Continuously variable transmission

Abstract: In a continuously variable transmission using a belt type continuously variable transmission unit which is composed of a primary pulley and a secondary pulley and whose torque transmitting direction is switched by torque circulation when a forward traveling is switched to a backward traveling and vice versa, when a low mode is selected in a D range, a somewhat high hydraulic pressure from a regulator valve 56 is supplied to a secondary hydraulic servo 33 through the ports f and i of a manual valve 59 as well as a somewhat low hydraulic pressure from a ratio control valve 57 is supplied to a primary hydraulic servo 32 through ports h and g. When the D range is switched to a R range by the manual valve 59, the supply of the above hydraulic pressures is reversed. With this arrangement, the axial forces of both the pulleys are changed by switching the forward traveling and backward traveling.

Description: The present invention relates to a continuously variable transmission using a belt type continuously variable transmission unit composed of a primary pulley and a secondary pulley composed of two sheaves and a belt trained around the primary and secondary pulleys, and more specifically, to control of the axial forces of the belt type continuously variable transmission unit when forward traveling and backward traveling are switched in a continuously variable transmission composed of the combination of the belt type continuously variable transmission unit and a planetary gear which is particularly suitable as a transmission in automobiles. Recently, an automatic continuously variable transmission having a belt type continuously variable transmission unit assembled therein has attracted attention as a transmission for automobiles to cope with the requirement for improving a fuel consumption ratio, a driving capability, and the like. As shown in. for example, Japanese Patent Application Laid-Open No. 56-52653. there is conventionally devised a continuously variable transmission having a planetary gear train capable of changing a speed in the two stages of directly-coupled rotation and forward speed reducing rotation, a belt type continuously variable transmission unit and a planetary gear unit for synthesizing the rotation output from the above planetary gear train and the rotation output from the belt type continuously variable transmission unit and outputting the synthesized rotation for continuously changing a speed over a forward direction and a reverse direction by the combination of the two-stage speed change effected by the above planetary gear train and the continuously variable speed change effected by the belt type continuously variable transmission unit. In the continuously variable transmission, when a select lever is set to a D position, the speed reducing forward rotation of the planetary gear train is executed. Then. the speed reducing forward rotation is coupled with the speed changing rotation of the belt type continuously variable transmission unit and taken out from the planetary gear unit as a forward continuously variable rotation. Whereas, when the select lever is set to a R position, the directly-coupled rotation of the planetary gear train is executed. Then. the directly-coupled rotation is coupled with the rotation of the belt type continuously variable transmission unit likewise and taken out from the planetary gear unit as a reverse continuously variable rotation. At the time, the direction in which power is transmitted by the belt type continuously variable transmission unit, that is, the driving/driven relationship between a primary pulley and a secondary pulley is switched in a forward traveling state (D position) and a reverse traveling state (R position) due to torque circulation. In general. the belt type continuously variable transmission unit changes a pulley ratio to thereby change a speed by changing axial forces (belt holding pressures) acting on both driving and driven side pulleys. Although the ratio of axial forces acting on both the pulleys must be set to a predetermined value to transmit power by setting both the pulleys to a predetermined pulley ratio, the axial force of the driving side pulley must be made larger than that of the driven side pulley at the time. In the aforesaid continuously variable transmission, however, since the axial forces of the pulleys are not taken into consideration regardless of that the power transmitting direction of the belt type continuously variable transmission unit is changed in forward rotation and reverse rotation due to the torque circulation, speed changing control is made troublesome when forward traveling and reverse traveling are switched and a time is consumed before the speed changing control is stabilized. thus there is a possibility that a response may be delayed. US-A-4 990 127, in accordance with the preamble of claim 1, discloses a dual range infinitely variable transmission utilizing a continuously variable speed mechanism in conjunction with a fixed ratio speed mechanism to provide two power paths from the input to the rotatable output. A jaw clutch is utilized to switch between the high range and low range of the transmission. A reaction type clutch is utilized to provide a positive neutral for the transmission. GB-A-1 603 853 discloses a continuously variable-ratio transmission, comprising a pair of pulleys of variable effective diameter interconnected by an endless drive band and each connected to drive a respective one of two of the elements of an epicyclic gearset. This transmission includes means for monitoring the torque in an input shaft of the transmission, means for controlling the ratio provided by the pulleys and feedback means operable in accordance with the torque in the output shaft to bias the control means in such a manner as to oppose the generation of the torque. Patent Abstracts of Japan, vol. 11, no. 57 (M-564), February 21, 1997 and JP-A-61-218860 relate to a stepless speed changer with V-belt. To control the torque ratio, these references describe that the oil pressure is supplied or exhausted to/from a number of oil chambers in pulleys by the use of a valve so as to change the effective belt diameter of each pulley. In response to the torque ratio control, an oil pressure in accordance with the axial force ratio is held in a first oil chamber in the first pulley during positive torque transmission, and an oil pressure in accordance with the axial force ratio is held in a second oil chamber in the second pulley during negative torque transmission. An object of the present invention is to provide a continuously variable transmission capable of solving the above problem by changing the axial forces acting on both the pulleys of a belt type continuously variable transmission unit. This object is achieved with the features of the claims. According to the arrangement of the present invention, the engine output from the input shaft is subjected to a proper speed changing operation effected by the belt type continuously variable transmission unit and transmitted to the second rotary element of the planetary gear as well as a constant rotation is transmitted to the first rotary element. then both the rotations are synthesized by the planetary gear and taken out from the third rotary element by the output chaft. Then, when the axial force actuating means is controlled by the control means, the axial force of the first or second pulley is changed, so that the pulley ratio of the belt type continuously variable transmission unit is changed to a predetermined pulley ratio so as to properly change the speed thereof. Further, when the forward traveling state or the backward traveling state is selected by the selection means, the direction of torque output from the output shaft is changed to a forward rotation of the reverse rotation as well as the torque transmitting direction of the belt type continuously variable transmission unit is changed. At the time, the relative relationship in magnitude of the axial forces acting on the above first and second pulleys is reversed by the switching mean, so as to coincide with the changed torque transmitting direction of the belt type continuously variable transmission unit. According to the present invention of claim 1, when the forward traveling state and backward traveling state of the vehicle is switched, the switching means switches the relative relationship in magnitude of the axial forces acting on both the pulleys so as to coincide with the relationship of the axial forces of both the pulleys of the belt type continuously variable transmission unit whose torque transmitting direction is reversed by the torque circulation, so that the forward traveling and backward traveling can be securely and promptly switched. Since the hydraulic pressures acting on the first and second hydraulic servos are reversed by the switching valve, the hydraulic pressures of both the hydraulic servos need not be individually controlled, so that the forward traveling and backward traveling can be simply and promptly switched by changing oil passages by the switching valve. Since the switching valve is composed of the manual valve, a driver can securely switch the forward traveling and backward traveling based on his intention. According to the present invention of claim 2, since each of the first and second hydraulic servos has a plurality of hydraulic chambers, the forward traveling and backward traveling can be easily and securely switched by reversing the effective pressure-receiving areas of both the hydraulic servos by reversing the hydraulic pressures supplied to the hydraulic chambers without the need of controlling the regulation of the hydraulic pressures acting on both the hydraulic servos. According to the present invention of claim 3, since hydraulic pressures are supplied to the first hydraulic chambers of both the hydraulic servos at all times and a hydraulic pressure supplied to the second hydraulic chamber of any one of both the hydraulic servos is switched, the forward traveling and backward traveling can be easily and securely switched. According to the present invention of claim 4, when the regulator valve and the ratio control valve are disposed downstream of the switching valve, the number of valves must be increased or the function of a single valve must be increased. However, when the regulator valve and the ratio control valve are interposed between the oil pump and the switching valve, the above increase is not needed and the arrangement can be simplified. According to the present invention of claim 5, since the two regulator valves regulate different hydraulic pressures, respectively and the supply of the regulated hydraulic pressures is switched between both the hydraulic servos in the forward traveling state and backward traveling state, the forward traveling and backward traveling can be securely and easily switched by reversing the axial forces of both the pulleys. According to the present invention of claim 6, the arrangement can be simplified without the need of an increase of the number of valves and an increase of the function of a single valve likewise claim 4. According to the present invention of claim 7 since the hydraulic pressure of the hydraulic servo on which a high hydraulic pressure acts is lowered by the relief valve, axial forces acting on the pulleys are changed so as to coincide with a torque transmitting direction even in a coast state in which the torque transmitting direction is reversed, thus the coast state can be securely maintained. According to the present invention of claim 8, torque circulation is securely executed by the planetary gear. In the following preferred embodiments of the present invention will be described with reference to the drawings, in which: FIG. 1 shows an embodiment of a power transmission mechanism in a continuously variable transmission according to the present invention, wherein FIG. 1(a) is a skeleton diagram and FIG. 1(b) is a graph showing a speed; FIG. 2 is a graph showing the change of output torque relating to the torque ratio of a belt type continuously variable transmission (CVT) unit of the continuously variable transmission; FIG. 3 is a graph showing the change of output r.p.m. relating to the torque ratio of the CVT unit; FIG. 4 is a cross sectional view showing double chamber type hydraulic servos applying axial forces to primary and secondary pulleys; FIG. 5 is a hydraulic piping diagram of a hydraulic control mechanism of the continuously variable transmission showing a low (L) mode state in a D range; FIG. 6 is a diagram showing a high (H) mode state in the D range in the hydraulic piping diagram; FIG. 7 is a diagram showing an R range in the hydraulic piping diagram; FIG. 8 is a table showing the operation of the hydraulic control mechanism; FIG. 9 is a cross sectional view showing double regulator type hydraulic servos; FIG. 10 is a hydraulic piping diagram of a hydraulic control mechanism of the double regulator type hydraulic servos showing a low (L) mode state in a D range; FIG. 11 is a diagram showing a high (H) mode state in the D range in the hydraulic piping diagram; FIG. 12 is a diagram showing an R range in the hydraulic piping diagram; FIG. 13 is a table showing the operation of the hydraulic control mechanism; FIG. 14 is a diagram showing a partially modified hydraulic control mechanism applicable to the above double chamber type hydraulic servos: FIG. 15 is a diagram showing a partially modified hydraulic control mechanism applicable to the above double chamber type hydraulic servos: FIG. 16 is a block diagram showing an electric control mechanism according to the embodiment: FIG. 17 is a general flowchart showing the electric control mechanism: FIG. 18 is a flowchart showing a reverse (R) range control subroutine of the electric control mechanism: FIG. 19 is a flowchart showing up-shift control of the electric control mechanism: FIG. 20 is a flowchart showing a Lo → Hi determining subroutine of the electric control mechanism: FIG. 21 is a flowchart showing down-shift control of the electric control mechanism: FIG. 22 is a flowchart showing a Hi → Lo determining subroutine of the electric control mechanism: FIG. 23 is a graph showing the balance between the axial forces of the primary and secondary pulleys; FIG. 24 shows another embodiment of the above power transmission mechanism, wherein FIG. 24(a) is a skeleton diagram and FIG. 24(b) is a graph showing a speed: FIG. 25 shows another embodiment of the above power transmission mechanism, wherein FIG. 25(a) is a skeleton diagram and FIG. 25(b) is a graph showing a speed: and FIG. 26 shows another embodiment of the above power transmission mechanism, wherein FIG. 26(a) is a skeleton diagram and FIG. 26(b) is a graph showing a speed. Embodiments of the present invention will be described below with reference to the drawings. First, a power transmission mechanism of a continuously variable transmission according to the present invention will be described with reference to FIG. 1(a). A vehicle-mounted automatic continuously variable transmission 1 has a first shaft 3 aligned with an engine crank shaft 2. a second shaft 5 and a third shaft 6 (a, b) aligned with a front wheel shaft. A primary pulley 7 is supported by the input shaft 3. a secondary pulley 9 is supported by the second shaft 5 and a belt 10 is trained around these two pulleys 7. 9 to thereby constitute a belt type continuously variable transmission unit 11. Further, the first shaft 3 is coupled with the engine crank shaft 2 through a damper 12 to thereby constitute an input shaft. The input side member 13 of a low clutch C L is fixed to the input shaft 3 as well as the output side member 15 of the low clutch C L is rotatably supported by the input shaft 3 and further a primary side sprocket 16 constituting the power transmission means is coupled with the output side member 15 integrally therewith. In addition, the fixed sheave 7a of the primary pulley 7 is fixed to the input shaft 3 as well as an oil pump 17 is coupled with the extreme end of the input shaft 3 and a movable sheave 7b is axially movably supported by the fixed sheave 7a. The secondary pulley 9 is rotatably supported by the second shaft 5 and has a fixed sheave 9a. a movable sheave 9b axially movably supported by the fixed sheave and a secondary shaft 9c integrally coupled with the fixed sheave. Further, a high clutch C H is interposed between the second shaft 5 and the secondary shaft 9c. a planetary gear 19 is disposed to the second shaft 5 as well as a secondary side sprocket 20 is rotatably supported by the output shaft 5. The output shaft 5 has an output gear 21 fixed at an end thereof to thereby constitute an output shaft. The planetary gear 19 is composed of a single pinion planetary gear having a sun gear 19s. a ring gear 19r and a carrier 19c rotatably supporting pinions 19p each meshed with both the sun gear and ring gear. The sun gear 19s is coupled with the secondary shaft 9c to thereby constitute a second rotary element, the ring gear 19r is coupled with the output shaft 5 to thereby constitute a third rotary element and the carrier 19c is coupled with the secondary side sprocket 20 to thereby constitute a first rotary element. Further, a training member 22 such as a silent chain, roller chain. timing belt or the like is trained around the primary side and secondary side sprockets 16. 20. The gear 21 fixed to the output shaft 5 is meshed with the large gear 23a of a speed reduction gear unit 23 and the small gear 23b of the unit is meshed with the ring gear 24 of a differential gear unit 25 which outputs a differential rotation to right and left axle shafts 6a, 6b constituting the third shaft. Next, operation based on the power transmission mechanism of the continuously variable transmission 1 will be described with reference to FIG. 1(a), FIG. 1(b), FIG. 2 and FIG. 3. The rotation of the engine crank shaft 2 is transmitted to the input shaft 3 through the damper 12. In a low mode in which the low clutch C L is connected and the high clutch C H is disconnected, the rotation of the input shaft 3 is transmitted to the primary pulley 7 as well as to the carrier 19c of the planetary gear 19 through the power transmission unit 26 composed of the primary side sprocket 16, training member 22 and secondary side sprocket 20. Whereas, the rotation of the primary pulley 7 is continuously varied by that the pulley ratio of the primary and secondary pulleys is properly adjusted by an axis force actuating means such as a hydraulic servo and the like to be described later and transmitted to the secondary pulley 9 and further the speed-changed rotation of the pulley 9 is transmitted to the sun gear 19s of the planetary gear 19. As shown in the graph of a speed of FIG. 1(b). the carrier 19c to which a constant speed rotation is transmitted through the power transmission unit 26 acts as a reaction element in the planetary gear 19 and a continuously variable rotation is transmitted from the belt type continuously variable transmission (hereinafter, referred to as CVT) unit 11 to the sun gear 19s, and the rotations of these carrier and sun gear are synthesized and transmitted to the output shaft 5 through the ring gear 19r. At the time, since the output shaft 5 is coupled with the ring gear 19r which is a rotary element other than a reaction support element, torque circulation is caused to the planetary gear 19 as well as the sun gear 19s and the carrier 19c are rotated in the same direction. As a result, the output shaft 5 is rotated in a forward (Lo) direction and a reverse (Rev) direction across a zero rotation. That is, when the output shaft 5 is rotated in the forward (advancing) direction based on the torque circulation, the belt type continuously variable transmission unit 11 transmits torque from the secondary pulley 9 to the primary pulley 7. whereas when the output shaft 5 is rotated in the backward (retracting) direction, the transmission unit 11 transmits torque from the primary pulley 7 to the secondary pulley 9. Then, in a high mode in which the low clutch C L is disconnected and the high clutch C H is connected, since power transmission to the planetary gear 19 through the power transmission unit 26 is shut off, the planetary gear 19 is rotated integrally with the high clutch C H by being engaged therewith. Consequently, the rotation of the input shaft 3 is entirely transmitted to the output shaft 5 through the belt type CVT unit 11 and the high clutch C H . That is, the CVT unit 11 transmits power from the primary pulley 7 to the secondary pulley 9. Further, the rotation of the output shaft 5 is transmitted to the differential gear unit 25 through the output gear 21 and speed reduction gear unit 23 and further transmitted to the right and left front wheels through the right and left axle shafts 6a, 6b. As shown in the graph of a speed of FIG. 1(b), the graph of output torque of FIG. 2 and the graph of output r.p.m. of FIG. 3, since the sun gear 19s is at its maximum rotation when the belt type CVT unit 11 is at the limit (O/D end) in a speed increasing direction (at the position a in the graph of FIG. 1) in the low mode, the ring gear 19r is rotated in a reverse direction with respect to the carrier 19c having a constant r.p.m. and the reverse rotation (REV) is transmitted to the output shaft 5. When the CVT unit 11 is shifted in a speed reducing (U/D) direction the r.p.m. of the reverse rotation is reduced, so that a neutral position (NEU) where the r.p.m. of the output shaft 5 is made zero is reached at a predetermined pulley ratio determined by the gear ratio of the planetary gear 19 and the power transmission unit 26. When the CVT unit 11 is further shifted in the speed reducing direction, since the ring gear 19r is switched to a forward rotating direction, the forward rotation, i.e, the rotation in a forward traveling direction is transmitted to the output shaft 5. At the time, as apparent from FIG. 2, the torque of the second shaft 5 diverges indefinitely in the vicinity of the above neutral position NEU. Next, when the CVT unit 11 reaches the end of the speed reducing direction (U/D), the high clutch C H is connected and an operation mode is switched to the high mode. In the high mode, since the rotation output from the CVT unit 11 is transmitted to the output shaft 5 as it is, a parallel line is obtained as shown by b in the graph of a speed of FIG. 1(b). Thereafter, as the CVT unit 11 is shifted in a speed increasing (O/D) direction, the rotation of the output shaft 5 is also shifted in the speed increasing direction and an amount of transmitted torque is reduced accordingly. Note, λ in FIG. 1(b) represents the ratio (Zs/Zr) of the number of teeth Zs of the sun gear to the number of teeth of the ring gear Zr. Next, an embodiment of a hydraulic control mechanism of the continuously variable transmission according to the present invention will be described. In the primary and secondary pulleys 7, 9, the movable sheaves 7b, 9b are axially movably supported by the bosses 7c, 9c of the fixed sheaves 7a, 9a through ball splines 30, 31 and hydraulic servos 32, 33 are disposed on the back surfaces of these movable sheaves 9b, 9b and constitute the axial force actuating means for applying axial forces to the pulleys, respectively, as shown in FIG. 4. Both of the hydraulic servos 32, 33 have partition members 35, 36 fixed to the bosses 7c, 9c of the fixed sheaves, cylinder members 37, 39, drum members 40, 41 fixed to the back surfaces of the movable sheaves 7, 9b and second piston members 42, 43. The partition members 35, 36 are engaged with the drum members 40, 41 in an oil tight state as well as the second piston members 42, 43 are engaged with the cylinder members 37, 39 and the partition members 35, 36 in an oil tight state to thereby constitute a double piston structure composed of first hydraulic chambers 45, 46 and second hydraulic chambers 47, 49, respectively. The first hydraulic chambers 45, 46 in the above hydraulic servos 32, 33 form piston surfaces on the back surfaces of the movable sheaves 7b, 9b, respectively and the piston surfaces have the same effective pressure-receiving areas on the primary side and the secondary side. Further, oil passages 50, 51, 52, 53 are defined to the bosses 7c, 9c of the fixed sheaves 7a, 9a on the primary side and secondary side and communicate with the first hydraulic chambers 45, 46 and second hydraulic chambers 47, 49, respectively. In addition, a preloading spring 55 is contractedly disposed to the first hydraulic chamber 45 of the primary side hydraulic servo 32. As shown in FIG. 5, the hydraulic control mechanism (means) 54 of this embodiment has a primary regulator valve 56, a ratio control valve 57, a down-shift relief valve 58, a manual valve (switching valve, selection means) 59 and a low/high control valve (low/high shifting means) 60. Thus, the hydraulic pressure from the oil pump 17 is properly regulated as well as switched and supplied to the first and second hydraulic chambers 45, 46, 47, 49 of the above hydraulic servos 32, 33, the low clutch (hydraulic servo) C L and high clutch (hydraulic servo) C H . Next, operation of the above hydraulic control mechanism 54 will be described with reference to FIG. 4 to FIG. 8. As shown in FIG. 8, when the low (L) mode is employed in a D range, a predetermined hydraulic pressure is supplied to the hydraulic chamber 45 of the primary side hydraulic servo 32 as well as a predetermined hydraulic pressure acts on both the first hydraulic chamber 46 and second hydraulic chamber 49 of the secondary side hydraulic servo 33 and further a hydraulic pressure acts on the low clutch C L to connect it. That is, in the low mode, the manual valve 59 is actuated to the position of the D range shown in FIG. 5 so as to cause a port d to communicate with a port e, a port f with a port g, a port h with a port i and a port k with a port l as well as a port j with a drain port Ex. Further, the low/high control valve 60 is located at a low (L) position and switched and maintained to cause a port o to communicate with a port p, a port q with a port r, a port s with a port t, a port u with a port v and a port w with a port x as well as a port y with the drain port Ex. Therefore, the hydraulic pressure from the output port c of the primary regulator valve 56 is supplied to the low clutch hydraulic servo C L through the ports d and e in the manual valve 59 located at the position of the D range (refer to FIG. 5) and further the ports o and p of the low/high control valve 60 and causes the clutch C L to be engaged. In addition, the above hydraulic pressure is supplied to the first hydraulic chamber 45 of the primary hydraulic servo 32 through the ports f and g of the manual valve and further the ports q and r of the low/high control valve 60. In a state other than down shift operation, the down-shift relief valve 58 is located at a normal position where the port m communicates with the port n and the hydraulic pressure from the above output port c is supplied to the first hydraulic chamber 46 of the secondary hydraulic servo 33 through the ports m and n, the ports h and i of the manual valve 59 and the ports s and t of the low/high control valve 60. The hydraulic pressure from the above output port c is properly regulated by the ratio control valve 57 so that it becomes a hydraulic pressure corresponding to a target pulley ratio and the hydraulic pressure from the output port z is supplied to the second hydraulic chamber 49 of the hydraulic servo 33 through the ports k and l of the manual valve 59 and the ports w and x of the low/high control valve 60. Note, in this state, the high clutch hydraulic servo C H is caused to communicate with the drain port Ex from the port y of the low/high control valve 60 and in a released state. Further, the second hydraulic chamber 47 of the primary hydraulic servo 32 communicates with the drain port Ex through the ports u and v of the low/high control valve 60 and the port j of the manual valve 59. With this operation, the low clutch C L is connected as well as the axial force caused by the secondary side hydraulic servo 33 in which a hydraulic pressure acts on both the first and second hydraulic chambers 46, 49 is made higher than the axial force caused by the primary side hydraulic servo 32 in which a hydraulic pressure acts on only the first hydraulic chamber 45 in the CVT unit 11. Thus, the opening/closing of the above ratio control valve 57 is properly adjusted in the state of the axial forces of both the pulleys 9, 7 corresponding to the torque transmitted from the secondary pulley 9 to the primary pulley 7. With this operation, the hydraulic pressure of the second hydraulic chamber 49 of the secondary hydraulic servo 33 is regulated to thereby properly adjust the axial force caused by the hydraulic servo 33, so that a pulley ratio (torque ratio) is properly changed. In this state, the engine torque which is transmitted from the input shaft 3 to the carrier 19c of the planetary gear 19 through the low clutch C L and power transmission unit 26 is taken out from the output shaft 5 through the ring gear 19r while being regulated by the CVT unit 11 through the sun gear 19s based on the above predetermined pulley ratio. In the high (H) mode in the D range, a predetermined hydraulic pressure is supplied to the first and second hydraulic chambers 45, 47 of the primary side hydraulic servo 32 as well as to the first hydraulic chamber 46 of the secondary side hydraulic servo 33 and further to the high clutch hydraulic servo C H as shown in FIG. 8. That is, in the high (H) mode, although the manual valve 59 is located at the D range position likewise the above low mode, since the low/high control valve 60 is switched to the high position (H), the port o communicates with the port y, the port q with the port t, the port s with the port r, the port x with the port v and the port w with the port u, respectively and the port p communicates with the drain port Ex as shown in FIG. 6. Therefore, the hydraulic pressure output from the primary regulator valve 56 is supplied to the high clutch hydraulic servo C H through the ports d and e of the manual valve 59 and the ports o and y of the low/high control valve 60 to thereby cause the clutch C H to engage. Further, the above hydraulic pressure is supplied to the first hydraulic chamber 46 of the secondary hydraulic servo 33 through the ports f and g of the manual valve 59 and the ports q and t of the low/high control valve 60. In addition, the hydraulic pressure is supplied to the first hydraulic chamber 45 of the primary hydraulic servo 32 through the ports m and n of the down-shift relief valve 58, the ports h and i of the manual valve 59 and the ports s and r of the low/high control valve 60 and to the second hydraulic chamber 47 of the primary hydraulic servo 32 through the output port z of the ratio control valve 57 in which the hydraulic pressure is properly regulated, the ports k and l of the manual valve 59 and the ports w and u of the low/high control valve 60. Note, the low clutch hydraulic servo C L is released by being caused to communicate with the drain port Ex from the port p of the low/high control valve 60 in this state. Further, the second hydraulic chamber 49 of the secondary hydraulic servo 33 communicates with the drain port Ex through the ports x and v of the low/high control valve 60 and the port j of the manual valve 59. With this operation, the high clutch C H is connected as well as the axial force caused by the primary side hydraulic servo 32 in which a hydraulic pressure is supplied to the first and second hydraulic chambers 45, 47 is made higher than the axial force caused by the secondary side hydraulic servo 33 in which a hydraulic pressure is supplied only to the first hydraulic chamber 46 in the CVT unit 11. Thus, above ratio control valve 57 is properly adjusted in the state of the axial force corresponding to the torque transmitted from the primary pulley 7 to the secondary pulley 9, With this operation, the hydraulic pressure of the second hydraulic chamber 47 of the primary hydraulic servo 32 is regulated to thereby adjust the axial force caused by the primary hydraulic servo 32, so that a proper pulley ratio (torque ratio) can be obtained. In this state, the torque transmitted from an engine to the input shaft 3 is properly changed by the CVT unit 11 by being transmitted from the primary pulley 7 to the secondary pulley 9 and further taken out from the output shaft 5 through the high clutch C H . In a reverse range (R), a predetermined hydraulic pressure is supplied to the first and second hydraulic chambers 45, 57 of the primary side hydraulic servo 32 as well as to the first hydraulic chamber 46 of the secondary side hydraulic servo 33 and further to the low clutch hydraulic servo C L as shown in FIG. 8. That is, in the reverse range, the manual valve 59 is located at a reverse (R) range position and the low/high control valve 60 is located at the low position (L) as shown in FIG. 7. In this state, the port d communicates with the port e, the port f with the port i and the port k with the port j, respectively in the manual valve 59 and the port 1 communicates with the drain port Ex. Further, the port o communicates with the port p, the port q with the port r, the port s with the port t, the port v with the port u and the port x with the port w, respectively in the low/high control valve 60 likewise the aforethe low (L) mode and the port y communicates with the drain port Ex. Therefore, the hydraulic pressure from the output port c of the primary regulator valve 56 is supplied to the low clutch hydraulic servo C L through the ports d and e of the manual valve 59 and the ports o and p of the low/high control valve 60 and further to the first hydraulic chamber 46 of the secondary hydraulic servo 33 through the ports f and i of the manual valve 59 and the ports s and t of the low/high control valve 60. Further, the above hydraulic pressure is supplied to the first hydraulic chamber 45 of the primary hydraulic servo 32 through the ports m and n of the down-shift relief valve 58, the ports h and g of the manual valve 59 and the ports q and r of the low/high control valve 60. In addition, the hydraulic pressure is properly regulated by the ratio control valve 57 and supplied from the output port z thereof to the second hydraulic chamber 47 of the primary hydraulic servo 32 through the ports k and j of the manual valve 59 and the ports v and u of the low/high control valve 60. With this operation, the low clutch C L is connected as well as the axial force caused by the primary side hydraulic servo 32 in which a hydraulic pressure acts on the first and second hydraulic chambers 45, 47 is made higher than the axial force caused by the secondary side hydraulic servo 33 in which a hydraulic pressure acts only on the first hydraulic chamber 46 in the CVT unit 11. Thus, the state of an axial force corresponding to the torque transmitted from the primary pulley 7 to the secondary pulley 9 is achieved as well as the hydraulic pressure in the second hydraulic chamber 47 of the primary hydraulic servo 32 is regulated by the adjustment of the ratio control valve 57 to thereby adjust the axial force caused by the hydraulic servo 32, so that a proper pulley ratio can be obtained. In this state, since the pulley ratio of the CVT unit 11 is in a predetermined speed increasing (O/D) state, the engine torque from the input shaft 3 is transmitted to the carrier 19c of the planetary gear 19 through the low clutch C L and power transmission unit 26 as well as to the sun gear 19s through the CVT unit 11 which transmits torque from the primary pulley 7 to the secondary pulley 9, thus the both of the torque are synthesized by the planetary gear 19 and taken out from the output shaft 5 through the ring gear 19r as a reverse rotation. In a parking position P and neutral position N of the manual valve 59, both the low clutch C L and high clutch C H are released as well as a predetermined hydraulic pressure is supplied to the first hydraulic chambers 45, 46 of the primary side and secondary side hydraulic servos 32, 33 as shown in FIG. 8. That is, the port f communicates with the port g, the port h with the port i in the manual valve 59 and further the ports e, j, l communicate with the drain port Ex, respectively. In addition, the low/high control valve 60 is kept at the aforesaid low position L. Therefore, the hydraulic pressure output from the primary regulator valve 56 is supplied to the first hydraulic chamber 45 of the primary hydraulic servo 32 through the ports f and g of the manual valve 59 and the ports q and r of the low/high control valve 60 as well as to the first hydraulic chamber 46 of the secondary hydraulic servo 33 through the ports m and n of the down-shift relief valve 58, the ports h and i of the manual valve 59 and the ports s and t of the low/high control valve 60. Further, the hydraulic pressure of the high clutch hydraulic servo C H is released through the port y of the low/high control valve 60 and the drain port Ex, the hydraulic pressure of the low clutch hydraulic servo C L is released through the ports p and o of the low/high control valve 60, the port e of the manual valve 59 and the drain port Ex, the hydraulic pressure of the second hydraulic chamber 47 of the primary hydraulic servo 32 is released through the ports u and v of the low/high control valve 60, the port j of the manual valve 59 and the drain port Ex, and the hydraulic pressure of the second hydraulic chamber 49 of the second hydraulic servo 33 is released through the ports x and w of the low/high control valve 60, the port l of the manual valve 59 and the drain port Ex. With this operation, since both of the low clutch C L and high clutch C H are released and the same hydraulic pressure acts only on the first hydraulic chambers 45, 46 of the primary hydraulic servo 32 and secondary hydraulic servo 33, approximately the same axial force acts on both the primary and secondary pulleys 7, 9. Note, since a predetermined hydraulic pressure is supplied from the primary regulator valve 56 to the first hydraulic chambers 45, 46 of both the primary and secondary hydraulic servos 32, 33, respectively in the aforesaid respective positions D, N and R and the low (L) mode and high (H) mode, a predetermined axial force corresponding to the transmitted torque is secured so that a belt does not slip, Further, since the regulated hydraulic pressure from the ratio control valve 57 is supplied to the second hydraulic chamber 47 or 49 of any one of the above hydraulic servos 32, 33, the ratio of the axial forces of both the pulleys 7, 9 is adjusted so that a speed changing operation is executed so as to obtain a predetermined pulley ratio. Further, when down-shift is effected in a coast state, the down-shift relief valve 58 is switched to cause the port n to communicate with the drain port Ex. With this operation, the hydraulic pressure supplied to the first hydraulic chamber 45 or 46 of both the hydraulic servos 32, 33 as described above is discharged through predetermined ports, the port n of the relief valve 58 and the drain port Ex and the hydraulic pressure in the hydraulic servo on which a high hydraulic pressure acts is made lower than the hydraulic pressure of the hydraulic servo on which a low hydraulic pressure acts. Next, the control mechanism according to another embodiment will be described with reference to FIG. 9 to FIG. 13. Although primary and secondary side hydraulic servos 32, 33 of this embodiment shown in FIG. 9 have first and second hydraulic chambers 73 1 , 75 1 , 73 2 , 75 2 , respectively likewise those shown in FIG. 4, holes 70, 70 are defined to the partition members 35, 36 so as to simply increase pressure receiving areas. Thus, this embodiment constitutes a hydraulic actuator composed of substantially each one of hydraulic chambers 73, 75 which are supplied with hydraulic pressures from oil passages 71, 72, respectively and the hydraulic chambers 73, 75 of both the hydraulic servos 32, 33 have the same pressure-receiving areas. Further, a preloading spring 55 is contractedly disposed to the first hydraulic chamber 75 1 of the secondary side hydraulic servo 33. As shown in FIG. 10 to FIG. 12, the control mechanism 54' has a primary regulator valve 56, a ratio control valve 57 constituting a secondary regulator valve and a down-shift relief valve 58 which are continuously disposed in series from an oil pump 17 and further a manual valve 59' and a low/high control valve 60' each having an oil passage structure different from that of the above embodiment. The control mechanism 54' operates as shown in the operation table of FIG. 13 In the low (L) mode in the D range, a relatively low line pressure (P L - L) is supplied to the hydraulic chamber 73 of the primary hydraulic servo 32 from the ratio control valve 57, a relatively high line pressure (P L - H) is supplied to the hydraulic chamber 75 of the secondary hydraulic servo 33 from the primary regulator valve 56 and a line pressure is supplied to a low clutch hydraulic servo C L , as shown in FIG. 13. That is, since the manual valve 59' is located at the D position, a port d communicates with a port e. a port f with a port i and a port h with a port g, and since the low/high control valve 60' is located at the low (L) position, a port o communicates with a port p, a port q with a port t and a port s with a port r as well as a port y communicates with a drain port Ex, as shown in FIG. 10. Therefore, the relatively high line pressure (P L - H) regulated by the primary regulator valve 56 is supplied to the low clutch hydraulic servo C L through the ports d and e of the manual valve 59' and the ports o and p of the low/high control valve 60' and further to the secondary hydraulic servo 33 through the ports m and n of the down-shift relief valve 58, the ports f and i of the manual valve 59' and the ports s and r of the low/high control valve 60'. Further, the pressure output from the primary regulator valve 56 is further properly regulated by the ratio control valve (secondary regulator valve) 57 and the resulting regulated relatively low line pressure (P L - L) is supplied from the output port z thereof to the primary hydraulic servo 32 through the ports h and g of the manual valve 59' and the ports q and t of the low/high control valve 60'. Note, the hydraulic pressure of a high clutch hydraulic servo C H is released through the port y and drain port Ex. With this operation, a low clutch C L is connected as well as the axial force of a secondary pulley 9 caused by the secondary hydraulic servo 33 on which the relatively high line pressure P L - H acts is made higher than the axial force of a primary pulley 7 caused by the primary hydraulic servo 32 on which the relatively low line pressure P L - L acts in a CVT unit 11. This state corresponds to the low (L) mode in the D range where torque is transmitted from the secondary pulley 9 to the primary pulley 7 as well as the axial force of the primary pulley 7 is adjusted by the proper pressure regulation effected by the ratio control valve 57 so that a pulley ratio (torque ratio) is properly changed. Further, in the high (H) mode in the D range, the relatively high line pressure P L - H is supplied to the primary hydraulic servo 32 as well as the relatively low line pressure P L - L is supplied to the secondary hydraulic servo 33 and the line pressure is supplied to the high clutch hydraulic servo C H as shown in FIG. 13. That is, although the manual valve 59' is maintained at the D range position likewise the above case, since the low/high control valve 60' is switched to a high (H) position, the port o communicates with the port y and the port q with the port r and the port s with the port t, respectively, and further the port p communicates with the drain port Ex, as shown in FIG. 11. Therefore, the relatively high regulated pressure (line pressure) P L - H from the primary regulator valve 56 is supplied to the high clutch hydraulic servo C H through the ports d and e of the manual valve 59' and the ports o and y of the low/high control valve 60' and further to the primary hydraulic servo 32 through the ports m and n of the down-shift relief valve 58, the ports f and i of the manual valve 59' and the ports s and t of the low/high control valve 60'. On the other hand, the relatively low regulated pressure (line pressure) P L - L obtained by further regulating the pressure output from the primary regulator valve 56 is supplied from the output port z thereof to the secondary hydraulic servo 33 through the ports h and g of the manual valve 59' and the ports q and r of the low/high control valve 60'. Note, the hydraulic pressure of the low clutch hydraulic servo C L is released through the port p and drain port Ex. With this operation, the high clutch C H is connected as well as the axial force of the primary pulley 7 caused by the primary hydraulic servo 32 on which the relatively high line pressure P L - H acts is made higher than the axial force of the secondary pulley 9 caused by the secondary hydraulic servo 33 on which the relatively low line pressure P L - L acts in the CVT unit 11. This state corresponds to the high (L) mode where torque is transmitted from the primary pulley 7 to the secondary pulley 9 as well as the axial force of the secondary pulley 9 is adjusted by the proper pressure regulation effected by the ratio control valve 57 so that the pulley ratio (torque ratio) is properly changed. In the reverse range (R), the relatively high line pressure P L - H acts on the primary hydraulic servo 32 and the relatively low line pressure P L - L acts on the secondary hydraulic servo 33 as well as the line pressure acts on the low clutch hydraulic servo C L as shown in FIG. 13. That is, the manual valve 59 is switched to a reverse (R) range position, the port d communicates with the port e, the port f with the port g and the port h with the port i and further the low/high control valve 60' is maintained at the aforesaid low (L) position, as shown in FIG. 12. Therefore, the relatively high regulated pressure (line pressure) P L - H from the primary regulator valve 56 is supplied to the low clutch hydraulic servo C L through the ports d and e of the manual valve 59' and the ports o and p of the low/high control valve 60' as well as to the primary hydraulic servo 32 through the ports m and n of the down-shift relief valve 58, the ports f and g of the manual valve 59' and the ports q and t of the low/high control valve 60'. On the other hand, the relatively low regulated pressure (line pressure) P L - L from the ratio control valve 57 is supplied to the secondary hydraulic servo 33 through the port h and i of the manual valve 59' and the ports s and r of the low/high control valve 60'. Note, the hydraulic pressure of the high clutch hydraulic servo C H is discharged from the port y to the drain port Ex. With this operation, the low clutch C L is connected as well as the relatively high hydraulic pressure P L - H based on the primary regulator valve 56 acts on the primary hydraulic servo 32 and the relatively low hydraulic pressure P L - L based on the ratio control valve 57 acts on the secondary hydraulic servo 33 in the CVT unit 11. In the reverse (R) range, although torque is transmitted from the primary pulley 7 to the secondary pulley 9 in the CVT unit 11, the axial force of the primary pulley 7 is made higher than that of the secondary pulley 9 in correspondence with the torque transmission as well as the axial force of the secondary pulley 9 is adjusted by the pressure regulation effected by the ratio control valve 57 so that the pulley ratio is properly changed. Further, in the neutral (N) range and parking (P) range, the relatively low hydraulic pressure P L - L acts on the primary hydraulic servo 32 as well as the relatively low hydraulic pressure P L - L acts on the secondary hydraulic servo 33 as shown in FIG. 13. That is, since the manual valve 59' is switched to the neutral or parking position, the port e communicates with the drain port Ex as well as the port h communicates with the ports g and i and the low/high control valve 60' is maintained at the aforesaid low (L) position. Therefore, the hydraulic pressure from the output port z of the ratio control valve 57 is supplied from the port h of the manual valve 59' to the primary hydraulic servo 32 through the port g and the ports q and t of the low/high control valve 60' as well as to the secondary hydraulic servo 33 through the port i and the ports s and r of the low/high control valve 60'. Further, the hydraulic pressure of the high clutch hydraulic servo C H is discharged from the port y of the low/high control valve 60' to the drain port Ex and the hydraulic pressure of the low clutch hydraulic servo C L is discharged through the ports p and o of the low/high control valve 60', the port e of the manual valve 59' and the drain port Ex in this state. With this operation, both the low and high clutches C L , C H are shut off as well as the same hydraulic pressure acts on the primary hydraulic servo 32 and secondary hydraulic servo 33 having the same pressure-receiving areas from the ratio control valve 57 so that approximately the same axial forces act on both the primary and secondary pulleys 7, 9, Note, when down-shift operation is executed, the down-shift relief valve 58 is switched to cause the port n to communicate with the drain port Ex, thus the hydraulic pressure of the hydraulic servo on which a high hydraulic pressure acts is made lower than the hydraulic pressure of the hydraulic servo on which a low hydraulic pressure acts. Next, a partially modified control mechanism will be described with reference to FIG. 14. The control mechanism 541 is applied to a so-called double chamber type control mechanism having the primary and secondary hydraulic servos 32, 33 shown in FIG 4, that is, the first and second hydraulic chambers 45, 46, 47, 49. Note, although a primary regulator valve 56, a ratio control valve 57, a down-shift relief valve 58 and a manual valve 59 are used which are similar to those of the embodiments shown in FIG. 5 to FIG. 7, a low/high control valve 601 is composed of a simple 2-position-switching solenoid valve with 4 ports. Then, a pilot oil passage 77a is branched from an oil passage 77 from the port y of the low/high control valve 601 to a high clutch servo C H and there is provided a switching valve 79 which is switched by a pilot pressure. In the low (L) mode in the D range, since the low/high control valve 601 is located at a low position, a port o communicates with a port p and a port y with a drain port Ex. Further, since the switching valve 79 is located at the illustrated position in the state that the oil passage 77 communicates with the drain port Ex, a port q with a port r, a port s with a port t, a port u with a port v and a port w with a port x. Note, the manual valve 59 is arranged in the D range position such that the ratio pressure from the ratio control valve 57 is introduced to an oil passage (2) and an oil passage (1) communicates with the drain port Ex. Therefore, the clutch pressure P c from the manual valve 59 is supplied to a low clutch hydraulic servo C L through the ports o and p of the low/high control valve 601 and the first and second line pressures P L 1, P L 2 from the manual valve are supplied to the first hydraulic chamber 45 of the primary hydraulic servo 32 and the first hydraulic chamber 46 of the secondary hydraulic servo 33 through the ports q, r, s, t of the switching valve 79. Further, the ratio pressure from the manual valve is supplied to the second hydraulic chamber 49 of the secondary hydraulic servo 33 through the oil passage (2) and the ports w and x of the switching valve 79 and the hydraulic pressure of the second hydraulic chamber 47 of the primary hydraulic servo 32 communicates with the drain port Ex of the manual valve through the ports u and v and the oil passage (1). In this state, the low clutch C L is connected as well as the axial force of the secondary pulley 9 is made higher than that of the primary pulley 7 by the ratio pressure acting on the second hydraulic chamber 49 of the secondary hydraulic servo 33 in accordance with the direction in which torque is transmitted in the CVT unit 11 likewise the above-mentioned. Further, in the high (H) mode in the D range, since the low/high control valve 60 1 is switched to the high (H) position, the port o communicates with the port y as well as the port p with the drain port Ex. Since the switching valve 79 is switched by the pilot pressure from the oil passage 77a in the state that the clutch pressure Pc is supplied to the oil passage 77 through the ports o and y, the port q communicates with the port t, the port s with the port r, the port x with the port v and the port w with the port u. Therefore, the clutch pressure P C is supplied to the high clutch hydraulic servo C H through the ports o and y of the low/high control valve 60 1 and the line pressures P L 1 and P L 2 are supplied to the first hydraulic chamber 46 of the secondary hydraulic servo 33 through the ports q and t of the switching valve 79 as well as to the first hydraulic chamber 45 of the primary hydraulic servo 32 through the ports s and r. Further, the ratio pressure from the manual valve 59 is supplied to the second hydraulic chamber 47 of the primary hydraulic servo 32 through the oil passage (2) and the ports w and u and the hydraulic pressure of the second hydraulic chamber 49 of the secondary hydraulic servo 33 is discharged from the drain port of the manual valve 59 through the ports x and v and oil passage (1). In this state, the high clutch C H is connected as well as the axial force of the primary pulley 7 caused by the primary hydraulic servo 32 in which the ratio pressure acts on the second hydraulic chamber 47 is made higher than that of the secondary pulley 9 in accordance with the direction in which torque is transmitted in the CVT unit 11. In the reverse (R) range, the manual valve 59 is switched and the low/high control valve 60 1 is maintained at the low (L) position as shown in FIG. 7. Therefore, the switching valve 79 is maintained at a lower position in the drawing in this state likewise the above-mentioned. Further, the ratio pressure is supplied to the oil passage (1) by switching the switching valve 79 as well as the oil passage (2) communicates with the drain port Ex. Therefore, the clutch pressure P C is supplied to the low clutch hydraulic servo C L through the ports o and p and the first and second line pressures are supplied to the first hydraulic chamber 5 of the primary hydraulic servo through the ports q and r as well as to the first hydraulic chamber 46 of the secondary hydraulic servo through the ports s and t. Further, the ratio pressure through the manual valve 59 is supplied to the second hydraulic chamber 47 of the primary hydraulic servo through the oil passage (1) and the ports v and u and the hydraulic pressure of the second hydraulic chamber of the secondary hydraulic servo is discharged from the drain port Ex of the manual valve 59 through the ports x and w and the oil passage (2). In this state, the low clutch C L is connected as well as the axial force of the primary pulley 7 caused by the primary hydraulic servo 32 in which the ratio pressure acts on the second hydraulic chamber 47 is made higher than that of the secondary pulley 9 in accordance with the direction in which torque is transmitted in the CVT unit 11. Next, another partially modified control mechanism will be described with reference to FIG. 15. The control mechanism 542 acts on a so-called double regulator type control mechanism having the primary and secondary hydraulic servos 32, 33 shown in FIG 9, that is, each one of a hydraulic chambers 73 and 75 having the same pressure-receiving areas for supplying different hydraulic pressures, respectively. Note, although a primary regulator valve 56, a ratio control valve (secondary regulator valve) 57, a down-shift relief valve 58 and a manual valve 59' are used which are similar to those of the embodiments shown in FIG. 10 to FIG. 12, a low/high control valve 60 1 is composed of a simple 2-position-switching solenoid valve with 4 ports as well as a switching valve 80 is provided which can switch a hydraulic pressure from the port y of the valve 60 1 to the high clutch hydraulic servo C H as a pilot pressure. In the low (L) mode in the D range, since the low/high control valve 60 1 is located at the low (L) position, the port o communicates with the port p and the port y with the drain port Ex. Further, the switching valve 80 is located at a position shown in the drawing in the state that the oil passage 77 communicates with the drain port Ex, the port q with the port t and the port s with the port r. Note, since the manual valve 59' is located at the D range position at the time, the relatively high line pressure P L - H from the primary regulator valve 56 is introduced to an oil passage (4) and the relatively low line pressure P L - L from the ratio control valve 57 is introduced to an oil passage (3). Therefore, the clutch pressure P C from the manual valve 59' is supplied to the low clutch hydraulic servo C L through the ports o and p of the low/high control valve 60 1 , the relatively low line pressure P L - L from the oil passage (3) is supplied to the primary hydraulic servo 32 through the ports q and t of the switching valve 80 and the relatively high line pressure P L - H from the oil passage (4) is supplied to the hydraulic servo 33 through the ports s and r. In this state, the low clutch C L is connected as well as the axial force of the secondary pulley 9 caused by the hydraulic servo 33 on which a somewhat higher line pressure acts is made higher than that of the primary pulley 7 in accordance with the direction in which torque is transmitted in the CVT unit 11 likewise the above-mentioned. Further, in the high (H) mode in the D range, since the low/high control valve 60 1 is switched to the high (H) position, the port o communicates with the port y as well as the port p with the drain port Ex. Since the switching valve 80 is switched by the pilot pressure from the oil passage 77a in the state that the clutch pressure P L is supplied to the oil passage 77 through the ports o and y, the port q communicates with the port r and the port s with the port t. Therefore, the clutch pressure P C is supplied to the high clutch hydraulic servo C H , the relatively low line pressure P L - L from the oil passage (3) is supplied to the secondary hydraulic servo 33 through the ports q and r of the switching valve 80 and the relatively high line pressure P L - H from the oil passage (4) is supplied to the primary hydraulic servo 32 through the ports s and t. In this state, the high clutch C H is connected as well as the axial force of the primary pulley 7 caused by the hydraulic servo 32 on which a somewhat high line pressure acts is made higher than that of the secondary pulley 9 in accordance with the direction in which torque is transmitted in the CVT unit 11. In the reverse (R) range, the manual valve 59' is switched and the low/high control valve 60 1 is maintained at the low (L) position as shown in FIG. 12. Therefore, the switching valve 80 is maintained at a lower position of the drawing in this state likewise the above-mentioned. In addition, the switching of the manual valve 59' causes the relatively high line pressure P L - H from the primary regulator valve 56 to be introduced to the oil passage (3) and the relatively low pressure line pressure P L - L to be introduced to the oil passage (4) from the ratio control valve 57. Further, the relatively low line pressure P L - L from the ratio control valve 57 is introduced to the oil passage (4). Therefore, the clutch pressure C L is supplied to the low clutch hydraulic servo C L through the ports o and p, the somewhat high line pressure P L - H from the oil passage (3) is supplied to the primary hydraulic servo 32 through the ports q and t and the somewhat low line pressure P L - L from the oil passage (4) is supplied to the secondary hydraulic servo 33 through the ports s and r. In this state, the low clutch C L is connected as well as the axial force of the primary pulley 7 caused by the hydraulic servo 32 on which a somewhat high line pressure P L - H acts is made higher than that of the secondary pulley 9 in accordance with the direction in which torque is transmitted in the CVT unit 11. Next, control of the continuously variable transmission of the embodiment will be described. FIG. 16 is a block diagram of an electronic control unit (ECU) 90, wherein numeral 91 denotes a sensor disposed to the continuously variable transmission 1 for detecting the r.p.m. of the input shaft 2 thereof, numeral 92 denotes a sensor for detecting the r.p.m. of the secondary pulley 9 of the CVT unit 11, numeral 93 denotes a vehicle speed sensor for detecting the r.p.m. of the output shaft 5 of the continuously variable transmission and numeral 94 denotes a similar vehicle speed sensor also used for a signal of a speed meter as well as used as a backup device when the vehicle speed sensor 93 fails. Numeral 95 denotes a sensor for detecting at which of the respective shift positions P. R. N. D the shift lever or the manual valve of the continuously variable transmission is located, numeral 96 denotes a sensor composed of a potentiometer disposed in the engine for detecting a throttle opening and numeral 97 denotes a sensor disposed in the throttle opening sensor 96 for detecting whether an accelerator is in an entirely closed state or not. The signals from the above respective sensors are fetched by a CPU, a ROM or a RAM through an input processing circuit and an input interface circuit, respectively. Numeral 101 denotes a solenoid for the low/high control valve 60 for switching to the low (L) mode and the high (H) mode and the solenoid executes ON/OFF operation. Numeral 102 denotes a solenoid used for the down-shift relief valve 58 for draining a high voltage side circuit and composed of a duty or linear solenoid. Numeral 103 denotes a solenoid used for the ratio control valve 57 for regulating a speed changing hydraulic pressure and composed of a duty or linear solenoid. Numeral 105 denotes a solenoid used for the primary regulator valve 56 for controlling a line pressure and composed of a linear solenoid. The above respective solenoids are driven through a solenoid drive circuit 106 for generating a predetermined voltage or output based on the signals from an output interface circuit, respectively and the operations of the respective solenoids are checked by a monitor circuit 107 to determine their failure as well as they determine their failure by themselves. Numeral 109 denotes an electronic control unit for controlling the engine and numeral 110 denotes a circuit for issuing a signal for temporarily reducing torque generated by the engine by delaying an ignition timing angle, shutting-off fuel supply etc, to ease shock caused in speed change. Numeral 111 denotes a processing circuit for inputting engine r.p.m. Numeral 112 denotes a checker composed of an indicator lamp and the like for outputting the result of self-diagnosis when the electronic control unit 90 fails and numeral 113 denotes a circuit for outputting the result of the self-diagnosis in the above failure. Numeral 115 denotes a display unit such as low (L) mode and high (H) mode display lamps and the like for displaying the state of the continuously variable transmission and numeral 116 denotes a drive circuit for driving the display unit 115. FIG. 17 is a general flowchart of the continuously variable transmission, wherein all the steps are initialized at the beginning of calculation at step S1. At step 2, r.p.m. is calculated based on the signals from the above r.p.m. sensors 91, 92 and 93. The signal from a neutral start switch is processed and a shifted position is read at step S3 and a throttle opening is calculated at step S4. Then, whether the manual valve is located in the reverse (R) range or not is determined based on the above step S3 at step S5, and when the determination is YES, an R range control to be described later will be executed (step S6). When the determination is NO, whether the D range is selected or not is determined (step S7), and when the D range is selected, it is determined whether or not a current vehicle speed V is equal to or less than a predetermined vehicle speed V1 (for example, 5 Km/h) (step S8). When the current vehicle speed is equal or less than the vehicle speed, start control is executed (step S9), whereas it is greater than the predetermined vehicle speed, line pressure control is executed to set a line pressure depending upon whether the low (L) mode is selected or the high (H) mode is selected, the throttle opening and the pulley ratio (step S10). Further, a target engine r.p.m. is read and determined from the map of a preset optimum fuel consumption curve based on the throttle opening at step S11. Then, which of down-shift and up-shift must be selected or otherwise the current state must be maintained is determined by comparing the current engine r.p.m. with the above target engine r.p.m. (step S12) and the down-shift control (step S13) or the up-shift control (step S14) to be described later will be executed respectively. FIG. 18 is a flowchart showing a subroutine of the reverse (R) range control, wherein whether a vehicle travels forward or not is determined first from the pulley ratio and vehicle speed at step S15. When the vehicle travels forward, whether the current vehicle speed V is greater than a predetermined vehicle speed V2 (for example, 5 Km/h) or not is determined (step S16). When the current vehicle speed is greater than the predetermined vehicle speed, control for prohibiting reverse operation is executed by, for example, the release of both the high clutch C H and low clutch C L , or the like (step S17), whereas when the vehicle speed is smaller than the predetermined vehicle speed, neutral (N) control is executed (step S18). In the neutral control, the axial forces of the primary pulley 7 and secondary pulley 9 are controlled to be substantially equalized. Otherwise, at least the difference between the axial force of the primary pulley and the axial force of the secondary pulley is set to a value smaller than the difference of the axial forces of both the pulleys which is determined by the input torque of the CVT and the pulley ratio at the time when output torque is in a positive direction within the range in which the relative relationship in magnitude of the axial forces is not reversed or to a value smaller than the difference of the axial forces of the primary and secondary pulleys which is determined by the input torque of the CVT unit and the pulley ratio at the time when the output torque is in a negative direction within the range in which the relative relationship in magnitude of the axial forces is not reversed. More specifically, in the double chamber type control mechanisms 54, 54 1 shown in FIG. 4 to FIG. 7 and FIG. 14, the hydraulic pressures in both the second hydraulic chambers 47, 49 of both the primary and secondary hydraulic servos 32, 33 are released in the state that hydraulic pressures are supplied to both the first hydraulic chambers 45, 46 of the primary and secondary hydraulic servos 32, 33 so as to equalize the axial forces of both the pulleys 7 and 9. Further, in the double regulator type control mechanisms 54' and 542 shown in FIG. 9 to FIG 12 and FIG. 15, equal hydraulic pressures are supplied to both the primary and secondary hydraulic servos 32 and 33 by entirely opening the ratio control valve (secondary regulator valve) 57 so as to equalize the axial forces of both the pulleys 7 and 9. With these operations, the CVT unit 11 is converged by itself so that the output from the output shaft 5 is made to zero regardless of the r.p.m. of the output shaft and stably maintained at the neutral position where the r.p.m. of the output shaft 5 is made to zero. Note, the aforesaid neutral control is described in the specification and drawings of Japanese Patent Application Laid-Open No. 7-66234 in detail and the content thereof is applicable to this embodiment as it is. When it is determined that the vehicle travels backward at step S15, the current vehicle speed V is compared with the predetermined vehicle speed V1 (for example, 5 Km/h) (step S19) and when the current vehicle speed V is smaller than the predetermined vehicle speed V1, start control is executed (step S20), whereas when the current vehicle speed V is greater than the predetermined vehicle speed V1, line pressure control is executed to set a line pressure based on the throttle opening and pulley ratio (step S21). Further, a target engine r.p.m. is determined by reading, the map likewise the above-mentioned (refer to step S10) at step S22. Then, any of up-shift, down-shift and the maintenance of the current state is determined likewise the above step S12 (step S23) and the down-shift control (step S24) or the up-shift control (step S25) to be described later will be executed, respectively. Note, the CVT unit 11 is located at the O/D end and the speed increasing rotation transmitted from the CVT unit to the sun gear 19s is synthesized with the constant rotation transmitted from the power transmission unit 26 to the carrier 19c by the planetary gear 19 and a reverse rotation is taken out from the output shaft 5 integrally arranged with the ring gear 19r in the above reverse (R) control. At the time, however, the torque output from the CVT unit 11 is reversed across the neutral position (output from the output shaft 5 is zero) based on the torque circulation and power is transmitted from the primary pulley 7 to the secondary pulley 9 in the CVT unit 11 as shown in FIG. 2. In the reverse control, the manual valves 59, 59' are located at the reverse (R) position as shown in FIG. 7, FIG. 8, FIG. 12, FIG. 13, FIG. 14, FIG. 15 and the hydraulic pressure of the hydraulic servo 32 is set higher than that of the hydraulic servo 33 when positive torque is transmitted (transmitted to the engine and wheels). Consequently, the axial force Fpr of the primary pulley 7 is made higher than the axial force Fsr of the secondary pulley 9 as shown in FIG. 23 so that the relationship of axial forces in coincidence with the aforesaid torque transmission direction is established. In this state, the axial forces Fpr and Fsr of both the pulleys 7, 9 are properly changed so as to properly change a speed by the proper control of the ratio control valve 57. FIG. 19 is a flowchart showing up-shift control in the D range shown at step S14, wherein it is determined whether a speed is to be changed from the low (L) mode to the high (H) mode to be described later at step S26. Note, R range control is not executed at step S26. Further, the difference between a current engine r.p.m. and a target engine r.p.m. and the alteration of the difference are calculated (step S27) and a speed changing speed is calculated based on the calculation (step S28), whereby the application control of the ratio control valve 57 is executed (step S29). FIG. 20 is a subroutine showing the determination of Lo → Hi at the above step S26. At step S30, a pulley ratio is calculated based on the signals from the r.p.m. sensors 91, 92 of the primary pulley 7 and secondary pulley 9. Then, it is determined whether the pulley ratio is within the range in which a speed change from the low mode to the high mode is permitted (step S31) and when the pulley ratio is within the range, the low/high control valve 60 is switched to thereby change the engagement with the low clutch C L to the engagement with the high clutch C H (step S32). FIG. 21 is a flowchart showing down-shift control shown at step S13, wherein it is determined at step S33 whether a speed change from the high (H) mode to the low (L) mode to be described later is to be executed or not (this is not executed in the R range control). Further, the difference between a current engine r.p.m. and a target engine r.p.m. and the alteration of the difference are calculated (step S34) and a speed changing speed is calculated based on the calculation (step S35). Further, whether down-shift is power-off down-shift or not is determined based on the signals from the throttle opening sensor 96 and idle switch 97 (step S36). Then, when a coast state is achieved by power-off, that is, by releasing depression of an accelerator pedal, or the like, the ratio control valve 57 is controlled as well as the down-shift relief valve 58 is controlled (step S37), whereas when down-shift is executed in a power- on state, that is, in an accelerated state in which the accelerator pedal is abruptly depressed, the ratio control valve 57 is controlled (step S38). That is, the coast state is in a negative torque state in which torque is transmitted from the wheels to the engine as shown at step S37. In the negative torque state, however, the relative relationship in magnitude between the axial force Fp' (Fpr') of the primary pulley 7 and the axial force Fs' (Fsr') of the secondary pulley 9 is reversed from the axial forces Fp, Fs (Fpr, Fsr) in a positive torque (engine wheels) state shown by the solid lines of FIG. 23, as shown by the dash lines of the drawing. At the time, since the down-shift relief valve 58 is drained in the double chamber type control mechanism 54 shown in FIG. 5, the first hydraulic chamber 46 of the secondary hydraulic servo 33 is released, whereby the axial force Fp' of the primary pulley 7 is made higher than the axial force Fs' of the secondary pulley 9, thus the speed change of the CVT unit 11 is executed by the proper adjustment of the ratio control valve 57 in this state. Further, in the double regulator type control mechanism 54' shown in FIG. 10, since the down-shift relief valve 58 is drained, the secondary hydraulic servo 33 is released, whereby the axial force Fp' of the primary pulley is made higher than the axial force Fs' of the secondary pulley. Note, although an oil supply passage to the primary hydraulic servo 32 is reversed from that to the secondary hydraulic servo 33 based on the manual valve in the reverse (R) control shown in FIG. 7 and FIG. 11, the pressure of the hydraulic servo having a higher hydraulic pressure is reduced so that the axial force Fsr' of the secondary pulley is made higher than the axial force Fpr' of the primary pulley. Further, down shift control executed at step S38 when power is turned on will be described. Since the ratio control valve 57 is controlled in a pressure reducing direction in the double chamber type control mechanism 54 shown in FIG. 5, the hydraulic pressure of the secondary hydraulic servo 33 is made relatively smaller than that of the primary hydraulic servo 32, whereas since the ratio control valve 57 is controlled in a pressure increasing direction in the double regulator type control mechanism 54' shown in FIG. 10, the hydraulic pressure of the secondary hydraulic servo 33 is made relatively smaller than that of the primary hydraulic servo 32. With this operation, since the ratio of the axial force FS of the secondary pulley and the axial force FP of the primary pulley is reduced, the speed of the CVT unit 11 is abruptly changed in a speed increasing (O/D) direction and thus the speed of the transmission 1 is abruptly changed in a speed reducing (U/D) direction as a whole. FIG. 22 shows a subroutine for determining Hi → Lo at the above step S33. A current pulley ratio is read based on the signals from the r.p.m. sensors of the primary pulley 7 (= input shaft 3) and secondary pulley 9 (step S40) and it is determined whether the pulley ratio is within the range in which a speed change is permitted from the high mode to the low mode (step S41). When the pulley ratio is within the range, the low/high control valve 60 is switched so as to change the engagement from the high clutch C H to the low clutch C L (step S42). The speed is changed between the low (L) mode and high (H) mode by the change of engagement of the low clutch C L and high clutch C H executed at the above steps S32 and S42. In the low mode, a positive rotation is taken out from the output shaft 5 by the aforesaid synthesization of both the rotations of the power transmission unit 26 and CVT unit 11 executed by the planetary gear 19. At the time, power is transmitted from the secondary pulley 9 to the primary pulley 7 by the torque circulation in the CVT unit 11. On the other hand, in the high mode, since the rotation output from the CVT unit 11 is transmitted to the output shaft 5 as it is, power is transmitted from the primary pulley 7 to the secondary pulley 9 in the CVT unit 11. At the time, the hydraulic pressures of the primary hydraulic servo 32 and secondary hydraulic servo 33 are switched by switching the low/high control valves 60, 60' and 60 1 as shown in FIG. 6, FIG. 11, FIG. 14 and FIG. 15. With this operation, the axial force FS of the secondary pulley 9 is made higher than the axial force FP of the primary pulley 7 in the low mode and the axial force of the primary pulley is made higher than that of the secondary pulley in the high mode in order to coincide with the above direction of power transmission. Note, the above description relates to the transmission of positive torque and this is reversed in the transmission of negative torque by the down-shift relief valve as described above. Next, a modified example of the power transmission mechanism of the continuously variable transmission will be described with reference to FIG. 24. FIG. 24 shows an embodiment using a dual pinion planetary gear 19. The ring gear 19r of the planetary gear can be coupled with the input shaft 3 through gears 130, 131 and the low clutch C L (first rotary element) the sun gear 19s thereof can be coupled with the secondary pulley 9 through gears 132, 133 (second rotary element), the carrier 19c thereof can be coupled with the output shaft 5 (third rotary element) and the ring gear 19r can be coupled with the sun gear 19s through the high clutch C H . Then, the ring gear 19r and sun gear 19s are rotated in the same direction. Note, the third rotary element may be composed of the sun gear in place of the carrier and the first and second rotary elements may be composed of the carrier, the sun gear and the ring gear which do not serve as the third rotary element, respectively. In this embodiment, the engine torque from the input shaft 3 is transmitted to the ring gear 19r of the planetary gear 19 through the low clutch C L and gears 131, 130 as well as to the sun gear 19s through the CVT unit 11 and gears 133, 132 in the low mode (Lo) as shown in the graph showing a speed of FIG. 24(b). At the time, torque circulation is caused by the planetary gear 19 likewise the above embodiment and an output is applied from the carrier 19c to the output shaft 5 in an increasing/decreasing direction opposite to that of the CVT unit 11 with the secondary pulley 9 serving as a driving side and the primary pulley 7 serving as a driven side in the CVT unit 11. Further, when the CVT unit 11 is in a predetermined speed increasing (O/D) state, the rotation of the carrier 19c is reversed, that is, made to a reverse (Rev) rotation as well as the direction in which torque is transmitted from the CVT unit 11 is reversed. In addition, in the high mode (Hi), the engine torque from the output shaft 5 is transmitted to a shaft 135 through the CVT unit 11 and gears 132, 133 as well as the planetary gear 19 is integrally rotated by the connection of the high clutch C H in the high mode, thus the speed changing torque of the above shaft 135 is transmitted to the output shaft 5 as it is in the CVT unit. In FIG. 25, the sun gear 19s of a single pinion planetary gear 19 is coupled with the input shaft 3 through the chain 22 and the low clutch C L (first rotary element), the ring gear 19r thereof is coupled with the secondary pulley 9 through the gears 132, 133 (second rotary element) and the carrier 19c thereof is coupled with the differential gear unit 25 through the output shaft 5 and the gear 21 and gear 24 (third rotary element). Then, the sun gear 19s and the carrier 19c are rotated in a reverse direction. Note, the first rotary element may be composed of the ring gear and the second rotary element may be composed of the sun gear. In this embodiment, the engine rotation from the input shaft 3 is transmitted to the sun gear 19s of the planetary gear 19 through the low clutch C L and chain 22 as well as to the ring gear through the CVT unit 11 and gears 133, 132 and both the rotations are synthesized and taken out from the output shaft 5 through the carrier 19c as shown in the graph showing a speed of FIG. 25(b). At the time, torque circulation is caused likewise the above-mentioned as well as the rotation is output from the output shaft 5 with the positive to reverse change of its rotational direction. In FIG. 26, a dual pinion planetary gear 19 is used by modifying the embodiment shown in FIG. 25. That is, the sun gear 19s of the dual pinion planetary gear 19 is coupled with the input shaft 3 through the chain 22 and low clutch C L (first rotary element), the carrier 19c thereof is coupled with the secondary pulley 9 through the gears 132, 133 (second rotary element), the ring gear 19r thereof is coupled with the differential gear unit 25 through the output shaft 5 and gears 21, 24 (third rotary element). Then, the sun gear 19s and carrier 19c are rotated in a reverse direction. Note, the first rotary element may be composed of the carrier and the second rotary element may be composed of the sun gear. In this embodiment, the engine rotation from the input shaft 3 is transmitted to the sun gear 19s of the dual planetary gear 19 through the low clutch C L and chain 22 as well as to the carrier 19c through the CVT unit 11 and gears 133, 132, then both the rotations are synthesized and taken out from the output shaft 5 through the carrier 19c as shown in the graph showing a speed of FIG. 26(b). At the time, torque circulation is caused likewise the above-mentioned as well as the rotation is output from the output shaft 5 with the positive to reverse change of its rotational direction.

Claims: A continuously variable transmission (1) comprising: an input shaft (3) associated with an engine output shaft (2): an output shaft (5) associated with wheels: a belt type continuously variable transmission unit (11) having a first pulley (7) associated with said input shaft (3), a second pulley (9), a belt (10) trained around both the pulleys (7, 9) and axial force actuating means (32, 33) for applying axial forces to both the pulleys (7, 9) for changing the pulley ratio of said first and second pulleys; a planetary gear (19) having at least first, second and third rotary elements (19s, 19c, 19r) with said first rotary element associated with said input shaft (3), said second rotary element associated with said second pulley (9) and said third rotary element associated with said output shaft (5), respectively; selection means for selecting at least a forward traveling state and a backward traveling state of a vehicle, wherein a torque transmitting direction is changed between said first and second pulleys (7, 9) said belt type continuously variable transmission unit (11) as well as the direction of torque output from said output shaft (15) is changed, based on switching between said forward traveling state and said backward traveling state of said selection means; control means (54) for controlling said axial force actuating means (32, 33) so that axial forces acting on said first and second pulleys (7, 9) produce a difference corresponding to the pulley ratio: and switching means (59) for changing the relative relationship in magnitude between the axial forces which are caused to act on said first and second pulleys (7, 9) by said control means (54) based on switching between said forward traveling state and said backward traveling state of said selection means so that said relative relationship in magnitude is reversed, wherein: said axial force actuating means (32, 33) has a first hydraulic servo (32) acting on said first pulley (7) and a second hydraulic servo (33) acting on said second pulley (9) and applies axial forces to said first and second pulleys (7, 9) by supplying hydraulic pressures from an oil pump (17) to said first and second hydraulic servos (32, 33); and said switching means (59) is a switching valve for switching the hydraulic pressure acting on said first hydraulic servo (32) and the hydraulic pressure acting on said second hydraulic servo (33), characterized in that said switching valve is a manual valve (59) integrally associating with said selection means. A continuously variable transmission (1) according to claim 1, wherein: each of said first and second hydraulic servos (32, 33) has a plurality of hydraulic chambers (45, 47, 46, 49): and said switching valve (59) selectively switches said hydraulic chambers to which hydraulic pressures are supplied to thereby reverse the effective pressure-receiving areas of said first and second hydraulic servos (32, 33). A continuously variable transmission (1) according to claim 2, wherein: each of said first and second hydraulic servos (32, 33) has at least a first hydraulic chamber (45, 46) and a second hydraulic chamber (47, 49), respectively as well as the first hydraulic chambers (45, 46) of said both hydraulic servos (32, 33) have the same effective pressure-receiving areas: said control means (54) has a regulator valve (56) and a ratio control valve (57), the hydraulic pressure from said regulator valve (56) is supplied to the first hydraulic chambers (45, 46) of said both hydraulic servos (32, 33) at all times as well as the hydraulic pressure from said ratio control valve (57) is supplied to the second hydraulic chamber (47, 49) of any one of said both hydraulic servos (32, 33); and said switching valve (59) reverses the communication of said ratio control valve (57) with any one of said second hydraulic chambers (32, 33) to the other of them. A continuously variable transmission (1) according to claim 3, wherein said regulator valve (56) and said ratio control valve (57) are interposed between said oil pump (17) and said switching valve (59). A continuously variable transmission (1) according to claim 1, wherein: said control means (54) has first and second regulator valves (56, 57) for adjusting the hydraulic pressure from said oil pump (17) to respective hydraulic pressures, any one of said first and second regulator valves (56, 57) is caused to communicate with any one of said first and second hydraulic servos (32, 33) and the other of said regulator valves (56, 57) is caused to communicate with the other of said hydraulic servos (32, 33), respectively; and said switching valve (59) reverses the communication of said first and second regulator valves (56, 57) with said first and second hydraulic servos (32, 33). A continuously variable transmission (1) according to claim 5, wherein said first and second regulator valves (56, 57) are interposed between said oil pump (17) and said switching valve (59). A transmission (1) according to any of claims 1 to 6, comprising a relief valve (58) which makes the hydraulic pressure of said hydraulic servo (32, 33) on which a high hydraulic pressure acts lower than the hydraulic pressure of said hydraulic servo (32, 33) on which a low hydraulic pressure acts when it is determined that said vehicle is in a coast state. A transmission (1) according to any of claims 1 to 7, wherein said input shaft (3) is coupled with said first pulley (7) and said first rotary element, respectively so that the direction in said input shaft (3) of the torque transmitted from said first pulley (7) and the direction in said input shaft (3) of the torque transmitted from said first rotary element are reversed each other, respectively.

Inventors: HATTORI MASASHI, INUZUKA TAKESHI, MIYAGAWA SHOICHI, MOROTO SHUZO, SAKAKIBARA SHIRO, TANIGUCHI TAKAO, TSUKAMOTO KAZUMASA, HATTORI, MASASHI, INUZUKA, TAKESHI, MIYAGAWA, SHOICHI, MOROTO, SHUZO, SAKAKIBARA, SHIRO, TANIGUCHI, TAKAO, TSUKAMOTO, KAZUMASA
Applicants: AISIN AW CO, AISIN AW CO., LTD.
Application year: 1996
Main classification: F16H61
IPC codes: F16H61, F16H37