JP2019173769A - Control device of power transmission device for vehicle - Google Patents

Abstract

To prevent the generation of the over-rotation of a power source after a finish of an upshift of a power transmission device for a vehicle.SOLUTION: When a target gear change ratio γcvttgt is deviated from the lowest-side gear change ratio γmax after a finish of the execution of a stepped upshift of a power transmission device 16, a value of a gear change FB integration term for use in gear change FB control is set to zero once, and therefore, primary thrust Win can be increased, that is, primary pressure Pin can be boosted in order to perform an upshift of a continuously variable transmission 24 without affecting a correction at a side for reducing the primary thrust Win by the value of the gear change FB integration term for maintaining a gear change ratio γcvt of the continuously variable transmission 24 at the lowest-side gear change ratio γmax during the execution of the stepped upshift of the power transmission device 16. By this constitution, it can be prevented that the over-rotation of an engine 12 occurs after a finish of the stepped upshift of the power transmission device 16.SELECTED DRAWING: Figure 10

Description

  The present invention relates to a control device for a vehicle power transmission device including a plurality of power transmission paths provided in parallel between a power source and drive wheels.

  A control device for a vehicle power transmission device having a power transmission path via a continuously variable transmission mechanism in which a transmission element is wound between a primary pulley and a secondary pulley is well known. For example, this is the control device for a continuously variable transmission for a vehicle described in Patent Document 1. In Patent Document 1, in order to reliably maintain the speed ratio of the continuously variable transmission mechanism at the lowest speed ratio, a control variation that is a control variation in hydraulic control related to the speed change of the continuously variable transmission mechanism is considered. Thus, it is disclosed that the instruction value of the pulley hydraulic pressure supplied to the hydraulic actuator of the primary pulley is lowered from the set value for maintaining the lowest speed ratio.

  Further, as in the vehicle drive device described in Patent Document 2, for example, it is provided in parallel between an input rotation member that transmits power from a power source and an output rotation member that outputs the power to drive wheels. A vehicle having a first power transmission path via a gear mechanism having a gear stage, and a second power transmission path via a continuously variable transmission mechanism in which a higher gear ratio than the first power transmission path is formed. Power transmission devices are well known.

Japanese Patent No. 5435137 Japanese Patent Laying-Open No. 2015-105708

  By the way, in the control device for a vehicle power transmission device having a plurality of power transmission paths as described above, the vehicle power for switching from the state where the first power transmission path is formed to the state where the second power transmission path is formed. An upshift of the transmission device is executed with the speed ratio of the continuously variable transmission mechanism set to the lowest speed ratio, and during the upshift, the speed ratio of the continuously variable transmission mechanism is changed to the lowest speed ratio. In order to maintain reliably, it is possible to make the pulley hydraulic pressure supplied to the hydraulic actuator of a primary pulley lower than the pulley hydraulic pressure when not considering variation. As in the case of starting the vehicle when the amount of driving requested by the driver is increased, the vehicle is started with the first power transmission path formed, and after the rotational speed of the power source reaches the high rotation range, the vehicle power transmission is performed. When the upshift of the device is executed, after the upshift is completed, the transmission ratio of the continuously variable transmission mechanism is switched from the lowest transmission ratio to the transmission ratio according to the vehicle speed or the like. Executed. In such an infinitely variable transmission upshift, the primary pulley pulley hydraulic pressure is lowered to ensure that the lowest transmission ratio is maintained during the upshift of the vehicle power transmission device. The operation to increase the pulley hydraulic pressure of the pulley may be delayed. Then, the upshift of the continuously variable transmission mechanism is delayed, and the rotational speed of the power source becomes excessive, that is, the power source may be excessively rotated.

  The present invention has been made against the background of the above circumstances, and its purpose is to prevent the over-rotation of the power source from occurring after the completion of the upshift of the vehicle power transmission device. An object of the present invention is to provide a control device for a vehicular power transmission device.

  The gist of the first invention is that: (a) the power provided in parallel between an input rotary member to which power of a power source is transmitted and an output rotary member for outputting the power to a drive wheel; A plurality of power transmission paths each capable of transmitting from the input rotation member to the output rotation member, the plurality of power transmission paths being a first power transmission path via a gear mechanism having a gear; and For a vehicle that is a second power transmission path through a continuously variable transmission mechanism in which a transmission element is wound between a primary pulley and a secondary pulley, in which a gear ratio higher than the first power transmission path is formed (B) an upshift of the vehicular power transmission device that switches from a state in which the first power transmission path is formed to a state in which the second power transmission path is formed; The continuously variable transmission The transmission gear ratio is applied by the hydraulic actuator of the primary pulley by feedback control so that the actual transmission ratio of the continuously variable transmission mechanism matches the target transmission ratio. A shift control unit that corrects thrust of the primary pulley that pinches the transmission element; and (c) the shift control unit completes execution of the upshift of the vehicle power transmission device, and When the target gear ratio is switched from the lowest gear ratio to a gear ratio higher than the lowest gear ratio, the value of the integral term in the predetermined feedback control equation used in the feedback control is once set to zero. There is to do.

  According to the first aspect, after the execution of the upshift of the vehicle power transmission device is completed, the target transmission ratio of the continuously variable transmission mechanism is higher than the lowest transmission ratio from the lowest transmission ratio. When the gear ratio is switched to the side gear ratio, the value of the integral term in a predetermined feedback control equation used in the feedback control is once reset to zero, so that the continuously variable transmission mechanism during the execution of the upshift of the vehicle power transmission device In order to carry out the upshift of the continuously variable transmission mechanism without being affected by the correction of the primary pulley thrust by the value of the integral term for maintaining the transmission gear ratio at the lowest side gear ratio, Increasing, that is, the pulley hydraulic pressure of the primary pulley can be increased. Therefore, it is possible to prevent the over-rotation of the power source from occurring after completion of the upshift of the vehicle power transmission device.

BRIEF DESCRIPTION OF THE DRAWINGS It is a figure explaining the schematic structure of the vehicle to which this invention is applied, and is a figure explaining the principal part of the control function and various control systems for various control in a vehicle. It is a figure for demonstrating the structure of a continuously variable transmission mechanism, and the structure of a hydraulic control circuit. It is a figure which shows an example for demonstrating a thrust required for transmission control. It is a figure which shows an example of the relationship of each thrust in t2 time of FIG. FIG. 5 is a block diagram showing a control structure for achieving both a target shift and belt slip prevention with a minimum necessary thrust. It is a figure which shows an example of the thrust ratio map for calculating the thrust ratio used for calculation of the thrust by the side of a secondary pulley. It is a figure which shows an example of the thrust ratio map for calculating the thrust ratio used for calculation of the thrust by the side of a primary pulley. It is a figure which shows an example of the difference thrust map for calculating a secondary difference thrust. It is a figure which shows an example of the difference thrust map for calculating a primary difference thrust. It is a flowchart explaining the control operation | movement for preventing that an engine overspeed generate | occur | produces after the main part of the control operation | movement of an electronic control apparatus, ie, completion of the stepped upshift of a power transmission device.

  In the embodiment of the present invention, the primary pulley that is an input pulley and the secondary pulley that is an output pulley are, for example, a fixed sheave, a movable sheave, and a groove width between the fixed sheave and the movable sheave. And a hydraulic actuator for applying a thrust for changing. A vehicle including the vehicle power transmission device includes a hydraulic pressure control circuit that independently controls pulley hydraulic pressure as hydraulic pressure supplied to the hydraulic actuator. The hydraulic control circuit may be configured to generate pulley hydraulic pressure as a result by controlling the flow rate of hydraulic oil to the hydraulic actuator, for example. By such a hydraulic control circuit, each thrust (= pulley hydraulic pressure × pressure receiving area) in the primary pulley and the secondary pulley is controlled, so that the target shift can be realized while preventing the transmission element from slipping. Thus, the shift control is executed. The transmission element wound between the primary pulley and the secondary pulley includes an endless annular hoop and an element that is a thick plate piece block that is connected in the thickness direction along the hoop. An endless annular compression type transmission belt having an end, or a tension type transmission belt constituting an endless annular link chain in which ends of alternately linked link plates are connected to each other by connecting pins. The continuously variable transmission mechanism is a known belt-type continuously variable transmission. In a broad sense, the concept of this belt-type continuously variable transmission includes a chain-type continuously variable transmission.

  The gear ratio is “the rotational speed of the input-side rotating member / the rotational speed of the output-side rotating member”. For example, the gear ratio of the continuously variable transmission mechanism is “the rotational speed of the primary pulley / the rotational speed of the secondary pulley”. The transmission ratio of the vehicle power transmission device is “the rotational speed of the input rotary member / the rotational speed of the output rotary member”. The high side in the gear ratio is the high vehicle speed side, which is the side on which the gear ratio decreases. The low side in the gear ratio is the low vehicle speed side on which the gear ratio is increased. For example, the lowest speed ratio is a speed ratio on the lowest vehicle speed side that is the lowest vehicle speed side, and is a maximum speed ratio that gives the largest value.

  The power source is, for example, an engine such as a gasoline engine or a diesel engine that generates power by burning fuel. Further, the vehicle may include an electric motor or the like as the power source in addition to or instead of the engine.

  Hereinafter, embodiments of the present invention will be described in detail with reference to the drawings.

  FIG. 1 is a diagram illustrating a schematic configuration of a vehicle 10 to which the present invention is applied, and also illustrates a control function for various controls in the vehicle 10 and a main part of a control system. In FIG. 1, a vehicle 10 includes an engine 12 that functions as a power source, drive wheels 14, and a vehicle power transmission device 16 that is provided in a power transmission path between the engine 12 and the drive wheels 14. . Hereinafter, the vehicle power transmission device 16 is referred to as a power transmission device 16.

  The power transmission device 16 is connected to a known torque converter 20 as a fluid transmission device connected to the engine 12, an input shaft 22 connected to the torque converter 20, and an input shaft 22 in a case 18 as a non-rotating member. The continuously variable transmission mechanism 24, the forward / reverse switching device 26 connected to the input shaft 22 and the gear mechanism connected to the input shaft 22 via the forward / backward switching device 26 and provided in parallel with the continuously variable transmission mechanism 24. 28, a reduction gear comprising a pair of gears that are provided in mesh with each other so as not to rotate relative to the output shaft 30, the counter shaft 32, the output shaft 30, and the counter shaft 32, which are common output rotation members of the continuously variable transmission mechanism 24 and the gear mechanism A gear device 34, a gear 36 that is not rotatable relative to the counter shaft 32, a differential gear 38 connected to the gear 36, and the like are provided. The power transmission device 16 includes left and right axles 40 connected to the differential gear 38. The input shaft 22 is an input rotating member to which the power of the engine 12 is transmitted. The output shaft 30 is an output rotating member that outputs the power of the engine 12 to the drive wheels 14. As for the power, torque and force are the same unless particularly distinguished.

  In the power transmission device 16 configured as described above, the power output from the engine 12 is sequentially transmitted through the torque converter 20, the forward / reverse switching device 26, the gear mechanism 28, the reduction gear device 34, the differential gear 38, the axle 40, and the like. Is transmitted to the left and right drive wheels 14. Alternatively, in the power transmission device 16, the power output from the engine 12 is transmitted to the left and right drive wheels 14 via the torque converter 20, the continuously variable transmission mechanism 24, the reduction gear device 34, the differential gear 38, the axle 40, and the like sequentially. Is done.

  As described above, the power transmission device 16 includes the gear mechanism 28 and the continuously variable transmission mechanism 24 provided in parallel with the power transmission path PT between the engine 12 and the drive wheels 14. Specifically, the power transmission device 16 includes a gear mechanism 28 and a continuously variable transmission mechanism 24 provided in parallel with the power transmission path PT between the input shaft 22 and the output shaft 30. That is, the power transmission device 16 is provided in parallel between the input shaft 22 and the output shaft 30, and is capable of transmitting the power of the engine 12 from the input shaft 22 to the output shaft 30. It has. The plurality of power transmission paths are a first power transmission path PT1 via the gear mechanism 28 and a second power transmission path PT2 via the continuously variable transmission mechanism 24. That is, the power transmission device 16 includes a plurality of power transmission paths, that is, a first power transmission path PT1 and a second power transmission path PT2, between the input shaft 22 and the output shaft 30 in parallel. The first power transmission path PT1 is a power transmission path for transmitting the power of the engine 12 from the input shaft 22 to the drive wheels 14 via the gear mechanism 28. The second power transmission path PT2 is a power transmission path for transmitting the power of the engine 12 from the input shaft 22 to the drive wheels 14 via the continuously variable transmission mechanism 24.

  In the power transmission device 16, the power transmission path for transmitting the power of the engine 12 to the drive wheels 14 is switched between the first power transmission path PT <b> 1 and the second power transmission path PT <b> 2 according to the traveling state of the vehicle 10. Therefore, the power transmission device 16 includes a plurality of engagement devices that selectively form the first power transmission path PT1 and the second power transmission path PT2. The plurality of engagement devices include a first clutch C1, a first brake B1, and a second clutch C2. The first clutch C1 is provided in the first power transmission path PT1, and is an engagement device that selectively connects or disconnects the first power transmission path PT1, and is engaged when the vehicle moves forward. Thus, the engaging device forms the first power transmission path PT1. The first brake B1 is provided on the first power transmission path PT1, and is an engagement device that selectively connects or disconnects the first power transmission path PT1, and is engaged when the vehicle is moving backward. Thus, the engaging device forms the first power transmission path PT1. The first power transmission path PT1 is formed by engagement of the first clutch C1 or the first brake B1. The second clutch C2 is provided in the second power transmission path PT2, and is an engagement device that selectively connects or disconnects the second power transmission path PT2, and is engaged when engaged. 2 is an engagement device that forms a power transmission path PT2. The second power transmission path PT2 is formed by engagement of the second clutch C2. The first clutch C1, the first brake B1, and the second clutch C2 are all known hydraulic wet friction engagement devices that are frictionally engaged by a hydraulic actuator. The first clutch C1 and the first brake B1 are each one of the elements that constitute the forward / reverse switching device 26, as will be described later.

  The engine 12 includes an engine control device 42 having various devices necessary for output control of the engine 12, such as an electronic throttle device, a fuel injection device, and an ignition device. The engine 12 is controlled by an electronic control device 90 described later according to an accelerator operation amount θacc, which is an operation amount of an accelerator pedal corresponding to an amount required to drive the vehicle 10 by a driver. The engine torque Te, which is 12 output torque, is controlled.

  The torque converter 20 includes a pump impeller 20 p connected to the engine 12 and a turbine impeller 20 t connected to the input shaft 22. The power transmission device 16 includes a mechanical oil pump 44 connected to the pump impeller 20p. The oil pump 44 is rotationally driven by the engine 12, thereby controlling the speed of the continuously variable transmission mechanism 24, generating belt clamping pressure in the continuously variable transmission mechanism 24, and engaging each of the plurality of engagement devices. The original pressure of the operating hydraulic pressure for switching the operating state such as engagement and release is supplied to the hydraulic control circuit 46 provided in the vehicle 10.

  The forward / reverse switching device 26 includes a double pinion type planetary gear device 26p, a first clutch C1, and a first brake B1. The planetary gear device 26p is a differential mechanism having three rotating elements: a carrier 26c as an input element, a sun gear 26s as an output element, and a ring gear 26r as a reaction force element. The carrier 26 c is connected to the input shaft 22. The ring gear 26r is selectively connected to the case 18 via the first brake B1. The sun gear 26s is connected to a small-diameter gear 48 provided around the input shaft 22 so as to be rotatable relative to the input shaft 22 coaxially. The carrier 26c and the sun gear 26s are selectively coupled via the first clutch C1.

  The gear mechanism 28 is provided with a small diameter gear 48, a gear mechanism counter shaft 50, and a large diameter which is provided around the gear mechanism counter shaft 50 so as not to rotate relative to the gear mechanism counter shaft 50 and is engaged with the small diameter gear 48. And a gear 52. The large diameter gear 52 has a larger diameter than the small diameter gear 48. The gear mechanism 28 is coaxial with the output shaft 30 around the output shaft 30 and the idler gear 54 provided around the gear mechanism counter shaft 50 so as to be rotatable relative to the gear mechanism counter shaft 50. An output gear 56 that is provided in the center so as not to be relatively rotatable and meshes with the idler gear 54 is provided. The output gear 56 has a larger diameter than the idler gear 54. Accordingly, the gear mechanism 28 forms one gear stage in the power transmission path PT between the input shaft 22 and the output shaft 30. The gear mechanism 28 is a gear mechanism having a gear stage. The gear mechanism 28 is further provided between the large-diameter gear 52 and the idler gear 54 around the gear mechanism counter shaft 50, and meshes to selectively connect or disconnect the power transmission path between them. A type clutch D1 is provided. The meshing clutch D1 is an engagement device that selectively connects or disconnects the first power transmission path PT1, and is an engagement device that forms the first power transmission path PT1 when engaged. is there. The meshing clutch D1 is an engaging device that forms the first power transmission path PT1 by being engaged with the first clutch C1 or the first brake B1, and is included in the plurality of engaging devices. The operating state of the meshing clutch D1 is switched by the operation of a hydraulic actuator (not shown) provided in the power transmission device 16.

  The first power transmission path PT1 is formed by engaging the meshing clutch D1 and the first clutch C1 or the first brake B1 provided closer to the input shaft 22 than the meshing clutch D1. . A forward power transmission path is formed by the engagement of the first clutch C1, while a reverse power transmission path is formed by the engagement of the first brake B1. In the power transmission device 16, when the first power transmission path PT <b> 1 is formed, the power transmission state in which the power of the engine 12 can be transmitted from the input shaft 22 to the output shaft 30 via the gear mechanism 28 is set. . On the other hand, when the first clutch C1 and the first brake B1 are both released, or when the meshing clutch D1 is released, the first power transmission path PT1 is in a neutral state in which power transmission is impossible.

  FIG. 2 is a diagram for explaining the configuration of the continuously variable transmission mechanism 24. 1 and 2, the continuously variable transmission mechanism 24 includes a primary shaft 58 that is provided coaxially with the input shaft 22 and is integrally connected to the input shaft 22, and an effective diameter that is connected to the primary shaft 58 is variable. Primary pulley 60, secondary shaft 62 provided coaxially with output shaft 30, secondary pulley 64 having a variable effective diameter coupled to secondary shaft 62, and these pulleys 60, 64 are wound around each other. And a transmission belt 66 as a transmission element. The continuously variable transmission mechanism 24 is a well-known belt-type continuously variable transmission in which power is transmitted through frictional forces between the pulleys 60 and 64 and the transmission belt 66, and uses the power of the engine 12 as driving wheels 14. To the side. The frictional force is the same as the clamping pressure, and is also referred to as the belt clamping pressure. This belt clamping pressure is a belt torque capacity Tcvt that is a torque capacity of the transmission belt 66 in the continuously variable transmission mechanism 24.

  The primary pulley 60 includes a fixed sheave 60a coupled to the primary shaft 58, a movable sheave 60b provided so as not to rotate relative to the fixed sheave 60a around the axis of the primary shaft 58, and to be movable in the axial direction. And a hydraulic actuator 60c that applies a primary thrust Win to the movable sheave 60b. The primary thrust Win is the thrust of the primary pulley 60 (= primary pressure Pin × pressure receiving area) for changing the V groove width between the fixed sheave 60a and the movable sheave 60b. That is, the primary thrust Win is the thrust of the primary pulley 60 that pinches the transmission belt 66 applied by the hydraulic actuator 60c. The primary pressure Pin is a hydraulic pressure supplied to the hydraulic actuator 60c by the hydraulic control circuit 46, and is a pulley hydraulic pressure that generates the primary thrust Win. The secondary pulley 64 includes a fixed sheave 64a connected to the secondary shaft 62, and a movable sheave 64b provided so as not to be relatively rotatable around the axis of the secondary shaft 62 and movable in the axial direction with respect to the fixed sheave 64a. And a hydraulic actuator 64c that applies secondary thrust Wout to the movable sheave 64b. The secondary thrust Wout is a thrust of the secondary pulley 64 (= secondary pressure Pout × pressure receiving area) for changing the V groove width between the fixed sheave 64a and the movable sheave 64b. That is, the secondary thrust Wout is a thrust of the secondary pulley 64 that pinches the transmission belt 66 applied by the hydraulic actuator 64c. The secondary pressure Pout is a hydraulic pressure supplied to the hydraulic actuator 64c by the hydraulic control circuit 46, and is a pulley hydraulic pressure that generates the secondary thrust Wout.

  In the continuously variable transmission mechanism 24, the primary thrust Pin and the secondary thrust Wout are respectively controlled by adjusting the primary pressure Pin and the secondary pressure Pout by a hydraulic control circuit 46 that is driven by an electronic control unit 90 described later. The Thereby, in the continuously variable transmission mechanism 24, the V-groove width of each pulley 60, 64 is changed to change the engagement diameter (= effective diameter) of the transmission belt 66, and the transmission ratio γcvt (= primary rotational speed Npri / secondary rotation). (Speed Nsec) is changed, and the belt clamping pressure is controlled so that the transmission belt 66 does not slip. That is, by controlling the primary thrust Win and the secondary thrust Wout, the speed ratio γcvt of the continuously variable transmission mechanism 24 is set to the target speed ratio γcvttgt while preventing belt slippage that is the slippage of the transmission belt 66. The primary rotational speed Npri is the rotational speed of the primary shaft 58, and the secondary rotational speed Nsec is the rotational speed of the secondary shaft 62.

  In the continuously variable transmission mechanism 24, when the primary pressure Pin is increased, the V groove width of the primary pulley 60 is narrowed and the speed ratio γcvt is decreased. Decreasing the gear ratio γcvt means that the continuously variable transmission mechanism 24 is upshifted. In the continuously variable transmission mechanism 24, the highest gear ratio γmin is formed where the V groove width of the primary pulley 60 is minimized. The highest gear ratio γmin is the highest vehicle speed side gear ratio γcvt within the range of the gear ratio γcvt that can be formed by the continuously variable transmission mechanism 24, and the gear ratio γcvt is the smallest value. The minimum gear ratio. On the other hand, in the continuously variable transmission mechanism 24, when the primary pressure Pin is lowered, the V groove width of the primary pulley 60 is widened and the gear ratio γcvt is increased. Increasing the gear ratio γcvt means that the continuously variable transmission mechanism 24 is downshifted. In the continuously variable transmission mechanism 24, the lowest speed ratio γmax is formed when the V groove width of the primary pulley 60 is maximized. This lowest speed gear ratio γmax is a speed ratio γcvt on the lowest vehicle speed side that is the lowest vehicle speed side in the range of the gear ratio γcvt that can be formed by the continuously variable transmission mechanism 24, and the gear ratio γcvt is the largest value. It is the maximum gear ratio. In the continuously variable transmission mechanism 24, belt slippage is prevented by the primary thrust Win and the secondary thrust Wout, and the target gear ratio γcvttgt is realized by the mutual relationship between the primary thrust Win and the secondary thrust Wout. The target shift is not realized with only one thrust. As will be described later, the continuously variable transmission is performed by changing the thrust ratio τ (= Wout / Win), which is the value of the ratio between the primary thrust Win and the secondary thrust Wout, due to the mutual relationship between the primary pressure Pin and the secondary pressure Pout. The speed ratio γcvt of the mechanism 24 is changed. The thrust ratio τ is a value of the ratio of the secondary thrust Wout to the primary thrust Win. For example, the gear ratio γcvt is increased as the thrust ratio τ is increased, that is, the continuously variable transmission mechanism 24 is downshifted.

  The output shaft 30 is disposed so as to be rotatable relative to the secondary shaft 62 coaxially. The second clutch C <b> 2 is provided in the power transmission path between the secondary pulley 64 and the output shaft 30. The second power transmission path PT2 is formed by engaging the second clutch C2. In the power transmission device 16, when the second power transmission path PT <b> 2 is formed, a power transmission possible state in which the power of the engine 12 can be transmitted from the input shaft 22 to the output shaft 30 via the continuously variable transmission mechanism 24. Is done. On the other hand, the second power transmission path PT2 is set to the neutral state when the second clutch C2 is released. The gear ratio γcvt of the continuously variable transmission mechanism 24 corresponds to the gear ratio in the second power transmission path PT2.

  In the power transmission device 16, the speed ratio EL of the gear mechanism 28, which is the speed ratio γ gear (= input shaft rotational speed Nin / output shaft rotational speed Nout) in the first power transmission path PT1, is the maximum speed change in the second power transmission path PT2. The ratio is set to a value larger than the lowest speed ratio γmax of the continuously variable transmission mechanism 24. That is, the gear ratio EL is set to a gear ratio on the lower side than the lowest gear ratio γmax. The gear ratio EL of the gear mechanism 28 corresponds to the first speed gear ratio γ1 in the power transmission device 16, and the lowest speed gear ratio γmax of the continuously variable transmission mechanism 24 corresponds to the second speed gear ratio γ2 in the power transmission device 16. Equivalent to. In this way, the second power transmission path PT2 has a higher gear ratio than the first power transmission path PT1. The input shaft rotational speed Nin is the rotational speed of the input shaft 22, and the output shaft rotational speed Nout is the rotational speed of the output shaft 30.

  The vehicle 10 can selectively perform traveling in the gear traveling mode and traveling in the belt traveling mode. The gear travel mode is a travel mode in which travel can be performed using the first power transmission path PT1, and is a travel mode in which the first power transmission path PT1 is formed in the power transmission device 16. The belt traveling mode is a traveling mode in which traveling can be performed using the second power transmission path PT2, and the second power transmission path PT2 is formed in the power transmission device 16. In the gear travel mode, when enabling forward travel, the first clutch C1 and the meshing clutch D1 are engaged, and the second clutch C2 and the first brake B1 are released. In the gear travel mode, when the reverse travel is enabled, the first brake B1 and the meshing clutch D1 are engaged, and the second clutch C2 and the first clutch C1 are released. In the belt running mode, the second clutch C2 is engaged and the first clutch C1 and the first brake B1 are released. In this belt travel mode, forward travel is possible.

  The gear travel mode is selected in a relatively low vehicle speed region including when the vehicle is stopped. The belt running mode is selected in a relatively high vehicle speed region including a medium vehicle speed region. The meshing clutch D1 is engaged in the belt traveling mode in the medium vehicle speed region of the belt traveling mode, while the meshing clutch D1 is released in the belt traveling mode in the high vehicle speed region of the belt traveling mode. . The meshing clutch D1 is released in the belt travel mode in the high vehicle speed region, for example, the drag of the gear mechanism 28 and the like during traveling in the belt travel mode is eliminated, and the gear mechanism 28 and the planetary gear device are operated at the high vehicle speed. This is to prevent, for example, a pinion that is a component of 26p from rotating at a high speed.

  The vehicle 10 includes an electronic control device 90 as a controller including a control device for the power transmission device 16. The electronic control unit 90 includes, for example, a so-called microcomputer having a CPU, a RAM, a ROM, an input / output interface, and the like. Various controls of the vehicle 10 are executed by performing signal processing. The electronic control unit 90 controls the output of the engine 12, the shift control of the continuously variable transmission mechanism 24, the belt clamping pressure control, and the hydraulic control for switching the operating states of the plurality of engagement devices (C1, B1, C2, D1). Etc. The electronic control unit 90 is configured separately for engine control, hydraulic control, and the like as necessary.

  The electronic control device 90 includes various sensors and the like (for example, various rotational speed sensors 70, 72, 74, 76, an accelerator operation amount sensor 78, a throttle opening sensor 80, a shift position sensor 82, etc.) provided in the vehicle 10. Detection signals and the like (for example, the engine rotation speed Ne, the primary rotation speed Npri that is the same as the input shaft rotation speed Nin, the secondary rotation speed Nsec, the output shaft rotation speed Nout corresponding to the vehicle speed V, and the magnitude of the driver's acceleration operation) The accelerator operation amount θacc, the throttle opening degree tap, the operation position POSsh of the shift lever 84 as a shift switching device provided in the vehicle 10, and the like are respectively supplied. Further, the electronic control device 90 sends various command signals (for example, an engine control command signal Se for controlling the engine 12) to each device (for example, the engine control device 42, the hydraulic control circuit 46, etc.) provided in the vehicle 10 Hydraulic control command signal Sccvt for controlling the shift of the step transmission mechanism 24, belt clamping pressure, etc., and a hydraulic control command signal Scbd for controlling the operating states of the plurality of engagement devices, respectively. The The input shaft rotational speed Nin (= primary rotational speed Npri) is also the turbine rotational speed, the primary rotational speed Npri is also the rotational speed of the primary pulley 60, and the secondary rotational speed Nsec is the rotational speed of the secondary pulley 64. But there is. Further, the electronic control unit 90 calculates an actual speed ratio γcvt (= Npri / Nsec) that is the actual speed ratio γcvt of the continuously variable transmission mechanism 24 based on the primary rotational speed Npri and the secondary rotational speed Nsec.

  The operation position POSsh of the shift lever 84 is, for example, a P, R, N, D operation position. The P operation position is a parking operation position for selecting the P position of the power transmission device 16 in which the power transmission device 16 is in a neutral state and the output shaft 30 is mechanically fixed so as not to rotate. The neutral state of the power transmission device 16 is realized by, for example, releasing the first clutch C1, the first brake B1, and the second clutch C2. That is, the neutral state of the power transmission device 16 is a state in which neither the first power transmission path PT1 nor the second power transmission path PT2 is formed. The R operation position is a reverse travel operation position for selecting the R position of the power transmission device 16 that enables reverse travel in the gear travel mode. The N operation position is a neutral operation position for selecting the N position of the power transmission device 16 in which the power transmission device 16 is in the neutral state. The D operation position allows the forward travel in the gear travel mode, or executes the automatic shift control of the continuously variable transmission mechanism 24 in the belt travel mode to enable the forward travel. This is the forward travel operation position for selecting the position. Accordingly, the D operation position and the R operation position are travel operation positions for setting the power transmission device 16 in a power transmission enabled state.

  The electronic control unit 90 includes an engine control unit, that is, an engine control unit 92 and a shift control unit, that is, a shift control unit 94 in order to realize various controls in the vehicle 10.

  The engine control unit 92 applies the accelerator operation amount θacc and the vehicle speed V to a driving force map that is a relationship that has been obtained experimentally or design in advance and stored, that is, a predetermined relationship, for example. Is calculated. The engine control unit 92 sets a target engine torque Tetgt from which the target driving force Fwtgt is obtained, and outputs an engine control command signal Se for controlling the engine 12 to the engine control device 42 so as to obtain the target engine torque Tetgt. .

  If the operation position POSsh is the P operation position or the N operation position while the vehicle is stopped, the shift control unit 94 is a hydraulic control command signal for engaging the meshing clutch D1 in preparation for shifting to the gear travel mode. Scbd is output to the hydraulic control circuit 46. When the operation position POSsh is changed from the P operation position or the N operation position to the D operation position while the vehicle is stopped, the shift control unit 94 sends a hydraulic control command signal Scbd for engaging the first clutch C1 to the hydraulic control circuit 46. Output. As a result, the travel mode is shifted to a gear travel mode that enables forward travel. When the operation position POSsh is changed from the P operation position or the N operation position to the R operation position while the vehicle is stopped, the shift control unit 94 transmits a hydraulic control command signal Scbd for engaging the first brake B1 to the hydraulic control circuit 46. Output. As a result, the travel mode is shifted to a gear travel mode that allows reverse travel.

  The shift control unit 94 executes switching control for switching between the gear travel mode and the belt travel mode when the operation position POSsh is the D operation position. Specifically, the shift control unit 94 corresponds to the first speed gear stage corresponding to the gear ratio EL of the gear mechanism 28 in the gear travel mode and the lowest speed ratio γmax of the continuously variable transmission mechanism 24 in the belt travel mode. The vehicle speed V and the accelerator operation amount θacc are applied to an upshift line and a downshift line as a stepped shift map having a predetermined relationship and having a predetermined hysteresis for switching to the second speed shift stage. Thus, whether or not shifting is necessary is determined, and the traveling mode is switched based on the determination result.

  When the shift control unit 94 determines an upshift during the traveling in the gear traveling mode and switches to the belt traveling mode, the clutch that disengages the first clutch C1 and engages the second clutch C2 is engaged. A hydraulic control command signal Scbd for performing a to-clutch shift is output to the hydraulic control circuit 46. Thereby, the power transmission path PT in the power transmission device 16 is switched from the first power transmission path PT1 to the second power transmission path PT2. As described above, the shift control unit 94 performs the first shift from the gear travel mode in which the first power transmission path PT1 is formed by stepped shift control by releasing the first clutch C1 and engaging the second clutch C2. Upshift of the power transmission device 16 is performed to switch to the belt travel mode in which the two power transmission paths PT2 are formed. In this embodiment, the upshift of the power transmission device 16 that switches from the gear travel mode to the belt travel mode is referred to as a stepped upshift.

  When the shift control unit 94 determines a downshift during the traveling in the belt traveling mode and switches to the gear traveling mode, the shift control unit 94 releases the second clutch C2 and clutches the clutch so as to engage the first clutch C1. A hydraulic control command signal Scbd for performing a to-clutch shift is output to the hydraulic control circuit 46. As a result, the power transmission path PT in the power transmission device 16 is switched from the second power transmission path PT2 to the first power transmission path PT1. As described above, the shift control unit 94 performs the operation from the belt travel mode in which the second power transmission path PT2 is formed by the stepped shift control by releasing the second clutch C2 and engaging the first clutch C1. A downshift of the power transmission device 16 is performed to switch to the gear travel mode in which the one power transmission path PT1 is formed. In this embodiment, the downshift of the power transmission device 16 that switches from the belt travel mode to the gear travel mode is referred to as a stepped downshift.

  In the switching control for switching between the gear running mode and the belt running mode, the torque is transferred by the clutch-to-clutch shift through the state of the belt running mode in the middle vehicle speed region where the meshing clutch D1 is engaged. Only the first power transmission path PT1 and the second power transmission path PT2 are switched, so that the switching shock is suppressed. The switching control for switching between the gear traveling mode and the belt traveling mode is a stepped shift control by switching the operating state of the first clutch C1 and switching the operating state of the second clutch C2. In this embodiment, the switching control for switching between the gear travel mode and the belt travel mode is referred to as clutch-to-clutch shift control, that is, CtoC shift control.

  In the belt running mode, the speed change control unit 94 is configured to adjust the primary pressure Pin and the secondary pressure Pout so as to achieve the target speed ratio γcvttgt of the continuously variable transmission mechanism 24 while preventing belt slippage of the continuously variable transmission mechanism 24. A hydraulic control command signal Scvt for controlling the control is output to the hydraulic control circuit 46, and the continuously variable transmission mechanism 24 is shifted. The hydraulic control command signal Sccv is a primary command pressure Spin for setting the primary pressure Pin as the target primary pressure Pintgt, and a secondary command pressure Spout for setting the secondary pressure Pout as the target secondary pressure Pouttgt.

  The target primary pressure Pintgt is a target value of the primary pressure Pin that generates the primary target thrust Wintgt that is the target thrust of the primary pulley 60, that is, the target value of the primary thrust Win. The target secondary pressure Pouttgt is a target value of the secondary pressure Pout that generates a target thrust of the secondary pulley 64, that is, a secondary target thrust Wouttgt that is a target value of the secondary thrust Wout. In the calculation of the primary target thrust Wintgt and the secondary target thrust Wouttgt, the necessary thrust, which is the thrust necessary for preventing belt slippage of the continuously variable transmission mechanism 24 with the minimum necessary thrust, is taken into consideration. This necessary thrust is the belt slip limit thrust Wlmt, which is the thrust immediately before the belt slip of the continuously variable transmission mechanism 24 occurs. In this embodiment, the belt slip limit thrust Wlmt is referred to as a slip limit thrust Wlmt.

  Specifically, the shift control unit 94 calculates a primary target thrust Wintgt and a secondary target thrust Wouttgt, respectively. The shift control unit 94 uses, as the secondary target thrust Wouttgt, the secondary thrust Wout calculated based on the primary side slip limit thrust Winlmt that is the slip limit thrust Wlmt in the primary pulley 60 and the secondary side that is the slip limit thrust Wlmt in the secondary pulley 64. The larger thrust of the slip limit thrust Woutlmt is selected. The secondary thrust Wout calculated based on the primary-side slip limit thrust Winlmt is a secondary-side shift control thrust Woutsh that is a thrust required for shift control on the secondary pulley 64 side, as will be described later.

  The shift control unit 94 sets the primary thrust Win calculated based on the secondary target thrust Wouttgt as the primary target thrust Wintgt. As will be described later, the primary thrust Win calculated based on the secondary target thrust Wouttgt is a primary shift control thrust Winsh that is a thrust required for shift control on the primary pulley 60 side. Further, as will be described later, the shift control unit 94 performs primary side shift control thrust by feedback control of the primary thrust Win based on the shift ratio deviation Δγcvt (= γcvttgt−γcvt) between the target speed ratio γcvttgt and the actual speed ratio γcvt. Winsh is corrected, that is, the primary target thrust Wintgt is corrected. In this embodiment, this feedback control is referred to as shift FB control.

  In the correction of the primary side shift control thrust Winsh described above, the deviation between the target value and the actual value in the parameter corresponding to the gear ratio γcvt and one-to-one may be used instead of the gear ratio deviation Δγcvt. For example, in the correction of the primary side shift control thrust Winsh, the deviation ΔXin (= Xintgt−Xin) between the target pulley position Xintgt and the actual pulley position Xin (see FIG. 2) in the primary pulley 60 and the target pulley position Xouttgt in the secondary pulley 64 Deviation ΔXout (= Xouttgt−Xout) from actual pulley position Xout (see FIG. 2), deviation ΔRin (= Rintgt−Rin) between target belt engagement diameter Rintgt and actual belt engagement diameter Rin (see FIG. 2) in primary pulley 60 , Deviation ΔRout (= Routtgt−Rout) between the target belt engagement diameter Routtgt and the actual belt engagement diameter Rout (see FIG. 2) in the secondary pulley 64, and the difference ΔNpri (= Npritgt between the target primary rotation speed Npritgt and the actual primary rotation speed Npri. -Npri) can be used.

  The thrust necessary for the shift control described above is a thrust necessary for realizing the target shift, and is a thrust necessary for realizing the target gear ratio γcvttgt and the target gear shift speed dγtgt. The transmission speed dγ is, for example, a change amount (= dγcvt / dt) of the transmission ratio γcvt per unit time. In this embodiment, the transmission speed dγ is defined as the amount of pulley position movement per element of the transmission belt 66 (= dX / dNelm). “DX” is the axial displacement amount [mm / ms] of the pulley per unit time, and “dNelm” is the number of elements biting into the pulley per unit time [piece / ms]. The shift speed dγ is represented by a primary shift speed dγin (= dXin / dNelmin) and a secondary shift speed dγout (= dXout / dNelmout).

  Specifically, the thrust of each pulley 60, 64 in a steady state where the gear ratio γcvt is constant is referred to as a balance thrust Wbl. The balance thrust Wbl is also a steady thrust. The balance thrust Wbl of the primary pulley 60 is the primary balance thrust Winbl, the balance thrust Wbl of the secondary pulley 64 is the secondary balance thrust Woutbl, and the ratio of these is the thrust ratio τ (= Woutbl / Winbl). On the other hand, when a certain thrust is added to or subtracted from any of the thrusts of the pulleys 60 and 64 in the steady state, the steady state is broken and the gear ratio γcvt changes, and the magnitude of the thrust that is added or subtracted. A shift speed dγ corresponding to is generated. This added or subtracted thrust is referred to as shift difference thrust ΔW. Hereinafter, the shift difference thrust ΔW is referred to as a difference thrust ΔW. The differential thrust ΔW is also a transient thrust. The differential thrust ΔW when the target gear shift is realized on the primary pulley 60 side is a primary differential thrust ΔWin as the differential thrust ΔW converted on the primary pulley 60 side. The differential thrust ΔW when realizing the target shift on the secondary pulley 64 side is the secondary differential thrust ΔWout as the differential thrust ΔW converted on the secondary pulley 64 side.

  The thrust required for the speed change control described above is to realize the target speed ratio γcvttgt corresponding to one thrust based on the thrust ratio τ for maintaining the target speed ratio γcvttgt when one thrust is set. The other balance thrust Wbl and the difference thrust ΔW for realizing the target speed dγtgt when the target speed ratio γcvttgt is changed. The target shift speed dγtgt is represented by a primary target shift speed dγintgt and a secondary target shift speed dγouttgt. The primary differential thrust ΔWin is a positive value exceeding zero in the upshift state, that is, “ΔWin> 0”, and a negative value less than zero in the downshift state, that is, “ΔWin <0”, so that the gear ratio is constant. Then, it becomes zero, that is, “ΔWin = 0”. Further, the secondary differential thrust ΔWout becomes a negative value less than zero in the upshift state, that is, “ΔWout <0”, and if it is in the downshift state, it becomes a positive value exceeding zero, that is, “ΔWout> 0”. In the steady state, it is zero, that is, “ΔWout = 0”.

  FIG. 3 is a diagram for explaining the thrust required for the above-described shift control. FIG. 4 is a diagram illustrating an example of the relationship between the thrusts at time t2 in FIG. 3 and 4 show, for example, the primary set when the target upshift is realized on the primary pulley 60 side when the secondary thrust Wout is set so as to realize the belt slip prevention on the secondary pulley 64 side. An example of the thrust Win is shown. In FIG. 3, before the time point t1 or after the time point t3, the target speed ratio γcvttgt is in a constant steady state and ΔWin = 0, so the primary thrust Win is only the primary balance thrust Winbl (= Wout / τ). Since the target gear ratio γcvttgt is reduced from time t1 to time t3, the primary thrust Win is the sum of the primary balance thrust Winbl and the primary differential thrust ΔWin, as shown in FIG. The shaded portion of each thrust shown in FIG. 4 corresponds to each balance thrust Wbl for maintaining the target speed ratio γcvttgt at time t2 in FIG.

  FIG. 5 is a block diagram showing a control structure for achieving both the target shift and the belt slip prevention with the minimum necessary thrust, and is a diagram for explaining the hydraulic control in the continuously variable transmission mechanism 24, that is, CVT hydraulic control. is there.

  In FIG. 5, the transmission control unit 94 calculates a target transmission ratio γcvttgt. Specifically, the shift control unit 94 calculates the target primary rotational speed Npritgt by applying the accelerator operation amount θacc and the vehicle speed V to a CVT shift map that has a predetermined relationship, for example. The shift control unit 94 calculates a post-shift target speed ratio γcvttgtl (= Npritgt / Nsec) that is a speed ratio γcvt to be achieved after the speed change of the continuously variable transmission mechanism 24 based on the target primary rotational speed Npritgt. The speed change control unit 94 is in a predetermined relationship so as to realize a quick and smooth speed change, for example, based on the speed ratio γcvt before the start of speed change, the target speed ratio γcvttgtl after the speed change, and the difference therebetween. The target speed ratio γcvttgt is determined as the target value of the transient speed ratio γcvt. For example, the speed change control unit 94 determines the target speed ratio γcvttgt to be changed during the speed change as a function of the elapsed time that changes along the smooth curve that changes from the start of speed change toward the target speed ratio γcvttgtl after the speed change. This smooth curve is, for example, a primary delay curve or a secondary delay curve. When determining the target speed ratio γcvttgt, the speed change control unit 94 calculates the target speed change dγtgt during the speed change based on the target speed ratio γcvttgt. When the gear shift is completed and the target gear ratio γcvttgt is in a constant steady state, the target gear shift speed dγtgt is set to zero.

  The shift control unit 94 calculates an input torque to the continuously variable transmission mechanism 24 used for calculating the primary target thrust Wintgt and the secondary target thrust Wouttgt. The input torque to the continuously variable transmission mechanism 24 includes a belt input torque Tb1 for thrust ratio calculation as a first input torque used for calculating the thrust ratio τ for realizing the target speed ratio γcvttgt of the continuously variable transmission mechanism 24, and A belt input torque Tb2 for preventing belt slippage as a second input torque used for calculating each of the primary side slip limit thrust Winlmt and the secondary side slip limit thrust Woutlmt. In the present embodiment, the belt input torque Tb1 for thrust ratio calculation is referred to as thrust ratio calculation input torque Tb1, and the belt input torque Tb2 for preventing belt slip is referred to as belt slip prevention input torque Tb2.

  Specifically, the shift control unit 94 calculates the estimated value of the engine torque Te by applying the throttle opening degree tap and the engine speed Ne to an engine torque map that has a predetermined relationship, for example. The shift control unit 94 calculates the turbine torque Tt based on the estimated value of the engine torque Te and, for example, the characteristic of the torque converter 20 that is a predetermined relationship. The turbine torque Tt is an estimated value of input torque to the continuously variable transmission mechanism 24. The transmission control unit 94 uses the turbine torque Tt as the thrust ratio calculating input torque Tb1.

  As the belt slip preventing input torque Tb2, basically, the thrust ratio calculating input torque Tb1 may be used. However, it is not preferable that the slip limit thrust Wlmt is zero when the thrust ratio calculation input torque Tb1 is zero in consideration of variations and the like. Therefore, as the belt slip prevention input torque Tb2, a torque obtained by applying a lower limit guard process to the absolute value of the thrust ratio calculation input torque Tb1 is used. The shift control unit 94 selects the larger torque of the absolute value of the thrust ratio calculation input torque Tb1 and the minimum guaranteed torque Tblim as the belt slip prevention input torque Tb2. This minimum guaranteed torque Tblim is a predetermined lower limit torque for increasing the belt slip prevention input torque Tb2 to the safe side in consideration of variations in order to prevent belt slip, for example, and is a positive value torque. . When the thrust ratio calculation input torque Tb1 is a negative value, in consideration of low torque accuracy, a predetermined torque corresponding to the thrust ratio calculation input torque Tb1 is used as the belt slip prevention input torque Tb2. May be. This predetermined torque is, for example, a positive value torque determined in advance as a value larger than the absolute value of the thrust torque calculating input torque Tb1. Thus, the belt slip prevention input torque Tb2 is a torque based on the thrust ratio calculation input torque Tb1.

  In block B1 and block B2 of FIG. 5, the transmission control unit 94 calculates the slip limit thrust Wlmt based on the actual transmission ratio γcvt and the belt slip prevention input torque Tb2. Specifically, the shift control unit 94 calculates the secondary side slip limit thrust Woutlmt using the following equation (1). The shift control unit 94 calculates the primary side slip limit thrust Winlmt using the following equation (2). In the following equations (1) and (2), “Tb2” is the belt slip prevention input torque Tb2, and “Tout” is the torque obtained by converting the belt slip prevention input torque Tb2 to the secondary pulley 64 side (= γcvt × Tb2). = (Rout / Rin) × Tb2), “α” is the sheave angle of each pulley 60, 64, “μin” is a predetermined element-pulley friction coefficient in the primary pulley 60, and “μout” is a predetermined value in the secondary pulley 64 The coefficient of friction between the element and the pulley, “Rin” is a belt engagement diameter in the primary pulley 60 that is uniquely calculated from the actual transmission ratio γcvt, and “Rout” is the belt in the secondary pulley 64 that is uniquely calculated from the actual transmission ratio γcvt. It is a hook diameter (see FIG. 2).

Woutlmt = (Tout × cosα) / (2 × μout × Rout)
= (Tb2 × cosα) / (2 × μout × Rin) (1)
Winlmt = (Tb2 × cosα) / (2 × μin × Rin) (2)

  As the slip limit thrust Wlmt, for example, a value obtained by performing a lower limit guard process on the calculated slip limit thrust Wlmt may be used. For example, as the primary side slip limit thrust Winlmt used in the block B3 of FIG. 5, the shift control unit 94 is larger of the primary side slip limit thrust Winlmt calculated using the equation (2) and the primary side minimum thrust Winmin. Select the thrust of the direction. The primary-side minimum thrust Winmin is a hard limit minimum thrust of the primary pulley 60 including a thrust generated due to variations in control of the primary pressure Pin and a thrust generated by centrifugal hydraulic pressure in the hydraulic actuator 60c. The variation in the control of the primary pressure Pin is, for example, a predetermined primary that may be supplied from the hydraulic control circuit 46 to the hydraulic actuator 60c even if the primary command pressure Spin with the primary pressure Pin set to zero is output. This is the maximum value of the pressure Pin. The same applies to the secondary side slip limit thrust Woutlmt.

  In block B3 and block B6 of FIG. 5, the shift control unit 94 calculates a balance thrust Wbl. That is, the shift control unit 94 calculates the secondary balance thrust Woutbl with respect to the primary-side slip limit thrust Winlmt and the primary balance thrust Winbl with respect to the secondary target thrust Wouttgt.

Specifically, the speed change control unit 94 applies the target speed ratio γcvttgt and the reciprocal SFin ⁻¹ of the primary-side safety factor SFin to the thrust ratio map map (τin) as shown in FIG. The thrust ratio τin for realizing γcvttgt is calculated. The thrust ratio map map (τin) is a diagram showing an example of the relationship between the reciprocal SFin ⁻¹ of the primary-side safety factor and the thrust ratio τin determined in advance using the target speed ratio γcvttgt as a parameter. The thrust ratio τin is a thrust ratio for secondary thrust calculation as a thrust ratio used when calculating the thrust on the secondary pulley 64 side based on the thrust on the primary pulley 60 side. The shift control unit 94 calculates the secondary balance thrust Woutbl based on the primary slip limit thrust Winlmt and the thrust ratio τin using the following equation (3). The primary side safety factor SFin is, for example, “Win / Winlmt” or “Tb2 / Tb1”, and the reciprocal number SFin− ¹ of the primary side safety factor is, for example, “Winlmt / Win” or “Tb1 / Tb2”. is there. Further, the speed change control unit 94 realizes the target speed change ratio γcvttgt by applying the target speed change ratio γcvttgt and the reciprocal SFout ⁻¹ of the secondary side safety factor SFout to the thrust ratio map map (τout) as shown in FIG. 7, for example. The thrust ratio τout to be calculated is calculated. The thrust ratio map map (τout) is a diagram illustrating an example of a relationship between the reciprocal SFout ⁻¹ of the secondary-side safety factor and the thrust ratio τout that are determined in advance using the target speed ratio γcvttgt as a parameter. The thrust ratio τout is a primary thrust calculation thrust ratio as a thrust ratio used when calculating the thrust on the primary pulley 60 side based on the thrust on the secondary pulley 64 side. The shift control unit 94 calculates the primary balance thrust Winbl based on the secondary target thrust Wouttgt and the thrust ratio τout using the following equation (4). The secondary side safety factor SFout is, for example, “Wout / Woutlmt” or “Tb2 / Tb1”, and the reciprocal number SFout− ¹ of the secondary side safety factor is, for example, “Woutlmt / Wout” or “Tb1 / Tb2”. is there. Since the belt slip prevention input torque Tb2 is always a positive value, when the vehicle 10 is in a driving state such that the thrust ratio calculation input torque Tb1 is a positive value, the reciprocals of the safety factors SFin ⁻¹ and SFout are used. ^{Since −1 is} also a positive value, the value of the drive region is used as the thrust ratio τ. On the other hand, when the vehicle 10 in which the input torque Tb1 for thrust ratio calculation has a negative value is in a driven state, the reciprocals SFin ⁻¹ and SFout ^{−1 of the} safety factors are also negative values, so the thrust ratio τ is The value of the driven area is used. The reciprocals SFin ⁻¹ and SFout ⁻¹ may be calculated each time the balance thrust Wbl is calculated. However, if a predetermined value (for example, about 1 to 1.5) is set for the safety factors SFin and SFout, respectively. For example, the reciprocal number may be set.

Woutbl = Winlmt × τin (3)
Winbl = Wouttgt / τout (4)

As described above, the slip limit thrusts Winlmt and Woutlmt are calculated based on the belt slip prevention input torque Tb2 based on the thrust ratio calculation input torque Tb1. The reciprocals SFin ⁻¹ and SFout ⁻¹ of the above safety factors, which are the basis for calculating the thrust ratios τin and τout, are values based on the thrust ratio calculating input torque Tb1. Therefore, the transmission control unit 94 calculates the thrust ratio τ that realizes the target speed ratio γcvttgt of the continuously variable transmission mechanism 24 based on the thrust ratio calculation input torque Tb1.

  In block B4 and block B7 in FIG. 5, the shift control unit 94 calculates a differential thrust ΔW. That is, the shift control unit 94 calculates the secondary differential thrust ΔWout and the primary differential thrust ΔWin.

  Specifically, the shift control unit 94 calculates the secondary differential thrust ΔWout by applying the secondary target shift speed dγouttgt to, for example, a differential thrust map map (ΔWout) as shown in FIG. The differential thrust map map (ΔWout) is a diagram illustrating an example of a relationship between a predetermined secondary shift speed dγout and the secondary differential thrust ΔWout. The shift control unit 94 calculates a secondary shift control thrust Woutsh (= Woutbl + ΔWout) obtained by adding the secondary differential thrust ΔWout to the secondary balance thrust Woutbl as a secondary thrust necessary for preventing belt slippage on the primary pulley 60 side. Further, the shift control unit 94 calculates the primary differential thrust ΔWin by applying the primary target shift speed dγintgt to, for example, the differential thrust map map (ΔWin) as shown in FIG. The differential thrust map map (ΔWin) is a diagram illustrating an example of a relationship between a predetermined primary shift speed dγin and a primary differential thrust ΔWin. The shift control unit 94 adds the primary differential thrust ΔWin to the primary balance thrust Winbl to calculate the primary shift control thrust Winsh (= Winbl + ΔWin).

  In the calculation in the blocks B3 and B4, a predetermined physical characteristic diagram such as a thrust ratio map map (τin) as shown in FIG. 6 or a differential thrust map map (ΔWout) as shown in FIG. 8 is used. Therefore, there are variations in the physical characteristics in the calculation results of the secondary balance thrust Woutbl and the secondary differential thrust ΔWout due to individual differences in the hydraulic control circuit 46 and the like. Therefore, when considering such variations in physical characteristics, the shift control unit 94 may add the control margin Wmgn to the primary-side slip limit thrust Winlmt. The control margin Wmgn is a predetermined predetermined thrust corresponding to a variation with respect to physical characteristics related to the calculation of the secondary balance thrust Woutbl and the secondary differential thrust ΔWout. When considering the variation with respect to the physical characteristics as described above, the shift control unit 94 uses the equation “Woutbl = (Winlmt + Wmgn) × τin” shown in FIG. The balance thrust Woutbl is calculated. Note that the variation with respect to the physical characteristics is different from the actual variation of the pulley hydraulic pressure with respect to the hydraulic control command signal Scvt. The variation in the pulley hydraulic pressure is a relatively large value depending on the hardware unit such as the hydraulic control circuit 46, but the variation with respect to the physical characteristic is an extremely small value compared with the variation in the pulley hydraulic pressure.

  In block B5 of FIG. 5, the shift control unit 94 selects the larger thrust of the secondary side slip limit thrust Woutlmt and the secondary side shift control thrust Woutsh as the secondary target thrust Wouttgt.

  In block B8 in FIG. 5, the shift control unit 94 calculates a feedback control amount (= FB control amount) Winfb. Specifically, the transmission control unit 94 matches the actual transmission ratio γcvt with the target transmission ratio γcvttgt by using a predetermined feedback control expression (= FB control expression) as shown in the following equation (5), for example. The FB control amount Winfb for this purpose is calculated. In the following equation (5), “Δγcvt” is a gear ratio deviation Δγcvt, “Kp” is a predetermined proportionality constant, and “Ki” is a predetermined integration constant. Therefore, “Kp × Δγcvt” in the first term on the right side is a proportional term in the FB control equation, and “Ki × (∫Δγcvtdt)” in the second term on the right side is an integral term in the FB control equation. The shift control unit 94 calculates the value (= Winsh + Winfb) after correcting the primary shift control thrust Winsh by the shift FB control as the primary target thrust Wintgt by adding the FB control amount Winfb to the primary shift control thrust Winsh. To do. In this way, the shift control unit 94 corrects the primary thrust Win by the shift FB control so that the actual speed ratio γcvt of the continuously variable transmission mechanism 24 matches the target speed ratio γcvttgt. This shift FB control is a known PI control. In this embodiment, the proportional term in the FB control formula is referred to as a shift FB proportional term, and the integral term in the FB control formula is referred to as a shift FB integral term.

  Winfb = Kp × Δγcvt + Ki × (∫Δγcvtdt) (5)

  In block B9 and block B10 of FIG. 5, the shift control unit 94 converts the target thrust into the target pulley pressure. Specifically, the shift control unit 94 sets the secondary target thrust Wouttgt and the primary target thrust Wintgt to the target secondary pressure Pouttgt (= Wouttgt / pressure receiving area) and the target based on the pressure receiving areas of the hydraulic actuators 60c and 64c, respectively. Each is converted into a primary pressure Pintgt (= Wintgt / pressure receiving area). The shift control unit 94 sets the target secondary pressure Pouttgt and the target primary pressure Pintgt as the secondary command pressure Spout and the primary command pressure Spin, respectively.

  The shift control unit 94 outputs the primary command pressure Spin and the secondary command pressure Spout to the hydraulic control circuit 46 as the hydraulic control command signal Sccv so that the target primary pressure Pintgt and the target secondary pressure Pouttgt are obtained. The hydraulic control circuit 46 regulates the primary pressure Pin and the secondary pressure Pout according to the hydraulic control command signal Sccv.

  Here, in the power transmission device 16, the operating state of the second clutch C <b> 2 represented by four states of complete release, complete engagement, release transition, and engagement transition differs depending on the travel mode and the like. For example, in the belt travel mode, the second clutch C2 is completely engaged, while in the gear travel mode, the second clutch C2 is completely disengaged. Further, in the CtoC shift control of the power transmission device 16, the second clutch C2 is temporarily brought into a release transition state or an engagement transition state. In the belt running mode, when a garage operation is performed in which the shift lever 84 is operated between the N operation position and the D operation position, the second clutch C2 is temporarily in a release transition state or an engagement transition state. It is said. If the operating state of the second clutch C2 is different, the input torque to the continuously variable transmission mechanism 24 is different. That is, the input torque to the continuously variable transmission mechanism 24 is a torque corresponding to the operating state of the second clutch C2.

  In other than the belt travel mode, it is desirable to realize the target speed ratio γcvttgt of the continuously variable transmission mechanism 24 while preventing belt slippage of the continuously variable transmission mechanism 24 as in the belt travel mode. Therefore, the shift control unit 94 calculates the input torque to the continuously variable transmission mechanism 24 used for calculating the primary target thrust Wintgt and the secondary target thrust Wouttgt according to the operating state of the second clutch C2. That is, the shift control unit 94 calculates the thrust ratio calculation input torque Tb1 and the belt slip prevention input torque Tb2 according to the operating state of the second clutch C2.

  The shift control unit 94 is a method for calculating the primary target thrust Wintgt and the secondary target thrust Wouttgt, which is described by exemplifying the case where the vehicle 10 is in the belt running mode when the second clutch C2 is in the fully engaged state. Further, the thrust torque calculating input torque Tb1 is set as the turbine torque Tt, and the belt slip preventing input torque Tb2 is set as the turbine torque Tt considering the minimum guaranteed torque Tblim.

  When the second clutch C2 is completely released, the shift control unit 94 sets the thrust ratio calculation input torque Tb1 to zero, the belt slip prevention input torque Tb2, and the drag torque of the second clutch C2 to the primary shaft 58. The torque value converted upward, that is, the drag torque of the second clutch C2 converted onto the input shaft 22 is used. The drag torque of the second clutch C2 is, for example, a predetermined torque when the second clutch C2 is in a fully released state.

  When the second clutch C2 is in an engagement transition state, the shift control unit 94 sets the thrust ratio calculation input torque Tb1 to the torque value obtained by converting the C2 clutch torque Tcltc2 onto the primary shaft 58, that is, on the input shaft 22. The converted C2 clutch torque Tcltc2 is used. When the second clutch C2 is in an engagement transition state, the speed change control unit 94 selects the larger one of the thrust ratio calculation input torque Tb1 and the minimum guaranteed torque Tblim as the belt slip prevention input torque Tb2. That is, the belt slip prevention input torque Tb2 is set to the C2 clutch torque Tcltc2 converted on the input shaft 22 in consideration of the minimum guaranteed torque Tblim. The shift control unit 94 calculates the C2 clutch torque Tcltc2 by applying the C2 command pressure to, for example, a predetermined C2 clutch torque map. The minimum guaranteed torque Tblim here may be the same value as when the second clutch C2 is in the fully engaged state, or a different value may be used.

  When the second clutch C2 is in a disengagement transition state, the transmission control unit 94 uses the smaller torque of the turbine torque Tt and the C2 clutch torque Tcltc2 converted on the input shaft 22 as the thrust ratio calculation input torque Tb1. Select. When the second clutch C2 is in the disengagement transition state, the speed change control unit 94 converts the thrust ratio calculation input torque Tb1, the turbine torque Tt, and the second converted into the input shaft 22 as the belt slip prevention input torque Tb2. The larger torque of the drag torques of the clutch C2 is selected.

  On the other hand, it is desirable that the target speed ratio γcvttgt of the continuously variable transmission mechanism 24 is set according to the difference in the travel mode and the control state. While traveling in the belt travel mode, the shift control unit 94 sets the target speed ratio γcvttgt of the continuously variable transmission mechanism 24 as the target speed ratio γcvttgt calculated using the CVT speed map as described above.

  In consideration of controllability, the stepped upshift and stepped downshift which are CtoC shift control of the power transmission device 16 are performed between two gear stages each having a fixed gear ratio like a known stepped transmission. Similar to the stepped transmission control, it is desirable that the transmission be performed between the transmission ratio EL of the gear mechanism 28 and the predetermined transmission ratio γcvt of the continuously variable transmission mechanism 24. In the present embodiment, considering the amount of change in the primary rotational speed Npri during the CtoC shift control of the power transmission device 16 or the continuity of the driving force, the predetermined shift of the continuously variable transmission mechanism 24 is taken into consideration. The ratio γcvt is set to the lowest speed ratio γmax that becomes the speed ratio γcvt closest to the speed ratio EL of the gear mechanism 28. Suppressing the change amount of the primary rotational speed Npri leads to, for example, reducing the differential rotational speed during the engagement transition of the second clutch C2 to suppress the heat generation amount.

  The speed change control unit 94 executes the CtoC speed change control of the power transmission device 16 in a state where the speed change ratio γcvt of the continuously variable transmission mechanism 24 is set to the lowest side speed change ratio γmax. The shift control unit 94 sets the speed ratio γcvt of the continuously variable transmission mechanism 24 to the lowest speed ratio γmax in preparation for the stepped upshift of the power transmission device 16 during travel in the gear travel mode. That is, the shift control unit 94 maintains the speed ratio γcvt of the continuously variable transmission mechanism 24 at the lowest speed ratio γmax during the execution of the CtoC shift control of the power transmission device 16 or during travel in the gear travel mode.

  The shift control unit 94 reliably maintains the speed ratio γcvt of the continuously variable transmission mechanism 24 at the lowest speed ratio γmax during the execution of the CtoC shift control of the power transmission device 16 or during travel in the gear travel mode. As the target speed ratio γcvttgt used in the FB control expression of the above formula (5) in the speed FB control described above, the lowest target speed ratio γmaxtgt in consideration of hardware variations is set. The lowest target speed ratio γmaxtgt is, for example, a γmax design nominal value, which is a design nominal value of the lowest speed ratio γmax, and an upper limit value for hardware variations due to individual differences of the continuously variable transmission mechanism 24 with respect to the γmax design nominal value. Is a predetermined value to which γmax hard upper limit variation is added.

  When the lowest target speed ratio γmaxtgt considering the hardware variation is set as the target speed ratio γcvttgt, the primary target thrust Wintgt is set in the speed FB control compared to the case where the target speed ratio γcvttgt is a γmax design nominal value. The primary pressure Pin is lowered by being reduced. As a result, the speed ratio γcvt of the continuously variable transmission mechanism 24 is easily maintained reliably at the lowest speed ratio γmax. In this embodiment, a structure is employed in which the movement of the movable sheave 60b in the direction of widening the V groove width of the primary pulley 60 is mechanically blocked. In the continuously variable transmission mechanism 24, the V-groove width is maximized at the position of the movable sheave 60b where the movement of the primary pulley 60 in the direction of increasing the V-groove width is mechanically blocked. γmax is formed. When the transmission gear ratio γcvt of the continuously variable transmission mechanism 24 is the lowest transmission gear ratio γmax, the movable sheave 60b is mechanically prevented from moving in the direction of increasing the V groove width of the primary pulley 60. Even if the primary pressure Pin is lowered below the value for realizing the low-side gear ratio γmax, a belt torque capacity Tcvt for preventing belt slip is secured.

  By the way, when starting from the vehicle stop state, the vehicle travels in the gear travel mode in order to ensure power performance, and thereafter, the CtoC shift control of the power transmission device 16 that is a stepped upshift is executed. In the shift control in the continuously variable transmission mechanism 24 during the execution of the CtoC shift control, the shift FB control is executed. Therefore, the target command ratio γcvttgt is set to the lowest target shift ratio γmaxtgt, and the primary command pressure that lowers the primary pressure Pin Spin output is continued and the value of the shift FB integral term is integrated. After the stepped upshift is completed and the transition to the belt travel mode is completed, a target gear ratio γcvttgt corresponding to the accelerator operation amount θacc and the vehicle speed V is calculated using the CVT shift map. That is, the target speed ratio γcvttgt is switched from the lowest speed ratio γmax to a calculated value corresponding to the vehicle speed V, for example, as the vehicle speed V increases after the transition to the belt travel mode is completed. Thereby, the continuously variable transmission mechanism 24 is upshifted. However, at this time, the value of the integrated shift FB integral term is affected by the value of the shift FB integral term integrated to the side where the primary pressure Pin is decreased in order to reliably maintain the lowest speed ratio γmax. Until it is swept out, the output of the primary command pressure Spin that increases the primary pressure Pin in the shift control in the belt running mode in the continuously variable transmission mechanism 24, that is, the output that instructs the upshift of the continuously variable transmission mechanism 24 is delayed. There is a possibility. If the upshift of the continuously variable transmission mechanism 24 is delayed, there is a possibility that the engine speed Ne becomes high enough to reach the overspeed region (= overrev region). Such a phenomenon is likely to occur when the stepped upshift is executed in a state where the engine rotational speed Ne is in the high rotational speed range, such as when the accelerator operation amount θacc is increased.

  In response to the possibility that the upshift of the continuously variable transmission mechanism 24 may be delayed after completion of the stepped upshift as described above, and the engine 12 may be over-revised, the shift control unit 94 performs the belt travel mode after the completion of the stepped upshift. When the target gear ratio γcvttgt deviates from the lowest speed gear ratio γmax, that is, when the upshift of the continuously variable transmission mechanism 24 is started, the integrated value that is the value of the shift FB integral term is once reset to zero, that is, cleared to zero. Thus, it is possible to execute the output of the primary command pressure Spin that increases the primary pressure Pin according to the target speed ratio γcvttgt, and to prevent the engine 12 from over-rotating.

  The electronic control unit 90 further includes a state determination unit, that is, a state determination unit 106, in order to realize a control function of once setting the value of the above-described shift FB integral term to zero.

  The state determination unit 106 determines that the shift control unit 94 performs the target shift of the continuously variable transmission mechanism 24 after the shift control unit 94 completes execution of the stepped upshift of the power transmission device 16 and completes the transition to the belt travel mode. It is determined whether or not the ratio γcvttgt has been switched from the lowest gear ratio γmax to the higher gear ratio γcvthi, which is a higher gear ratio than the lowest gear ratio γmax. The state determination unit 106 determines whether or not the condition 1 that the target speed ratio γcvttgt deviates from the lowest speed ratio γmax is satisfied, so that the speed change control unit 94 sets the target speed ratio γcvttgt to the maximum. It is determined whether or not the low side gear ratio γmax is switched to the high side gear ratio γcvthi. When determining that the condition 1 is satisfied, the state determination unit 106 determines that the transmission control unit 94 has switched the target transmission ratio γcvttgt from the lowest transmission ratio γmax to the higher transmission ratio γcvthi.

  When the state determination unit 106 determines that the condition 1 is satisfied after the execution of the stepped upshift of the power transmission device 16 is completed, the transmission control unit 94 uses an FB control formula used in the transmission FB control. The value of the shift FB integral term at is once set to zero. In the subsequent shift control of the continuously variable transmission mechanism 24 in the belt running mode, the value of the shift FB integral term in the FB control formula used in the shift FB control is newly added.

  FIG. 10 is a flowchart for explaining a main part of the control operation of the electronic control unit 90, that is, a control operation for preventing the engine 12 from over-rotating after completion of the stepped upshift of the power transmission device 16. Yes, for example, after the execution of the stepped upshift of the power transmission device 16 is completed, it is repeatedly executed.

  In FIG. 10, first, when the condition 1 that the target speed ratio γcvttgt deviates from the lowest speed ratio γmax is satisfied in step (hereinafter, step is omitted) S10 corresponding to the function of the state determination unit 106. It is determined whether or not there is. For example, in the control operation repeatedly executed in this flowchart, the condition 1 is that the previous value of the target speed ratio γcvttgt is the lowest speed ratio γmax, and the current value of the target speed ratio γcvttgt is the lowest speed ratio γmax. It is a condition that it is not. If the determination at S10 is negative, this routine is terminated. If the determination in S10 is affirmative, in S20 corresponding to the function of the shift control unit 94, the integrated value that is the value of the shift FB integral term is once set to zero.

  As described above, according to the present embodiment, after the execution of the stepped upshift of the power transmission device 16 is completed, the target speed ratio γcvttgt of the continuously variable transmission mechanism 24 is increased from the lowest speed ratio γmax to the high side. When the gear ratio γcvthi is switched, that is, when the target gear ratio γcvttgt deviates from the lowest gear ratio γmax, the value of the gear shift FB integral term used in the gear shift FB control is once reset to zero. During execution of the stepped upshift, it is affected by the correction on the side of reducing the primary thrust Win by the value of the shift FB integral term for maintaining the speed ratio γcvt of the continuously variable transmission mechanism 24 at the lowest speed ratio γmax. Instead, the primary thrust Win can be increased, that is, the primary pressure Pin can be increased in order to perform the upshift of the continuously variable transmission mechanism 24. Therefore, it is possible to prevent the engine 12 from over-rotating after completion of the stepped upshift of the power transmission device 16.

  As mentioned above, although the Example of this invention was described in detail based on drawing, this invention is applied also in another aspect.

  For example, in the above-described embodiment, the second clutch C <b> 2 is provided in the power transmission path between the secondary pulley 64 and the output shaft 30, but is not limited to this aspect. For example, the secondary shaft 62 may be coupled to the output shaft 30 integrally, and the primary shaft 58 may be coupled to the input shaft 22 via the second clutch C2. That is, the second clutch C <b> 2 may be provided in a power transmission path between the primary pulley 60 and the input shaft 22.

  In the above-described embodiment, the gear mechanism 28 is a gear mechanism in which one gear stage having a lower gear ratio than the lowest gear ratio γmax of the continuously variable transmission mechanism 24 is formed. It is not restricted to an aspect. For example, the gear mechanism 28 may be a gear mechanism in which a plurality of gear stages having different gear ratios are formed. That is, the gear mechanism 28 may be a stepped transmission that is shifted to two or more stages. Alternatively, the gear mechanism 28 is a gear mechanism that forms a gear ratio higher than the highest gear ratio γmin of the continuously variable transmission mechanism 24 and a gear ratio lower than the lowest gear ratio γmax. Also good.

  Moreover, in the above-mentioned Example, although the driving mode of the power transmission device 16 was switched using the predetermined upshift line and downshift line, it is not restricted to this aspect. For example, the driving mode of the power transmission device 16 may be switched by calculating the target driving force Fwtgt based on the vehicle speed V and the accelerator operation amount θacc and setting a gear ratio that can satisfy the target driving force Fwtgt. .

  Moreover, in the above-mentioned Example, although the torque converter 20 was used as a fluid type transmission device, it is not restricted to this aspect. For example, instead of the torque converter 20, other fluid transmission devices such as a fluid coupling having no torque amplification function may be used. Alternatively, the fluid transmission device need not necessarily be provided. In addition, although the meshing clutch D1 is provided in the first power transmission path PT1 via the gear mechanism 28, the meshing clutch D1 is not necessarily provided in carrying out the present invention.

  The above description is only an embodiment, and the present invention can be implemented in variously modified and improved forms based on the knowledge of those skilled in the art.

12: Engine (power source)
14: Drive wheel 16: Vehicle power transmission device 22: Input shaft (input rotation member)
24: continuously variable transmission mechanism 28: gear mechanism 30: output shaft (output rotating member)
60: Primary pulley 60c: Hydraulic actuator 64: Secondary pulley 66: Transmission belt (transmission element)
90: Electronic control device (control device)
94: Shift control unit PT1: First power transmission path PT2: Second power transmission path

Claims (1)

Transmitting the power from the input rotating member to the output rotating member, which are provided in parallel between the input rotating member to which the power of the power source is transmitted and the output rotating member that outputs the power to the drive wheels. A plurality of power transmission paths, and the plurality of power transmission paths have a first power transmission path via a gear mechanism having a gear and a gear ratio on a higher side than the first power transmission path. A control device for a vehicle power transmission device, which is a second power transmission path through a continuously variable transmission mechanism in which a transmission element is wound between a primary pulley and a secondary pulley,
In the upshift of the vehicle power transmission device that switches from the state in which the first power transmission path is formed to the state in which the second power transmission path is formed, the gear ratio of the continuously variable transmission mechanism is the lowest side gear ratio. Of the primary pulley that clamps the transmission element applied by the hydraulic actuator of the primary pulley by feedback control so that the actual speed ratio of the continuously variable transmission mechanism matches the target speed ratio. Including a shift control unit for correcting thrust,
The shift control unit, after completing the upshift of the vehicle power transmission device, changes the target gear ratio of the continuously variable transmission mechanism from the lowest gear ratio to a higher gear than the lowest gear ratio. A control device for a vehicle power transmission device, characterized in that, when switched to a gear ratio, a value of an integral term in a predetermined feedback control equation used in the feedback control is once set to zero.

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