JP2018003971A - Control device of power transmission device for vehicle - Google Patents

Abstract

PROBLEM TO BE SOLVED: To solve a problem that, in a vehicle in which a crankshaft of an engine and a primary sheave are connected to each other, the primary sheave rotates accompanied by a start of the engine, therefore, in order to prevent a slide of a transmission belt accompanied by the start of the engine, it is necessary to fill oil into a secondary sheave in a short time, and the oil filling is accelerated by raising hydraulic pressure to the secondary sheave, however, there arises a risk that the durability of the belt is lowered due to a rise of the grip pressure of the transmission belt.SOLUTION: An oil amount Qneed necessary for filling oil into a secondary sheave 70 is calculated, and hydraulic pressure Psa necessary for completing the filling within a regulated time tb is calculated. The hydraulic pressure is increased for the regulated time tb, after the completion of the filling of the secondary sheave 70, a slide of the transmission belt 72 is prevented by returning the hydraulic pressure to normal hydraulic pressure Psb, and the lowering of the durability of the transmission belt 72 can be suppressed without applying excessive pressure to the transmission belt 72.SELECTED DRAWING: Figure 5

Description

  The present invention relates to a control device for a vehicle power transmission device including a belt-type continuously variable transmission, and more particularly to hydraulic control for suppressing slippage of a transmission belt of the belt-type continuously variable transmission at the start of an engine. .

  In the belt-type continuously variable transmission comprising a primary sheave and a secondary sheave each having a fixed sheave fixed to the rotating shaft and a movable sheave provided so as not to be relatively rotatable and movable in the direction of the rotating shaft. A vehicle power transmission device is disclosed in which the primary sheave is provided on an input shaft to which engine output torque is always input during operation of the engine. The vehicle of patent document 1 is it. In Patent Document 2, in a vehicle having a continuously variable transmission provided with a clutch that connects and disconnects the input shaft to which the output torque of the engine is always input during operation of the engine and the primary sheave, In order to prevent belt slippage due to insufficient sheave pressure resulting from insufficient oil filling into the sheave after starting, a technique for waiting for engagement of the clutch until oil filling after the engine is started is disclosed. Yes.

Japanese Patent No. 5444739 JP 2009-156317 A

  In a vehicle in which the primary sheave is provided on the input shaft of Patent Document 1, in a type in which engine output torque is always input to the input shaft during engine operation, the primary sheave also rotates when the engine is started. It becomes. For this reason, the technology disclosed in Patent Document 2, that is, after the engine is started, the rotation of the primary sheave is stopped by waiting for the engagement of the clutch until the oil filling to the primary sheave is completed, The technique disclosed in Patent Document 2 that prevents belt slippage due to insufficient sheave pressure resulting from insufficient oil filling cannot be used. Therefore, in such a vehicle, in order to avoid belt slipping, if the oil supply hydraulic pressure into the sheave is set high so that oil is quickly filled in the sheave immediately after the engine is started, a short time is required. Thus, the filling is completed and belt slip can be avoided. On the other hand, when the hydraulic pressure to the belt becomes high, the durability of the belt may be lowered.

  The present invention has been made in the background of the above circumstances, and the object of the present invention is to complete the oil filling into the sheave in a short time in order to avoid belt slip and to the belt. The purpose is to suppress the load and suppress the deterioration of the durability.

  The gist of the present invention is that: (a) a primary sheave provided on an input shaft to which an output torque of the engine is always inputted during operation of the engine, and a secondary sheave around which a transmission belt is wound together with the primary sheave A vehicle-type power transmission device comprising: a belt-type continuously variable transmission including: an oil pump that is rotationally driven by the engine; and a hydraulic control circuit that is supplied with an original pressure from the oil pump. Immediately after, a control device for a vehicle power transmission device that temporarily increases the hydraulic pressure supplied to the primary sheave and / or the secondary sheave, (b) for filling the primary sheave and / or the secondary sheave. The required oil inflow amount is set after the filling calculated from the transmission ratio of the continuously variable transmission. The amount of oil remaining in the primary sheave and / or the secondary sheave immediately after the start of the engine, the amount of oil discharged from the oil pump, and the hydraulic pressure control are determined from the oil filling amount held in the sheave and / or the secondary sheave. (C) The hydraulic pressure supplied to the primary sheave and / or the secondary sheave by subtracting the amount of oil flowing into the primary sheave and / or the secondary sheave estimated from the amount of oil consumed in the circuit Is calculated from the oil inflow amount and the oil temperature required to fill the primary sheave and / or the secondary sheave.

  In this case, the hydraulic pressure supplied to the primary sheave and / or the secondary sheave is calculated from the oil inflow amount and the oil temperature necessary to fill the primary sheave and / or the secondary sheave. By setting the oil pressure to be temporarily increased, the oil filling to the sheave can be completed in a short time to avoid belt slip, and after the oil filling is completed, an excessive load is applied to the electric belt. By setting the hydraulic pressure not to be applied, slippage of the electric belt can be avoided and a decrease in durability can be effectively suppressed.

It is a figure explaining the schematic structure of the vehicle to which the present invention is applied. It is a figure for demonstrating the switching of the running pattern of the power transmission device in the vehicle of FIG. It is a figure explaining the principal part of the control function and various control systems for various control in the vehicle of FIG. It is a figure explaining the part which controls the hydraulic pressure regarding a belt type continuously variable transmission, a 1st clutch, and a 2nd clutch among the hydraulic control circuits of FIG. It is a flowchart which shows the control action regarding the setting of the oil_pressure | hydraulic of the primary sheave and the secondary sheave of FIG.

  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. In FIG. 1, a vehicle 10 includes an engine 12 such as a gasoline engine or a diesel engine that functions as a driving source for traveling, a driving wheel 14, and a power transmission device 16 provided between the engine 12 and the driving wheel 14. It has. The power transmission device 16 is connected to the torque converter 20 as a fluid transmission device connected to the engine 12, the input shaft 22 connected to the torque converter 20, and the input shaft 22 in a housing 18 as a non-rotating member. A belt type continuously variable transmission 24 (hereinafter referred to as a continuously variable transmission), a forward / reverse switching device 26 connected to the input shaft 22, and a continuously variable transmission connected to the input shaft 22 via the forward / reverse switching device 26. Gear transmission mechanism 28 as a gear transmission provided in parallel with the machine 24, the output shaft 30, which is a common output rotating member of the continuously variable transmission 24 and the gear transmission mechanism 28, the counter shaft 32, the output shaft 30, and the counter shaft A reduction gear device 34 composed of a pair of gears that are provided in mesh with each other and engaged with each other, and a differential 36 connected to a gear 36 provided with the counter shaft 32 so as not to be relatively rotatable. Ya 38 includes an axle 40 or the like of the pair coupled to a differential gear 38. In the power transmission device 16 configured as described above, the power of the engine 12 (the torque and the force are synonymous unless otherwise distinguished) is transmitted to the torque converter 20, the continuously variable transmission 24 or the forward / reverse switching device 26, and the gear transmission mechanism 28. The reduction gear device 34, the differential gear 38, the axle 40, and the like are sequentially transmitted to the pair of drive wheels 14. During operation of the engine 12, the output torque of the engine 12 is always input to the input shaft 22.

  As described above, the power transmission device 16 transmits the power of the engine 12 to the engine 12 (here, the input shaft 22 which is an input rotating member to which the power of the engine 12 is transmitted) and the driving wheel 14 (here, the driving wheel 14 is transmitted). A gear transmission mechanism 28 as a first transmission unit and a continuously variable transmission 24 as a second transmission unit, which are provided in parallel with the output shaft 30 which is an output rotating member for output. Therefore, the power transmission device 16 transmits the power of the engine 12 from the input shaft 22 to the drive wheel 14 side (that is, the output shaft 30) via the gear transmission mechanism 28, and the power of the engine 12 is transmitted. A plurality of power transmission paths PT, which are the second power transmission path PT2 that is transmitted from the input shaft 22 to the drive wheel 14 side (that is, the output shaft 30) via the continuously variable transmission 24, are connected to the input shaft 22 and the output shaft 30. In parallel between. The power transmission device 16 is switched between the first power transmission path PT1 and the second power transmission path PT2 in accordance with the traveling state of the vehicle 10. Therefore, the power transmission device 16 has a plurality of engagements for selectively switching the power transmission path PT for transmitting the power of the engine 12 to the drive wheel 14 side between the first power transmission path PT1 and the second power transmission path PT2. Equipment. This engagement device includes a first clutch C1 that connects and disconnects the first power transmission path PT1, and a second clutch C2 that serves as a second engagement device that connects and disconnects the second power transmission path PT2.

  The torque converter 20 is provided coaxially with the input shaft 22 around the input shaft 22, and includes a pump impeller 20 p connected to the engine 12 and a turbine impeller 20 t connected to the input shaft 22. ing. The pump impeller 20p is supplied with hydraulic pressure for controlling the transmission of the continuously variable transmission 24, operating the plurality of engagement devices, and supplying lubricating oil to each part of the power transmission device 16. A mechanical oil pump 42 that is generated by being driven by rotation and supplied to the hydraulic control circuit 80 is connected. During operation of the engine 12, the output torque of the engine 12 is constantly input to the input shaft 22 via the torque converter 20.

  The forward / reverse switching device 26 is provided coaxially with the input shaft 22 around the input shaft 22 in the first power transmission path PT1, and includes a double pinion planetary gear device 26p, a first clutch C1, and a first clutch C1. One brake B1 is provided. 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 26c is integrally connected to the input shaft 22, the ring gear 26r is selectively connected to the housing 18 via the first brake B1, and the sun gear 26s is coaxial with the input shaft 22 around the input shaft 22. It is connected to a small-diameter gear 44 provided so as to be relatively rotatable. The carrier 26c and the sun gear 26s are selectively connected via the first clutch C1. Therefore, the first clutch C1 is an engagement device that selectively connects two of the three rotating elements for forward gear travel, and the first brake B1 is used for the reverse travel. This is an engagement device for selectively connecting a ring gear 26r as a reaction force element to the housing 18.

  The gear transmission mechanism 28 includes a small-diameter gear 44 and a large-diameter gear 48 that is provided around the gear mechanism counter shaft 46 so as not to rotate relative to the gear mechanism counter shaft 46 and meshes with the small-diameter gear 44. ing. The gear transmission mechanism 28 includes an idler gear 50 provided around the gear mechanism counter shaft 46 so as to be relatively rotatable coaxially with the gear mechanism counter shaft 46, and the output shaft 30 with respect to the output shaft 30. An output gear 52 that is provided on the coaxial center so as not to rotate relative to the idler gear 50 is provided. The output gear 52 has a larger diameter than the idler gear 50. Accordingly, the gear transmission mechanism 28 is a gear transmission mechanism in which one speed ratio (speed stage) as a predetermined speed ratio (speed stage) is formed in the power transmission path PT between the input shaft 22 and the output shaft 30. It is. Around the gear mechanism counter shaft 46, a meshing clutch D <b> 1 is provided between the large-diameter gear 48 and the idler gear 50 to selectively connect and disconnect between them. The meshing clutch D1 is provided in the power transmission device 16 and is disposed in a power transmission path between the forward / reverse switching device 26 (the first friction clutch also agrees) and the output shaft 30 (in other words, the above-described clutch). A third engagement device (provided on the output shaft 30 side of the first clutch C1) and the first power transmission path PT1 (in other words, the first power transmission by being engaged with the first clutch C1). The third engagement device that forms the path PT1) is included in the plurality of engagement devices.

  Specifically, the meshing clutch D1 includes a clutch hub 54 provided around the gear mechanism counter shaft 46 so as not to rotate relative to the gear mechanism counter shaft 46, an idler gear 50, and a clutch hub 54. A clutch gear 56 disposed between and fixed to the idler gear 50 is spline-fitted to the clutch hub 54 so that the gear mechanism counter shaft 46 cannot rotate relative to the shaft center and is parallel to the shaft center. And a cylindrical sleeve 58 provided so as to be relatively movable in various directions. The sleeve 58 that is always rotated integrally with the clutch hub 54 is moved to the clutch gear 56 side and meshed with the clutch gear 56, whereby the idler gear 50 and the gear mechanism counter shaft 46 are connected. Further, the meshing clutch D1 includes a known synchromesh mechanism S1 as a synchronizing mechanism that synchronizes rotation when the sleeve 58 and the clutch gear 56 are engaged. In the meshing clutch D1 configured as described above, the fork shaft 60 is operated by the hydraulic actuator 62, whereby the sleeve 58 is connected to the shaft of the gear mechanism counter shaft 46 via the shift fork 64 fixed to the fork shaft 60. It is slid in a direction parallel to the center, and the engaged state and the released state are switched.

  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, and 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 transmission mechanism 28 is set. The On the other hand, the first power transmission path PT1 is in a neutral state (power transmission) that interrupts power transmission when at least the first clutch C1 and the first brake B1 are both released or at least the meshing clutch D1 is released. It is said that it is in a cut-off state.

  The continuously variable transmission 24 is provided on the input shaft 22 that rotates together with the engine 12 even when the engine 12 is in operation while the engine 12 is connected to the engine via the torque converter 20 and the second clutch C2 is released. A primary sheave (primary pulley) 66 having a variable effective diameter, a secondary sheave (secondary pulley) 70 having a variable effective diameter provided on a rotary shaft 68 coaxial with the output shaft 30, and the sheaves 66, 70. And a transmission belt 72 wound around the belt, and power is transmitted through frictional force (belt clamping pressure) between the sheaves 66, 70 and the transmission belt 72. In the primary sheave 66, the sheave hydraulic pressure supplied to the primary sheave 66 (that is, the primary pressure Pin supplied to the primary hydraulic actuator 66c) is driven by an electronic control device 90 (see FIGS. 3 and 4) corresponding to the control device. By controlling the pressure regulation by the control circuit 80 (see FIGS. 3 and 4), a primary thrust Win (= primary pressure Pin × pressure receiving area) for changing the V groove width between the fixed sheave 66a and the movable sheave 66b is applied. . In the secondary sheave 70, the sheave hydraulic pressure supplied to the secondary sheave 70 (that is, the secondary pressure Pout supplied to the secondary hydraulic actuator 70c) is regulated by the hydraulic control circuit 80, so that the fixed sheave 70a and the movable sheave 70 Secondary thrust Wout (= secondary pressure Pout × pressure receiving area) for changing the V groove width between 70 b is applied. In the continuously variable transmission 24, the primary thrust Win (primary pressure Pin) and the secondary thrust Wout (secondary pressure Pout) are controlled, so that the V-groove widths of the sheaves 66 and 70 change and the transmission belt 72 is engaged. The diameter (effective diameter) is changed, the gear ratio γcvt (= primary sheave rotation speed Npri / secondary sheave rotation speed Nsec) is changed, and the sheaves 66 and 70 and the transmission belt are prevented from slipping. The frictional force with 72 is controlled.

  The output shaft 30 is disposed around the rotation shaft 68 so as to be rotatable relative to the rotation shaft 68 coaxially. The second clutch C2 is provided on the drive wheel 14 (here, the output shaft 30 also agrees) side with respect to the continuously variable transmission 24 (that is, provided between the secondary sheave 70 and the output shaft 30). The secondary sheave 70 (rotary shaft 68) and the output shaft 30 are selectively connected or disconnected. 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 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 operation of the power transmission device 16 will be described below. FIG. 2 is a diagram for explaining the switching of the travel pattern using the engagement table of the engagement device for each travel pattern (travel mode) of the power transmission device 16 switched by the electronic control device 90. In FIG. 2, C1 corresponds to the operating state of the first clutch C1, C2 corresponds to the operating state of the second clutch C2, B1 corresponds to the operating state of the first brake B1, and D1 corresponds to the meshing clutch D1. Corresponding to the operating state, “◯” indicates engagement (connection), and “×” indicates release (cutoff).

  FIG. 3 is a diagram for explaining the main functions of the control function and the control system for various controls in the vehicle 10. In FIG. 3, the vehicle 10 includes an electronic control device 90 including a control device for the power transmission device 16, for example. Therefore, FIG. 3 is a diagram showing an input / output system of the electronic control unit 90, and is a functional block diagram for explaining a main part of a control function by the electronic control unit 90. 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. The CPU uses a temporary storage function of the RAM and follows a program stored in the ROM in advance. Various controls of the vehicle 10 are executed by performing signal processing. For example, the electronic control unit 90 executes output control of the engine 12, shift control of the continuously variable transmission 24, travel pattern switching control of the power transmission device 16, and the like. 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 based on detection signals from various sensors provided in the vehicle 10, such as various rotational speed sensors 110, 112, 114, 116, an oil temperature sensor 118, a line oil pressure sensor 120, an accelerator opening sensor 122, and the like. For example, the engine speed Ne (rpm), the primary sheave speed Npri (rpm) that is the input shaft speed Nin (rpm), the secondary sheave speed Nsec (rpm) that is the speed of the rotary shaft 68, and the vehicle speed V Corresponding output shaft rotation speed Nout (rpm), oil temperature Toil (° C.), line pressure Pl (MPa), accelerator opening θacc (%), etc. are supplied. The electronic control unit 90 also outputs an engine output control command signal Se for output control of the engine 12, a hydraulic control command signal Sccv for hydraulic control related to the shift of the continuously variable transmission 24, and a travel pattern of the power transmission device 16. The hydraulic control command signal Sswt and the like for controlling the first clutch C1, the first brake B1, the second clutch C2, and the meshing clutch D1 are output. For example, as a hydraulic control command signal Sswt, for driving each solenoid valve that regulates each hydraulic pressure supplied to each hydraulic actuator of the first clutch C1, the first brake B1, the second clutch C2, and the meshing clutch D1. A command signal (hydraulic command) is output to the hydraulic control circuit 80.

  FIG. 4 is a diagram for explaining a portion of the hydraulic control circuit 80 provided in the power transmission device 16 that controls the hydraulic pressure related to the start of the engine 12 to the start of the vehicle 10. The hydraulic control circuit 80 uses a modulator pressure Pm as a primary pressure, a primary solenoid valve SLP that controls a primary pressure Pin supplied to the primary pulley 66, and a secondary pressure Pout that supplies the modulator pressure Pm as a primary pressure to the secondary sheave 70. The secondary solenoid valve SLS for controlling the pressure, the C1 solenoid valve SL1 for controlling the C1 pressure Pc1, which is the hydraulic pressure supplied to the first clutch C1, using the line pressure Pl as a source pressure, and the line pressure Pl as the source pressure, And a C2 solenoid valve SL2 that controls a C2 pressure Pc2 that is a hydraulic pressure supplied to the clutch C2. In addition, a relief pressure type line pressure regulating valve that regulates the line pressure Pl using the discharge pressure of the oil pump 42 not shown as a source pressure, and a second pressure to the torque converter 20 using the relief pressure from the line pressure regulating valve as the source pressure. The second line pressure regulating valve that regulates the line pressure, the modulator valve that regulates a constant modulator pressure Pm using the discharge pressure of the oil pump 42 as a source pressure, the torque converter 20, the synchromesh mechanism S1, and the hydraulic pressure are supplied to the lock-up clutch. A solenoid valve (not shown) is provided. Further, in the valve body 74 or the hydraulic control circuit 80, for example, the line pressure Pl is adjusted using the hydraulic pressure discharged from the oil pump 42 as a base pressure, and the second line pressure and the modulator pressure Pm are adjusted, whereby the hydraulic pressure is increased. Is consumed. In FIG. 4, a two-dot chain line indicates a spool set, a switching valve, a pressure regulating valve, etc. (not shown) that supply hydraulic oil supplied from the oil pump 42 to the primary sheave 66, the secondary sheave 70, and the like. The valve body 74 is included.

  Each of the solenoid valves SLP, SLS, SL1, and SL2 is a linear solenoid valve that is driven by a hydraulic control command signal (drive current) output from the electronic control unit 90. The solenoid valves SLP and SLS are both normally open solenoid valves. The electromagnetic valves SL1 and SL2 are both normally closed electromagnetic valves. The solenoid valves SLP and SLS each output, for example, a hydraulic pressure using the modulator pressure Pm as a source pressure, and the solenoid valves SL1 and SL2 each output a hydraulic pressure using, for example, the line pressure Pl as a source pressure. The primary pressure control valve 82 is operated based on the hydraulic pressure Pslp output from the primary solenoid valve SLP, thereby adjusting the primary pressure Pin using the line pressure Pl as a source pressure. The secondary pressure control valve 84 is operated based on the hydraulic pressure Psls output from the secondary solenoid valve SLS, thereby adjusting the secondary pressure Pout using the line pressure Pl as a source pressure. The hydraulic pressure Pc1 output from the C1 solenoid valve SL1 is supplied to the first clutch C1. The hydraulic pressure Pc2 output from the C2 solenoid valve SL2 is supplied to the second clutch C2.

  The electronic control unit 90 calculates the amount of oil for filling the primary side hydraulic actuator 66c of the primary sheave 66 and the secondary side hydraulic actuator 70c of the secondary sheave 70, and sets the engine 12 in advance so as not to cause belt slip. By increasing the hydraulic pressure Pin and Pout by a predetermined specified time tb from the start of the engine, the primary hydraulic actuator 66c and the secondary hydraulic actuator 70c are set to a predetermined specified time tb from the start of the engine 12 set in advance. A series of controls is performed so that the belt clamping pressure is not excessively increased by immediately returning to the normal hydraulic pressure Ps. Since the gear ratio γcvt at the start of the engine 12 is set to γmax by increasing the hydraulic pressure to the secondary hydraulic actuator 70c immediately before the vehicle stops, the gear ratio γcvt immediately after the start of the engine 12 is set to γmax. Will be described.

  In FIG. 3, the elapsed time after starting determination unit 92 determines that the elapsed time from the start of the engine 12 is within a predetermined time ta. When the elapsed time from the start of the engine 12 is within the predetermined time ta, that is, immediately after the start, the sheave filling amount calculation means 98 is placed in the secondary sheave 70 based on the speed ratio γcvt, that is, γmax instructed by the speed ratio setting means 94. The hydraulic oil filling amount Qtgt in the secondary actuator 70c, which is the oil filling amount to be held, that is, the volume in the secondary actuator 70c is calculated based on the actual gear ratio γcvt (= γmax). The engine starting residual amount calculation means 102 is a relationship (map) between the soak time tc stored in advance based on the stop time that the engine 12 has been stopped at the time of starting the engine 12, that is, the soak time tc, and the residual oil amount Qa. ) To calculate a residual oil amount Qa that is an oil residual amount in the secondary actuator 70c when the engine 12 is started. The soak time counting means 96 counts the stop time during which the engine 12 has been stopped when the engine 12 is started, that is, the soak time tc. The inflow amount calculating means 100 calculates the discharge capacity Qop of the oil pump 42 based on a relationship (map) stored in advance from the engine speed Ne and the line pressure Pl. Further, based on the relationship (map) stored in advance from the line pressure Pl and the oil temperature Toil, etc., the amount of hydraulic oil consumed by the hydraulic control circuit 80, that is, the valve body 74, for example, the illustrated solenoid valves SLP, SLS, The amount of hydraulic oil consumed by SL1, SL2, and control valves 82, 84, a spool set, a switching valve, a pressure regulating valve, etc. (not shown) and an oil passage in the valve body is calculated. Further, the inflow amount calculating means 100 is based on the relationship (map) stored in advance from the discharge oil amount Qop of the oil pump 42 and the consumed oil amount Qvb consumed in the valve body 74 or the hydraulic circuit 80 at the specified time tb. An inflow amount Qin of hydraulic oil flowing into the secondary side actuator 70c is calculated.

  Further, the required oil amount calculating means 104 is calculated from the sieve filling amount Qtgt calculated by the sieve filling amount calculating means 98 from the residual amount Qa calculated by the engine starting residual amount calculating means 102 and the inflow amount calculating means 100. The required oil amount Qneed, which is the oil inflow amount required to fill the secondary sheave 70, is calculated by subtracting the inflow amount Qin. The required amount of oil Qneed is set to the normal pressure Psb set so that the secondary pressure Pout supplied to the secondary sheave 70 can prevent belt slippage at the start of the engine 12 and the life of the transmission belt 72 can be sufficiently maintained up to a specified time tb. The amount of hydraulic oil that is insufficient when maintained, that is, the amount of oil until the secondary actuator 70c reaches filling is shown. The required sheave hydraulic pressure calculating means 105 supplies the required oil quantity Qneed to the secondary actuator 70c by the predetermined specified time tb from the required oil quantity Qneed calculated by the required oil quantity calculating means 104 and the oil temperature Toil of the hydraulic oil. And the sheave hydraulic pressure setting means 106 controls the secondary control valve 84 to maintain the secondary pressure Pout at the sheave required hydraulic pressure Psa, based on the relationship (map) stored in advance. When the required oil pressure prescribed time determining means 108 determines the elapse of a prescribed time tb determined in advance from the start of the engine 12, the sheave oil pressure setting means 106 changes the sheave oil pressure Ps to the normal oil pressure Psb.

  FIG. 5 shows a main part of the control operation of the electronic control unit 90, that is, when the engine 12 is started, the secondary pressure Pout supplied to the secondary side actuator 70c of the secondary sheave 70 is temporarily increased to enter the secondary side actuator 70c. FIG. 6 is a flowchart for explaining a control operation in which the filling of the hydraulic oil is completed at a predetermined time tb and the hydraulic oil is filled into the secondary side actuator 70c and decreases to the normal pressure Psb, and is repeatedly executed when the engine 12 is started. Is done.

  In FIG. 5, in a step (hereinafter, step is omitted) S10 corresponding to the function of the elapsed time determining means 92 after starting, it is determined whether or not the elapsed time from the start of the engine 12 is within a predetermined time ta. If the determination in S10 is negative, this routine is terminated. If the determination in S10 is affirmative, in S20 corresponding to the functions of the transmission ratio setting means 94 and the sheave filling amount calculating means 98, the sheave filling amount Qtgt of the hydraulic oil in the secondary side actuator 70c, that is, in the secondary side actuator 70c. The volume of is calculated. When the gear ratio γ is set to γmax, the oil filling amount Qtgt of the hydraulic oil in the secondary actuator 70c is the maximum volume in the secondary actuator 70c. In S30 corresponding to the functions of the soak time measuring means 96 and the engine starting residual quantity calculating means 102, the engine 12 is started based on the soak time tc, that is, the time that the engine 12 has been stopped before the engine 12 is started. A residual oil amount Qa in the secondary actuator 70c is calculated. In S40 corresponding to the function of the inflow amount calculating means 100, the secondary actuator 70c is started from the start of the engine 12 to a predetermined specified time tb from the discharge capacity Qop of the oil pump 42 and the consumed oil amount Qvb consumed by the valve body 74. An inflow amount Qin that flows in is calculated. In S50 corresponding to the function of the required oil amount calculation means 104, the required oil amount Qneed is calculated by subtracting the residual oil amount Qa and the inflow amount Qin from the sheave filling amount Qtgt. In S60 corresponding to the function of the sheave required oil pressure calculating means 105, it is necessary to fill the secondary side actuator 70c from the start of the engine 12 to a predetermined specified time tb based on the required oil amount Qneed and the oil temperature Toil. The sheave required oil pressure Psa is calculated. In S70 corresponding to the function of the sheave oil pressure setting means 106, the secondary pressure Pout is adjusted to the sheave required oil pressure Psa by the secondary pressure control valve 84. In S80 corresponding to the function of the required oil pressure prescribed time determining means 108, it is determined whether or not a predetermined prescribed time tb has been reached since the start of the engine 12, and the sheave required oil pressure Psa is maintained until the prescribed time tb is reached. . When this determination is affirmative, that is, when the specified time tb has elapsed, the secondary pressure Pout is changed to the normal pressure Psb by the secondary pressure control valve 84 in S90 corresponding to the function of the sheave oil pressure setting means 106.

  In the vehicle 12 in which the input shaft 22 and the primary sheave 66 are connected, the primary sheave 12 also rotates as soon as the engine 12 is started. For this reason, after the engine 12 is started, the hydraulic oil is filled in the secondary hydraulic actuator 70c in the secondary sheave 70, and the sheave oil pressure Ps is maintained, so that the frictional force between the sheaves 66, 70 and the transmission belt 72 (belt clamping) Pressure) must be secured. However, if the oil supply hydraulic pressure into the sheave is set high, the filling of the secondary hydraulic actuator 70c is completed in a short time. On the other hand, when the hydraulic pressure to the transmission belt 72 increases, the durability of the belt decreases. There is a risk of this. For this reason, according to the present embodiment, the secondary pressure Pout is temporarily increased to fill the hydraulic actuator 70c by a predetermined time tb that needs to be filled in the secondary hydraulic actuator, and the filling is performed. After a predetermined time tb from the start of the engine 12 that is expected to be completed, the secondary pressure Pout is added to the normal sheave oil pressure Psb that prevents belt slippage and suppresses the decrease in the life of the transmission belt 72 caused by excessive pressure. By reducing this, it is possible to avoid slipping of the transmission belt 72 and to effectively suppress a decrease in durability of the transmission belt 72.

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

  In the above-described embodiment, the hydraulic pressure to the second-side hydraulic actuator 70c, that is, the secondary pressure Pout is increased to set the transmission ratio γcvt to γmax. However, the present invention is not limited to this mode. For example, the second-side hydraulic actuator of this embodiment Instead of 70c, the hydraulic pressure to the primary hydraulic actuator 66c, that is, the primary pressure Pin may be increased to prevent belt slippage. In the above-described embodiment, the gear ratio γcvt is set to γmax. However, the gear ratio is not limited to γmax and is set to a predetermined gear ratio γ, and the primary pressure Pin and the secondary pressure Pout are increased by a predetermined time tb. The normal hydraulic pressure may be reduced after the filling of the primary hydraulic actuator 66c and the secondary hydraulic actuator 70c is completed, that is, after the specified time tb has elapsed.

  In the above-described embodiment, the power of the engine 12 is connected from the input shaft 22 to the input shaft 22 to the drive wheel 14 side via the gear transmission mechanism 28 as a gear transmission portion provided in parallel with the continuously variable transmission 24. There are a plurality of power transmission paths PT1 including a first power transmission path PT1 for transmission and a second power transmission path PT2 for transmitting the power of the engine 12 from the input shaft 22 to the drive wheel 14 via the continuously variable transmission 24. The vehicle power transmission device is not limited to this mode, but may be a vehicle power transmission device having only a second power transmission path PT2 for transmitting power via the continuously variable transmission 24, for example.

  In the above-described embodiment, the engine 12 is exemplified as the driving force source, but the present invention is not limited thereto. For example, the driving force source may employ another prime mover such as an electric motor alone or in combination with the engine 12. Further, the power of the engine 12 is transmitted to the input shaft 22 via the torque converter 20, but the present invention is not limited to this. For example, instead of the torque converter 20, other transmission devices such as a fluid coupling (fluid coupling) and an electromagnetic clutch that do not have a torque amplification function may be used. Or this torque converter 20 does not necessarily need to be provided.

  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 16: vehicle power transmission device 22: input shaft 24: continuously variable transmission (belt type continuously variable transmission)
42: Oil pump 66: Primary sheave 70: Secondary sheave 72: Transmission belt 80: Hydraulic control circuit 90: Electronic control device (control device)
Toil: Oil temperature Qin: Inflow oil amount Qtgt: Oil filling amount Qa: Oil residual amount Qneed: Necessary oil amount Qop: Oil pump discharge oil amount Qvb: Oil consumption amount of valve body γcvt: Gear ratio

Claims (1)

A belt-type continuously variable transmission comprising a primary sheave provided on an input shaft to which output torque of the engine is always input during operation of the engine, and a secondary sheave around which a transmission belt is wound together with the primary sheave; In a vehicle power transmission device comprising an oil pump that is driven to rotate by an engine and a hydraulic pressure control circuit that is supplied with an original pressure from the oil pump, the primary sheave and / or the secondary sheave immediately after the engine is started. A control device for a vehicle power transmission device that temporarily raises the hydraulic pressure to be supplied,
Oil that is retained in the primary sheave and / or the secondary sheave after filling the amount of oil inflow required to fill the primary sheave and / or the secondary sheave calculated from the transmission ratio of the continuously variable transmission The primary sheave estimated from the amount of oil remaining in the primary sheave and / or the secondary sheave immediately after starting the engine, the amount of oil discharged from the oil pump and the amount of oil consumed by the hydraulic control circuit / Or calculated by subtracting the amount of oil flowing into the secondary sheave,
The hydraulic pressure supplied to the primary sheave and / or the secondary sheave is calculated from the oil inflow amount and the oil temperature required to fill the primary sheave and / or the secondary sheave. Control device for power transmission device.

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