JPH03181662A - Controller for continuously variable transmission - Google Patents

Abstract

PURPOSE:To improve the responsiveness of a controller by acting control pressure corresponding to solenoid current depending on a proportional electromagnetic relief valve onto one side of a spool of a secondary control valve to make setting pressure variable for linearly controlling secondary pressure. CONSTITUTION:Discharge pressure of a pump 34 is regulated to make control pressure corresponding to solenoid current Is depending on a proportional electromagnetic relief valve 55 act on one side of a spool 52 of a secondary control valve 50 to control secondary pressure Ps of a secondary cylinder 24, and setting pressure produced by a spring 53 therefore becomes variable for linearly controlling secondary pressure Ps to the solenoid current Is. The electromagnetic control unit 70 determines target secondary pressure, and supplies the solenoid current Is corresponding to the above secondary pressure to linearly control the secondary pressure Ps. Thus troubles such as the abnormal rising of the secondary pressure or the like does not occur, and moreover volume and oil pathes for the control pressure can be reduced to produce the better responsiveness and downsizing of a controller.

Description

[Detailed Description of the Invention] [Field of Industrial Application] The present invention relates to a control device for electronically controlling secondary pressure and speed change in a belt-type continuously variable transmission for a vehicle. This invention relates to secondary pressure control using a pilot-type secondary control valve.

[Conventional technology]

This type of continuously variable transmission applies primary pressure to a primary pulley on the input side, and applies secondary pressure to a secondary pulley on the output side, thereby applying a pressing force to a belt wrapped around both pulleys. The secondary pressure is controlled to provide a pressing force that does not cause belt slip to the transmitted torque, and the primary pressure can shift the belt to the primary or secondary boost to obtain a predetermined gear ratio. Controlled to the possible pressing force.

Regarding the secondary pressure control valve, since the transmitted torque depends on the gear ratio, there is a method of mechanically controlling the secondary pressure by applying a spring force according to the gear ratio to balance it. be. However, according to this method, it is necessary to always assume a state where the input torque is at its maximum, leading to an increase in secondary pressure, pump loss, etc. From this, there is a tendency to estimate the input torque, take into account fluctuations in the discharge pressure of the oil pump, and control the secondary pressure to exactly correspond to the transmitted torque.

By the way, an important problem in electronic control of secondary pressure is the structure of valves and the like that are hydraulically controlled by electric signals. Therefore, based on the premise of a mechanical control valve structure, the present application is based on a pilot-type system in which a duty solenoid valve is used to generate a control pressure with a duty ratio according to an electrical signal, and this control pressure is applied to the control valve. suggested by people. However, this method requires a control pressure circuit,
Measures such as correcting fluctuations in control pressure due to oil temperature, etc. are also required, which complicates the hydraulic circuit. For this reason, a control valve that variably controls the secondary pressure directly using an electronic signal and an electronic control system related thereto have been developed.

Conventionally, regarding electronic control of the continuously variable transmission, particularly the secondary pressure, there is a prior art, for example, disclosed in Japanese Patent Application Laid-Open No. 63-31533. Here, it is shown to have an output oil pressure return type main oil pressure control solenoid valve, and to control the secondary pressure by stroking the spool based on the balance between the electromagnetic force of the coil and the output oil pressure.

[Problem to be solved by the invention]

By the way, in the prior art mentioned above, since the valve is a three-way valve, the spool may close off the input boat on the pump discharge port side, and in this case, an excessive load may be applied to the pump and it may be damaged. Therefore, such a structure is not preferable. Furthermore, since the output hydraulic pressure is applied to the end of the spool, the structure is complicated, the hydraulic pressure increases, and the electromagnetic force of the coil and power consumption also increase. Furthermore, if the electrical signal of the coil is interrupted due to a wire break or the like, the spool will be located on the drain side and no secondary pressure will be generated, causing problems such as belt slip and the inability to run.

Here, as a method of preventing the pump discharge side from being closed, it is conceivable to configure the secondary control valve in the relief valve hole. If this relief valve hole is configured as a direct-acting type in which the spool is directly operated by the electromagnetic force of the solenoid, the structure will be simplified, which is preferable. The two-liquid force acts, for example, in a direction that increases the secondary pressure. This fluid force causes the balance of forces on the spool to become unstable, and has a particularly large effect when the engine speed is high and a large amount of hydraulic pressure is drained.

Therefore, one possible method for reducing the influence of the fluid force is to increase the control force of the spool to relatively reduce the influence of the fluid force. In this case, if the electromagnetic force of the solenoid is increased, the solenoid becomes larger and the power consumption increases, so it is necessary to devise another method.

The present invention has been made in view of the above points, and its purpose is to provide a control valve using electronic control of secondary pressure, which is small and highly safe, is less affected by fluid force, and has a fail-safe function. An object of the present invention is to provide a control device for a continuously variable transmission that can be equipped with a continuously variable transmission.

[Means to solve the problem]

In order to achieve the above object, the control device for a continuously variable transmission of the present invention regulates the pump discharge pressure by a secondary control valve to which an electric signal is input, and controls the secondary pressure of the secondary cylinder. A control pressure according to a solenoid current by a proportional electromagnetic relief valve is applied to one of the spools of the control valve to make the set pressure variable, and the secondary pressure is linearly controlled with respect to the solenoid current.

[For production]

Based on the above configuration, the secondary control valve provided in the hydraulic control system of the continuously variable transmission linearly controls the secondary pressure with respect to the solenoid current using the control pressure of the proportional electromagnetic relief valve. Therefore, if the electronic control system calculates the target secondary pressure and outputs the solenoid current according to this target secondary pressure, the secondary pressure can be accurately and optimally controlled to be the same as the target value by the secondary control valve.

〔Example〕

Embodiments of the present invention will be described below based on the drawings.

Referring to FIG. 1, an outline of the drive system of a continuously variable transmission with a lock-up converter will be described. Reference numeral 1 indicates an engine, and the crankshaft 2 is connected to a torque converter device 31, a forward/reverse switching device 4. Continuously variable transmission 5 and differential device 6
The transmission is configured sequentially.

In the torque converter device 3, the crankshaft 2 is connected to a converter cover 11 and a pump impeller 12a of a torque converter 12 via a drive plate 10. A turbine runner 12b of the torque converter 12 is connected to a turbine shaft 13, and a stator 12c is guided by a one-way clutch 14. A lock-up clutch I5 integral with the turbine runner 12b engages the drive plate IO. is releasably installed to transfer engine power to the torque converter 12 or lock-up clutch 1.
5.

The reverse switching device 4 has a double binion planetary gear 1B, the turbine shaft 13 is input to the sun gear lea, and output is output from the carrier 1.6b to the primary shaft 20.

A forward clutch I7 is provided between the sun gear 16a and the ring gear 16c, and a reverse brake I8 is provided between the ring gear 18c and the case, and when the forward clutch 17 is engaged, the planetary gear I6 is integrated to connect the turbine shaft 13 and the primary shaft. 20 is directly connected. Further, by engaging the reverse brake 18, reversed power is output to the primary shaft 20, and by releasing the forward clutch 17 and reverse brake I8, the planetary gear 16 is made free.

The continuously variable transmission 5 has a hydraulic cylinder 21 on a primary shaft 20.
The primary pulley 22 with variable pulley spacing has
A secondary pulley 25 having a hydraulic cylinder 24 is similarly provided on the secondary shaft 23, and the primary pulley 2
A drive belt 26 is wound between the pulley 2 and the secondary pulley 25. Here, the pressure-receiving area of the primary cylinder 21 is set to be larger than that of the primary cylinder 21, and the primary pressure of the primary cylinder 21 causes the primary pulley 22. The winding around the secondary pulley 25 is configured to change the diameter ratio so as to be continuously variable.

In the differential device 6, an output shaft 28 is connected to a secondary shaft 23 via a pair of compression gears 27.
The drive gear 29 of this output shaft 28 is the final gear 3.
meshes with 0. And the differential device 31 of the final gear 30
is connected to left and right wheels 33 via an axle 32.

On the other hand, in order to obtain a hydraulic pressure source for controlling the continuously variable transmission, a main oil pump 34 is arranged adjacent to the torque converter 12, and this main oil pump 34 is connected to the converter cover ll by a pump drive shaft 35, so that the main oil pump 34 is always connected to the converter cover ll by a pump drive shaft 35. The pump is driven by engine power to generate hydraulic pressure. In the continuously variable transmission 4, since the oil pressure is controlled over a wide range of high and low levels, the oil pump 34 is of a roller vane type, for example, and has a plurality of sets of suction and discharge ports, and is configured as a variable displacement type.

Next, a continuously variable transmission control system will be described as a hydraulic control system.

First, a pressure passage 41 from the oil pump 34 that communicates with the oil pan 40 communicates with the secondary control valve 50 to generate a predetermined secondary pressure Ps, and this secondary pressure Ps is constantly supplied to the secondary cylinder 24 through the oil passage 42. Supplied. The secondary pressure Ps is guided to the primary control valve BO via an oil passage 43, and is supplied to and discharged from the primary cylinder 2I via an oil passage 44 to generate a primary pressure Pp.

The primary control valve 60 uses a flow rate control method using a duty signal, and a spool 62 is pushed into the valve body 61.
An initial setting spring 63 is biased to one side of the spool 62 . The control unit 70 also includes a duty solenoid valve 64 to which the secondary pressure Ps of the oil passage 43 is guided.
A pulse-like control pressure Pc is generated using the secondary pressure Ps as a source pressure in response to a duty signal from the spool 62, and this pulse-like control pressure Pc is guided to a port etd at one end of the spool 62 by an oil passage 45. Furthermore, the secondary pressure Ps of the oil passage 63 is guided to the boat 81a at the other end of the spool 62. In this way, due to the relationship between the pulsed control pressure Pc and the constant high secondary pressure Ps, the spool 62 is set to the supply position where the boat old a of the oil passage 43 is connected to the boat 61b of the oil passage 44, and the boat Blb is set to the drain position. The oil drain position communicates with the boat Blc. Then, the duty ratio determines the proportion of oil supply and drainage time, that is, the inflow and outflow flow rates as well as the primary pressure P.
It is configured to perform speed change control by changing p.

The secondary control valve 50 is a pilot type equipped with a proportional electromagnetic relief valve, and a stepped spool 52 is attached to the valve body 51.
is inserted, and a spring 53 is biased against one side of the spool 52. In addition, a control chamber 54 is provided in the spring 53 portion.
is formed, and a proportional electromagnetic relief valve 55 is integrally provided adjacent thereto.

The spool 52 is connected to a large diameter land 5 from the spring 53 side.
2a and a small-diameter land 52b, and in a hydraulic chamber 51a communicating with the oil passage 41, a reaction force of the secondary pressure Ps is applied to both lands 52a, 521) in opposition to the spring 53. In addition, the hydraulic pressure is drained by the land 52b) 5
It drains to 1b to regulate the pressure, and is also capable of self-feedback.

The proportional electromagnetic relief valve 55 is provided in such a manner that a plunger 58 equipped with a spring 57 is attracted to a proportional solenoid 56 by electromagnetic force, and a valve body 59 with a spring of this plunger 58 is installed in the drain port 54a of the control chamber 54. . Further, an oil passage 47 branching from the oil passage 41 has an orifice 48 and communicates with the control chamber 54 so as to introduce secondary pressure. According to the electromagnetic force of the proportional solenoid 56, a control pressure Pc is generated in the control chamber 54 using the secondary pressure Ps as a source pressure, and the set pressure of the spool 52 is made variable.

As a result, in the spool 52, the lands 52a, 52
The force Fs of the spring 53 and the force of the control pressure Pc acting on the area SI of the land 52a oppose the reaction force of the secondary pressure Ps acting on the pressure receiving area difference ΔS of b, and the pressure is adjusted so that both are balanced. . Therefore, the balance equation in this case is as follows.

Pc-3, +Fs -Ps ·ΔS Also, the force F of the proportional electromagnetic relief valve 55 and the spring 57. A suppressing force (-IS on Fo-) is generated in accordance with the electromagnetic force caused by the solenoid current Is and the constant K. As a result, the opening area Sc of the port 54a is changed by the valve body 59, and the control pressure P is increased.
Since c occurs, the following equation holds true.

Pce Sc -Fo K .Is Therefore, the secondary pressure Ps is expressed by the following equation with respect to the solenoid current Is.

Ps-8I -Fo/Sc・Δ3+FS/ΔS-(K-
3,/Sc-ΔS)・IS That is, the control pressure Pc is inversely proportional to the solenoid current 1s as shown in FIG. 3(a), and the secondary pressure Ps is inversely proportional to the control pressure Pc as shown in FIG. 3(b). ), and as a result, as shown in Figure 3 (C), the solenoid current 1
The secondary pressure Ps linearly decreases in a one-to-one relationship as s increases. and solenoid current 1s
When is zero, the secondary pressure Ps becomes maximum, providing a fail-safe function that prevents belt slip due to wire breakage, etc. Due to this proportional pilot type characteristic, it is only necessary for the control unit 70 to determine the operation amount of the solenoid current for the target secondary pressure using this map and output it to the proportional solenoid 56.

Note that the oil pump 34 is of a variable displacement type, and a relatively high lubricating pressure is always generated in the oil passage 4B on the drain side of the secondary control valve 50. Therefore, this lubricating pressure is applied to the torque converter 12. Forward/forward switching device 4. The circuit is configured so that the lubricant is supplied to the lubricating section of the belt 24, etc.

In FIG. 2, the electronic control system will be described.

First, regarding the speed change control system, the primary pulley rotation speed sensor 71. Secondary pulley rotation speed sensor 72
．． It has an engine rotation speed sensor 73 and a throttle opening sensor 74. In the control unit 70, the primary pulley rotation speed Np and the secondary pulley rotation speed Ns of the primary pulley rotation speed sensor 71° and the secondary pulley rotation speed sensor 72 are input to the actual gear ratio calculation unit 75,
The actual speed ratio i is calculated by 1-Np/Ne. The actual gear ratio I and the throttle opening θ of the throttle opening sensor 74 are input to the target brake rotation speed search unit 7B,
The target primary pulley rotation speed NPD is determined based on the relationship I-θ. The target primary pulley rotation speed NPD and the secondary pulley rotation speed Ns are input to the target gear ratio calculation unit 77,
The target gear ratio 1s is calculated by 1s-NPD/Ns.

In this way, based on the speed change pattern, the target speed change ratio ls is determined based on the elements of the secondary boost rotation speed Ns, the actual speed change ratio I, and the throttle opening degree θ according to each driving and traveling condition.

Here, the flow rate that determines the required amount of oil for the primary cylinder 21 has a one-to-one correspondence with the rate of change de/dt of the actual pulley position e, and the rate of change de/dt of the actual pulley position is the same as the rate of change between the target pulley position es and the actual pulley position. It is expressed as the deviation from position 0 (es-e). In addition, the flow rate when the valve is operated according to the duty ratio is a function of the duty ratio and the actual pulley position e. Therefore, the duty ratio is a function of the actual pulley position change speed de/dt and the actual pulley position e. It can be determined by

Therefore, the actual gear ratio I and the target gear ratio 1s are determined by the actual pulley position converter 78. The target pulley position converter 79 inputs the input and converts it into an actual pulley position e and a target pulley position es, and these actual pulley positions 0. The target pulley position es is input to the actual pulley position change rate calculating section 80, and the actual pulley position change rate de/
dt is calculated as follows.

de/dt=to, (es-e)+to 2 ・de
s/di In the above formula, 2 is a constant, and des/
di is a phase advance element.

Then, the actual pulley position change speed de/dt and the actual pulley position e are input to the duty ratio search unit 81, and D-f(d
A duty ratio is searched based on the relationship e/dt, e), and a duty ratio corresponding to an upshift or a downshift and corresponding to the deviation of (es - e) is determined. This duty signal is output to the duty solenoid valve 64 via the drive section 82.

Next, the secondary pressure control system will be described.

First, an engine torque calculation unit 83 is provided which inputs the throttle opening degree θ and the engine speed Ne, and estimates the engine torque Te based on the torque characteristic of θ-Ne. In addition, the torque converter, output side engine rotation speed Ne, and primary pulley rotation speed Np are input to the torque amplification factor calculation unit 84, and a torque amplification factor t is determined according to the speed ratio n (Np/Ne). Te,) The torque amplification factor t is input to the input torque calculation unit 85 to calculate the input torque TI, TI
- Calculated by Te・t.

On the other hand, the actual boolean position e is input to the required secondary pressure setting section 86 to determine the secondary pressure Psu at the slip limit necessary for unit torque transmission, and the secondary pressure Psu at the slip limit necessary for unit torque transmission, the input torque TI is input to the target secondary pressure calculation unit 87 to obtain the target secondary pressure Ps.
s is calculated by Pss-TI-Psu. Further, the secondary pulley rotation speed Ns is input to the target secondary pressure calculation unit 87, and the centrifugal extraction bone etc. are corrected to decrease according to the secondary pulley rotation speed Ns.

Here, in order to perform feedback control of the secondary pressure control, a pressure sensor 88 is provided, and the actual secondary pressure Ps detected by the pressure sensor 88 and the target secondary pressure P
ss is input to the solenoid current calculation section 89. And the difference ΔP between the target secondary pressure Pss and the actual secondary pressure Ps
(-Pss-Ps) is determined, the solenoid current Is is corrected accordingly, and is outputted to the proportional solenoid 5B via the drive section 90.

Next, the operation of the continuously variable transmission control device having such a configuration will be described.

First, by operating the engine 1, the torque converter 12
converter cover 11. The oil pump 34 is driven by the rear drive shaft 35 to generate extraction pressure, and this oil pressure is guided to the secondary control valve 50. Therefore, when the vehicle is stopped, the target gear ratio Is of the transmission control system is set to a value larger than 2.5 as the mechanically maximum gear ratio of the continuously variable transmission 5, and a corresponding duty signal is sent to the duty solenoid valve 64. By inputting and operating the primary control valve 60 to the oil drain side, the primary pressure Pp is not generated. Therefore, all of the secondary pressure Ps from the secondary control valve 50 is supplied only to the secondary cylinder 24, and the continuously variable transmission 5 is in the low speed stage with the maximum gear ratio in which the belt 26 is shifted most toward the secondary pulley 25.

At this time, the lock-up clutch 15 is released by a hydraulic control system (not shown), and the torque converter I2 is supplied with oil. Therefore, for example, when shifting to the drive range, the forward clutch 17 of the forward/reverse switching device 4 is engaged by refueling and becomes the forward position.

Therefore, the power of the engine 1 is transferred to the torque converter 12,
It is input to the primary shaft 20 of the continuously variable transmission 5 via the forward/reverse switching device 4, and the primary pulley 22. The power of the maximum gear ratio is outputted to the secondary shaft 23 by the secondary pulley 25 and the belt 2B, which is transmitted to the differential device 6.
The signal is transmitted to the wheels 33 via , and the vehicle can be started.

On the other hand, in the secondary pressure control system, the engine torque T
o is estimated, and during idling before starting, both the engine torque Te and the input torque Ti are small, and the target secondary pressure Pss is set relatively low. Therefore,
In accordance with the target secondary pressure Pss, a relatively large solenoid current Is in the map shown in FIG. The pressing force of the valve body 59 is reduced. Therefore, the control pressure Pc decreases and the set pressure of the spool 52 is set low.
The spool 52 strokes to the left, more oil is drained by the land 52b, and the secondary pressure Ps is controlled to be low.

Next, when starting by pressing the accelerator, the engine torque T
e, torque amplification factor t, because the secondary pressure Psu at the slip limit required for unit torque transmission is large, the target secondary pressure Pss is also calculated to increase rapidly, and the solenoid current Is increases due to the difference with the secondary pressure Ps of the pressure sensor 88. significantly reduced. Therefore, in the secondary control valve 50, the pressing force of the valve body 59 is increased in response to the decrease in the electromagnetic force of the proportional solenoid 5B of the proportional electromagnetic relief valve 55, a high control pressure Pc is generated, and the set pressure of the spool 52 is increased. For this reason, spool 5
2 operates to reduce the drain amount by stroking to the right, and the secondary pressure Ps is controlled to be high. Then, after the vehicle is started, the speed is controlled and the lock-up clutch 15 is engaged to reduce the calculated value of the target secondary pressure Pss. In the secondary control valve 50, the set pressure is sequentially decreased together with the control pressure Pc by the proportional electromagnetic relief valve 55. The pressure Ps is controlled to decrease.

In this way, the characteristics of the secondary pressure Ps are summarized as shown in FIG. 4, and the control is optimally performed so as to always ensure the minimum boob pressing force that does not cause belt slippage with respect to the transmitted torque.

Here, due to the influence of pump discharge pressure etc., the primary pressure Ps
When this changes, the stroke of the spool 52 changes and is corrected by the hydraulic pressure in the hydraulic chamber 51a of the secondary control valve 50, thus creating a self-feedback effect. Furthermore, when the solenoid current Is is cut off due to a disconnection or the like during driving, the spring 57 in the secondary control valve 50 causes the set pressure to reach the maximum together with the control pressure Pc, and the secondary pressure Ps is controlled to be the highest. This provides a failsafe to prevent belt slip.

Also, after starting, Is<2.5 depending on driving and driving conditions.
When the shift start condition is satisfied, the shift control system changes the actual gear ratio l.
, target gear ratio Is is actual pulley position e, target pulley position e
S, and the actual pulley position change speed de/dt is calculated. Then, a duty signal corresponding to the actual boolean position change rate de/dt and the actual boolean position e is output to the duty solenoid valve 64, and the control pressure Pc and the secondary pressure Ps of the source pressure are controlled by the primary control. It acts opposite to the valve 60 and controls the amount of oil supplied to the primary cylinder 21 to increase the primary pressure Pp. So primary pulley 2
2, the belt 2B is pressed against the primary pulley 22
, and upshift control is performed to a high speed gear with a small gear ratio. Further, during deceleration, the displacement amount of the primary cylinder 21 is controlled by the primary control valve 60 in accordance with the decrease in vehicle speed, the primary pressure Pp is reduced, and downshift control is performed to a low speed gear.

Referring to FIG. 5, a second embodiment of the present invention will be described.
In the secondary control valve 50, a passage 52c with a choke throttle is provided in the spool 52 so as to communicate the hydraulic chamber 51a and the control chamber 54. Therefore, in this embodiment, the secondary pressure Ps, which is the source pressure, is always introduced into the control chamber 54 through the passage 52c, and further acts on the orifice to generate the control pressure Pc in the same manner as described above.

Although the embodiments of the present invention have been described above, the electronic control system is not limited to the embodiments.

〔Effect of the invention〕

As described above, according to the present invention, in the secondary piezoelectric control system of the continuously variable transmission, the secondary control valve is configured as a proportional electromagnetic relief valve hole, and the pump discharge pressure is relieved and regulated. Therefore, inconveniences such as an abnormal increase in secondary pressure do not occur.

In addition, the secondary control valve is equipped with a proportional electromagnetic relief valve and has a large control force, so the influence of fluid force acting on the spool is small, and no correction is required for this, simplifying the control system, and increasing the spool operation to prevent pulp stick. , less likely to cause dust to get caught.

Furthermore, since the secondary control valve is provided with a proportional electromagnetic relief valve near the spool, the volume of control pressure and the number of oil passages are small, resulting in good responsiveness and miniaturization.

In addition, the setting pressure of the secondary pressure is made variable by the control pressure of the proportional electromagnetic relief valve, and the secondary pressure is controlled linearly with respect to the solenoid current, so the solenoid current is set relative to the target secondary pressure calculated by the electronic control system. This makes feedback control easier.

Furthermore, since the secondary control valve has a self-feedback function, it is possible to perform stable hydraulic control and prevent belt slip due to a fail-safe function in the event of wire breakage.

Then, the electronic control system calculates the target secondary pressure according to each transmitted torque based on the input torque and secondary pressure per unit torque, and operates the secondary control valve proportionally based on this to accurately and optimally control the secondary pressure. It is possible to simplify the control system and improve accuracy.

Furthermore, in the second embodiment, a choke is formed in the passage of the spool, and since the same effect as an orifice can be produced with a diameter larger than that of an orifice, there is less clogging with dirt.

[Brief explanation of drawings]

Figure 1 is an overall configuration diagram showing an embodiment of the control device for a continuously variable transmission of the present invention, Figure 2 is a block diagram of the electronic control system, and Figures 3 (a), (b), and (c) are secondary FIG. 4 is a characteristic diagram of the secondary pressure; FIG. 5 is a sectional view showing a second embodiment of the secondary control valve. 5...Continuously variable transmission, 21...Primary cylinder,
24...Secondary cylinder, 41.42.43...
・Secondary pressure oil path, 44...Primary pressure oil path, 5
0...Secondary control valve, 52...Spool, 53
...Spring, 54...Control room, 55...Proportional electromagnetic relief valve, 56...Proportional solenoid, 59...
・Valve body, 60...Primary control valve, 7o...Control unit

Claims (6)

[Claims] (1) In a control system that regulates the pump discharge pressure by a secondary control valve to which an electric signal is input and controls the secondary pressure of the secondary cylinder, one of the spools of the secondary control valve has a proportional solenoid relief valve that controls the solenoid current. A control device for a continuously variable transmission, characterized in that the control device is configured to apply a control pressure to make the set pressure variable, and linearly control a secondary pressure with respect to the solenoid current. (2) Claim (1) characterized in that the proportional electromagnetic relief valve is arranged coaxially adjacent to one end of the spool of the secondary control valve, and is controlled so that the control pressure is generated using the secondary pressure as the source pressure. A control device for the continuously variable transmission described above. (3) The control device for a continuously variable transmission according to claim (1), wherein the secondary control valve has a characteristic of controlling the secondary pressure in a decreasing function manner with respect to the solenoid current. (4) A claim characterized in that the proportional electromagnetic relief valve is configured so that the control pressure is controlled in a decreasing function relative to the solenoid current, and the secondary pressure at the spool is controlled in an increasing function relative to the control pressure. (1) A control device for a continuously variable transmission as described in (1). (5) The electric signal according to claim (1), wherein the electric signal is a solenoid current according to a target secondary pressure calculated by an electronic control system or a deviation between the target secondary pressure and the actual secondary pressure. Control device for gear transmission. (6) The secondary control valve is provided with a land that applies hydraulic reaction force due to secondary pressure to the spool and drains oil, and in the control chamber at the end facing the hydraulic reaction force of the spool, applies a spring and performs control. 2. The control device for a continuously variable transmission according to claim 1, wherein the control device is configured to apply pressure.