JPH03194244A - Continuously variable transmission - Google Patents

Abstract

PURPOSE:To increase durability of a belt by setting initial alignment in such a way that deviation becomes zero when gear change ratio is the maximum and transmission torque is at the maximum load when alignment in the axial direction between drive and follow-up pulleys is set. CONSTITUTION:When alignment in the axial direction between a drive pulley 41 and a follow-up pulley 42 of a continuously variable transmission 4 is set, initial alignment is set in such a way that deviation becomes zero when gear change ratio is the maximum and transmission torque of a belt 43 is at the maximum load. Since deviation in alignment in the axial direction between the drive pulley 41 and the follow-up pulley 42 when high load is applied becomes zero, the torsional stress which is applied on the belt 43 is reduced even if high load is applied. Thus, it is possible to improve durability and reliability of the belt.

Description

DETAILED DESCRIPTION OF THE INVENTION (Industrial Application Field) The present invention relates to an improvement in a belt type (pulley type) continuously variable transmission in which the gear ratio can be changed continuously using hydraulic pressure.

(Prior art) Conventionally, as a belt type continuously variable transmission, for example,
As disclosed in Japanese Patent Publication No. 2-4958, it includes driving and driven pulleys with variable effective radii, a belt wound between the two pulleys, and a gear ratio control valve and a line pressure regulating valve. The gear ratio control valve supplies and discharges hydraulic pressure to and from the hydraulic cylinder of the drive pulley to continuously adjust its effective radius, thereby continuously variable control of the gear ratio, and a line pressure regulating valve. Line pressure is supplied to the hydraulic cylinder of the driven pulley, and the magnitude of the supplied line pressure is adjusted to control the tension of the belt, thereby suppressing or preventing rotational slippage of the belt. It has been known.

(Problem to be Solved by the Invention) In this type of continuously variable transmission, in order to increase the durability of the belt and ensure its reliability, it is necessary to improve the axial alignment between the driving and driven pulleys. It is necessary to set the deviation to zero. That is, if the alignment between both pulleys is out of alignment, an inappropriate stress in the torsion direction is applied to the belt, which adversely affects durability. For this reason, usually
Setting is performed so that the above-mentioned alignment deviation becomes zero when the speed ratio, which is the ratio of the rotation speed of the driving pulley to the rotation speed of the driven pulley, is at its maximum (at low speed).

However, for example, when hydraulic pressure acts on a hydraulic cylinder that opens and closes the drive and driven pulleys, or in a structure in which helical gears are connected to the rotational shafts of the drive and driven pulleys, the rotational shafts will change as the helical gears mesh. When an axial load is applied to the rotary shaft, the two rotating shafts undergo relative displacement in the axial direction, albeit slightly. At that time, due to deformation of the case supporting the rotary shaft, play of the bearing between the two, etc., the relative displacement between the two rotary shafts is expanded, and it is inevitable that the misalignment between the two pulleys becomes large.

If the torque transmitted between both pulleys is large, the belt will be twisted under high tension, which will have a considerable negative effect on its reliability.

The present invention has been made in view of the above, and its purpose is to improve the setting conditions for the axial alignment between the driving and driven pulleys, thereby making it possible to improve the alignment when maximum torque is applied between both pulleys. Make sure there is no misalignment,
Therefore, the objective is to improve the durability and reliability of the belt.

(Means for Solving the Problem) In order to achieve the above object, the invention according to claim (1) assumes in advance the axial alignment deviation between the driving and driven pulleys in the maximum load state and corrects the deviation. The initial positions of both pulleys were set so that the deviation was zero.

That is, the present invention includes drive and driven pulleys that are variable pulleys with variable effective radii, and a belt wound between both pulleys, so that the speed ratio between the drive and driven pulleys can be continuously adjusted steplessly. For a continuously variable transmission that is configured to change the initial axial alignment between both pulleys, the shift in axial alignment between both pulleys becomes zero when the gear ratio is maximum and the transmission torque is maximum. The feature is that the alignment is set in advance in a direction opposite to the direction in which the alignment changes.

(Function) With the above configuration, in the present invention, the gear ratio is maximum,
In addition, the axial alignment between the drive and driven pulleys changes at high loads when the transmitted torque is at its maximum, but the initial alignment is set to deviate in the opposite direction to the direction in which the alignment changes at high loads. Therefore, the misalignment becomes zero at the time of the above-mentioned high load. Therefore, even under high loads, the torsional stress applied to the belt is reduced, thereby improving its durability and reliability.

(Example) Embodiments of the present invention will be described below with reference to the drawings.

FIG. 1 shows the overall structure of a continuously variable transmission according to an embodiment in which the present invention is applied to an automobile. The transmission shown in the figure basically includes a torque converter 2 connected to an output shaft 1 of an on-vehicle engine 1, a forward/reverse switching mechanism 3, a continuously variable transmission mechanism 4, a reduction mechanism 5, and a differential mechanism. 6 (differential mechanism).

As shown in enlarged detail in FIG. 2, the torque converter 2 includes a pump cover 2 coupled to the engine output shaft 11.
1, and a pump impeller 22 that is fixed to one side of the pump cover 21 and rotates integrally with the engine output shaft 11.
, a turbine runner 23 rotatably provided inside the pump cover 21 to face the pump impeller 22 , and the turbine runner 23 and the pump impeller 22
a stator 24 that is interposed between the
and a turbine shaft 25 fixed to the turbine launch 23. The stator 24 is connected to the transmission case 7 via a one-way clutch 26 and a cylindrical stator shaft 27.

A lock-up piston 28 slidably attached to the turbine shaft 25 is provided between the turbine runner 23 and the pump cover 21, and lock-up fastening chambers 29 are formed on both sides of the lock-up piston 28.
The lock-up piston 28 and the pump cover 21 are connected and opened by introducing and discharging hydraulic pressure into the lock-up opening chamber 29b and the lock-up opening chamber 29b.

The forward/reverse switching mechanism 3 includes a ring gear 35 that is spline-coupled to the turbine shaft 25 of the torque converter 2, and a pinion 32 that meshes with the ring gear 35. 32.・
...and the binion 32. 32°..., and a sun gear 34 spline-coupled to a primary shaft 4]1 of a continuously variable transmission mechanism 4, which will be described later, and meshing with the pinion gear 32. A forward clutch 36 is disposed between the ring gear 35 and the carrier 31 to connect and disconnect them, and a forward clutch 36 is provided between the carrier 3 and the transmission case 7 to selectively fix the carrier 31 to the transmission case 7. A reverse brake 37 is provided. When the forward clutch 36 is engaged and the reverse brake 37 is released, the ring gear 35 and carrier 31 are connected to the rotating body, and the rotation of the turbine shaft 25 is directly transmitted to the continuously variable transmission mechanism 4. On the other hand, when the reverse brake 37 is engaged and the forward clutch 36 is released, the carrier 31 is fixed to the case 7 so as not to rotate, and the rotation of the ring gear 35 is transmitted to the pinion gear 32. 32. ... to the sun gear 34 to reverse the rotation of the turbine shaft 25 and transmit it to the primary shaft 411 of the continuously variable transmission mechanism 4. Also, a forward clutch 36 and a reverse brake 37
When both are opened, a neutral state and a parking state are established in which the driving force of the engine is not transmitted from the turbine shaft 25 to the primary shaft 41 of the continuously variable transmission mechanism 4.

The continuously variable transmission mechanism 4 includes a primary pulley 41 as a driving pulley, a secondary pulley 42 as a driven pulley, and these pulleys 4].

42 and a V-belt 43 wound around between the belts 42 and 42. The primary pulley 41 includes a primary shaft 411 as a drive shaft disposed coaxially with the turbine shaft 25, a fixed sheave 412 rotatably fixed to the primary shaft 411, and a fixed sheave 412 disposed opposite to the fixed sheave 412. The V-belt 43 has a movable sheave 413 rotatably and slidably supported by the primary shaft 411, and as the movable sheave 413 moves, the pitch line of the V-belt 43 changes to change the effective pitch diameter (effective radius). ) is set to change. That is, when the movable sheave 413 approaches the fixed sheave 412, the effective pitch diameter becomes large, and conversely, when the movable sheave 4]-3 moves away from the fixed sheave 412, the effective pitch diameter becomes small.

On the other hand, the secondary pulley 42 basically has the same configuration as the primary pulley 41 described above.

In other words, a secondary shaft 42 as a driven shaft arranged parallel to the primary shaft 411 and this secondary shaft 4
21 and a movable sheave 423 that is slidably supported.
3 changes the effective pitch diameter.

Each movable sheave 413 in each of these pulleys 41 and 42
, 423 are provided with movable sheaves 413, 4, respectively.
Hydraulic cylinders 414 and 424 for sliding the 23 are provided. To specifically explain the configuration of the hydraulic cylinder 414 of the primary pulley 41, the hydraulic cylinder 414 is a first cylinder formed on the back side of the movable sheave 413.
A connecting member between the oil chamber 14a and the outer end of the movable sheave 413]
Movable pistons 1 and 4 are movably connected together via 4b.
c, and a second oil chamber 1.4d formed on the back side of the movable piston 14c, the first and second oil chambers 14a,
, ], and 4d to move the movable sheave 413 to the right in FIG. 2.

On the other hand, the configuration of the hydraulic cylinder 424 of the secondary pulley 42 is similar to that described above, and is connected to the first oil chamber 15a formed on the back side of the movable sheave 423 and the outer end of the movable sheave 423 via the connecting member 15b. movable piston 15c
, a second oil chamber 15d formed on the back side of the movable piston 15c, and a spring 15e compressed within the second oil chamber 15d to bias the movable piston 15c leftward in FIG. , the movable sheave 423 is moved to the left in FIG. 2 by supplying hydraulic pressure to the first and second oil chambers 15a and 15d.

The hydraulic cylinder 41 of the primary pulley 41
The total pressure receiving area of the first oil chamber 14a and the second oil chamber 14d of
The area is set to be approximately twice the total pressure receiving area of the oil chamber 15a and the second oil chamber 15d, and when hydraulic pressure is introduced into the hydraulic cylinder 414 of the primary pulley 41, the movable sheave 4
13 moves to the right in FIG. 2 and approaches the fixed sheave 4]2, the pitch line of the V belt 43 in the primary pulley 41 moves toward the outer circumferential side, and the effective pitch diameter of the primary pulley 4] increases. Due to the accompanying displacement of the V-belt 43, the movable sheave 423 of the secondary pulley 42 moves to the right in FIG. As the secondary pulley 42 moves, the effective pitch diameter of the secondary pulley 42 becomes smaller, and the gear ratio between the primary and secondary shafts 411 and 421 changes to a smaller value (in the direction of speed increase). Conversely, when the hydraulic pressure is discharged from the hydraulic cylinder 414, the movable sheave 413 of the primary pulley 41 moves to the left in FIG. 2 and separates from the fixed sheave 412, and its effective pitch diameter becomes smaller. The movable sheave 423 of the secondary pulley 42 also moves to the right in FIG. 2 and approaches the fixed sheave 422, and its effective pitch diameter increases, causing the primary and secondary shafts 411, . The gear ratio between 421 and 421 changes greatly (in the direction of deceleration).

Further, the speed reduction mechanism 5 and the differential mechanism 6 have a known structure,
Although not described in detail here, the rotation of the secondary shaft 421 is transmitted to the axle 61.

Next, the torque converter 2 in the above-mentioned continuously variable transmission
, the forward clutch 36 and reverse brake 37 of the forward/reverse switching mechanism 3, and the primary pulley 41 and secondary pulley 4 of the continuously variable transmission mechanism 4.
A hydraulic circuit for controlling each operation of the motor 2 will be explained based on FIG.

The hydraulic circuit shown in the figure includes an oil pump 8 driven by the engine 1. The hydraulic oil discharged from the oil pump 8 is first adjusted to a predetermined line pressure at the line pressure regulating valve 82 and then sent to the hydraulic cylinder 4 of the secondary pulley 42 via the line 101.
24, and is finally supplied to the hydraulic cylinder 4]4 of the primary pulley 41 via a line 102 branched from line ]01.

The line pressure regulating valve 82 has a spool 8 composed of a main spool 821 and a sub spool 822 arranged in series.
It has 20. A main spool 821 and a sub-spool 822 constituting the spool 820 are connected such that one end of the sub-spool 822 is brought into contact with one end of the main spool 821. At the other end of the sub spool 822,
A large diameter portion 822a is provided which has a larger cross-sectional area than the contact area (cross-sectional area of the connecting portion) with the main spool 821. At a position corresponding to the center of the main spool 821 is a pressure regulating port 82 through which oil discharged from the oil pump 81 is introduced.
3 and a drain port 824 that communicates with the suction side of the oil pump 81 are installed, and when the main spool 82 moves to the left side in the figure, the connection between the pressure regulating port 823 and the drain port 824 is cut off, and the main spool 82 is closed off. When the spool 821 moves to the right side in the figure, the pressure regulating port 823 and the drain port 824 are cut off, and the main spool 821
On the other hand, when the pressure adjustment port 823 and the drain port 824 are moved to the right side in the figure, the pressure regulation port 823 and the drain port 824 are communicated with each other. A first pilot chamber 825 is formed at a position corresponding to the connecting portion between the main spool 821 and the sub spool 822.
A spring 826 is interposed in the pilot chamber 825 to bias the main spool 821 to the left in the figure. Also,
A second pilot chamber 827 communicating with the first pilot chamber 825 is formed in the large diameter portion 822a of the sub spool 822. The first pilot chamber 825 and the second pilot chamber 827 are supplied with hydraulic oil that is branched from the line 102 and then reduced to a predetermined pressure by the reducing valve 83 while passing through the line 103.
The pilot pressure is adjusted by the first duty solenoid valve 91 while passing through the air. While this pilot pressure acts in the same direction as the biasing force of the spring 826,
The hydraulic pressure in the line 101 acts on the other end of the main spool 821 to counteract the biasing force and pilot pressure, and the spool 820 moves due to the force relationship between the pressure regulating port 823 and the drain port 824. By communicating and cutting off the line pressure, the line pressure is controlled to a value corresponding to the pilot pressure regulated by the first duty solenoid valve 91.

A speed ratio control valve 85 is provided in the line 102 . This gear ratio control valve 85 includes a spool 85] and a spring 8 that biases this spool 851 in the right direction in the figure.
52, a line pressure port 853 connected to the upstream part of the line 102, a drain port 854, and a spring 85.
2 to the installation side and connected to the shift valve 8 via line 104.
Reverse port 855 connected to 7 and spring 8
A pilot chamber 856 is formed on the opposite side of the installation side of 52 and into which pilot pressure is introduced. The pilot chamber 856 is connected to a second duty solenoid valve 92 via a pitot valve 86, and a pitot pressure generating means 9 that generates a pitot pressure corresponding to the rotational speed of the engine 1.
Connected to 0. Therefore, the pitot pressure generated by the pitot pressure generating means 90 and the pressure adjusted by the second duty solenoid valve 92 can be selectively introduced into the pilot chamber 856 as pilot pressure by the pitot valve 86. Even if the second duty solenoid valve 92 fails, the pitot pressure can be introduced from the pitot pressure generating means 90 into the pilot chamber 856 as a pilot pressure.

When the gear ratio control valve 85 is moving forward (when the shift valve 87 is in any of the shift positions 2.1)
Since hydraulic pressure is drained from the reverse port 855 via the shift valve 87, the spool 851 moves due to the force relationship between the pilot pressure introduced into the pilot chamber 856 and the biasing force of the spring 852, and the line pressure port 853 and drain port 854 are the primary pulley 41
6 of the hydraulic cylinders 414. In this way, when moving forward, the pilot chamber 85
The gear ratio between the primary pulley 41 and the secondary pulley 42 of the continuously variable transmission mechanism 4 is varied by controlling the supply and discharge of hydraulic pressure to the hydraulic cylinder 414 of the primary pulley 41 according to the pilot pressure introduced into the hydraulic cylinder 414 of the continuously variable transmission mechanism 4. It is configured to adjust to.

On the other hand, when traveling in reverse (when the shift valve 87 is in the R shift position), hydraulic pressure (operating pressure to be described later) is introduced from the reverse port 855, and this operating pressure causes the spool 85 to
1 is fixed in a state where it is pressed to the right side in the figure. Therefore, when traveling in reverse, the hydraulic cylinder 414 of the primary pulley 41 and the drain port 854 are always communicated with each other, and the gear ratio is fixedly maintained at the maximum gear ratio.

In addition, even in neutral and parking states (when the shift valve 87 is in the N and P shift positions), where the driving force of the engine 1 is not transmitted to the axle 61 by the forward/reverse switching mechanism 3, the same state as when reversing is maintained. become.

The hydraulic oil whose pressure is regulated by the line pressure regulating valve 82 is
In addition to line 101, it is also sent out to line 105. The hydraulic fluid sent to line 105 is adjusted to a predetermined operating pressure by an operating pressure regulating valve 88, and then supplied to line 106 and line 1.07.

The operating pressure regulating valve 88 has a spool 881 and a spool 88.
a pilot chamber 882 formed on one end side of 1; a spring 883 interposed in this pilot chamber 882;
It has a first pressure regulating boat 884 connected to line 105, a second pressure regulating boat 885 connected to line 107, and a drain port 886. The pilot chamber 882 is connected to the first duty solenoid valve 91 via the pilot passage 103a. Therefore, the pilot chamber 882 includes the first duty solenoid valve 91.
The hydraulic oil whose pressure was regulated in step 1 is introduced as pilot pressure. 8 While this pilot pressure acts in the same direction as the biasing force of the spring 883, the hydraulic pressure in the line 105 acts on the other end of the spool 881 to counteract the biasing force and the pilot pressure. Spool 88 depending on the power relationship
1 moves to communicate and disconnect between the first and second pressure regulating ports 884 and 885 and the train port 886, so that the operating pressure of the forward clutch 36 and the reverse brake 37 is reduced to the first duty solenoid valve. The pressure is controlled to a value according to the pilot pressure regulated at step 91.

The hydraulic oil supplied to the line 106 is transferred to the shift valve 87
Gari, 2. When in the shift position 1, the forward clutch 3 of the forward/reverse switching mechanism 3 is connected via the line 109.
When the shift valve 87 is in the R shift position, it is supplied to the hydraulic chamber 37a of the reverse brake 37 of the forward/reverse switching mechanism 3 via the line 108, and is also supplied via the line 104. gear ratio control valve 85
It is designed to be supplied to the report port 855.

]9 On the other hand, the hydraulic oil in each hydraulic chamber 36a, 37a of the forward clutch 36 and reverse brake 37 of the forward/reverse switching mechanism 3 flows through the line 109 when the shift valve 87 is in the R, N, or P shift position. , 108. Therefore, the forward clutch 36 of the forward/reverse switching mechanism 3
The reverse brake 37 is engaged and released according to the shift position of the shift valve 87, and as described above, the gear ratio of the continuously variable transmission mechanism 4 reaches the maximum gear ratio at the R, N, and P shift positions. It is held fixed in the state of .

Further, the hydraulic oil supplied to the line 107 is supplied to the lockup engagement chamber 29a or lockup release chamber 29b of the lockup clutch 28 via the lockup control valve 89. The lock-up control valve 89 is configured such that the operation of the spool 891 is controlled by the pilot pressure regulated by the third duty solenoid valve 93. That is, when the pilot pressure decreases, the spool 891 moves to the right in the figure, so that hydraulic oil is supplied from the line 107 to the lockup engagement chamber 029a, and hydraulic oil in the lockup opening chamber 29b is drained. The lock-up clutch 28 is then engaged.

On the other hand, when the pilot pressure increases, the spool 891
moves to the left in the figure, hydraulic oil is supplied from the line 107 to the lock-up opening chamber 29b, the hydraulic oil in the lock-up engagement chamber 29a is drained, and the lock-up clutch 28 is disengaged. In this embodiment, the lock-up control valve 89 and the third
Engagement force control means 120 that controls the engagement force of the lock-up clutch 28 by the duty solenoid valve 93
is configured.

In addition, in the figure, 94 is the first duty solenoid valve 91.
95.96 is an accumulator that relieves the shock when the forward clutch 36 and reverse brake 37 are engaged, respectively. 97 is a relief valve. It is. Also, 98
is a pressure holding valve that maintains a predetermined low constant pressure when draining the pressure oil in the hydraulic cylinders 414 of the four immediate pulleys.

FIG. 7 shows an electric control circuit for the above-mentioned continuously variable transmission. In this figure, a control unit 110 containing a microcomputer etc. receives a shift position signal from a shift position sensor 111 that detects the shift position (D, 1.2°R, N, P) operated by the driver; The primary pulley rotation speed signal from the primary rotation speed sensor 112 that detects the rotation speed NP of the primary shaft 4110, the secondary pulley rotation speed signal from the secondary rotation speed sensor 113 that detects the rotation speed NS of the secondary shaft 421, and the A throttle valve opening signal from a throttle opening sensor 114 that detects the throttle valve opening TvO, an engine rotational speed signal from an engine rotational speed sensor 115 that detects the rotational speed Ne of the engine 1, and a turbine shaft of the torque converter 2. The turbine rotation speed signal from the turbine rotation speed sensor 116 which detects the rotation speed Nt of 25 is inputted.

The control unit 110 controls the duty solenoid valves 9 to 9 based on these input signals.
L 92 and 93 are duty-controlled, thereby controlling the line pressure adjustment valve 82, the operating pressure adjustment valve 88, and the gear ratio control valve 85.
and each pilot pressure introduced into the lock-up control valve 89.

As a feature of the present invention, the initial alignment in the axial direction between the primary and secondary pulleys 41 and 42 is set as follows.

That is, as the gear ratio changes, the belt 43 moves between the primary and secondary pulleys 41 . ．． At this time, as shown in FIG. 4, for example, the belt 43 wound around the primary pulley 4 is partially The deviation of the plane P passing through the width direction center of the belt 43 at the secondary pulley 42 with respect to the primary side plane PP passing through the width direction center of the belt 43 and orthogonal to the primary axis 411 is expressed as an alignment deviation (misalignment). , side planes pp and P coincide, the misalignment is considered to be zero. In this embodiment, in anticipation of deformation of the case 7, play of the bearing, etc., the initial alignment in the axial direction between the two pulleys 4142 is set so that the two pulleys 41,42 when the gear ratio is maximum and the transmission torque is maximum. It is set in advance to be shifted in the opposite direction to the direction in which the alignment changes so that the shift in the axial alignment between them becomes zero.

Specifically, the gear 51 on the secondary shaft 421 in the speed reduction mechanism 5 and the gear 52 on the intermediate shaft 53 that meshes with the gear 51 are both helical gears. 421, when a leftward shaft driving force is applied to the secondary pulley 421 in the figure, the above-mentioned alignment deviation is such that a leftward deviation of the secondary pulley 42 with respect to the primary pulley 41 in the drawing is a positive direction, and a rightward deviation is a negative direction. It is set. As shown in Figure 5, this misalignment changes in response to changes in the gear ratio and transmission torque, and reaches its maximum in the positive direction when the gear ratio is ``1'', and from there the gear ratio decreases. It becomes smaller as it increases or becomes larger.

Therefore, in this embodiment, the difference A between the alignment deviation at no load when the gear ratio is maximum but there is no transmitted torque, and the alignment deviation when the transmission torque is maximum at the same maximum gear ratio is A. has been determined through experiments, etc.
The initial alignment is deviated in the negative direction by the difference A, and the setting is made so that the alignment deviation under load becomes zero when the gear ratio is at the maximum.

As shown in enlarged detail in FIG. 3, in the hydraulic cylinder 4 which slides the movable sheave 413 of the primary pulley 41, an outer cylinder 415 and an inner plunger 41 define both oil chambers 14a and 14d.
6 is fixed to the primary shaft 41 through a cylindrical holder 419 serving as a collar so as not to be movable in the axial direction. That is, a holder 419 is fitted on the outer periphery of the primary shaft 411, and the outer cylinder 415 and the inner plunger 4
16 is press-fitted and fixed to this holder 419. Due to this structure, the primary pulley 4 due to manufacturing errors of the outer cylinder 415 and the inner plunger 4]6 is fixed.
This is to reduce the misalignment in the axial direction of 1.

Therefore, in the above embodiment, when changing the gear ratio small, the second duty solenoid valve 9
2, the pilot pressure introduced into the pilot chamber 856 of the gear ratio control valve 85 increases and the spool 851 moves to the left in FIG. Line pressure flows into the hydraulic cylinder 414 of the primary pulley 41, and the amount of the inflow increases, so that the hydraulic pressure acting on the hydraulic cylinder 414 increases. Therefore, the effective pitch diameter of the primary pulley 41 becomes large, and the effective pitch diameter of the secondary pulley 42 becomes small. On the other hand, when changing the gear ratio to a larger value, the spool 851 moves to the right in FIG. 6, contrary to the above, the amount of line pressure flowing into the hydraulic cylinder 414 decreases, and the oil pressure decreases. So, primary pulley 4
As the effective pitch diameter of the secondary pulley 42 becomes smaller, the effective pitch diameter of the secondary pulley 42 becomes larger.

In this case, when the gear ratio 1 is at its maximum and the transmission torque is at its maximum under load, the secondary pulley 42 is displaced relative to the primary pulley 41 to the left in FIG. 2, and the primary and secondary pulleys 4 . Since they are set to be shifted, the alignment shift is zero under the above load. Therefore, even when the belt 43 is loaded with a large tension, the torsional stress applied to the belt 43 is reduced, thereby improving its durability and reliability.

(Effect of the invention) As explained above, according to the invention according to claim (1), when setting the axial alignment between the driving and driven pulleys of a continuously variable transmission, the gear ratio is maximum and the transmission torque is By anticipating the deviation of alignment 7 at the maximum high load and setting the initial alignment so that the deviation becomes zero, it is possible to eliminate the alignment deviation and reduce the inappropriate stress applied to the belt at the time of high load. Therefore, it has an excellent practical effect in that it can contribute to improving the durability and reliability of the belt.

[Brief explanation of drawings]

The drawings show an embodiment of the present invention, and FIG. 1 is a skeleton diagram showing the overall structure of the transmission, FIG. 2 is a sectional view showing the detailed structure, and FIG. 3 is an enlarged view of the ■ part in FIG. 2. Figure 4 is a conceptual diagram showing alignment deviations, Figure 5 is a characteristic diagram showing alignment change characteristics due to changes in gear ratio, Figure 6 is a hydraulic control circuit diagram, and Figure 7 is a block diagram of the electrical control system. It is. 41...Primary pulley (drive pulley) 42...
Secondary pulley (driven pulley) 43...Belt 411...Primary shaft (drive shaft) 421...Secondary shaft (driven shaft) 414.424...Hydraulic cylinder 8 A...Maximum gear ratio and load Misalignment 9 41...Primary pulley (drive pulley) 42...
Secondary pulley (driven pulley) 43...Belt 411...Primary shaft (drive shaft) 421...Secondary shaft (driven shaft) 414.424...Hydraulic cylinder gruel 4 No. 7 Figure 331-

Claims (1)

[Claims] (1) Equipped with drive and driven pulleys consisting of variable pulleys with variable effective radii and a belt wrapped around both pulleys, so that the speed ratio between the drive and driven pulleys can be changed steplessly and continuously. In the continuously variable transmission, the initial axial alignment between both pulleys is adjusted in advance so that the deviation in axial alignment between both pulleys is zero when the gear ratio is maximum and the transmission torque is maximum. A continuously variable transmission characterized by being shifted in the direction opposite to the direction in which alignment changes.