oil passage a2. 113, the port 113a connected to the ball shuttle valve 138 via the oil passage c3, the drain port 113b, and the signal pressure PSS2 inputted via the oil passage h2 are applied to the spool 113f. There is an oil chamber 113d that acts in a direction opposite to the urging force of the spring, and a feedback oil chamber 113e in which the hydraulic pressure output from the port 113a acts as a feedback pressure in the urging direction of the spring via the oil path i2. Configured.

  As the signal pressure PSS2 input to the oil chamber 113d increases, the second reaction pressure control valve 113 is feedback-controlled by the feedback pressure input to the oil chamber 113e and resists the biasing force of the spring. The position of the spool 113f is driven to move toward the left half position. As a result, the opening amount of the port 113c and the port 113a is increased and the opening amount of the drain port 113b is decreased (or closed), that is, the port 113a is based on the line pressure PL input to the port 113c. Pressure is adjusted so that the servo pressure PS2 output to the oil passage c3 becomes larger. When the signal pressure PSS2 is not output (or reduced), the position of the spool 113f is driven to move toward the right half position by the biasing force of the spring. As a result, the opening amount of the port 113c and the port 113a is reduced (or closed) and the opening amount of the drain port 113b is increased. That is, the servo pressure PS2 is discharged from the drain port 113b to the lubricating oil passage.

  The hydraulic actuator 40A includes a cylinder 40As and a piston 40Ap that is slidably contained in the cylinder 40As and connected to the piston rod 46A. Oil chambers 40Aa and 40Ab that are opposed to each other are formed, thereby constituting two hydraulic servos (power roller control hydraulic servos) 40A1 and 40A2 that act in opposite directions.

  The oil chamber 40Aa of the hydraulic servo 40A1 is connected to the port 113a of the second reaction pressure control valve 113 through oil passages c1, c4, c3, and the oil chamber 40Ab of the hydraulic servo 40A2 The first reaction pressure control valve 112 is connected to the port 112a via the paths b1, b4, and b3. That is, the servo pressure PS2 output from the port 113a acts on the oil chamber 40Aa, and the servo pressure PS1 output from the port 112a acts on the oil chamber 40Ab.

  When the servo pressure PS1 is adjusted larger than the servo pressure PS2 (or when the servo pressure PS2 is drained so as to be smaller than the servo pressure PS1), the oil pressure in the oil chamber 40Ab is higher than the oil pressure in the oil chamber 40Aa. The piston 40Ap is pushed upward in the figure, and the lever 45A rotating around the support shaft 43 via the piston rod 46A is pushed up in the figure to drive the carriage 41 and the spherical bearing 42. Then, the power roller 4a is driven in the ω1 direction.

  On the other hand, when the servo pressure PS2 is regulated to be larger than the servo pressure PS1 (or when the servo pressure PS1 is drained so as to be smaller than the servo pressure PS2), the oil pressure of the oil chamber 40Aa is changed to the oil chamber 40Ab. The piston 40Ap is pressed and driven downward in the drawing, and the lever 45A that rotates about the support shaft 43 via the piston rod 46A is pulled and driven downward in the drawing, and the carriage 41 and The power roller 4a is driven in the ω2 direction via the spherical bearing 42.

  Similarly to the hydraulic actuator 40A, the hydraulic actuator 40B includes a cylinder 40Bs and a piston 40Bp slidably contained in the cylinder 40Bs and connected to the piston rod 46B. The cylinder 40Bs and the piston 40Bp forms oil chambers 40Ba and 40Bb opposed to each other through the piston 40Bp, thereby constituting two hydraulic servos 40B1 and 40B2 that act in opposite directions.

  The oil chamber 40Ba of the hydraulic servo 40B1 is connected to the port 112a of the first reaction pressure control valve 112 via oil passages b2, b4, b3, and the oil chamber 40Bb of the hydraulic servo 40B2 The second reaction pressure control valve 113 is connected to the port 113a via the paths c2, c4, and c3. That is, the servo pressure PS1 output from the port 112a acts on the oil chamber 40Ba, and the servo pressure PS2 output from the port 113a acts on the oil chamber 40Bb.

  When the servo pressure PS2 is adjusted higher than the servo pressure PS1 (or when the servo pressure PS1 is drained so as to be smaller than the servo pressure PS2), the oil pressure in the oil chamber 40Bb is higher than the oil pressure in the oil chamber 40Ba. The piston 40Bp is pushed upward in the figure, and the lever 45B that rotates about the support shaft 43 via the piston rod 46B is pushed up in the figure to drive the carriage 41 and the spherical bearing 42. Then, the power roller 4a is driven in the ω2 direction.

  On the other hand, when the servo pressure PS1 is adjusted to be larger than the servo pressure PS2 (or when the servo pressure PS2 is drained so as to be smaller than the servo pressure PS1), the oil pressure in the oil chamber 40Ba is changed to the oil chamber 40Bb. The piston 40Bp is pressed and driven downward in the drawing, and the lever 45B that rotates about the support shaft 43 via the piston rod 46B is pulled and driven downward in the drawing, and the carriage 41 and The power roller 4a is driven in the ω1 direction via the spherical bearing 42.

  In FIG. 8, two hydraulic actuators 40A and 40B that drive and control the two power rollers 4a and 4c are illustrated for the sake of space, but in reality, other similar hydraulic actuators are used. In other words, all the six power rollers 4 are driven and controlled by all six hydraulic actuators that are driven by the servo pressures PS1 and PS2 supplied by the two reaction pressure control valves 112 and 113. Yes.

  The end stop lever 48A has an end portion 48Aa disposed opposite to an end portion 45Aa of a lever (lever member) 45A, and the end portion 48Ac opposes an impact absorbing spring 115d of a first end stop valve 115 described later in detail. It is arrange | positioned and is arrange | positioned so that rotation is possible by using the support part 48Ab fixed to the case 20 as a fulcrum. When the power roller 4a exceeds a predetermined moving range, the end 45Aa of the lever 45A presses the end 48Aa of the end stop lever 48A, and the opposite end 48Ac presses the spring 115d of the first end stop valve 115. Is forming.

  The first end stop valve 115 includes a spool 115f, a spring 115g that urges the spool 115f to the right side in the drawing, and a shock absorbing spring 115d. When the spool 115f is driven to the left side in the drawing by the portion 48Ac via the spring 115d and the pressing force of the end stop lever 48A exceeds the urging force of the spring 115g, the position of the spool 115f is the right half in the drawing. Switched to position. When the spool 115f is switched to the right half position in the figure, the port 115a and the port 115b are communicated, and the line pressure PL input to the port 115b via the oil passage a4 is output from the port 115a to the oil passage d1. Further, the action of the line pressure PL from the port 115a input to the oil chamber 115e via the oil passage i5 and the urging force of the spring 115g are not pressed by the end stop lever 48A or the urging force of the spring 115g. When the pressing force is exceeded, the position of the spool 115f is switched to the left half position in the figure. When the spool 115f is switched to the left half position in the figure, the port 115a and the drain port 115c are communicated, and the oil pressure in the oil passage d1 is discharged from the drain port 115c.

  Similarly, the end stop lever 48B is disposed so that the end 48Ba faces the end 45Ba of the lever 45B, and the end 48Bc opposes an impact absorbing spring 116d of the second end stop valve 116 described later in detail. It is a member that is rotatably arranged with a support portion 48Bb fixed to the case 20 as a fulcrum. When the power roller 4c exceeds a predetermined movement range, the end portion 45Ba of the lever 45B is moved to the end stop lever 48B. The end portion 48Ba of the second end stop valve 116 is pressed by the opposite end portion 48Bc to form a mechanism.

  The second end stop valve 116 includes a spool 116f, a spring 116g that urges the spool 116f to the left in the drawing, and a shock absorbing spring 116d. When the spool 116f is driven to the right side in the drawing by the portion 48Bc via the spring 116d and the pressing force of the end stop lever 48B exceeds the urging force of the spring 116g, the position of the spool 116f is the left half in the drawing. Switched to position. When the spool 116f is switched to the left half position in the figure, the port 116a and the port 116b are communicated, and the line pressure PL input to the port 116b via the oil passage a1 is output from the port 116a to the oil passage d2. Further, the action of the line pressure PL from the port 116a input to the oil chamber 116e via the oil passage i6 and the biasing force of the spring 116g are not pressed by the end stop lever 48B or the biasing force of the spring 116g. When the pressing force is exceeded, the position of the spool 116f is switched to the right half position in the figure. When the spool 116f is switched to the right half position in the figure, the port 116a communicates with the drain port 116c, and the oil pressure in the oil passage d2 is discharged from the drain port 116c.

  The pilot shuttle check valve 131 has a check ball 131b that receives the servo pressure PS2 from the oil passage f1 and the servo pressure PS1 from the oil passage f2, and receives the servo pressure PS2 and the oil passage f2 in the oil passage f1. The check ball 131b is moved by the higher pressure of the servo pressure PS1 and communicates with the higher pressure of the oil passage f1 and the oil passage f2 and the oil passage f3. That is, the pilot shuttle check valve 131 outputs the higher hydraulic pressure of the servo pressure PS1 and the servo pressure PS2 to the oil passage f3.

  The pilot shuttle check valve 132 has a check ball 132b that receives the low clutch clutch pressure PCL from the oil passage f5 and the high clutch clutch pressure PCH from the oil passage f6 so as to face each other. The check ball 132b is moved by the higher pressure of the low clutch clutch pressure PCL and the high clutch clutch pressure PCH of the oil passage f6, and the higher pressure of the oil passage f5 and the oil passage f6 communicates with the oil passage f4. To do. That is, the pilot shuttle check valve 132 outputs the higher hydraulic pressure of the low clutch clutch pressure PCL and the high clutch clutch pressure PCH to the oil passage f4.

  Further, the pilot shuttle check valve 133 has a check ball 133b that receives the oil pressure from the oil passage f3 and the oil pressure from the oil passage f4 in opposition to each other. The check ball 133b is moved by the pressure higher than the hydraulic pressure from the path f4, and the oil path f3 and the higher pressure in the oil path f4 communicate with the oil path f7. That is, the pilot shuttle check valve 133 outputs the higher hydraulic pressure of the oil passage f3 and the oil passage f4 to the oil passage f7.

  The ball shuttle valve 138 has a check ball 138b that receives the oil pressure from the oil passage d2 and the servo pressure PS2 from the oil passage c3 so as to face each other, and the oil pressure from the oil passage d2 and the servo pressure from the oil passage c3. The check ball 138b is moved by the higher pressure with PS2, and communicates with the higher pressure in the oil passage d2 and the oil passage c3 and the oil passage c4. That is, the ball shuttle valve 138 outputs the higher hydraulic pressure of the oil passage d2 and the servo pressure PS2 to the oil passage c4.

  Similarly, the ball shuttle valve 139 has a check ball 139b that receives the oil pressure from the oil passage d1 and the servo pressure PS1 from the oil passage b3 in opposition to each other, and the oil pressure from the oil passage d1 and the oil passage b3. The check ball 139b is moved by the higher pressure from the servo pressure PS1 to communicate with the higher pressure in the oil passages d1 and b3 and the oil passage b4. That is, the ball shuttle valve 139 outputs the higher hydraulic pressure of the oil passage d1 and the servo pressure PS1 to the oil passage b4.

  The bypass check valve 135 has a check ball 135b that receives the servo pressure PS2 from the oil passage c4 and the lubrication pressure Plube from the oil passage e so that the servo pressure PS2 is output to the normal oil passage c4. However, when the lubrication pressure Plube becomes higher than the servo pressure PS2, the check ball 135b is moved, and the oil passage c4 and the oil passage e communicate with each other. That is, the bypass check valve 135 outputs the lubricating pressure Plube to the oil passage c4 when the lubricating pressure Plube becomes higher than the servo pressure PS2.

  Similarly, the bypass check valve 136 has a check ball 136b that receives the servo pressure PS1 from the oil passage b4 and the lubrication pressure Plube from the oil passage e so as to face each other, and the normal oil passage b4 has the servo pressure PS1. Although it is output, when the lubricating pressure Plube becomes higher than the servo pressure PS1, the check ball 136b is moved, and the oil passage b4 and the oil passage e are communicated. That is, the bypass check valve 136 outputs the lubricating pressure Plube to the oil passage b4 when the lubricating pressure Plube becomes higher than the servo pressure PS1.

  Next, the operation of the continuously variable transmission 1 based on the hydraulic circuit will be described with reference to FIG. For example, when the vehicle equipped with the continuously variable transmission 1 starts or reverses, a low mode state is set by operating a shift lever (not shown) to the drive (D) range. That is, the line pressure PL input to the low clutch clutch pressure control valve 120 is adjusted to the low clutch clutch pressure PCL and supplied to the hydraulic servo 101 of the low clutch L to engage the low clutch L. The clutch pressure PCH is not supplied from the high clutch clutch pressure control valve 121 to the hydraulic servo 105 of the high clutch H, and the high clutch H is released. Further, when the speed increases and the speed becomes relatively high, and the shift is determined, the supply of the clutch pressure PCL from the low clutch clutch pressure control valve 120 to the hydraulic servo 101 of the low clutch L is stopped, and the drain port 120b. Accordingly, the clutch pressure PCL is discharged, the low clutch L is released, and the line pressure PL input to the high clutch clutch pressure control valve 121 is adjusted to the high clutch clutch pressure PCH. The high clutch H is supplied to the hydraulic servo 105 and the high clutch H is engaged, and the high mode is set.

  On the other hand, in the continuously variable transmission 1, when a shift lever (not shown) is operated and the drive (D) range is set, the low-mode state is set. The servo pressure PS1 is adjusted to the servo pressure PS1 based on the signal pressure PSS1, and the servo pressure PS1 is supplied to the oil chamber 40Ab of the hydraulic servo 40A2 through the oil passages b3, b4, b1, and the oil passages b3, b4, b2 To the oil chamber 40Ba of the hydraulic servo 40B1. Further, in the second reaction pressure control valve 113, the line pressure PL is similarly adjusted to the servo pressure PS2 based on the signal pressure PSS2 from the linear solenoid valve, and the servo pressure PS2 passes through the oil passages c3, c4, c1. Is supplied to the oil chamber 40Aa of the hydraulic servo 40A1, and is also supplied to the oil chamber 40Bb of the hydraulic servo 40B2 through the oil passages c3, c4, and c2. Thereby, the power rollers 4a and 4c are hydraulically controlled by the hydraulic actuators 40A and 40B.

  Further, when the output disk 3 rotates in the ω1 direction in the driving state of the variator 5 (the input disk 2A rotates in the ω2 direction), the reaction force acts on the power roller 4a as described above, and the hydraulic pressure In the actuator 40A, the servo pressure PS1 in the oil chamber 40Ab is controlled in a state larger than the servo pressure PS2 in the oil chamber 40Aa.

  At this time, when shifting the variator 5 to the speed increasing side, as described above, the linear solenoid valve for adjusting the servo pressure PS1 (not shown) is a signal based on an electric signal from a control unit (ECU) (not shown). The pressure PSS1 is output, and the first reaction pressure control valve 112 outputs the servo pressure PS1 based on the signal pressure PSS1. When the servo pressure PS1 is supplied to the oil chamber 40Ab of the hydraulic servo 40A2, the piston 40Ap and the piston rod 46A of the hydraulic actuator 40A are pressed upward in the figure, and the lever 45A rotates counterclockwise around the support shaft 43. The end portion 45Aa moves the power roller 4a to the ω1 direction side through the spherical bearing 42 and the carriage 41.

  Further, the servo pressure PS1 output from the first reaction pressure control valve 112 based on the signal pressure PSS1 is also supplied to the oil chamber 40Ba of the hydraulic servo 40B1, and the piston 40Bp and the piston rod 46B of the hydraulic actuator 40B are moved downward in the figure. The power roller 4c is also moved to the ω1 direction side via the lever 45B, the spherical bearing 42, and the carriage 41. As a result, as described above, the power roller 4a (4c) has the input disk 2A side (front side of the paper surface) on the outer peripheral side of the input disk 2A, and the output disk 3 side (back side of the paper surface) is the center of the output disk 3. The output disc 3 is accelerated with respect to the input disc 2A.

  When shifting the variator 5 to the deceleration side, a linear solenoid valve for adjusting servo pressure PS2 (not shown) outputs a signal pressure PSS2 based on an electric signal from a control unit (ECU) not shown, Based on the signal pressure PSS2, the second reaction pressure control valve 113 outputs the servo pressure PS2. When the servo pressure PS2 is supplied to the oil chamber 40Aa of the hydraulic servo 40A1, the piston 40Ap and the piston rod 46A of the hydraulic actuator 40A are pressed downward in the figure, and the lever 45A rotates clockwise around the support shaft 43. The end portion 45Aa moves the power roller 4a to the ω2 direction side through the spherical bearing 42 and the carriage 41.

  The servo pressure PS2 output from the second reaction pressure control valve 113 based on the signal pressure PSS2 is also supplied to the oil chamber 40Bb of the hydraulic servo 40B2, and the piston 40Bp and the piston rod 46B of the hydraulic actuator 40B are moved upward in the figure. The power roller 4c is also moved to the ω2 direction side through the lever 45B, the spherical bearing 42, and the carriage 41. As a result, as described above, the power roller 4a (4c) has the input disk 2A side (front side of the paper) at the center side of the input disk 2A, and the output disk 3 side (back side of the paper) at the outer periphery of the output disk 3. The output disk 3 is decelerated with respect to the input disk 2A.

  In the continuously variable transmission 1, when a change occurs in the relative rotation (acceleration) of the input disks 2 </ b> A and 2 </ b> B and the output disk 3 of the variator 5, the balance of the balance between the traction force acting on the power roller 4 and the reaction force. Collapses. The hydraulic control in this case will be described.

  For example, when a force for decelerating the output disk 3 is applied, such as when moving from a flat road to an uphill road, the traction force acting on the power roller 4 increases, and the traction force exceeds the reaction force.

  That is, even in the power roller 4a, the traction force exceeds the reaction force, and the power roller 4a starts to move toward the ω2 direction side of the output disk 3. Then, the power roller 4a acts on the end 45Aa of the lever 45A via the carriage 41 and the spherical bearing 42, and the lever 45A rotates in the clockwise direction around the support shaft 43 and is connected to the lever 45A. The piston 40Ap is moved downward in the figure through the piston rod 46A.

  Similarly, the traction force of the power roller 4c exceeds the reaction force, and the power roller 4c starts to move toward the ω2 direction side of the output disk 3. The power roller 4c acts on the end 45Ba of the lever 45B via the carriage 41 and the spherical bearing 42. The lever 45B rotates in the clockwise direction around the support shaft 43 and is connected to the lever 45B. The piston 40Bp is moved upward in the figure via the piston rod 46B.

  Then, the servos of the oil chambers 40Aa and 40Bb on the side where the oil chambers of the hydraulic actuators 40A and 40B expand (the side to which the servo pressure PS2 is supplied) and the oil passages c1, c2 and c4 connected to the oil chambers 40Aa and 40Bb are provided. The pressure PS2 is reduced. At this time, if the servo pressure PS2 of the oil passages c1, c2, c4 connected to the oil chambers 40Aa, 40Bb and the oil chambers 40Aa, 40Bb is reduced to be smaller than the lubrication pressure Plube, the bypass is performed. In the check valve 135, the check ball 135b is moved as described above, the oil passage c4 and the oil passage e communicate with each other, and the lubrication pressure Plube is supplied to the oil chambers 40Aa and 40Bb through the oil passages e, c4, c1, and c2. Is done.

  At the same time, the oil chambers 40Ab, 40Ba on the side where the oil chambers of the hydraulic actuators 40A, 40B are narrowed (the side to which the servo pressure PS1 is supplied) and the oil passages b1, b2, b4 connected to the oil chambers 40Ab, 40Ba Servo pressure PS1 is increased. At this time, the oil pushed out from the oil chambers 40Ab and 40Ba is input to the port 112a of the first reaction pressure control valve 112 via the oil passages b1, b2 and b4 and at the same time the first reaction pressure via the oil passage i1. It is also input to the feedback oil chamber 112e of the control valve 112. In the first reaction pressure control valve 112, the action of the oil input to the feedback oil chamber 112e and the urging force of the spring are driven to move the position of the spool 112f toward the left half position against the signal pressure PSS1. . As a result, the port 112a and the port 112b communicate with each other, and the oil pushed out from the oil chambers 40Ab and 40Ba is discharged to the oil passage e.

  On the other hand, when a force that accelerates the output disk 3 is applied, such as when moving from a flat road to a downhill road, the traction force acting on the power roller 4 is weakened and the reaction force exceeds the traction force.

  That is, even in the power roller 4c, the reaction force exceeds the traction force, and the power roller 4c starts to move toward the ω1 direction side of the output disk 3. Then, the power roller 4c acts on the end 45Ba of the lever 45B via the carriage 41 and the spherical bearing 42, and the lever 45B rotates in the counterclockwise direction around the support shaft 43 and is connected to the lever 45B. The piston 40Bp is moved downward in the figure through the piston rod 46B.

  Similarly, the reaction force of the power roller 4 a exceeds the traction force, and the power roller 4 a starts to move toward the ω1 direction side of the output disk 3. The power roller 4a acts on the end portion 45Aa of the lever 45A via the carriage 41 and the spherical bearing 42. The lever 45A rotates in the counterclockwise direction around the support shaft 43 and is connected to the lever 45A. The piston 40Ap is moved upward in the figure via the piston rod 46A.

  Then, the servos of the oil passages b1, b2, b4 connected to the oil chambers 40Ab, 40Ba on the side where the oil chambers of the hydraulic actuators 40A, 40B expand (the side to which the servo pressure PS1 is supplied) and the oil chambers 40Ab, 40Ba are connected. The pressure PS1 is reduced. At this time, when the servo pressure PS1 of the oil passages b1, b2, b4 connected to the oil chambers 40Ab, 40Ba and the oil chambers 40Ab, 40Ba is reduced to be smaller than the lubrication pressure Plube, In the check valve 136, as described above, the check ball 136b is moved so that the oil passage b4 and the oil passage e communicate with each other, and the lubrication pressure Plube is supplied to the oil chambers 40Ab and 40Ba via the oil passages e, b4, b1, and b2. Is done.

  At the same time, the oil chambers 40Aa, 40Bb on the side where the oil chambers of the hydraulic actuators 40A, 40B are narrowed (the side to which the servo pressure PS2 is supplied) and the oil passages c1, c2, c4 connected to the oil chambers 40Aa, 40Bb Servo pressure PS2 is increased. At this time, the oil pushed out from the oil chambers 40Aa and 40Bb is input to the port 113a of the second reaction pressure control valve 113 via the oil passages c1, c2 and c4, and at the same time the second reaction pressure via the oil passage i2. It is also input to the feedback oil chamber 113e of the control valve 113. In the second reaction pressure control valve 113, the action of oil input to the feedback oil chamber 113e and the biasing force of the spring are driven to move the position of the spool 113f toward the right half position against the signal pressure PSS2. . As a result, the port 113a and the port 113b communicate with each other, and the oil pushed out from the oil chambers 40Aa and 40Bb is discharged to the oil passage e.

  As described above, in the continuously variable transmission 1, when the expanding oil chamber and the servo pressure PS1 or PS2 in the supply oil passage become lower than the lubricating pressure Plube, the expanding oil chamber and oil It is possible to prevent a negative pressure from being generated by supplying the lubrication pressure Plube to the road. Further, since the oil pushed out from the narrowing oil chamber is discharged to the oil passage e through the first or second reaction pressure control valves 112 and 113, the bypass check valve 135 or the bypass check valve is discharged from the oil passage e. It is possible to circulate the hydraulic oil to the oil chamber on the expanding side via 136, that is, to immediately increase the operating pressure on the expanding side, and to improve the hydraulic response. As a result, the continuously variable transmission 1 does not require a bypass circuit, a bypass valve control device, or the like, and is inexpensive and can be made compact.

  Furthermore, when a force larger than the state in which the force to be decelerated acts on the output disk 3 as described above is applied to the power roller 4, that is, when the output disk 3 is suddenly decelerated due to sudden braking by a foot brake, for example. The traction force acting on the power roller 4 may exceed the reaction force, and the power roller 4 may move greatly in the ω2 direction and exceed a predetermined movement range.

  At this time, the power roller 4a moves greatly in the ω2 direction as described above, and acts on the end 45Aa of the lever 45A via the carriage 41 and the spherical bearing 42, and the lever 45A rotates clockwise around the support shaft 43. Rotate in the direction. The lever 45A is rotated and moved so that the end 45Aa presses the end 48Aa of the end stop lever 48A, and the end stop lever 48A is the first end stop at the other end 48Ac via the fulcrum 48Ab. The spring 115d of the valve 115 is pressed, and the spool 115f is driven to move to the right half position in the figure. Thereby, in the first end stop valve 115, the ports 115a and 115b communicate with each other, and the line pressure PL input to the port 115b via the oil passage a4 is output from the port 115a to the oil passage d1. The line pressure PL output to the oil passage d1 is input to the ball shuttle valve 139, and the line pressure PL input to the ball shuttle valve 139 is higher than the servo pressure PS1 of the oil passage b3. 139b is moved, the oil passage d1 and the oil passage b4 are communicated, and the line pressure PL is supplied to the oil chamber 40Ab of the hydraulic servo 40A2 and the oil chamber 40Ba of the hydraulic servo 40B1 through the oil passages b4, b1, and b2. The

  Further, the line pressure PL input to the oil chambers 40Ab and 40Ba presses the piston 40Ap upward in the drawing in the hydraulic servo 40A2, and the piston rod 46A, lever 45A, and spherical bearing 42 connected to the piston 40Ap. The power roller 4a is driven in the ω1 direction so as to move within a predetermined movement range via the carriage 41 and the like. Further, the line pressure PL presses the piston 40Bp downward in the figure in the hydraulic servo 40B1, and through the piston rod 46B, the lever 45B, the spherical bearing 42, the carriage 41, etc. connected to the piston 40Bp, The power roller 4c is pressed and driven in the ω1 direction so as to move within a predetermined movement range.

  When the power roller 4a is driven in the ω1 direction, the end portion 45Aa of the lever 45A is weakened in the force to press the end portion 48Aa of the end stop lever 48A as the end portion 45Aa moves, and finally the end portion 48Ac is separated from the spring 115d of the first end stop valve 115. Then, the spool 115f is switched rightward in the drawing by the action of the first end stop valve 115 from the oil chamber 115e and the urging force of the spring 115g, and the port 115b to which the line pressure PL is input is closed, 115a communicates with the drain port 115c to drain the oil pressure in the oil passages d1 and i5 and the oil chamber 115e.

  Thus, the first end stop valve 115 stops the supply of the line pressure PL to the oil chamber 40Ab of the hydraulic servo 40A2 and the oil chamber 40Ba of the hydraulic servo 40B1 via the oil passages d1, b4, b1, b2, and the oil passage d1. , I5 and the oil pressure in the oil chamber 115e are drained from the drain port 115c of the first end stop valve 115. When the oil pressure is drained and the oil pressure in the oil passage d1 becomes smaller than the servo pressure PS1 in the oil passage b3, the check ball 139b of the ball shuttle valve 139 is moved, and the oil passage b3 and the oil passage b4 are communicated again. The state is controlled by the servo pressure PS1.

  Even when a large force acts on the output disk 3 so that the reaction force acting on the power roller 4 exceeds the traction force, the power roller 4 may move greatly in the ω1 direction and exceed a predetermined moving range. is there.

  At this time, the power roller 4c moves greatly in the ω1 direction as described above, and acts on the end 45Ba of the lever 45B via the carriage 41 and the spherical bearing 42. The lever 45B rotates counterclockwise around the support shaft 43. Rotate in the direction. The lever 45B is largely rotated, so that the end 45Ba presses the end 48Ba of the end stop lever 48B, and the end stop lever 48B is the second end stop at the other end 48Bc via the fulcrum 48Bb. The spring 116d of the valve 116 is pressed, and the spool 116f is driven to move to the left half position in the figure. Thereby, in the second end stop valve 116, the port 116a and the port 116b communicate with each other, and the line pressure PL input to the port 116b through the oil passage a1 is output from the port 116a to the oil passage d2. The line pressure PL output to the oil passage d2 is input to the ball shuttle valve 138, and the line pressure PL input to the ball shuttle valve 138 is higher than the servo pressure PS2 of the oil passage c3. 138b is moved, the oil passage d2 and the oil passage c4 are communicated, and the line pressure PL is supplied to the oil chamber 40Bb of the hydraulic servo 40B2 and the oil chamber 40Aa of the hydraulic servo 40A1 through the oil passages c4, c1, and c2. The

  Further, the line pressure PL input to the oil chambers 40Bb and 40Aa presses the piston 40Bp upward in the drawing in the hydraulic servo 40B2, and the piston rod 46B, lever 45B, and spherical bearing 42 connected to the piston 40Bp. The power roller 4c is pressed and driven in the ω2 direction so as to move within a predetermined movement range via the carriage 41 and the like. Further, the line pressure PL presses the piston 40Ap downward in the figure in the hydraulic servo 40A1, and through the piston rod 46A, the lever 45A, the spherical bearing 42, the carriage 41, etc. connected to the piston 40Ap, The power roller 4a is pressed and driven in the ω2 direction so as to move within a predetermined movement range.

  Further, when the power roller 4c is driven to be pressed in the ω2 direction, the end 45Ba of the lever 45B weakens the force that presses the end 48Ba of the end stop lever 48B as the end 45Ba moves, and finally the end 45Ba 48 Bc is separated from the spring 116 d of the second end stop valve 116. Then, the spool 116f is switched to the left in the figure by the action of the second end stop valve 116 from the oil chamber 116e and the urging force of the spring 116g, and the port 116b to which the line pressure PL is input is closed, and the port 116a communicates with the drain port 116c to drain the oil pressure in the oil passages d2 and i6 and the oil chamber 116e.

  As a result, the second end stop valve 116 stops supplying the line pressure PL to the oil chamber 40Bb of the hydraulic servo 40B2 and the oil chamber 40Aa of the hydraulic servo 40A1 via the oil passages d2, c4, c1, and c2, and the oil passage d2. , I6 and the oil pressure of the oil chamber 116e are drained from the drain port 116c of the second end stop valve 116. When the oil pressure is drained and the oil pressure in the oil passage d2 becomes smaller than the servo pressure PS2 in the oil passage c3, the check ball 138b of the ball shuttle valve 138 is moved, and the oil passage c3 and the oil passage c4 are again communicated with each other. The state is controlled by the servo pressure PS2.

  In the above description, the two hydraulic actuators 40A and 40B that drive and control the two power rollers 4a and 4c are shown and described, but in reality, there are four other similar hydraulic actuators, That is, all the six power rollers 4 are driven and controlled by all six hydraulic actuators that are driven by the line pressure PL being supplied by the two end stop valves 115 and 116.

  Further, in the hydraulic control device A of the continuously variable transmission 1, the operating pressure is input to the primary regulator valve 111 as the signal pressure PS so that the line pressure PL becomes an appropriate pressure, and a line higher than the operating pressure by a predetermined pressure. It operates to be regulated to the pressure PL. That is, as shown in FIG. 8, servo pressure PS1, servo pressure PS2, low clutch clutch pressure PCL, and high clutch clutch pressure PCH are output. The servo pressure PS1 is output from the first reaction pressure control valve 112 and input to the pilot shuttle check valve 131 via the oil passages b4 and f2. The servo pressure PS2 is output from the second reaction pressure control valve 113 and the oil passage c4. , F1 to the pilot shuttle check valve 131.

  As described above, the higher operating pressure is output to the oil passage f3. Similarly, the low clutch clutch pressure PCL is output from the low clutch clutch pressure control valve 120 and input to the pilot shuttle check valve 132 via the oil passages g1 and f5, and the high clutch clutch pressure PCH is the high clutch clutch pressure. It is output from the control valve 121 and input to the pilot shuttle check valve 132 via the oil passages g2 and f6. As described above, the higher operating pressure is output to the oil passage f4.

  Furthermore, the higher servo pressure is input to the pilot shuttle check valve 133 via the oil passage f3, and the higher clutch pressure is input to the pilot shuttle check valve 133 via the oil passage f4. The higher operating pressure of the servo pressure and the clutch pressure is output to the oil passage f7, and is input as the signal pressure PS to the oil chamber 111c of the primary regulator valve 111 via the oil passage f7. As a result, the highest operating pressure among the servo pressure PS1, servo pressure PS2, low clutch clutch pressure PCL, and high clutch clutch pressure PCH is input to the primary regulator valve 111 as the signal pressure PS. The valve 111 outputs a line pressure PL adjusted so as to be higher than the signal pressure PS by a predetermined pressure.

  As described above, when the first end stop valve 115 or the second end stop valve 116 is operated (see FIG. 8) and the line pressure PL is output as the operating pressure PS1 or the operating pressure PS2, the oil passage is similarly used. Since the primary regulator valve 111 regulates the line pressure PL so as to be a predetermined pressure higher than the signal pressure PS, that is, input as the signal pressure PS via f1, f2, f3, f7, that is, the first end stop valve 115. Or when the 2nd end stop valve 116 act | operates, the line pressure PL can be raised rapidly. As a result, the power roller 4 exceeding the predetermined range can be immediately pushed back.

  Thus, it becomes possible to regulate the line pressure PL corresponding to all the operating pressures, and it becomes possible to eliminate the need to regulate the line pressure PL using, for example, an electromagnetic valve. It is possible to eliminate the pressure regulating solenoid valve.

  Further, two of the plurality of operating pressures are input oppositely to the check balls 131b, 132b, and 133b, and the pilot shuttle check valves 131 and 132 that conduct the larger operating pressure depending on the positions of the check balls 131b, 132b, and 133b. , 133 in a hierarchical manner, the maximum operating pressure among the plurality of operating pressures can be guided to the primary regulator valve 111.

  As described above, in the present invention, the first end stop valve 115 and the second end stop valve are detected when the end stop levers 48A and 48B detect that the rotation center d of the power roller 4 exceeds the predetermined movement range. 116 bypasses the first reaction pressure control valve 112 and the second reaction pressure control valve 113 and supplies the original pressure PL as the operating pressure of the hydraulic servos 40A1, 40A2, 40B1, and 40B2. When the pressure exceeds a predetermined movement range, the operating pressure of the hydraulic servos 40A1, 40A2, 40B1, and 40B2 can be rapidly increased. As a result, when the power roller 4 exceeds the predetermined movement range, the power roller 4 can be immediately pushed back into the predetermined movement range. The control can be immediately returned to the control under the press drive control of the hydraulic servos 40A1, 40A2, 40B1, and 40B2.

  The toroidal continuously variable transmission 1 includes a first reaction pressure control valve 112 and a second reaction pressure control valve 113. The hydraulic servos 40A1, 40A2, 40B1, and 40B2 are two oils facing the pistons 40Ap and 40Bp. Different operating pressures PS1 and PS2 are input to the chambers 40Aa, 40Ab, and 40Ba, 40Bb, and the rotation center d of the power roller 4 is moved through the carriage 41, the spherical bearing 42, the lever 45, and the piston rod 46 to both the disks 2A, 2B. , And the hydraulic servo for controlling the power roller that can be driven in the direction of the surface 3, so that the original pressure is applied to the pistons 40Ap, 40Bp in the direction in which the power roller 4 is pressed within the predetermined movement range. Working oil chambers 40Aa, 40Ab, and 40Ba It can be supplied to the 40Bb. Thus, while the power roller 4 can be immediately pushed back into the predetermined movement range, the shift control of the power roller 4 is controlled by the press drive control of the hydraulic servos 40A1, 40A2, 40B1, and 40B2. Can be restored immediately to the bottom.

  The out-of-range detection means is composed of levers 45A and 45B, and the bypass communication valve converts the original pressure PL to the operating pressure of the hydraulic servos 40A1, 40A2, 40B1, and 40B when the end stop levers 48A and 48B are rotated. As the first end stop valve 115 and the second end stop valve 116 are supplied as follows, it is detected that the power roller 4 has exceeded a predetermined movement range, and the operating pressures of the hydraulic servos 40A1, 40A2, 40B1, 40B are set. The operation up to the sudden rise can be mechanically performed without detecting the electrical pressure and electrically increasing the hydraulic pressure, for example, and can be made inexpensive and compact.

  Further, since the variator 5 is a full toroidal type, the power roller 4 can be autonomously shifted by moving and driving the power roller 4 on a surface parallel to both the disks 2A, 2B and 3. it can.

  In the hydraulic control apparatus according to the present embodiment described above, the pilot shuttle check valve and the ball shuttle valve are assumed to guide the maximum operating pressure of the plurality of operating pressures to the original pressure regulating valve (primary regulator valve). However, the present invention is not limited to these, but any type of system can be used as long as the maximum operating pressure of the plurality of operating pressures is led to the original pressure regulating valve. It may be. Furthermore, the pilot shuttle check valve, the ball shuttle valve, the bypass check valve, etc., which mainly use the check ball as the pressure receiving member have been described as an example, but the larger one of the two input hydraulic pressures For example, a spool may be used as the pressure receiving member, that is, any shuttle valve can be used in the present invention.

The skeleton figure which shows the continuously variable transmission which concerns on this invention. The speed diagram of the continuously variable transmission which concerns on this invention. 1 is an overall cross-sectional view showing a continuously variable transmission according to the present invention. The schematic cross-sectional view which shows a toroidal type continuously variable transmission part. The schematic diagram which shows a part of toroidal type continuously variable transmission part. The schematic diagram which shows a part of toroidal type continuously variable transmission part. The schematic diagram which shows a part of toroidal type continuously variable transmission part. The schematic diagram which shows a part of toroidal continuously variable transmission part and a part of hydraulic control apparatus.

Explanation of symbols

1 continuously variable transmission 2A, 2B input disk 3 output disk 4 power roller 5 continuously variable transmission (variator)
40Aa, 40Ab, 40Ba, 40Bb Oil chamber 40A1, 40A2, 40B1, 40B2 Hydraulic servo for hydraulic control (hydraulic servo)
40Ap, 40Bp Piston 41 Roller carriage 41, 42, 45, 46 Link mechanism 45A, 45B Lever member (lever)
48A, 48B Out-of-range detection means (end stop lever)
110 Oil pressure source (oil pump)
111 source pressure regulator (primary regulator valve)
112 Working pressure regulator (first reaction pressure control valve)
113 Working pressure regulator (second reaction pressure control valve)
115 Detour communication valve (first end stop valve)
115a, 116a Output port 115b, 116b Input port 115f, 116f Spool 115g, 116g Spring 116 Detour communication valve (second end stop valve)
B Roller drive device d Rotation center PL Original pressure (line pressure)
PS1, PS2 Operating pressure (servo pressure)

Claims (4)

A variator device having an input disk, an output disk, and a power roller sandwiched between the two disks;
It has a hydraulic servo and a link mechanism that supports the power roller, and based on the pressure drive of the hydraulic servo, the rotational center of the power roller can be driven to move with respect to the surface direction of the two disks via the link mechanism A roller driving device;
A source pressure regulating valve that regulates the hydraulic pressure of the hydraulic pressure source as the source pressure;
A toroidal continuously variable transmission comprising: an operating pressure regulating valve that regulates the original pressure as an operating pressure to be supplied to the hydraulic servo;
Out-of-range detection means for detecting that the rotation center of the power roller exceeds a predetermined movement range;
When the out-of-range detecting means detects that the rotation center of the power roller exceeds a predetermined moving range, the detour bypassing the operating pressure regulating valve and supplying the original pressure as the operating pressure of the hydraulic servo A road communication valve,
A toroidal-type continuously variable transmission. A plurality of the operating pressure regulating valves are provided,
The hydraulic servo is a power roller capable of inputting different operating pressures to two oil chambers opposed to a piston and pressing the rotation center of the power roller against the surface direction of the two disks via the link mechanism. Hydraulic servo for control,
The bypass communication valve supplies the original pressure to an oil chamber that acts on the piston in a direction in which the power roller is pressed and driven within the predetermined movement range.
The toroidal continuously variable transmission according to claim 1. The link mechanism is connected to a lever member that can be driven to rotate on surfaces parallel to the two disks by the pressing drive of the hydraulic servo, and is pivotably and rotatably connected to the lever member, and the rotation of the power roller. A roller carriage that rotatably supports the center,
The out-of-range detection means is configured such that when the lever member rotates more than a predetermined angle, the rotation center of the power roller exceeds a predetermined movement range by rotating by being abutted and pressed against the lever member. It consists of an end stop lever that detects
When the end stop lever is rotated, the bypass communication valve directly communicates with the input port for inputting the original pressure and the hydraulic servo when the spool is pressed against the spring by the end stop lever. An output port that communicates with each other, and includes an end stop valve that bypasses the operating pressure regulating valve and supplies the original pressure as the operating pressure of the hydraulic servo.
The toroidal continuously variable transmission according to claim 1 or 2. The variator device is a full toroidal type,
The toroidal continuously variable transmission according to any one of claims 1 to 3.

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