JP2009036299A - Controller of vehicular hydraulic transmission - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a controller capable of restraining heat-generation, and restraining increase of a load of a power source at the time of starting without complicating contents of the control. <P>SOLUTION: This controller of a vehicular hydraulic transmission comprises a variable-capacity type hydraulic pump, a variable-capacity type hydraulic motor, and a pressure control valve capable of controlling a hydraulic pressure between the hydraulic pump and the hydraulic motor. In this controller, a torque transmitted to an output member changes according to extrusion capacities and hydraulic pressures of the hydraulic pump and the hydraulic motor. This controller comprises start detection means (step S4) for detecting an operator's start will, elapsed time detection means (steps S4, S5) for detecting an elapsed time from a point at which the start will is detected, pressure control valve flow rate calculation means (step S6) for calculating an oil volume to be passed through the pressure control valve based on the elapsed time and a required driving force, and pump-motor control means (steps S7, S8, S9) for controlling the hydraulic pump and the hydraulic motor to emit the oil volume. <P>COPYRIGHT: (C)2009,JPO&INPIT

Description

  The present invention relates to a control apparatus for a vehicular hydraulic transmission capable of continuously changing a gear ratio by transmitting power using 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 applied can be changed appropriately and continuously. That is, it is possible to configure a continuously variable transmission that transmits hydraulic power using hydraulic pressure and that can continuously change the gear ratio. One example thereof is described in Patent Document 1. The vehicle hydraulic continuously variable transmission described in Patent Document 1 is connected to a variable displacement hydraulic pump driven by an engine, and the hydraulic pump via a hydraulic closed circuit and drives a drive wheel. A fixed displacement type or variable displacement type hydraulic motor, and a relief valve means for regulating the hydraulic pressure of the hydraulic closed circuit, receiving the driving force of the hydraulic motor driven by the pressure oil supplied from the hydraulic pump, The vehicle is configured to be driven to travel.

  The hydraulic continuously variable transmission for a vehicle described in Patent Document 1 suppresses a reduction in fuel consumption, an increase in idle rotation vibration, an increase in hydraulic vibration, and the like when a brake is operated in a creep state. When a state where the brake is released is detected, the hydraulic pressure in the oil passage on the discharge side of the hydraulic pump in the hydraulic closed circuit is regulated to a hydraulic pressure higher than a predetermined hydraulic pressure by the relief valve means.

  In Patent Document 2, a short circuit between a suction port and a discharge port of a hydraulic pump is provided in a hydraulic closed circuit formed between a hydraulic pump connected to an input shaft and a hydraulic motor connected to an output shaft. A clutch device for a hydraulic transmission for a vehicle is described which is connected and provided with a clutch valve (relief valve) for opening and closing the short-circuit path. The clutch device for a hydraulic transmission for a vehicle described in Patent Document 2 is based on the throttle opening and the engine speed in order to prevent the engine from blowing up when the vehicle starts. Is configured to change.

  Patent Document 3 discloses at least a pair of variable displacement fluid pressure pump motors that can exchange pressure fluid with each other and at least two transmissions that transmit torque transmitted by each variable displacement fluid pump motor to an output member. And a switching mechanism for switching each transmission mechanism between a state where power transmission is possible and a state where power transmission is impossible, a fixed gear stage determined by a gear ratio of one of the transmission mechanisms, and each variable displacement type A transmission capable of setting a continuously variable transmission state by changing power transmitted between fluid pump motors via pressure fluid is described.

JP-A-11-166620 JP-A 61-207228 JP 2006-266493 A

  The transmission described in Patent Document 1 controls the driving force by changing the opening of a relief valve provided in a hydraulic closed circuit by a relief valve means. That is, the relief valve provided in the hydraulic closed circuit as described above sets the upper limit of the transmission torque, and the transmission torque, that is, the driving force is controlled by controlling the opening degree of the relief valve. For example, when the vehicle is braked by a brake, the relief valve is opened, and the pressure in the hydraulic closed circuit is reduced. Then, as the brake is released and the vehicle starts and the vehicle speed increases, the relief valve is controlled to close gradually.

  Therefore, if it takes a long time for the vehicle speed to rise after the brake has been released and the vehicle has started, such as when starting with a high running resistance such as uphill or uneven roads, As compared with the case, the state in which the relief valve is opened (including the half-open state) becomes longer. As a result, it takes a long time for the pressure oil in the hydraulic closed circuit to be relieved through the relief valve, which may reduce the efficiency or increase the heat generation in the relief valve.

  On the other hand, by changing the opening degree of the relief valve to limit the flow rate of the relief valve, problems such as the above-described decrease in efficiency and heat generation of the relief valve can be avoided. However, in that case, the load on the engine increases as the hydraulic pressure in the relief valve increases, and it is necessary to execute control to increase the engine output in response to the increase in the load. As a result, the control content of the engine becomes complicated, and there is a risk of fuel consumption deterioration due to an increase in engine output.

  The present invention has been made paying attention to the above technical problem, and in a hydraulic transmission for a vehicle using a hydraulic pump and a hydraulic motor connected to each other in a closed circuit, the heat generation and efficiency in the closed circuit are improved. It is an object of the present invention to provide a control device that can suppress a decrease and suppress an increase in load on a power source at the start without complicating the control contents.

  In order to achieve the above object, a first aspect of the present invention is a closed circuit capable of transmitting and receiving pressure oil to and from a variable displacement hydraulic pump driven by power output from a power source. A variable displacement hydraulic motor that outputs power to an output member by being supplied with and driven by pressure oil output from the hydraulic pump, and a hydraulic pressure between the hydraulic pump and the hydraulic motor. In a control device for a hydraulic transmission for a vehicle, the driver includes a pressure regulating valve capable of regulating pressure, and the torque transmitted to the output member varies according to the extrusion volume and hydraulic pressure of the hydraulic pump and the hydraulic motor. A start detection means for detecting the vehicle's start intention, an elapsed time detection means for detecting an elapsed time from the time when the start intention is detected by the start detection means, and an elapsed time detection means. Based on the detected elapsed time and the required driving force for the vehicle, pressure regulation valve flow rate calculation means for calculating the amount of oil to be passed through the pressure regulation valve; and the oil amount calculated by the pressure regulation valve flow rate calculation means And a pump / motor control means for controlling the hydraulic pump and the hydraulic motor so as to discharge.

  According to a second aspect of the present invention, in the first aspect of the invention, the start detection means detects that the braking by the braking device is released when the vehicle is stopped by being braked by the braking device. Thus, the control device includes means for detecting the intention to start.

  According to a third aspect of the present invention, in the first or second aspect of the present invention, the hydraulic pump and the hydraulic motor are each a variable displacement hydraulic pump motor having both a function as a hydraulic pump and a function as a hydraulic motor. It is a control device characterized by including.

  According to the first aspect of the present invention, when the driver's intention to start the vehicle is detected, the elapsed time from when the driver's intention to start is detected is measured. The amount of oil to be passed through the pressure regulating valve is determined based on the required driving force represented by degrees. Then, the hydraulic pump and the hydraulic motor are controlled so that the determined amount of oil actually passes through the pressure regulating valve. Therefore, by controlling the hydraulic pump and the hydraulic motor, the driving force of the vehicle is appropriately controlled according to the elapsed time from the intention to start without changing the set pressure of the pressure regulating valve. Therefore, heat generation at the relief valve and reduction in efficiency can be avoided or suppressed, and an increase in load on the power source at the start can be suppressed without complicating the control contents.

  According to the invention of claim 2, it is determined that the driver is willing to start the vehicle by releasing the braking by the braking device while the vehicle is being braked and stopped. Therefore, by detecting the operating state of the braking device, it is possible to easily detect the intention to start the vehicle by the driver.

  According to the invention of claim 3, the variable displacement type hydraulic pump and the hydraulic motor are respectively constituted by the variable displacement type hydraulic pump motor having both functions. Therefore, it is possible to increase the degree of freedom in the transmission direction of the hydraulic pressure in the hydraulic circuit in which they are provided, and to diversify the form in which torque is transmitted to the output member.

  Next, the present invention will be described based on specific examples. First, the vehicle hydraulic transmission that is the subject of the present invention will be described. As an example, the vehicle hydraulic transmission that is the subject of the present invention is a continuously variable transmission of a variable displacement hydraulic pump motor type, At least two power transmission paths are provided, and the torque can be transmitted from the power source to the output member via both of the power transmission paths. As a result, the rotation speed ratio between the power source and the output member This is a transmission capable of continuously changing the gear ratio. More specifically, each power transmission path includes a variable displacement hydraulic pump motor that functions as a pump and a motor, and is configured to transmit torque according to its capacity (extrusion volume). The variable displacement hydraulic pump motors communicate with each other so that hydraulic fluid can be exchanged between them. Therefore, by causing one variable displacement hydraulic pump motor to function as a pump, torque corresponding to the extrusion volume is transmitted from the power source to the output member, and at the same time, from one variable displacement hydraulic pump motor to the other variable displacement. Pressure oil is supplied to the mold hydraulic pump motor, and the other variable displacement hydraulic pump motor functions as a motor. That is, power transmission via pressure oil is performed in parallel. The torque is transmitted to the output member via the other power transmission path. As a result, the torque transmitted to the output member is the sum of the torque transmitted via each power transmission path, and the torque transmitted via the pressure oil changes according to each extrusion volume. Eventually, the gear ratio changes continuously.

  Each power transmission path can be provided with a transmission mechanism such as a gear pair or a wrapping transmission mechanism with different gear ratios. When torque is transmitted to the output member via only one power transmission path, the transmission speed is changed. The speed ratio of the entire machine is determined by the speed ratio of the transmission mechanism in the power transmission path. If such a gear ratio is referred to as a fixed gear ratio, transmission of power via pressure oil does not occur in a state where the fixed gear ratio is set, so that power loss is unlikely to occur and an efficient transmission state. It becomes. Note that a switching mechanism such as a clutch mechanism is preferably included in each transmission mechanism so that only one of the transmission mechanisms is involved in torque transmission, or between the power source or output member and the transmission mechanism. It is preferable to provide a switching mechanism.

  Since the variable displacement hydraulic pump motor type continuously variable transmission targeted by the present invention is configured to transmit power via pressure oil, it is a transmission configured as a hydrostatic transmission (HST). However, as described above, it is preferably configured as a hydrostatic mechanical transmission (HMT) having a function of setting a gear ratio by mechanical power transmission. The mechanical transmission portion can be appropriately configured as necessary, and a mechanism configured to select a gear pair that is always meshed by a clutch mechanism or a synchronous coupling mechanism, or a plurality of planetary gear mechanisms or compound planetary gears. A configuration in which a plurality of gear ratios can be set by a mechanism can be employed. Further, the variable displacement hydraulic pump motor may be configured to use a variable displacement hydraulic pump motor as a reaction force means in addition to the configuration in which the variable displacement hydraulic pump motor is interposed in series between the power source and the output member.

  A configuration of a variable displacement hydraulic pump motor type continuously variable transmission targeted in the present invention will be described with reference to FIG. The configuration example shown in FIG. 1 is an example configured as a transmission TM for a vehicle VE, and uses a differential mechanism as a power distribution mechanism and a plurality of gear pairs as a transmission mechanism, and thus a variable displacement hydraulic pressure. In this example, the pump motor is a reaction force mechanism, and three forward speeds and one reverse speed are set as so-called fixed speed ratios that can be set without changing the form of power (energy) to be transmitted. This is a configured example. That is, in FIG. 1, a mechanism for distributing, transmitting, and interrupting power is disposed on the same axis as the input member 2 connected to the power source 1 or on an axis parallel to the same.

  Here, the power source 1 may be a general power source used in the vehicle VE, such as an internal combustion engine, an electric motor, or a combination thereof. In the following description, the power source 1 is referred to as the engine 1 temporarily. 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 the following description, the input member 2 is referred to as an input shaft 2. Appropriate transmission means such as a damper, a clutch, and a torque converter can be interposed between the engine 1 and the input shaft 2. Reference numeral 3 denotes an oil pump called a sub-pump or a charge pump, which supplies lubricating oil to each part in the transmission TM and to oil passages formed between hydraulic pump motors described later. It is used for replenishment of pressure oil.

  The mechanism arranged on each axis outputs the input power as it is, or outputs a part of it as it is, and outputs other power by converting the energy form, and further idles. It is a kind of transmission means configured not to transmit power. The configuration example shown in FIG. 1 includes a differential mechanism and a reaction force mechanism that applies a reaction force to the differential mechanism and can change the reaction force. The differential mechanism may be any mechanism that performs a differential action by three rotating elements, and is a mechanism that uses gears and rollers as rotating elements. Among them, a single pinion type planetary gear is used as the geared differential mechanism. A mechanism or a double pinion type planetary gear mechanism can be used. The reaction force mechanism may be a mechanism that can selectively output torque, and a hydraulic pump motor such as a hydraulic pressure or a motor / generator that operates electrically may be used.

  In the configuration example shown in FIG. 1, a single pinion type planetary gear mechanism is used as a differential mechanism, and a variable displacement hydraulic pump motor is used as a reaction force mechanism (corresponding to the hydraulic motor of the present invention) for generating a reaction force. Is used. In the following description, a planetary gear mechanism disposed on the same axis as the first drive shaft 4 parallel to the engine 1 and the input shaft 2 is temporarily referred to as a first planetary gear mechanism 5, and a hydraulic pump motor is temporarily referred to as a first pump. This will be referred to as motor 6. Further, similarly to the first planetary gear mechanism 5, a planetary gear mechanism disposed on the same axis as the second drive shaft 7 parallel to the engine 1 and the input shaft 2 is temporarily referred to as a second planetary gear mechanism 8, and is hydraulically The pump motor is referred to as a second pump motor 9. The first pump motor 6 is indicated as PM1 in the figure, and the second pump motor 9 is indicated as PM2 in the figure.

  The first planetary gear mechanism 5 rotates and rotates a sun gear S1 that is an external gear, a ring gear R1 that is an internal gear disposed concentrically with the sun gear S1, and a pinion gear that meshes with the sun gear S1 and the ring gear R1. This is a single-pinion type planetary gear mechanism that rotates a carrier C1 that is revolving freely. A counter drive gear 10A of the first counter gear pair 10 is attached to the input shaft 2, and one counter driven gear 10B meshing with the counter drive gear 10A is connected to the ring gear R1. That is, the input shaft 2 is connected to the ring gear R1 via the first counter gear pair 10. Therefore, the ring gear R1 is an input element.

  Further, the rotor shaft 6A of the first pump motor 6 as a reaction force mechanism is connected to the sun gear S1. Therefore, the sun gear S1 is a reaction force element. The first drive shaft 4 is connected to the carrier C1. And this 1st drive shaft 4 is made into the state which can selectively transmit torque between the driven shaft 11 used as the output shaft of this transmission TM by the several transmission mechanism and switching mechanism which are mentioned later. It has a configuration. That is, the carrier C1 is connected to the driven shaft 11 via the first drive shaft 4, each transmission mechanism, and the switching mechanism. Therefore, the carrier C1 is an output element. The first drive shaft 4 is disposed on the opposite side in the axial direction from the first pump motor 6 with the first planetary gear mechanism 5 interposed therebetween.

  The first pump motor 6 is a variable displacement type that can change the extrusion volume. In the configuration example shown in FIG. 1, the first pump motor 6 is a so-called single swing type that can change the extrusion volume from zero to either positive or negative. The first planetary gear mechanism 5 is disposed on the same axis as the first planetary gear mechanism 5 on the engine 1 side (left side in FIG. 1). As this type of first pump motor 6, various types of pumps can be employed. For example, a swash plate pump, a swash shaft pump, or a radial piston pump can be used.

  On the other hand, the second planetary gear mechanism 8 has the same configuration as that of the first planetary gear mechanism 5 described above, and the carrier C2 holds the sun gear S2, the ring gear R2, and the pinion gear meshing with them in a freely rotating and revolving manner. Is a single-pinion type planetary gear mechanism that performs differential action by these three rotating elements.

  Similarly to the first planetary gear mechanism 5 described above, the other counter driven gear 10C meshed with the counter drive gear 10A attached to the input shaft 2 is connected to the ring gear R2 via the start (S) synchro 12. Yes. This start synchronizer 12 constitutes a so-called start switching mechanism, which makes it possible to selectively transmit torque between the ring gear R2 of the second planetary gear mechanism 8 and the engine 1, and to rotate the ring gear R2. It is configured to be able to regulate, that is, to fix the ring gear R2. Therefore, the ring gear R2 is an input element.

  Further, the rotor shaft 9A of the second pump motor 9 as a reaction force mechanism is connected to the sun gear S2. Therefore, the sun gear S2 is a reaction force element. The second drive shaft 7 is connected to the carrier C2. A counter drive gear 13A of the second counter gear pair 13 is attached to the second drive shaft 7, and a counter driven gear 13B meshing with the counter drive gear 13A is connected to the driven shaft 11. That is, the carrier C <b> 2 is connected to the driven shaft 11 via the second drive shaft 7 and the second counter gear pair 13. Therefore, the carrier C2 is an output element.

  The first counter gear pair 10 and the second counter gear pair 13 constitute a so-called input transmission mechanism and output transmission mechanism, respectively. It is also possible to replace it with a winding transmission mechanism using a belt or the like.

  The second pump motor 9 is a variable displacement type that can change the extrusion volume. In the configuration example shown in FIG. 1, the second pump motor 9 is a so-called double swing type that can change the extrusion volume from zero to both positive and negative directions. The second planetary gear mechanism 8 is disposed on the same axis and adjacent to the outside in the radial direction of the first pump motor 6 described above. As this type of second pump motor 9, various types of pumps can be employed as in the case of the first pump motor 6. For example, a swash plate pump, a swash shaft pump, or a radial piston pump can be used. it can.

  Here, the start synchronizer 12 as the start switching mechanism will be described. The start synchronizer 12 is composed of, for example, a synchronous coupling mechanism (synchronizer), a meshing clutch (dog clutch), or a friction clutch. Describes a start sync 12 comprising a synchronous coupling mechanism. The start sync 12 includes a sleeve 12S that is spline-fitted to a hub integrated with the ring gear R2 of the second planetary gear mechanism 8, and the counter of the first counter gear pair 10 is disposed on both sides of the sleeve 12S. A spline integrated with the driven gear 10C and a fixed member 14 fixed to, for example, a casing (not shown) of the transmission TM is disposed.

  Specifically, a spline integrated with the counter driven gear 10C of the first counter gear pair 10 is arranged on the left side of the sleeve 12S in FIG. 1, and integrated with the fixing member 14 on the right side of the sleeve 12S in FIG. Splines are placed. Accordingly, the start sync 12 moves the sleeve 12S to the left side in FIG. 1 to connect the counter driven gear 10C of the first counter gear pair 10 to the ring gear R2 of the second planetary gear mechanism 8, and the sleeve 12S in FIG. By moving the ring gear R2 of the second planetary gear mechanism 8 to the fixing member 14 to restrict the rotation of the ring gear R2, that is, fixing the ring gear R2 and positioning the sleeve 12S in the center. The counter driven gear 10C or the fixed member 14 is not engaged with the neutral state.

  The driven shaft 11 to which power is transmitted from the drive shafts 4 and 7 is arranged in parallel with the drive shafts 4 and 7 and on the same axis as the input shaft 2. Therefore, the transmission TM shown in FIG. 1 has a so-called two-shaft structure in which the main part is composed of two shafts, in particular, the first drive shaft 4 and the driven shaft 11. A plurality of transmission mechanisms for setting different gear ratios are provided between the first drive shaft 4 and the second drive shaft 7 and the driven shaft 11. Each of these transmission mechanisms is for setting a gear ratio between the input shaft 2 and the driven shaft 11 in accordance with the respective rotation speed ratio when involved in torque transmission. A hanging transmission mechanism, a mechanism using a friction wheel, or the like can be employed. In the configuration example shown in FIG. 1, three gear pairs 15, 13, 16 for forward traveling and a gear pair 17 for backward traveling are provided.

  The driven gears 15B, 16B, and 17B in the gear pairs 15, 16, and 17 attached to the first drive shaft 4 are rotatably fitted to and supported by the driven shaft 11. That is, the reverse driven gear 17B is rotatably fitted to the driven shaft 11 while meshing with the idle gear 17C disposed between the reverse driven gear 17B and the reverse drive gear 17A, and the reverse driven gear 17B rotates. The direction and the rotation direction of the reverse drive gear 17A are the same. Further, the first speed driven gear 15B is rotatably fitted to the driven shaft 11 while being meshed with the first speed drive gear 15A, and is disposed adjacent to the reverse driven gear 17B. Further, the third speed driven gear 16B is rotatably fitted to the driven shaft 11 while being engaged with the third speed drive gear 16A, and is disposed adjacent to the first speed driven gear 15B.

  A switching mechanism is provided for making these gear pairs 15, 16, and 17 selectively transmit power. This switching mechanism is a mechanism that selectively couples each gear pair 15, 16, and 17 to either the first drive shaft 4 or the driven shaft 11, and therefore, a synchronous coupling mechanism (synchronized) in a conventional manual transmission or the like. Nizer) can be used, or a mesh clutch (dog clutch), a friction clutch, or the like can be used. Further, when the driven gear is integrally attached to the driven shaft 11, the drive gear is rotatable with respect to the first drive shaft 4, and the drive gear is selectively selected with respect to the first drive shaft 4. A switching mechanism can be provided on the first drive shaft 4 side so as to be coupled.

  In the configuration example shown in FIG. 1, a synchronous coupling mechanism is used as a switching mechanism, and a first sync 18 is arranged adjacent to the reverse driven gear 17B. In addition, a second sync 19 is disposed between the first speed driven gear 15B and the third speed driven gear 16B. These synchros 18 and 19 are the same as those used in conventional manual transmissions, and a sleeve is spline-fitted to a hub integrated with the driven shaft 11, and the sleeve is moved in the axial direction. Chamfers or splines that are gradually fitted to the spline are integrally provided on each driven gear, and a ring that synchronizes rotation by gradually frictionally contacting a predetermined member on the driven gear side as the sleeve moves is provided. .

  Accordingly, the first sync 18 moves the sleeve 18S to the left side of FIG. 1 to connect the reverse driven gear 17B to the driven shaft 11, and by positioning the sleeve 18S in the center, the reverse driven gear 17B is It is comprised so that it may be in the neutral state which does not engage. Further, the second synchro 19 connects the first speed driven gear 15B to the driven shaft 11 by moving the sleeve 19S to the right side in FIG. 1, and moves the sleeve 19S to the left side in FIG. By connecting the third speed driven gear 16B to the driven shaft 11 and further positioning the sleeve 19S in the center, it is configured to be in a neutral state that does not engage any of the driven gears 15B, 16B.

  Each of the syncs 18 and 19 and the sleeves 18S and 19S and the sleeve 12S of the start sync 12 can be configured to be switched by manual operation via a linkage (not shown). A switching operation can be performed by an actuator (not shown) provided individually for each. Further, the extrusion volume of each pump motor 6, 9 or the operation of each actuator is electrically controlled by an electronic control unit (ECU) 29 described later.

  Further, the second counter gear pair 13 is arranged between the second drive shaft 7 and the driven shaft 11 as described above. That is, the counter drive gear 13A of the second counter gear pair 13 is attached to the right end of the second drive shaft 7 in FIG. 1, and the counter driven gear 13B meshed with the counter drive gear 13A is connected to the driven shaft 11. It is attached to the left end in FIG.

  Next, a hydraulic circuit for controlling the pump motors 6 and 9 will be described. The pump motors 6 and 9 are communicated with each other by oil passages 20 and 21 so that the pressure oil can be transferred to each other. That is, the respective suction ports (suction ports) 6S and 9S are communicated with each other by the oil passage 20, and the discharge ports (discharge ports) 6D and 9D are communicated with each other through the oil passage 21. Accordingly, a closed circuit is formed by the oil passages 20 and 21.

  Each oil passage 20, 21 forming this closed circuit is provided with a charge pump (sometimes called a boost pump) 22 for replenishing oil. The charge pump 22 is used to compensate for the shortage of oil due to leakage from the closed circuit, and is driven by the power source 1 or a motor (not shown) to pump oil from the oil pan 23. It is designed to supply a closed circuit.

  The discharge port of the charge pump 22 communicates with the oil passage 20 and the oil passage 21 in the closed circuit via check valves 24 and 25, respectively. The check valves 24 and 25 are configured to open in the discharge direction from the charge pump 22 and close in the opposite direction. Further, a relief valve 26 for adjusting the discharge pressure of the charge pump 22 is communicated with the discharge port of the charge pump 22. The relief valve 26 is configured to open and discharge the oil to the oil pan 23 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 discharge pressure is set to a pressure corresponding to the pilot pressure.

  Further, a relief valve 27 is provided between the suction port 6 </ b> S of the first pump motor 6 and the oil passage 21. That is, a relief valve 27 is provided in parallel with the first pump motor 6 so as to communicate the oil passages 20 and 21. When the pressure oil is discharged from the suction port 6S of the first pump motor 6 or the suction port 9S of the second pump motor 9, the relief valve 27 maintains the discharge pressure at a preset pressure, that is, a set pressure. It is configured. In other words, the relief valve 27 is configured to open and discharge when the pressure in the oil passage 21 is higher than a preset pressure (set pressure).

  A relief valve 28 is provided between the discharge port 9 </ b> D of the second pump motor 9 and the oil passage 20. That is, a relief valve 28 is provided in parallel with the second pump motor 9 so as to communicate the oil passages 20 and 21. The relief valve 28 maintains the discharge pressure at a preset pressure (set pressure) when pressure oil is discharged from the discharge port 9D of the second pump motor 9 or the discharge port 6D of the first pump motor 6. It is configured as follows. In other words, the relief valve 28 is configured to open and discharge when the pressure in the oil passage 20 is higher than a preset pressure (set pressure).

  The relief valves 27 and 28 apply a control pressure to a valve body (not shown) such as a spool pressed by a spring (not shown) in the valve opening direction in a direction opposite to the spring. Furthermore, the oil pressure of the oil passage 20 or the oil passage 21 is configured to act on the valve body in the same direction as the spring. The control pressure (control signal) is generated by an electromagnetic force generated by a solenoid or a hydraulic pressure controlled by a solenoid valve, although not particularly shown.

  More specifically, these relief valves 27 and 28 increase the oil pressure of the oil passage 20 or the oil passage 21 by increasing the control pressure to increase the so-called pressure regulation level, that is, the set pressure (relief pressure). The differential pressure (hydraulic pressure difference) between the oil passage 20 and the oil passage 21 can be increased. In other words, by increasing the set pressure of the relief valve 27 (or 28), the differential pressure between the upstream side and the downstream side of the relief valve 27 (or 28) is increased.

  In addition, when the control pressure is lowered to a predetermined lower limit pressure, the relief valves 27 and 28 directly connect the oil passage 20 and the oil passage 21 and the differential pressure between the oil passage 20 and the oil passage 21. Is configured to be almost zero. Therefore, the configuration shown in FIG. 1 is configured such that the discharge pressure of the pump motor 6 (or 9) functioning as a pump or the associated shaft torque can be electrically controlled via a solenoid valve or the like. That is, the relief valves 27 and 28 are control valves corresponding to the pressure regulating valve of the present invention, which can regulate the hydraulic pressure between the pump motor 6 (or 9) and the pump motor 9 (or 6), respectively.

  An electronic control unit (ECU) 29 for electrically controlling the extrusion volume of each pump motor 6, 9 and each synchro 12, 18, 19 is provided. The electronic control unit 29 is configured mainly with a microcomputer, and receives detection signals such as the number of rotations of a predetermined rotating member and the stroke of an operating member, and stores those input signals and prestores them. The calculation is performed based on the information and the program being operated, and the extrusion volume of each pump motor 6, 9 is set according to the calculation result, or a command signal for operating each synchro 12, 18, 19 is output. Is configured to do.

  Next, the operation of the transmission TM described above will be described. FIG. 2 shows the pump motors (PM1, PM2) 6, 9 and the synchros 12, 18, 19 at the time of setting the respective gears determined by the gear ratio of any one of the gear pairs 15, 13, 16, 17. FIG. 2 is a chart (operation table) collectively showing the operation state, and “0” for each of the pump motors 6 and 9 in FIG. 2 makes the capacity (extrusion volume) substantially zero and the rotor shaft rotates. No pressure oil is generated even if it is applied, and the output shaft does not rotate (free) even if hydraulic pressure is supplied, and “LOCK” indicates that the rotation of the rotor is stopped (locked) Is shown. Further, “PUMP” indicates a state in which the pump capacity is made larger than substantially zero and the pressure oil is discharged, and thus the corresponding first or second pump motor 6 or 9 functions as a pump. “MOTOR” indicates a state in which the pressure oil discharged from one of the pump motors 6 (or 9) is supplied and functions as a motor. Therefore, the corresponding hydraulic pump motor 9 (or 6) has a shaft torque. It has occurred.

  The “right” and “left” for each of the synchros 12, 18, and 19 indicate the positions of the sleeves 12S, 18S, and 19S in FIG. 1, and the parentheses indicate a standby state for downshifting, the key. The parentheses indicate a standby state for upshifting, and “N” indicates a state where the corresponding synchros 12, 18, and 19 are set to the OFF state (neutral position), and italic “N” reduces dragging. Therefore, it indicates that the OFF state (neutral position) is set.

  When the neutral position is selected and the neutral state is set, the extrusion volumes of the pump motors 6 and 9 are set to zero, and the synchros 12, 18, and 19 are turned off. That is, each sleeve 12S, 18S, 19S is set at the center position. Accordingly, when the first sync 18 and the second sync 19 are set to the OFF state, the gear pairs 15, 16, and 17 disposed between the first drive shaft 4 and the driven shaft 11 are all driven. Power is transmitted to the driven shaft 11 from the power transmission path from the engine 1 or the first pump motor 6 to the driven shaft 11 via the first planetary gear mechanism 5 and the first drive shaft 4. Is not transmitted.

  Further, when the start sync 12 is set to the OFF state, the power from the engine 1 is not transmitted to the second planetary gear mechanism 8 and the second pump motor 9, and therefore, from the engine 1 or the second pump motor 9. Power is not transmitted to the driven shaft 11 from the power transmission path that reaches the driven shaft 11 via the second planetary gear mechanism 8 and the second drive shaft 7. Therefore, the driven shaft 11 is in a neutral state in which power is not transmitted from any path.

  At this time, since the first pump motor 6 is in a so-called idling state, even if torque is transmitted from the engine 1 to the ring gear R1 of the first planetary gear mechanism 5, no reaction force acts on the sun gear S1, so that the output element is an output element. Torque is not transmitted to the first drive shaft 4 connected to the carrier C1. Then, the torque from the engine 1 is not transmitted to the second planetary gear mechanism 8, and the second pump motor 9 is stopped without inputting or outputting torque, so the second planetary gear mechanism. Torque is not transmitted to the second drive shaft 7 connected to the carrier C2, which is the output element 8. As a result, the transmission TM is in a neutral state as described above.

  When the shift position is switched to a travel position such as a drive position, the sleeve 19S of the second sync 19 and the sleeve 12S of the start sync 12 are respectively on the right side of FIG. Moved. Accordingly, the first speed driven gear 15B is connected to the driven shaft 11, and the ring gear R2 of the second planetary gear mechanism 8 is connected to the fixed member 14. As a result, the first drive shaft 4 and the driven shaft 11 are connected via the first speed gear pair 15, and the ring gear R2 of the second planetary gear mechanism 8 is fixed.

  That is, the connected state of the gear pair is a state where the first speed is set. As shown in the collinear diagram of FIG. 3, since the ring gear R2 of the second planetary gear mechanism 8 is fixed, the second planetary gear mechanism 8 is connected to the sun gear S2 via the rotor shaft 9A and the second pump motor 9. When the torque output by the second planetary gear mechanism 8 is reduced with respect to the rotational speed of the sun gear S2, in other words, the speed reduction mechanism, in other words, the sun gear S2 is connected to the rotor shaft. When the torque output from the second pump motor 9 is input via 9A, it functions as a speed reduction mechanism that amplifies the torque of the carrier C2, which is the output element of the second planetary gear mechanism 8, with respect to the torque of the sun gear S2. It becomes a state to do.

  Therefore, when the vehicle VE is started, the shift position is switched to the travel position, so that the power of the engine 1 is transferred to the driven shaft 11 via the first planetary gear mechanism 5, the first drive shaft 4, and the first speed gear pair 15. The transmitted power transmission path and the torque output from the second pump motor 9 are amplified by the second planetary gear mechanism 8 and passed through the second drive shaft 7 and the second speed gear pair (second counter gear pair) 13. Two power transmission paths, that is, a power transmission path transmitted to the driven shaft 11, are formed.

  In this state, since the vehicle VE is still stopped, in the first planetary gear mechanism 5, the power is input from the engine 1 to the ring gear R1 while the carrier C1 is stopped. Therefore, the sun gear S1 rotates the ring gear R1. Rotate in the opposite direction. In this state, the extrusion volumes of the pump motors 6 and 9 are gradually increased. First, the first pump motor 6 functions as a pump to generate hydraulic pressure. Then, since the reaction force accompanying it acts on the sun gear S1 in the first planetary gear mechanism 5, a torque appears to rotate the carrier C1 in the same direction as the ring gear R1. As a result, power is transmitted to the driven shaft 11 via the first speed gear pair 15.

  Since the first pump motor 6 rotates in reverse so as to function as a pump, the pressure oil is discharged from the suction port 6S and supplied to the suction port 9S of the second pump motor 9. As a result, the second pump motor 9 functions as a motor, a so-called forward rotation torque is output from the rotor shaft 9A, and the torque is input to the sun gear S2 in the second planetary gear mechanism 8. At this time, the second planetary gear mechanism 8 functions as a reduction mechanism having the ring gear R2 fixed and the carrier C2 as an output element as described above, so that the torque input to the sun gear S2 is the second planetary gear mechanism 8. And is transmitted to the driven shaft 11 via the second drive shaft 7 and the second speed gear pair (second counter gear pair) 13. That is, the torque output from the second pump motor 9 is amplified and transmitted to the driven shaft 11.

  As described above, when the vehicle VE starts, a part of the power input from the engine 1 is transmitted to the driven shaft 11 via the first planetary gear mechanism 5 and the first speed gear pair 15, and other power is compressed. The energy is converted into a flow of oil, which is transmitted to the second pump motor 9, and further from the second pump motor 9 via the second planetary gear mechanism 8 and the second speed gear pair (second counter gear pair) 13. Thus, torque is amplified and transmitted to the driven shaft 11. That is, at the start of the vehicle VE, so-called mechanical power transmission and power transmission through the fluid are performed, and torque is amplified at the time of power transmission through the fluid. Output to the driven shaft 11.

  In the power transmission state as described above, the torque appearing on the driven shaft 11 is larger than the torque in the case of only mechanical transmission via the first speed gear pair 15, and therefore the overall gear ratio of the transmission TM is Therefore, it becomes larger than a so-called fixed transmission gear ratio determined by the first speed gear pair 15. Further, the gear ratio changes in accordance with the transmission ratio of power through the fluid. Therefore, as the rotational speed of the sun gear S2 in the second planetary gear mechanism 8 and the second pump motor 9 connected to the sun gear S2 gradually approaches zero, the rate of power transmission via the fluid decreases, and the entire transmission TM The gear ratio as follows approaches the fixed gear ratio of the first speed. And the extrusion volume of the 1st pump motor 6 increases to the maximum, and that rotation stops, and it becomes the 1st speed which is a fixed gear ratio.

  In this state, the extrusion volume of the second pump motor 9 is set to zero, so that the second pump motor 9 idles and the first pump motor 6 is locked to stop its rotation. That is, since the closed circuit connecting the pump motors 6 and 9 is closed by the second pump motor 9, the first pump motor 6 having the maximum extrusion volume can supply and discharge the pressure oil. Disappears and its rotation is stopped. As a result, a torque for fixing the sun gear S1 of the first planetary gear mechanism 5 acts. Therefore, in the first planetary gear mechanism 5, power is input to the ring gear R1 while the sun gear S1 is fixed. Therefore, a torque is generated in the carrier C1, which is an output element, to rotate it in the same direction as the ring gear R1. It is transmitted to the driven shaft 11 as the output shaft via the 1 drive shaft 4 and the first speed gear pair 15. Thus, the first speed that is the fixed gear ratio is set.

  If the start sync 12 is set to the OFF state in this first speed state, that is, if the sleeve 12S is set to the neutral position, the second pump motor 9 will not be rotated, so that power loss due to so-called drag is avoided. can do. Further, the sleeve 12S of the second synchro 19 is moved to the right side in FIG. 1 and the sleeve 12S of the start sync 12 is moved to the left side in FIG. If the counter driven gear 10C of the one counter gear pair 10 is connected to the ring gear R2 of the second planetary gear mechanism 8, the input shaft 2 is connected to the first counter gear pair 10, the second planetary gear mechanism 8, the second drive shaft 7, and the second drive shaft 7. Since it is connected to the driven shaft 11 via the second speed gear pair (second counter gear pair) 13, it enters an upshift standby state to the second speed, which is a fixed gear ratio. On the other hand, the sleeve 12S of the start sync 12 is moved to the right in FIG. 1 to fix the ring gear R2 of the second planetary gear mechanism 8, and the second planetary gear mechanism 8 is output from the carrier C2 with respect to the input to the sun gear S2. If it is in a state of functioning as a speed reduction mechanism, the downshift standby state in which a gear ratio larger than the first speed is set.

  In the upshift standby state from the first speed to the second speed, the second pump motor 9 and the sun gear S2 connected thereto rotate in the opposite direction to the ring gear R2. Therefore, when the extrusion volume of the second pump motor 9 is increased in the positive direction, the second pump motor 9 functions as a pump, and a reaction force associated therewith acts on the sun gear S2. As a result, a torque obtained by synthesizing the torque input to the ring gear R2 and the reaction force acting on the sun gear S2 acts on the carrier C2, which rotates in the forward direction, and the rotational speed gradually increases. In other words, the rotational speed of the engine 1 is gradually reduced. Torque is transmitted from the carrier C 2 to the driven shaft 11 via the second drive shaft 7 and the second speed gear pair (second counter gear pair) 13.

  Pressure oil generated when the second pump motor 9 functions as a pump is supplied from the suction port 9S to the suction port 6S of the first pump motor 6. Therefore, the first pump motor 6 functions as a motor and outputs torque in the forward rotation direction, which acts on the sun gear S 1 of the first planetary gear mechanism 5. Since power is input from the engine 1 to the ring gear R1 of the first planetary gear mechanism 5, the torque and torque acting on the sun gear S1 are combined and output from the carrier C1 to the first drive shaft 4. That is, power transmission via hydraulic pressure is generated in parallel with mechanical power transmission, and the combined power is transmitted to the driven shaft 11.

  As the rotational speed of the second pump motor 9 gradually decreases, the rate of mechanical power transmission through the second planetary gear mechanism 8 and the second speed gear pair (second counter gear pair) 13 gradually increases. The transmission ratio of the transmission TM as a whole gradually decreases from the transmission ratio determined by the first speed gear pair 15 to the transmission ratio determined by the second speed gear pair (second counter gear pair) 13. The change is a continuous change as in the case of changing to the first speed, which is a fixed gear ratio, after starting. That is, continuously variable transmission is achieved. And the extrusion volume of the 2nd pump motor 9 increases to the maximum, and that rotation stops, and it becomes the 2nd speed which is a fixed gear ratio.

  In this state, the extrusion volume of the first pump motor 6 is set to zero, so that the first pump motor 6 idles and the second pump motor 9 is locked to stop its rotation. That is, since the closed circuit connecting the pump motors 6 and 9 is closed by the first pump motor 6, the second pump motor 9 having the maximum extrusion volume can supply and discharge the pressure oil. Disappears and its rotation is stopped. As a result, the torque for fixing the sun gear S2 of the second planetary gear mechanism 8 acts. Therefore, in the second planetary gear mechanism 8, since power is input to the ring gear R2 with the sun gear S2 fixed, a torque is generated in the carrier C2, which is an output element, to rotate it in the same direction as the ring gear R2. It is transmitted to a driven shaft 11 as an output shaft via a two-drive shaft 7 and a second speed gear pair (second counter gear pair) 13. Thus, the second speed, which is a fixed gear ratio, is set.

  If the second synchro 19 is set to the OFF state in the second speed state, that is, if the sleeve 19S is set to the neutral position, the first pump motor 6 is not rotated, so that the power loss due to so-called drag is reduced. It can be avoided. Further, when the sleeve 19S of the second synchro 19 is moved to the left side in FIG. 1 and the third speed driven gear 16B is connected to the driven shaft 11, an upshift standby state to the third speed, which is a fixed gear ratio, is established. On the other hand, if the sleeve 19S of the second synchro 19 is moved to the right side in FIG. 1 and the first speed driven gear 15B is connected to the driven shaft 11, a downshift standby state to the first speed is established.

  In the upshift standby state from the second speed to the third speed, the first pump motor 6 and the sun gear S1 connected thereto rotate in the opposite direction to the ring gear R1. Therefore, when the extrusion volume of the first pump motor 6 is increased in the positive direction, the first pump motor 6 functions as a pump, and the reaction force associated therewith acts on the sun gear S1. As a result, a torque obtained by synthesizing the torque input to the ring gear R1 and the reaction force acting on the sun gear S1 acts on the carrier C1 and rotates in the forward direction, and the torque rotates to the first drive shaft 4 and the third speed gear pair 16. Is transmitted to the driven shaft 11 which is the output shaft. Further, the rotational speed of the engine 1 is gradually reduced as the speed ratio decreases.

  Pressure oil generated when the first pump motor 6 functions as a pump is supplied from the suction port 6S to the suction port 9S of the second pump motor 9. Therefore, the second pump motor 9 functions as a motor and outputs torque in the forward rotation direction, which acts on the sun gear S <b> 2 of the second planetary gear mechanism 8. Since the power from the engine 1 is input to the ring gear R2 of the second planetary gear mechanism 8, the torque and the torque acting on the sun gear S2 are combined to generate the second drive shaft 7 and the second counter gear pair from the carrier C2. Is output to (second speed gear pair) 13. That is, power transmission via hydraulic pressure is generated in parallel with mechanical power transmission, and the combined power is transmitted to the driven shaft 11.

  And since the rotational speed of the 1st pump motor 6 falls gradually, the ratio of the mechanical power transmission via the 1st planetary gear mechanism 5 and the 3rd speed gear pair 16 will increase gradually, and the transmission TM as a whole. Is gradually reduced from the speed ratio determined by the second speed gear pair (second counter gear pair) 13 to the speed ratio determined by the third speed gear pair 16. The change is a continuous change as in the case of changing to the first speed that is the fixed gear ratio after the start and the case of upshifting from the first speed to the second speed. That is, continuously variable transmission is achieved. And when the extrusion volume of the 1st pump motor 6 increases to the maximum and the rotation stops, it becomes the 3rd speed which is a fixed gear ratio.

  In this state, the extrusion volume of the second pump motor 9 is set to zero, so that the second pump motor 9 idles and the first pump motor 6 is locked to stop its rotation. That is, since the closed circuit connecting the pump motors 6 and 9 is closed by the second pump motor 9, the first pump motor 6 having the maximum extrusion volume can supply and discharge the pressure oil. Disappears and its rotation is stopped. As a result, a torque for fixing the sun gear S1 of the first planetary gear mechanism 5 acts. Therefore, in the first planetary gear mechanism 5, power is input to the ring gear R1 while the sun gear S1 is fixed. Therefore, a torque is generated in the carrier C1, which is an output element, to rotate it in the same direction as the ring gear R1. It is transmitted to the driven shaft 11 as the output shaft via the 1 drive shaft 4 and the third speed gear pair 16. Thus, the third speed, which is a fixed gear ratio, is set.

  Next, the reverse gear will be described. When an instruction to set the reverse gear is issued, for example, when the shift position is switched from the neutral position to the reverse position, the sleeve 12S of the start sync 12 is moved to the right side in FIG. The ring gear R2 is connected to the fixing member 14, and the ring gear R2 is fixed. Further, the sleeve 18S of the first sync 18 is moved to the left in FIG. 1, the reverse driven gear 17B is connected to the driven shaft 11, and the second sync 19 is set to the OFF state. That is, a power transmission path from the input shaft 2 to the driven shaft 11 via the first planetary gear mechanism 5, the first drive shaft 4 and the reverse gear pair 17, and from the rotor shaft 9 A of the second pump motor 9 to the second Two power transmission paths are formed: a planetary gear mechanism 8, a second drive shaft 7, and a power transmission path reaching the driven shaft 11 via the second speed gear pair (second counter gear pair) 13.

  In this state, the extrusion volume of the first pump motor 6 is gradually increased. Further, the extrusion volume of the second pump motor 9 is gradually increased in the negative direction opposite to the above-described forward stage (forward travel). Since the driven shaft 11 does not rotate when the vehicle VE is stopped, the second pump motor 9 connected thereto is stopped. On the other hand, in the first planetary gear mechanism 5, power is input from the engine 1 to the ring gear R1 in a state where the carrier C1 connected to the first drive shaft 4 is fixed. The first pump motor 6 is rotating in the opposite direction to the ring gear R1.

  Therefore, when the torque capacity of the first pump motor 6 is gradually increased, the first pump motor 6 functions as a pump and generates hydraulic pressure. The accompanying reaction force acts on the sun gear S <b> 1, so that a torque is generated in the carrier C <b> 1 that is an output element in the same direction as when traveling forward, and this is transmitted to the first drive shaft 4. Since the reverse gear pair 17 disposed between the first drive shaft 4 and the driven shaft 11 includes an idle gear 17C, when the first drive shaft 4 rotates in the same direction as when traveling forward, the driven gear pair 17 is driven. The shaft 11 rotates in the opposite direction and therefore travels backward.

  Next, an example of control by the control device for the transmission TM of the present invention will be described. As described above, the variable displacement hydraulic pump motor type transmission TM that is the subject of the present invention increases the vehicle speed after the brake is released, such as when starting on an uphill or uneven road. When it takes a long time to complete, the state where the relief valve is opened or semi-opened is continued for a long time compared to the normal start. As a result, there are problems such as a decrease in efficiency and heat generation of the relief valve. Therefore, in the control device of the present invention, the elapsed time after the brake is released at the start of the vehicle VE is detected, and from the elapsed time from the release of the brake and the representative value of the required driving force such as the accelerator opening, A target value of the ratio of the flow rate of oil relieved by the relief valves 27 and 28 with respect to the extrusion volumes (discharge flow rates) of the pump motors 6 and 9 is set, and the actual relief valves 27 and 28 are set to the target values. The pump volumes of the pump motors 6 and 9 are configured to be controlled so that the flow rates of oil passing therethrough coincide with each other.

  FIG. 4 is a flowchart for explaining an example of control by the control device of the present invention. The routine shown in this flowchart is repeatedly executed every predetermined short time. In FIG. 4, first, it is determined whether or not the shift range of the transmission TM is a drive range (drive position) (step S1). The drive range here is a shift range that is set when the vehicle VE moves forward or backward, and is, for example, a D (drive) range, an L (low) range, or an R (reverse) range.

  If the determination is negative in step S1 because the shift range of the transmission TM is not the drive range, for example, the shift range of the transmission TM is the N (neutral) range or the P (parking) range. In step S2, the brake release time, that is, the measured value of the brake release elapsed time timer for measuring the elapsed time after the brake is released is cleared, and then this routine is temporarily ended.

  On the other hand, the shift range of the transmission TM is a drive range, that is, the shift range of the transmission TM is a D (drive) range, an L (low) range, an R (reverse) range, or the like. Thus, if the determination in step S1 is affirmative, the process proceeds to step S3, and it is determined whether or not the relief valves 27 and 28 are open.

  Here, the open state of the relief valves 27 and 28 is a state including a half-open state in which the relief valves 27 and 28 are opened at a predetermined ratio. If the result of step S3 is negative, the routine proceeds to step S2 where the measured value of the elapsed time timer after release of the brake is cleared, and this routine is once terminated.

  On the other hand, the fact that each relief valve 27, 28 is not completely closed, that is, each relief valve 27, 28 is fully opened or half-opened, is affirmative in step S3. If it is determined in step S4, the process proceeds to step S4, and it is determined whether or not the brake of the vehicle VE is released. The determination of the brake release state of the vehicle VE can be made based on, for example, an ON / OFF signal of a brake switch (not shown). That is, it can be determined that the brake of the vehicle VE is released when the brake switch is OFF, and the brake of the vehicle VE is operated when the brake switch is ON.

  If a negative determination is made in step S4 because the brake of the vehicle VE has not been released, the process proceeds to step S2, the measured value of the elapsed time timer after release of the brake is cleared, and this routine is once ended. .

  On the other hand, if a positive determination is made in step S4 because the brake of the vehicle VE is released, the process proceeds to step S5, and the elapsed time timer after the brake release is increased by the control period.

  Subsequently, the target relief ratio Rrlf_ref of each relief valve 27, 28 is obtained (step S6). The relief ratio Rrlf is the ratio of the flow rate of oil relieved from the relief valves 27 and 28 to the flow rate of oil passing through the relief valves 27 and 28. In other words, it is the ratio of the flow rate of oil relieved by the relief valves 27 and 28 to the oil discharge flow rate of the pump motors 6 and 9. The target relief ratio Rrlf_ref is stored in a map as shown in FIG. 5 in advance, for example, for each accelerator opening, the transition with respect to the time of the relief ratio Rrlf of the relief valves 27 and 28 when starting on a flat road. Further, it can be obtained using the map shown in FIG. 5 based on the detected value of the accelerator opening and the elapsed time after brake release.

Further, the target gear ratio γpm_ref of the transmission TM is calculated (step S7). Specifically, the extrusion volume of the first pump motor 6 and the second pump when assuming that the input shaft rotational speed of the transmission TM is Nin, the output shaft rotational speed is Nout, and the relief valves 27 and 28 are closed. Assuming that the transmission ratio of the transmission TM (ratio between the input shaft rotation speed and the output shaft rotation speed), which is a value determined from the ratio with the extrusion volume of the motor 9, is γpm, the relief ratio Rrlf of the relief valves 27 and 28 is
Rrlf = Nin−Nout × γpm
From this relational expression, the gear ratio for realizing the target relief ratio Rrlf_ref, that is, the target gear ratio γpm_ref is
γpm_ref = {Nin × (1-Rrlf_ref)} / Nout
Can be obtained as

  When the target gear ratio γpm_ref is calculated, the extrusion volumes Q1 and Q2 of the pump motors 6 and 9 that realize the target gear ratio γpm_ref are obtained (step S8). These extrusion volumes Q1 and Q2 can be obtained from a map showing the relationship between the extrusion volumes Q1 and Q2 of the pump motors 6 and 9 with respect to the transmission ratio γpm of the transmission TM as shown in FIG. 6, for example.

  Then, the pump motors 6 and 9 are controlled based on the extrusion volumes Q1 and Q2 obtained in step S8. Specifically, for example, linear solenoid current control for operating the control units of the pump motors 6 and 9 is executed so that the extrusion volumes of the pump motors 6 and 9 become the extrusion volumes Q1 and Q2, respectively. Thereafter, this routine is once terminated.

  As described above, according to the control device of the present invention, when the brake of the vehicle VE is activated and the vehicle VE is braked and stopped, the braking by the brake is released, so that the vehicle by the driver is The intention to start is detected. For example, when a brake switch ON / OFF signal is detected and it is detected that the brake switch has been turned OFF, it can be determined that the driver has the intention to start the vehicle.

  When the driver's intention to start the vehicle VE is detected, the elapsed time from when the driver's intention to start is detected is measured, and the elapsed time, for example, the accelerator opening or the throttle opening of the engine 1 is measured. Based on the representative required driving force, the flow rate of oil to be passed through the relief valves 27 and 28 is obtained. And the extrusion volume of each pump motor 6 and 9 is controlled so that the oil of the calculated | required flow volume may actually pass each relief valve 27 and 28, respectively. Accordingly, by controlling the push-out volume of each pump motor 6, 9, after the intention to start is detected without changing the set pressure of each relief valve 27, 28, that is, after the brake of the vehicle VE is released. The driving force of the vehicle VE is appropriately controlled according to the elapsed time. Therefore, heat generation and reduction in efficiency of the relief valves 27 and 28 can be avoided or suppressed, and an increase in load on the engine 1 when the vehicle VE is started without complicating the control contents of the engine 1. Can be suppressed.

  For example, when starting on a road surface with high running resistance such as an uphill road or rough terrain, a large driving force is required compared to normal starting on a flat road, etc., and the vehicle speed is released after the brake is released for starting. Takes longer than usual to rise. At this time, in the conventional control, the elapsed time after releasing the brake increases, so that the time that the oil is continuously relieved by the relief valves 27 and 28 becomes longer, and as a result, the heat generation in the hydraulic circuit increases. In addition, the efficiency is lowered due to a large amount of oil being continuously relieved. On the other hand, according to the control of the present invention, the target gear ratio γpm_ref is set large in order to bring the relief ratio Rrlf close to the target value. As a result, the ratio of the output torque to the input torque of the transmission TM is As a result, the driving force of the vehicle VE can be increased. Therefore, even when a large driving force is required at the time of starting, the driving force of the vehicle VE can be automatically controlled to an appropriate magnitude without performing an adjustment operation by the driver.

  Further, when the driving force of the vehicle VE is controlled, the set pressures of the relief valves 27 and 28 are not changed, and therefore the load on the engine 1 is substantially constant and does not change, so that the driving force is controlled. There is no need to increase or decrease the output of the engine 1. As a result, a reduction in fuel consumption of the engine 1 and an engine stall can be avoided, and the control structure of the engine 1 can be simplified.

  Further, for example, information such as road surface gradient, road surface μ, road surface unevenness, or steering angle is not necessary for correcting the driving force of the vehicle VE, so that there is no need to provide special sensors for detecting them. It is possible to simplify the structure of the apparatus and avoid an increase in cost.

  Here, the relationship between the above-described specific example and the present invention will be briefly described. The functional means of step S4 described above corresponds to the start detection means of the present invention, and the functional means of steps S4 and S5 are this. This corresponds to the elapsed time detecting means of the invention. The functional means in step S6 corresponds to the pressure regulating valve flow rate calculating means of the present invention, and the functional means in steps S7, S8, and S9 correspond to the pump / motor control means of the present invention.

  The present invention is not limited to the above specific example, and the target transmission may be other than the configuration shown in FIG. 1, for example, as shown in FIG. The transmission may be configured to transmit the power of No. 1 to the drive wheel 31 via the differential 30 and perform a shift. That is, it may be a hydrostatic transmission (hydrostatic transmission: HST). Further, as shown in FIG. 8, the transmission may be configured such that the HST is provided in parallel with the speed change mechanism mainly including the gear mechanisms 32 and 33 so that the speed can be changed steplessly as a whole. In addition, in the example shown in FIG. 1, the fixed transmission gear ratio is set so that three forward speeds and one reverse speed can be set. It may be more or less.

  Also, when the pump motor is used as a reaction force mechanism for a differential mechanism such as a single pinion type planetary gear mechanism or a double pinion type planetary gear mechanism, it is a so-called one-way swing type that can increase its pushing volume only in one direction from zero. Not limited to this, a so-called double swing type pump motor that can be changed in both positive and negative directions can also be used. In that case, the gear mechanism may have a configuration different from that shown in FIG.

  Further, the arrangement of transmission mechanisms such as a pump motor, a differential mechanism, and a gear pair can be appropriately changed as necessary. Furthermore, the power source may be connected to the idle gear of the counter gear pair described above instead of being directly connected to one of the differential mechanisms. Further, a mechanism such as a belt or a chain may be used instead of the gear pair. The power source in the present invention does not have to be an engine, and may be an electric motor or a hybrid drive device that combines an internal combustion engine and an electric motor.

It is a skeleton figure which shows typically an example of the transmission made into object by this invention. FIG. 2 is a chart collectively showing operation states of pump motors and synchros when setting respective gear ratios in the transmission shown in FIG. 1. FIG. It is a collinear diagram which shows operation | movement of each rotating member at the time of start of the transmission made into object by this invention. It is a flowchart for demonstrating the example of control in the control apparatus of this invention. It is an example of the map used by control of step S6 in the flowchart of FIG. It is an example of the map used by control of step S8 in the flowchart of FIG. It is a skeleton figure which shows typically the other example of the transmission made into object by this invention. It is a skeleton figure which shows typically the other example of the transmission made into object by this invention.

Explanation of symbols

  DESCRIPTION OF SYMBOLS 1 ... Engine (power source), 6 ... 1st pump motor (hydraulic pump or hydraulic motor), 9 ... 2nd pump motor (hydraulic pump or hydraulic motor), 11 ... Driven shaft (output member), 20, 21 ... Oil Road, 27, 28 ... Relief valve (pressure regulating valve), 29 ... Electronic control unit (ECU).

Claims (3)

A variable displacement hydraulic pump driven by the power output from the power source is connected to the hydraulic pump in a closed circuit so that pressure oil can be exchanged between them, and the pressure oil output from the hydraulic pump is supplied. And a variable displacement hydraulic motor that outputs power to the output member by being driven, and a pressure regulating valve that can regulate the hydraulic pressure between the hydraulic pump and the hydraulic motor, and is transmitted to the output member. In a control device for a hydraulic transmission for a vehicle in which the torque to be changed changes according to the extrusion volume and hydraulic pressure of these hydraulic pumps and hydraulic motors,
Start detection means for detecting the vehicle's start intention by the driver;
An elapsed time detection means for detecting an elapsed time from the time when the start intention is detected by the start detection means;
A pressure regulating valve flow rate calculating means for calculating the amount of oil to be passed through the pressure regulating valve based on the elapsed time detected by the elapsed time detecting means and the required driving force for the vehicle;
Control of a vehicle hydraulic transmission, comprising: a pump / motor control means for controlling the hydraulic pump and the hydraulic motor so as to discharge the oil amount calculated by the pressure regulating valve flow rate calculation means. apparatus.   The start detection means includes means for detecting the intention to start by detecting that the braking by the braking device is released when the vehicle is stopped by being braked by the braking device. The control device for a hydraulic transmission for a vehicle according to claim 1.   3. The vehicle hydraulic system according to claim 1, wherein each of the hydraulic pump and the hydraulic motor includes a variable displacement type hydraulic pump motor having both a function as a hydraulic pump and a function as a hydraulic motor. Transmission control device.

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