JP2007040479A - Transmission for vehicle - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a transmission having less power loss, enabling easy gear shift, and having excellent acceleration performance. <P>SOLUTION: This transmission for a vehicle comprises at least two fluid pressure pumps 12, 13 in which shaft torques are set according to at least one of an extruding volume and a delivery pressure, first transmission mechanisms 19-21 transmitting a torque output from a power source 1 to an output shaft 18 according to the shaft torque of the first fluid pressure pump 12, and second transmission mechanisms 22, 23 transmitting the torque output from the power source 1 to the output shaft 18 according to the shaft torque of the second fluid pressure pump 13 at a reduction gear ratio different from that of the first transmission mechanisms 19-21. The transmission for the vehicle further comprises an acceleration control means 17 setting the shaft torques of the fluid pressure pumps 12, 13 at predetermined values or higher when an increase in the torque output from the output shaft 18 is requested and transmitting the torque to the output shaft 18 through the transmission mechanisms 19-23. <P>COPYRIGHT: (C)2007,JPO&INPIT

Description

  The present invention relates to a transmission for a vehicle capable of controlling the torque transmitted to an output member in accordance with the extrusion volume or discharge pressure of a fluid pressure pump and capable of continuously changing a gear ratio. It is.

  The transmission selectively forms a plurality of power transmission paths between the input member and the output member, and makes the speed of rotation between the input member and the output member different by changing the acceleration / deceleration ratio in each power transmission path. It is a power transmission device configured to set a speed ratio as a ratio to a plurality of speed ratios. It is well known that this type of transmission is mounted on a vehicle. As a transmission for a vehicle, there are a large number of gear ratios that can be set, a small size and light weight, and power transmission efficiency. It is required to be expensive. Thus, for example, Patent Document 1 describes a transmission that can set seven or more shift stages and that can be downsized.

  The transmission described in Patent Document 1 is a so-called twin clutch type stepped transmission, which is connected to an engine via a first clutch and a first input shaft connected to the engine via a first clutch. The second input shaft, the output shaft, the auxiliary shaft connected to the first input shaft via a gear pair, and the first input shaft and the auxiliary shaft are provided and selectively engaged by the mesh clutch mechanism. And a plurality of gear pairs that are provided between the second input shaft and the output shaft and that are selectively connected by a meshing clutch mechanism. The transmission includes a gear stage that transmits torque from any one of the input shafts to the output shaft through a predetermined gear pair, and any output shaft from the input shaft to the output shaft through the predetermined gear pair and the sub shaft. It is configured to set a gear stage for transmitting torque, and as a result, it is configured to set seven or more gear stages including the reverse gear.

  Patent Document 2 describes a hydraulic power transmission device that functions as a clutch between an engine and a transmission.

The ultimate structure with as many gears as possible is a continuously variable transmission that can continuously change the gear ratio. According to the continuously variable transmission, the speed of the power source such as the engine is operated. It can be set to the optimum number of revolutions considering efficiency and the driving force can be changed in various ways. As an example of the continuously variable transmission, there is a configuration in which a stepped transmission unit using a gear train and a continuously variable transmission unit using hydraulic pressure are arranged in parallel between an input shaft and an output shaft. It is described in Patent Document 4.
Japanese Patent Application Laid-Open No. 2003-120764 JP-A-8-284777 JP 11-51150 A JP 2000-320644 A

  In the transmission described in Patent Document 1 described above, since the number of shift speeds that can be set is large, the engine can be operated in a state with good fuel efficiency, and the auxiliary shaft is effectively used. The transmission is reduced in size and weight as a whole, and as a result, the fuel consumption of the vehicle can be improved.

  However, the first clutch and the second clutch used for setting and changing the power transmission path are constituted by a hydraulic friction clutch to absorb the inertial force at the time of shifting transition, Therefore, there is room for improvement in terms of energy efficiency and shift response. In other words, since the hydraulic friction clutch is engaged by pressing the friction plate with hydraulic pressure, the clutch is engaged even in a steady state where the vehicle is traveling with a predetermined gear set. It is necessary to generate the hydraulic pressure, and power for that is always consumed.

  In addition, the clutch that is not involved in the transmission of torque is controlled in a so-called released state, but drag torque is generated by relative rotation of the friction plate, and power loss is caused by the accompanying friction. In addition, since heat is generated at that time, it is necessary to always supply lubricating oil for cooling, and power is consumed for the lubrication, which may increase power loss.

  Further, when the clutch in the released state is engaged, after the clearance between the friction plates is clogged, the friction plates are substantially engaged to transmit torque. Therefore, the time until the clearance is blocked becomes a delay time. In particular, the transmission described in Patent Document 1 is a so-called clutch-to-clutch shift in which the release of one clutch and the engagement of the other clutch proceed in a coordinated manner. Thus, the engagement or disengagement is advanced, so that not only complicated control is forced, but also the shift response is not always good.

  On the other hand, although the power transmission device described in Patent Document 2 has a function as a clutch, Patent Document 2 does not disclose the use of the power transmission device for shifting gear ratios or the technology therefor. . Further, even if the power transmission device described in Patent Document 2 is used in the transmission described in Patent Document 1, the transmission is a so-called stepped transmission, and the gear ratio is continuously changed. It is not possible.

  Further, the transmission described in Patent Document 3 or Patent Document 4 can function as a continuously variable transmission that continuously changes the gear ratio. However, in a continuously variable transmission using hydraulic pressure, While generating hydraulic pressure, the pressure oil is supplied to the motor. Therefore, loss due to oil agitation and friction or loss due to leakage occurs constantly and unavoidably, and as a result, power transmission efficiency is not necessarily good, and the fuel consumption of the vehicle as a whole may deteriorate. .

  The present invention has been made paying attention to the above technical problem, and as a whole, the energy efficiency of the vehicle is good, the gear ratio can be continuously changed, and the speed change for a vehicle with good acceleration performance. The purpose is to provide a machine.

  In order to achieve the above object, the invention of claim 1 is directed to at least two fluid pressure pumps in which shaft torque is set in accordance with at least one of extrusion volume and discharge pressure, and torque output from a power source. Is transmitted to the output member according to the axial torque of the first fluid pressure pump, and the torque output from the power source is determined according to the shaft torque of the second fluid pressure pump and the first transmission. In a vehicle transmission having a second transmission mechanism that transmits to the output member at a gear ratio different from that of the mechanism, the shaft torque of each of the fluid pressure pumps exceeds a predetermined value when an increase in torque output from the output member is requested. And an acceleration control means for transmitting torque to the output member via each transmission mechanism.

  According to a second aspect of the invention, in the first aspect of the invention, each of the fluid pressure pumps includes a variable displacement fluid pressure pump having a variable push-out volume, and the acceleration control means includes the variable displacement fluid pressure pump. It is a vehicle transmission characterized by including the means to change the extrusion volume of this.

  The invention of claim 3 is the vehicle transmission according to claim 1 or 2, wherein the acceleration control means includes means for controlling the discharge pressure of each fluid pressure pump.

  According to a fourth aspect of the present invention, in the third aspect of the present invention, when the acceleration control means controls the discharge pressure of each fluid pressure pump, the pushing capacity of the variable displacement fluid pressure pump is fixed to the maximum. The vehicle transmission further includes means.

  According to the first aspect of the present invention, the torque transmitted from the first transmission mechanism to the output member and the torque transmitted from the second transmission mechanism to the output member are provided corresponding to each transmission mechanism. The torque corresponds to the extrusion volume or discharge pressure of the fluid pressure pump. Therefore, if the push-out volume or discharge pressure of one of the fluid pressure pumps is set to zero and the push-out volume and discharge pressure of the other fluid pressure pump are set to predetermined values, the transmission mechanism corresponding to the other fluid pressure pump As a result, torque is transmitted to the output member, and as a result, the transmission ratio by the other transmission mechanism is set. Further, in the process of changing the state in which torque is transmitted to the output member via one transmission mechanism to the state in which torque is transmitted via the other transmission mechanism, an intermediate value of the gear ratio of each transmission mechanism However, since the transmission ratio as a whole of the transmission is obtained, the transmission ratio as a vehicle can be continuously changed. Further, when there is an acceleration request for starting or the like, the shaft torque of each fluid pressure pump is set to a predetermined value or more, and as a result, torque is output to the output member via both the first and second transmission mechanisms. Is transmitted. Therefore, the torque transmitted to the output member is increased, and the acceleration performance as a vehicle is improved. In addition, when a large torque is required, since a plurality of fluid pressure pumps are responsible for the shaft torque so as to satisfy the request, the capacity required for each fluid pressure pump is reduced. As a result, the fluid pressure pump and the transmission are reduced. Can be reduced in size and weight.

  According to the invention of claim 2, since the shaft torque of the fluid pressure pump is controlled by the pushing volume of the fluid pressure pump, unlike the case where pressure control is performed, the speed ratio or drive torque is controlled stably. In addition, the valve that regulates the discharge pressure of the fluid pressure pump does not have to include a pressure variable mechanism, and thus the configuration of the valve and the hydraulic control mechanism can be simplified.

  Furthermore, according to the invention of claim 3, since the shaft torque of the fluid pressure pump is controlled by the discharge pressure, the responsiveness of torque control can be improved.

  According to the invention of claim 4, since the discharge pressure is controlled in a state where the extrusion volume of the fluid pressure pump is maximized, the discharge pressure for controlling the shaft torque or the drive torque can be relatively lowered. As a result, the leakage of the pressure fluid can be suppressed, the power transmission efficiency and the vehicle fuel consumption can be improved, and the controllability can be improved.

  Next, the present invention will be described based on specific examples. FIG. 1 shows an example of the present invention in a skeleton diagram, and the example shown here is an example in which five transmission ratios are set as so-called fixed transmission ratios that can be set by transmitting torque without using fluid. It is. That is, the input member 2 is connected to the power source (E / G) 1, and the torque is transmitted from the input member 2 to the first planetary gear mechanism 3 and the second planetary gear mechanism 4.

  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. Further, an appropriate transmission means such as a damper, a clutch, or a torque converter may be interposed between the power source 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. These planetary gear mechanisms 3 and 4 are constituted by single pinion type planetary gear mechanisms, which are sun gears 3S and 4S that are external gears, and internal gears that are arranged concentrically with the sun gears 3S and 4S. There are provided ring gears 3R, 4R, and carriers 3C, 4C holding pinion gears meshed with the sun gears 3S, 4S and the ring gears 3R, 4R so as to freely rotate and revolve. 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, and a first rotating shaft 9 is rotatably inserted therein, and one end portion of the first rotating shaft 9 is a reaction force in the first planetary gear mechanism 3. The sun gear 3S as an element is connected so as to rotate integrally.

  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 second rotating shaft 11 is rotatably inserted. One end of the second rotating shaft 11 is a reaction force in the second planetary gear mechanism 4. The sun gear 4S, which is an element, is connected to rotate integrally.

  The other end of the first rotary shaft 9 is connected to the output shaft of the variable displacement pump motor 12 corresponding to the first fluid pressure pump in the present invention. The variable displacement pump motor 12 is a fluid pressure (hydraulic) pump capable of changing the discharge capacity (push-out volume) such as a slant shaft pump, a swash plate pump, or a radial piston pump, and applies torque to its output shaft. By rotating, it discharges pressure fluid (pressure oil) functioning as a pump, and by supplying pressure fluid from a discharge port when functioning as a pump and discharging it from an intake port when functioning as a pump It is designed to function as a motor. In the following description, the variable displacement pump motor 12 is referred to as a first pump motor 12, and is indicated as P / M1 in the figure.

  The other end of the second rotating shaft 11 is connected to the output shaft of the variable displacement pump motor 13 corresponding to the second fluid pressure pump in the present invention. The variable displacement pump motor 13 is a fluid pressure (hydraulic) pump capable of changing the discharge capacity, such as an oblique shaft pump, a swash plate pump, or a radial piston pump, and is rotated by applying torque to its output shaft. Functions as a motor by discharging pressure fluid (pressure oil) functioning as a pump, supplying pressure fluid from a discharge port when functioning as a pump, and discharging from a suction port when functioning as a pump It is supposed to be. In the following description, the variable displacement pump motor 13 is referred to as a second pump motor 13 and is indicated as P / M2 in the figure.

  The pump motors 12 and 13 are communicated with each other by 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. The hydraulic control device 16 mainly includes a valve for controlling the amount and pressure of the pressure oil flowing through the oil passages 14 and 15, that is, the pushing volume and pressure of the pump motors 12 and 13, and the oil passages 14 and 15. Is intervened. Furthermore, an electronic control unit (ECU) 17 for controlling the discharge capacity of the hydraulic control device 16 and the pump motors 12 and 13 is provided. That is, a command is sent from the electronic control unit 17 to an actuator (not shown) for changing the angle of the swash plate or the oblique axis for setting the discharge capacity or the phase angle of the cam ring (not shown) of the radial piston pump. A signal is output.

  An output shaft 18 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 18 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 19, 20, 21, 22, and 23 are used. Specifically, on the first intermediate shaft 8, a fifth speed drive gear 19A, a third speed drive gear 20A, and a first speed drive gear 21A are arranged in order from the first planetary gear mechanism 3 side. The fifth speed drive gear 19A and the third speed drive gear 20A are rotatably fitted to the first intermediate shaft 8, and the first speed drive gear 21A is integrated with the first intermediate shaft 8. It is attached to rotate.

  A fifth speed driven gear 19B meshed with the fifth speed drive gear 19A and a third speed driven gear 20B meshed with the third speed drive gear 20A are attached to the output shaft 18 so as to rotate integrally. Yes. A first speed driven gear 21B meshed with the first speed drive gear 21A is rotatably fitted to the output shaft 18. Therefore, the fifth speed gear pair 19, the third speed gear pair 20, and the first speed gear pair 21 comprising the drive gears 19A, 20A, 21A and the driven gears 19B, 20B, 21B are the first transmission in the present invention. It corresponds to the mechanism.

  Further, a fourth speed drive gear 22A meshed with the fifth speed driven gear 19B and a second speed drive gear 23A meshed with the third speed driven gear 20B are rotatable on the second intermediate shaft 10. It is made to fit. Accordingly, the fifth speed driven gear 19B also serves as the fourth speed driven gear, and the third speed driven gear 20B also serves as the second speed driven gear. Accordingly, the fourth speed gear pair 22 and the second speed gear pair 23 composed of the drive gears 22A and 23A for the fourth speed and the second speed and the driven gears 19B and 20B meshing with the drive gears 22A and 23A for the fourth speed are provided in the present invention. This corresponds to the second transmission mechanism in FIG.

  Here, the gear ratio of each gear pair 19, 20, 21, 22, 23 (ratio of the number of teeth of the driven gear to the number of teeth of the respective drive gear) will be described. 21, a second speed gear pair 23, a third speed gear pair 20, a fourth speed gear pair 22, and a fifth speed gear pair 19.

  In order to make each of the first-speed to fifth-speed gear pairs 19, 20, 21, 22, and 23 transmit torque between any of the intermediate shafts 8 and 10 and the output shaft 18. The engagement mechanism is provided. In short, this engagement mechanism is a mechanism that selectively transmits torque, and conventionally known mechanisms such as a dog clutch mechanism and a synchronous coupling mechanism (synchronizer) can be employed. An example employing a synchronizer is shown.

  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 connected by synchronizing the rotating shaft and the rotating member by the frictional contact of the knit ring with the rotating member. A first synchronizer (hereinafter referred to as a first synchronizer) 25 is provided on the output shaft 18 at a position adjacent to the first speed driven gear 21B. The first sync 25 moves the sleeve to the left side of FIG. 1 to connect the first speed driven gear 21B to the output shaft 18, and the first speed gear pair 21 is connected to the first intermediate shaft 8 and the output. Torque is transmitted to and from the shaft 18.

  A second synchronizer (hereinafter referred to as a second synchronizer) 26 is provided on the second intermediate shaft 10 between the fourth speed drive gear 22A and the second speed drive gear 23A. The second sync 26 moves the sleeve to the left side of FIG. 1 to connect the second speed drive gear 23A to the second intermediate shaft 10, and the second speed gear pair 23 is connected to the second intermediate shaft 10. And the output shaft 18 are configured to transmit torque. On the other hand, the fourth speed drive gear 22A is connected to the second intermediate shaft 10 by moving the sleeve to the right in FIG. 1, and the fourth speed gear pair 22 is connected to the second intermediate shaft 10 and the output shaft. Torque is transmitted to and from 18.

  Further, on the first intermediate shaft 8, a third synchronizer (hereinafter referred to as a third synchronizer) 27 is provided between the third speed drive gear 20A and the fifth speed drive gear 19A. The third synchro 27 connects the third-speed drive gear 20A to the first intermediate shaft 8 by moving the sleeve to the left side in FIG. 1, and the third-speed gear pair 20 is connected to the first intermediate shaft 8. And the output shaft 18 are configured to transmit torque. On the other hand, by moving the sleeve to the right in FIG. 1, the fifth speed drive gear 19A is connected to the first intermediate shaft 8, and the fourth speed gear pair 19 is connected to the first intermediate shaft 8 and the output shaft. Torque is transmitted to and from 18.

  These synchros 25, 26, and 27 can be configured to be switched by manual operation, but can be configured to perform so-called automatic control instead. In this case, for example, the above-described sleeve is used. An appropriate actuator (not shown) that moves in the axial direction may be provided, and the actuator may be configured to operate the command signal of the electronic control device 17 described above.

  As described above, the transmission shown in FIG. 1 is configured such that the torque output from the power source 1 is transmitted to the output shaft 18 via one of the intermediate shafts 8 and 10. A differential 29 is connected to the output shaft 18 via a transmission mechanism 28 such as a gear mechanism or a winding transmission mechanism such as a chain, and outputs power to right and left wheels (not shown) from here. It has become.

  The pump motors 12 and 13 described above increase the discharge pressure if the oil passage 15 on the discharge ports 12D and 13D side is throttled in a state where the extrusion volume is a predetermined value of zero or more. Therefore, in this state, the torque required to rotate the output shafts of the pump motors 12 and 13, that is, the shaft torque increases, and in the configuration shown in FIG. 1, the reaction torque of the planetary gear mechanisms 3 and 4 increases. That is, when the pump motors 12 and 13 are of the fixed capacity type, the shaft torque can be controlled by the discharge pressure.

  On the other hand, when the extrusion volume of each pump motor 12 and 13 is increased, if the discharge pressure is maintained (fixed), the shaft torque of each pump motor 12 and 13 increases. Further, when functioning as a motor, if the supplied hydraulic pressure is constant, the torque appearing on the output shaft, that is, the shaft torque increases according to the pushing volume.

  An example of a hydraulic control device 16 that controls the pump motors 12 and 13 having such characteristics is shown in FIG. In the example shown here, the discharge pressures of the pump motors 12 and 13 are kept constant, and the push-out volumes are changed to control the shaft torque. The discharge ports 12D and 13D are connected to each other. Two relief valves 31 and 32 for communicating the oil passages 14 and 15 are provided between the oil passage 15 and the oil passage 14 communicating with the suction ports 12S and 13S. The first relief valve 31 opens when the pressure of the oil passage 15 on the discharge ports 12D, 13D side exceeds a preset pressure, and releases the hydraulic pressure to the oil passage 14 on the suction ports 12S, 13S side. Is configured to do. The second relief valve 32 opens when the pressure of the oil passage 14 on the suction ports 12S, 13S side exceeds a preset pressure, and the oil passage 15 on the discharge ports 12D, 13D side. Configured to release.

  Moreover, the boost pump 33 which supplies pressure oil to each said pump motors 12 and 13 and each oil path 14 and 15 is provided. The boost pump 33 is configured to pump and pressurize the pressure oil from the oil pan 34 by being driven by the power source 1 or an electric motor (not shown) described above, and its discharge port communicates with the boost pressure control valve 35. ing. The boost pressure control valve 35 is basically a pressure regulating valve, and obtains an output pressure corresponding to the signal pressure by causing the signal pressure and the feedback pressure to act against a valve body (not shown). It is configured as follows.

  The boost pressure control valve 35 and the oil passage 15 on the discharge port 12D, 13D side are communicated with each other via a check valve 36 that enables the flow of the pressure oil toward the oil passage 15. Further, the boost pressure control valve 35 and the oil passage 14 on the suction port 12S, 13S side are communicated with each other via a check valve 37 that enables the flow of the pressure oil toward the oil passage 14.

  Therefore, in the configuration shown in FIG. 2, when one or both of the pump motors 12 and 13 function as a pump and a reaction torque is applied to one or both of the planetary gear mechanisms 3 and 4, the pushing volume is increased. Gradually increase. In this case, since the discharge ports 12D and 13D of the pump motors 12 and 13 are communicated with each other, the discharge pressure gradually increases, and the shaft torque of the pump motors 12 and 13 increases accordingly. When the discharge port exceeds the pressure set by the first relief valve 31, the relief valve 31 is opened and the hydraulic pressure is released to the other oil passage 14 side.

  In addition, you may use the thing of the structure which can adjust setting pressure for the 1st relief valve which releases discharge pressure. An example thereof is shown in FIG. 3, and the first relief valve 31A shown here is configured such that a release pressure is set in accordance with a signal pressure Psol supplied from an electromagnetic valve (not shown) or the like. Therefore, when the hydraulic control device having the configuration shown in FIG. 3 is used, not only the extrusion volume of each pump motor 12 and 13 but also its discharge pressure can be controlled. It becomes possible to control also by. The other configuration in FIG. 3 is the same as the configuration shown in FIG. 2, and therefore, the same reference numerals as those in FIG.

  Next, the operation of the transmission described above will be described. FIG. 4 is a chart collectively showing the operating states of the oil pumps (P / M1, P / M2) 12, 13 and the synchros 25, 26, 27 when setting the respective gear positions. “OFF” for the oil pumps 12 and 13 in FIG. 4 makes the pump capacity substantially zero, does not generate pressure oil even if its output shaft is rotated, and outputs even if hydraulic pressure is supplied. Indicates that the shaft does not rotate (free), and “LOCK” indicates a state where the pump capacity is maximized and the oil discharge is restricted and torque appears on the output shaft (in other words, pressure oil is confined). Show. Furthermore, “hydraulic pressure generation” indicates a state in which the pump capacity is made larger than substantially zero and pressure oil is discharged, and therefore the corresponding oil pumps 12 and 13 function as pumps. “Hydraulic pressure recovery” indicates a state in which pressure oil discharged from one oil pump 13 (or 12) is supplied and functions as a motor, and therefore the corresponding oil pump 13 (or 12) has a shaft torque. And driving torque is transmitted to the corresponding intermediate shafts 8 and 10.

  In addition, “right” and “left” for each of the syncs 25, 26, and 27 indicate the positions of the sleeves in the respective syncs 25, 26, and 27 in FIG. 1, and parentheses indicate a standby state for downshifting. , Square brackets indicate a waiting state for upshifting, and "-" indicates that the sleeve is in the center and is in a neutral state.

  When the neutral (N) state is set by selecting a neutral position with a shift device (not shown), the oil pumps 12 and 13 are set to the “OFF” state, and the synchros 25, 26, and 27 are turned on. The sleeve is set to the center position. Accordingly, none of the gear pairs 19, 20, 21, 22, and 23 is in a neutral state that is not connected to the output shaft 18. That is, the oil pumps 12 and 13 are controlled so that the pump capacity becomes substantially zero, and as a result, a so-called idling state is established. Therefore, the power source 1 is connected to the ring gears 3R and 4R of the planetary gear mechanisms 3 and 4, respectively. No torque is transmitted to the intermediate shafts 8 and 10 connected to the carriers 3C and 4C as output elements because no reaction force acts on the sun gears 3S and 4S even if torque is transmitted from the motor.

  The operation states of the planetary gear mechanisms 3 and 4 in this neutral state are shown in a collinear diagram in FIG. In this nomograph, ρ1 and ρ2 indicate the gear ratios of the planetary gear mechanisms 3 and 4 (ratio of the number of teeth of the sun gear and the ring gear), and the arrows indicate the directions of torque. In the neutral state in which the power source 1 is being driven, the ring gears 3R and 4R of the planetary gear mechanisms 3 and 4 are rotating forward, and the carriers 3C and 4C connected to the output shaft 18 are stopped. Therefore, the sun gears 3S and 4S, which are the reaction force elements, and the pump motors 12 and 13 connected to the sun gears 3S and 4S are reversely rotated in a free state.

  When the shift position is switched to a travel position such as a drive position, the sleeve of the first sync 25 and the sleeve of the second sync 26 are moved to the left in FIG. 1 so that the first speed driven gear 21B is moved to the output shaft 18. The second speed drive gear 23 </ b> A is connected to the second intermediate shaft 10 while being connected. That is, the first speed and the second speed, which are fixed gear ratios, are set. In addition, the pump motors 12 and 13 are controlled so as to function as pumps. That is, hydraulic pressure is generated. Specifically, control is performed so that the pushing volumes of the pump motors 12 and 13 are increased to a predetermined value larger than zero. In the hydraulic control device 16 shown in FIG. 3, the set pressure (release pressure) of the first relief valve 31A is set to a high pressure that is substantially zero or higher. That is, when there is an acceleration request such as starting, the transmission mechanisms in each of the intermediate shafts 8 and 10 are set to a state where fixed gear ratios adjacent to each other are set.

  As described above, since the discharge ports 12D and 13D of the pump motors 12 and 13 are communicated with each other, the discharge pressure of the pump motors 12 and 13 reaches the pressure set by the first relief valves 31 and 31A. Increasing gradually. Along with this, the shaft torques of the pump motors 12 and 13 gradually increase, and this acts on the sun gears 3S and 4S of the planetary gear mechanisms 3 and 4 as a reaction torque. The state is shown in a collinear diagram in FIG.

  That is, in the first planetary gear mechanism 3, the torque input from the power source 1 and the reaction torque input from the first pump motor 12 are combined to form the carrier 3C and the first intermediate coupled to the carrier 3C. A torque that causes the shaft 8 to rotate forward is generated, and the torque is transmitted to the output shaft 18 via the first speed gear pair 21. In the second planetary gear mechanism 4, the torque input from the power source 1 and the reaction torque input from the second pump motor 13 are combined to generate the carrier 4 </ b> C and the second intermediate coupled thereto. A torque that causes the shaft 10 to rotate in the forward direction is generated, and the torque is transmitted to the output shaft 18 via the second speed gear pair 23.

Therefore, the torque that is the sum of the torque transmitted through the first speed gear pair 21 and the torque transmitted through the second speed gear pair 23 appears on the output shaft 18. This situation is indicated by thick arrows in FIG. The output torque To is
To ≒ (1 + ρ1) / ρ1 ・ κ1 ・ P15 ・ q1 / 2π + (1 + ρ2) / ρ2 ・ κ2 ・ P15 ・ q2 / 2π
It becomes. The first term on the right side of this equation corresponds to the torque transmitted via the first speed gear pair 21, and the second term corresponds to the torque transmitted via the second speed gear pair 23. Ρ1 is the gear ratio of the first planetary gear mechanism 3, ρ2 is the gear ratio of the second planetary gear mechanism 4, κ1 is the gear ratio of the first speed gear pair 21, κ2 is the gear ratio of the second speed gear pair 23, P15 is the oil pressure of the oil passage 15, q1 is the pushing volume of the first pump motor 12, and q2 is the pushing volume of the second pump motor 13. Since this output torque To is transmitted to the differential 29 via the transmission mechanism 28, a large driving force can be obtained at the time of start. When the vehicle is stopped, it appears as creep force.

  After the vehicle has started, the pushing volume and / or discharge pressure of the first pump motor 12 is gradually increased to control the “LOCK” state, and the pushing volume and / or the second pump motor 13 Alternatively, the discharge pressure is gradually reduced to control the “OFF” state. The states of the planetary gear mechanisms 3 and 4 in this process are shown in a collinear diagram in FIG. 7, and the torque transmitted through the first speed gear pair 21 gradually increases, whereas the second speed gear is increased. The torque transmitted through the pair 23 gradually decreases. When the torque required to drive the first pump motor 12 becomes larger than the input torque and the first motor pump 12 is completely in the “LOCK” state, the rotation of the sun gear 3 </ b> S of the first planetary gear mechanism 3. Is stopped, and the second pump motor 13 is in the “OFF” state (free state), so that the reaction force does not act on the sun gear 4S of the second planetary gear mechanism 4, so that the first speed, which is the fixed gear ratio, is increased. Is set. This state is shown in FIG. 8 as an alignment chart for the planetary gear mechanisms 3 and 4.

  In the first speed state, the second pump motor 13 is in a free state and does not generate shaft torque, so torque transmission through the second planetary gear mechanism 4 and the second intermediate shaft 10 does not occur. Therefore, the sleeve of the second synchro 26 can be set to either the neutral position or the left side of FIG. If the neutral position is set, torque is not transmitted between the output shaft 18 and the second intermediate shaft 10 via the second speed gear pair 23, so that the second intermediate shaft 10 and the second planetary gear mechanism 4 are There is no so-called rotation around the output shaft 18, and as a result, power loss due to drag can be reduced. This state is shown in FIG. 9 as a collinear diagram for the planetary gear mechanisms 3 and 4. Further, if the sleeve of the second sync 26 is positioned on the left side of FIG. 1, a standby state is prepared for shifting to the second speed.

  The second speed is set from the power source 1 via the second planetary gear mechanism 4, the second intermediate shaft 10, and the second speed gear pair 23, so that the pump capacity (push capacity) of the second pump motor 13 in the free state is set. Is gradually increased and the discharge amount is gradually reduced (the discharge pressure is gradually increased) to execute the shift to the second speed. Since the second pump motor 13 rotates in the reverse direction at the first speed, if the pushing volume is increased, the pressure oil generated by functioning as a pump is discharged, and the reaction torque associated therewith is increased by the second planetary gear mechanism 4. Acts on the sun gear 4S. Therefore, the torque from the power source 1 and the reaction torque from the second pump motor 13 are combined by the second planetary gear mechanism 4 and transmitted to the output shaft 18 via the second intermediate shaft 10 and the second speed gear pair 23. Is done. A collinear diagram of the second planetary gear mechanism 4 in this state is shown in FIG.

  Further, since the pressure oil discharged from the second pump motor 13 is supplied to the first pump motor 12, the first pump motor 12 functions as a motor, and the torque is transmitted to the sun gear 3 </ b> S of the first planetary gear mechanism 3. The Therefore, in the first planetary gear mechanism 3, the torque transmitted from the power source 1 and the torque transmitted from the first pump motor 12 are combined, and the combined torque is combined with the first intermediate shaft 8 and the first speed gear pair 21. Is transmitted to the output shaft 18 via. A collinear diagram of the first planetary gear mechanism 3 in this state is shown in FIG.

  Thus, by continuously increasing the push-out volume of the second pump motor 13 and gradually reducing the discharge, the shift to the second speed proceeds, so that a continuously variable shift in which the gear ratio and torque continuously change is executed. The Further, the second pump motor 13 functions as a pump and generates pressure oil in the process of the speed change, but the pressure oil is supplied to the first pump motor 12 and recovered as power. This is possible and is advantageous in improving the fuel consumption of the vehicle.

  As described above, the second speed is achieved by gradually restricting the discharge of the second pump motor 13 to finally zero, that is, by locking. Further, in this second speed state, particularly in a state immediately after being upshifted from the first speed or in a standby state in preparation for a downshift to the first speed, the first pump motor 12 can be freely rotated with zero pushing volume. Is set to a free state. Further, the stable second speed state that is not provided for either the upshift or the downshift is a state in which the second sync 26 is set to the neutral position.

  Thereafter, similarly, the pump motors 12 and 13 are alternately switched between the “LOCK” state and the “OFF” state, and in the process, one of them functions as a pump and the other functions as a motor. , 26, and 27 are appropriately set to the left and right or neutral positions in FIG. 1, and the gear ratio can be continuously changed up to the fifth speed which is the fixed gear ratio.

  By the way, as described with reference to FIGS. 2 and 3 above, the shaft torque of each pump motor 12 and 13 can be appropriately set by changing at least one of the extrusion volume and the discharge pressure. it can. Therefore, in the starting process until the speed ratio of the first speed is set, each pump motor 12 is maintained with the discharge pressure PA, which is the set pressure in the first relief valve 31 (or 31A), being kept constant. , 13 by changing the extrusion volumes q1, q2 and, on the contrary, the discharge pressure, i.e., the first pressure, in the state where the extrusion volumes q1, q2 of the pump motors 12, 13 are kept constant. A method of controlling the shaft torque by changing the set pressure of the one relief valve 31A is possible.

  Hereinafter, each control will be described. FIG. 11 is a time chart for explaining an example of control for changing the pushing volumes q1 and q2 of the pump motors 12 and 13 while the discharge pressure PA is maintained at the predetermined value P1. Therefore, the push-out volumes q1, q2 of the pump motors 12, 13 are gradually increased based on the start control start command signal (at time t1). Accordingly, output torques To1 and To2 transmitted to the output shaft 18 gradually increase in accordance with the shaft torques of the pump motors 12 and 13, respectively. The shaft torque in this type of pump is represented by (q · P / 2π) where q is the extrusion volume and P is the discharge pressure. Accordingly, the output torques To1 and To2 increase in accordance with the pushing volumes q1 and q2.

  When the pushing volumes q1, q2 of the pump motors 12, 13 reach preset volumes, the output torques To1, To2 corresponding to the shaft torques of the pump motors 12, 13 reach the target torque (at time t2). In other words, the extrusion volumes q1 and q2 are set according to the target torque. Further, the output torques To1 and To2 are torques transmitted to the output shaft 18 and are torques after being decelerated by the gear pairs 21 and 23. Therefore, the gear ratios of these gear pairs 21 and 23 are the same. The torques differ from each other based on the difference between κ1 and κ2 (κ1> κ2), and the output torque To1 transmitted through the first speed gear pair 21 is greater than the output torque To2 transmitted through the second speed gear pair 23. growing.

  Thereafter, with the start of control for forming the first speed (at time t3), the pushing volume q1 of the first pump motor 12 increases toward the maximum value so that the first pump motor 12 is in the “LOCK” state. In contrast, the pushing volume q2 of the second pump motor 13 is decreased toward zero so that the second pump motor 13 is in the “OFF” state. In this case, the output torques To1 and To2 reach the maximum values determined by the discharge pressure P1, and therefore the output torque To1 based on the shaft torque of the first pump motor 12 is maintained at the previous torque. The other output torque To2 based on the shaft torque of the two-pump motor 13 decreases as the push-out volume q2 decreases. Then, the first speed is formed at the time point t4 when the pushing volume q2 of the second pump motor 13 becomes zero, and at the same time, the pushing volume q1 is further increased in order to bring the first pump motor 12 into the "LOCK" state. The end control is started, and the discharge pressure PA is lowered to a predetermined value P2. As a result, the first pump motor 12 is “LOCKed”, and the discharge pressure PA is reduced to a predetermined value P2, so that the start control ends (time t5).

  Therefore, in such a start control process, the pump motors 12 and 13 function as pumps, and output torques To1 and To2 corresponding to the respective shaft torques are transmitted to the output shaft 18. Therefore, the torque output from the output shaft 18 is the sum of these two output torques To1 and To2, as represented by the above-described equation. As a result, the output torque To at the time of starting becomes large and the starting acceleration of the vehicle becomes good.

  For comparison, the pushing volume q1 and the output torque To1 when starting using only the first pump motor 12 are indicated by broken lines in FIG. When the pushing volume q1 of the first pump motor 12 is gradually increased as the start control is started, the shaft torque of the first pump motor 12 is gradually increased accordingly. However, in this case, the torque transmitted to the output shaft 18 is limited to the torque transmitted only from the first speed gear pair 21, so that the output torque To1 reaches the target value at time t21 later than the time t2 described above. To do. Accordingly, the pushing volume q1 of the first pump motor 12 is larger than that in the above case where the two pump motors 12 and 13 are used. Thereafter, when the first speed is established (at time t4), the end control is executed in the same manner as in the above-described example according to the present invention.

  As described above, when starting using only one pump motor 12, the output torque To obtained is limited to the torque corresponding to the shaft torque of the pump motor 12, and becomes a relatively small torque. , Starting acceleration is not always sufficient. Further, when only one pump motor 12 is used, it is necessary to increase the pushing volume q1, but in the example according to the present invention described above, two pump motors 12 and 13 are used. The extrusion volumes q1 and q2 can be reduced, so that the pump motors 12 and 13 can be reduced in size, and an efficient operation with less power loss due to leakage or the like can be performed. Furthermore, since the volume is controlled, the control becomes easy and stable control is possible.

  On the other hand, FIG. 12 illustrates an example of control for gradually changing the discharge pressure PA in a state where the pushing volumes q1 and q2 of the pump motors 12 and 13 are fixed to a constant value (specifically, a maximum value). It is a time chart, and the set pressure, that is, the discharge pressure PA by the first relief valve 31A is gradually increased based on the start control start command signal (at time t11). Accordingly, output torques To1 and To2 transmitted to the output shaft 18 gradually increase in accordance with the shaft torques of the pump motors 12 and 13, respectively.

  When the discharge pressure PA reaches a preset pressure P3, the output torques To1 and To2 corresponding to the shaft torques of the pump motors 12 and 13 reach the target torque (at time t12). In other words, the discharge pressure PA is controlled according to the target torque.

  Thereafter, when the control for forming the first speed is started (at time t13), the discharge pressure PA is increased toward the maximum value so that the first pump motor 12 is in the “LOCK” state. At the same time, the pushing volume q2 of the second pump motor 13 is lowered toward zero so that the second pump motor 13 is in the “OFF” state. In this case, the output torques To1 and To2 reach the maximum values determined by the discharge pressure P1, and therefore the output torque To1 based on the shaft torque of the first pump motor 12 is maintained at the previous torque. The other output torque To2 based on the shaft torque of the two-pump motor 13 decreases as the push-out volume q2 decreases. Then, the first speed is formed at time t14 when the pushing volume q2 of the second pump motor 13 becomes zero, and the end control is started at the same time. Thereafter, the start control ends (time t15).

  Changes in the output torques To1 and To2 in the process of such control are the same as in the control example described with reference to FIG.

  For comparison, the pushing volume q1 and the output torque To1 when starting using only the first pump motor 12 are shown by broken lines in FIG. When the discharge pressure PA is gradually increased with the start of the start control, the shaft torque of the first pump motor 12 is gradually increased accordingly. However, in this case, the torque transmitted to the output shaft 18 is limited to the torque transmitted only from the first speed gear pair 21, so that the output torque To1 reaches the target value at time t22 later than the time t12 described above. To do. Accordingly, the discharge port PA is higher than in the above case where the two pump motors 12 and 13 are used.

  As described above, when starting using only one pump motor 12, the output torque To obtained is limited to the torque corresponding to the shaft torque of the pump motor 12, and becomes a relatively small torque. , Starting acceleration is not always sufficient. When only one pump motor 12 is used, it is necessary to increase the discharge pressure PA. However, in the example according to the present invention described above, since the two pump motors 12 and 13 are used, the discharge pressure PA. Therefore, the pump motors 12 and 13 can be downsized, and an efficient operation with less power loss due to leakage or the like can be performed.

  Here, the relationship between the present invention and the above-described specific example will be briefly described. The electronic control unit 17 for controlling the pushing volumes q1, q2 and the discharge pressure PA of the pump motors 12, 13 as described above is described below. This corresponds to the acceleration control means of the invention, and also corresponds to means for fixing the extrusion volume to the maximum value.

  The present invention is not limited to the specific examples described above. That is, the present invention includes a plurality of fluid pressure pumps whose axial torque changes according to the fluid extrusion volume and discharge pressure, and the torque according to each axial torque is transmitted to the output member via the transmission mechanism. Therefore, the transmission may be a transmission configured to transmit torque from each fluid pressure pump directly to the transmission mechanism without using the planetary gear mechanisms 3 and 4 described above. An example is shown below.

  FIG. 13 shows an example of the present invention in a skeleton diagram, and an input member 51 is connected to two fluid pressure pumps 52 and 53. The input member 51 is for transmitting torque of a power source such as an engine or an electric motor (not shown) to the fluid pressure pumps 52 and 53, and is constituted by a rotating shaft, a gear, a winding transmission mechanism, or the like. In the following description, the input member 51 is referred to as an input shaft 51.

  The fluid pressure pumps 52 and 53 are mainly for transmitting torque between the input side member and the output side member. By confining the fluid, the input side member and the output side member are used. One example is a positive displacement oil pump that can limit the relative rotation of the oil pump. In the following description, the fluid pressure pump is referred to as an oil pump.

  14 and 15 show the oil pumps 52 and 53, and these oil pumps 52 and 53 are arranged side by side in the axial direction on the same axis, and are configured as a so-called tandem type twin pump. . That is, the input-side member to which the input shaft 51 is connected is a cylindrical housing 54, and the inner peripheral surface thereof is a wavy curved surface that is smoothly uneven in the radial direction. This inner peripheral surface is a cam surface 55 as will be described later.

  Inside the housing 54, two rotors 56 and 57, which are output side members, are arranged side by side in the axial direction on the same axis. Each of the rotors 56 and 57 is a disk-shaped member having a predetermined thickness, and is supported inside the housing 54 so as to rotate about the central axis of the housing 54. In each of the rotors 56 and 57, a plurality of concave portions 58 that are open on the outer peripheral surface are formed radially about the rotation center axis. These concave portions 58 are portions corresponding to cylinders, and therefore, pistons (or plungers) 59 are accommodated therein so as to reciprocate in the radial direction of the rotors 56 and 57. In the following description, the concave portion 58 is referred to as a cylinder portion 58.

  A rotating body that is in contact with the cam surface 55 that is the inner peripheral surface of the housing 54 is rotatably held at the top portion of each piston 59, that is, the outer end portion in the radial direction of the rotors 56 and 57. The rotating body functions as a cam follower that moves along the cam surface 55, and includes a ball, a roller, or the like. In the following description, an example in which the ball 60 is used as a sphere will be described.

  In order to bring the ball 60 into contact with the cam surface 55 and move along the cam surface 55, a spring 61 that presses the piston 59 toward the cam surface 55 is provided on the bottom side of the piston 59 (the rotors 56 and 57). (Center side in the radial direction) Accordingly, an oil chamber 62 partitioned by the piston 59 is formed inside the cylinder portion 58, and the volume of the oil chamber 62 is changed by the reciprocating motion of the piston 59, so that fluid such as oil is supplied to the oil chamber. The fluid is sucked into the interior of 62 and the fluid is pressurized and discharged. If a positive cam such as a groove cam is formed between the rotors 56 and 57 and the piston 59 instead of the cam surface 55, the return spring 61 can be eliminated.

  The flow path for oil suction and discharge will be described. A suction hole 63 and a discharge hole 64 penetrating in the radial direction are formed in the bottom of each oil chamber 62 (portion on the rotation center side of the rotors 56 and 57). Has been. The suction hole 63 is provided with a suction-side check valve 65 that opens at the time of suction when the volume of the oil chamber 62 increases, and the discharge hole 64 has a discharge-side check that opens at the time of discharge when the volume of the oil chamber 62 decreases. A valve 66 is provided.

  Accordingly, the oil pumps 52 and 53 are cam surfaces formed on the inner peripheral surface of the housing 54 by the relative rotation of the housing 54 that is the input side rotating member and the rotors 56 and 57 that are the output side rotating members. Since each piston 59 is reciprocated in the radial direction of the rotors 56 and 57 by 55 and the spring 61, it is configured as a so-called differential pump that sucks oil and pressurizes and discharges the oil. Further, when the suction or discharge of the oil into the oil chamber 62 is restricted, that is, when the discharge pressure is increased, the oil is sealed in the oil chamber 62, and the oil is substantially incompressible. Reciprocation is limited or prevented. As a result, the piston 59 and the oil function as a so-called wedge between the housing 54 and the rotors 56 and 57, and relative rotation between the housing 54 and the rotors 56 and 57 is prevented or restricted. That is, torque (or transmission torque capacity) transmitted between the housing 54 and the rotors 56 and 57 increases.

  Intermediate shafts 67 and 68 for transmitting torque from the oil pumps 52 and 53 to a gear mechanism to be described later are provided. That is, the rotor 56 in one oil pump (on the input shaft 51 side or the left side oil pump in FIG. 14) 52 has an intermediate shaft (hereinafter referred to as a first intermediate shaft) 67 extending along the rotation center axis. The first intermediate shaft 67 passes through the rotor 57 in the other oil pump 53 (the oil pump on the side opposite to the input shaft 51 or the right side in FIG. 14) 53, and the input shaft 51 extends in the opposite direction to the input shaft 51 on the same axis line as 51 and protrudes outside the housing 54. Another intermediate shaft (hereinafter referred to as a second intermediate shaft) 68 is fitted on the outer peripheral side of the first intermediate shaft 67 so as to be relatively rotatable. That is, the second intermediate shaft 68 is a hollow shaft, and is attached so as to be integrated with the central portion of one side surface of the rotor 57 in the oil pump 53. As with the first intermediate shaft 67, the second intermediate shaft 68 extends on the same axis as the input shaft 51 and in the opposite direction to the input shaft 51, and protrudes outside the housing 54.

  These intermediate shafts 67 and 68 are rotatably supported by a fixed portion or casing portion such as a casing (not shown) of the entire transmission, and are in close contact with the fixed portion or the casing portion at or near the supporting portion. It is mated. The suction hole 63 and the discharge hole 64 extend through the inside of each of the intermediate shafts 67 and 68 to a fixed fitting portion or a close fitting portion with the casing portion, and from here to the fixing portion or the casing portion. It communicates with a formed oil passage (not shown).

  Here, a hydraulic circuit for controlling the discharge state of oil from the oil pumps 52 and 53 will be described. As shown in FIG. 16, the discharge holes 64 of the oil pumps 52 and 53 have respective oil pumps. Discharge oil passages 69 and 70 are provided so as to communicate with each other 52 and 53, and the discharge oil passages 69 and 70 are connected to control valves 71 and 72 respectively provided corresponding thereto. These control valves 71, 72 include a valve body (not shown) and return springs 73, 74 that press the valve body in the valve opening direction, and the inflow side hydraulic pressure is from the same direction as the return springs 73, 74. It is configured to act on the valve body, and control force (or control oil pressure) 75, 76 is applied to the valve body in the valve closing direction.

  The control forces 75 and 76 are not particularly shown, but may be an electromagnetic force generated by a solenoid, a hydraulic pressure controlled by a solenoid valve, a spring force that can be changed by a cam, or the like, preferably electrically controlled. Possible pressing force. An electronic control unit (ECU) 77 for controlling the control forces 75 and 76 is provided. This electronic control unit 77 is mainly composed of a microcomputer, and is configured to perform calculations according to input data, prestored data and programs, and output a predetermined control signal according to the result of the calculation. Has been.

  On the other hand, the suction holes 63 of the oil pumps 52 and 53 are provided with suction oil passages 78 and 79 communicating with the oil pumps 52 and 53, respectively. It communicates with an oil reservoir 80 such as an oil pan. Further, ports (not shown) on the discharge side of the control valves 71 and 72 are communicated with the corresponding intake oil passages 78 and 79, respectively. Note that a temperature sensor 81 for detecting the temperature of the oil is disposed in the oil reservoir 80, and the detection signal is input to the electronic control unit 77.

  In addition, when making each control valve 71 and 72 into the structure which can be electrically controlled, it is preferable to set it as the valve of what is called a normal open structure which will be in a fully open state in an electrical OFF state. When an electrical failure such as disconnection occurs, the oil pump 52, 53 is automatically released and the discharge flow rate of each oil pump 52, 53 is not limited. As a result, each oil pump 52, 53 transmits torque. It is because it can disappear and fail safe can be established.

  The transmission according to the present invention is configured to output the torque transmitted from the input shaft 51 via the oil pumps 52 and 53 via the gear mechanism. An example of the gear mechanism will be described. As shown in FIG. 13, an output shaft 82 and a secondary shaft 83 are arranged in parallel with the intermediate shafts 67 and 68. The first intermediate shaft 67 protrudes from the distal end side (right side in FIG. 13) of the second intermediate shaft 68, and the first speed gear pair 84 is sequentially arranged from the distal end side to the proximal end side of the first intermediate shaft 67. A third speed gear pair 85 and a fifth speed gear pair 86 are provided. A second speed gear pair 87 and a fourth speed gear pair 88 are provided in order from the distal end side to the proximal end side of the second intermediate shaft 68. The first speed gear pair 84, the second speed gear pair 87, the third speed gear pair 85, the fourth speed gear pair 88, and the fifth speed gear pair 86 have a gear ratio that decreases in the order listed here. It is configured.

  More specifically, the first speed drive gear 84A and the third speed drive gear 85A are attached to the first intermediate shaft 67 adjacent to each other, and are engaged with the first speed drive gear 84A. The speed driven gear 84B and the third speed driven gear 85B meshing with the third speed drive gear 85A are rotatably fitted to the output shaft 82 in a state where they are adjacent to each other. A clutch mechanism for selectively connecting the driven gears 84B and 85B to the output shaft 82 is disposed between the driven gears 84B and 85B. As an example, this clutch mechanism is configured by a conventionally known synchronous coupling mechanism (synchronizer) 89, and a sleeve that rotates together with the output shaft 82 is moved in the axial direction so that any one of the driven gears 84B and 85B is moved. The driven gears 84B and 85B are selectively connected to the output shaft 82 by spline fitting.

  A fifth speed drive gear 86A is attached to the first intermediate shaft 67 adjacent to the third speed drive gear 85A, and a fifth speed driven gear 86B meshed with the fifth speed drive gear 86A is provided. The counter shaft 83 is rotatably fitted and held.

  Further, a second speed drive gear 87A and a fourth speed drive gear 88A are attached in order from the tip end side of the second intermediate shaft 68 located on the outer peripheral side of the first intermediate shaft 67, and the second speed drive gear 88A is attached. The second speed driven gear 87B meshed with the speed drive gear 87A and the fourth speed driven gear 88B meshed with the fourth speed drive gear 88A are rotatably fitted to the output shaft 82 in a state of being adjacent to each other. ing. A clutch mechanism for selectively connecting the driven gears 87B and 88B to the output shaft 82 is disposed between the driven gears 87B and 88B. As an example, this clutch mechanism is configured by a conventionally known synchronous coupling mechanism (synchronizer) 90, and a sleeve that rotates together with the output shaft 82 is moved in the axial direction so that any one of the driven gears 87B and 88B is moved. The driven gears 87B and 88B are selectively connected to the output shaft 82 by spline fitting.

  An idle gear 91 meshed with the second speed drive gear 87A is disposed on the outer peripheral side of the second speed drive gear 87A, and a reverse driven gear 92B meshed with the idle gear 91 is attached to the countershaft 83. It is supported by being rotatably fitted. Accordingly, the reverse driven gear 92B and the fifth speed driven gear 86B are adjacent to each other on the auxiliary shaft 83, and a clutch mechanism for selectively connecting the driven gears 86B and 92B to the auxiliary shaft 83 is provided. , Between the driven gears 86B and 92B. As an example, this clutch mechanism is configured by a conventionally known synchronous coupling mechanism (synchronizer) 93, and a sleeve that rotates together with the auxiliary shaft 83 is moved in the axial direction so that any one of the driven gears 86B and 92B is moved. The driven gears 86B and 92B are selectively connected to the output shaft 82 by spline fitting. Accordingly, the second speed drive gear 87A has a function of a drive gear in a gear pair for setting the reverse gear (reverse gear).

  A transmission mechanism for transmitting power between the auxiliary shaft 83 and the output shaft 82 is provided. As this transmission mechanism, a gear mechanism, a winding transmission mechanism, or the like can be employed as necessary. In the example shown in FIG. 13, a gear mechanism using an idle gear 94 is employed. The gear ratio in the gear mechanism is set to “1”, so that acceleration / deceleration does not occur between the auxiliary shaft 83 and the output shaft 82.

  Each of the synchronous coupling mechanisms 89, 90, 93 (hereinafter referred to as first sync 89, second sync 90, and third sync 93) corresponds to the switching mechanism of the present invention, and moves the sleeve to the left or right. As a result, one of the driven gears is connected to the output shaft 82 or the auxiliary shaft 83, and in a state where the sleeve is located at the center, the connection is released to be neutral. Although such a movement of the sleeve can be directly performed by a manual operation, it is preferable that the sleeve is operated by an electric actuator or a hydraulic actuator. This is because by operating this type of actuator in accordance with a control signal from the electronic control unit 77, electrical shift control can be performed.

  Therefore, even in the transmission configured as shown in FIG. 13, when starting at the first speed from the stopped state, both the oil pumps 52 and 53 function as pumps to output shaft torque, and If the first and second synchros 89 and 90 are operated so as to set the first speed and the second speed, their shaft torques are output to the output shaft via the first speed gear pair 84 and the second speed gear pair 87. Therefore, a large driving torque can be obtained when starting. The contents of the control for setting other speed ratios and the control between the fixed speed ratios are the same as those in the above-described example, and are collectively shown in FIG. The meaning of each symbol in FIG. 17 is the same as that in FIG. 4 described above. Therefore, even with the configuration shown in FIG. 13, a so-called continuously variable transmission in which the gear ratio is continuously changed is possible, and when a predetermined fixed gear ratio is set, the oil pump is set to the “LOCK” state and the oil is supplied. Since it is confined, power consumption can be improved without consuming any power.

  In addition, the present invention is not limited to the configuration using two fluid pressure pump motors, and may be a configuration of three or more so-called multi-axis types. In such a configuration, there is a fluid pressure pump motor that is not involved in setting the fixed gear ratio and shifting between the fixed gear ratio and another fixed gear ratio. The driving torque can be increased. Further, in the present invention, the shaft torque of the fluid pressure pump motor can be controlled by the discharge pressure. Therefore, in the present invention, a fixed displacement type fluid pressure pump motor can be used instead of the variable displacement type fluid pressure pump motor. it can. Further, the transmission mechanism in the present invention is not limited to a gear pair, and may be a winding transmission mechanism such as a chain or a belt, or a friction transmission mechanism capable of continuously changing a gear ratio. There may be.

It is a skeleton figure which shows an example of this invention typically. FIG. 3 is a hydraulic circuit diagram schematically showing an example of the hydraulic control device. FIG. 6 is a hydraulic circuit diagram schematically showing another example of the hydraulic control device. FIG. 2 is a chart collectively showing an operation state of the transmission shown in FIG. 1. FIG. It is a collinear diagram about each planetary gear mechanism in the stop state which creep torque does not generate | occur | produce. It is a collinear diagram about each planetary gear mechanism in the stop state which creep torque generate | occur | produces. It is a collinear diagram about each planetary gear mechanism in the process from the start to the first speed. It is a collinear diagram about each planetary gear mechanism in the 1st speed in the upshift waiting state to the 2nd speed. It is a collinear diagram about each planetary gear mechanism in the 1st speed in the state which set the sleeve of the 2nd synchro to the neutral position. It is a collinear diagram about each planetary gear mechanism in the speed change process from 1st speed to 2nd speed. It is a time chart for demonstrating the example which controls the axial torque of each pump motor by changing discharge volume by making discharge pressure constant at the time of start. It is a time chart for demonstrating the example which controls the axial torque of each pump motor by discharge pressure, making extrusion volume constant at the time of start. It is a skeleton figure which shows the other example of this invention typically. It is a schematic sectional view along the central axis showing the configuration of the oil pump. FIG. 3 is a schematic cross-sectional view along a plane perpendicular to the central axis showing one configuration of the oil pump. It is a block diagram which shows typically the hydraulic circuit about the oil pump. FIG. 14 is a table collectively showing the operating state of the transmission shown in FIG. 13.

Explanation of symbols

  DESCRIPTION OF SYMBOLS 1 ... Power source, 2 ... Input member 3, 4 ... Planetary gear mechanism, 8 ... 1st intermediate shaft, 9 ... 1st rotating shaft, 10 ... 2nd intermediate shaft, 11 ... 2nd rotating shaft, 12 ... 1st Pump motor, 13 ... 2nd pump motor, 17, 77 ... Electronic control unit (ECU), 18, 82 ... Output shaft, 19, 86 ... 5th speed gear pair, 20, 85 ... 3rd speed gear pair, 21, 84 ... 1st speed gear pair, 22, 88 ... 4th speed gear pair, 23, 87 ... 2nd speed gear pair, 25, 26, 27 ... Synchronous coupling mechanism (synchro), 52, 53 ... Oil pump, 67, 68: Intermediate shaft.

Claims (4)

At least two fluid pressure pumps whose shaft torque is set according to at least one of the extrusion volume and the discharge pressure, and the torque output from the power source to the output member according to the shaft torque of the first fluid pressure pump A first transmission mechanism for transmission, and a second transmission for transmitting torque output from the power source to the output member in accordance with the shaft torque of the second fluid pressure pump and at a speed ratio different from that of the first transmission mechanism. In a vehicle transmission having a transmission mechanism,
And an acceleration control means for setting the shaft torque of each fluid pressure pump to a predetermined value or more and transmitting the torque to the output member via each transmission mechanism when an increase in torque output from the output member is requested. A vehicle transmission characterized by the above. Each of the fluid pressure pumps includes a variable displacement fluid pressure pump having a variable extrusion volume,
2. The vehicle transmission according to claim 1, wherein the acceleration control means includes means for changing a push-out volume of the variable displacement fluid pressure pump.   The vehicle transmission according to claim 1 or 2, wherein the acceleration control means includes means for controlling a discharge pressure of each fluid pressure pump.   4. The apparatus according to claim 3, further comprising means for fixing a displacement volume of the variable displacement fluid pressure pump to a maximum when the acceleration control means controls a discharge pressure of each fluid pressure pump. Vehicle transmission.

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