JP4853275B2 - Hydraulic control device for continuously variable transmission - Google Patents

Description

  The present invention is for a continuously variable transmission capable of supplying oil to an oil required portion where oil is required when torque transmission is performed by the continuously variable transmission and cooling the oil supplied to the oil required portion. The present invention relates to a hydraulic control device.

  Generally, when a continuously variable transmission is provided in a power path from a power source to a driven member in a vehicle, a transport machine, a machine tool, etc., heat generation and seizure between components constituting the continuously variable transmission is suppressed. In order to achieve this, lubricating oil is supplied to the continuously variable transmission for lubrication and cooling. An example of a continuously variable transmission having a lubrication / cooling device is described in Patent Document 1.

  The lubrication / cooling device for a continuously variable transmission for a vehicle described in Patent Document 1 has a first pump and a second pump driven by an engine, and is discharged from a first oil pump. The hydraulic oil is supplied to the first oil passage, and a part thereof is drained to the second oil passage by the primary regulator valve. The hydraulic oil in the second oil passage is supplied to the torque converter. Further, a part of the hydraulic oil in the second oil passage is drained to the first lubricating oil passage by the secondary regulator valve, and the hydraulic oil supplied to the first lubricating oil passage becomes the pulley of the belt-type continuously variable transmission. And the belt are configured to be supplied to the contact portion. On the other hand, the hydraulic oil discharged from the second pump to the no-load circulation oil passage is configured to be supplied to the first oil passage via the first check valve. Further, a part of the hydraulic oil discharged from the second pump to the no-load circulation oil path is configured to be supplied to the second circulation oil path via the on-off valve. Part of the hydraulic oil in the second circulation oil passage returns to the suction side of the second pump via the second check valve. Further, part of the hydraulic oil in the second circulation oil passage is supplied to the first lubricating oil passage via an oil cooler. Note that a third check valve is provided between the outlet of the cooler and the first lubricating oil passage, and hydraulic oil in the first lubricating oil passage is not supplied to the oil cooler side.

  In the invention described in Patent Document 1, when the engine speed is low, the on-off valve is closed, and the hydraulic oil discharged from the second pump is supplied to the first oil passage. In contrast, when the engine speed increases, the on-off valve is opened, and the hydraulic oil discharged from the second pump is supplied to the second lubricating oil passage. A part of the hydraulic oil in the second lubricating oil passage is cooled by an oil cooler to lubricate and cool the belt type continuously variable transmission. When the hydraulic pressure of the hydraulic oil in the second circulation oil passage increases, the second check valve is opened and the hydraulic oil is returned to the suction side of the second pump.

JP 2005-180620 A

  However, in the lubrication / cooling device described in Patent Document 1, the flow rate of oil cooled by the oil cooler is controlled by controlling the on-off valve. When a large amount is required, there is a possibility that the supply speed of the oil supplied to the oil cooler is insufficient.

  The present invention has been made against the background of the above circumstances, and an object thereof is to provide a hydraulic control device for a continuously variable transmission capable of increasing the supply speed of oil supplied to a cooling device.

In order to achieve the above object, the invention of claim 1 is a continuously variable transmission in which torque is input from a power source and the gear ratio between the input element and the output element can be changed continuously. An oil required part that requires oil when performing torque transmission with this continuously variable transmission, an oil pump that discharges oil to be supplied to the oil required part, and has a plurality of discharge ports, A hydraulic control device for a continuously variable transmission having a cooling device that cools oil discharged from a discharge port before the oil is supplied to the oil required portion. and before the oil is supplied to the cooling device, the boost in the minimum oil pressure by remote high hydraulic pressure pressure regulated by the minimum oil pressure or the secondary regulator valve in the hydraulic circuit of the entire device the hydraulic pressure of the oil A boosting equipment to a flow control valve for controlling the flow rate of oil supplied to the cooling device from the one of the discharge ports to generate a signal for controlling the boosting function of the booster, and the signal And a composite control device that controls the flow rate control function of the flow rate control valve .

According to a second aspect of the present invention, in addition to the configuration of the first aspect, the required oil amount determining means for determining the required amount of oil in the required oil portion, and the greater the required amount of oil in the required oil portion, And hydraulic control means for performing control to increase the hydraulic pressure of the oil boosted by the apparatus.

According to a third aspect of the present invention, in addition to the structure of the first or second aspect , the cooling device is provided between an oil passage through which oil discharged from a plurality of oil discharge ports merges and the oil required portion. Is provided.

According to a fourth aspect of the present invention, in addition to the structure of any one of the first to third aspects, an input disk and an output disk that are arranged to be rotatable about the same axis, and between the input disk and the output disk. When the torque of the power source is transmitted to the input disk, the toroidal type continuously variable transmission has the power roller connected to the input disk. And the traction oil is supplied to a portion in contact with the output disk, and the transmission force is transmitted between the input disk and the output disk by the shearing force of the traction oil. A continuously variable transmission corresponds to the continuously variable transmission, the input disk corresponds to the input element, and the output disk requires the output. And the toroidal-type continuously variable transmission is provided between the input disk and the output disk based on the contact radius of the power roller with respect to the input disk and the contact radius of the power roller with respect to the output disk. In this case, a portion where the power roller contacts the input disc and the output disc is included in the oil required portion.

  According to the first aspect of the present invention, the torque of the power source is transmitted to the continuously variable transmission, and when torque transmission is performed by the continuously variable transmission, oil is required at the oil required portion. On the other hand, of the plurality of discharge ports, any discharged oil is boosted to a pressure higher than the minimum pressure and supplied to the cooling device. Therefore, the supply speed of the oil cooled by the cooling device can be increased, and the amount of oil cooled by the cooling device can be increased.

  According to the invention of claim 2, in addition to obtaining the same effect as that of the invention of claim 1, the flow rate of oil supplied to the cooling device from any one of the discharge ports can be controlled by the flow control valve. it can. Further, a signal for controlling the boosting function in the boosting device can be shared as a signal for controlling the flow control function in the flow control valve.

  According to the invention of claim 3, in addition to obtaining the same effect as that of the invention of claim 2, the required amount of oil in the oil required part is judged, and the greater the required amount of oil in the oil required part, the more the pressure increase Control is performed to increase the oil pressure of the oil boosted by the device. Accordingly, it is possible to more reliably suppress “the flow rate of the cooled oil being insufficient in the oil required portion”.

  According to the invention of claim 4, in addition to obtaining the same effect as that of any of the inventions of claims 1 to 3, the oil discharged from a plurality of oil discharge ports merges, and the merged oil becomes the cooling device Cooled by. Therefore, the amount of oil cooled by the cooling device can be further increased.

  According to the fifth aspect of the invention, in addition to obtaining the same effect as any of the first to fourth aspects of the invention, when the torque of the power source is transmitted to the input disk of the toroidal type continuously variable transmission. In the toroidal type continuously variable transmission, traction oil is supplied to a portion where the power roller contacts the input disk and the output disk, and the shearing force of the traction oil causes a gap between the input disk and the output disk. Power transmission is performed. That is, in the toroidal type continuously variable transmission, torque transmission is performed by the function or characteristic of the oil. Further, the toroidal continuously variable transmission has a gear ratio between the input disk and the output disk based on a contact radius of the power roller with respect to the input disk and a contact radius of the power roller with respect to the output disk. Is determined.

  Next, an embodiment of the present invention will be described. The present invention can be used for vehicles, transport machines, machine tools, and the like. In the present invention, the power source is a device that generates power to be transmitted to the driven member. As this power source, a single driving force source or a plurality of types of power sources having different power generation principles can be used. For example, an engine, an electric motor, a hydraulic motor, a flywheel system, or the like can be used as a plurality of types of power sources having different power generation principles. The driven member performs an operation such as a rotational motion, a reciprocating motion, and a linear motion by the power transmitted from the power source. When the present invention is used in a vehicle, a continuously variable transmission is disposed in a power transmission path from a power source to wheels. That is, the wheel corresponds to the driven member. When the present invention is used for a vehicle, the vehicle having a power train configured to transmit the power of the power source to either the front wheels or the rear wheels, that is, the two-wheel drive vehicle, or the power of the power source is the front wheels. It can be applied to any vehicle having a power train configured to be transmitted to both the rear wheel and the four-wheel drive vehicle.

  In the present invention, the oil required portion that requires oil when torque is transmitted by the continuously variable transmission is a part of the elements constituting the continuously variable transmission or the power from the power source to the driven member. It may be provided in any of the elements constituting the transmission path. In the present invention, the continuously variable transmission includes a toroidal continuously variable transmission and a belt-type continuously variable transmission. Further, some of the elements constituting the power transmission path from the power source to the driven member include meshing portions of gears, bearings for supporting the rotating elements, and clutches for controlling the power transmission state between the rotating elements. And mechanisms such as brakes. Examples of the oil required portion include a portion where torque is transmitted by the shearing force of the oil, a portion which is lubricated and cooled by the oil, an actuator which operates using the oil as hydraulic fluid, a control device, a hydraulic device, and the like. In other words, the oil required part in the present invention is a part that requires oil to generate a phenomenon necessary for power generation, a part that requires oil as hydraulic oil, and is unavoidable when transmitting the generated power. In particular, parts that require oil are included. The concept of “oil required part” in the present invention is extracted from these various parts. Further, as a configuration for driving the oil pump in the present invention, a configuration in which the oil pump is driven by torque of a power source that transmits power to the driven member can be employed.

  As another configuration, a power source that does not transmit power to the driven member, that is, a dedicated driving power source for driving the oil pump is provided, and torque of the driving power source is transmitted to the oil pump. A driving structure can also be adopted. An example of the driving power source is an electric motor. In addition to the oil passage itself, the oil passage in the present invention includes configurations of a valve port provided in the oil passage, the valve itself, and the like. That is, the oil passage in this invention may be a mechanism that constitutes a passage through which oil flows. The booster in the present invention is a device capable of controlling the oil pressure, and is a mechanism capable of further increasing the oil pressure from the minimum oil pressure. The composite control device according to the present invention is a mechanism that generates a common signal for controlling the booster and the flow rate control valve. This signal is, for example, a signal pressure. The composite control device includes a solenoid valve that generates the signal pressure, an electronic control device that controls the energization current of the solenoid valve, and the like. Furthermore, when this invention is used for a machine tool, a cutting tool for cutting a workpiece, a chuck for holding the workpiece, and the like correspond to the driven member. Further, in the present invention, the input element and the output element transmit torque, and include mechanisms such as a disk-shaped disk, a connecting drum, a gear, a shaft, a carrier, and a member. In the present invention, a plurality of discharge ports are provided in a single oil pump, or a plurality of oil pumps that can be driven and stopped separately are provided, and each oil pump has one discharge port. It may be provided.

  Next, a specific example in which the present invention is used in a vehicle equipped with a continuously variable transmission will be described with reference to FIG. FIG. 2 is a conceptual diagram showing a power train of a vehicle having a continuously variable transmission and a control system of the vehicle. In the vehicle 1 shown in FIG. 2, an engine 2 is mounted as a power source, and torque output from the engine 2 is converted into a fluid transmission device 3, a forward / reverse switching device 4, a continuously variable transmission 5, and a differential 6. It is comprised so that it may be transmitted to the wheel 7 via. In FIG. 2, an engine 2 is mounted on the front portion of a vehicle 1, and a power train vehicle configured to transmit torque of the engine 2 to wheels 7 as rear wheels, so-called front engine / rear drive. An (FR) type vehicle is shown as an example. The engine 2 is a power unit that converts thermal energy generated when fuel is burned into kinetic energy and outputs the kinetic energy. The engine 2 includes an output shaft of the crankshaft 8, the fluid transmission unit 3, and the like. Are connected to transmit power. The fluid transmission device 3 is a clutch capable of transmitting power by the kinetic energy of the fluid. In the configuration example of FIG. 2, a torque converter having a torque amplification function is used as the fluid transmission device 3. Hereinafter, “fluid transmission device 3” will be referred to as “torque converter 3” for convenience.

  In this torque converter 3, a pump impeller 9 as an input element and a turbine runner 10 as an output element are disposed so as to be relatively rotatable and coaxially. The pump impeller 9 is connected to the casing 11 so as to rotate integrally therewith, and the casing 11 is connected to the crankshaft 8 so that power can be transmitted. Further, the turbine runner 10 is connected to the torque transmission shaft 12 so as to rotate integrally. The pump impeller 9 and the turbine runner 10 are provided with a large number of blades (not shown), and a converter oil chamber 13 is formed between the pump impeller 9 and the turbine runner 10. Then, hydraulic oil is supplied via the converter oil chamber 13, and power is transmitted from the pump impeller 9 to the turbine runner 10 by the kinetic energy of the hydraulic oil generated by the rotation of the pump impeller 9. Further, a stator 14 that selectively changes the flow direction of the hydraulic oil fed from the turbine runner 10 and flows into the pump impeller 9 is disposed on the inner periphery of the pump impeller 9 and the turbine runner 10. ing. By the function of the stator 14, the torque transmitted between the pump impeller 9 and the turbine runner 10 can be amplified.

  Furthermore, a lock-up clutch 15 that selectively connects and releases the casing 11 and the torque transmission shaft 12 is provided. Further, the lock-up clutch 15 is a mechanism provided for transmitting power between the casing 11 and the torque transmission shaft 12 by a frictional force. The lockup clutch 15 is configured such that the transmission torque is controlled by pressing a friction material rotating together with the torque transmission shaft 12 against the inner wall surface of the casing 11. In order to control the transmission torque of the lockup clutch 15, an engagement hydraulic chamber 16 and a release hydraulic chamber 17 are provided. Based on a pressure difference between the engagement hydraulic chamber 16 and the release hydraulic chamber 17. The lock-up clutch 15 is engaged or released. Specifically, when the hydraulic pressure in the engagement hydraulic chamber 16 becomes higher than the hydraulic pressure in the release hydraulic chamber 17, the friction material is pressed against the casing 11 to increase the frictional force. In this way, the transmission torque of the lockup clutch 15 is increased (engaged).

  On the other hand, when the hydraulic pressure in the engagement hydraulic chamber 16 becomes lower than the hydraulic pressure in the release hydraulic chamber 17, the friction material separates from the casing 11 and the frictional force decreases. In this way, the transmission torque of the lockup clutch 15 is reduced (released). In this embodiment, when the lockup clutch 15 is released, power transmission by the frictional force of the lockup clutch 15 is impossible. On the other hand, when the lockup clutch 15 is engaged, power transmission by the frictional force of the lockup clutch 15 is possible. In this configuration example, “engagement of the lock-up clutch 15” includes “slip of the lock-up clutch 15”. The converter oil chamber 13 communicates with the engagement hydraulic chamber 16, and when the oil pressure in the converter oil chamber 13 increases, the oil pressure in the engagement hydraulic chamber 16 increases, When the hydraulic pressure decreases, the hydraulic pressure in the engagement hydraulic chamber 16 decreases.

  On the other hand, the forward / reverse switching device 4 is a device that switches the rotational direction of the input shaft 18 of the continuously variable transmission 5 between forward and reverse with respect to the rotational direction of the torque transmission shaft 12. As the forward / reverse switching device 4, a double pinion type planetary gear mechanism is employed in the example shown in FIG. That is, a sun gear 19 that rotates integrally with the torque transmission shaft 12 and a ring gear 20 that is arranged coaxially with the sun gear 19 are provided. The pinion gears 22 are provided, and the two pinion gears 21 and 22 are held by the carrier 23 so as to rotate and revolve freely. Further, the forward / reverse switching device 4 has a forward clutch 24 for selectively connecting and releasing the torque transmission shaft 12 and the carrier 23 so that power can be transmitted. Further, the forward / reverse switching device 4 includes a reverse brake 25 that selectively fixes the ring gear 20 to switch the rotational direction of the input shaft 18 relative to the rotational direction of the torque transmission shaft 12 between normal and reverse. . In this embodiment, as the forward clutch 24 and the reverse brake 25, hydraulically controlled clutches and brakes are used. That is, a clutch hydraulic chamber (described later) for controlling the transmission torque of the forward clutch 24 is provided, and a brake hydraulic chamber (described later) for controlling the braking force or torque capacity of the reverse brake 25 is provided. .

  Next, the configuration of the continuously variable transmission 5 will be described. In FIG. 2, a toroidal continuously variable transmission is shown as the continuously variable transmission 5. The continuously variable transmission 5 is a so-called double cavity type continuously variable transmission having two sets of transmission units 28 with the input disk 26 and the output disk 27 as a set of transmission units 28. Two output disks 27 are arranged between the two input disks 26, and the input disk 26 and the output disk 27 are all arranged coaxially. The two input disks 26 rotate integrally with the input shaft 18 and are attached so as to be relatively movable along the axis of the input shaft 18. A pinching pressure generating mechanism (not shown) is provided that generates a pinching force in a direction that brings the input disks 26 closer to each other in the direction along the axis of the input shaft 18. For example, an actuator such as a hydraulic cylinder, a pneumatic cylinder, or a cam mechanism can be used as the clamping pressure generating mechanism. In this configuration example, a hydraulic cylinder is used. Further, a toroidal surface 29 is formed on each of the two input disks 26. On the other hand, the two output disks 27 are configured to be rotatable relative to the input shaft 18, and are connected so that the two output disks 27 rotate integrally. Further, a toroidal surface 30 is formed on each of the two output disks 27. Each transmission unit 28 is provided with a plurality of power rollers 31, and the outer peripheral surfaces of the power rollers 31 are the toroidal surfaces 29 of the input disks 26 and the toroidal surfaces of the output disks 27 constituting each transmission unit 28. 30 is contacted via traction oil.

  Each transmission 28 is provided with a trunnion (not shown) for supporting each power roller 31, and each power roller 31 is orthogonal to the axis of the input shaft 18 in a substantially horizontal plane. It is configured to be rotatable about a rotation center line. Each trunnion is configured to reciprocate in a direction along a substantially vertical center line, and rotates (tilts) within a certain angle range around the center line in a substantially horizontal plane. It is configured to be possible. Furthermore, an actuator (not shown) for controlling the operation and stop position of the trunnion in the direction along the center line is provided. In this specific example, a hydraulically controlled actuator is used, and the operation and stop position of the trunnion are controlled by controlling the hydraulic pressure of a plurality of hydraulic chambers for shift control (described later). Yes. Further, an output gear 32 that rotates integrally with the two output disks 27 is provided, and an intermediate shaft 33 is provided in parallel with the input shaft 18. Gears 34 and 35 are formed on the intermediate shaft 33, and the output gear 32 and the gear 34 are meshed with each other. An output shaft 37 having a gear 36 is provided, and an intermediate gear 38 that meshes with the gear 35 and the gear 36 is provided. Further, the output shaft 37 is connected to the differential 6 so as to be able to transmit power.

  Next, a control system of the vehicle 1 shown in FIG. 2 will be described. First, an electronic control unit 39 is provided, and the electronic control unit 39 is used for engine speed, input disk 26 speed, output disk 27 speed, vehicle speed, acceleration request, braking request, shift position, and the like. The signal shown is input. The electronic control device 39 outputs a signal for controlling the engine 2, a signal for controlling the hydraulic control device 40, and the like. By the function of the hydraulic control device 40, the lockup clutch 15 is engaged / released, the forward clutch 24 of the forward / reverse switching device 4 is engaged / released, the reverse brake 25 is engaged / released, a continuously variable transmission. 5 and the transmission torque of the continuously variable transmission 5 are controlled.

  Further, the outline of the control performed in the vehicle 1 will be described. The torque output from the engine 2 is transmitted to the torque converter 3. Here, when the hydraulic pressure in the engagement hydraulic chamber 16 is increased and the lock-up clutch 15 is engaged, power is transmitted between the crankshaft 8 and the torque transmission shaft 12 by frictional force. It is carried out. On the other hand, when the hydraulic pressure in the release hydraulic chamber 17 is increased and the lockup clutch 15 is released, the kinetic energy of the hydraulic oil between the pump impeller 9 and the turbine runner 10 is increased. Power transmission is performed. Thus, when the lockup clutch 15 is released, the speed converter 3 has a speed ratio between the pump impeller 9 and the turbine runner 10 in a region (torque converter range) of less than 1.0. In some cases, torque amplification by the function of the stator 14 is performed. On the other hand, when the speed ratio between the pump impeller 9 and the turbine runner 10 is in a region (fluid coupling range) closer to 1.0 than the torque converter range, torque amplification is not performed.

  In this way, engine torque is transmitted to the torque transmission shaft 12. Next, the principle and control of torque transmission in the forward / reverse switching device 4 will be described. When a forward position, for example, a D (drive) position is selected as the shift position, the hydraulic pressure in the clutch hydraulic chamber is increased, the forward clutch 24 is engaged, and the brake hydraulic pressure is engaged. The hydraulic pressure in the chamber is lowered, and the reverse brake 25 is released. Then, the torque transmission shaft 12 and the carrier 23 rotate integrally, and the torque of the torque transmission shaft 12 is transmitted to the input shaft 18. In contrast, when the reverse position is selected, the hydraulic pressure in the clutch hydraulic chamber is reduced, the forward clutch 24 is released, and the hydraulic pressure in the brake hydraulic chamber is increased. The reverse brake 25 is engaged. That is, the ring gear 20 is fixed. When the engine torque is transmitted to the sun gear 19, the ring gear 20 becomes a reaction force element, and the torque of the sun gear 19 is transmitted to the input shaft 18 via the carrier 23. Here, the rotation direction of the input shaft 18 is opposite to that in the forward position.

  Next, the principle and control of torque transmission in the continuously variable transmission 5 will be described. A load in a direction to bring the input disks 26 closer to each other is given by the clamping pressure generating mechanism. Further, traction oil is supplied to a portion where the power roller 31 contacts the toroidal surfaces 29 and 30. As described above, when torque is transmitted to the input shaft 18, the traction oil is pressurized by the load to cause a glass transition, and a large shearing force is applied to the torque from the input disk 26 to the output disk 27. Communicated. Here, according to the ratio of the radius of the contact portion S1 between the power roller 31 and the input disc 26 and the radius of the contact portion S1 between the output disc 27 and the power roller 31, the rotational speed of the input disc 26 and the output disc The ratio with the rotational speed of 27 changes, and the ratio of the rotational speed (rotational speed) becomes the speed ratio of the continuously variable transmission 5.

  Furthermore, the control of the gear ratio of the continuously variable transmission 5 will be described. The operation of the trunnion is controlled by controlling the hydraulic pressure in a hydraulic chamber for gear ratio control (described later). When the trunnion is stopped in a state where the axis of the power roller 31 is orthogonal to the axis of the input shaft 18, the transmission ratio is maintained substantially constant. Thus, “a position where the trunnion is stopped in a state where the axis of the power roller 31 is orthogonal to the axis of the input shaft 18” is a “neutral position”. On the other hand, when changing the gear ratio, the trunnion stopped at the neutral position is operated in one direction along the center line. Then, a force (side slip force) for tilting the power roller 31 is generated at the contact point between the power roller 31 and each of the disks 26 and 27, and each power roller 31 tilts. Here, the target stroke amount or target stroke position of the trunnion is obtained based on the target change amount of the gear ratio, and the actual stroke amount of the trunnion reaches the target stroke amount or the actual stroke position is When the target stroke position is reached, the trunnion is operated in the opposite direction along the center line, the trunnion is returned to the neutral position, and the tilting of the power roller 31 is stopped.

  In this way, the target transmission ratio is maintained in the continuously variable transmission 5. In the above-described control of the trunnion operation, when the trunnion stopped at the neutral position is operated in the same direction as the rotation direction of the input disk 31, the power roller 31 is tilted by the side slip force, A shift in which the contact radius on the input disk 26 is reduced and the contact radius on the output disk 27 is increased, that is, a downshift in which the transmission ratio is increased is performed. On the other hand, when the trunnion stopped at the neutral position is operated in a direction opposite to the rotation direction of the input disk 26, the power roller 31 is tilted by the side slip force, and the input disk 26 is tilted. A shift in which the contact radius of the output disk 27 becomes large and the contact radius of the output disk 27 becomes small, that is, an upshift in which the speed ratio becomes small is performed. When torque is transmitted from the input disk 26 to the output disk 27, the torque is transmitted to the intermediate shaft 33 via the output gear 32. The torque of the intermediate shaft 33 is transmitted to the output shaft 37 via the intermediate gear 38. The torque of the output shaft 37 is transmitted to the wheel 7 via the differential 6 and a driving force is generated.

  As described above, in the continuously variable transmission 5, power is transmitted by the shearing force of the traction oil. Therefore, by keeping the temperature of the traction oil low, the traction coefficient can be increased, and the power roller 31. The friction coefficient at the contact portion S1 is increased, and the power roller 31 can be prevented from slipping. As a result, a decrease in power transmission efficiency in the continuously variable transmission 5 can be suppressed, and the durability of the parts can be improved. In this embodiment, a mechanism for cooling the traction oil supplied to the continuously variable transmission 5 is provided, and an example of the cooling device will be described with reference to FIG. FIG. 1 is a specific configuration example of the hydraulic control device 40 described above. First, a main oil pump 41 and a sub oil pump 42 driven by engine torque are provided. The main oil pump 41 and the sub oil pump 42 are configured such that they can be driven and stopped separately. A strainer 46 is attached to the suction port 43 of the main oil pump 41 and the suction port 44 of the sub oil pump 42 via an oil passage 45. The strainer 46 is provided in the oil pan 47. An oil passage 49 is connected to the discharge port 48 of the main oil pump 41, and the oil passage 49 is connected to the shift control hydraulic chambers 50A and 50B via a shift control valve (not shown). Based on the hydraulic pressure in the shift control hydraulic chambers 50A and 50B, a force for operating one trunnion in the center line direction is generated. The oil passage 49 is connected to a hydraulic chamber 51 of the clamping pressure generating mechanism. Further, a clutch hydraulic chamber 52 and a brake hydraulic chamber 53 are connected to the oil passage 49.

  Further, the oil passage 49 is connected to the input port 55 of the primary regulator valve 54, and an oil passage 57 is connected to the output port 56 of the primary regulator valve 54. The primary regulator valve 54 is a pressure control valve that controls the oil pressure in the oil passage 49 by discharging the oil in the oil passage 49 to the oil passage 57. The oil passage 57 is connected to the first input port 59 of the lockup relay valve 58 via the oil passage 118. In addition to the first input port 59, the lockup relay valve 58 has a second input port 60, a first output port 61, a second output port 62, a drain port 63, and a signal pressure port 120. The lockup relay valve 58 can be switched to an on / off two-stage operation mode according to the signal pressure input to the signal pressure port 120. Specifically, when the signal pressure input to the signal pressure port 120 is equal to or lower than a predetermined pressure, the operation mode of the lockup relay valve 58 is turned off. When the operation mode of the lockup relay valve 58 is OFF, the second output port 62 and the drain port 63 are connected, and the first input port 59 and the first output port 61 are connected. And the second input port 60 is closed. When the signal pressure input to the signal pressure port 120 increases and the signal pressure exceeds a predetermined pressure, the operation mode of the lockup relay valve 58 is turned on. When the operation mode of the lockup relay valve 58 is turned on, the first input port 59 and the second output port 62 are connected, and the second input port 60 and the first output port 61 are connected. And the drain port 63 is closed.

  An oil passage 119 branched from the oil passage 118 is provided, and the oil passage 119 is connected to the second input port 60 via a lockup control valve 65. The lock-up control valve 65 linearly controls the oil pressure supplied to the second input port 60 based on the signal pressure input to the signal pressure port 66. Specifically, it has a control characteristic that increases the hydraulic pressure supplied to the second input port 60 in proportion to an increase in the signal pressure input to the signal pressure port 66. The first output port 61 is connected to the release hydraulic chamber 17, and the second output port 62 is connected to the converter oil chamber 13. An oil passage 64 is connected to the drain port 63, and a warmer 67 is provided on a route from the oil passage 64 to the oil pan 47. The warmer 67 is a heat exchanger that heats the oil by causing heat exchange between cooling water of a cooling device (not shown) that cools the engine 1 and oil flowing through the oil passage 64. . A relief valve 68 is provided between the drain port 63 and the warmer 67 in the oil passage 64.

  On the other hand, an oil passage 70 is connected to the discharge port 69 of the sub oil pump 42, and a check valve 72 is provided in the oil passage 71 connecting the oil passage 70 and the oil passage 49. The check valve 72 is opened when the oil pressure in the oil passage 70 becomes higher than the oil pressure in the oil passage 49 by a predetermined value or more, and supplies the oil in the oil passage 70 to the oil passage 49. The check valve 72 prevents oil in the oil passage 49 from flowing into the oil passage 70. Further, an oil pump switching valve 73 having a port connected to the oil passages 49 and 70 is provided. The oil pump switching valve 73 has an input port 74, a first output port 75, a second output port 76, and a signal pressure port 77, and the input port 74 is connected to the oil passage 70. The first output port 75 is connected to the oil passage 49, and the oil passage 78 is connected to the second output port 76. The oil pump switching valve 73 is switched in two operation modes, on and off, according to the level of the signal pressure input to the signal pressure port 77. Specifically, when the signal pressure input to the signal pressure port 77 is high, the operation mode of the oil pump switching valve 73 is turned on. When the operation mode of the oil pump switching valve 73 is turned on, the first output port 75 is closed, and the input port 74 and the second output port 76 are connected. On the other hand, when the signal pressure input to the signal pressure port 77 is low, the operation mode of the oil pump switching valve 73 is turned off. When the operation mode of the oil pump switching valve 73 is turned off, the second output port 76 is closed, and the input port 74 and the first output port 75 are connected.

  Further, the oil passage 78 is connected to the inlet 80 of the cooling device 79. The cooling device 79 is a device that cools oil flowing from the inlet 80 and discharges it from the outlet 81, and may be either air-cooled or water-cooled. The outlet 81 is connected to an oil passage 83 that supplies oil to the first oil required portion 82. The first oil required portion 82 is a portion that requires oil when torque transmission is performed by the continuously variable transmission 5. In other words, the first oil required portion 82 includes a portion where oil is used as a medium for torque transmission in the continuously variable transmission 5, and a portion which is cooled and lubricated by oil when torque transmission is performed in the continuously variable transmission 5. included. Examples of the first oil required portion 82 include a contact portion S1 between the power roller 31 and the toroidal surfaces 29, 30, a sliding portion of the bearing that supports the power roller 31, and a sliding portion of the bearing that supports the trunnion. Etc. The oil passage 83 is provided with a relief valve 115.

  The oil passage 78 is provided with a throttle portion 84, and an oil passage 85 connected to the oil passage 78 and bypassing the throttle portion 84 is provided. The throttle portion 84 may be either an orifice or a choke. The oil passage 85 is provided with a cooling flow rate control valve 86. The cooling flow rate control valve 86 is a mechanism for controlling (increasing / decreasing) the amount of oil supplied to the cooling device 79 via the oil passage 85. Specifically, the cooling flow control valve 86 is provided with a valve body (not shown) for adjusting the cross-sectional area of a port (not shown) through which oil flows, and a signal for controlling the operation of the valve body. A pressure port 87 is provided. As shown in FIG. 3, the cross-sectional area of the port increases in proportion to the increase in the signal pressure input to the signal pressure port 87, and the oil flow rate Q increases linearly. Yes. As the cooling flow rate control valve 86, for example, a spool valve can be used. A secondary regulator valve 88 is provided in a path connecting the oil path 78 and the oil path 57. The secondary regulator valve 88 has an input port 89, a first drain port 90, and a second drain port 91. The oil path 56 is connected to the input port 89, and the oil path 78 is connected to the first drain port 90 via the oil path 92. Specifically, the oil passage 92 is connected to a portion of the oil passage 78 between the second output port 76 and the throttle portion 84.

  A check valve 93 is provided in the oil passage 92. The check valve 93 is opened when the oil discharged from the first output port 90 is supplied to the oil passage 78 via the oil passage 92, and the oil in the oil passage 78 flows to the first output port 90. It has the structure which prevents this. Further, the second output port 91 is connected to the oil passage 45 through an oil passage 94. The secondary regulator valve 88 configured as described above has a function of controlling the oil pressure of the oil passage 57 by controlling the amount of oil discharged from the oil passage 57 to the oil passages 92 and 95. It is a valve. The secondary regulator valve 88 controls the oil pressure in the oil passage 57 by discharging the oil in the oil passage 57 to the oil passage 92 first when the oil pressure in the oil passage 57 rises. When the oil pressure rises, the oil in the oil passage 57 is also discharged to the oil passage 94. That is, the oil pressure in the oil passage 57 that opens the first drain port 90 is lower than the oil pressure in the oil passage 57 that opens the second drain port 91.

  An oil passage 116 is connected between the check valve 93 and the first output port 90 in the oil passage 92, and oil is supplied from the oil passage 116 to the second oil required portion 117. Yes. The second oil required portion 117 is a portion that is lubricated and cooled by oil. For example, the second oil required portion 117 is provided in a meshing portion of the gears constituting the forward / reverse switching device 4 and a path from the continuously variable transmission 5 to the output gear 37. The meshing portion between the gears corresponds to the second oil required portion 117. When the first oil required portion 82 and the second oil required portion 117 are compared, the required oil temperature is different. Specifically, it is desirable that the temperature of the oil supplied to the first oil required portion 82 is lower. This is because when the traction oil is supplied to the contact portion S1 of the power roller 31, the higher the viscosity is, the higher the traction coefficient is, and the power transmission efficiency and durability are excellent.

  Further, a cooling hydraulic pressure control valve 95 is provided in a path connecting the oil path 78 and the oil path 94. The cooling hydraulic control valve 95 has an input port 96, a drain port 97, a signal pressure port 98 and a feedback port 99. The input port 96 and the feedback port 99 are connected to the oil passage 78, the drain port 97 is connected to the oil passage 94, and the oil passage 100 is connected to the signal pressure port 98. The cooling oil pressure control valve 95 is a mechanism for controlling the oil pressure of the oil supplied to the cooling device 79 via the oil passage 78. Specifically, the cooling oil pressure control valve 95 controls the amount of oil discharged from the oil passage 78 to the oil passage 94 based on the signal pressure input to the signal pressure port 98, and the oil passage 78. The hydraulic pressure is controlled. FIG. 4 shows the hydraulic control characteristics of the cooling hydraulic control valve 95. That is, the cooling oil pressure control valve 95 has a characteristic that the oil pressure (cooling oil pressure) of the oil passage 78 increases in proportion to the increase of the signal pressure input to the signal pressure port 98. The cooling hydraulic pressure control valve 95 can be constituted by, for example, a spool valve.

  An oil passage 102 is connected to the oil passage 49 via a modulator valve 101. The modulator valve 101 controls the oil pressure of the oil passage 102, and a solenoid valve 103 is provided in a path from the oil passage 102 to the signal pressure port 77 of the oil pump switching valve 73. The solenoid valve 103 has an input port 104, an output port 105, and a drain port 106. The input port 104 is connected to the oil passage 102, the output port 105 is connected to the signal pressure port 77, and the drain port 106 is connected to the oil pan 47. When the solenoid valve 103 is energized (turned on), the input port 104 and the output port 105 are connected, and the drain port 106 is closed. Therefore, the signal pressure input from the oil passage 102 to the signal pressure port 77 becomes high. On the other hand, when the solenoid valve 103 is de-energized (off), the input port 104 is closed and the output port 105 and the drain port 106 are connected. Therefore, the oil in the signal pressure port 77 is discharged via the drain port 106, and the signal pressure input to the signal pressure port 77 becomes a low pressure.

  Further, the oil passage 102 is connected to two duty solenoid valves 107 and 108. First, the first duty solenoid valve 107 has an input port 109, an output port 110, and a drain port 111. The input port 109 is connected to the oil passage 102, the output port 105 is connected to the oil passage 112, and the drain port 111 is connected to the oil pan 47. The oil pressure passage 112 is connected to the signal pressure ports 120 and 66 in parallel. The first duty solenoid valve 107 is a mechanism for controlling the signal pressure output from the output port 110 by controlling the energization amount. Specifically, the first duty solenoid valve 107 has a characteristic that the signal pressure output from the output port 110 increases linearly in proportion to an increase in energization amount.

  On the other hand, the second duty solenoid valve 108 has an input port 113, an output port 114 and a drain port 115. The input port 113 is connected to the oil passage 102, the output port 114 is connected to the oil passage 110, and the drain port 115 is connected to the oil pan 47. A signal pressure port 87 of the cooling flow rate control valve 86 is connected to the oil passage 100. That is, the signal pressure port 87 and the signal pressure port 98 are connected in parallel. The second duty solenoid valve 108 is a mechanism for controlling the signal pressure output from the output port 114 by controlling the energization amount. Specifically, the second duty solenoid valve 108 has a characteristic that the signal pressure output from the output port 114 increases linearly in proportion to an increase in energization amount. The current values of the solenoid valves are controlled by the electronic control unit 39 based on the vehicle speed, the required driving force, the transmission ratio of the continuously variable transmission 5, the engine torque, the lockup clutch control map, and the like. For example, when the required driving force is high, when the gear ratio of the continuously variable transmission 5 is changed, or when the engine torque is high, ON can be selected as the operation mode of the solenoid valve 103. On the other hand, when the required driving force is low, when the transmission ratio of the continuously variable transmission 5 is maintained substantially constant, or when the engine torque is low, the solenoid valve 103 is selected to be turned off. Is possible. When the engagement pressure of the lockup clutch 15 is increased, the signal pressure of the first duty solenoid valve 107 is increased. On the other hand, when the engagement pressure of the lockup clutch 15 is reduced, the signal pressure of the first duty solenoid valve 107 is reduced. Further, when the transmission torque in the continuously variable transmission 5 is high, the signal pressure of the second duty solenoid valve 108 is increased, and when the transmission torque in the continuously variable transmission 5 is low, the second duty solenoid valve 108 The signal pressure is reduced.

  Next, the operation and control of the hydraulic control device 40 shown in FIG. 1 will be described. When the main oil pump 41 is driven by the torque of the engine 2, the oil stored in the oil pan 47 is drawn into the main oil pump 41 and discharged to the oil passage 49. The primary regulator valve 54 controls the amount of oil discharged from the oil passage 49 to the oil passage 57, thereby controlling the oil pressure (line pressure) of the oil passage 49. The oil in the oil passage 49 is supplied to the shift control hydraulic chambers 50A and 50B, and the shift control valves (not shown) control the hydraulic pressures in the shift control hydraulic chambers 50A and 50B. The gear ratio of the continuously variable transmission 5 is controlled. Further, the oil pressure supplied from the oil passage 49 to the clutch hydraulic chamber 52 and the brake hydraulic chamber 53 is controlled, and the engagement / release of the forward clutch 24 and the reverse brake 25 is controlled. The

  On the other hand, the oil in the oil passage 57 is also supplied to the first input port 59 and the second input port 60 of the lockup relay valve 58. Here, the oil pressure of the oil supplied to the second input port 60 is controlled by the lockup control valve 65. This is because the oil in the oil passage 49 is supplied to the oil passage 102, and the first duty solenoid valve 107 regulates the oil pressure in the oil passage 102 and outputs a signal pressure from the output port 110, The signal pressure is input to the signal pressure port 66 of the lockup control valve 65, and the pressure regulation characteristic of the lockup control valve 65 is controlled. Specifically, the oil pressure of the oil input to the second input port 60 increases in proportion to the increase of the signal pressure input to the signal pressure port 66.

  The signal pressure output from the signal pressure port 110 of the first duty solenoid valve 107 is also input to the signal pressure port 120 of the lockup relay valve 58, and the signal input to the signal pressure port 120 The operation mode of the lockup relay valve 58 is switched based on the pressure. When the signal pressure input to the signal pressure port 120 is not more than a predetermined value, the operation mode of the lockup relay valve 58 is turned off, and the oil in the oil passage 57 is supplied to the release hydraulic chamber 17, As the hydraulic pressure in the release hydraulic chamber 17 increases, the oil in the engagement hydraulic chamber 16 is discharged to the oil passage 64 and the hydraulic pressure in the engagement hydraulic chamber 16 decreases. In this way, the transmission torque of the lockup clutch 15 is reduced, and the lockup clutch 15 is released or slipped. The oil discharged to the oil passage 64 is warmed by the warmer 67 and then returned to the oil pan 47.

  On the other hand, when the signal pressure input to the signal pressure port 120 exceeds a predetermined value, the operation mode of the lockup relay valve 58 is turned on, and a part of the oil in the oil passage 57 is The oil pressure is supplied to the engagement hydraulic chamber 16 via the second output port 62, and the hydraulic pressure in the engagement hydraulic chamber 16 increases. A part of the oil in the oil passage 57 is supplied to the release hydraulic chamber 17 via the first output port 61. Here, when the hydraulic pressure in the engagement hydraulic chamber 16 is higher than the hydraulic pressure in the release hydraulic chamber 17, the transmission torque of the lockup clutch 15 is increased and the lockup clutch 15 is engaged. As described above, in the hydraulic control device 40 of FIG. 1, the oil discharged from the torque converter 3 to the oil passage 64 can be warmed by the warmer 67. Accordingly, it is possible to suppress an increase in the viscosity of the oil, and when supplying the oil returned to the oil pan 47 into the casing 11 of the torque converter 3 again, the relative relationship between the pump impeller 9 and the turbine runner 10 can be reduced. An increase in drag torque generated by rotation can be suppressed. When the vehicle 1 travels at a high speed or travels at a high load, the operation mode of the lockup relay valve 58 is turned on, the lockup clutch 15 is engaged, and the torque converter 3 is engaged. Therefore, oil is not discharged to the oil passage 64.

  On the other hand, when the sub oil pump 42 is driven by engine torque, the oil in the oil pan 47 is sucked into the sub oil pump 42 and discharged to the oil passage 70. When the signal pressure input to the signal pressure port 77 of the oil pump switching valve 73 is low, the operation mode of the sub oil pump switching valve 73 is turned off. Then, the oil in the oil passage 70 is supplied to the oil passage 49. On the other hand, when the signal pressure input to the signal pressure port 77 of the oil pump switching valve 73 is high, the operation mode of the sub oil pump switching valve 73 is turned on. Then, the oil in the oil passage 70 is supplied to the oil passage 78. When the oil pressure in the oil passage 70 is higher than the oil pressure in the oil passage 49 by a predetermined value or more, the relief valve 72 is opened and the oil in the oil passage 70 is supplied to the oil passage 49. The oil supplied to the oil passage 78 passes through the throttle portion 84 and is supplied to the cooling device 79. The oil cooled by the cooling device 79 is supplied to the first oil required portion 82 via the oil passage 83.

  Thus, when the oil in the oil passage 78 is supplied to the cooling device 79, the oil pressure can be controlled by the cooling oil pressure control valve 95. That is, the oil in the oil passage 78 is discharged to the oil passage 94 via the cooling oil pressure control valve 95. Here, as shown in FIG. 4, even when the signal pressure input to the signal pressure port 98 is the lowest value, the oil pressure in the oil passage 78 is ensured to be the lowest pressure P1. In other words, when the oil circulation area in the cooling oil pressure control valve 95 is the maximum, the oil pressure in the oil passage 78 is ensured to be the minimum pressure P1. As the signal pressure input to the signal pressure port 98 increases, the amount of oil discharged from the oil passage 78 to the oil passage 94 decreases, and the oil pressure of the oil passage 78 increases. . Further, it is possible to control the amount of oil supplied to the cooling device 79 via the oil passage 78. When the signal pressure input to the signal pressure port 87 of the cooling flow rate control valve 86 is the lowest value, the cooling flow rate control valve 86 is closed, but the oil in the oil passage 78 passes through the throttle portion 84. Then, the cooling device 79 is supplied. Thus, the amount of oil supplied to the cooling device 79 through the throttle portion 84 is shown in FIG. 3 as the minimum flow rate Q1. When the signal pressure input to the signal pressure port 87 of the cooling flow control valve 86 rises, the amount of oil supplied to the cooling device 79 via the cooling flow control valve 86 increases, as shown in FIG. The flow characteristics are as shown.

  Here, a control example of the second duty solenoid valve 108, the cooling hydraulic pressure control valve 95, and the cooling flow rate control valve 86 will be described with reference to FIG. First, the required oil amount in the first oil required portion 82 is determined (step S1). In step S1, the required oil amount can be determined based on parameters such as the vehicle speed, the torque input to the continuously variable transmission 5, the gear ratio of the continuously variable transmission 5, and the like. For example, the required oil amount increases as the vehicle speed increases, the required oil amount increases as the torque input to the continuously variable transmission 5 increases, and the required oil amount increases as the gear ratio of the continuously variable transmission 5 increases. Become more. In step S2 performed after step S1, the following control is performed. First, based on the required oil amount obtained in step S1, a target value for the amount of oil supplied to the cooling device 79 via the oil passage 78 and a target value for the oil pressure in the oil passage 78 are obtained. Based on the target value of the oil amount and the target value of the hydraulic pressure, the signal pressure output from the second duty solenoid valve 108 is controlled, and the flow rate of oil supplied from the oil passage 78 to the cooling device 79 and The oil pressure in the oil passage 78 is controlled.

  In the process of step S2, the process of increasing the target hydraulic pressure is performed as the target flow rate of oil increases. Assuming that the cross-sectional area of the port through which oil passes through the cooling flow rate control valve 86 is the same, the oil flow rate increases as the oil pressure in the oil passage 78 increases, and is supplied to the cooling device 79 as a result. This is because the amount of oil increases. In the processes of steps S1 and S2, the map may be prepared in advance, or the calculation process may be performed every time the routine is executed. After the process of step S2, the control routine is terminated. By executing the control of FIG. 5, the amount of oil that is actually supplied can be brought close to the required amount of oil, and the supply of cooled oil can be more reliably suppressed. As described above, the second duty solenoid valve 108 has a function of controlling the oil pressure of the oil passage 78 and a function of controlling the flow rate of the oil supplied from the oil passage 78 to the cooling device 79. Yes. In other words, the signal pressure output from the second duty solenoid valve 108 is shared by the control of the cooling hydraulic pressure control valve 95 and the control of the cooling flow rate control valve 86. Therefore, it is not necessary to provide a separate solenoid valve for controlling the cooling hydraulic pressure control valve 95 and the cooling flow rate control valve 86, an increase in the number of parts can be suppressed, and the manufacturing cost of the hydraulic control device 40 can be reduced. .

  Further, in the hydraulic control device 40 of FIG. 1, when the oil pressure in the oil passage 57 increases, the first drain port 90 is opened. At this time, the second drain port 91 is closed. The oil discharged from the oil passage 57 to the oil passage 92 is supplied to the cooling device 79 via the oil passage 78. 1 supplies oil to the cooling device 79 via the oil passage 78, and oil passages 49, 92 provided in parallel to the oil passage 78. The cooling device 79 can be supplied via Therefore, the amount of oil cooled by the cooling device 79 and supplied to the first oil required portion 82 can be increased as much as possible, and the cooling efficiency of the traction oil supplied to the continuously variable transmission 5 is improved. When the oil pressure in the oil passage 57 further increases after the first drain port 90 is opened, the second drain port 91 is opened, and a part of the oil in the oil passage 57 passes through the oil passage 94. It returns to the oil path 45 via. Further, when the vehicle 1 travels at a high speed or travels at a high load, the operation mode of the lockup relay valve 58 is turned on, and no oil is discharged from the torque converter 3 to the oil passage 64. Therefore, by controlling the flow rate of the oil cooled by the cooling device 79, the oil temperature of the oil in the entire hydraulic control device 40 can be controlled.

  Here, the correspondence between the configuration described with reference to FIGS. 1 and 2 and the configuration of the present invention will be described. The engine 1 corresponds to the power source of the present invention, and the input disk 26 and the input shaft 18 are The output disk 27 and the output gear 32 correspond to the output element of the present invention, the discharge ports 48 and 69 correspond to the plurality of discharge ports in the present invention, and the discharge port 69 corresponds to the input element of the present invention. The oil passage 78 corresponds to “one of the discharge ports” in the present invention, and the oil passage 78 corresponds to “an oil passage through which oil discharged from the plurality of oil discharge ports merges” in the present invention. The cooling flow control valve 86 corresponds to the flow control valve in the present invention, the cooling hydraulic control valve 95 corresponds to the booster in the present invention, and the electronic control unit 39 Oh Beauty second duty solenoid valve 108 is equivalent to control complex of the present invention, the contact portion S1 is included in the oil needs 82 in the present invention. Instead of the torque converter, a fluid coupling that does not have a torque amplification function can be used. Further, as the forward / reverse switching device 4, a parallel shaft gear type can be used instead of the planetary gear mechanism. Furthermore, although the main oil pump 41 and the sub oil pump 42 can be driven / stopped individually, a single oil pump may be provided with a plurality of discharge ports. Furthermore, the present invention can also be applied to a hydraulic control device provided with three or more discharge ports. Further, the correspondence between the functional means shown in FIG. 5 and the configuration of claim 3 will be described. Step S1 corresponds to the oil required amount judging means in the present invention, and step S2 is the hydraulic pressure in the present invention. It corresponds to the control means.

It is a figure which shows the specific example of the hydraulic control apparatus for continuously variable transmission in this invention. It is a figure which shows an example of the power train of the vehicle which has a continuously variable transmission in this invention, and its control system. It is a diagram which shows the characteristic of the cooling flow control valve used for the hydraulic control apparatus of FIG. It is a diagram which shows the characteristic of the cooling hydraulic control valve used for the hydraulic control apparatus of FIG. 3 is a flowchart illustrating an example of control that can be executed by the hydraulic control device of FIG. 1.

Explanation of symbols

  DESCRIPTION OF SYMBOLS 2 ... Power source (engine), 5 ... Continuously variable transmission (toroidal type continuously variable transmission), 18 ... Input shaft, 26 ... Input disk, 27 ... Output disk, 31 ... Power roller, 32 ... Output gear, 39 ... Electronic control device 40 ... Hydraulic control device 41 ... Main oil pump 42 ... Sub oil pump 48,69 ... Discharge port 78 ... Oil passage 82 ... First oil required part 108 ... Second duty solenoid valve, S1 ... contact part.

Claims (4)

A continuously variable transmission in which torque is input from the power source and the gear ratio between the input element and the output element can be changed steplessly, and oil is required for torque transmission with this continuously variable transmission. The required oil portion, the oil supplied to the required oil portion, and an oil pump having a plurality of discharge ports, and the oil discharged from any of the discharge ports are supplied to the required oil portion Before, in the hydraulic control device for continuously variable transmission, having a cooling device for cooling the oil,
Before the oil said discharged from one discharge port is supplied to the cooling device, the minimum oil pressure pressure regulated by the minimum oil pressure or the secondary regulator valve the oil pressure of the oil in the hydraulic circuit of the entire device a boosting equipment boosting the remote high pressure,
A flow rate control valve for controlling the flow rate of oil supplied to the cooling device from any one of the discharge ports;
A composite control device that generates a signal for controlling a boosting function in the boosting device and controls the flow rate control function in the flow rate control valve using the signal;
Hydraulic control system for a continuously variable transmission and having a. And the oil required amount determining means for determining the required amount of oil before Symbol oil required portion,
Hydraulic control means for performing control to increase the hydraulic pressure of the oil boosted by the booster device as the required amount of oil in the oil required portion increases;
Hydraulic control system for a continuously variable transmission according to claim 1, characterized in that it comprises a. 3. The continuously variable transmission according to claim 1 , wherein the cooling device is provided between an oil passage through which oil discharged from a plurality of oil discharge ports merges and the oil required portion. Hydraulic control device for machine. A toroidal continuously variable transmission having an input disk and an output disk that are rotatably arranged around the same axis, and a power roller interposed between the input disk and the output disk; When the torque is transmitted to the input disk, the toroidal continuously variable transmission is configured such that the traction oil is supplied to the portion where the power roller contacts the input disk and the output disk, and the shear force of the traction oil The power transmission is performed between the input disk and the output disk,
The toroidal continuously variable transmission corresponds to the continuously variable transmission, the input disk corresponds to the input element, the output disk corresponds to the output element, and the toroidal continuously variable transmission includes: A speed ratio between the input disk and the output disk is determined based on a contact radius of the power roller with respect to the input disk and a contact radius of the power roller with respect to the output disk; 4. The hydraulic control device for a continuously variable transmission according to claim 1 , wherein a portion where the power roller contacts the input disk and the output disk is included in the oil required portion .

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