JP4962077B2 - Hydraulic control device for continuously variable transmission - Google Patents

Description

  The present invention relates to a continuously variable transmission capable of continuously changing a gear ratio by transmitting power using hydraulic pressure, and more particularly to an apparatus for controlling the hydraulic pressure.

  If the hydraulic pump is driven by a power device such as an engine and the pressure oil generated by the hydraulic pump is supplied to the hydraulic motor, the power can be transmitted via the hydraulic pressure, and the hydraulic pressure can be controlled by transmitting the hydraulic pressure. The torque or power to be changed can be appropriately changed. One example thereof is described in Patent Document 1. In the transmission described in Patent Document 1, variable displacement hydraulic pump motors (hydraulic pump motors) are connected to reaction force elements in each of a pair of planetary gear mechanisms, and the discharge of each variable displacement hydraulic pump motor is performed. The outlets and the suction ports are connected to each other to form a closed circuit. Moreover, the power output from a power source such as an engine is input to the input element in each planetary gear mechanism. Furthermore, a drive gear for setting a so-called fixed stage is arranged on an intermediate shaft integral with the output element of each planetary gear mechanism, and a driven gear meshing with each drive gear is arranged on the output shaft. Yes. A synchronous coupling mechanism (so-called synchro) is provided for switching each gear pair composed of the drive gear and the driven gear between a state where torque can be transmitted and a state where torque is not transmitted.

  Therefore, if one of the variable displacement hydraulic pump motors is locked and the reaction force element is fixed, the power output from the power source is transmitted to one intermediate shaft through the planetary gear mechanism having the reaction force element. Further, power is transmitted to the output shaft through a gear pair that is connected to the intermediate shaft by synchronization. In this case, the gear ratio is a gear ratio according to the gear ratio of the gear pair involved in power transmission.

  The lock of the variable displacement hydraulic pump motor in this case is set by making the extrusion volume (capacity) of the other variable displacement hydraulic pump motor zero, that is, minimum. That is, since each variable displacement hydraulic pump motor is connected by a closed circuit, if the extrusion volume of the other variable displacement hydraulic pump motor is reduced to zero, the flow of pressure oil does not occur. By setting the extrusion volume of the hydraulic pump motor to a maximum, such as maximizing the extrusion volume of the hydraulic pump motor, one of the variable displacement hydraulic pump motors is locked and prevented from rotating.

  In addition, the displacement volume of each variable displacement hydraulic pump motor is set to be larger than zero, the predetermined gear pair is set to be capable of transmitting torque by synchronization on the side of one variable displacement hydraulic pump motor, and the other variable displacement hydraulic pressure is set. When the other gear pairs are made capable of transmitting torque by synchronization on the pump motor side, a gear ratio having an intermediate value of the gear ratio determined according to the gear ratio of each gear pair is set. That is, one variable displacement hydraulic pump motor generates pressure oil, which is supplied to the other variable displacement hydraulic pump motor, which operates as a motor, and that power is transmitted to the output shaft through the other gear pair. Is done. As a result, power that is a combination of the power transmitted through such a fluid and the power mechanically transmitted through one variable displacement hydraulic pump motor appears on the output shaft. The power through the fluid can be continuously changed by continuously changing the extrusion volume of each variable displacement hydraulic pump motor. It can be set continuously, i.e. steplessly.

  In addition, the transmission described in Patent Document 1 has a control pressure by an electrically controllable solenoid valve (electromagnetic control valve) in an oil passage that communicates discharge ports of two variable displacement hydraulic pump motors. Is provided so that the discharge pressure of each variable displacement hydraulic pump motor or the associated shaft torque can be electrically controlled via a solenoid valve.

  Patent Document 2 describes an invention relating to a fluid machine including a rotating body configured to switch between a cooling state and a non-cooling state of hydraulic oil according to a load acting on the rotating body. The device described in Patent Document 2 is adjacent to the HST (Hydro Static Transmission) in which the hydraulic pump mechanism is housed in the upper part and the hydraulic motor mechanism is housed in the lower part. When a cooling water channel is formed and a solenoid valve is provided in the cooling water pipe connected to the cooling water channel, and the circuit pressure of the hydraulic circuit is below a predetermined value, the solenoid valve is closed while the circuit pressure is predetermined. When the value is exceeded, the solenoid valve is opened to cool the lubricating oil.

  Further, in Patent Document 3, the amount of heat generated in the transmission mechanism of the automatic transmission is calculated, the required amount of lubricating oil is determined based on the amount of generated heat, and the lubricating oil is supplied. An invention relating to a lubricating oil control device for an automatic transmission configured to prevent a temperature increase of a friction engagement element is described.

JP 2006-266493 A JP 2001-59562 A Japanese Patent Laid-Open No. 10-141480

  In the transmission described in Patent Document 1, an oil pump called a charge pump or a boost pump for replenishing oil is provided in a closed circuit in which the variable displacement hydraulic pump motors communicate with each other. Is provided. This oil pump is driven by a power source such as an engine to replenish closed-circuit hydraulic pressure that connects the variable displacement hydraulic pump motors to each other, and to change the speed of a gear mechanism of a transmission or a clutch mechanism such as a synchro. Oil for cooling and lubrication is supplied to the parts requiring lubrication of the machine.

  For this reason, in the transmission described in Patent Document 1 described above, the amount of oil supplied to the lubrication required portion of the transmission by the oil pump may vary depending on the rotational state of the variable displacement hydraulic pump motor. That is, this oil pump supplies pressure oil to the closed circuit in order to compensate for the shortage of hydraulic pressure in the closed circuit due to unavoidable oil leakage in the closed circuit that connects the variable displacement hydraulic pump motors. Therefore, for example, when the rotational speed of the variable displacement hydraulic pump motor is high and oil leakage in the closed circuit increases, a large amount of oil is supplied to the closed circuit, and conversely, the rotational speed of the variable displacement hydraulic pump motor If the oil leakage is low in the closed circuit, a small amount of oil is supplied to the closed circuit.

  When the amount of oil supplied to the closed circuit by the oil pump increases, the amount of oil supplied to the parts requiring lubrication of the transmission decreases. Conversely, when the amount of oil supplied to the closed circuit by the oil pump decreases, Increases the amount of oil supplied to the machine's lubricating parts. Therefore, when the rotational state of the variable displacement hydraulic pump motor varies, the amount of oil supplied to the lubrication required part of the transmission also varies. Depending on the rotational state of the variable displacement hydraulic pump motor, lubrication may be necessary. There has been a problem that the oil supply to the part is insufficient, or conversely, the oil supply to the part requiring lubrication becomes excessive and the transmission efficiency of the transmission decreases.

  The present invention has been made paying attention to the above technical problem, and in a continuously variable transmission using a variable displacement hydraulic pump motor, for cooling or lubrication regardless of the rotational state of the hydraulic pump motor. It is an object of the present invention to provide a hydraulic control device for a continuously variable transmission that can appropriately and easily control the amount of oil to be supplied.

In order to achieve the above object, the invention of claim 1 is a variable displacement hydraulic pump driven by power output from a power source, and is supplied with pressure oil output from the hydraulic pump and driven. And a variable displacement hydraulic motor that outputs power to the output member, and the torque transmitted to the output member varies according to the capacity and hydraulic pressure of the hydraulic pump and the hydraulic motor. In the control device, a closed circuit that circulates pressure oil between the hydraulic pump and the hydraulic motor, and supplying hydraulic fluid for cooling or lubrication to the lubricating pump, the hydraulic motor, and the lubrication-necessary parts of the continuously variable transmission. A lubrication circuit, a replenishment pump that replenishes the closed circuit with pressure oil and supplies the hydraulic fluid to the lubrication circuit, a capacity of the hydraulic pump and the hydraulic motor, and a replenishment pump. A hydraulic control equipment of the continuously variable transmission that include a working device and for varying in conjunction with the flow rate of the hydraulic pump and the hydraulic oil supplied to the hydraulic motor from the flop, the hydraulic pump and the hydraulic motor A variable displacement type first hydraulic pump motor and a second hydraulic pump motor having both functions of a pump and a motor, wherein the continuously variable transmission includes an input element to which power is input from the power source, A first differential mechanism that performs a differential action between an output element that outputs power to the output member via a transmission mechanism and a reaction force element that is connected to the first hydraulic pump motor, and power from the power source. Differential action between another input element that is input, another output element that outputs power to the output member via another transmission mechanism, and another reaction element that is connected to the second hydraulic pump motor. Second differential mechanism A switching mechanism that selectively transmits torque to each of the transmission mechanisms, and the lubrication-required portion is always connected to the first and second differential mechanisms and the first and second differential mechanisms. Alternatively, the lubrication circuit may include a first flow rate control valve whose flow rate changes in conjunction with a change in the capacity of the first hydraulic pump motor. In addition, the flow rate changes in conjunction with the change in the capacity of the first supply oil passage for supplying the hydraulic oil from the supply pump to the first hydraulic pump motor or the second hydraulic pump motor, and the second hydraulic pump motor. A second flow rate control valve and a second supply oil path for supplying the hydraulic oil from the replenishment pump to the second hydraulic pump motor or the first hydraulic pump motor; and the first hydraulic pump A first connection oil passage for flowing the hydraulic oil supplied to the motor to the lubrication-required portion; a second connection oil passage for flowing the hydraulic oil supplied to the second hydraulic pump motor to the lubrication-required portion; The actuating device includes a first electromagnetic control valve that can be electrically controlled by linking a capacity of the first hydraulic pump motor and a flow rate of the first flow control valve, and a second hydraulic pump motor. a hydraulic control system of the continuously variable transmission, wherein it to contain electrically controllable second electromagnetic control valve in conjunction with the flow capacity and the second flow control valve.

According to a fourth aspect of the present invention, in the third aspect of the invention, the first electromagnetic control valve is configured such that the flow rate of the first supply oil passage decreases as the capacity of the first hydraulic pump motor increases. The first hydraulic pump motor and the first flow control valve are configured to be controllable, and the second electromagnetic control valve decreases the flow rate of the second supply oil passage as the capacity of the second hydraulic pump motor increases. Thus, the hydraulic control device for a continuously variable transmission is configured to be able to control the second hydraulic pump motor and the second flow rate control valve.

Further, the invention of claim 3, in the invention of claim 1 or 2, wherein the lubricating circuit, connects the hydraulic pump motor of a relatively low speed and the supply pump by the first oil supply passage And a first supply state in which the replenishment pump and the hydraulic pump motor having a relatively high rotation speed are connected by the second supply oil path, and the replenishment pump and the relative position by the first supply oil path. A second switching state is selectively switched to a second supply state in which a high-speed hydraulic pump motor is connected and the supply pump and the relatively low-speed hydraulic pump motor are connected by the second supply oil passage. A one-way switching valve is connected to the relatively low rotational speed hydraulic pump motor and the relatively low rotational speed hydraulic pump motor side by the first connecting oil passage; And by the second connecting oil passage A first connection state in which the relatively high rotational speed hydraulic pump motor is connected to the lubrication-required portion disposed on the relatively high rotational speed hydraulic pump motor side, and the first connection oil passage allows the A relatively low rotational speed hydraulic pump motor is connected to the lubrication-needed portion disposed on the relatively high rotational speed hydraulic pump motor side, and the relatively high rotational speed is achieved by the second connection oil passage. A second directional control valve that selectively switches between a number of the hydraulic pump motors and a second connection state in which the lubrication-needed portions disposed on the relatively low rotational speed hydraulic pump motor side are connected, Oil temperature detecting means for detecting the temperature of the hydraulic oil, and switching valve control means for controlling the switching state of the first and second directional switching valves based on the oil temperature detected by the oil temperature detecting means. More A hydraulic control system of the continuously variable transmission according to claim.

According to a fourth aspect of the present invention, in the third aspect of the present invention, the switching valve control means is configured such that when the oil temperature detected by the oil temperature detecting means is equal to or lower than a predetermined temperature, the first A drag reduction mode is set in which the supply oil path and the second supply oil path are set to the first supply state, and the first connection oil path and the second connection oil path are set to the first connection state. When the temperature is higher than the predetermined temperature, the first supply oil path and the second supply oil path are set to the second supply state, and the first connection oil path and the second connection oil path are set to the second connection state. A hydraulic control device for a continuously variable transmission, comprising means for setting a cooling promotion mode.

According to a fifth aspect of the present invention, there is provided a relief valve according to any one of the first to fourth aspects, further comprising a relief valve that sets a hydraulic pressure discharged from the replenishing pump to a predetermined pressure, and the lubricating circuit includes the relief circuit. The hydraulic control device for a continuously variable transmission includes a circuit for supplying pressure oil discharged from a valve as the hydraulic oil to the hydraulic pump and a hydraulic motor.

  According to the first aspect of the present invention, the variable displacement hydraulic pump is driven by the power source, and the hydraulic pressure is supplied from the hydraulic pump to the variable displacement hydraulic motor to drive it. That is, power is transmitted through the fluid, and the transmission ratio is continuously changed by continuously changing the transmission amount of power through the hydraulic pressure. By transmitting the power in this way, heat is generated in the hydraulic pump, the hydraulic motor or the power transmission part in the continuously variable transmission, so that cooling or lubricating hydraulic oil is supplied to the lubrication circuit to cool them. Then, the power transmission part of the hydraulic pump, the hydraulic motor, or the continuously variable transmission, that is, the lubrication required part is cooled or lubricated by the hydraulic oil. In this case, if the capacity of the variable displacement hydraulic pump or hydraulic motor is changed, and the rotational state of the hydraulic pump or hydraulic motor changes, the required amount of hydraulic oil in the hydraulic pump, hydraulic motor, or lubrication required part However, when the capacity of a variable displacement hydraulic pump or hydraulic motor is changed, the amount of hydraulic oil, that is, the flow rate of hydraulic oil supplied to the hydraulic pump, hydraulic motor, or lubrication required part changes the capacity. It can be changed simultaneously in conjunction with. Therefore, according to changes in the rotation state of the hydraulic pump and hydraulic motor, the amount of hydraulic oil supplied to the hydraulic pump, hydraulic motor, or lubrication required part can be easily controlled without providing a special control system. can do. As a result, even when the rotational state of the variable displacement hydraulic pump or hydraulic motor changes, the hydraulic oil supplied to the hydraulic pump, hydraulic motor, or lubrication required part is insufficient, or the hydraulic oil is excessive. The supply of hydraulic oil can be performed efficiently by avoiding being supplied.

Further, variable displacement hydraulic pump motor gives a reaction force to the differential mechanism, the by the torque input from the reaction force and the power source is synthesized, through a transmission mechanism and the switching mechanism from the output element Power is transmitted to the output member. Since at least two such power transmission systems are provided, if one hydraulic pump motor runs idle and the other hydraulic pump motor outputs a reaction force, power is output via one power transmission system. Thus, the transmission gear ratio determined by the transmission mechanism in the power transmission system can be set. In addition, if the two power transmission systems are in a state where torque can be transmitted by the respective switching mechanisms, and the power is transmitted by exchanging pressure oil between each hydraulic pump motor in that state, torque can be transmitted by the switching mechanism. It is possible to set a gear ratio that is an intermediate value of the gear ratio determined by the transmission mechanism. Since the intermediate gear ratio changes according to the rate of fluid transmission, the overall gear ratio changes steplessly. In the state where the gear ratio is set in this way, for example, the amount of heat generated by one of the hydraulic pump motors is greater than the amount of heat generated by the other hydraulic pump motor, so There is a difference in the required amount of hydraulic fluid for cooling or lubrication between the lubrication required parts connected to the pump motor, but the change in the rotation state of each hydraulic pump motor, that is, the required hydraulic fluid amount for each hydraulic pump motor In accordance with this change, the flow rate of hydraulic oil supplied to each hydraulic pump motor or each lubrication required site is controlled to change simultaneously in conjunction with the change in the capacity of each hydraulic pump motor. Therefore, the amount of hydraulic oil supplied to each hydraulic pump motor or the lubrication required portion can be appropriately and easily controlled in accordance with a change in the rotation state of each hydraulic pump motor.

Then, the flow rate of the cooling or lubrication hydraulic oil supplied to each hydraulic pump motor is adjusted by each flow control valve provided in each supply oil path connected from the replenishment pump to each hydraulic pump motor. Control of the flow rate of each of these flow control valves is performed simultaneously with the control of the capacity of each hydraulic pump motor by an electromagnetic control valve. Therefore, it is possible to easily control the amount of hydraulic oil supplied to each hydraulic pump motor or each necessary lubrication site without providing a special control system.

According to the second aspect of the present invention, for example, when the capacity of the first hydraulic pump motor is set to the maximum, the first hydraulic pump motor generates less heat by locking its rotation. Accordingly, the amount of cooling hydraulic oil required by the first pump motor may be small. On the other hand, when the capacity of the first hydraulic pump motor is set to the minimum, the first hydraulic pump motor idles and is driven at a high speed to increase the amount of heat generation. Accordingly, the first pump motor requires a large amount of hydraulic fluid for cooling. That is, the required amount of cooling hydraulic oil for the hydraulic pump motor changes in inverse proportion to the change in the capacity of the hydraulic pump motor. On the other hand, in the invention of claim 2, when controlling each flow control valve provided in each supply oil passage by each electromagnetic control valve in conjunction with the capacity control unit of each hydraulic pump motor, The capacity of the first hydraulic pump motor and the flow rate of the hydraulic oil in the first supply oil passage are changed in inverse proportion, and the capacity of the second hydraulic pump motor and the flow rate of the hydraulic oil in the second supply oil passage are inversely proportional. Thus, the flow rate of each flow control valve and the capacity of each hydraulic pump motor are controlled in conjunction with each other. Therefore, an electromagnetic control valve that controls the capacity of each hydraulic pump motor is used for cooling or lubrication operation in response to changes in the required amount of hydraulic oil for cooling or lubrication due to changes in the capacity of each hydraulic pump motor. The amount of oil supply can be controlled appropriately and easily.

According to the invention of claim 3, the flow rate of the hydraulic oil supplied to each hydraulic pump motor based on the oil temperature of the cooling or lubricating hydraulic oil supplied to each hydraulic pump motor and the lubrication required portion, In addition, the flow direction of the hydraulic oil from each hydraulic pump motor to each lubrication necessary portion connected to each hydraulic pump motor can be appropriately switched and set according to the rotation state of each hydraulic pump motor.

Further, according to the invention of claim 4, by appropriately controlling each direction switching valve according to the oil temperature of the hydraulic oil, for example, the oil temperature of the hydraulic oil is low and the viscosity of the hydraulic oil is high. In this case, it is necessary to lubricate the hydraulic oil that is supplied to the relatively high speed hydraulic pump motor and has a relatively high oil temperature and that is idling at high speed without transmitting power. It is possible to set a drag reduction mode to be distributed to the part. Alternatively, when the oil temperature of the hydraulic oil is high and it is necessary to suppress an increase in the oil temperature of the hydraulic oil, the oil temperature is supplied to the hydraulic pump motor that is relatively low in rotation and the oil temperature is relatively low. It is possible to set a cooling acceleration mode in which the operating oil is circulated to the lubrication-needed portion on the side where heat is generated by performing power transmission. Therefore, drag loss due to the viscous resistance of the hydraulic oil can be reduced, and an increase in the oil temperature of the hydraulic oil supplied to the lubrication-required part can be suppressed to improve the cooling effect of the lubrication-needed part.

And, according to the invention of claim 5, since the pressure oil discharged from the relief valve is used as a working oil for cooling or lubrication, it becomes possible to perform cooling using existing equipment and devices, As a result, efficient cooling or lubrication is possible such that no power is newly consumed for cooling or lubrication, and simplification and miniaturization of the apparatus can be achieved.

  Next, the present invention will be described based on specific examples. First, the transmission targeted by the present invention will be described. In the example shown in FIG. 1, four forward speeds and one reverse speed are set as so-called fixed gear ratios that can be set without changing the form of power (energy) to be transmitted. This is an example configured to set the stage, and in particular, an example configured to be suitable for a so-called FF vehicle (front engine / front drive vehicle) in which the power source 1 such as an engine is mounted in the width direction of the vehicle. It is.

  In FIG. 1, an input member 2 is connected to a power source (E / G) 1, and a first planetary gear mechanism (PG1) 3 corresponding to the first and second differential mechanisms in the present invention is connected from the input member 2. The second planetary gear mechanism (PG2) 4 is configured to transmit torque. Here, the power source 1 may be a general power source used in a vehicle such as an internal combustion engine, an electric motor, or a combination thereof. In the following description, the power source 1 is temporarily referred to as the engine 1. Moreover, the input member 2 should just be a member which can transmit the motive power which the engine (E / G) 1 output, and may be a drive plate or an input shaft. In addition, appropriate transmission means such as a damper, a clutch, a torque converter, and the like can be interposed between the engine 1 and the input member 2.

  The first planetary gear mechanism 3 is disposed on the same axis as the input member 2, the second planetary gear mechanism 4 is separated outward in the radial direction of the first planetary gear mechanism 3, and the respective central axes are parallel to each other. They are arranged in parallel. As these planetary gear mechanisms 3 and 4, a planetary gear mechanism of an appropriate type such as a single pinion type or a double pinion type can be adopted. The example shown in FIG. 1 is an example constituted by a single pinion type planetary gear mechanism, and is a sun gear 3S, 4S that is an external gear, and a ring gear 3R that is an internal gear arranged concentrically with the sun gear 3S, 4S. , 4R, and carriers 3C, 4C holding pinion gears meshed with these sun gears 3S, 4S and ring gears 3R, 4R so as to be rotatable and revolved. The input member 2 is connected to a ring gear 3R in the first planetary gear mechanism 3, and the ring gear 3R serves as an input element.

  Further, a counter drive gear 5 is attached to the input member 2, and an idle gear 6 is engaged with the counter drive gear 5, and a counter driven gear 7 is engaged with the idle gear 6. The counter driven gear 7 is disposed on the same axis as the second planetary gear mechanism 4 and is connected to the ring gear 4R of the second planetary gear mechanism 4 so as to rotate together. Therefore, in the second planetary gear mechanism 4, the ring gear 4R serves as an input element. The ring gears 3R and 4R, which are input elements of the planetary gear mechanisms 3 and 4, are configured so that the counter gear pair includes the idle gear 6, and thus rotate in the same direction.

  The carrier 3C in the first planetary gear mechanism 3 serves as an output element, and the first intermediate shaft 8 is connected to the carrier 3C so as to rotate together. The first intermediate shaft 8 is a hollow shaft into which a motor shaft 9 is rotatably inserted. One end of the motor shaft 9 is a sun gear that is a reaction force element in the first planetary gear mechanism 3. It is connected to 3S so as to rotate together.

  The second planetary gear mechanism 4 has the same configuration, and the carrier 4C serves as an output element, and the second intermediate shaft 10 is connected to the carrier 4C so as to rotate together. The second intermediate shaft 10 is a hollow shaft into which a motor shaft 11 is rotatably inserted. One end of the motor shaft 11 is a sun gear that is a reaction force element in the second planetary gear mechanism 4. 4S is connected to rotate integrally.

  The other end of the motor shaft 9 is connected to the output shaft of the variable displacement hydraulic pump motor 12. The variable displacement hydraulic pump motor 12 is a fluid pressure (hydraulic) pump capable of changing the discharge capacity, such as a slant shaft pump, a swash plate pump, or a radial piston pump, and rotates by giving torque to its output shaft. Therefore, the pressure fluid (pressure oil) is discharged by functioning as a pump, and the pressure fluid is supplied from the discharge port or the suction port, thereby functioning as a motor. In the following description, the variable displacement hydraulic pump motor 12 is referred to as a first pump motor 12 and is indicated as PM1 in the figure.

  On the other hand, the other end of the motor shaft 11 is connected to the output shaft of the variable displacement hydraulic pump motor 13. The variable displacement hydraulic pump motor 13 has the same configuration as the first pump motor 12 on the motor shaft 59 side. In the following description, the variable displacement hydraulic pump motor 13 is referred to as a second pump motor 13 and is indicated as PM2 in the figure.

  The hydraulic pump motors 12 and 13 are communicated with each other through oil passages 14 and 15 so that the pressure oil, which is a pressure fluid, can be transferred to each other. That is, the suction ports 12 </ b> S and 13 </ b> S are communicated with each other through the oil passage 14, and the discharge ports 12 </ b> D and 13 </ b> D are communicated with each other through the oil passage 15. Accordingly, a closed circuit C0 is formed by the oil passages 14 and 15.

  An output shaft 16 corresponding to the output member of the present invention is arranged in parallel with each of the intermediate shafts 8 and 10 described above. A transmission mechanism for setting a predetermined gear ratio is provided between the output shaft 16 and each of the intermediate shafts 8 and 10. The transmission mechanism in the present invention is not limited to a mechanism that transmits power at a fixed gear ratio, and a mechanism with a variable gear ratio can be adopted. In the example shown in FIG. 1, power is transmitted at a fixed gear ratio. A plurality of gear pairs 17, 18, 19, and 20 for transmitting the above are employed.

  More specifically, a fourth speed drive gear 17A and a second speed drive gear 18A are arranged on the first intermediate shaft 8 in order from the first planetary gear mechanism 3 side. 17A and the second speed drive gear 18A are rotatably fitted to the first intermediate shaft 8. A fourth speed driven gear 17B meshed with the fourth speed drive gear 17A and a second speed driven gear 18B meshed with the second speed drive gear 18A are attached to the output shaft 16 so as to rotate integrally. Yes.

  Further, a third speed drive gear 19A meshed with the fourth speed driven gear 17B and a first speed drive gear 20A meshed with the second speed driven gear 18B are rotatable on the second intermediate shaft 10. It is made to fit. Accordingly, the fourth speed driven gear 17B also serves as the third speed driven gear, and the second speed driven gear 18B also serves as the first speed driven gear. Here, the gear ratio of each gear pair 17, 18, 19, and 20 (ratio of the number of teeth of the driven gear to the number of teeth of each drive gear) will be described. The second speed gear pair 18, the third speed gear pair 19, and the fourth speed gear pair 17 are configured to decrease in order.

  Furthermore, a starting gear pair 21 is provided. The starting gear pair 21 is used to transmit the power to the output shaft 16 together with the first speed gear pair 20 so that the driving force at the time of starting is sufficiently large. A start drive gear 21A attached to the motor shaft 9 on the pump motor 12 side and a start driven gear 21B attached to the output shaft 16 so as to be rotatable are provided.

  A clutch mechanism is provided for selectively allowing each of the gear pairs 17, 18, 19, 20, and 21 described above to transmit power. This clutch mechanism is a mechanism for selectively connecting each gear pair 17, 18, 19, 20, 21 to any one of the intermediate shafts 8, 10 and the output shaft 16, and corresponds to the switching mechanism of the present invention. is doing. Therefore, a synchronous coupling mechanism (synchronizer) in a conventional manual transmission or the like can be used, or a meshing clutch (dog clutch), a friction clutch, or the like can be used. When the driven gear is integrally attached to the output shaft 16, the drive gear is rotatable with respect to the intermediate shafts 8 and 10, and the drive gear is selectively selected with respect to the intermediate shafts 8 and 10. A switching mechanism can be provided on the intermediate shafts 8 and 10 side so as to be connected. FIG. 1 shows an example in which a synchronizer is employed.

  The synchronizer basically moves the sleeve that rotates together with the rotating shaft in the axial direction, and engages with the spline of the rotating member that is mounted so as to rotate relative to the rotating shaft. The rotating shaft and the rotating member are synchronized with each other by gradually bringing the kniter ring into frictional contact with the rotating member, thereby connecting the rotating shaft and the rotating member. On the output shaft 16, a first synchronizer (hereinafter referred to as a first synchronizer) 22 is provided at a position adjacent to the starter driven gear 21B. The first sync 22 moves the sleeve 22S to the left side of FIG. 1 to connect the starter driven gear 21B to the output shaft 16, and the starter gear pair 21 is connected between the motor shaft 9 and the output shaft 16. It is configured to transmit torque.

  On the second intermediate shaft 10, a second synchronizer (hereinafter referred to as a second synchronizer) 23 is provided between the third speed drive gear 19A and the first speed drive gear 20A. The second synchronizer 23 connects the first speed drive gear 20A to the second intermediate shaft 10 by moving the sleeve 23S to the left side in FIG. 1, and the first speed gear pair 20 is connected to the second intermediate shaft 10. And the output shaft 16 are configured to transmit torque. On the other hand, by moving the sleeve 23S to the right in FIG. 1, the third speed drive gear 19A is connected to the second intermediate shaft 10, and the third speed gear pair 19 is connected to the second intermediate shaft 10 and the output shaft. 16 is configured to transmit torque.

  Further, on the first intermediate shaft 8, a third synchronizer (hereinafter referred to as a third synchronizer) 24 is provided between the second speed drive gear 18A and the fourth speed drive gear 17A. The third sync 24 moves the sleeve 24S to the left side in FIG. 1 to connect the second speed drive gear 18A to the first intermediate shaft 8 and the second speed gear pair 18 is connected to the first intermediate shaft 8. And the output shaft 16 are configured to transmit torque. On the other hand, by moving the sleeve 24S to the right in FIG. 1, the fourth speed drive gear 17A is connected to the first intermediate shaft 8, and the fourth speed gear pair 17 is connected to the first intermediate shaft 8 and the output shaft. 16 is configured to transmit torque.

  Furthermore, a reverse synchronizer (hereinafter referred to as “R synchro”) 25 is provided on the motor shaft 11 on the second pump motor 13 side at a position adjacent to the shaft end of the second intermediate shaft 10. The R synchro 25 connects the motor shaft 11 and the second intermediate shaft 10, that is, the sun gear 4S and the carrier 4C in the second planetary gear mechanism 4 by moving the sleeve 25S to the right in FIG. The entire planetary gear mechanism 4 is configured to rotate integrally.

  Each of the synchros 22, 23, 24, and 25 can be configured to be switched by a manual operation, but can be configured to perform so-called automatic control instead. In that case, for example, an appropriate actuator (not shown) for moving the above-described sleeve in the axial direction may be provided, and the actuator may be electrically controlled.

  As described above, the transmission shown in FIG. 1 is configured such that the torque output from the engine 1 is transmitted to the output shaft 16 via any one of the intermediate shafts 8 and 10 or the motor shafts 9 and 11. Yes. A differential 27 is connected to the output shaft 16 through a transmission means 26 such as a gear mechanism or a winding transmission mechanism such as a chain, and power is output to the left and right axles 28 from here.

  A sensor for detecting the operating state of the transmission is provided. Specifically, the input rotation speed sensor 29 for detecting the rotation speed of the input member 2 or the counter drive gear 5 integrated therewith, the output rotation speed sensor 30 for detecting the rotation speed of the axle 28, and the first pump motor. 12 or the rotation speed sensor 31 for detecting the rotation speed of the sun gear 3S to which the first pump motor 12 is connected, and the rotation for detecting the rotation speed of the sun gear 4S to which the second pump motor 13 or the second pump motor 13 is connected. A number sensor 32, an oil temperature sensor 33 for detecting the temperature of hydraulic oil in the transmission, and the like are provided.

  As described above, the hydraulic control device for a continuously variable transmission according to the present invention appropriately controls the amount of oil supplied for cooling or lubrication regardless of the operating state of the transmission and the rotational state of the hydraulic pump motor. The hydraulic control circuit for this purpose is configured as follows.

(First embodiment)
FIG. 1 is a diagram schematically showing a first embodiment of a hydraulic control circuit constituting the hydraulic control device of the present invention. In FIG. 1, the suction ports 12 </ b> S and 13 </ b> S of the hydraulic pump motors 12 and 13 are communicated with each other through an oil passage 14, and the discharge ports 12 </ b> D and 13 </ b> D are communicated with each other through an oil passage 15. Here, the suction ports 12S and 13S are ports for sucking in pressure oil by rotating forward with the extrusion volume (pump capacity) set in the positive direction, and the ports for discharging the pressure oil are discharge ports. 12D and 13D. Accordingly, the hydraulic pump motors 12 and 13 are communicated with each other by a closed circuit C0 that circulates pressure oil between them.

  The closed circuit C0 is provided with a charge pump (sometimes referred to as a boost pump) 35 for supplying pressure oil. The charge pump 35 compensates for oil shortage due to leakage from the closed circuit C0 and corresponds to the replenishment pump of the present invention. The charge pump 35 is driven by the above-described engine 1, a motor (not shown), etc. Oil is pumped from 36 and supplied to the closed circuit C0.

  Therefore, the discharge port of the charge pump 35 communicates with the oil passage 14 and the oil passage 15 in the closed circuit C0 via the check valves 37 and 38, respectively. The check valves 37 and 38 are configured to open in the discharge direction from the charge pump 35 and close in the opposite direction. Further, a charge pressure control valve (boost pressure control valve) 39 for adjusting the discharge pressure of the charge pump 35 is communicated with the discharge port of the charge pump 35. The charge pressure control valve 39 is configured to open and discharge oil when a pressure higher than the sum of the elastic force of the spring and the pilot pressure or the pressing force of the solenoid is applied. The pressure is set to a pressure corresponding to the pilot pressure.

  Further, an electromagnetic relief valve 40 is provided between the suction port 12 </ b> S of the first pump motor 12 and the oil passage 15. In other words, the electromagnetic relief valve 40 is provided in parallel with the first pump motor 12 so as to communicate the oil passages 14 and 15. The electromagnetic relief valve 40 is configured to maintain the discharge pressure at a preset pressure when pressure oil is discharged from the suction port 12S of the first pump motor 12 or the suction port 13S of the second pump motor 13. Has been. An electromagnetic relief valve 41 is provided between the discharge port 13 </ b> D of the second pump motor 13 and the oil passage 14. In other words, the electromagnetic relief valve 41 is provided in parallel with the second pump motor 13 so as to communicate the oil passages 14 and 15. The electromagnetic relief valve 41 is configured to maintain the discharge pressure at a preset pressure when pressure oil is discharged from the discharge port 12D of the first pump motor 12 or the discharge port 13D of the second pump motor 13. Has been.

  Further, an oil passage is provided for supplying the hydraulic oil discharged from the charge pressure control valve 39 to the hydraulic pump motors 12 and 13 as hydraulic or oil for cooling or lubrication. That is, this oil passage communicates the discharge port of the charge pressure control valve 39 with the oil passage in the casing of each hydraulic pump motor 12, 13, and the pressure oil discharged from the discharge port of the charge pressure control valve 39. Is branched and supplied to the hydraulic pump motors 12 and 13. That is, one supply oil passage 42 corresponding to the first supply oil passage of the present invention is connected to a cooling oil passage (not shown) formed in the casing 12 </ b> C of the first pump motor 12. The supply oil passage 42 is provided with a flow rate control valve 43 that can control the flow rate of the flowing hydraulic oil and corresponds to the first flow rate control valve of the present invention. Further, a return oil passage 44 is communicated with the cooling oil passage in the first pump motor 12, and the pressure oil for cooling discharged from the first pump motor 12 is guided to the oil sump 45.

  Further, a connection oil passage 46 corresponding to the first connection oil passage of the present invention is formed by branching from a connection portion of the supply oil passage 42 with the cooling oil passage formed in the casing 12C. The other end of the connection oil passage 46 is connected to the second planetary gear mechanism 4 connected to the second pump motor 13 via the motor shaft 11 and the sun gear 4S, and to the second planetary gear mechanism 4. It is connected to a lubrication required portion D2 such as a gear pair or a synchro which is connected or selectively connected. Therefore, the oil supplied to the cooling oil passage of the first pump motor 12 via the supply oil passage 42 circulates through the cooling oil passage, and then the lubrication required portion D2 via the connection oil passage 46. To be supplied.

  Similarly, the other supply oil passage 47 branched from the one supply oil passage 42 and corresponding to the second supply oil passage of the present invention is formed in the casing 13C of the second pump motor 13 for cooling. Are connected to an oil passage (not shown). Further, the supply oil passage 47 is provided with 48 corresponding to the second flow rate control valve of the present invention, which can control the flow rate of the flowing hydraulic oil. Further, a return oil passage 49 is communicated with the cooling oil passage in the second pump motor 13 so that the cooling pressure oil discharged from the second pump motor 13 is guided to the oil sump 45.

  Further, a connection oil passage 50 corresponding to the second connection oil passage of the present invention is formed by branching from a connection portion of the supply oil passage 47 with the cooling oil passage formed in the casing 13C. The other end of the connection oil passage 50 is connected to the first planetary gear mechanism 3 connected to the first pump motor 12 via the motor shaft 9 and the sun gear 3S, and to the first planetary gear mechanism 3. It is connected to a lubrication-required part D1 such as a gear pair or a synchro that is connected or selectively connected. Therefore, the oil supplied to the cooling oil passage of the second pump motor 13 via the supply oil passage 47 circulates through the cooling oil passage, and then the lubrication required portion D1 via the connection oil passage 50. To be supplied.

  Exhaust pressure from the electromagnetic relief valves 40 and 41 is guided to the oil reservoir 45, and pressure oil is further returned from the oil reservoir 45 to the oil pan 36. Further, a relief valve 51 is provided branching from the oil passage between the charge pressure control valve 39 and the supply oil passages 42 and 47, and the pressure oil discharged from the relief valve 51 is stored in the oil sump 45. Configured to lead to. Further, the pressure oil in the oil passage 14 and the oil passage 15 in the closed circuit C0 is guided to the oil sump 45 through the check valves 52 and 53 and the flow rate control valve 54, respectively. The check valves 52 and 53 are configured to open in the flow direction in which the pressure oil flows from the oil passages 14 and 15 to the oil sump 45 and to close in the opposite direction. An oil cooler 55 is provided in the oil path between the oil sump 45 and the oil pan 36.

  As described above, the supply oil passages 42 and 47, the connection oil passages 46 and 50, the flow rate control valves 43 and 48, the return oil passages 44 and 49, and the like constitute the circuit C1 corresponding to the lubricating circuit of the present invention. Has been.

  Actuators 56 and 57 for adjusting the extrusion volumes or capacities of the hydraulic pump motors 12 and 13 are provided in the hydraulic pump motors 12 and 13, respectively. That is, the first pump motor 12 is provided with an electromagnetic control valve (linear solenoid valve) 58 corresponding to the first electromagnetic control valve of the present invention and an actuator 59 operated by the electromagnetic control valve 58. The electromagnetic control valve 58 is supplied with hydraulic pressure from the oil passage 14 or the oil passage 15 in the closed circuit C0, and the control hydraulic pressure is supplied from the electromagnetic control valve 58 to the actuator 59, whereby the operation of the actuator 59 is controlled.

  The actuator 59 is configured to operate a capacity control unit (not shown) of the first pump motor 12, and when the actuator 59 is operated, the capacity control unit of the first pump motor 12 operates. The extrusion volume of one pump motor 12 is changed. Therefore, the variable displacement control of the first pump motor 12 is executed by controlling the electromagnetic control valve 58 and operating the actuator 59.

  The actuator 59 is configured to operate the capacity control unit of the first pump motor 12 as described above and to operate the flow rate control valve 43 provided in the supply oil passage 42. That is, when the actuator 59 operates, the capacity control unit of the first pump motor 12 operates, and at the same time, the flow control valve 43 operates. In other words, when the actuator 59 operates, the capacity control unit of the first pump motor 12 and the flow rate control valve 43 operate in conjunction with each other. Therefore, by controlling the electromagnetic control valve 58 and operating the actuator 59, the variable displacement control of the first pump motor 12 and the flow rate control of the supply oil passage 42 by the flow rate control valve 43 are simultaneously performed in conjunction with each other. It is configured to be able to.

  In the variable displacement control of the first pump motor 12 and the flow rate control by the flow rate control valve 43, the increase / decrease direction of the capacity during the variable displacement control is opposite to the increase / decrease direction of the flow rate during the flow rate control. ing. That is, as described above, the operating device 56 constituted by the electromagnetic control valve 58 and the actuator 59 controls the electromagnetic control valve 58 to operate the actuator 59, as shown in FIG. 12, the flow rate of the flow control valve 43, that is, the flow rate of the supply oil passage 42 decreases, and conversely, the flow rate of the flow control valve 43, that is, the supply oil passage increases as the capacity of the first pump motor 12 decreases. It is comprised so that the flow volume of 42 may increase.

  On the other hand, the second pump motor 13 has an electromagnetic control valve 60 corresponding to the second electromagnetic control valve of the present invention and an actuator 61 operated by the electromagnetic control valve 60, similarly to the first pump motor 12. Is provided. The electromagnetic control valve 60 is supplied with hydraulic pressure from the oil passage 14 or the oil passage 15 in the closed circuit C0, and the control hydraulic pressure is supplied from the electromagnetic control valve 60 to the actuator 61, whereby the operation of the actuator 61 is controlled.

  The actuator 61 is configured to operate a capacity control unit (not shown) of the second pump motor 13, and when the actuator 61 is operated, the capacity control unit of the second pump motor 13 operates. The extrusion volume of the two-pump motor 13 is changed. Therefore, the variable displacement control of the second pump motor 13 is executed by operating the actuator 61 by controlling the electromagnetic control valve 60.

  The actuator 61 is configured to operate the capacity control unit of the second pump motor 13 as described above and to operate the flow control valve 48 provided in the supply oil passage 47. That is, when the actuator 61 operates, the capacity control unit of the second pump motor 13 operates, and at the same time, the flow control valve 48 operates. In other words, when the actuator 61 operates, the capacity control unit of the second pump motor 13 and the flow rate control valve 48 operate in conjunction with each other. Therefore, by controlling the electromagnetic control valve 60 and operating the actuator 61, the variable displacement control of the second pump motor 13 and the flow rate control of the supply oil passage 47 by the flow rate control valve 48 are simultaneously performed in conjunction with each other. It is configured to be able to.

  In the variable capacity control of the second pump motor 13 and the flow control by the flow control valve 48, the direction of increase / decrease in capacity during the variable capacity control is opposite to the direction of increase / decrease in flow during the flow control. It has become. That is, as described above, the actuator 57 composed of the electromagnetic control valve 60 and the actuator 61 controls the second pump motor as shown in FIG. 2 when the actuator 61 is operated by controlling the electromagnetic control valve 60. 13, the flow rate of the flow control valve 48, that is, the flow rate of the supply oil passage 47 decreases, and conversely, the flow rate of the flow control valve 48, that is, the supply oil passage increases, as the capacity of the second pump motor 13 decreases. 47 is configured to increase the flow rate.

  The control pressures of the electromagnetic control valves 58 and 60 (that is, the capacities of the hydraulic pump motors 12 and 13 and the flow rates in the supply oil passages 42 and 47), the switching operation of the synchros 22, 23 and 24, and the charge The pressure control valve 39 and the relief pressures of the electromagnetic relief valves 40 and 41 can be electrically controlled, and an electronic control unit (ECU) 62 is provided for that purpose. The electronic control unit 62 is configured mainly with a microcomputer, and receives the number of rotations of a predetermined rotating member and other detection signals, and the input signals and information stored in advance. In addition, the calculation is performed based on the program, and the command signal is output according to the calculation result.

  Here, the operation of the above-described transmission will be described. FIG. 3 is a chart collectively showing the operating states of the hydraulic pump motors (PM1, PM2) 12, 13 and the synchros 22, 23, 24, 25 when setting the respective gear positions. “OFF” for each of the hydraulic pump motors 12 and 13 in FIG. 4 causes the pump capacity to be substantially zero, does not generate pressure oil even if its output shaft is rotated, and is output even if hydraulic pressure is supplied. A state where the shaft does not rotate (free) is indicated, and “LOCK” indicates a state where the rotation of the rotor is stopped. Furthermore, “hydraulic pressure generation” indicates a state in which the pump capacity is made larger than substantially zero and pressure oil is discharged, and thus the corresponding hydraulic pump motors 12 and 13 function as pumps. “Hydraulic recovery” indicates a state in which the hydraulic oil discharged from one hydraulic pump motor 13 (or 12) is supplied and functions as a motor. Therefore, the corresponding hydraulic pump motor 12 (or 13) is a shaft. Torque is generated and driving torque is transmitted to the corresponding motor shafts 9 and 11 and intermediate shafts 8 and 10.

  In addition, “right” and “left” for each of the synchros 22, 23, 24, and 25 indicate the positions of the sleeves in the respective synchros 22, 23, 24, and 25 in FIG. 1, and the parentheses are downshifted. The stand-by state, the brackets indicate the stand-by state for upshifting, and “◯” reduces drag by setting the corresponding synchro 22, 23, 24, 25 to the OFF state (neutral position). “●” indicates that the corresponding synchro 22, 23, 24, 25 is set to the OFF state (neutral position) and is in the neutral state.

  When the neutral position is selected and the neutral (N) state is set, the hydraulic pump motors 12 and 13 are set to the “OFF” state, and the sleeves of the synchros 22, 23, 24, and 25 are set to the center positions. Is done. Therefore, none of the gear pairs 17, 18, 19, 20, 21 is in a neutral state that is not connected to the output shaft 16. That is, the hydraulic pump motors 12 and 13 are controlled so that the extrusion volume becomes substantially zero. As a result, a so-called idling state is established, so that even if torque is transmitted from the engine 1 to the ring gears 3R, 4R of the planetary gear mechanisms 3, 4, no reaction force acts on the sun gears 3S, 4S. Therefore, torque is not transmitted to the intermediate shafts 8 and 10 connected to the carriers 3C and 4C that are output elements.

  When the shift position is switched to a travel position such as a drive position, the sleeve 22S of the first sync 22 is moved to the left in FIG. 1, and the sleeve 23S of the second sync 23 is moved to the left in FIG. Therefore, the starting drive gear 21A is connected to the motor shaft 9, the first pump motor 12 and the output shaft 16 are connected, and the first speed drive gear 20A is connected to the second intermediate shaft 10 to be the second planetary gear mechanism. 4 is the output element 4 and the output shaft 16 is connected. That is, the first speed that is the fixed gear ratio is set. At the same time, the extrusion volumes of the hydraulic pump motors 12 and 13 are controlled to be larger than zero.

  Therefore, the second pump motor 13 is driven by the power of the engine 1 distributed by the second planetary gear mechanism 4 and functions as a pump. Therefore, the second pump motor 13 gives reaction force torque accompanying generation of hydraulic pressure to the motor shaft 11 and the sun gear 4S. This is described as “hydraulic pressure generation” in FIG. Therefore, torque is transmitted to the carrier 4 </ b> C by the differential action of the second planetary gear mechanism 4, and the torque is transmitted to the output shaft 16 via the first speed gear pair 20.

  On the other hand, the hydraulic pressure generated by the second pump motor 13 is discharged from the suction port 13S and supplied to the suction port 12S of the first pump motor 12. As a result, the first pump motor 12 functions as a motor. This is shown in FIG. 3 as “hydraulic pressure recovery”. In this way, the power transmitted to the first pump motor 12 is transmitted to the output shaft 16 via the starting gear pair 21. Therefore, in the driving state from the start to the first speed, both so-called mechanical power transmission via the second planetary gear mechanism 4 and power transmission via the hydraulic pressure are generated, and the combined power of these powers is generated. Appears on the output shaft 16. Further, the gear ratio in this process becomes a value larger than the first speed which is a fixed gear ratio, and the gear ratio changes continuously or steplessly.

  Thus, when the rotation speed of the engine 1 and the vehicle speed change to the first gear ratio, the extrusion volume of the first pump motor 12 is set to zero, and the extrusion volume of the second pump motor 13 is set to the maximum. As a result, the rotation of the second pump motor 13 is substantially locked. That is, the motor shaft 11 and the second pump motor 13 connected thereto are fixed. At the same time, the first sync 22 is set to the OFF state. As a result, the sun gear 4S of the second planetary gear mechanism 4 is fixed, and the first planetary gear mechanism 3 is not involved in the transmission of power to the output shaft 16, so that the power output from the engine 1 is the second planetary gear mechanism. 4 and the first speed gear pair 20 are transmitted to the output shaft 16. That is, a fixed gear ratio determined by the gear ratio of the first speed gear pair 20 is set. In this case, since the first pump motor 12 and the motor shaft 9 connected thereto are idled, no torque appears on the first intermediate shaft 8.

  In the case of upshifting from the first speed that is the fixed gear ratio, the sleeve 24S of the third sync 24 is moved to the left side in FIG. 1 and the second speed drive gear 18A is connected to the first intermediate shaft 8. The R synchro 25 is kept in a neutral state. When the sleeve 24S of the third synchro 24 is engaged with the second speed drive gear 18A, synchronous control is performed so that the rotation speed of the sleeve 24S of the third synchro 24 matches the rotation speed of the second speed drive gear 18A. . The synchronization control is performed in the same manner when the sleeves of the synchros 22, 23, 24, and 25 are engaged with the mating members.

  In this state, the extrusion volume of the first pump motor 12 is gradually increased toward the maximum. In the standby state for upshifting to the second speed, the first pump motor 12 rotates in the reverse direction, and thus functions as a pump by gradually increasing the extrusion volume. That is, hydraulic pressure is generated (indicated as “hydraulic pressure generation” in FIG. 3), and at the same time, a reaction force torque associated therewith appears on the motor shaft 9. As a result, transmission of power through the first planetary gear mechanism 3 and the second speed gear pair 18 is gradually performed. Further, the hydraulic pressure generated by the first pump motor 12 is supplied to the second pump motor 13 and functions as a motor (denoted as “hydraulic pressure recovery” in FIG. 3), so the second pump motor 13 and the second planetary gear. Power is transmitted through the gear mechanism 4 and the first speed gear pair 20. Therefore, the speed ratio in the process of shifting from the first speed to the second speed is a value between the speed ratio of the first speed and the speed ratio of the second speed and is a continuously changing speed ratio. . That is, a continuously variable transmission state in which the gear ratio continuously changes is obtained. This is the same during the period from the start to the speed ratio of the first speed and between the fixed speed ratios. Therefore, the above-described transmission can function as a continuously variable transmission.

  The motor shaft 9 is substantially fixed by making the extrusion volume of the first pump motor 12 substantially maximum as described above and stopping its rotation or becoming nearly stopped. In addition, the second pump motor 13 is set to the OFF state. Therefore, in the first planetary gear mechanism 3, since the sun gear 3S is fixed, the power input to the ring gear 3R is output from the carrier 3C to the second speed drive gear 18A via the intermediate shaft 8. On the other hand, the second pump motor 13 is in the OFF state, and the R synchro 25 and the second synchro 23 arranged coaxially with the second pump motor 13 are in the OFF state and the sleeve 23S is in the neutral position. The pump motor 13 and the second planetary gear mechanism 4 are not involved in power transmission. Accordingly, the second speed, which is a fixed gear ratio determined by the gear ratio of the second speed gear pair 18, is set.

  Hereinafter, similarly, the third speed is achieved by moving the sleeve 23S of the second synchro 23 to the right side in FIG. 1 to connect the third speed drive gear 19A to the second intermediate shaft 10 and pushing the second pump motor 13 out. By maximizing the volume, similarly to the case of the first speed, the motor shaft 11 and the second pump motor 13 are fixed, and the other synchros 22 and 24 are turned off. Accordingly, the power is transmitted to the output shaft 16 via the third speed gear pair 19, and the third speed, which is a fixed gear ratio, is set. In the fourth speed, the sleeve 24S of the third sync 24 is moved to the right in FIG. 1 to connect the fourth speed drive gear 17A to the first intermediate shaft 8, and the first pump motor 12 has a maximum pushing volume. Thus, as in the case of the second speed, the motor shaft 9 and the first pump motor 12 are fixed, and the other synchros 23 and 25 are turned off. Therefore, power is transmitted to the output shaft 16 via the fourth speed gear pair 17 and the fourth speed, which is a fixed gear ratio, is set.

  Further, the reverse gear will be described. When the reverse position is selected, the sleeve 22S of the first sync 22 is moved to the left in FIG. 1, and the sleeve 25S of the R sync 25 is moved to the right in FIG. Further, the other synchros 23 and 24 are set to the OFF state. Therefore, when the second intermediate shaft 10 and the motor shaft 11 are connected by the R sync 25, the sun gear 4S of the second planetary gear mechanism 4 and the carrier 4C are connected, and the entire second planetary gear mechanism 4 is substantially the same. Integrated. The start drive gear 21 </ b> A is connected to the motor shaft 9, that is, the rotor of the first pump motor 12.

  Therefore, the motive power transmitted from the engine 1 to the second planetary gear mechanism 4 is transmitted to the second pump motor 13 as it is, and this is driven, and hydraulic pressure is generated by the second pump motor 13. Since the second synchro 23 is in the OFF state, power is not transmitted from the second planetary gear mechanism 4 or the second intermediate shaft 10 to the output shaft 16. On the other hand, the extrusion volume of the first pump motor 12 is controlled to a volume larger than zero, for example, the maximum volume. As a result, the first pump motor 12 functions as a motor by the hydraulic pressure supplied from the second pump motor 13 and outputs torque to the motor shaft 9. In this case, since the hydraulic pressure is supplied to the first pump motor 12 from the discharge port 12D, the first pump motor 12 rotates in the reverse direction. Then, the torque is transmitted to the output shaft 16 via the starting gear pair 21, so that a reverse state is established. That is, in the reverse speed, power is transmitted via hydraulic pressure, and in FIG. 3, this is indicated as “hydraulic pressure recovery” for the first pump motor 12 and “hydraulic pressure generation” for the second pump motor 13.

  In the above transmission according to the present invention, at least one of the hydraulic pump motors 12 and 13 rotates while the engine 1 is driven. Therefore, at least one of the hydraulic pump motors 12 and 13 generates heat due to pressure oil being pressurized or subjected to a shearing action, friction is generated, and pressure loss such as leakage occurs. These hydraulic pump motors 12, 13 are supplied with the pressure oil discharged from the relief valve 33 through the supply oil passages 42, 47, so that heat is taken away by the pressure oil and the hydraulic pump motors 12, 13 are supplied. 13 is cooled.

  Further, the amount of heat generated by each of the hydraulic pump motors 12 and 13 and the planetary gear mechanisms 3 and 4 varies depending on the operating state of the transmission or the set gear ratio. As described above, in the state where the fixed gear ratio is set, one of the hydraulic pump motors 12 and 13 is locked and stopped rotating, and the other hydraulic pump motors 13 and 12 are idled to pressurize the pressure oil. And there is not a little friction. When a so-called intermediate speed ratio between the fixed speed ratios is set, one of the hydraulic pump motors 12 and 13 functions as a pump and rotates at a rotational speed corresponding to the pushing volume and hydraulic pressure. The other hydraulic pump motors 13 and 12 function as motors by receiving the hydraulic pressure, rotate at a rotational speed corresponding to the extrusion volume, and output torque.

  FIG. 4 and FIG. 5 show this situation in a state between the first speed and the second speed, which are the fixed gear ratio. In FIG. 4, at the first speed, the extrusion volume q2 of the second pump motor 13 is set to zero, whereas the extrusion volume q1 of the first pump motor 12 is set to the maximum. The second pump motor 13 is idling at a predetermined rotational speed NPM2, while the first pump motor 12 is stopped. That is, the rotational speed NPM1 is zero.

  When upshifting from this state toward the second speed, which is the fixed gear ratio, as described above, the pushing volume q2 of the second pump motor 13 is gradually increased to function as a pump. In this case, the extrusion volume q1 of the first pump motor 12 is maintained at the maximum until the extrusion volume q2 of the second pump motor 13 is maximized. The extrusion volume q1 of the first pump motor 12 may be changed simultaneously with the extrusion volume q2 of the second pump motor 13.

  The rotation speed NPM2 of the second pump motor 13 gradually decreases as the extrusion volume q2 increases. On the other hand, the first pump motor 12 is supplied with pressure oil from the second pump motor 13. By functioning as a motor, its rotational speed NPM1 gradually increases.

  When the extrusion volume q2 of the second pump motor 13 is increased to the maximum, the extrusion volume q1 of the first pump motor 12 is gradually decreased and finally decreased to zero. In this process, the rotational speed NPM2 of the second pump motor 13 continues to decrease, and the rotational speed NPM1 of the first pump motor 12 continues to increase.

  Such a situation is the same in other shift states such as between the second speed and the third speed and between the third speed and the fourth speed. Therefore, except for the case where the extrusion volumes q1 and q2 of the hydraulic pump motors 12 and 13 are equal, the rotational speeds NPM1 and NPM2 are different and the heat generation amounts are different. That is, the amount of heat generated by the hydraulic pump motors 12 and 13 having a relatively high rotational speed is increased. Since it is preferable to set the cooling or lubricating amount according to the heat generation amount, the necessary lubricating oil amount (cooling oil amount) VPM1 and VPM2 to be supplied to the hydraulic pump motors 12 and 13 are the above-described rotational speeds NPM1 and NPM2. It becomes the same relationship.

  On the other hand, in FIG. 5, at the first speed, the extrusion volume q2 of the second pump motor 13 is set to zero, while the first pump motor 12 has the extrusion volume q1 set to the maximum. The second pump motor 13 is idling at a predetermined rotational speed NPM2, while the first pump motor 12 is stopped. That is, the rotational speed NPM1 is zero.

  In this state, as described above, the sleeve 23S of the second synchro 23 is moved to the left side in FIG. 1, so that the second intermediate shaft 10 and the output shaft 16 are interposed between the first speed gear pair 20. The power is transmitted. That is, a load is applied to the second planetary gear mechanism 4. The magnitude of the load (planetary load) LPG2 acting on the second planetary gear mechanism 4 is set at the first speed which is a fixed speed ratio at which the power transmission performed via the first speed gear pair 20 is maximized. Maximum.

  On the other hand, the first planetary gear mechanism 3 is idling at a predetermined rotational speed because the third sync 24 is OFF, and no load is applied to the first planetary gear mechanism 3. That is, the magnitude of the load (planetary load) LPG1 acting on the first planetary gear mechanism 3 is minimized at the first speed which is the fixed gear ratio.

  When upshifting from this state toward the second speed, which is the fixed gear ratio, as described above, the pushing volume q2 of the second pump motor 13 is gradually increased to function as a pump. In this case, the extrusion volume q1 of the first pump motor 12 is maintained at the maximum until the extrusion volume q2 of the second pump motor 13 is maximized. The rotation speed of the second pump motor 13 gradually decreases as the extrusion volume q2 increases. On the other hand, the first pump motor 12 is supplied with pressure oil from the second pump motor 13 and the motor. , The rotational speed gradually increases.

  In this case, as described above, the sleeve 24S of the third sync 24 is moved to the left side in FIG. 1, so that the second intermediate shaft 8 and the output shaft 16 are interposed between the second speed gear pair 18. Power begins to be transmitted. That is, the load LPG1 starts to act on the first planetary gear mechanism 3. The load LPG1 acting on the first planetary gear mechanism 3 gradually increases as the extrusion volume q2 of the second pump motor 13 increases, and after the extrusion volume q2 of the second pump motor 13 becomes maximum. Continuously increases as the extrusion volume q1 of the first pump motor 12 decreases. The load LPG1 acting on the first planetary gear mechanism 3 is maximized when the extrusion volume q1 of the first pump motor 12 becomes minimum (that is, zero), that is, when the second speed that is the fixed gear ratio is set. become.

  On the other hand, the load LPG2 applied by the second planetary gear mechanism 4 gradually decreases from the maximum load state as the extrusion volume q2 of the second pump motor 13 increases, and the second pump motor 13 pushes out. After the volume q2 reaches the maximum, it gradually decreases as the extrusion volume q1 of the first pump motor 12 decreases. When the extrusion volume q1 of the first pump motor 12 becomes minimum (that is, zero), that is, when the second speed that is the fixed gear ratio is set, the load LPG2 acting on the first planetary gear mechanism 3 is minimum. become.

  Such a situation is the same in other shift states such as between the second speed and the third speed and between the third speed and the fourth speed. Therefore, except when the extrusion volumes q1 and q2 of the hydraulic pump motors 12 and 13 are equal, the loads LPG1 and LPG2 acting on the planetary gear mechanisms 3 and 4, that is, the planetary gear mechanisms 3 and 3 by the loads LPG1 and LPG2 are used. The calorific value of 4 is different. That is, like the heat generation amounts of the hydraulic pump motors 12 and 13 described above, the heat generation amounts of the planetary gear mechanisms 3 and 4 on the side connected to the hydraulic pump motors 12 and 13 having a relatively high rotational speed are increased. Yes. Since it is preferable to set the cooling or lubricating amount according to the heat generation amount, the necessary lubricating oil amounts (cooling oil amounts) VPG1 and VPG2 to be supplied to the planetary gear mechanisms 3 and 4 are the above-described loads LPG1 and LPG2, respectively. Similar relationship.

  Therefore, in the first embodiment of the hydraulic control apparatus according to the present invention, the relationship between the rotational speeds NPM1, NPM2 and the heat generation amount, that is, the required lubricating oil amounts VPM1, VPM2, and the loads LPG1, LPG2 and the heat generation amount, that is, necessary lubrication Considering the relationship between the oil amounts VPG1 and VPG2, the extrusion volumes (capacities) q1 and q2 of the hydraulic pump motors 12 and 13, the required lubricating oil amounts VPM1 and VPM2 of the hydraulic pump motors 12 and 13, and the planetary gears. The required lubricating oil amounts VPG1 and VPG2 of the mechanisms 3 and 4 are configured to change in conjunction with each other. Specifically, as described above, the first hydraulic pump is specifically configured so that the extrusion volume (capacity) of the hydraulic pump motors 12 and 13 and the flow rates of the supply oil passages 42 and 47 are changed in conjunction with each other. The flow rate of the supply oil passage 42 decreases as the extrusion volume (capacity) q1 of the motor 12 increases, and the flow rate of the supply oil passage 47 decreases as the extrusion volume (capacity) q2 of the second hydraulic pump motor 13 increases. It is configured as follows.

  Thus, according to the first embodiment of the hydraulic control apparatus of the present invention, the amount of lubricating oil (the amount of cooling oil) is changed according to the change in the operating state of the transmission, that is, the rotational state of the hydraulic pump motors 12 and 13. ) Is set. Therefore, the hydraulic pump motors 12 and 13 and the planetary gear mechanisms 3 and 4 can be cooled or lubricated without excess or deficiency. The total amount of lubricating oil (cooling oil amount) required is the sum of the lubricating oil amounts VPM1, VPM2 required by the hydraulic pump motors 12, 13, and the planetary gear mechanisms 3, 4, and the lubricating oil amount VPG1. , VPG2 is shown in FIGS. 6 and 7. In FIG. 6, the sum of the lubricating oil amount VPM1 indicated by the alternate long and short dash line and the lubricating oil amount VPM2 indicated by the alternate long and two short dashes line is indicated by a solid line. Compared to the oil amount VMAX (dotted line), the amount is significantly smaller. In FIG. 7, the sum of the lubricating oil amount VPG1 indicated by the alternate long and short dash line and the lubricating oil amount VPG2 indicated by the alternate long and two short dashes line is indicated by the solid line. The amount of lubricating oil VMAX (dotted line) in this case is significantly smaller. As a result, since the amount of oil to be discharged by the charge pump 35 can be reduced, the charge pump 35 can be miniaturized, and the power consumed can be reduced and efficient cooling or lubrication can be performed.

(Second embodiment)
FIG. 8 is a diagram schematically showing a second embodiment of the hydraulic control circuit constituting the hydraulic control device of the present invention. The second embodiment is different from the first embodiment shown in FIG. 1 in that the first direction switching valve for switching the connection state of the supply oil passages 42 and 47 and the connection oil passages 46 and 50 are connected. It is an example of the structure which further provided the 2nd direction switching valve which switches a connection state. Accordingly, the configuration other than these directional control valves and the parts related thereto is the same as that of the first embodiment shown in FIG. 1, and the same parts as those of the first embodiment shown in FIG. The same reference numerals are attached and detailed description thereof is omitted.

  In FIG. 8, between the discharge port of the charge pressure control valve 39 and the oil passage in the casing 12C of the first pump motor 12, and in the discharge port of the charge pressure control valve 39 and the casing 13C of the second pump motor 13. Between the oil passage, there is provided a direction switching valve 63 corresponding to the first direction switching valve of the present invention. That is, one end is connected to the discharge port of the charge pressure control valve 39, and the other end of the supply oil passage 42 is connected to the discharge port of the charge pressure control valve 39. The other end of the supply oil passage 47 is connected to one connection port of the direction switching valve 63. The other connection port of the direction switching valve 63 is connected to the supply oil passage 64 connected to the oil passage in the casing 12C of the first pump motor 12 and the oil passage in the casing 13C of the second pump motor 13. A supply oil passage 65 is connected.

  The direction switching valve 63 includes a first supply state in which the supply oil passage 42 and the supply oil passage 64 are communicated, and the supply oil passage 47 and the supply oil passage 65 are in communication, and the supply oil passage 42 and the supply oil passage. 65, and the second supply state in which the supply oil passage 47 and the supply oil passage 64 are communicated can be selectively switched and set. That is, the direction switching valve 63 connects the discharge port of the charge pressure control valve 39 and the oil passage in the casing 12C of the first pump motor 12 via the supply oil passage 42, and via the supply oil passage 47. The first supply state in which the discharge port of the charge pressure control valve 39 and the oil passage in the casing 13C of the second pump motor 13 are connected, and the discharge port of the charge pressure control valve 39 and the first pump through the supply oil passage 47. A second supply state in which the oil passage in the casing 12C of the motor 12 is connected and the discharge port of the charge pressure control valve 39 and the oil passage in the casing 13C of the second pump motor 13 are connected via the supply oil passage 42. And can be selectively switched and set.

  The flow rate control valves 43 and 48 for adjusting the flow rates of the supply oil passages 42 and 47, respectively, are arranged in the middle of the supply oil passages 42 and 47, and are similar to the configuration according to the first embodiment described above. Further, the electromagnetic control valves 58 and 60 of the actuators 56 and 57 and the actuators 59 and 61 provided in the hydraulic pump motors 12 and 13 can be controlled in conjunction with the capacities of the hydraulic pump motors 12 and 13, respectively. It is configured as follows.

  That is, similar to the configuration according to the first embodiment described above, the first pump motor 12 and the flow rate control valve 43 are configured so that the capacity increase / decrease direction during the variable displacement control and the flow rate increase / decrease direction during the flow rate control. Is supposed to be the opposite. Specifically, the flow rate of the supply oil passage 42 decreases as the capacity of the first pump motor 12 increases, and conversely, the flow rate of the supply oil passage 42 increases as the capacity of the first pump motor 12 decreases. Further, the second pump motor 13 and the flow rate control valve 48 are configured such that the direction of increase / decrease in capacity during variable displacement control is opposite to the direction of increase / decrease in flow rate during flow control. ing. Specifically, the flow rate of the supply oil passage 47 decreases as the capacity of the second pump motor 13 increases, and conversely, the flow rate of the supply oil passage 47 increases as the capacity of the second pump motor 12 decreases. It is configured.

  Further, a directional switching valve 66 corresponding to the second directional switching valve of the present invention is provided between the oil passages in the casings 12C and 13C of the hydraulic pump motors 12 and 13 and the lubrication required portions D1 and D2. ing. That is, one end is connected to the other end of the connection oil passage 67 connected to the oil passage in the casing 12C of the first pump motor 12, and one end is connected to the discharge port of the charge pressure control valve 39. The other end of the connecting oil passage 68 is connected to one connection port of the direction switching valve 66. The other connection port of the direction switching valve 66 has a connection oil passage 69 connected to a lubrication required portion D2 arranged on the second pump motor 13 side and a lubrication arranged on the first pump motor 12 side. A connection oil passage 70 connected to the necessary portion D1 is connected.

  The direction switching valve 66 includes a first connection state in which the connection oil passage 67 and the connection oil passage 69 communicate with each other, and the connection oil passage 68 and the connection oil passage 70 communicate with each other, and the connection oil passage 67 and the connection oil passage. 70 and the second connection state in which the connection oil passage 68 and the connection oil passage 69 are communicated can be selectively switched and set. That is, the direction switching valve 66 connects and connects the oil passage in the casing 12C of the first pump motor 12 and the lubrication required portion D2 arranged on the second pump motor 13 side via the connection oil passage 67. A first connection state in which the oil passage in the casing 13C of the second pump motor 13 and the lubrication required portion D1 arranged on the first pump motor 12 side are connected via the oil passage 68, and the connection oil passage 67 is provided. Then, the casing of the second pump motor 13 is connected to the oil passage in the casing 12C of the first pump motor 12 and the lubrication required portion D1 disposed on the first pump motor 12 side via the supply oil passage 68. The second connection state in which the oil passage in 13C and the lubrication required portion D2 arranged on the second pump motor 13 side are connected can be selectively switched and set. .

  The switching control of the direction switching valves 63 and 66 is appropriately performed based on the temperature of the hydraulic oil supplied to the hydraulic pump motors 12 and 13 and the lubrication required portions D1 and D2. ing. An example of the switching control is shown in the flowchart of FIG. The routine shown in this flowchart is repeatedly executed every predetermined short time. In FIG. 9, first, the temperature (oil temperature) t of the hydraulic oil supplied to each hydraulic pump motor 12, 13 and each lubrication required portion D1, D2, and the rotational speed NPM1, NPM2 of each hydraulic pump motor 12, 13 are determined. Are detected (step S1). The oil temperature t of the hydraulic oil in this case can be detected by the oil temperature sensor 33, for example. Further, the oil temperature t can be estimated based on a detected value of the temperature of the transmission by a temperature sensor (not shown) installed at a predetermined position of the transmission. Further, the rotation speeds NPM1, NPM2 can be detected by, for example, the rotation speed sensors 31, 32.

  Subsequently, it is determined whether or not the obtained oil temperature t of the hydraulic oil is equal to or lower than a predetermined temperature α (step S2). The predetermined temperature α is a predetermined value set in advance as a threshold value for determining the warm-up state of the transmission or the viscosity state of the hydraulic oil.

  If the oil temperature t is equal to or lower than the predetermined temperature α, if the determination in step S2 is affirmative, the process proceeds to step S3, and the switching position of each direction switching valve 63, 66 is set to the drag reduction mode. . In this drag reduction mode, when the hydraulic oil temperature t is low and the viscosity of the hydraulic oil is high, the hydraulic oil is supplied to one of the hydraulic pump motors 12 and 13 that are rotating at a relatively high speed. Each directional control valve 63 is circulated so that the hydraulic oil having a relatively high oil temperature is circulated to one of the lubrication-requiring parts D1, D2 on the side that is idling at high speed without transmitting power. , 66, the respective supply states and connection states are set.

  Specifically, in this case, for example, if the rotation speed NPM1 of the first pump motor 12 is higher than the rotation speed NPM2 of the second pump motor 13, that is, the first pump motor 12 has a relatively high rotation speed. If so, the direction switching valve 63 is set to the second supply state, and the direction switching valve 66 is set to the second connection state. That is, the hydraulic oil that is supplied to the first pump motor 12 that is relatively high in rotation and has a relatively high oil temperature t is idling at high speed without performing power transmission. A drag reduction mode is set for distribution to the lubrication required portion D1 disposed on the pump motor 12 side.

  In this case, that is, the state of the drag reduction mode set when the oil temperature t of the hydraulic oil is low and the rotation speed NPM1 of the first pump motor 12 is higher than the rotation speed NPM2 of the second pump motor 13 is shown. 10. The state shown in FIG. 10 is a state in which the transmission is set to the first speed, which is a fixed gear ratio, and the first pump motor 12 is turned off and the second pump motor 13 is placed in the LOCK state. Therefore, the first pump motor 12 rotates at a high speed, and the first planetary gear mechanism 3 disposed on the first pump motor 12 side does not transmit power (that is, the first planetary gear mechanism 3 Rotating at high speed (with no load), the rotation of the second pump motor 13 is stopped, and the second planetary gear mechanism 4 disposed on the second pump motor 13 side is transmitting power It is.

  In this state, as shown in FIG. 10, the first planetary gear mechanism 3 rotating at high speed with no load is disposed on the first planetary gear mechanism 3 side (that is, the first pump motor 12 side). The hydraulic oil which is supplied to the first pump motor 12 rotating at a high rotational speed and is relatively high in temperature and having a low viscosity is supplied to the lubrication required portion D1. Therefore, compared with the case where the hydraulic oil having a low oil temperature t and a high viscosity is supplied to the lubrication-required part D1, the hydraulic oil having a relatively high oil temperature t and a low viscosity is supplied. The drag loss due to the viscous resistance of the hydraulic oil at D1 is reduced.

  On the other hand, if the oil temperature t is higher than the predetermined temperature α and the determination is negative in the above-described step S2, the process proceeds to step S4, where the switching positions of the direction switching valves 63, 66 are set in the cooling acceleration mode. Set to This cooling acceleration mode is relative when the oil temperature t of the hydraulic oil is high and it is necessary to suppress an increase in the temperature of the hydraulic oil and any of the lubrication required portions D1 and D2 to which the hydraulic oil is supplied. The hydraulic oil supplied to one of the hydraulic pump motors 12 and 13 that is low in rotation and having a relatively low oil temperature has a relatively high power transmission rate, and therefore has a relatively high temperature. Each supply state and each connection state by each direction switching valve 63, 66 are set so as to be distributed to any of the lubrication required parts D1, D2.

  Specifically, in this case, for example, if the rotation speed NPM1 of the first pump motor 12 is higher than the rotation speed NPM2 of the second pump motor 13, that is, the first pump motor 12 has a relatively high rotation speed. If so, the direction switching valve 63 is set to the first supply state, and the direction switching valve 66 is set to the first connection state. In other words, the hydraulic oil that is supplied to the second pump motor 13 that is relatively low in rotation and has a relatively low oil temperature t is relatively hot due to power transmission. A cooling promotion mode is set for distribution to the lubrication required portion D2 disposed on the pump motor 13 side.

  In this case, that is, the state of the cooling acceleration mode set when the oil temperature t of the hydraulic oil is low and the rotation speed NPM1 of the first pump motor 12 is higher than the rotation speed NPM2 of the second pump motor 13. 11. The state shown in FIG. 11 is a state in which the transmission is set to the first speed that is the fixed gear ratio, as in FIG. 10 described above, and the first pump motor 12 is turned off, Since the pump motor 13 is in the LOCK state, the first pump motor 12 rotates at a high speed, and the first planetary gear mechanism 3 disposed on the first pump motor 12 side does not transmit power and is high. The second planetary gear mechanism 4 disposed on the second pump motor 13 side transmits power (rotates to the second planetary gear mechanism 4). The load is applied.

  In this state, as shown in FIG. 11, the second planetary gear mechanism 4 performing power transmission and the lubrication required portion D2 disposed on the side of the first planetary gear mechanism 4 (that is, the second pump motor 13 side) are provided. The hydraulic oil which is supplied to the second pump motor 13 whose rotation is locked and which has a relatively low temperature is supplied. Therefore, a load is applied by performing power transmission, and lubrication is required by supplying hydraulic oil having a relatively low oil temperature t to the lubrication-required portion D2 that is relatively hot due to heat generation at that time. Cooling at the site D2 is promoted.

  When the switching position of each direction switching valve 63, 66 is set to either the drag reduction mode or the cooling promotion mode in step S3 or step S4, this routine is temporarily terminated.

  Here, the relationship between each of the above specific examples and the present invention will be briefly described. Each of the hydraulic pump motors 12 and 13 described above corresponds to the hydraulic pump or the hydraulic motor of the present invention, and in the second embodiment. The supply oil passages 42 and 47, the supply oil passages 64 and 65, the connection oil passages 46 and 50, the connection oil passages 69 and 70, the flow control valves 43 and 48, and the return oil passages 44 and 49 shown in FIG. The circuit C1 constituted by the direction switching valves 63, 66 and the like also corresponds to the lubricating circuit of the present invention. On the other hand, the functional means for executing the control of step S1 shown in FIG. 9 corresponds to the oil temperature detecting means of the present invention, and the functional means for executing the controls of steps S2, S3, S4 is the switching valve of the present invention. It corresponds to the control means.

  Thus, according to the second embodiment of the hydraulic control apparatus of the present invention, by appropriately controlling the direction switching valves 63 and 66 according to the oil temperature t of the hydraulic oil, for example, the oil of the hydraulic oil When the temperature t is low and the viscosity is high, the hydraulic oil that is supplied to any of the hydraulic pump motors 12 and 13 that are relatively high in rotation and has a relatively high oil temperature t Further, it is possible to set a drag reduction mode for flowing to any of the lubrication-required portions D1 and D2 on the side that is idling at high speed without performing power transmission. Therefore, drag loss due to the viscous resistance of the hydraulic oil can be reduced.

  Also, when the oil temperature t of the hydraulic oil is high and it is necessary to suppress an increase in the oil temperature t, the hydraulic oil is supplied to any of the hydraulic pump motors 12 and 13 that are relatively low in rotation. It is possible to set a cooling promotion mode in which hydraulic oil having a low oil temperature is circulated to any of the lubrication required parts D1, D2 on the side where heat is generated by performing power transmission. . Therefore, it is possible to improve the cooling effect of the lubrication-required portions D1, D2 by suppressing an increase in the temperature of the hydraulic oil supplied to the lubrication-required portions D1, D2.

  The present invention is not limited to the specific examples described above, and the target continuously variable transmission is, in essence, performing power transmission via a fluid and continuously changing the transmitted torque. Any transmission that can be used. Therefore, instead of using a variable displacement hydraulic pump or hydraulic motor as a reaction force mechanism for each differential mechanism, the power source is directly connected to these variable displacement hydraulic pumps or hydraulic motors, and their outputs The present invention can also be applied to a hydraulic control apparatus for a so-called HST (Hydro Static Transmission) that is configured to directly transmit the power to the transmission mechanism or the output member. In the present invention, when the fixed gear ratio is set by the continuously variable transmission, the number of steps is not limited to four forward steps and one reverse step, and may be more or less.

It is a schematic diagram for demonstrating the structure of the 1st Example of the transmission and hydraulic control apparatus made into object by this invention. It is a figure for demonstrating the relationship between the variable displacement control of the variable displacement hydraulic pump motor in the 1st Example shown in FIG. 1, and the flow control by a flow control valve. It is a chart which shows collectively the operation state of the transmission made into object by this invention. The figure which shows the change of the extrusion volume in the process of setting the gear ratio of the intermediate value of 1st speed and 2nd speed which is a fixed gear ratio, the change of the rotation speed of a hydraulic pump motor, and the change of required hydraulic fluid amount It is. This shows changes in the extrusion volume, changes in the load acting on the planetary gear mechanism, and changes in the required amount of hydraulic fluid in the process of setting the speed ratio of the intermediate value between the first speed and the second speed, which is the fixed speed ratio. FIG. It is a diagram which shows the total hydraulic oil amount required for cooling of the hydraulic pump motor by this invention. It is a diagram which shows the total hydraulic oil amount required for cooling of the planetary gear mechanism by this invention. It is a schematic diagram for demonstrating the structure of the 2nd Example of the transmission and hydraulic control apparatus made into object by this invention. FIG. 9 is a flowchart for explaining an example of control of the switching state of the direction switching valve according to the present invention in the second embodiment shown in FIG. 8. FIG. 9 is a schematic diagram for explaining an example of a drag reduction mode according to the present invention in the second embodiment shown in FIG. 8. FIG. 9 is a schematic diagram for explaining an example of a cooling acceleration mode according to the present invention in the second embodiment shown in FIG. 8.

Explanation of symbols

  DESCRIPTION OF SYMBOLS 1 ... Engine (power source) 3, 4 ... Planetary gear mechanism (1st, 2nd differential mechanism) 12, 13 ... Hydraulic pump motor (variable capacity type hydraulic pump or hydraulic motor), 16 ... Output shaft (output) Members), 17, 18, 19, 20 ... gear pairs (transmission mechanism), 22, 23, 24, 25 ... synchro (switching mechanism), 33 ... oil temperature sensor, 35 ... charge pump (replenishment pump), 42, 47 ... Supply oil passage (first and second supply oil passages), 43, 48 ... Flow control valve (first and second flow control valves), 46, 50 ... Connection oil passage (first and second connection oil passages) 51 ... Relief valve 56, 57 ... Actuator 58, 60 ... Electromagnetic control valve (first and second electromagnetic control valve) 62 ... Electronic control unit (ECU) 63, 66 ... Directional switching valve (first , Second direction switching valve), C0 ... closed circuit, C1 ... circuit (lubricating circuit) , D1, D2 ... lubrication necessary site.

Claims (5)

A variable displacement hydraulic pump driven by power output from a power source, and a variable displacement hydraulic motor that outputs power to an output member by being supplied and driven by pressure oil output from the hydraulic pump. A torque control device for a continuously variable transmission in which torque transmitted to the output member changes according to the capacity and hydraulic pressure of these hydraulic pumps and hydraulic motors,
A closed circuit for circulating pressure oil between the hydraulic pump and the hydraulic motor;
A lubricating circuit for supplying cooling or lubricating hydraulic fluid to the hydraulic pump and the hydraulic motor and the lubrication-necessary part of the continuously variable transmission;
A replenishment pump that replenishes the closed circuit with pressure oil and supplies the hydraulic oil to the lubrication circuit;
Wherein the capacity of the hydraulic pump and the hydraulic motor, wherein the supply pump that provides an actuator and changing in conjunction with the flow rate of the hydraulic pump and the hydraulic oil supplied to the hydraulic motor CVT hydraulic pressure control instrumentation of Where
The hydraulic pump and the hydraulic motor include a variable displacement type first hydraulic pump motor and a second hydraulic pump motor having functions of both the pump and the motor,
The continuously variable transmission includes a reaction force element in which an input element that receives power from the power source, an output element that outputs power to the output member via a transmission mechanism, and the first hydraulic pump motor are connected. A first differential mechanism that performs a differential action, and another input element that receives power from the power source and another output element that outputs power to the output member via another transmission mechanism. A second differential mechanism that performs a differential action with another reaction force element to which the second hydraulic pump motor is coupled, and a switching mechanism that selectively allows each of the transmission mechanisms to transmit torque.
The lubrication required portion includes the first and second differential mechanisms, and the transmission mechanisms and the switching mechanisms that are always or selectively connected to the first and second differential mechanisms,
The lubrication circuit is provided with a first flow rate control valve whose flow rate changes in conjunction with a change in the capacity of the first hydraulic pump motor, and from the supply pump to the first hydraulic pump motor or the second hydraulic pump motor. A first supply oil path for supplying the hydraulic oil and a second flow rate control valve for changing a flow rate in conjunction with a change in the capacity of the second hydraulic pump motor are provided, and the second hydraulic pump motor is supplied from the replenishment pump. Alternatively, a second supply oil passage that supplies the hydraulic oil to the first hydraulic pump motor, a first connection oil passage that distributes the hydraulic oil supplied to the first hydraulic pump motor to the lubrication required portion, and A second connection oil passage for flowing the hydraulic oil supplied to the second hydraulic pump motor to the lubrication-required part;
The actuating device includes a first electromagnetic control valve that can be electrically controlled by linking a capacity of the first hydraulic pump motor and a flow rate of the first flow control valve, a capacity of the second hydraulic pump motor, and the A second electromagnetic control valve that can be electrically controlled in conjunction with the flow rate of the second flow control valve
A hydraulic control device for a continuously variable transmission. Before Symbol first electromagnetic control valve, controls the first hydraulic pump as the capacity of the motor increases from the first hydraulic pump motor so that the flow rate of the first oil supply passage is decreased first flow control valve Configured and possible
The second electromagnetic control valve can control the second hydraulic pump motor and the second flow control valve such that the flow rate of the second supply oil passage decreases as the capacity of the second hydraulic pump motor increases. Is configured to
Hydraulic control system for a continuously variable transmission according to claim 1, wherein the this. The lubrication circuit connects the replenishment pump and the hydraulic pump motor having a relatively low rotation speed through the first supply oil passage, and relatively high rotation speed from the replenishment pump through the second supply oil passage. A first supply state in which the hydraulic pump motor is connected, the replenishment pump and the hydraulic pump motor having a relatively high rotational speed are connected by the first supply oil passage, and the second supply oil passage is used. A first direction switching valve that selectively switches to a second supply state in which the replenishment pump and the hydraulic pump motor having a relatively low rotational speed are connected, and the relatively low rotational speed by the first connecting oil passage. A hydraulic pump motor having a relatively high rotation speed and a portion requiring lubrication disposed on the side of the hydraulic pump motor having a relatively low rotation speed, and the relatively high rotation speed hydraulic pump motor by the second connection oil passage. Its relatively high A first connection state in which the portion requiring lubrication arranged on the hydraulic pump motor side of the rotation is connected, and the relatively low rotation number of the hydraulic pump motor and the relatively high rotation by the first connection oil passage. The hydraulic pump motor having the relatively high rotational speed and the hydraulic pump having the relatively low rotational speed are connected to the portion requiring lubrication disposed on the side of the hydraulic pump motor and the relatively high rotational speed hydraulic pump motor by the second connection oil passage. A second directional control valve that selectively switches to a second connection state in which the lubrication-required portion disposed on the motor side is connected ;
Oil temperature detecting means for detecting a temperature of said work aggressive media,
Switching valve control means for controlling the switching state of the first and second directional switching valves based on the oil temperature detected by the oil temperature detection means;
It is further equipped with
Hydraulic control system for a continuously variable transmission according to claim 1 or 2, characterized and this. Before SL switching valve control unit, when a predetermined temperature below which is detected the oil temperature predetermined by the oil temperature detecting means, the first said supply oil passage and the second supply oil passage first supply state And a drag reduction mode in which the first connection oil path and the second connection oil path are set to the first connection state, and the first supply oil path when the oil temperature is higher than the predetermined temperature. and the second supply oil passage to the second supply state, and characterized that you including means for setting a cooling demonstration mode in which the first connecting oil passage and the second connecting oil passage to the second connection state The hydraulic control device for a continuously variable transmission according to claim 3. Further comprising a relief valve for setting the hydraulic pressure discharged from the pre-Symbol supply pump to a predetermined pressure,
The lubrication circuit has no according to any one of claims 1 to 4, characterized in it to contain circuitry to be supplied to said hydraulic pump and the hydraulic motor a hydraulic fluid which is discharged as the working oil from said relief valve Hydraulic control device for step transmission.

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