JP5373922B2 - Oil pump drive - Google Patents

Abstract

The disclosed oil-pump-driving device (2) is provided with: a first transmission pathway (25) that transmits the rotation of a drive shaft (24) to a driven shaft (22) at a first speed ratio; a second transmission pathway (23) that transmits the rotation of the drive shaft (24) to the driven shaft (22) at a second speed ratio that is greater than the first speed ratio; and a hydraulic engagement mechanism (25c) that switches between the first transmission pathway (25) and the second transmission pathway (23). An oil pathway (L81) that is for the engagement mechanism and that supplies hydraulic pressure to the hydraulic engagement mechanism (25c) is provided to a lock-up shift valve (94). Hydraulic pressure is supplied to the oil pathway (L81) for the engagement mechanism via the first transmission pathway (25) when the lock-up shift valve (94) is in a first shift state, and via the second transmission pathway (23) when the lock-up shift valve (94) is in a second shift state, in a manner such that the rotation of the drive shaft (24) is transmitted to the driven shaft (22).

Description

  The present invention relates to an oil pump drive device that drives an oil pump using the power of a drive source that drives drive wheels of a vehicle.

  2. Description of the Related Art Conventionally, an oil pump driving device that drives an oil pump that supplies lubricating oil to a friction engagement mechanism such as a hydraulic clutch or a hydraulic brake provided in an automatic transmission by using a driving force of an internal combustion engine that is a driving source of the vehicle. Is known (for example, see Japanese Patent Application Laid-Open No. 2008-121724).

  Here, the amount of lubricating oil required for the friction engagement mechanism is substantially constant regardless of the rotational speed of the drive source. However, when the oil pump is driven using the driving force of the driving source, the amount of lubricant discharged by the oil pump is proportional to the rotational speed of the driving source.

  Here, the capacity of the oil pump needs to be set so that a sufficient discharge amount of the lubricating oil can be ensured even when the drive source has a low rotational speed, for example, when the drive source is in an idling state. However, if the capacity of the oil pump is set large, when the rotational speed of the drive source is high, the discharge amount of the oil pump increases more than necessary, which may result in a drive loss of the oil pump and decrease fuel consumption. .

  In order to solve this problem, the oil pump driving device disclosed in Japanese Patent Application Laid-Open No. 2008-121724 described above has a driving gear and a driven gear, and has a gear ratio (number of teeth of the driven gear / tooth of the driving gear). Two gear trains having different numbers) and a switching mechanism for selectively switching the two gear trains are provided. When the internal combustion engine has a high rotational speed, a gear train having a large gear ratio is selected by the switching mechanism. The engine speed is reduced and transmitted to the oil pump through the gear train having a large gear ratio. Thereby, even if the rotation speed of an internal combustion engine is high, the discharge amount of an oil pump can be suppressed.

  The conventional oil pump drive device needs to be provided with a plurality of gear trains and a switching mechanism. Furthermore, it is necessary to provide a dedicated control unit for switching the switching mechanism according to the number of rotations of the drive source, and also to provide a dedicated actuator for operating the switching mechanism, which increases the size of the oil pump drive device. is there.

  In view of the above points, the present invention provides an oil pump drive device that can switch a switching mechanism without providing a dedicated control unit, and that can be reduced in size by reducing the number of components. With the goal.

  The present invention relates to an oil pump drive device that drives an oil pump that supplies lubricating oil to a transmission by using power of a drive source that drives drive wheels of a vehicle. In the vehicle, the power of the drive source is locked. The lockup clutch is transmitted to a transmission via a torque converter having an upclutch, and the lockup clutch is permitted to be engaged by setting a lockup shift valve provided in a hydraulic circuit to a first shift state. Fastening is prohibited by setting the lock-up shift valve to the second shift state, and a drive shaft that is rotated by transmission of power from the drive source, a driven shaft connected to the oil pump, and the drive A first transmission path for defining the value obtained by dividing the rotational speed of the shaft by the rotational speed of the driven shaft as a speed ratio, and transmitting the rotation of the drive shaft to the driven shaft at a predetermined first speed ratio. A second transmission path for transmitting to the driven shaft at a second speed ratio larger than the first speed ratio, and a hydraulic engagement mechanism for switching between the first transmission path and the second transmission path, The lockup shift valve is provided with an oil passage for an engagement mechanism that supplies hydraulic pressure to the hydraulic engagement mechanism. When the lockup shift valve is in the first shift state, the rotation of the drive shaft is the first. When the lock-up shift valve is in the second shift state, the rotation of the drive shaft is transmitted to the driven shaft via the second transmission path. In addition, hydraulic pressure is supplied to the oil passage for the engagement mechanism.

  The lock-up clutch is engaged when the driving source and the input shaft of the transmission are directly connected to improve fuel efficiency, and conversely, is released when the driving force is amplified by the torque converter. When the lockup clutch is engaged, the rotational speed of the drive source is generally low.

  In the present invention, when the lockup shift valve is in the first shift state in which the lockup clutch is permitted to be engaged, the rotation of the drive shaft is transmitted to the driven shaft via the first transmission path. Conversely, when the lock-up shift valve is in the second shift state in which the lock-up clutch is prohibited, the rotation of the drive shaft is transmitted to the driven shaft via the second transmission path.

  That is, in the oil pump drive device of the present invention, the first transmission path and the second transmission path are switched in conjunction with switching of the lockup shift valve that switches between engagement and disengagement of the lockup clutch. For this reason, it is not necessary to provide a dedicated control unit for the hydraulic engagement mechanism as the switching mechanism, and the oil pump drive device can be easily controlled.

  In addition, by providing an oil passage for the engagement mechanism in the lock-up shift valve, the lock-up shift valve can have a function as a shift valve that switches supply / cutoff of hydraulic pressure to the hydraulic engagement mechanism. There is no need to separately provide a shift valve dedicated to the type engagement mechanism in the hydraulic circuit. Therefore, it is possible to reduce the size of the oil pump drive device.

  Furthermore, since the hydraulic engagement mechanism is switched to the first transmission path or the second transmission path in conjunction with the engagement / release of the lock-up clutch, even when the drive source is rotating at low speed, The pump can be rotated at a high speed, and the oil pump can be downsized.

  As a first specific aspect of the present invention, the first transmission path is engaged with the first shaft support gear supported on the drive shaft or the driven shaft, and meshed with the first shaft support gear and on the driven shaft or the drive shaft. The second transmission path is configured by a first gear train including a fixed first fixed gear, and the second transmission path meshes with the second shaft support gear supported by the drive shaft or the driven shaft. In addition, it can be configured by a second gear train including a second fixed gear fixed to the driven shaft or the drive shaft.

  In the first specific aspect of the present invention, there is provided one of the first support gear and the second support gear, and the drive shaft or the driven shaft that supports the one support gear. A one-way clutch is provided between the first and second shaft-supporting gears, and the drive shaft or driven shaft that supports the other shaft-supporting gear. Are preferably configured to be connectable.

  According to this, by making the hydraulic engagement mechanism in a connected state, one of the support gears sandwiching the one-way clutch and the shaft supporting the gear can be rotated relative to each other, and the hydraulic engagement mechanism is opened. By doing so, it is possible to configure such that one of the shaft support gears sandwiching the one-way clutch and the shaft that supports the gears are integrally rotated. As a result, it is not necessary to arrange an oil passage or the like on one of the shaft support gears, compared to the case where a hydraulic engagement mechanism is provided instead of the one-way clutch, and the configuration of the oil pump drive device is simplified and miniaturized. Can be achieved.

  As a second specific embodiment of the present invention, a pump driving planetary gear mechanism including three elements including a sun gear, a carrier, and a ring gear is provided, and the three elements of the pump driving planetary gear mechanism are combined with the three elements. In the collinear chart in which the ratio of the relative rotational speeds can be represented by a straight line, the first element, the second element, and the third element from one side in the order of arrangement at intervals corresponding to the gear ratio of the planetary gear mechanism for driving the pump As a hydraulic coupling mechanism, a fixed state in which the driven shaft is connected to the first element or the second element, the drive shaft is connected to the second element or the first element, and the third element is fixed to the housing. And a pump driving brake that can be switched to an open state in which the fixing is cut off, a connected state in which any two of the first to third elements are connected, and an open state in which the connection is cut off. For pump drive By providing a latch, the pump drive brake is fixed, and the pump drive clutch is released, so that the rotation of the drive shaft is driven via one of the first transmission path and the second transmission path. When the pump drive brake is released and the pump drive clutch is connected, the rotation of the drive shaft is driven via the other transmission path of the first transmission path and the second transmission path. It can also be configured to be transmitted to the shaft.

  Also by this, the controllability and size reduction of the oil pump drive device can be achieved. Further, if the driven shaft is connected to the first element and the drive shaft is connected to the second element, the rotation of the drive shaft is changed to the first state by setting the pump drive brake to the fixed state and the pump drive clutch to the open state. The rotation of the drive shaft is transmitted to the driven shaft via the second transmission path by being transmitted to the driven shaft via the transmission path, and the pump drive brake being released and the pump drive clutch being connected. It becomes. That is, the rotation of the drive shaft can be increased and transmitted to the driven shaft. Therefore, even if the capacity of the oil pump is reduced, a sufficient discharge amount can be ensured, and the oil pump drive device can be further reduced in size.

The skeleton figure which shows the automatic transmission to which the oil pump drive device of 1st Embodiment of this invention is applied. 1 is a hydraulic circuit diagram of an automatic transmission according to a first embodiment. The alignment chart of the double planetary gear mechanism of 1st Embodiment. Explanatory drawing which shows the state of the hydraulic pressure regulation valve and shift valve in each gear stage of the automatic transmission of 1st Embodiment. The elements on larger scale of the hydraulic circuit of 1st Embodiment. The elements on larger scale of the hydraulic circuit of 1st Embodiment. The skeleton figure which shows the automatic transmission to which the oil pump drive device of 2nd Embodiment of this invention is applied. The alignment chart of the planetary gear mechanism of 2nd Embodiment.

[First Embodiment]
FIG. 1 shows an automatic transmission 1 to which an oil pump drive device 2 according to the first embodiment is applied. The automatic transmission 1 is provided with a torque converter 11 to which the driving force of the engine 3 as a driving source for driving the driving wheels of the vehicle is transmitted.

  The torque converter 11 includes a pump impeller PO to which driving force is transmitted from a flywheel 5 provided on the crankshaft 4 of the engine 3 via a damper 6 for absorbing rotational vibration, and power from the pump impeller PO via an internal fluid. A turbine ST to be transmitted, a stator ST that is disposed between the pump impeller PO and the turbine TB and is connected to a transmission case 13 as a housing via a one-way clutch 12, a turbine TB and a pump impeller PO. And a lockup clutch LC that can be freely connected.

  The automatic transmission 1 includes an input shaft 14 that is rotatably supported in a transmission case 13 as a housing and connected to the turbine TB of the torque converter 11, and an output gear 15 that is disposed concentrically with the input shaft 14. And. The rotation of the output gear 15 is transmitted to the left and right drive wheels of the vehicle via a differential gear and a propeller shaft (not shown). In FIG. 1, only the upper half of the automatic transmission 1 is mainly shown. A single planetary gear mechanism 16 for input and a double planetary gear mechanism 17 for shifting are arranged in the transmission case 13 around the input shaft 14.

  The single planetary gear mechanism 16 for input is a single pinion type planetary gear mechanism that includes a sun gear Sa, a ring gear Ra, and a carrier Ca that rotatably and revolves a pinion Pa that meshes with the sun gear Sa and the ring gear Ra. Composed. The carrier Ca of the single planetary gear mechanism 16 is connected to the input shaft 14, and the sun gear Sa is fixed to the transmission case 13. Therefore, the input element of the single planetary gear mechanism 16 is the carrier Ca, the fixed element is the sun gear Sa, and the output element is the ring gear Ra.

  When the gear ratio of the single planetary gear mechanism 16 (the number of teeth of the ring gear / the number of teeth of the sun gear) is k, the rotational speed of the ring gear Ra as an output element, that is, the output speed of the single planetary gear mechanism 16 is the rotational speed of the input shaft 14. Is “1”, and becomes (k + 1) / k, and the rotation of the input shaft 14 is accelerated and output. Hereinafter, (k + 1) / k is defined as “N1” for convenience of explanation.

  The double planetary gear mechanism 17 includes a pair of pinions that mesh with the first sun gear Sb1, the second sun gear Sb2, and the ring gear Rb, one meshing with the first sun gear Sb1, and the other meshing with the second sun gear Sb2 and the ring gear Rb. It is composed of a Ravigneaux type planetary gear mechanism comprising a carrier Cb that pivotally supports Pb and Pb ′ so as to rotate and revolve.

  FIG. 3 is a collinear diagram that can represent the ratio of the relative rotational speeds of the four elements Sb1, Sb2, Cb, and Rb of the double planetary gear mechanism 17 as a straight line. As indicated by vertical lines in FIG. 3, the four elements Sb1, Sb2, Cb, and Rb can be arranged in order of the second sun gear Sb2, the carrier Cb, the ring gear Rb, and the first sun gear Sb1 from the left side.

  The gear ratio between the ring gear Rb and the first sun gear Sb1 (number of teeth of the ring gear Rb / the number of teeth of the first sun gear Sb1) is m, and the gear ratio of the ring gear Rb and the second sun gear Sb2 (number of teeth of the ring gear Rb / first sun gear Sb2). When the number of teeth) is n, the interval between the second sun gear Sb2, the carrier Cb, the ring gear Rb, and the first sun gear Sb1 is a ratio of n: 1: m-1.

  In the alignment chart of FIG. 3, the lower horizontal line indicates that the rotational speed is “0”, and the upper horizontal line indicates that the rotational speed is “1” and the rotational speed of the input shaft is “1”. ".

  In the automatic transmission 1 according to the first embodiment, the first sun gear Sb1 of the double planetary gear mechanism 17 and the carrier Ca, which is the input element of the single planetary gear mechanism 16, can be released via the input shaft 14 as engaging elements. A second clutch C2 that releasably connects a carrier Cb of the double planetary gear mechanism 17 and a ring gear Ra that is an output element of the single planetary gear mechanism 16, and a second clutch C2 that is connected to the first planetary gear mechanism 17. A third clutch C3 that releasably connects the sun gear Sb2 and the carrier Ca that is an input element of the single planetary gear mechanism 16 via the input shaft 14 is provided.

  Further, a first brake B1 for releasably fixing the second sun gear Sb2 to the transmission case 13 and a second brake B2 for releasably fixing the carrier Cb to the transmission case 13 are provided.

  The automatic transmission 1 is supplied with working oil pressure to each of the engagement elements C1 to C3, B1, and B2, and is also supplied with working oil (lubricating oil) to lubricate a lubrication portion that requires lubrication of the automatic transmission 1. An oil pump 21 is provided. Further, the automatic transmission 1 is provided with an oil pump driving device 2 that drives the oil pump 21 by using the driving force of the engine 3.

  The oil pump drive device 2 includes a drive shaft 24 disposed in parallel with the input shaft 14, a driven shaft 22 disposed in parallel with the drive shaft 24, and a first and second drive gear and a driven gear. These two gear trains 25, 23, a one-way clutch 23o, and a hydraulic engagement mechanism 25c are provided.

  The rotation of the pump impeller PO is transmitted to the drive shaft 24 via the idle gear train 31. The idle gear train 31 is engaged with the first idle gear 32 connected to the pump impeller PO, the second idle gear 33 meshed with the first idle gear 32, meshed with the second idle gear 33 and fixed to the drive shaft 24. And a third idle gear 34.

  The first drive gear 25 b of the first gear train 25 is rotatably supported on the drive shaft 24. That is, the first drive gear 25b corresponds to the first shaft support gear of the present invention. The first driven gear 25 a of the first gear train 25 that meshes with the first drive gear 25 b is fixed to the driven shaft 22. That is, the first driven gear 25a corresponds to the first fixed gear of the present invention. The second drive gear 23 b of the second gear train 23 is rotatably supported on the drive shaft 24. That is, the second drive gear 23b corresponds to the second shaft support gear of the present invention. The second driven gear 23 a of the second gear train 23 that meshes with the second drive gear 23 b is fixed to the driven shaft 22. That is, the second driven gear 23a corresponds to the second fixed gear of the present invention.

  The hydraulic engagement mechanism 25c is composed of a wet multi-plate clutch, and is configured to be switchable between a fixed state in which the first drive gear 25b is fixed to the drive shaft 24 and an open state in which this fixation is released. When the rotational speed of the drive shaft 24 exceeds the rotational speed of the second drive gear 23b, the one-way clutch 23o is in a connected state in which the drive shaft 24 and the second drive gear 23b are coupled. When the speed is lower than the rotational speed of the second drive gear 23b, the connection between the drive shaft 24 and the second drive gear 23b is cut off and an open state is established.

  Assuming that the gear ratio of the first gear train 25 (the number of teeth of the driven gear / the number of teeth of the drive gear) is j and the gear ratio of the second gear train 23 is i, in this embodiment, the gear ratio i of the second gear train 23. Is set to “1”, and the gear ratio j of the first gear train 25 is set to be smaller than “1”. That is, when the rotation of the drive shaft 24 is transmitted to the driven shaft 22 via the second gear train 23, the driven shaft 22 rotates at “1” (1 / i) at the same speed as the drive shaft 24, When the rotation of the drive shaft 24 is transmitted to the driven shaft 22 via the first gear train 25, the rotation of the drive shaft 24 is increased and the rotational speed of the driven shaft 22 becomes “1 / j”. In the present embodiment, the gear ratio j of the first gear train 25 corresponds to a predetermined first speed ratio, and the gear ratio i of the second gear train 23 corresponds to a second speed ratio. The “speed ratio” is defined as a value obtained by dividing the rotational speed of the drive shaft 24 by the rotational speed of the driven shaft 22. The gear ratio i is not limited to “1”. What is necessary is just to set so that the gear ratio i and the gear ratio j may satisfy | fill the relationship of i / j> = 1.

  Next, with reference to FIG. 3, the case where each gear stage of the automatic transmission 1 of this embodiment is established is demonstrated. When establishing the first gear, the first clutch C1 is in a connected state, and the second brake B2 is in a fixed state. By setting the first clutch C1 to the connected state, the rotational speed of the first sun gear Sb1 becomes “1”. Further, by setting the second brake B2 in a fixed state, the rotational speed of the carrier Cb becomes “0”. The ring gear Rb to which the output gear 15 is connected rotates at a speed of 1 / m (“1st” in FIG. 3), and the first gear is established.

  When establishing the second gear, the first clutch C1 is in a connected state and the first brake B1 is in a fixed state. By setting the first clutch C1 to the connected state, the rotational speed of the first sun gear Sb1 becomes “1”. Further, by setting the first brake B1 in a fixed state, the rotation speed of the second sun gear Sb2 becomes “0”. The ring gear Rb to which the output gear 15 is coupled rotates at a speed (n + 1) / (m + n) (“2nd” in FIG. 3), and the second gear is established.

  When establishing the third speed, the first clutch C1 and the third clutch C3 are brought into a connected state. When the first clutch C1 and the third clutch C3 are in the connected state, the first sun gear Sb1 and the second sun gear Sb2 are connected via the input shaft 14 and rotate at “1” that is the same speed. As a result, the four elements Sb1, Sb2, Cb, and Rb of the double planetary gear mechanism 17 are locked so that they cannot rotate relative to each other, and the rotational speeds of all the elements Sb1, Sb2, Cb, and Rb become “1”. Accordingly, the ring gear Rb to which the output gear 15 is connected rotates at “1” (“3rd” in FIG. 3), and the third gear is established.

  In order to establish the fourth speed, the first clutch C1 and the second clutch C2 are connected. By setting the first clutch C1 to the connected state, the rotational speed of the first sun gear Sb1 becomes “1”. Further, by setting the second clutch C2 in the connected state, the rotation speed of the carrier Cb rotates at the output speed “N1” of the single planetary gear mechanism 16. Then, the ring gear Rb to which the output gear 15 is coupled rotates at a speed of {N1 (m−1) +1} / m (“4th” in FIG. 3), and the fourth speed stage is established.

  When establishing the fifth gear, the second clutch C2 and the third clutch C3 are brought into a connected state. By bringing the second clutch C2 into the connected state, the rotation speed of the carrier Cb rotates at the output speed “N1” of the single planetary gear mechanism 16. In addition, the rotational speed of the second sun gear Sb2 becomes “1” by bringing the third clutch C3 into the connected state. The ring gear Rb to which the output gear 15 is connected rotates at a speed of {N1 (n + 1) -1} / n ("5th" in FIG. 3), and the fifth gear is established.

  When establishing the sixth speed, the second clutch C2 is set in the connected state, and the first brake B1 is set in the fixed state. By bringing the second clutch C2 into the connected state, the rotation speed of the carrier Cb rotates at the output speed “N1” of the single planetary gear mechanism 16. Further, by setting the first brake B1 in a fixed state, the rotation speed of the second sun gear Sb2 becomes “0”. Then, the ring gear Rb to which the output gear 15 is connected rotates at a speed of {N1 (n + 1)} / n (“6th” in FIG. 3), and the sixth gear is established.

  In order to establish the reverse gear, the third clutch C3 is engaged and the second brake B2 is fixed. By setting the third clutch C3 in the connected state, the rotational speed of the second sun gear Sb2 becomes “1”. Further, by setting the second brake B2 in a fixed state, the rotational speed of the carrier Cb becomes “0”. The ring gear Rb to which the output gear 15 is connected rotates at a speed of “−1 / n” (“Rev” in FIG. 3) in the reverse rotation direction (the rotation direction when the vehicle moves backward), and the reverse gear is established. The

  Next, a hydraulic circuit L that controls the automatic transmission 1 having the above-described configuration will be described with reference to FIG.

  The hydraulic circuit L is driven by the engine 3 and discharges hydraulic oil from the oil tank T to the oil passage L21, a regulator valve 42 that regulates the line hydraulic pressure PL of the hydraulic circuit, and a select lever (not shown). , A manual valve 43 interlocked with the first, fifth hydraulic pressure regulating valves 81a, 81b, 81c, 81d, 81e comprising linear solenoid valves, and first to third open / close solenoid valves 82a, 82b. , 82 c, first to third three shift valves 91, 92, 93, a lockup shift valve 94, and a lockup control valve 95.

  The regulator valve 42 and the oil passage L21 are connected via an oil passage L21a. The manual valve 43 and the oil passage L21 are connected via an oil passage L21b. The three solenoid valves 82a, 82b, 82c, the second shift valve 92, the lockup control valve 95, the third hydraulic pressure adjustment valve 81c, and the fifth hydraulic pressure adjustment valve 81e are connected to the manual valve 43 via an oil passage L31. ing. The five hydraulic pressure adjusting valves 81a, 81b, 81c, 81d, 81e adjust the hydraulic pressure based on an instruction from a transmission control unit TCU (not shown).

  The first hydraulic pressure adjustment valve 81a, the second hydraulic pressure adjustment valve 81b, the fourth hydraulic pressure adjustment valve 81d, and the first to third shift valves 91, 92, 93 are connected to the manual valve 43 via an oil passage L32. ing. When the manual valve 43 is switched to “D”, which is the automatic shift position from the first gear to the sixth gear, by a select lever (not shown), the oil passage L21b is communicated with the oil passage L31 and the oil passage L32. The line oil pressure PL supplied from the path L21b is supplied to the oil paths L31 and L32.

  When energized, the first hydraulic pressure adjustment valve 81a adjusts the line hydraulic pressure PL supplied from the oil passage L32 to a predetermined hydraulic pressure instructed by the transmission control unit TCU, and locks it via the oil passage L41. Supply to the upshift valve 94. The first hydraulic pressure adjusting valve 81a shuts off the supply of hydraulic pressure to the lock-up shift valve 94 via the oil passage L41 when the energization is off.

  When the energization is on, the second hydraulic pressure adjusting valve 81b adjusts the line hydraulic pressure PL supplied from the oil passage L32 to a predetermined hydraulic pressure instructed by the transmission control unit TCU, and the second hydraulic pressure regulating valve 81b passes through the oil passage L42. The first shift valve 91 and the second shift valve 92 are supplied. The second hydraulic pressure regulating valve 81b blocks the supply of hydraulic pressure to the first shift valve 91 and the second shift valve 92 via the oil passage L42 when the energization is off.

  When the energization is on, the third hydraulic pressure adjusting valve 81c adjusts the line hydraulic pressure PL supplied from the oil passage L31 to a predetermined hydraulic pressure instructed by the transmission control unit TCU, and the third hydraulic pressure regulating valve 81c is connected via the oil passage L43. A three-shift valve 93 is supplied. The third hydraulic pressure adjusting valve 81c blocks the supply of hydraulic pressure to the third shift valve 93 via the oil passage L43 when the energization is off.

  When the energization is on, the fourth hydraulic pressure adjustment valve 81d adjusts the line hydraulic pressure PL supplied from the oil passage L32 to a predetermined hydraulic pressure instructed by the transmission control unit TCU, and passes through the oil passage L44. 1 brake is supplied to B1. When the energization is off, the fourth hydraulic pressure adjustment valve 81d blocks the supply of hydraulic pressure to the first brake B1 via the oil passage L44. The oil passage L44 is provided with a fifth accumulator Acc5 that suppresses pulsation of hydraulic pressure and a first brake switch SW5 that detects whether or not hydraulic pressure is supplied in parallel with the first brake B1. Yes.

  When the energization is on, the fifth hydraulic pressure regulating valve 81e adjusts the line hydraulic pressure PL supplied from the oil passage L31 to a predetermined hydraulic pressure instructed by the transmission control unit TCU, and the regulator via the oil passage L45. Supply to valve 42. The fifth hydraulic pressure regulating valve 81e shuts off the supply of hydraulic pressure to the regulator valve 42 via the oil passage L45 when energization is off.

  In addition, the transmission control unit TCU closes the fifth hydraulic pressure regulating valve 81e by turning off the energization at the time of in-gear, stall, and sudden acceleration accompanied by kick-down, and the line hydraulic pressure PL is controlled by the regulator valve 42. The low pressure region is set only due to the urging force of the spring 42b. Further, at the time of shifting, the transmission control unit TCU opens the fifth hydraulic pressure adjustment valve 81e so that the pressure can be freely adjusted by energization, and sets the line hydraulic pressure PL in a high pressure region.

  The first shift valve 91 is connected to the second shift valve 92 through two oil passages, an oil passage L51 and an oil passage L52. The first shift valve 91 is connected to the lockup shift valve 94 via an oil passage L53. A second accumulator Acc2 is connected to the oil passage L53 via a branching oil passage L53a. The first shift valve 91 is connected to the second brake B2 via the oil passage L71. The oil path L71 is provided with a second brake switch SW2 for detecting whether or not hydraulic pressure is supplied in parallel with the second brake B2.

  The first shift valve 91 has a plurality of annular grooves, and includes a spool 91a that is slidably inserted in the left-right direction in FIG. The spool 91a is urged to the right by a spring 91b.

  In this state, the oil pressure supplied from the second oil pressure adjusting valve 81b via the oil passage L42 is supplied to the second shift valve 92 via the annular groove of the spool 91a and the oil passage L51. L42 communicates with the oil passage L51. Further, the oil passage L32 and the oil passage L52 are communicated so that the line oil pressure PL supplied from the oil passage L32 is supplied to the second shift valve 92 via the annular groove of the spool 91a and the oil passage L52. Further, the oil passage L53 and the oil passage L71 are communicated.

  The first open / close solenoid valve 82a opens when energized, and supplies the line hydraulic pressure PL supplied from the oil passage L31 to the right end portion of the first shift valve 91 via the oil passage L61. As a result, the spool 91a of the first shift valve 91 moves to the left against the urging force of the spring 91b. In this state, the oil passage L52 and the oil passage L53 are communicated.

  The second shift valve 92 is connected to the first clutch C1 via the oil passage L72 and is connected to the second clutch C2 via the oil passage L73. The oil path L72 is provided with a first accumulator Acc1 that suppresses pulsation of hydraulic pressure and a first clutch switch SW1 that detects whether or not hydraulic pressure is supplied in parallel with the first clutch C1. The oil path L73 is provided with a third accumulator Acc3 that suppresses pulsation of hydraulic pressure and a second clutch switch SW3 that detects whether or not hydraulic pressure is supplied, in parallel with the second clutch C2.

  The second shift valve 92 includes a spool 92a that has a plurality of annular grooves and is slidably inserted in the left-right direction. The spool 92a is urged to the right by a spring 92b. In this state, the oil passage L51 and the oil passage L73 communicate with each other, and the oil passage L52 and the oil passage L72 communicate with each other.

  The second open / close solenoid valve 82b opens when energized, and supplies the line oil pressure PL supplied from the oil passage L31 to the right end portion of the second shift valve 92 via the oil passage L62. As a result, the spool 92a of the second shift valve 92 moves to the left against the urging force of the spring 92b. In this state, the oil passage L52 and the oil passage L73 communicate with each other, and the output hydraulic pressure of the second hydraulic pressure adjustment valve 81b supplied via the oil passage L42 is supplied to the first clutch C1 via the oil passage L72. As described above, the oil passage L42 and the oil passage L72 communicate with each other.

  The third shift valve 93 includes a spool 93a that has a plurality of annular grooves and is slidably inserted in the left-right direction. The spool 93a is biased to the right by a spring 93b. In this state, the oil path L43 and the oil path L74 are communicated so that the output hydraulic pressure of the third hydraulic pressure adjusting valve supplied via the oil path L43 is supplied to the third clutch C3 via the oil path L74. To do. The oil path 74 is provided with a fourth accumulator Acc4 for suppressing hydraulic pressure pulsation and a third clutch switch SW4 for detecting whether or not hydraulic pressure is supplied in parallel with the third clutch C3.

  In addition, when the second open / close solenoid valve 82b is opened when the second open / close solenoid valve 82b is energized, the third shift valve 93 is supplied with the line hydraulic pressure PL to the right end portion of the third shift valve 93 via the oil passage L62, and the spool 93a is spring-loaded. It moves to the left against the urging force of 93b.

  The lock-up shift valve 94 includes a spool 94a that has a plurality of annular grooves and is slidably inserted in the left-right direction. The spool 94a is urged to the right by a spring 94b. In this state, the oil passage L41 and the oil passage L53 communicate with each other, and the oil pressure adjusted by the first oil pressure adjusting valve 81a is supplied to the first shift valve via the oil passage L41, the annular groove of the spool 94a, and the oil passage L53. 91, and is supplied to the second accumulator Acc2 via an oil passage L53a branched from the oil passage L53.

  When the third open / close solenoid valve 82c is opened by energization, the line oil pressure PL is supplied to the right end portion of the lock-up shift valve 94 via the oil passage L63. As a result, the spool 94a of the lock-up shift valve 94 moves to the left against the urging force of the spring 94b. In this state, the oil passage L41 communicates with the oil passage L54 connecting the lockup shift valve 94 and the lockup control valve 95, and the hydraulic pressure output from the first hydraulic pressure adjustment valve 81a is Supplied to the right end.

  The lock-up control valve 95 includes a spool 95a that has a plurality of annular grooves and is slidably inserted in the left-right direction. The spool 95a is urged to the right by a spring 95b. The spool 95a of the lockup control valve 95 is moved to the left against the urging force of the spring 95a by the hydraulic pressure of the first hydraulic pressure regulating valve 81a supplied through the oil passage L41, the lockup shift valve 94, and the oil passage L54. Moving.

  In this state, the line hydraulic pressure PL supplied from the manual valve 43 via the oil passage L31 is supplied to the lockup clutch LC via the oil passage L75. The amount of movement of the lock-up control valve 95 to the left side of the spool 95a is controlled by adjusting the hydraulic pressure supplied by the first hydraulic pressure adjusting valve 81a. Thus, the lockup clutch LC is engaged by adjusting the hydraulic pressure supplied from the oil passage L75 to the lockup clutch LC and supplying a hydraulic pressure higher than the internal pressure of the torque converter 11. Conversely, if a hydraulic pressure lower than the internal pressure of the torque converter 11 is supplied from the oil passage L75, the lockup clutch LC is released without being engaged.

  That is, the state in which the spool 94a of the lock-up shift valve 94 is positioned on the right side by the biasing force of the spring 94b corresponds to the “second shift state” in which the lock-up clutch of the present invention is prohibited. The lock-up clutch of the present invention is permitted to be engaged when the spool 94a of the lock-up shift valve 94 is positioned on the left side against the urging force of the spring 94b from the opening of the third open / close solenoid valve 82c. This corresponds to the “first shift state”.

  Next, with reference to FIG. 4, the operation of the hydraulic pressure adjusting valves 81a to 81e, the open / close solenoid valves 82a to 82c, the shift valves 91 to 93, and the lock-up shift valve 94 at each shift speed will be described.

  In FIG. 4, “O” in the hydraulic pressure adjustment valve indicates a state in which the line hydraulic pressure PL supplied when energization is turned on is output as it is or adjusted and output, and “×” is supplied when energization is turned off. The state which interrupts | blocks line hydraulic pressure PL is shown. “-” Indicates a state in which the line hydraulic pressure PL is not supplied to the hydraulic pressure adjusting valve. In the open / close solenoid valve, “◯” indicates that the valve is opened when the power is turned on, and “×” indicates that the valve is closed when the power is turned off. “◯” in the shift valve indicates a state where the spool is moved to the left side, and “X” indicates a state where the spool is positioned on the right side.

  Further, “LC / ×” in the first hydraulic pressure regulating valve 81a indicates that “◯” and “x” in the state of the hydraulic pressure regulating valve are switched depending on whether or not the lockup clutch LC is engaged. . That is, “◯” indicates that the lockup clutch LC is engaged, and “X” indicates that the lockup clutch LC is not engaged.

  “O / X” in the third open / close solenoid valve 82c indicates that “O” and “X” are switched depending on whether or not the lockup clutch LC is engaged, and “O / X” in the lockup shift valve 94 is switched. “X” indicates “◯” when the lock-up clutch LC is engaged, and “X” when the lock-up clutch LC is not engaged, and indicates that the lock-up clutch LC is synchronized with the third open / close solenoid valve 82c. That is, the first hydraulic pressure regulating valve 81a, the third on-off solenoid valve 82c, and the lockup shift valve 94 are all “O” when the lockup clutch LC is engaged, and when the lockup clutch LC is not engaged. Becomes “x”.

  The first hydraulic pressure regulating valve 81a controls the hydraulic pressure applied to the first clutch C1 or the second brake B2 when the third open / close solenoid valve 82c and the lock-up shift valve 94 are “x”.

  Further, “●” of the second shift valve 92 (5th speed and 6th speed) indicates a case where the state where the spool is moved to the left side is maintained by the action of the self-lock function described later.

  Note that the details of the engagement of the lock-up clutch LC of the present embodiment are the same as those of the prior art, and the description thereof is omitted. The lock-up control valve 95 is also not described because the spool 95a moves depending on the engagement state of the lock-up clutch LC.

  When in-gear when the spool 7a of the manual valve 43 is switched from "N", which is the neutral position, to "D", which is the automatic transmission position from the first gear to the sixth gear, by the select lever, the manual valve 43 is the hydraulic circuit pump. The line oil pressure PL supplied from P through the oil passages L21 and L21b is in a state of being supplied to the oil passage L31 and the oil passage L32. Also, the first open / close solenoid valve 82a opens when energized, and the spool 91a of the first shift valve 91 moves to the left against the urging force of the spring 91b.

  Further, when the first hydraulic pressure adjusting valve 81a is energized, the hydraulic pressure supplied from the oil passage L32 is adjusted and output to the oil passage L41. The oil pressure output from the first oil pressure adjusting valve 81a is passed through the oil passage L41, the lock-up shift valve 94, the oil passage L53, the first shift valve 91, the oil passage L52, the second shift valve 92, and the oil passage L72. , Supplied to the first clutch C1 and the first accumulator Acc1. Further, the oil is also supplied to the second accumulator Acc2 via an oil passage L53a branched from the oil passage L53.

  At the time of in-gear, the line oil pressure PL is set in a low pressure region, but in this embodiment, it is possible to appropriately suppress long-period pulsation of the oil pressure by using two accumulators, the first accumulator Acc1 and the second accumulator Acc2. it can. Further, since the second accumulator Acc2 is an accumulator for the second brake B2, it is not necessary to provide a new accumulator, and the size can be reduced as compared with a new accumulator for in-gear. .

  When the first gear is established, the first open / close solenoid valve 82a is closed when the power is off, and the first shift valve 91 is in a state where the spool 91a is positioned on the right side. As a result, the line oil pressure PL supplied from the oil passage L32 is supplied to the first clutch C1 and the first accumulator Acc1 via the first shift valve 91, the oil passage L52, the second shift valve 92, and the oil passage L72. The The first hydraulic pressure adjusting valve 81a is energized and the output hydraulic pressure is supplied to the second brake B2 via the oil passage L41, the lockup shift valve 94, the oil passage L53, the first shift valve 91, and the oil passage L71. To be supplied. Further, the oil is also supplied to the second accumulator Acc2 via an oil passage L53a branched from the oil passage L53. In this way, the first hydraulic clutch C1 and the second brake B2 are engaged, and the first gear is established.

  When establishing the second speed stage, the fourth hydraulic pressure regulating valve 81d is opened by energizing it so that the pressure can be freely adjusted, and the hydraulic pressure output from the fourth hydraulic pressure regulating valve 81d is set via the oil passage L44. By supplying the brake B1, the first brake B1 is engaged. Further, the line hydraulic pressure PL is supplied to the first clutch C1 in the same manner as when the first gear is established. In this way, the first clutch C1 and the first brake B1 are engaged, and the second gear is established.

  When establishing the third speed stage, the third hydraulic pressure regulating valve 81c is opened by energizing it so that the pressure can be freely adjusted, and the hydraulic pressure output from the third hydraulic pressure regulating valve 81c is changed to the oil passage L43, the third shift valve. 93, the oil pressure is supplied to the third clutch C3 via the oil passage L74, and the engagement pressure of the third clutch C3 is controlled by the third hydraulic pressure regulating valve 81c. Further, the line hydraulic pressure PL is supplied to the first clutch C1 in the same manner as the first speed stage and the second speed stage. In this way, the first clutch C1 and the third clutch C3 are engaged, and the third gear is established.

  There are two patterns when establishing the fourth gear. In the first pattern, when the fourth speed is established by upshifting from the third gear, or when the vehicle is predicted to shift down from the driving state to the third gear while driving at the fourth gear, This is a case where the unit TCU makes a determination, and this first pattern will be described as being defined as the fourth speed stage Low. In the second pattern, when the 4th speed is established by downshifting from the 5th speed, or when the vehicle is predicted to be shifted up to the 5th speed while driving at the 4th speed, This is a case where the control unit TCU makes a determination, and this second pattern will be described as being defined as the fourth speed stage High.

  In the case of the fourth speed Low, the second hydraulic pressure adjusting valve 81b is opened by energizing it so that the pressure can be freely adjusted. The second clutch C2 is engaged by supplying it to the second clutch C2 via the oil path L51, the second shift valve 92, and the oil path L73. Further, the line hydraulic pressure PL is supplied to the first clutch C1 in the same manner as in the first to third gears. In this way, the first clutch C1 and the second clutch C2 are engaged, and the fourth speed (Low) is established.

  In the case of the fourth speed High, the second open / close solenoid valve 82b is opened by energization, and the second shift valve 92 is moved to the left side. At this time, the line oil pressure PL supplied from the oil passage L32 is supplied to the locking annular groove 92c provided in the spool 92a, and the spool 92a is moved to the left side even when the second open / close solenoid valve 82b is closed with the power off. The self-lock function that keeps moving is activated.

  The second hydraulic pressure adjustment valve 81b is opened in a freely adjustable manner when energized, and the hydraulic pressure output from the second hydraulic pressure adjustment valve 81b passes through the oil passage L42, the second shift valve 92, and the oil passage L72. 1 clutch C1 is supplied. The line oil pressure PL supplied from the oil passage L32 is supplied to the second clutch C2 via the first shift valve 91, the oil passage L52, the second shift valve 92, and the oil passage L73. Thus, the first clutch C1 and the second clutch C2 are engaged, and the fourth speed (High) is established.

  Note that the self-lock function of the second shift valve 92 needs to be canceled in order to shift from the fourth speed stage High state to the fourth speed stage Low state. In this case, the first open / close solenoid valve 82a is opened when energized, and the line hydraulic pressure PL is supplied to the left end portion of the second shift valve 92 via the oil passage L61. Is released. Further, by opening the first open / close solenoid valve 82a, the line hydraulic pressure PL is supplied to the right end portion of the first shift valve 91 through the oil passage L61, but the third open / close solenoid valve 82c is energized on. And the line oil pressure PL is supplied to the left end portion of the first shift valve 91 via the oil passage L63, so that the spool 91a of the first shift valve 91 is prevented from moving to the left side and is positioned on the right side. Maintained.

  When establishing the fifth gear stage, the third hydraulic pressure regulating valve 81c is opened to energize and the pressure can be freely adjusted, and the hydraulic pressure output from the third hydraulic pressure regulating valve 81c is changed to the oil passage L43, the third shift valve. 93, the oil pressure is supplied to the third clutch C3 via the oil passage L74, and the engagement pressure of the third clutch C3 is controlled by the third hydraulic pressure regulating valve 81c. Further, the line hydraulic pressure PL is supplied to the second clutch C2 in the same manner as the fourth speed high. In this way, the second clutch C2 and the third clutch C3 are engaged, and the fifth gear is established.

  When the sixth speed is established, the fourth hydraulic pressure regulating valve 81d is opened by energizing it so that the pressure can be freely adjusted, and the hydraulic pressure output from the fourth hydraulic pressure regulating valve 81d is set via the oil passage L44. By supplying the brake B1, the engagement pressure of the first brake B1 is controlled by the fourth hydraulic pressure regulating valve 81d. Further, the line hydraulic pressure PL is supplied to the second clutch C2 in the same manner as the fourth speed stage High and the fifth speed stage. In this way, the second clutch C2 and the first brake B1 are engaged, and the sixth speed is established.

  At the time of reverse in-gear when the manual valve 43 is switched to the reverse gear position “R”, the manual valve 43 shuts off the communication between the oil passage L21b and the oil passage L32 by the movement of the spool 43a. The line oil pressure PL supplied from the L21b is supplied to the oil passage L31 and the oil passage L33. The oil passage L33 is connected to the first shift valve 91, the third shift valve 93, and the lock-up shift valve 94.

  Further, the first open / close solenoid valve 82a is opened when energized, and the line hydraulic pressure PL is supplied to the right end portion of the first shift valve 91 via the oil passage L61. 91a moves to the left, and the oil passage L33 and the oil passage L71 communicate with each other. Thereby, the line oil pressure PL supplied to the first shift valve 91 via the oil passage L33 is supplied to the second brake B2 via the oil passage L71.

  In addition, the third hydraulic pressure adjusting valve 81c is opened to be adjustable when energized, and the output hydraulic pressure of the third hydraulic pressure adjusting valve 81c is supplied to the third clutch C3 via the oil passage L43, the third shift valve 93, and the oil passage L74. And the fourth accumulator Acc4.

  When the reverse gear is established, the second open / close solenoid valve 82b is opened by energization on from the reverse in-gear state, and the line hydraulic pressure PL is supplied to the right end portion of the third shift valve 93. 93 is in a state in which the spool 93a is moved to the left side. As a result, the oil passage L33 and the oil passage L74 communicate with each other, and the line oil pressure PL is supplied to the third clutch C3 and the fourth accumulator Acc4 via the oil passage L33, the third shift valve 93, and the oil passage L74. Is done. The second brake B2 is engaged in the same manner as the reverse in-gear. In this way, the third clutch C3 and the second brake B2 are engaged, and the reverse speed is established.

  Next, the operation of the oil pump drive device 2 will be described with reference to FIGS. The rotational speed of the driven shaft 22 connected to the oil pump 21 is transmitted to the driven shaft 22 through the second gear train 23 or through the first gear train 25 as the driving force of the drive shaft 24 is transmitted. It changes depending on how you do it. That is, the rotational speed of the driven shaft 22 varies depending on whether or not hydraulic pressure is supplied to the hydraulic engagement mechanism 25c.

  The lock-up shift valve 94 includes an oil passage L55 to which a hydraulic pressure such as a line hydraulic pressure PL is supplied, an oil passage L81 that supplies hydraulic pressure to the hydraulic engagement mechanism 25c, and an oil passage L82 that is blocked midway. And are connected.

  When the third open / close solenoid valve 82c is opened by energization and the line hydraulic pressure PL is supplied to the right end of the lockup shift valve 94, the spool 94a is moved to the left. In this state, the oil passage L55 and the oil passage L82 are communicated, and the communication between the oil passage L55 and the oil passage L81 is blocked.

  When the communication between the oil passage L55 and the oil passage L81 is interrupted, the supply of hydraulic pressure to the hydraulic engagement mechanism 25c is interrupted, so the hydraulic engagement mechanism 25c includes the drive shaft 24 and the first drive gear 25b. It will be in the open state where the connection of was cut. When the hydraulic engagement mechanism 25 c is in an open state, a driving force is transmitted between the drive shaft 24 and the driven shaft 22 via the second gear train 23. Since the gear ratio i of the second gear train 23 is set to “1” and is larger than the gear ratio j of the first gear train 25, the rotational speed of the driven shaft 22 is the same as the rotational speed of the drive shaft 24.

  On the other hand, when the third open / close solenoid valve 82c is closed by energization off, the supply of the line hydraulic pressure PL to the right end of the lockup shift valve 94 is cut off, and the spool 94a is moved to the right. In this state, the oil passage L55 communicates with the hydraulic engagement mechanism 25c via the oil passage L81, and hydraulic pressure is supplied to the hydraulic engagement mechanism 25c.

  The hydraulic engagement mechanism 25c supplied with the hydraulic pressure is in a connected state, and the drive shaft 24 and the first drive gear 25b are connected. At this time, since the gear ratio j of the first gear train 25 is set to be larger than “1” that is the gear ratio i of the second gear train 23, the rotational speed of the drive shaft 24 is the first gear train 25. The speed is increased and transmitted to the driven shaft 22. In this way, the rotational speed of the driven shaft 22 to which the oil pump 21 is connected is controlled by turning on and off the third open / close solenoid valve 82c.

  On the other hand, the lock-up clutch LC is also switched between the engaged state and the released state by turning on and off the third open / close solenoid valve 82c.

  The lock-up clutch LC involves in-gear when the select lever (not shown) is switched to the D range, when the vehicle does not advance despite the accelerator pedal being depressed, and with kick-down. At the time of rapid acceleration, it is not fastened and is in an open state. Therefore, the hydraulic pressure is supplied to the hydraulic engagement mechanism 25c via the oil path L81 during in-gear, stall, and sudden acceleration with kick-down. At this time, the rotation of the drive shaft 24 is transmitted to the driven shaft 22 via the first gear train 25, and the rotational speed of the driven shaft 22 connected to the oil pump 21 is set to "1". Then, the speed is increased to 1 / j (where j> 1).

  Further, during steady running in which the vehicle runs at a constant or substantially constant speed, the lockup clutch LC is engaged because the driving force required for the vehicle is small. Accordingly, during steady running, the hydraulic pressure supplied from the oil passage L55 is discharged from the discharge passage L82 and is not supplied to the hydraulic engagement mechanism 25c. Accordingly, the rotation of the drive shaft 24 is transmitted to the driven shaft 22 via the second gear train 23, and the rotational speed of the driven shaft 22 connected to the oil pump 21 is the same as the rotational speed of the drive shaft 24.

  Further, when the vehicle is traveling at a low speed, the lockup clutch LC is released because the driving force required for the vehicle is large.

  Further, when the vehicle is traveling at a high speed, the lockup clutch LC is engaged because the driving force required for the vehicle is small.

  Further, at the time of sudden acceleration accompanied by kick-down, the shift shock is suppressed by releasing the lock-up clutch LC. Therefore, at the time of sudden acceleration accompanied by kickdown, the rotational speed of the driven shaft 22 connected to the oil pump 21 is 1 / j when the rotational speed of the drive shaft 24 is “1”. Since the gear ratio j is set to be smaller than “1”, the rotation of the drive shaft 24 is accelerated and transmitted to the driven shaft 22 via the first gear train 25. Therefore, the rotation speed of the oil pump 21 is faster than the rotation speed of the drive shaft 24. During rapid acceleration with kickdown, the engine speed increases rapidly, and thus fuel consumption is lower than during steady running. Therefore, the effect of fuel consumption reduction due to an increase in the rotational speed of the oil pump 21 is small.

  According to the automatic transmission 1 using the oil pump drive device 2 of the first embodiment, when the lockup shift valve 94 is in the first shift state in which the engagement of the lockup clutch LC is permitted, The rotation is transmitted to the driven shaft 22 connected to the oil pump 21 via the first gear train 25 serving as the first transmission path. Conversely, when the lock-up shift valve 94 is in the second shift state in which the lock-up clutch LC is prohibited from being engaged, the rotation of the drive shaft 24 is performed via the second gear train 23 serving as the second transmission path. It is transmitted to the driven shaft 22 connected to the pump 21.

  That is, the oil pump drive device 2 according to the first embodiment interlocks with the switching of the lockup shift valve 94 that switches between engagement and disengagement of the lockup clutch LC, and the first gear train and the second gear as power transmission paths. The column is switched. For this reason, it is not necessary to provide a dedicated control unit for the hydraulic engagement mechanism as the switching mechanism, and the oil pump drive device can be easily controlled.

  Further, by providing the lockup shift valve 94 with an oil passage L81 that is an oil passage for the engagement mechanism, the lockup shift valve 94 has a function as a shift valve that switches supply / cutoff of hydraulic pressure to the hydraulic engagement mechanism 25c. Therefore, it is not necessary to separately provide a shift valve dedicated to the hydraulic engagement mechanism in the hydraulic circuit L. Therefore, it is possible to reduce the size of the oil pump drive device.

  Further, in conjunction with the engagement / release of the lock-up clutch LC, the hydraulic engagement mechanism switches to the first transmission path (first gear train 25) or the second transmission path (second gear train 23). Even when the engine 3 is rotating at a low speed, the oil pump 21 can be rotated at a high speed, and the oil pump 21 can be downsized.

  In the oil pump drive device 2 of the first embodiment, the drive shaft 24 and the second drive gear 23b are connected when the rotation speed of the drive shaft 24 exceeds the rotation speed of the second drive gear 23b. Conversely, when the rotational speed of the drive shaft 24 is lower than the rotational speed of the second drive gear 23b, a one-way clutch 23o is provided in which the connection between the drive shaft 24 and the second drive gear 23b is broken. Therefore, the number of oil passages provided in the drive shaft 24 can be reduced as compared with the case where a separate hydraulic engagement mechanism is provided instead of the one-way clutch 23o. As a result, the drive shaft 24 can be configured to be as small as possible (thin), and the configuration of the oil pump drive device can be simplified and downsized.

  In the first embodiment, the one-way clutch 23o is used to connect the second drive gear 23b and the drive shaft 24 of the second gear train 23. However, instead of the one-way clutch 23o, a hydraulic engagement mechanism is used. May be used. In this case, the oil pressure may be supplied to a hydraulic engagement mechanism that can connect the second drive gear 23b and the drive shaft 24 without closing the oil passage L82.

  In the first embodiment, the first driven gear 25a is the first fixed gear, the first drive gear 25b is the first shaft support gear, the second driven gear 23a is the second fixed gear, and the second drive gear 23b is the second. Although what was used as the shaft support gear was demonstrated, it is not restricted to this. For example, the first driven gear 25a may be a first support gear, the first drive gear 25b may be a first fixed gear, the second driven gear 23a may be a second support gear, and the second drive gear 23b may be a second fixed gear. . In this case, the one-way clutch 23o may be provided between the first driven gear 25a and the driven shaft 22, and the hydraulic engagement mechanism 25c may be provided so that the second driven gear 23a and the driven shaft 22 can be connected. Also by this, the same effect as 1st Embodiment can be acquired.

[Second Embodiment]
Next, an automatic transmission to which the oil pump drive device of the second embodiment is applied will be described with reference to FIGS. In addition, about the same content as 1st Embodiment, the same number is attached | subjected and the description is abbreviate | omitted.

  FIG. 7 shows an automatic transmission 1 according to the second embodiment. The automatic transmission 1 is configured in the same manner as the automatic transmission 1 of the first embodiment except for the oil pump drive device. The oil pump drive device 7 of 2nd Embodiment drives the oil pump 21 using the driving force of the engine 3 similarly to 1st Embodiment. A driven shaft 22 is connected to the oil pump 21.

  The oil pump drive device 7 includes a planetary gear mechanism 41 for driving the pump. The planetary gear mechanism 41 for driving the pump includes a sun gear 41s, a ring gear 41r, and a single-pinion type planetary planetary gear 41c that supports a plurality of pinions 41p that mesh with the sun gear 41s and the ring gear 41r so as to rotate and revolve, respectively. Consists of a gear device.

  FIG. 8 is a collinear diagram that can represent the ratio of the relative rotational speeds of the three elements of the sun gear 41s, the ring gear 41r, and the carrier 41c of the planetary gear mechanism 41 for driving the pump by a straight line. Referring to the collinear diagram of FIG. 8, the three elements of the sun gear 41s, the ring gear 41r and the carrier 41c of the planetary gear mechanism 41 for driving the pump are represented by the gear ratio (the number of teeth of the ring gear / the sun gear) in the collinear diagram of FIG. The first element is the sun gear 41s, the second element is the carrier 41c, and the third element is the ring gear 41r. Become.

  The ratio of the distances between the sun gear 41s (first element), the carrier 41c (second element), and the ring gear 41r (third element) is h: 1, where h is the gear ratio of the planetary gear mechanism 41 for driving the pump. Is set.

  The sun gear 41 s (first element) is connected to a driven shaft 22 connected to the oil pump 21. The carrier 41 c (second element) is connected to a drive shaft 24 to which the driving force of the engine 3 is transmitted via the idle gear train 31.

  Further, the oil pump drive device 7 is provided with a pump drive brake B3 that can be switched between a fixed state in which the ring gear 41r (third element) is fixed to the transmission case 13 and an open state in which the fixing is cut off. .

  The oil pump driving device 7 has a pump driving clutch C4 that can be switched between a connected state in which the carrier 41c (second element) and the ring gear 41r (third element) are connected and an open state in which the connection is cut off. Is provided.

  In the second embodiment, the pump drive brake B3 and the pump drive clutch C4 correspond to the hydraulic engagement mechanism of the present invention.

  When the pump drive brake B3 is fixed and the pump drive clutch C4 is released, the rotational speed of the ring gear 41r (third element) becomes “0”, the ring gear 41r (third element) and the carrier 41c (second element). The element) is in a state where relative rotation is possible. At this time, the rotational speed of the driven shaft 22 connected to the oil pump 21 is h + 1, and the rotation of the drive shaft 24 is increased and transmitted to the driven shaft 22. Thereby, the discharge amount of the hydraulic oil (lubricating oil) of the oil pump 21 can be increased. The transmission path of the driving force at this time corresponds to the first transmission path of the present invention.

  Further, when the pump drive brake B3 is in the released state and the pump drive clutch C4 is in the connected state, the rotational speed of the ring gear 41r (third element) is “1”, which is the same as the rotational speed of the carrier 41c (second element). It becomes. As a result, the first to third elements 41s, 41c, 41r of the planetary gear mechanism 41 for driving the pump are brought into a locked state in which relative rotation is impossible, and the rotational speed of the sun gear 41s (first element) is also changed to the ring gear 41r ( It becomes “1” which is the same as the third element) and the carrier 41c (second element). Thereby, the discharge amount of the hydraulic oil (lubricating oil) of the oil pump 21 can be reduced. The transmission path of the driving force at this time corresponds to the second transmission path of the present invention.

  In the hydraulic circuit L that controls the oil pump drive device 7 described above, the pump drive brake B3 is connected to the oil passage L81 instead of the hydraulic engagement mechanism 25c, as compared with the first embodiment, and the oil passage L82. Is the same as that of the first embodiment except that it is not closed and connected to the pump driving clutch C4. Therefore, the supply destination that supplies hydraulic pressure from the lock-up shift valve 94 via the oil passages L81 and L82 will be described with reference to FIGS. 5 and 6 of the first embodiment.

  The oil passage L81 of the second embodiment is connected to the pump drive brake B3. The oil passage L82 of the second embodiment is connected to the pump drive clutch C4.

  The third open / close solenoid valve 82c is opened when energized, and the line hydraulic pressure PL is supplied to the right end portion of the lockup shift valve 94, whereby the spool 94a is moved to the left side. In this state, the oil passage L55 is communicated with an oil passage L82 that connects the lockup shift valve 94 and the fourth clutch C4, and the hydraulic pressure output from the lockup shift valve 94 is supplied to the fourth clutch C4. . When the hydraulic pressure is supplied to the fourth clutch C4, the fourth clutch C4 enters a connected state. Then, since the communication between the oil passage L55 and the oil passage L81 is interrupted, the pump drive brake B3 is opened.

  Further, the third open / close solenoid valve 82c is closed by energization off, and the supply of the line hydraulic pressure PL to the lock-up shift valve 94 is stopped, whereby the spool 94a is moved to the right side. In this state, the oil passage L55 is communicated with an oil passage L81 connecting the lockup shift valve 94 and the third brake B3, and the hydraulic pressure output from the lockup shift valve 94 is supplied to the third brake B3. Thus, the pump drive brake B3 is in a fixed state. Then, since the communication between the oil passage L55 and the oil passage L82 is cut off, the pump driving clutch C4 is released.

  Also with the oil pump drive device 7 of the second embodiment, the same operational effects as those of the first embodiment can be obtained (however, the operational effects obtained by providing the one-way clutch 23o of the first embodiment can be obtained). except).

  In the second embodiment, the driven shaft 22 is connected to the sun gear 41s (first element) and the drive shaft 24 is connected to the carrier 41c (second element). However, the sun gear 41s (first element) is described. ) And the driven shaft 22 may be connected to the carrier 41c (second element). However, in this case, the capacity of the oil pump 21 needs to be increased.

  Further, the pump driving planetary gear mechanism 41 of the second embodiment is configured by a single pinion type planetary gear mechanism, but is not limited to this, and for example, the pump driving planetary gear mechanism includes a sun gear, a ring gear, and the like. A double-pinion type planetary gear mechanism including a pair of pinions that mesh with each other and one that meshes with the sun gear and the other meshes with the ring gear so as to rotate and revolve freely.

  In this case, in the collinear diagram shown in FIG. 8, the positions of the carrier and the ring gear are interchanged, the first element is the sun gear, the second element is the ring gear, and the third element is the carrier.

DESCRIPTION OF SYMBOLS 1 ... Automatic transmission, 2, 7 ... Oil pump drive device, 3 ... Engine (drive source), 4 ... Crankshaft, 11 ... Torque converter, 14 ... Input shaft, 16 ... Single planetary gear mechanism, 17 ... Double planetary gear Mechanism: 21 ... Oil pump, 22 ... Drive shaft, 23 ... Second gear train (second transmission path), 23a ... Second driven gear (second fixed gear), 23b ... Second drive gear (second shaft support gear) ), 24 ... drive shaft, 25 ... first gear train (first transmission path), 25a ... first driven gear (first fixed gear), 25b ... first drive gear (first shaft support gear), 25c ... hydraulic pressure. Type engagement mechanism, 41 ... planetary gear mechanism for driving the pump, 41s ... sun gear (first element), 41c ... carrier (second element), 41r ... ring gear (third element), 81a ... first hydraulic pressure regulating valve, 82c ... 3rd open / close solenoid valve, 94 ... Lock-up Shift valve, L ... hydraulic circuit, LC ... lockup clutch, C4 ... pump drive clutch (hydraulic engagement mechanism), B3 ... pump drive brake (hydraulic engaging mechanism).

Claims (4)

In an oil pump drive device that drives an oil pump that supplies lubricating oil to a transmission by using power of a drive source that drives a drive wheel of a vehicle,
In the vehicle, the power of the drive source is transmitted to the transmission via a torque converter having a lock-up clutch,
The lock-up clutch is permitted to be engaged when the lock-up shift valve provided in the hydraulic circuit is in the first shift state, and the engagement is prohibited when the lock-up shift valve is set to the second shift state. And
A drive shaft that rotates when the power of the drive source is transmitted;
A driven shaft connected to the oil pump;
A value obtained by dividing the rotational speed of the drive shaft by the rotational speed of the driven shaft is defined as a speed ratio, and a first transmission path for transmitting the rotation of the drive shaft to the driven shaft at a predetermined first speed ratio; ,
A second transmission path for transmitting to the driven shaft at a second speed ratio greater than the first speed ratio;
A hydraulic engagement mechanism that switches between the first transmission path and the second transmission path;
The lock-up shift valve is provided with an engagement mechanism oil passage for supplying hydraulic pressure to the hydraulic engagement mechanism,
When the lock-up shift valve is in the first shift state, rotation of the drive shaft is transmitted to the driven shaft via the first transmission path, and when the lock-up shift valve is in the second shift state An oil pump driving device, wherein hydraulic pressure is supplied to the oil passage for the engagement mechanism so that rotation of the driving shaft is transmitted to the driven shaft through the second transmission path. In the oil pump drive device according to claim 1,
The first transmission path includes a first support gear that is supported by the drive shaft or the driven shaft, and a first fixed gear that meshes with the first support gear and is fixed to the driven shaft or the drive shaft. A first gear train comprising gears,
The second transmission path includes a second shaft-supporting gear that is supported by the drive shaft or the driven shaft, and a second fixed gear that meshes with the second shaft-supporting gear and is fixed to the driven shaft or the driving shaft. An oil pump drive device comprising a second gear train comprising gears. In the oil pump drive device according to claim 2,
A one-way clutch is provided between any one of the first support gear and the second support gear and the drive shaft or the driven shaft that supports the one support gear;
The hydraulic engagement mechanism includes either the first shaft support gear or the second shaft support gear, and the drive shaft or the driven shaft that supports the other shaft support gear. An oil pump drive device characterized by being configured to be freely connectable. In the oil pump drive device according to claim 1,
A planetary gear mechanism for driving the pump including three elements including a sun gear, a carrier, and a ring gear;
The three elements of the planetary gear mechanism for driving the pump correspond to the gear ratio of the planetary gear mechanism for driving the pump in a collinear chart in which the ratio of the relative rotational speeds of the three elements can be represented by a straight line. As the first element, the second element, and the third element from one side in the order of arrangement at intervals,
The driven shaft is connected to the first element or the second element, the drive shaft is connected to the second element or the first element,
As the hydraulic engagement mechanism, a pump driving brake which can be switched between a fixed state in which the third element is fixed to the housing and an open state in which the fixing is cut off, and any one of the first to third elements A pump drive clutch that is switchable between a connected state connecting the two and an open state disconnecting the connection;
By rotating the pump driving brake and fixing the pump driving clutch, the rotation of the drive shaft is caused to rotate through one of the first transmission path and the second transmission path. Transmitted to the driven shaft,
By rotating the pump driving brake and bringing the pump driving clutch into a connected state, the rotation of the drive shaft is performed via the other transmission path of the first transmission path and the second transmission path. An oil pump drive device characterized by being transmitted to a driven shaft.

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