JPH0751980B2 - Hydraulic control device for continuously variable transmission - Google Patents

Description

Detailed Description of the Invention [Industrial applications]

The present invention relates to a hydraulic control device for a belt type continuously variable transmission for a vehicle, and more particularly to line pressure control in a low engine load and low rotation range. Regarding the hydraulic control of this type of continuously variable transmission, there is, for example, the basic control disclosed in JP-A-55-65755. Here, regarding the line pressure control, the pressure adjusting valve changes the spring force according to the gear ratio and the pitot pressure according to the engine speed to obtain the pulley pressing force required for torque transmission. That is, as shown in FIG. 5, the line pressure in this case has a characteristic that the gear ratio is higher toward the lower speed side, and when the engine speed is toward the higher speed side, the line pressure is downward. Therefore, in the idling state before starting, since the gear ratio is at the position of the maximum low speed stage, the line pressure is P ₁ on the maximum line pressure curve l _{1 in} FIG. The line pressure of the curve l ₁ is set so that the belt does not slip when the vehicle starts with the accelerator fully opened. Therefore, the line pressure becomes unnecessarily high for low engine load with a small accelerator opening, starting in a low speed range, and coasting, resulting in greater belt friction and greater drive loss of the oil pump. ， Deteriorate fuel efficiency. Therefore, in line pressure control, it is desired to promote optimization by taking into consideration factors other than speed change and engine speed.

[Prior art]

Therefore, conventionally, regarding the line pressure control of the continuously variable transmission, the power transmission efficiency between the belt and the input / output ports is detected as disclosed in, for example, JP-A-59-19756, and JP-A-58- It is disclosed that the belt slip is detected as in Japanese Patent No. 214054, and the control is performed in relation to the engine torque as in Japanese Patent Laid-Open No. 60-73160.

[Problems to be Solved by the Invention]

By the way, in any of the above-mentioned prior art methods, the line pressure is electronically controlled to always perform the minimum line pressure control. Therefore, the line pressure control has to be electronically controlled to electrically actuate various control valves, which is more complicated than the basic conventional hydraulic control. The present invention has been made in view of the above point, and in the conventional hydraulic control system of the basic line pressure, a simple device is added to optimize the line pressure in the low load and low rotation range of the engine. It is an object of the present invention to provide a hydraulic control device for a continuously variable transmission configured as described above.

[Means for solving problems]

In order to achieve the above object, the present invention provides a main shaft on the engine side with a variable pulley distance main pulley, and a wheel side auxiliary shaft arranged in parallel with the main shaft with a variable pulley distance auxiliary pulley. A drive belt is wound between the main pulley and the main pulley, and a pressure control valve that controls the line pressure in the oil passage from the hydraulic pressure source and supplies the line pressure to the cylinder of the sub pulley to apply the pulley pressing force is provided. Is equipped with a speed change control valve for supplying and discharging line pressure in the oil passage to the cylinder, and the ratio of the winding diameter of the drive belt to both pulleys is changed by the speed change control valve to supply and discharge the line pressure to achieve stepless operation. In a continuously variable transmission that shifts to, the pressure regulating valve acts on one of the spools with a spring force according to the gear ratio, and on the other side with a pitot pressure according to the engine speed and a line pressure of a predetermined pressure receiving area. , In these relationships Adjacent to the line pressure receiving port of the pressure control valve, a line pressure correction port with a small pressure receiving area is provided, and the line pressure port and the correction port are connected via a three-way solenoid valve to reduce the engine load. The line pressure is introduced into the correction port by the three-way solenoid valve in the low rotation speed range, and the line pressure is controlled to decrease.

[Action]

By controlling the three-way solenoid valve based on the above configuration, the line pressure controlled by the pressure adjusting valve to change the speed and the engine speed corresponds to the transmission torque in the case of low engine load and low engine speed. Then, it will be lowered, and it will come out of the area and return under other conditions. Thus, according to the present invention, it is possible to improve running performance and fuel efficiency without excessive line pressure in the case of low load of the engine, starting in a low rotation range, and coasting.

【Example】

An embodiment of the present invention will be specifically described below with reference to the drawings. Referring to FIG. 1, an example of a belt type continuously variable transmission to which the present invention is applied will be described. Reference numeral 1 is an electromagnetic powder clutch, 2 is a continuously variable transmission, and the continuously variable transmission 2 is roughly classified. Forward / reverse switching unit 3, pulley ratio conversion unit 4
The final deceleration unit 5 is configured by transmission. The electromagnetic powder clutch 1 is housed in one side of the clutch housing 6, the other side of the clutch housing 6 is joined to the main case 7, and the side of the main case 7 opposite to the clutch housing 6 is joined. Inside the side case 8, the switching unit 3, the pulley ratio conversion unit 4, and the final reduction unit 5 of the continuously variable transmission 2 are assembled. The electromagnetic powder clutch 1 includes a ring-shaped drive member 12 integrally connected to a crankshaft 10 from an engine via a drive plate 11, and a disk-shaped driven member integrally spline-connected to a transmission input shaft 13 in a rotational direction. Have 14. The coil 15 is built in the outer peripheral side of the driven member 14, and a gap 1 is formed between the members 12 and 14 along the circumference.
6 is formed, and this gap 16 communicates with the powder chamber 17 having the electromagnetic powder therein. In addition, a power feeding brush 19 is in sliding contact with the slip ring 18 of the hub portion of the driven member 14 including the coil 15, and the slip ring 18 further passes through the inside of the driven member 14 and is connected to the coil 15.
A clutch current circuit is configured. Thus, when the clutch current is applied to the coil 15, the gap
Due to the magnetic field lines generated between the drive and the driven members 12 and 14 via the magnetic particles, electromagnetic powder is coupled and accumulated in the gap 16 in a chain shape, and the coupling force by this causes the drive member 12
On the other hand, the driven member 14 is slid and integrally coupled, and the clutch is engaged. On the other hand, if the clutch current is cut off, the drive and driven members 12, 1
The coupling force of 4 disappears and the clutch is disengaged. If the control of the clutch current in this case is performed in conjunction with the operation of the switching unit 3 of the continuously variable transmission 2, the forward drive D from the P (parking) or N (neutral) range is performed.
(Drive) .Ds (Sporty drive) or clutch 1 automatically when switching to reverse R (reverse) range
Is disconnected and the clutch pedal operation becomes unnecessary. Next, in the continuously variable transmission 2, the switching portion 3 includes the input shaft 13 from the clutch 1 and the main shaft 20 coaxially arranged with the input shaft 13.
It is provided between and. That is, a reverse drive gear 21 that doubles as a forward engaged side is formed on the input shaft 13, and a reverse engaged side gear 22 is rotatably fitted to the main shaft 20. , 22 supported by shaft 23, counter gear 24, shaft 25
Are meshed with each other via an idler gear 26 supported by. A switching mechanism is provided between the main shaft 20 and the gears 21 and 22.
27 are provided. Here, the gear 2 that is always meshed
1, 24, 26, 22 are connected to the driven member 14 having the coil 15 of the clutch 1, and the switching mechanism 27 corresponds to the fact that the inertial mass of this portion when the clutch is disengaged is relatively large.
A sleeve 29 that is spline-fitted to the hub 28 of the main shaft 20 is configured to mesh with and be coupled to the gears 21 and 22 via the synchronizing mechanisms 30 and 31, respectively. As a result, at the neutral position of the P or N range, the sleeve 29 of the switching mechanism 27 fits only with the hub 28, and the main shaft 20 is separated from the input shaft 13. Next, when the sleeve 29 is meshed with the gear 21 side via the synchronizing mechanism 30, the main shaft 20 is directly connected to the input shaft 13 and the D or Ds range forward state is established. On the other hand, when the sleeve 29 is engaged with the gear 22 side via the synchro mechanism 31 in reverse, the input shaft 13 is connected to the main shaft 20 via the gears 21, 24, 26, 22 and the engine power is decelerated and reversed. The R range goes backward. In the pulley ratio conversion section 4, a sub shaft 35 is arranged in parallel with the main shaft 20, a main pulley 36 and a sub pulley 37 are provided on both shafts 20 and 35, respectively, and an endless structure is provided between the two pulleys 36 and 37. The drive belt 34 of is hung over. Each of the pulleys 36 and 37 is divided into two parts, and one pulley half body 36a, 37a is movable with respect to the other pulley half body 36b, 37b so that the pulley interval can be varied. Hydraulic servo devices 38 and 39, which also serve as pistons, are attached to 36b and 37b, and a spring 40 is urged to the movable pulley half 37b of the sub-pulley 37 in a direction to reduce the pulley interval. Further, an oil pump 41 as an operation source is installed next to the main pulley 36 as a hydraulic control system. This oil pump 41 is a high-pressure gear pump, and the pump drive shaft 42 has a main pulley 3
6, crankshaft 10 passing through the inside of main shaft 20 and input shaft 13
It is directly connected to and produces hydraulic pressure during engine operation. Then, the oil pressure of the oil pump 41 is controlled to supply and discharge oil to the respective hydraulic servo devices 38, 39, and the pulley spacing between the main pulley 36 and the sub pulley 37 is changed to the opposite relationship to drive the drive belt 34.
The pulley ratio of the pulleys 36, 37 is continuously converted, and the continuously variable power is output to the auxiliary shaft 35. The final deceleration unit 5 has a minimum pulley ratio of the pulley conversion unit 4 on the high-speed stage side of, for example, 0.5, which is very small. Therefore, the rotation speed of the sub shaft 35 is large. The output shaft 44 is connected via the gear 43. The final gear 46 meshes with the drive gear 45 of the output shaft 44, and the final gear 46 is transmitted to the left and right drive wheel axles 48, 49 via the differential mechanism 47. Referring to FIG. 2, the hydraulic system for gear shift control will be described. In the main pulley side hydraulic servo device 38, the movable pulley half 36b is fitted to the cylinder 38a integrated with the main shaft 20, and the line pressure is introduced into the cylinder 38a. Main pulley servo chamber 38b
Have. Also, in the sub-pulley side hydraulic servo device 39, the movable side pulley half body 3 is attached to the cylinder 39a integrated with the sub-shaft 35.
7b is fitted and has a sub-pulley servo chamber 39b into which the line pressure is introduced into the cylinder 39a, where the pulley half 36b has a larger line pressure receiving area than the pulley half 37b. . Then, the oil pumped from the oil sump 50 by the oil pump 41 is guided to the pressure adjusting valve 60 via the oil passage 51, and the oil passage 52 of the line pressure from the pressure adjusting valve 60 is constantly fed to the sub-pulley servo chamber 39b. It communicates to introduce the line pressure and further communicates with the gear ratio control valve 70. An oil passage 53 for supplying and discharging the line pressure communicates between the gear ratio control valve 70 and the main pulley servo chamber 38b. , 70 drain oil passages 54, 55 communicate with the oil reservoir side.
Further, a rotation signal detecting device 56 for extracting a control signal pressure of the pitot pressure according to the engine speed in the shift control after the clutch engagement is installed at the position of the cylinder 38a on the main pulley side. Pressure through oil passage 57 to each valve 6
Guided to 0,70. Further, for the D range in which the shift control is performed over a wide range including the low engine speed, the shift control is limited to the high range of the engine speed to obtain the Ds range in which the engine brake works when the accelerator is released. As a system, a ball check valve 58 is provided in the drain oil passage 54 from the pressure adjusting valve 60, and the oil passage 59 of the lubricating hydraulic circuit branched from the upstream side of this valve 58 communicates with the select position detection valve 90, The oil path 68 further branched from 59 is the actuator 10 of the gear ratio control valve 70.
It communicates with 0. The pressure regulating valve 60 has a spring 64 that is biased between the valve body 61, the spool 62, and one bush 63 of the spool 62, and engages with the main pulley movable pulley half 36b to change the actual speed. A sensor shoe 65 for detecting the ratio is movably supported by a shaft tube 66 that also serves as a lubricating oil passage and is connected to the bush 63. In the valve body 61, the Pitot pressure of the oil passage 57 and the line pressure of the oil passage 57 are introduced to the port 61a at the end of the spool 62 opposite to the spring 64. Further, the oil passage 51 on the pump side and the oil passage 52 for taking out the line pressure communicate with the port 61c.
The drain oil passage 54 communicates with the port 61d on the side of the spring 64 of c, and the port 61c is connected to the chamfer portion of the land 62a of the spool 62.
And 61d are connected to regulate pressure. That is, the pitot pressure and the line pressure act on the spool 62 in the direction to open the drain port 61d, while the load of the spring 64 depending on the gear ratio by the sensor shoe 65 acts in the direction to close the drain port 61d. . This allows
The line pressure is increased by increasing the spring load on the side of the low speed stage where the gear ratio is large, and conversely the line pressure is decreased by increasing the pitot pressure on the side of the high speed stage of the engine speed.
In this way, the pulley pressing force that does not cause belt slip is always maintained. Reference numeral 70 is a gear ratio control valve that controls the gear shift by supplying or discharging the line pressure to the main pulley servo chamber 38b according to the relationship between the accelerator opening and the engine speed, and 90 is a select position for extracting the operating hydraulic pressure according to each select position. The detection valve 100 is an actuator that shifts down when the accelerator of the Ds range is opened. Therefore, in the pressure adjusting valve 60 having the above structure, the line pressure correction port 61e is provided between the line pressure receiving port 61b and the pitot pressure port 61a. Then, the diameters of the lands 62b, 62c, 62d of the spool 62 at the ports 61b and 61e are set to D ₁ , D ₂ , D ₃
Then, there is a relation of D ₁ > D ₂ > D ₃ . Because of this, the port
The pressure receiving area S ₁ when the line pressure is introduced only into 61b is S ₁ = (D ₁ ² −D ₂ ² ) π / 4, and the pressure receiving area S ₂ when the line pressure is introduced into the ports 61b and 61e. Is S ₂ = (D ₁ ² −D ₃ ² ) π / 4, and S ₁ <S ₂ . In this way, the increase in the area for receiving the line pressure acts in the direction of decreasing the line pressure. On the other hand, the line pressure port 61c and the port 61e communicate with each other through an oil passage 115 having a three-way solenoid valve 110. The three-way solenoid valve 110 has an inlet port 112 when the coil 111 is off.
To connect the outlet port 113 to the drain port 114 and close the drain port 114 when the coil 111 is on.
Switch to connect 112 and 113. Drain port 114
Communicates with the oil sump 50 through an oil passage 116. Then, the solenoid valve 110 is configured to perform a switching operation by a signal from the control unit 120. Referring to FIG. 3, the control unit 120 will be described. The control unit 120 has an engine speed sensor 121, a brake switch 122, and an accelerator fully open switch 123. The signal Ne of the engine speed sensor 121 is input to the engine speed determination unit 124,
As shown in FIG. 4, it is determined whether the engine speed is lower than the set speed Ns which is slightly higher than the minimum shift line l ₂ as the low load and low speed range of the engine. The signal from the engine speed sensor 121 is input to the rapid acceleration determination unit 125, and the presence / absence of rapid acceleration is determined based on the engine speed increase rate dNe / dt. These signals are input to the operation mode determination unit 126, the solenoid valve 110 is turned on when Ne <Ns, and the solenoid valve 110 is turned off at the time of Ne ≧ Ns, rapid acceleration, braking, and full opening of the accelerator. Next, the operation of the hydraulic control device thus configured will be described. First, in the idling state before traveling, while the line is introduced into the sub pulley servo chamber 39b, the main pulley servo chamber 3b
8b is drained by the gear ratio control valve 70. Therefore, in the pulley ratio conversion unit 4 of the continuously variable transmission 2, the belt 34 is connected to the auxiliary pulley 37.
It shifts to the full side, and the low speed stage with the maximum gear ratio is reached. By the way, in this state, Ne <Ns is satisfied, so the control unit 120 turns on the solenoid valve 110 to introduce the line pressure into the port 61e of the pressure regulating valve 60. Therefore, the pressure adjusting valve 60 is acted on by the line pressure having a large pressure receiving area S ₂ , so that the line pressure decreases as indicated by point P ₂ in FIG. When the electromagnetic powder clutch 1 is engaged due to an increase in the engine speed, the engine power is transmitted and the vehicle starts. However, as long as the relationship of Ne <Ns is maintained during this starting and coasting, the same as above. Line pressure can be kept low. On the other hand, when the accelerator opening is large and Ne ≧ Ns, the vehicle starts, or when Ne <Ns after the vehicle starts and suddenly accelerates and dNe / dt exceeds the set value, the brake is operated or the accelerator is fully opened.
When 122 or 123 is turned on, the control unit 120 turns off the solenoid valve 110. Therefore, the port 61e of the pressure regulating valve 60 is drained so that the line pressure acts on the small pressure receiving area S ₁ , and thus the line pressure is immediately restored to the original state. Although the embodiment of the present invention has been described above, the solenoid valve 110 may be directly attached to the pressure regulating valve 60. The control method is not limited to the embodiment.

【The invention's effect】

As described above, according to the present invention, in the basic hydraulic control system of the line pressure by the pressure regulating valve, the line pressure is reduced in the low load and low rotation range of the engine, so that the belt at the time of starting and coasting is reduced. The friction and pump drive loss do not increase more than necessary, improving runnability and fuel efficiency. The line pressure receiving area is changed, and since the land diameter of the spool and the correction port are used, it is possible to reduce the change of the pressure adjusting valve. Since the electronic control uses a three-way solenoid valve, various elements can be added, and in the case of failure, normal line pressure control is performed and no inconvenience occurs.

[Brief description of drawings]

FIG. 1 is a sectional view showing an example of a continuously variable transmission to which the present invention is applied, FIG. 2 is a circuit diagram showing an embodiment of a hydraulic control device of the present invention, FIG. 3 is a block diagram of a control unit, and FIG. FIG. 4 is a diagram showing a shift pattern, and FIG. 5 is a diagram showing line pressure characteristics. 60 ... Pressure regulating valve, 61 ... Valve body, 61a ... Pitot pressure port, 61b
Line pressure receiving port, 61c Line pressure port, 61d Drain port, 61e Correction port, 62 Spool, 62b to 62
d ... Land, 110 ... Three-way solenoid valve, 115 ... Line pressure oil passage, 116 ... Drain oil passage, 120 ... Control unit.

Claims (1)

[Claims] 1. An engine-side main shaft is provided with a variable pulley-spacing main pulley, a wheel-side auxiliary shaft provided in parallel with the main shaft is provided with a variable-pulley-spacing auxiliary pulley, and a drive belt is provided between both pulleys. An oil passage to the cylinder of the main pulley is provided that is wound and controls the line pressure in the oil passage from the hydraulic pressure source, and supplies the line pressure to the cylinder of the sub pulley to apply the pulley pressing force. A speed change control valve that supplies and discharges line pressure is provided in the stepless speed change control that changes the ratio of the winding diameter of the drive belt to both pulleys by the line pressure supply and discharge action of the speed change control valve In the machine, the pressure regulating valve acts on one of the spools with a spring force according to the gear ratio, and on the other side with a pitot pressure according to the engine speed and a line pressure of a predetermined pressure receiving area, and the line pressure is related by these relationships. With pressure control Then, a line pressure correction port with a small pressure receiving area is provided next to the line pressure receiving port of the pressure control valve, and the line pressure port and its correction port are connected via a three-way solenoid so that it can be used in low engine load and low rotation speed regions. A hydraulic control device for a continuously variable transmission that controls the line pressure by introducing it into the correction port with a three-way solenoid valve.