WO2009101826A1 - Control system and control method for automatic transmission - Google Patents

Control system and control method for automatic transmission Download PDF

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Publication number
WO2009101826A1
WO2009101826A1 PCT/JP2009/050283 JP2009050283W WO2009101826A1 WO 2009101826 A1 WO2009101826 A1 WO 2009101826A1 JP 2009050283 W JP2009050283 W JP 2009050283W WO 2009101826 A1 WO2009101826 A1 WO 2009101826A1
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WO
WIPO (PCT)
Prior art keywords
pressure
friction engagement
hydraulic pressure
engagement elements
calculated
Prior art date
Application number
PCT/JP2009/050283
Other languages
French (fr)
Japanese (ja)
Inventor
Takaaki Tokura
Hideaki Otsubo
Original Assignee
Toyota Jidosha Kabushiki Kaisha
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Publication of WO2009101826A1 publication Critical patent/WO2009101826A1/en

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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16H—GEARING
    • F16H61/00—Control functions within control units of change-speed- or reversing-gearings for conveying rotary motion ; Control of exclusively fluid gearing, friction gearing, gearings with endless flexible members or other particular types of gearing
    • F16H61/0021—Generation or control of line pressure
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16H—GEARING
    • F16H61/00—Control functions within control units of change-speed- or reversing-gearings for conveying rotary motion ; Control of exclusively fluid gearing, friction gearing, gearings with endless flexible members or other particular types of gearing
    • F16H61/68—Control functions within control units of change-speed- or reversing-gearings for conveying rotary motion ; Control of exclusively fluid gearing, friction gearing, gearings with endless flexible members or other particular types of gearing specially adapted for stepped gearings
    • F16H61/684—Control functions within control units of change-speed- or reversing-gearings for conveying rotary motion ; Control of exclusively fluid gearing, friction gearing, gearings with endless flexible members or other particular types of gearing specially adapted for stepped gearings without interruption of drive
    • F16H61/686—Control functions within control units of change-speed- or reversing-gearings for conveying rotary motion ; Control of exclusively fluid gearing, friction gearing, gearings with endless flexible members or other particular types of gearing specially adapted for stepped gearings without interruption of drive with orbital gears
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16H—GEARING
    • F16H61/00—Control functions within control units of change-speed- or reversing-gearings for conveying rotary motion ; Control of exclusively fluid gearing, friction gearing, gearings with endless flexible members or other particular types of gearing
    • F16H2061/0015—Transmission control for optimising fuel consumptions

Definitions

  • the present invention relates to control of an automatic transmission for a vehicle, and more particularly to control of an automatic transmission that maintains a frictional engagement element for shifting with a linear solenoid valve.
  • a shift control device for an automatic transmission that forms a predetermined shift stage by supplying output hydraulic pressure of a linear solenoid valve controlled by a drive signal (drive current) from an electronic control device to a hydraulic friction engagement element.
  • a drive signal drive current
  • Patent Document 1 an example of hydraulic control in the case where the friction engagement element is maintained in the engaged state is disclosed in, for example, Japanese Patent Application Laid-Open No. 2003-42280 (Patent Document 1).
  • Japanese Patent Application Laid-Open No. 2003-42280 discloses an electronic control device for shifting which is a normally open type (a type in which the output hydraulic pressure is maximized when no power is supplied and the output hydraulic pressure decreases as the amount of current supplied increases).
  • a normally open type a type in which the output hydraulic pressure is maximized when no power is supplied and the output hydraulic pressure decreases as the amount of current supplied increases.
  • the linear solenoid valve is driven to maximize the output hydraulic pressure of the linear solenoid valve when the clutch is maintained in an engaged state.
  • the current is set to 0 (the duty ratio of the drive signal is 0%).
  • the present invention has been made to solve the above-described problems, and an object of the present invention is to provide a hydraulic shift friction engagement element and a friction engagement element by increasing an output hydraulic pressure with an increase in driving current.
  • a control device and a control method capable of reducing power consumption of a pressure regulating valve when maintaining a friction engagement element in an engaged state in an automatic transmission having a pressure regulating valve for outputting to a combined element. It is.
  • the control device includes an automatic transmission including a hydraulic shift friction engagement portion and a pressure adjusting portion that increases an output hydraulic pressure as the drive current increases and outputs the output hydraulic pressure to the friction engagement portion.
  • this control device is a hydraulic pressure that can be maintained in a state in which the friction engagement portion is engaged, and from the maximum value of the target command pressure of the output hydraulic pressure.
  • a calculation unit that calculates a small hydraulic pressure, and a pressure regulation control unit that outputs a drive current corresponding to the hydraulic pressure calculated by the calculation unit to the pressure regulation unit to control the pressure regulation unit.
  • the friction engagement portion includes a plurality of friction engagement elements.
  • the pressure regulating unit includes a plurality of pressure regulating valves respectively corresponding to the plurality of friction engagement elements.
  • the hydraulic pressure is calculated for each factor.
  • the pressure adjustment control unit outputs a drive current corresponding to the maximum pressure of the calculated hydraulic pressure to a pressure adjustment valve corresponding to two or more friction engagement elements that need to be maintained in an engaged state.
  • the friction engagement portion includes a plurality of friction engagement elements.
  • the pressure regulating unit includes a plurality of pressure regulating valves respectively corresponding to the plurality of friction engagement elements.
  • the hydraulic pressure is calculated for each factor.
  • the pressure regulation control unit outputs each drive current corresponding to the calculated hydraulic pressure to each pressure regulating valve corresponding to the calculated hydraulic pressure.
  • the automatic transmission is provided with a source pressure regulating valve that regulates the source pressure input to the pressure regulating unit.
  • a drive current corresponding to the calculated hydraulic pressure is output to the pressure adjustment unit by the pressure adjustment control unit, the control device causes the original pressure to be larger than the calculated hydraulic pressure by a predetermined value. It further includes an original pressure control unit that controls the pressure regulating valve.
  • the friction engagement element (friction engagement portion) is maintained in the engaged state by the normally closed type (type in which the output hydraulic pressure increases as the drive current increases).
  • a drive current corresponding to the hydraulic pressure that can be maintained in the engaged state and is smaller than the maximum value of the target command pressure of the output hydraulic pressure is output to the pressure regulating valve. Therefore, the drive current supplied to the pressure regulating valve is smaller than the maximum current value corresponding to the maximum value of the target command pressure. Thereby, the power consumption of a pressure regulation valve can be reduced, maintaining the state which engaged the friction engagement element.
  • FIG. (1) which shows the control structure of the program which ECU performs. It is a timing chart (the 1) of target command pressure computed by ECU. It is FIG. (2) which shows the control structure of the program which ECU performs. It is a timing chart (the 2) of the target command pressure computed by ECU.
  • This vehicle is a FR (Front engine Rear drive) vehicle.
  • FR Front engine Rear drive
  • a vehicle other than FR may be used.
  • the vehicle includes an engine 1000, an automatic transmission 2000, a torque converter 2100, a planetary gear unit 3000 that forms part of the automatic transmission 2000, a hydraulic circuit 4000 that forms part of the automatic transmission 2000, a propeller shaft 5000, A differential gear 6000, a rear wheel 7000, and an ECU (Electronic Control Unit) 8000 are included.
  • the control device according to the present embodiment is realized, for example, by executing a program recorded in a ROM (Read Only Memory) of ECU 8000.
  • Engine 1000 is an internal combustion engine that burns a mixture of fuel and air injected from an injector (not shown) in a combustion chamber of a cylinder. The piston in the cylinder is pushed down by the combustion, and the crankshaft is rotated. An auxiliary machine 1004 such as an alternator and an air conditioner is driven by the driving force of the engine 1000. A motor may be used as a power source instead of or in addition to engine 1000.
  • Automatic transmission 2000 is connected to engine 1000 via torque converter 2100. Automatic transmission 2000 shifts the rotational speed of the crankshaft to a desired rotational speed by forming a desired gear stage.
  • the driving force output from the automatic transmission 2000 is transmitted to the left and right rear wheels 7000 via the propeller shaft 5000 and the differential gear 6000.
  • the ECU 8000 includes a vehicle speed sensor 8002, a position switch 8006 of a shift lever 8004, an accelerator opening sensor 8010 of an accelerator pedal 8008, a pedaling force sensor 8014 of a brake pedal 8012, a throttle opening sensor 8018 of an electronic throttle valve 8016, An engine speed sensor 8020, an input shaft speed sensor 8022, and an output shaft speed sensor 8024 are connected via a harness or the like.
  • the vehicle speed sensor 8002 detects the vehicle speed V from the rotational speed of the drive shaft 6100.
  • the position switch 8006 detects the position (shift position) SP of the shift lever 8004.
  • the accelerator opening sensor 8010 detects the opening (accelerator opening) ACC of the accelerator pedal 8008.
  • the pedaling force sensor 8014 detects the pedaling force of the brake pedal 8012 (force that the driver steps on the brake pedal 8012).
  • the throttle opening sensor 8018 detects the opening (throttle opening) TH of the electronic throttle valve 8016.
  • the engine speed sensor 8020 detects a crankshaft speed (engine speed) NE of the engine 1000.
  • Input shaft rotational speed sensor 8022 detects input shaft rotational speed (turbine rotational speed of torque converter 2100) NT of automatic transmission 2000.
  • Output shaft rotational speed sensor 8024 detects the rotational speed (output shaft rotational speed) NOUT of the output shaft of automatic transmission 2000.
  • Each of these sensors transmits a signal representing the detection result to ECU 8000
  • ECU 8000 is sent from vehicle speed sensor 8002, position switch 8006, accelerator opening sensor 8010, pedal effort sensor 8014, throttle opening sensor 8018, engine speed sensor 8020, input shaft speed sensor 8022, output shaft speed sensor 8024, and the like. Based on the received signal, the map stored in the ROM, and the program, the devices are controlled so that the vehicle is in a desired running state.
  • the ECU 8000 when the shift lever 8004 is in the D (drive) position and the D (drive) range is selected as the shift range of the automatic transmission 2000, out of the first to eighth forward gears.
  • the automatic transmission 2000 is controlled so that any one of the gear positions is formed.
  • the automatic transmission 2000 can transmit the driving force to the rear wheels 7000 by forming any one of the first to eighth forward gears. In the D range, it may be possible to form a higher gear than the eighth gear.
  • the gear stage to be formed is determined based on a shift diagram created in advance by experiments or the like using the vehicle speed and the accelerator opening as parameters.
  • the ECU 8000 is described as one unit, but the ECU 8000 may be divided into two or more units.
  • the ECU 8000 includes an engine ECU that controls the engine 1000 and an ECT (Electronic Controlled Transmission) _ECU that controls the automatic transmission 2000 so that the engine ECU and the ECT_ECU can transmit and receive signals to and from each other. It may be.
  • ECT Electronic Controlled Transmission
  • Planetary gear unit 3000 will be described with reference to FIG. Planetary gear unit 3000 is connected to a torque converter 2100 having an input shaft 2102 coupled to the crankshaft.
  • the planetary gear unit 3000 includes a front planetary 3100, a rear planetary 3200, a C1 clutch 3301, a C2 clutch 3302, a C3 clutch 3303, a C4 clutch 3304, a B1 brake 3311, a B2 brake 3312, and a one-way clutch (F). 3320.
  • the front planetary 3100 is a double pinion type planetary gear mechanism.
  • Front planetary 3100 includes a first sun gear (S1) 3102, a pair of first pinion gears (P1) 3104, a carrier (CA) 3106, and a ring gear (R) 3108.
  • the first pinion gear (P1) 3104 meshes with the first sun gear (S1) 3102 and the first ring gear (R) 3108.
  • the first carrier (CA) 3106 supports the first pinion gear (P1) 3104 so that it can revolve and rotate.
  • the first sun gear (S1) 3102 is fixed to the gear case 3400 so as not to rotate.
  • First carrier (CA) 3106 is coupled to input shaft 3002 of planetary gear unit 3000.
  • the rear planetary 3200 is a Ravigneaux type planetary gear mechanism.
  • the rear planetary 3200 includes a second sun gear (S2) 3202, a second pinion gear (P2) 3204, a rear carrier (RCA) 3206, a rear ring gear (RR) 3208, a third sun gear (S3) 3210, a third Pinion gear (P3) 3212.
  • the second pinion gear (P2) 3204 meshes with the second sun gear (S2) 3202, the rear ring gear (RR) 3208, and the third pinion gear (P3) 3212.
  • Third pinion gear (P3) 3212 meshes with third sun gear (S3) 3210 in addition to second pinion gear (P2) 3204.
  • the rear carrier (RCA) 3206 supports the second pinion gear (P2) 3204 and the third pinion gear (P3) 3212 so that they can revolve and rotate.
  • Rear carrier (RCA) 3206 is coupled to one-way clutch (F) 3320.
  • the rear carrier (RCA) 3206 becomes non-rotatable when driving the first gear (when traveling using the driving force output from the engine 1000).
  • Rear ring gear (RR) 3208 is coupled to output shaft 3004 of planetary gear unit 3000.
  • the one-way clutch (F) 3320 is provided in parallel with the B2 brake 3312. That is, the outer race of the one-way clutch (F) 3320 is fixed to the gear case 3400, and the inner race is connected to the rear carrier (RCA) 3206.
  • FIG. 3 shows an operation table showing the relationship between each gear position and the operation state of each clutch and each brake. By operating the brakes and the clutches in the combinations shown in this operation table, forward 1st to 8th gears and reverse 1st and 2nd gears are formed.
  • the forward third speed gear stage is formed by engaging the C1 clutch 3301 and the C3 clutch 3303 and releasing the other clutches and brakes.
  • the main part of the hydraulic circuit 4000 will be described with reference to FIG.
  • the hydraulic circuit 4000 is not limited to the one described below.
  • the hydraulic circuit 4000 includes an oil pump 4004, a primary regulator valve 4006, a manual valve 4100, a solenoid modulator valve 4200, an SL1 linear solenoid (hereinafter referred to as SL (1)) 4210, and an SL2 linear solenoid (hereinafter referred to as “the solenoid valve”).
  • SL2 (described as SL (4)) 4220, SL3 linear solenoid (hereinafter referred to as SL (3)) 4230, SL4 linear solenoid (hereinafter referred to as SL (4)) 4240, and SL5 linear solenoid (hereinafter referred to as SL (3)).
  • SL (5)) 4250 SLT linear solenoid (hereinafter referred to as SLT) 4300, and B2 control valve 4500.
  • Oil pump 4004 is connected to the crankshaft of engine 1000. As the crankshaft rotates, the oil pump 4004 is driven to generate hydraulic pressure.
  • the hydraulic pressure generated by the oil pump 4004 is regulated by the primary regulator valve 4006 to generate a line pressure.
  • Primary regulator valve 4006 operates using the throttle pressure regulated by SLT 4300 as a pilot pressure.
  • the line pressure is supplied to the manual valve 4100 via the line pressure oil passage 4010.
  • Manual valve 4100 includes a drain port 4105. From the drain port 4105, the oil pressure in the D range pressure oil passage 4102 and the R range pressure oil passage 4104 is discharged.
  • both the D range pressure oil passage 4102 and the R range pressure oil passage 4104 are connected to the drain port 4105, and the D range pressure and R of the D range pressure oil passage 4102 are communicated.
  • the R range pressure of the range pressure oil passage 4104 is discharged from the drain port 4105.
  • the B2 control valve 4500 selectively supplies hydraulic pressure from one of the D range pressure oil passage 4102 and the R range pressure oil passage 4104 to the B2 brake 3312.
  • a D range pressure oil passage 4102 and an R range pressure oil passage 4104 are connected to the B2 control valve 4500.
  • the B2 control valve 4500 is controlled by the hydraulic pressure supplied from the SLU solenoid valve (not shown) and the urging force of the spring.
  • the B2 control valve 4500 When the SLU solenoid valve is on, the B2 control valve 4500 is in the state on the left side in FIG.
  • the B2 brake 3312 is supplied with the hydraulic pressure adjusted from the D range pressure using the hydraulic pressure supplied from the SLU solenoid valve as a pilot pressure.
  • the hydraulic pressure supplied to the D range pressure oil passage 4102 is finally supplied to the C1 clutch 3301, the C2 clutch 3302, and the C3 clutch 3303.
  • the B2 control valve 4500 When the SLU solenoid valve is off, the B2 control valve 4500 is in the state on the right side in FIG. In this case, the R range pressure of the R range pressure oil passage 4104 is supplied to the B2 brake 3312.
  • SL (1) 4210, SL (2) 4220, SL (3) 4230, SL (4) 4240, and SL (5) 4250 are normally closed linear solenoid valves.
  • the output hydraulic pressure POUT of these linear solenoid valves is controlled by a drive current I output from the ECU 8000 to each linear solenoid valve as a drive signal.
  • the output hydraulic pressure of these linear solenoid valves becomes minimum (“0”) when not energized, and the output hydraulic pressure increases as each drive current I from ECU 8000 increases.
  • the output hydraulic pressure POUT (C1) of SL (1) 4210 is supplied to the C1 clutch 3301, the output hydraulic pressure POUT (C2) of SL (2) 4220 is supplied to the C2 clutch 3302, and the output hydraulic pressure of SL (3) 4230.
  • POUT (C3) is supplied to the C3 clutch 3303, the output hydraulic pressure POUT (C4) of the SL (4) 4240 is supplied to the C4 clutch 3304, and the output hydraulic pressure POUT (B1) of the SL (5) 4250 is supplied to the B1 brake. 3311.
  • ECU 8000 Based on input torque (turbine torque) and the like, ECU 8000 has target command pressure P (C1) of output hydraulic pressure POUT (C1), target command pressure P (C2) of output hydraulic pressure POUT (C2), and output hydraulic pressure POUT (C3).
  • the method for calculating the target command pressure will be described in detail later.
  • the ECU 8000 includes a drive current I (C1) corresponding to the target command pressure P (C1), a drive current I (C2) corresponding to the target command pressure P (C2), and a drive current I corresponding to the target command pressure P (C3).
  • (C3), drive current I (C4) corresponding to the target command pressure P (C4), and drive current I (B1) corresponding to the target command pressure P (B1) are SL (1) 4210 and SL (2), respectively. 4220, SL (3) 4230, SL (4) 4240, and SL (5) 4250.
  • Solenoid modulator valve 4200 adjusts the hydraulic pressure (solenoid modulator pressure) supplied to SLT 4300 to a constant pressure using the line pressure as the original pressure.
  • the SLT 4300 adjusts the solenoid modulator pressure in accordance with the drive current I (T) from the ECU 8000 based on the accelerator opening degree ACC detected by the accelerator opening degree sensor 8010, and generates a throttle pressure.
  • the throttle pressure is supplied to the primary regulator valve 4006 via the SLT oil passage 4302.
  • the throttle pressure is used as a pilot pressure for the primary regulator valve 4006.
  • SL (1) 4210 the structure and characteristics of SL (1) 4210 will be described. Note that SL (2) 4220, SL (3) 4230, SL (4) 4240, SL (5) 4250, and SLT 4300 have the same structure and characteristics as SL (1) 4210. The description of will not be repeated.
  • FIG. 5 shows the internal state of SL (1) 4210 when drive current I (C1) from ECU 8000 is maximum value I (MAX).
  • SL (1) 4210 utilizes the phenomenon that the force by which the spool 4212 pushes the valve 4213 increases in proportion to the drive current I (C1) flowing through the coil 4211, and the stroke of the spool 4212 is driven by the drive current I (C1).
  • the flow rate cross section of the input port 4217 is changed by adjusting the amount.
  • the output hydraulic pressure POUT (C1) in which the line pressure PL is regulated is output from the output port 4218.
  • the stroke amount of the spool 4212 increases substantially in proportion to the drive current I (C1).
  • the drive current I (C1) reaches the maximum current value I (MAX)
  • the force by which the spool 4212 pushes the valve 4213 becomes maximum, and the spring 4214 side of the valve 4213 is against the biasing force of the spring 4214 as shown in FIG. Is brought into contact with the stopper 4215 (full stroke state).
  • FIG. 6 shows the relationship between the target command pressure P (C1) and the drive current I (C1).
  • the drive current I (C1) becomes 0 (minimum current value) when the target command pressure P (C1) is 0 (minimum pressure), and increases substantially in proportion to the target command pressure P (C1).
  • the drive current I (C1) when the target command pressure P (C1) is the line pressure PL is I (PL).
  • the drive current I (C1) becomes the maximum current value I (MAX).
  • the relationship between the target command pressure P (C1) and the drive current I (C1) is not limited to that shown in FIG.
  • the drive current I (C1) may be set to the maximum current value I (MAX).
  • FIG. 7 shows the relationship between the drive current I (C1) and the output hydraulic pressure POUT (C1). As shown in FIG. 7, in the range where the drive current I (C1) is smaller than I (PL), the output hydraulic pressure POUT (C1) increases substantially in proportion to the drive current I (C1). At this time, SL (1) 4210 is in a pressure regulation state in which the line pressure PL can be regulated according to the drive current I (C1).
  • the target command pressure P (C1) is set to the maximum pressure P (MAX), and the drive current I (C1) is set to the maximum current value I (MAX). (See FIG. 6).
  • the target command pressure P of the output hydraulic pressure of the linear solenoid valve corresponding to the friction engagement element (clutch or brake) to be maintained in the engaged state can be maintained in the engaged state and is maximum.
  • FIG. 8 shows a functional block diagram of the ECU 8000 which is a vehicle control apparatus according to this embodiment.
  • ECU 8000 includes an input interface (hereinafter referred to as input I / F) 8100, a calculation processing unit 8200, a storage unit 8300, and an output interface (hereinafter referred to as output I / F) 8400.
  • input I / F input interface
  • calculation processing unit 8200 calculation processing unit
  • storage unit 8300 a storage unit 8300
  • output I / F output interface
  • the input I / F 8100 includes a vehicle speed V from the vehicle speed sensor 8002, an accelerator opening ACC from the accelerator opening sensor 8010, a shift position SP from the position switch 8006, an engine speed NE from the engine speed sensor 8020, and an input shaft rotation.
  • the input shaft rotational speed NT from the number sensor 8022, the output shaft rotational speed NOUT from the output shaft rotational speed sensor 8024, and the throttle opening TH from the throttle opening sensor 8018 are received and transmitted to the arithmetic processing unit 8200.
  • the storage unit 8300 stores various types of information, programs, threshold values, maps, and the like, and data is read from or stored in the arithmetic processing unit 8200 as necessary.
  • the arithmetic processing unit 8200 includes a shift control determination unit 8210, a necessary hydraulic pressure calculation unit 8220, an engagement hydraulic pressure control unit 8230, and a line pressure control unit 8240.
  • the shift control determination unit 8210 determines whether or not a non-shift is in progress.
  • non-shifting means a state in which shift control is not performed.
  • the shift control determination unit 8210 determines that a non-shift is in progress when the shift determination based on the vehicle speed V, the accelerator opening degree ACC, and the shift diagram has not been performed and there is no ongoing shift.
  • the required hydraulic pressure calculation unit 8220 determines friction engagement elements that should be maintained in the engaged state during non-shifting based on the operation table of FIG. 3 described above, and maintains these friction engagement elements in the engaged state.
  • Required hydraulic pressure PR (required hydraulic pressure PR (C1) of the C1 clutch 3301, required hydraulic pressure PR (C2) of the C2 clutch 3302, required hydraulic pressure PR (C3) of the C3 clutch 3303, required hydraulic pressure PR (C4) of the C4 clutch 3304, B1
  • the required hydraulic pressure PR (B1) of the brake 3311) is calculated.
  • the required oil pressure calculation unit 8220 calculates the minimum pressure required to maintain each friction engagement element in the engaged state as (input torque) ⁇ (torque sharing ratio), and adds a safety factor (for example, 1.2). A value obtained by multiplying is calculated as the required hydraulic pressure PR. In this way, the required hydraulic pressure PR is calculated to a value slightly higher (larger by the safety factor) than the minimum pressure required to maintain each friction engagement element in the engaged state. The value is smaller than the maximum pressure P (MAX). The required hydraulic pressure PR is calculated for each friction engagement element to be maintained in the engaged state.
  • a safety factor for example, 1.2
  • the input torque is a torque (turbine torque) input from the engine 1000 to the input shaft 3002 of the planetary gear unit 3000 via the torque converter 2100.
  • the input torque is calculated based on, for example, the engine speed NE, the accelerator opening ACC, or the throttle opening TH.
  • the torque sharing ratio is a ratio that the input torque is shared by each friction engagement element maintained in the engaged state, and is set in advance according to the combination of the friction engagement elements maintained in the engaged state. Is done.
  • the engagement hydraulic pressure control unit 8230 has target command pressures P (P (C1), P (C2), P (C3), P (C4), P according to each required hydraulic pressure PR calculated by the required hydraulic pressure calculation unit 8220. (At least one of (B1)) is calculated, and the drive current I (I (C1), I (C2), I (C3), I (C4), I (B1) is determined according to the calculated target command pressure P. And outputs the calculated drive current I to each linear solenoid valve via the output I / F 8400. The drive current I is not output to each linear solenoid valve corresponding to the released frictional engagement element.
  • the line pressure control unit 8240 generates a drive current that generates a throttle pressure (a pilot pressure of the primary regulator valve 4006) at which the line pressure PL is higher than the required hydraulic pressure PR calculated by the required hydraulic pressure calculation unit 8220 by a predetermined value ⁇ .
  • I (T) is output to the SLT 4300 via the output I / F 8400.
  • the shift control determination unit 8210, the required hydraulic pressure calculation unit 8220, the engagement hydraulic pressure control unit 8230, and the line pressure control unit 8240 are all calculated by the CPU that is the arithmetic processing unit 8200 by the storage unit 8300.
  • the program functions as software that is realized by executing the program stored in the program, it may be realized by hardware. Such a program is recorded on a storage medium and mounted on the vehicle.
  • step (hereinafter, step is abbreviated as S) 100 ECU 8000 determines whether or not a non-shifting operation is being performed. If not (YES at S100), the process proceeds to S102. Otherwise (NO in S100), this process ends.
  • step S102 the ECU 8000 calculates a required hydraulic pressure PR for each friction engagement element to be maintained in the engaged state. As described above, the ECU 8000 calculates the value calculated by the input torque ⁇ the torque sharing ratio ⁇ the safety factor as the required hydraulic pressure PR for each friction engagement element.
  • the ECU 8000 uses the required hydraulic pressure PR (C1) of the C1 clutch 3301 as an input torque.
  • X torque sharing ratio of C1 clutch 3301)
  • necessary hydraulic pressure PR (C3) of C3 clutch 3303 is calculated as input torque x (torque sharing ratio of C3 clutch 3303) x safety coefficient.
  • ECU 8000 outputs to SLT 4300 drive current I (T) in which line pressure PL is higher by a predetermined value ⁇ than the maximum pressure of required hydraulic pressure PR calculated in S102.
  • T drive current
  • the required hydraulic pressure PR (C1) of the C1 clutch 3301 is the maximum pressure (that is, higher than the required hydraulic pressure PR (C3))
  • the line pressure PL is required.
  • the hydraulic pressure is controlled to PR (C1) + ⁇ .
  • ECU 8000 calculates the maximum pressure of necessary hydraulic pressure PR calculated in the process of S102 as target command pressure P for each friction engagement element to be maintained in the engaged state. For example, when the required hydraulic pressure PR (C1) is the maximum pressure during traveling at the third forward gear, ECU 8000 uses the required hydraulic pressure PR (C1) as the target command pressure P (C1) and the target command pressure P (C3). ).
  • ECU 8000 calculates drive current I based on calculated target command pressure P. For example, ECU 8000 calculates drive current I (C1) and drive current I (C3) corresponding to target command pressure P (C1) based on the above-described relationship of FIG. 6 when traveling at the third forward gear. calculate.
  • ECU 8000 outputs each calculated drive current I to each linear solenoid valve. For example, ECU 8000 outputs drive current I (C1) to SL (1) 4210 and drive current I (C3) to SL (3) 4230 when traveling at the third forward gear.
  • the target command pressure P controlled by the ECU 8000 which is the control device according to the present embodiment based on the structure and flowchart as described above, will be described with reference to FIG.
  • both the required oil pressure PR (C1) and the required oil pressure PR (C3) are smaller than the maximum pressure P (MAX) of the target command pressure P.
  • the target command pressure P (C1) and the target command pressure P (C3) are both the required hydraulic pressure PR (C1) (S106). That is, as shown in FIG. 10, both the target command pressure P (C1) and the target command pressure P (C3) are set to values smaller than the maximum pressure P (MAX).
  • a drive current I (C1) smaller than the maximum current value I (MAX) is output to SL (1) 4210, and a drive current I (C3) smaller than the maximum current value I (MAX) is SL (3). ) 4230 (S108, S110).
  • the line pressure PL is controlled to be higher than the required hydraulic pressure PR (C1) by a predetermined value ⁇ (S104). Therefore, even if the line pressure PL pulsates or varies slightly with respect to the command, the target command pressure P (1) and the target command pressure P (3) are unlikely to exceed the line pressure PL, and the drive current I Is suppressed to the maximum current value I (MAX) (see FIGS. 6 and 7).
  • the C1 clutch 3301 and the C3 clutch 3303 can be maintained in the pressure regulation state instead of the non-pressure regulation state. As a result, the control responsiveness of the C1 clutch 3301 and the C3 clutch 3303 during the subsequent shift can be improved.
  • the spools of the SL (1) 4210 and the SL (3) 4230 are controlled to the pressure adjustable positions.
  • the time required to move the spool to the pressure adjustable position is not required, and the shift time can be shortened accordingly.
  • SL (1) 4210 and SL (3) 4230 output drive currents I (C1) and I (C3) smaller than the maximum current value I (MAX). Therefore, at the time of the subsequent shift, there is no need for time to decrease the drive current I from the maximum current value I (MAX) to the current value that can be regulated (that is, I (PL)). Time can be shortened.
  • each friction engagement element when the friction engagement elements are maintained in the engaged state by the normally closed solenoid valve, each friction engagement element can be maintained in the engaged state.
  • the drive current I corresponding to the hydraulic pressure smaller than the maximum value P (MAX) of the target command pressure is output to each linear solenoid valve. Therefore, the drive current I supplied to each linear solenoid valve is smaller than the maximum current value I (MAX). Thereby, the electric power consumed by the linear solenoid valve for maintaining the engaged state can be reduced while maintaining the engaged state of the friction engagement element.
  • the line pressure PL is set to a value higher than the maximum pressure of the required hydraulic pressure PR by a predetermined value ⁇ in the process of S104 in FIG. 9, but the line pressure PL is set to the required hydraulic pressure PR.
  • the maximum pressure itself may be set to a value. Even in this case, the power consumed by the linear solenoid valve can be reduced.
  • the control of the linear solenoid valve corresponding to the friction engagement element that should be maintained in the engaged state during non-shifting has been described, but the present invention is not limited to being non-shifting.
  • the present invention is used to control a solenoid valve that supplies hydraulic pressure to a friction engagement element (for example, a C1 clutch 3301 in a shift performed between the first speed to the fifth speed) to be kept engaged even during a shift. May be applied.
  • FIG. 11 a control structure of a program executed by ECU 8000 which is a control device according to this modification will be described.
  • ECU 8000 which is a control device according to this modification.
  • the same steps as those in the flowchart shown in FIG. 9 are given the same step numbers. The processing for them is the same. Therefore, detailed description thereof will not be repeated here.
  • ECU 8000 calculates each target command pressure P of the linear solenoid valve corresponding to the frictional engagement element to be maintained in the engaged state in correspondence with the required hydraulic pressure PR calculated in the process of S102. For example, when the required hydraulic pressure PR (C1) and the required hydraulic pressure PR (C3) are calculated in the process of S102 during traveling at the third forward gear, ECU 8000 uses the required hydraulic pressure PR (C1) as the target command pressure P. (C1), and the required oil pressure PR (C3) is set as the target command pressure P (C3).
  • ECU 8000 calculates drive current I for each linear solenoid for each calculated target command pressure P. For example, ECU 8000 provides drive current I (C1) corresponding to target command pressure P (C1) and drive current I (C3) corresponding to target command pressure P (C3) during traveling at the third forward gear. , Based on the relationship of FIG. 6 described above.
  • the target command pressure P (C3) can be set to a required oil pressure PR (C3) lower than the required oil pressure PR (C1) (S200). That is, the drive current I (C3) is further reduced as compared with the case where the target command pressure P (C1) and the target command pressure P (C3) are both set to the required hydraulic pressure PR (C1) as in the above-described embodiment. Can do. Therefore, the power consumed by SL (3) 4230 can be further reduced.

Abstract

In case shifting frictional engagement elements are kept in engaging states by the output oil pressure of a normal close type linear solenoid valve, an ECU calculates (at S102) a necessary oil pressure (PR) for each frictional engagement element kept in the engaging state. The ECU calculates (at S106) the maximum pressure of the necessary oil pressures (PR) as a target command pressure (P) for each frictional engagement element, calculates (at S108) a drive current (I) on the basis of the target command pressure (P) calculated, and outputs (at S110) the calculated drive current (I) to each linear solenoid valve. The ECU calculates the necessary oil pressure (PR) as a value larger than the minimum pressure necessary for keeping each frictional engagement element in the engaging state and smaller than the maximum pressure (P(MAX)) of the target command pressure (P).

Description

自動変速機の制御装置および制御方法Control device and control method for automatic transmission
　本発明は、車両用の自動変速機の制御に関し、特に、リニアソレノイド弁で変速用の摩擦係合要素を係合状態に維持する自動変速機の制御に関する。 The present invention relates to control of an automatic transmission for a vehicle, and more particularly to control of an automatic transmission that maintains a frictional engagement element for shifting with a linear solenoid valve.
　電子制御装置からの駆動信号（駆動電流）によって制御されるリニアソレノイド弁の出力油圧を油圧式の摩擦係合要素に供給することによって、所定の変速段を形成する自動変速機の変速制御装置がある。このような制御装置において、摩擦係合要素を係合状態に維持する場合の油圧制御の一例が、たとえば特開２００３－４２２８０号公報（特許文献１）に開示されている。 A shift control device for an automatic transmission that forms a predetermined shift stage by supplying output hydraulic pressure of a linear solenoid valve controlled by a drive signal (drive current) from an electronic control device to a hydraulic friction engagement element. is there. In such a control device, an example of hydraulic control in the case where the friction engagement element is maintained in the engaged state is disclosed in, for example, Japanese Patent Application Laid-Open No. 2003-42280 (Patent Document 1).
　特開２００３－４２２８０号公報に開示された変速用電子制御装置は、ノーマルオープン型（非通電時に出力油圧が最大になり、通電量の増大に伴なって出力油圧が減少する型）のリニアソレノイド弁と、リニアソレノイド弁の出力油圧が供給されるクラッチとを備えた自動変速機において、クラッチを係合状態に維持する際、リニアソレノイド弁の出力油圧を最大にするためにリニアソレノイド弁の駆動電流を０（駆動信号のデューティ比を０％）に設定している。
特開２００３－４２２８０号公報 Japanese Patent Application Laid-Open No. 2003-42280 discloses an electronic control device for shifting which is a normally open type (a type in which the output hydraulic pressure is maximized when no power is supplied and the output hydraulic pressure decreases as the amount of current supplied increases). In an automatic transmission having a valve and a clutch to which the output hydraulic pressure of the linear solenoid valve is supplied, the linear solenoid valve is driven to maximize the output hydraulic pressure of the linear solenoid valve when the clutch is maintained in an engaged state. The current is set to 0 (the duty ratio of the drive signal is 0%).
Japanese Patent Laid-Open No. 2003-42280
　ところで、ノーマルクローズ型（非通電時に出力油圧が最小になり、通電量の増大に伴ない出力油圧が増大する型）のリニアソレノイド弁の出力油圧で摩擦係合要素を係合状態に維持する際、リニアソレノイド弁の出力油圧を最大にするためにリニアソレノイド弁の駆動電流を最大値に設定すると、駆動電流の最大値に応じた多量の電力が、摩擦係合要素を係合状態に維持する間継続してリニアソレノイド弁で消費されてしまい、最終的には燃料消費率が悪化してしまうという問題がある。 By the way, when the friction engagement element is maintained in the engaged state by the output hydraulic pressure of the linear solenoid valve of the normally closed type (the output hydraulic pressure is minimized when the power is not supplied and the output hydraulic pressure increases as the amount of current is increased). When the drive current of the linear solenoid valve is set to the maximum value in order to maximize the output hydraulic pressure of the linear solenoid valve, a large amount of electric power corresponding to the maximum value of the drive current maintains the friction engagement element in the engaged state. There is a problem that the linear solenoid valve is continuously consumed for a long time, and the fuel consumption rate eventually deteriorates.
　本発明は、上述の課題を解決するためになされたものであって、その目的は、油圧式の変速用摩擦係合要素と、駆動電流の増大に伴なって出力油圧を増大させて摩擦係合要素に出力する調圧弁とを備えた自動変速機において、摩擦係合要素を係合した状態に維持する場合の調圧弁の消費電力を低減することができる制御装置および制御方法を提供することである。 The present invention has been made to solve the above-described problems, and an object of the present invention is to provide a hydraulic shift friction engagement element and a friction engagement element by increasing an output hydraulic pressure with an increase in driving current. Provided is a control device and a control method capable of reducing power consumption of a pressure regulating valve when maintaining a friction engagement element in an engaged state in an automatic transmission having a pressure regulating valve for outputting to a combined element. It is.
　この発明に係る制御装置は、油圧式の変速用摩擦係合部と、駆動電流の増大に伴なって出力油圧を増大させて摩擦係合部に出力する調圧部とを備えた自動変速機を制御する。この制御装置は、摩擦係合部を係合した状態に維持する必要がある場合、摩擦係合部を係合した状態に維持可能な油圧であってかつ出力油圧の目標指令圧の最大値よりも小さい油圧を算出する算出部と、算出部によって算出された油圧に応じた駆動電流を調圧部に出力して、調圧部を制御する調圧制御部とを含む。 The control device according to the present invention includes an automatic transmission including a hydraulic shift friction engagement portion and a pressure adjusting portion that increases an output hydraulic pressure as the drive current increases and outputs the output hydraulic pressure to the friction engagement portion. To control. When it is necessary to maintain the friction engagement portion in an engaged state, this control device is a hydraulic pressure that can be maintained in a state in which the friction engagement portion is engaged, and from the maximum value of the target command pressure of the output hydraulic pressure. A calculation unit that calculates a small hydraulic pressure, and a pressure regulation control unit that outputs a drive current corresponding to the hydraulic pressure calculated by the calculation unit to the pressure regulation unit to control the pressure regulation unit.
　好ましくは、摩擦係合部は、複数の摩擦係合要素を含む。調圧部は、複数の摩擦係合要素にそれぞれ対応する複数の調圧弁を含む。算出部は、複数の摩擦係合要素のうち少なくとも２つ以上の摩擦係合要素を係合した状態に維持する必要がある場合、係合した状態に維持する必要がある２つ以上の摩擦係合要素ごとに油圧を算出する。調圧制御部は、算出された油圧のうちの最大圧に応じた駆動電流を、係合した状態に維持する必要がある２つ以上の摩擦係合要素に対応する調圧弁に出力する。 Preferably, the friction engagement portion includes a plurality of friction engagement elements. The pressure regulating unit includes a plurality of pressure regulating valves respectively corresponding to the plurality of friction engagement elements. When the calculation unit needs to maintain at least two friction engagement elements among the plurality of friction engagement elements in an engaged state, the calculation unit needs to maintain two or more friction engagement elements that need to be maintained in the engagement state. The hydraulic pressure is calculated for each factor. The pressure adjustment control unit outputs a drive current corresponding to the maximum pressure of the calculated hydraulic pressure to a pressure adjustment valve corresponding to two or more friction engagement elements that need to be maintained in an engaged state.
　さらに好ましくは、摩擦係合部は、複数の摩擦係合要素を含む。調圧部は、複数の摩擦係合要素にそれぞれ対応する複数の調圧弁を含む。算出部は、複数の摩擦係合要素のうち少なくとも２つ以上の摩擦係合要素を係合した状態に維持する必要がある場合、係合した状態に維持する必要がある２つ以上の摩擦係合要素ごとに油圧を算出する。調圧制御部は、算出された油圧に対応するそれぞれの駆動電流を、算出された油圧に対応するそれぞれの調圧弁に出力する。 More preferably, the friction engagement portion includes a plurality of friction engagement elements. The pressure regulating unit includes a plurality of pressure regulating valves respectively corresponding to the plurality of friction engagement elements. When the calculation unit needs to maintain at least two friction engagement elements among the plurality of friction engagement elements in an engaged state, the calculation unit needs to maintain two or more friction engagement elements that need to be maintained in the engagement state. The hydraulic pressure is calculated for each factor. The pressure regulation control unit outputs each drive current corresponding to the calculated hydraulic pressure to each pressure regulating valve corresponding to the calculated hydraulic pressure.
　さらに好ましくは、自動変速機には、調圧部に入力される元圧を調圧する元圧調圧弁が備えられる。制御装置は、算出された油圧に応じた駆動電流が調圧部に調圧制御部によって出力された場合、元圧が算出された油圧よりも予め定められた値だけ大きくなるように、元圧調圧弁を制御する元圧制御部をさらに含む。 More preferably, the automatic transmission is provided with a source pressure regulating valve that regulates the source pressure input to the pressure regulating unit. When a drive current corresponding to the calculated hydraulic pressure is output to the pressure adjustment unit by the pressure adjustment control unit, the control device causes the original pressure to be larger than the calculated hydraulic pressure by a predetermined value. It further includes an original pressure control unit that controls the pressure regulating valve.
　本発明によれば、ノーマルクローズ型（駆動電流の増大に伴なって出力油圧が増大する型）の調圧弁（調圧部）で摩擦係合要素（摩擦係合部）を係合状態に維持する場合、係合状態に維持可能な油圧であってかつ出力油圧の目標指令圧の最大値よりも小さい油圧に応じた駆動電流が調圧弁に出力される。そのため、調圧弁に供給される駆動電流は、目標指令圧の最大値に応じた最大電流値よりも小さくなる。これにより、摩擦係合要素を係合した状態に維持しつつ、調圧弁の消費電力を低減することができる。 According to the present invention, the friction engagement element (friction engagement portion) is maintained in the engaged state by the normally closed type (type in which the output hydraulic pressure increases as the drive current increases). In this case, a drive current corresponding to the hydraulic pressure that can be maintained in the engaged state and is smaller than the maximum value of the target command pressure of the output hydraulic pressure is output to the pressure regulating valve. Therefore, the drive current supplied to the pressure regulating valve is smaller than the maximum current value corresponding to the maximum value of the target command pressure. Thereby, the power consumption of a pressure regulation valve can be reduced, maintaining the state which engaged the friction engagement element.
車両のパワートレーンを示す概略構成図である。It is a schematic block diagram which shows the power train of a vehicle. オートマチックトランスミッションのプラネタリギヤユニットを示すスケルトン図である。It is a skeleton figure which shows the planetary gear unit of an automatic transmission. オートマチックトランスミッションの作動表を示す図である。It is a figure which shows the operation | movement table | surface of an automatic transmission. オートマチックトランスミッションの油圧回路を示す図である。It is a figure which shows the hydraulic circuit of an automatic transmission. リニアソレノイドバルブの構造を示す図である。It is a figure which shows the structure of a linear solenoid valve. リニアソレノイドバルブの目標指令圧と駆動電流との関係を示す図である。It is a figure which shows the relationship between the target command pressure and drive current of a linear solenoid valve. リニアソレノイドバルブの駆動電流と出力油圧との関係を示す図である。It is a figure which shows the relationship between the drive current of a linear solenoid valve, and output hydraulic pressure. ＥＣＵの機能ブロック図である。It is a functional block diagram of ECU. ＥＣＵが実行するプログラムの制御構造を示す図（その１）である。It is FIG. (1) which shows the control structure of the program which ECU performs. ＥＣＵによって算出された目標指令圧のタイミングチャート（その１）である。It is a timing chart (the 1) of target command pressure computed by ECU. ＥＣＵが実行するプログラムの制御構造を示す図（その２）である。It is FIG. (2) which shows the control structure of the program which ECU performs. ＥＣＵによって算出された目標指令圧のタイミングチャート（その２）である。It is a timing chart (the 2) of the target command pressure computed by ECU.
符号の説明Explanation of symbols
　１０００　エンジン、１００４　補機、２０００　オートマチックトランスミッション、２１００　トルクコンバータ、２１０２　入力軸、３０００　プラネタリギヤユニット、３００２　入力軸、３００４　出力軸、３１００　フロントプラネタリ、３２００　リアプラネタリ、３３０１　Ｃ１クラッチ、３３０２　Ｃ２クラッチ、３３０３　Ｃ３クラッチ、３３０４　Ｃ４クラッチ、３３１１　Ｂ１ブレーキ、３３１２　Ｂ２ブレーキ、３４００　ギヤケース、４０００　油圧回路、５０００　プロペラシャフト、６０００　デファレンシャルギヤ、６１００　ドライブシャフト、７０００　後輪、８０００　ＥＣＵ、８００２　車速センサ、８００４　シフトレバー、８００６　ポジションスイッチ、８００８　アクセルペダル、８０１０　アクセル開度センサ、８０１２　ブレーキペダル、８０１４　踏力センサ、８０１６　電子スロットルバルブ、８０１８　スロットル開度センサ、８０２０　エンジン回転数センサ、８０２２　入力軸回転数センサ、８０２４　出力軸回転数センサ、８１００　入力Ｉ／Ｆ、８２００　演算処理部、８２１０　変速制御判断部、８２２０　必要油圧算出部、８２３０　係合油圧制御部、８２４０　ライン圧制御部、８３００　記憶部、８４００　出力Ｉ／Ｆ。 1000 engine, 1004 auxiliary machine, 2000 automatic transmission, 2100 torque converter, 2102 input shaft, 3000 planetary gear unit, 3002 input shaft, 3004 output shaft, 3100 front planetary, 3200 rear planetary, 3301 C1 clutch, 3302 C2 clutch, 3303 C3 clutch 3304 C4 clutch, 3311 B1 brake, 3312 B2 brake, 3400 gear case, 4000 hydraulic circuit, 5000 propeller shaft, 6000 differential gear, 6100 drive shaft, 7000 rear wheel, 8000 ECU, 8002, vehicle speed sensor, 8004 shift lever, 8006 position switch , 8008 Accelerator pedal, 8 10 accelerator opening sensor, 8012 brake pedal, 8014 pedal force sensor, 8016 electronic throttle valve, 8018 throttle opening sensor, 8020 engine speed sensor, 8022 input shaft speed sensor, 8024 output shaft speed sensor, 8100 input I / F , 8200 arithmetic processing unit, 8210 shift control determination unit, 8220 required hydraulic pressure calculation unit, 8230 engagement hydraulic pressure control unit, 8240 line pressure control unit, 8300 storage unit, 8400 output I / F.
　以下、図面を参照しつつ、本発明の実施例について説明する。以下の説明では、同一の部品には同一の符号を付してある。それらの名称および機能も同一である。したがって、それらについての詳細な説明は繰返さない。 Hereinafter, embodiments of the present invention will be described with reference to the drawings. In the following description, the same parts are denoted by the same reference numerals. Their names and functions are also the same. Therefore, detailed description thereof will not be repeated.
　図１を参照して、本発明の実施例に係る制御装置を搭載した車両について説明する。この車両は、ＦＲ（Ｆｒｏｎｔ　ｅｎｇｉｎｅ　Ｒｅａｒ　ｄｒｉｖｅ）車両である。なお、ＦＲ以外の車両であってもよい。 A vehicle equipped with a control device according to an embodiment of the present invention will be described with reference to FIG. This vehicle is a FR (Front engine Rear drive) vehicle. A vehicle other than FR may be used.
　車両は、エンジン１０００と、オートマチックトランスミッション２０００と、トルクコンバータ２１００と、オートマチックトランスミッション２０００の一部を構成するプラネタリギヤユニット３０００と、オートマチックトランスミッション２０００の一部を構成する油圧回路４０００と、プロペラシャフト５０００と、デファレンシャルギヤ６０００と、後輪７０００と、ＥＣＵ（Ｅｌｅｃｔｒｏｎｉｃ　Ｃｏｎｔｒｏｌ　Ｕｎｉｔ）８０００とを含む。本実施の形態に係る制御装置は、たとえばＥＣＵ８０００のＲＯＭ（Ｒｅａｄ　Ｏｎｌｙ　Ｍｅｍｏｒｙ）に記録されたプログラムを実行することによって実現される。 The vehicle includes an engine 1000, an automatic transmission 2000, a torque converter 2100, a planetary gear unit 3000 that forms part of the automatic transmission 2000, a hydraulic circuit 4000 that forms part of the automatic transmission 2000, a propeller shaft 5000, A differential gear 6000, a rear wheel 7000, and an ECU (Electronic Control Unit) 8000 are included. The control device according to the present embodiment is realized, for example, by executing a program recorded in a ROM (Read Only Memory) of ECU 8000.
　エンジン１０００は、インジェクタ（図示せず）から噴射された燃料と空気との混合気を、シリンダの燃焼室内で燃焼させる内燃機関である。燃焼によってシリンダ内のピストンが押し下げられて、クランクシャフトが回転させられる。エンジン１０００の駆動力によって、オルタネータおよびエアコンディショナーなどの補機１００４が駆動される。なお、エンジン１０００の代わりにもしくは加えて、動力源にモータを用いるようにしてもよい。 Engine 1000 is an internal combustion engine that burns a mixture of fuel and air injected from an injector (not shown) in a combustion chamber of a cylinder. The piston in the cylinder is pushed down by the combustion, and the crankshaft is rotated. An auxiliary machine 1004 such as an alternator and an air conditioner is driven by the driving force of the engine 1000. A motor may be used as a power source instead of or in addition to engine 1000.
　オートマチックトランスミッション２０００は、トルクコンバータ２１００を経由してエンジン１０００に連結される。オートマチックトランスミッション２０００は、所望のギヤ段を形成することによって、クランクシャフトの回転数を所望の回転数に変速する。 Automatic transmission 2000 is connected to engine 1000 via torque converter 2100. Automatic transmission 2000 shifts the rotational speed of the crankshaft to a desired rotational speed by forming a desired gear stage.
　オートマチックトランスミッション２０００から出力された駆動力は、プロペラシャフト５０００およびデファレンシャルギヤ６０００を経由して、左右の後輪７０００に伝達される。 The driving force output from the automatic transmission 2000 is transmitted to the left and right rear wheels 7000 via the propeller shaft 5000 and the differential gear 6000.
　ＥＣＵ８０００には、車速センサ８００２と、シフトレバー８００４のポジションスイッチ８００６と、アクセルペダル８００８のアクセル開度センサ８０１０と、ブレーキペダル８０１２の踏力センサ８０１４と、電子スロットルバルブ８０１６のスロットル開度センサ８０１８と、エンジン回転数センサ８０２０と、入力軸回転数センサ８０２２と、出力軸回転数センサ８０２４とがハーネスなどを経由して接続されている。 The ECU 8000 includes a vehicle speed sensor 8002, a position switch 8006 of a shift lever 8004, an accelerator opening sensor 8010 of an accelerator pedal 8008, a pedaling force sensor 8014 of a brake pedal 8012, a throttle opening sensor 8018 of an electronic throttle valve 8016, An engine speed sensor 8020, an input shaft speed sensor 8022, and an output shaft speed sensor 8024 are connected via a harness or the like.
　車速センサ８００２は、ドライブシャフト６１００の回転数から車速Ｖを検出する。ポジションスイッチ８００６は、シフトレバー８００４の位置（シフトポジション）ＳＰを検出する。アクセル開度センサ８０１０は、アクセルペダル８００８の開度（アクセル開度）ＡＣＣを検出する。踏力センサ８０１４は、ブレーキペダル８０１２の踏力（運転者がブレーキペダル８０１２を踏む力）を検出する。スロットル開度センサ８０１８は、電子スロットルバルブ８０１６の開度（スロットル開度）ＴＨを検出する。エンジン回転数センサ８０２０は、エンジン１０００のクランクシャフトの回転数（エンジン回転数）ＮＥを検出する。入力軸回転数センサ８０２２は、オートマチックトランスミッション２０００の入力軸回転数（トルクコンバータ２１００のタービン回転数）ＮＴを検出する。出力軸回転数センサ８０２４は、オートマチックトランスミッション２０００の出力軸の回転数（出力軸回転数）ＮＯＵＴを検出する。これらの各センサは、検出結果を表わす信号をＥＣＵ８０００に送信する。 The vehicle speed sensor 8002 detects the vehicle speed V from the rotational speed of the drive shaft 6100. The position switch 8006 detects the position (shift position) SP of the shift lever 8004. The accelerator opening sensor 8010 detects the opening (accelerator opening) ACC of the accelerator pedal 8008. The pedaling force sensor 8014 detects the pedaling force of the brake pedal 8012 (force that the driver steps on the brake pedal 8012). The throttle opening sensor 8018 detects the opening (throttle opening) TH of the electronic throttle valve 8016. The engine speed sensor 8020 detects a crankshaft speed (engine speed) NE of the engine 1000. Input shaft rotational speed sensor 8022 detects input shaft rotational speed (turbine rotational speed of torque converter 2100) NT of automatic transmission 2000. Output shaft rotational speed sensor 8024 detects the rotational speed (output shaft rotational speed) NOUT of the output shaft of automatic transmission 2000. Each of these sensors transmits a signal representing the detection result to ECU 8000.
　ＥＣＵ８０００は、車速センサ８００２、ポジションスイッチ８００６、アクセル開度センサ８０１０、踏力センサ８０１４、スロットル開度センサ８０１８、エンジン回転数センサ８０２０、入力軸回転数センサ８０２２、出力軸回転数センサ８０２４などから送られてきた信号、ＲＯＭに記憶されたマップおよびプログラムに基づいて、車両が所望の走行状態となるように、機器類を制御する。 ECU 8000 is sent from vehicle speed sensor 8002, position switch 8006, accelerator opening sensor 8010, pedal effort sensor 8014, throttle opening sensor 8018, engine speed sensor 8020, input shaft speed sensor 8022, output shaft speed sensor 8024, and the like. Based on the received signal, the map stored in the ROM, and the program, the devices are controlled so that the vehicle is in a desired running state.
　本実施例において、ＥＣＵ８０００は、シフトレバー８００４がＤ（ドライブ）ポジションであることによって、オートマチックトランスミッション２０００のシフトレンジにＤ（ドライブ）レンジが選択された場合、前進１速～８速ギヤ段のうちのいずれかのギヤ段が形成されるように、オートマチックトランスミッション２０００を制御する。前進１速～８速ギヤ段のうちのいずれかのギヤ段が形成されることによって、オートマチックトランスミッション２０００は後輪７０００に駆動力を伝達し得る。なお、Ｄレンジにおいて、８速ギヤ段よりも高速のギヤ段を形成可能であるようにしてもよい。形成するギヤ段は、車速とアクセル開度とをパラメータとして実験等によって予め作成された変速線図に基づいて決定される。 In the present embodiment, the ECU 8000, when the shift lever 8004 is in the D (drive) position and the D (drive) range is selected as the shift range of the automatic transmission 2000, out of the first to eighth forward gears. The automatic transmission 2000 is controlled so that any one of the gear positions is formed. The automatic transmission 2000 can transmit the driving force to the rear wheels 7000 by forming any one of the first to eighth forward gears. In the D range, it may be possible to form a higher gear than the eighth gear. The gear stage to be formed is determined based on a shift diagram created in advance by experiments or the like using the vehicle speed and the accelerator opening as parameters.
　なお、本実施例においては、ＥＣＵ８０００が１つのユニットとして説明するが、ＥＣＵ８０００を２つ以上のユニットに分割するようにしてもよい。たとえば、ＥＣＵ８０００がエンジン１０００を制御するエンジンＥＣＵと、オートマチックトランスミッション２０００を制御するＥＣＴ（Ｅｌｅｃｔｒｏｎｉｃ　Ｃｏｎｔｒｏｌｌｅｄ　Ｔｒａｎｓｍｉｓｓｉｏｎ）＿ＥＣＵとを含むようにし、エンジンＥＣＵとＥＣＴ＿ＥＣＵとが互いに信号を送受信可能であるように構成するようにしてもよい。 In this embodiment, the ECU 8000 is described as one unit, but the ECU 8000 may be divided into two or more units. For example, the ECU 8000 includes an engine ECU that controls the engine 1000 and an ECT (Electronic Controlled Transmission) _ECU that controls the automatic transmission 2000 so that the engine ECU and the ECT_ECU can transmit and receive signals to and from each other. It may be.
　図２を参照して、プラネタリギヤユニット３０００について説明する。プラネタリギヤユニット３０００は、クランクシャフトに連結された入力軸２１０２を有するトルクコンバータ２１００に接続されている。 The planetary gear unit 3000 will be described with reference to FIG. Planetary gear unit 3000 is connected to a torque converter 2100 having an input shaft 2102 coupled to the crankshaft.
　プラネタリギヤユニット３０００は、フロントプラネタリ３１００と、リアプラネタリ３２００と、Ｃ１クラッチ３３０１と、Ｃ２クラッチ３３０２と、Ｃ３クラッチ３３０３と、Ｃ４クラッチ３３０４と、Ｂ１ブレーキ３３１１と、Ｂ２ブレーキ３３１２と、ワンウェイクラッチ（Ｆ）３３２０とを含む。 The planetary gear unit 3000 includes a front planetary 3100, a rear planetary 3200, a C1 clutch 3301, a C2 clutch 3302, a C3 clutch 3303, a C4 clutch 3304, a B1 brake 3311, a B2 brake 3312, and a one-way clutch (F). 3320.
　フロントプラネタリ３１００は、ダブルピニオン型の遊星歯車機構である。フロントプラネタリ３１００は、第１サンギヤ（Ｓ１）３１０２と、１対の第１ピニオンギヤ（Ｐ１）３１０４と、キャリア（ＣＡ）３１０６と、リングギヤ（Ｒ）３１０８とを含む。 The front planetary 3100 is a double pinion type planetary gear mechanism. Front planetary 3100 includes a first sun gear (S1) 3102, a pair of first pinion gears (P1) 3104, a carrier (CA) 3106, and a ring gear (R) 3108.
　第１ピニオンギヤ（Ｐ１）３１０４は、第１サンギヤ（Ｓ１）３１０２および第１リングギヤ（Ｒ）３１０８と噛合っている。第１キャリア（ＣＡ）３１０６は、第１ピニオンギヤ（Ｐ１）３１０４が公転および自転可能であるように支持している。 The first pinion gear (P1) 3104 meshes with the first sun gear (S1) 3102 and the first ring gear (R) 3108. The first carrier (CA) 3106 supports the first pinion gear (P1) 3104 so that it can revolve and rotate.
　第１サンギヤ（Ｓ１）３１０２は、回転不能であるようにギヤケース３４００に固定される。第１キャリア（ＣＡ）３１０６は、プラネタリギヤユニット３０００の入力軸３００２に連結される。 The first sun gear (S1) 3102 is fixed to the gear case 3400 so as not to rotate. First carrier (CA) 3106 is coupled to input shaft 3002 of planetary gear unit 3000.
　リアプラネタリ３２００は、ラビニヨ型の遊星歯車機構である。リアプラネタリ３２００は、第２サンギヤ（Ｓ２）３２０２と、第２ピニオンギヤ（Ｐ２）３２０４と、リアキャリア（ＲＣＡ）３２０６と、リアリングギヤ（ＲＲ）３２０８と、第３サンギヤ（Ｓ３）３２１０と、第３ピニオンギヤ（Ｐ３）３２１２とを含む。 The rear planetary 3200 is a Ravigneaux type planetary gear mechanism. The rear planetary 3200 includes a second sun gear (S2) 3202, a second pinion gear (P2) 3204, a rear carrier (RCA) 3206, a rear ring gear (RR) 3208, a third sun gear (S3) 3210, a third Pinion gear (P3) 3212.
　第２ピニオンギヤ（Ｐ２）３２０４は、第２サンギヤ（Ｓ２）３２０２、リアリングギヤ（ＲＲ）３２０８および第３ピニオンギヤ（Ｐ３）３２１２と噛合っている。第３ピニオンギヤ（Ｐ３）３２１２は、第２ピニオンギヤ（Ｐ２）３２０４に加えて、第３サンギヤ（Ｓ３）３２１０と噛合っている。 The second pinion gear (P2) 3204 meshes with the second sun gear (S2) 3202, the rear ring gear (RR) 3208, and the third pinion gear (P3) 3212. Third pinion gear (P3) 3212 meshes with third sun gear (S3) 3210 in addition to second pinion gear (P2) 3204.
　リアキャリア（ＲＣＡ）３２０６は、第２ピニオンギヤ（Ｐ２）３２０４および第３ピニオンギヤ（Ｐ３）３２１２が公転および自転可能であるように支持している。リアキャリア（ＲＣＡ）３２０６は、ワンウェイクラッチ（Ｆ）３３２０に連結される。リアキャリア（ＲＣＡ）３２０６は、１速ギヤ段の駆動時（エンジン１０００から出力された駆動力を用いた走行時）に回転不能となる。リアリングギヤ（ＲＲ）３２０８は、プラネタリギヤユニット３０００の出力軸３００４に連結される。 The rear carrier (RCA) 3206 supports the second pinion gear (P2) 3204 and the third pinion gear (P3) 3212 so that they can revolve and rotate. Rear carrier (RCA) 3206 is coupled to one-way clutch (F) 3320. The rear carrier (RCA) 3206 becomes non-rotatable when driving the first gear (when traveling using the driving force output from the engine 1000). Rear ring gear (RR) 3208 is coupled to output shaft 3004 of planetary gear unit 3000.
　ワンウェイクラッチ（Ｆ）３３２０は、Ｂ２ブレーキ３３１２と並列に設けられる。すなわち、ワンウェイクラッチ（Ｆ）３３２０のアウターレースはギヤケース３４００に固定され、インナーレースはリアキャリア（ＲＣＡ）３２０６に連結される。 The one-way clutch (F) 3320 is provided in parallel with the B2 brake 3312. That is, the outer race of the one-way clutch (F) 3320 is fixed to the gear case 3400, and the inner race is connected to the rear carrier (RCA) 3206.
　図３に、各変速ギヤ段と、各クラッチおよび各ブレーキの作動状態との関係を表した作動表を示す。この作動表に示された組み合わせで各ブレーキおよび各クラッチを作動させることによって、前進１速～８速のギヤ段と、後進１速および２速のギヤ段が形成される。 FIG. 3 shows an operation table showing the relationship between each gear position and the operation state of each clutch and each brake. By operating the brakes and the clutches in the combinations shown in this operation table, forward 1st to 8th gears and reverse 1st and 2nd gears are formed.
　たとえば、Ｃ１クラッチ３３０１およびＣ３クラッチ３３０３を係合させ、他のクラッチおよびブレーキを解放させることによって、前進３速のギヤ段が形成される。 For example, the forward third speed gear stage is formed by engaging the C1 clutch 3301 and the C3 clutch 3303 and releasing the other clutches and brakes.
　図４を参照して、油圧回路４０００の要部について説明する。なお、油圧回路４０００は、以下に説明するものに限られない。 The main part of the hydraulic circuit 4000 will be described with reference to FIG. The hydraulic circuit 4000 is not limited to the one described below.
　油圧回路４０００は、オイルポンプ４００４と、プライマリレギュレータバルブ４００６と、マニュアルバルブ４１００と、ソレノイドモジュレータバルブ４２００と、ＳＬ１リニアソレノイド（以下、ＳＬ（１）と記載する）４２１０と、ＳＬ２リニアソレノイド（以下、ＳＬ（２）と記載する）４２２０と、ＳＬ３リニアソレノイド（以下、ＳＬ（３）と記載する）４２３０と、ＳＬ４リニアソレノイド（以下、ＳＬ（４）と記載する）４２４０と、ＳＬ５リニアソレノイド（以下、ＳＬ（５）と記載する）４２５０と、ＳＬＴリニアソレノイド（以下、ＳＬＴと記載する）４３００と、Ｂ２コントロールバルブ４５００とを含む。 The hydraulic circuit 4000 includes an oil pump 4004, a primary regulator valve 4006, a manual valve 4100, a solenoid modulator valve 4200, an SL1 linear solenoid (hereinafter referred to as SL (1)) 4210, and an SL2 linear solenoid (hereinafter referred to as “the solenoid valve”). SL2 (described as SL (4)) 4220, SL3 linear solenoid (hereinafter referred to as SL (3)) 4230, SL4 linear solenoid (hereinafter referred to as SL (4)) 4240, and SL5 linear solenoid (hereinafter referred to as SL (3)). , SL (5)) 4250, SLT linear solenoid (hereinafter referred to as SLT) 4300, and B2 control valve 4500.
　オイルポンプ４００４は、エンジン１０００のクランクシャフトに連結されている。クランクシャフトが回転することによって、オイルポンプ４００４が駆動し、油圧を発生する。 Oil pump 4004 is connected to the crankshaft of engine 1000. As the crankshaft rotates, the oil pump 4004 is driven to generate hydraulic pressure.
　オイルポンプ４００４で発生した油圧は、プライマリレギュレータバルブ４００６によって調圧され、ライン圧が生成される。プライマリレギュレータバルブ４００６は、ＳＬＴ４３００によって調圧されたスロットル圧をパイロット圧として作動する。ライン圧は、ライン圧油路４０１０を経由してマニュアルバルブ４１００に供給される。 The hydraulic pressure generated by the oil pump 4004 is regulated by the primary regulator valve 4006 to generate a line pressure. Primary regulator valve 4006 operates using the throttle pressure regulated by SLT 4300 as a pilot pressure. The line pressure is supplied to the manual valve 4100 via the line pressure oil passage 4010.
　マニュアルバルブ４１００は、ドレンポート４１０５を含む。ドレンポート４１０５から、Ｄレンジ圧油路４１０２およびＲレンジ圧油路４１０４の油圧が排出される。 Manual valve 4100 includes a drain port 4105. From the drain port 4105, the oil pressure in the D range pressure oil passage 4102 and the R range pressure oil passage 4104 is discharged.
　マニュアルバルブ４１００のスプールがＤポジションにある場合、ライン圧油路４０１０とＤレンジ圧油路４１０２とが連通させられ、Ｄレンジ圧油路４１０２に油圧が供給される。このとき、Ｒレンジ圧油路４１０４とドレンポート４１０５とが連通させられ、Ｒレンジ圧油路４１０４のＲレンジ圧がドレンポート４１０５から排出される。 When the spool of the manual valve 4100 is in the D position, the line pressure oil passage 4010 and the D range pressure oil passage 4102 are communicated, and the oil pressure is supplied to the D range pressure oil passage 4102. At this time, the R range pressure oil passage 4104 and the drain port 4105 are communicated, and the R range pressure of the R range pressure oil passage 4104 is discharged from the drain port 4105.
　マニュアルバルブ４１００のスプールがＲポジションにある場合、ライン圧油路４０１０とＲレンジ圧油路４１０４とが連通させられ、Ｒレンジ圧油路４１０４に油圧が供給される。このとき、Ｄレンジ圧油路４１０２とドレンポート４１０５とが連通させられ、Ｄレンジ圧油路４１０２のＤレンジ圧がドレンポート４１０５から排出される。 When the spool of the manual valve 4100 is in the R position, the line pressure oil passage 4010 and the R range pressure oil passage 4104 are communicated, and the oil pressure is supplied to the R range pressure oil passage 4104. At this time, the D range pressure oil passage 4102 and the drain port 4105 are communicated, and the D range pressure in the D range pressure oil passage 4102 is discharged from the drain port 4105.
　マニュアルバルブ４１００のスプールがＮポジションにある場合、Ｄレンジ圧油路４１０２およびＲレンジ圧油路４１０４の両方と、ドレンポート４１０５とが連通させられ、Ｄレンジ圧油路４１０２のＤレンジ圧およびＲレンジ圧油路４１０４のＲレンジ圧がドレンポート４１０５から排出される。 When the spool of the manual valve 4100 is in the N position, both the D range pressure oil passage 4102 and the R range pressure oil passage 4104 are connected to the drain port 4105, and the D range pressure and R of the D range pressure oil passage 4102 are communicated. The R range pressure of the range pressure oil passage 4104 is discharged from the drain port 4105.
　Ｂ２コントロールバルブ４５００は、Ｄレンジ圧油路４１０２およびＲレンジ圧油路４１０４のいずれか一方からの油圧を選択的に、Ｂ２ブレーキ３３１２に供給する。Ｂ２コントロールバルブ４５００に、Ｄレンジ圧油路４１０２およびＲレンジ圧油路４１０４が接続されている。Ｂ２コントロールバルブ４５００は、ＳＬＵソレノイドバルブ（図示せず）から供給された油圧とスプリングの付勢力とによって制御される。 The B2 control valve 4500 selectively supplies hydraulic pressure from one of the D range pressure oil passage 4102 and the R range pressure oil passage 4104 to the B2 brake 3312. A D range pressure oil passage 4102 and an R range pressure oil passage 4104 are connected to the B2 control valve 4500. The B2 control valve 4500 is controlled by the hydraulic pressure supplied from the SLU solenoid valve (not shown) and the urging force of the spring.
　ＳＬＵソレノイドバルブがオンの場合、Ｂ２コントロールバルブ４５００は、図４において左側の状態となる。この場合、Ｂ２ブレーキ３３１２には、ＳＬＵソレノイドバルブから供給された油圧をパイロット圧として、Ｄレンジ圧を調圧した油圧が供給される。なお、Ｄレンジ圧油路４１０２に供給された油圧は、最終的には、Ｃ１クラッチ３３０１、Ｃ２クラッチ３３０２およびＣ３クラッチ３３０３に供給される。 When the SLU solenoid valve is on, the B2 control valve 4500 is in the state on the left side in FIG. In this case, the B2 brake 3312 is supplied with the hydraulic pressure adjusted from the D range pressure using the hydraulic pressure supplied from the SLU solenoid valve as a pilot pressure. The hydraulic pressure supplied to the D range pressure oil passage 4102 is finally supplied to the C1 clutch 3301, the C2 clutch 3302, and the C3 clutch 3303.
　ＳＬＵソレノイドバルブがオフの場合、Ｂ２コントロールバルブ４５００は、図４において右側の状態となる。この場合、Ｂ２ブレーキ３３１２には、Ｒレンジ圧油路４１０４のＲレンジ圧が供給される。 When the SLU solenoid valve is off, the B2 control valve 4500 is in the state on the right side in FIG. In this case, the R range pressure of the R range pressure oil passage 4104 is supplied to the B2 brake 3312.
　ＳＬ（１）４２１０、ＳＬ（２）４２２０、ＳＬ（３）４２３０、ＳＬ（４）４２４０、ＳＬ（５）４２５０は、ノーマルクローズ型のリニアソレノイドバルブである。これらのリニアソレノイドバルブの出力油圧ＰＯＵＴは、駆動信号としてＥＣＵ８０００からそれぞれのリニアソレノイドバルブに出力される駆動電流Ｉによって制御される。すなわち、これらのリニアソレノイドバルブは、非通電時に出力油圧が最小（「０」）になり、ＥＣＵ８０００からの各駆動電流Ｉの増大に伴ない出力油圧が増大する。 SL (1) 4210, SL (2) 4220, SL (3) 4230, SL (4) 4240, and SL (5) 4250 are normally closed linear solenoid valves. The output hydraulic pressure POUT of these linear solenoid valves is controlled by a drive current I output from the ECU 8000 to each linear solenoid valve as a drive signal. In other words, the output hydraulic pressure of these linear solenoid valves becomes minimum (“0”) when not energized, and the output hydraulic pressure increases as each drive current I from ECU 8000 increases.
　ＳＬ（１）４２１０の出力油圧ＰＯＵＴ（Ｃ１）は、Ｃ１クラッチ３３０１に供給され、ＳＬ（２）４２２０の出力油圧ＰＯＵＴ（Ｃ２）は、Ｃ２クラッチ３３０２に供給され、ＳＬ（３）４２３０の出力油圧ＰＯＵＴ（Ｃ３）は、Ｃ３クラッチ３３０３に供給され、ＳＬ（４）４２４０の出力油圧ＰＯＵＴ（Ｃ４）は、Ｃ４クラッチ３３０４に供給され、ＳＬ（５）４２５０の出力油圧ＰＯＵＴ（Ｂ１）は、Ｂ１ブレーキ３３１１に供給される。 The output hydraulic pressure POUT (C1) of SL (1) 4210 is supplied to the C1 clutch 3301, the output hydraulic pressure POUT (C2) of SL (2) 4220 is supplied to the C2 clutch 3302, and the output hydraulic pressure of SL (3) 4230. POUT (C3) is supplied to the C3 clutch 3303, the output hydraulic pressure POUT (C4) of the SL (4) 4240 is supplied to the C4 clutch 3304, and the output hydraulic pressure POUT (B1) of the SL (5) 4250 is supplied to the B1 brake. 3311.
　ＥＣＵ８０００は、入力トルク（タービントルク）などに基づいて、出力油圧ＰＯＵＴ（Ｃ１）の目標指令圧Ｐ（Ｃ１）、出力油圧ＰＯＵＴ（Ｃ２）の目標指令圧Ｐ（Ｃ２）、出力油圧ＰＯＵＴ（Ｃ３）の目標指令圧Ｐ（Ｃ３）、出力油圧ＰＯＵＴ（Ｃ４）の目標指令圧Ｐ（Ｃ４）、出力油圧ＰＯＵＴ（Ｂ１）の目標指令圧Ｐ（Ｂ１）をそれぞれ算出する。なお、目標指令圧の算出方法は後に詳述する。 Based on input torque (turbine torque) and the like, ECU 8000 has target command pressure P (C1) of output hydraulic pressure POUT (C1), target command pressure P (C2) of output hydraulic pressure POUT (C2), and output hydraulic pressure POUT (C3). Target command pressure P (C3), target command pressure P (C4) of output hydraulic pressure POUT (C4), and target command pressure P (B1) of output hydraulic pressure POUT (B1). The method for calculating the target command pressure will be described in detail later.
　ＥＣＵ８０００は、目標指令圧Ｐ（Ｃ１）に応じた駆動電流Ｉ（Ｃ１）、目標指令圧Ｐ（Ｃ２）に応じた駆動電流Ｉ（Ｃ２）、目標指令圧Ｐ（Ｃ３）に応じた駆動電流Ｉ（Ｃ３）、目標指令圧Ｐ（Ｃ４）に応じた駆動電流Ｉ（Ｃ４）、目標指令圧Ｐ（Ｂ１）に応じた駆動電流Ｉ（Ｂ１）を、それぞれＳＬ（１）４２１０、ＳＬ（２）４２２０、ＳＬ（３）４２３０、ＳＬ（４）４２４０、ＳＬ（５）４２５０に出力する。 The ECU 8000 includes a drive current I (C1) corresponding to the target command pressure P (C1), a drive current I (C2) corresponding to the target command pressure P (C2), and a drive current I corresponding to the target command pressure P (C3). (C3), drive current I (C4) corresponding to the target command pressure P (C4), and drive current I (B1) corresponding to the target command pressure P (B1) are SL (1) 4210 and SL (2), respectively. 4220, SL (3) 4230, SL (4) 4240, and SL (5) 4250.
　ソレノイドモジュレータバルブ４２００は、ライン圧を元圧とし、ＳＬＴ４３００に供給する油圧（ソレノイドモジュレータ圧）を一定の圧力に調圧する。 Solenoid modulator valve 4200 adjusts the hydraulic pressure (solenoid modulator pressure) supplied to SLT 4300 to a constant pressure using the line pressure as the original pressure.
　ＳＬＴ４３００は、アクセル開度センサ８０１０によって検出されたアクセル開度ＡＣＣなどに基づいたＥＣＵ８０００からの駆動電流Ｉ（Ｔ）に応じて、ソレノイドモジュレータ圧を調圧し、スロットル圧を生成する。スロットル圧は、ＳＬＴ油路４３０２を経由して、プライマリレギュレータバルブ４００６に供給される。スロットル圧は、プライマリレギュレータバルブ４００６のパイロット圧として利用される。 The SLT 4300 adjusts the solenoid modulator pressure in accordance with the drive current I (T) from the ECU 8000 based on the accelerator opening degree ACC detected by the accelerator opening degree sensor 8010, and generates a throttle pressure. The throttle pressure is supplied to the primary regulator valve 4006 via the SLT oil passage 4302. The throttle pressure is used as a pilot pressure for the primary regulator valve 4006.
　図５－７を参照して、ＳＬ（１）４２１０の構造および特性について説明する。なお、ＳＬ（２）４２２０、ＳＬ（３）４２３０、ＳＬ（４）４２４０、ＳＬ（５）４２５０およびＳＬＴ４３００は、ＳＬ（１）４２１０と同様の構造および特性を有するため、それらの構造および特性についての説明は繰返さない。 Referring to FIG. 5-7, the structure and characteristics of SL (1) 4210 will be described. Note that SL (2) 4220, SL (3) 4230, SL (4) 4240, SL (5) 4250, and SLT 4300 have the same structure and characteristics as SL (1) 4210. The description of will not be repeated.
　図５は、ＥＣＵ８０００からの駆動電流Ｉ（Ｃ１）が最大値Ｉ（ＭＡＸ）である場合のＳＬ（１）４２１０の内部の状態を示している。ＳＬ（１）４２１０は、スプール４２１２がバルブ４２１３を押す力がコイル４２１１に流れる駆動電流Ｉ（Ｃ１）にほぼ比例して大きくなる現象を利用して、駆動電流Ｉ（Ｃ１）によってスプール４２１２のストローク量を調整して、入力ポート４２１７の流路断面を変化させる。これにより、ライン圧ＰＬが調圧された出力油圧ＰＯＵＴ（Ｃ１）が出力ポート４２１８から出力される。 FIG. 5 shows the internal state of SL (1) 4210 when drive current I (C1) from ECU 8000 is maximum value I (MAX). SL (1) 4210 utilizes the phenomenon that the force by which the spool 4212 pushes the valve 4213 increases in proportion to the drive current I (C1) flowing through the coil 4211, and the stroke of the spool 4212 is driven by the drive current I (C1). The flow rate cross section of the input port 4217 is changed by adjusting the amount. As a result, the output hydraulic pressure POUT (C1) in which the line pressure PL is regulated is output from the output port 4218.
　スプール４２１２のストローク量は、駆動電流Ｉ（Ｃ１）にほぼ比例して大きくなる。駆動電流Ｉ（Ｃ１）が最大電流値Ｉ（ＭＡＸ）になると、スプール４２１２がバルブ４２１３を押す力が最大となり、図５に示すように、スプリング４２１４の付勢力に反してバルブ４２１３のスプリング４２１４側の先端がストッパ４２１５に当接する状態（フルストローク状態）となる。 The stroke amount of the spool 4212 increases substantially in proportion to the drive current I (C1). When the drive current I (C1) reaches the maximum current value I (MAX), the force by which the spool 4212 pushes the valve 4213 becomes maximum, and the spring 4214 side of the valve 4213 is against the biasing force of the spring 4214 as shown in FIG. Is brought into contact with the stopper 4215 (full stroke state).
　図６は、目標指令圧Ｐ（Ｃ１）と駆動電流Ｉ（Ｃ１）との関係を示している。駆動電流Ｉ（Ｃ１）は、目標指令圧Ｐ（Ｃ１）が０（最小圧）であると０（最小電流値）となり、目標指令圧Ｐ（Ｃ１）にほぼ比例して大きくなる。目標指令圧Ｐ（Ｃ１）がライン圧ＰＬのときの駆動電流Ｉ（Ｃ１）は、Ｉ（ＰＬ）となる。目標指令圧Ｐ（Ｃ１）が目標指令値の最大圧Ｐ（ＭＡＸ）になると、駆動電流Ｉ（Ｃ１）は最大電流値Ｉ（ＭＡＸ）となる。 FIG. 6 shows the relationship between the target command pressure P (C1) and the drive current I (C1). The drive current I (C1) becomes 0 (minimum current value) when the target command pressure P (C1) is 0 (minimum pressure), and increases substantially in proportion to the target command pressure P (C1). The drive current I (C1) when the target command pressure P (C1) is the line pressure PL is I (PL). When the target command pressure P (C1) becomes the maximum pressure P (MAX) of the target command value, the drive current I (C1) becomes the maximum current value I (MAX).
　なお、目標指令圧Ｐ（Ｃ１）と駆動電流Ｉ（Ｃ１）との関係は、図６に示したものに限定されない。たとえば、目標指令圧Ｐ（Ｃ１）がライン圧ＰＬより大きくなった場合に、駆動電流Ｉ（Ｃ１）を最大電流値Ｉ（ＭＡＸ）とするようにしてもよい。 It should be noted that the relationship between the target command pressure P (C1) and the drive current I (C1) is not limited to that shown in FIG. For example, when the target command pressure P (C1) becomes larger than the line pressure PL, the drive current I (C1) may be set to the maximum current value I (MAX).
　図７は、駆動電流Ｉ（Ｃ１）と出力油圧ＰＯＵＴ（Ｃ１）との関係を示している。図７に示すように、駆動電流Ｉ（Ｃ１）がＩ（ＰＬ）より小さい範囲では、出力油圧ＰＯＵＴ（Ｃ１）が駆動電流Ｉ（Ｃ１）にほぼ比例して大きくなる。このとき、ＳＬ（１）４２１０は、ライン圧ＰＬを駆動電流Ｉ（Ｃ１）に応じて調圧することが可能な調圧状態となる。 FIG. 7 shows the relationship between the drive current I (C1) and the output hydraulic pressure POUT (C1). As shown in FIG. 7, in the range where the drive current I (C1) is smaller than I (PL), the output hydraulic pressure POUT (C1) increases substantially in proportion to the drive current I (C1). At this time, SL (1) 4210 is in a pressure regulation state in which the line pressure PL can be regulated according to the drive current I (C1).
　一方、駆動電流ＩがＩ（ＰＬ）を超えた場合、出力油圧ＰＯＵＴ（Ｃ１）は、駆動電流Ｉ（Ｃ１）を変化させても最大値（すなわちライン圧ＰＬそのままの値）となる。このとき、ＳＬ（１）４２１０は、ライン圧ＰＬを調圧できない状態（非調圧状態）となる。非調圧状態になる要因は、駆動電流ＩがＩ（ＰＬ）を超えた場合には、調圧面４２１６が間隔Ａに示した範囲に位置する状態となる（図５参照）ため、駆動電流Ｉ（Ｃ１）を変化させても入力ポート４２１７の流路断面が変化しないためである。 On the other hand, when the drive current I exceeds I (PL), the output hydraulic pressure POUT (C1) becomes the maximum value (that is, the value of the line pressure PL as it is) even if the drive current I (C1) is changed. At this time, SL (1) 4210 enters a state where the line pressure PL cannot be regulated (non-regulated state). When the drive current I exceeds I (PL), the factor causing the non-pressure regulation state is that the pressure regulation surface 4216 is located in the range indicated by the interval A (see FIG. 5). This is because the flow path cross section of the input port 4217 does not change even if (C1) is changed.
　従来においては、たとえばＣ１クラッチ３３０１を係合状態に維持する際、目標指令圧Ｐ（Ｃ１）を最大圧Ｐ（ＭＡＸ）に設定して、駆動電流Ｉ（Ｃ１）を最大電流値Ｉ（ＭＡＸ）にしていた（図６参照）。 Conventionally, for example, when the C1 clutch 3301 is maintained in the engaged state, the target command pressure P (C1) is set to the maximum pressure P (MAX), and the drive current I (C1) is set to the maximum current value I (MAX). (See FIG. 6).
　しかしながら、図７に示したように、駆動電流Ｉ（Ｃ１）がＩ（ＰＬ）からＩ（ＭＡＸ）の範囲である場合、ＳＬ（１）４２１０は非調圧状態となり、ライン圧ＰＬがそのまま出力される。すなわち、駆動電流Ｉ（Ｃ１）のうち、Ｉ（ＰＬ）を超えた部分は出力油圧ＰＯＵＴ（Ｃ１）に反映されず、無駄に消費されることになる。 However, as shown in FIG. 7, when the drive current I (C1) is in the range from I (PL) to I (MAX), SL (1) 4210 is in a non-regulated state, and the line pressure PL is output as it is. Is done. That is, the portion of the drive current I (C1) that exceeds I (PL) is not reflected in the output hydraulic pressure POUT (C1) and is wasted.
　そこで、本発明においては、係合状態に維持すべき摩擦係合要素（クラッチあるいはブレーキ）に対応するリニアソレノイドバルブの出力油圧の目標指令圧Ｐを、係合状態に維持可能であってかつ最大圧Ｐ（ＭＡＸ）よりも小さい油圧とすることによって、リニアソレノイドバルブに供給される駆動電流を低減する。 Therefore, in the present invention, the target command pressure P of the output hydraulic pressure of the linear solenoid valve corresponding to the friction engagement element (clutch or brake) to be maintained in the engaged state can be maintained in the engaged state and is maximum. By making the hydraulic pressure smaller than the pressure P (MAX), the drive current supplied to the linear solenoid valve is reduced.
　図８に、本実施例に係る車両の制御装置であるＥＣＵ８０００の機能ブロック図を示す。ＥＣＵ８０００は、入力インターフェイス（以下、入力Ｉ／Ｆと記載する）８１００と、演算処理部８２００と、記憶部８３００と、出力インターフェイス（以下、出力Ｉ／Ｆと記載する）８４００とを含む。 FIG. 8 shows a functional block diagram of the ECU 8000 which is a vehicle control apparatus according to this embodiment. ECU 8000 includes an input interface (hereinafter referred to as input I / F) 8100, a calculation processing unit 8200, a storage unit 8300, and an output interface (hereinafter referred to as output I / F) 8400.
　入力Ｉ／Ｆ８１００は、車速センサ８００２からの車速Ｖ、アクセル開度センサ８０１０からのアクセル開度ＡＣＣ、ポジションスイッチ８００６からのシフトポジションＳＰ、エンジン回転数センサ８０２０からのエンジン回転数ＮＥ、入力軸回転数センサ８０２２からの入力軸回転数ＮＴ、出力軸回転数センサ８０２４からの出力軸回転数ＮＯＵＴ、スロットル開度センサ８０１８からのスロットル開度ＴＨを受信して、演算処理部８２００に送信する。 The input I / F 8100 includes a vehicle speed V from the vehicle speed sensor 8002, an accelerator opening ACC from the accelerator opening sensor 8010, a shift position SP from the position switch 8006, an engine speed NE from the engine speed sensor 8020, and an input shaft rotation. The input shaft rotational speed NT from the number sensor 8022, the output shaft rotational speed NOUT from the output shaft rotational speed sensor 8024, and the throttle opening TH from the throttle opening sensor 8018 are received and transmitted to the arithmetic processing unit 8200.
　記憶部８３００には、各種情報、プログラム、しきい値、マップ等が記憶され、必要に応じて演算処理部８２００からデータが読み出されたり、格納されたりする。 The storage unit 8300 stores various types of information, programs, threshold values, maps, and the like, and data is read from or stored in the arithmetic processing unit 8200 as necessary.
　演算処理部８２００は、変速制御判断部８２１０と、必要油圧算出部８２２０と、係合油圧制御部８２３０と、ライン圧制御部８２４０とを含む。 The arithmetic processing unit 8200 includes a shift control determination unit 8210, a necessary hydraulic pressure calculation unit 8220, an engagement hydraulic pressure control unit 8230, and a line pressure control unit 8240.
　変速制御判断部８２１０は、非変速中であるか否かを判断する。なお、非変速中とは、変速制御が行なわれていない状態を意味する。たとえば、変速制御判断部８２１０は、車速Ｖ、アクセル開度ＡＣＣおよび変速線図に基づく変速判断が行われておらず、かつ進行中の変速がない場合に、非変速中であると判断する。 The shift control determination unit 8210 determines whether or not a non-shift is in progress. Note that non-shifting means a state in which shift control is not performed. For example, the shift control determination unit 8210 determines that a non-shift is in progress when the shift determination based on the vehicle speed V, the accelerator opening degree ACC, and the shift diagram has not been performed and there is no ongoing shift.
　必要油圧算出部８２２０は、前述した図３の作動表に基づいて非変速中において係合状態に維持すべき摩擦係合要素を判断し、これらの摩擦係合要素を係合状態に維持するための必要油圧ＰＲ（Ｃ１クラッチ３３０１の必要油圧ＰＲ（Ｃ１）、Ｃ２クラッチ３３０２の必要油圧ＰＲ（Ｃ２）、Ｃ３クラッチ３３０３の必要油圧ＰＲ（Ｃ３）、Ｃ４クラッチ３３０４の必要油圧ＰＲ（Ｃ４）、Ｂ１ブレーキ３３１１の必要油圧ＰＲ（Ｂ１）の少なくともいずれか）を算出する。 The required hydraulic pressure calculation unit 8220 determines friction engagement elements that should be maintained in the engaged state during non-shifting based on the operation table of FIG. 3 described above, and maintains these friction engagement elements in the engaged state. Required hydraulic pressure PR (required hydraulic pressure PR (C1) of the C1 clutch 3301, required hydraulic pressure PR (C2) of the C2 clutch 3302, required hydraulic pressure PR (C3) of the C3 clutch 3303, required hydraulic pressure PR (C4) of the C4 clutch 3304, B1 The required hydraulic pressure PR (B1) of the brake 3311) is calculated.
　必要油圧算出部８２２０は、各摩擦係合要素を係合状態に維持するために必要な最小圧力を（入力トルク）×（トルク分担率）として算出し、これに安全係数（たとえば１．２）を乗じた値を、必要油圧ＰＲとして算出する。このように、必要油圧ＰＲは、各摩擦係合要素を係合状態に維持するために必要な最小圧力よりもやや高い（安全係数分だけ大きい）値に算出されるため、目標指令圧Ｐの最大圧Ｐ（ＭＡＸ）よりも小さい値となる。必要油圧ＰＲは、係合状態に維持すべき摩擦係合要素ごとに算出される。 The required oil pressure calculation unit 8220 calculates the minimum pressure required to maintain each friction engagement element in the engaged state as (input torque) × (torque sharing ratio), and adds a safety factor (for example, 1.2). A value obtained by multiplying is calculated as the required hydraulic pressure PR. In this way, the required hydraulic pressure PR is calculated to a value slightly higher (larger by the safety factor) than the minimum pressure required to maintain each friction engagement element in the engaged state. The value is smaller than the maximum pressure P (MAX). The required hydraulic pressure PR is calculated for each friction engagement element to be maintained in the engaged state.
　ここで、入力トルクとは、エンジン１０００からトルクコンバータ２１００を経由してプラネタリギヤユニット３０００の入力軸３００２に入力されたトルク（タービントルク）である。入力トルクは、たとえば、エンジン回転数ＮＥ、アクセル開度ＡＣＣあるいはスロットル開度ＴＨに基づいて算出される。 Here, the input torque is a torque (turbine torque) input from the engine 1000 to the input shaft 3002 of the planetary gear unit 3000 via the torque converter 2100. The input torque is calculated based on, for example, the engine speed NE, the accelerator opening ACC, or the throttle opening TH.
　また、トルク分担率とは、係合状態に維持される各摩擦係合要素に入力トルクが分担される割合であって、係合状態に維持される摩擦係合要素の組み合わせに応じて予め設定される。 Further, the torque sharing ratio is a ratio that the input torque is shared by each friction engagement element maintained in the engaged state, and is set in advance according to the combination of the friction engagement elements maintained in the engaged state. Is done.
　係合油圧制御部８２３０は、必要油圧算出部８２２０で算出された各必要油圧ＰＲに応じて目標指令圧Ｐ（Ｐ（Ｃ１）、Ｐ（Ｃ２）、Ｐ（Ｃ３）、Ｐ（Ｃ４）、Ｐ（Ｂ１）の少なくともいずれか）を算出し、算出された目標指令圧Ｐに応じて駆動電流Ｉ（Ｉ（Ｃ１）、Ｉ（Ｃ２）、Ｉ（Ｃ３）、Ｉ（Ｃ４）、Ｉ（Ｂ１）の少なくともいずれか）を算出し、算出された各駆動電流Ｉを出力Ｉ／Ｆ８４００を経由して各リニアソレノイドバルブに出力する。なお、解放される摩擦係合要素に対応する各リニアソレノイドバルブには、駆動電流Ｉは出力されない。 The engagement hydraulic pressure control unit 8230 has target command pressures P (P (C1), P (C2), P (C3), P (C4), P according to each required hydraulic pressure PR calculated by the required hydraulic pressure calculation unit 8220. (At least one of (B1)) is calculated, and the drive current I (I (C1), I (C2), I (C3), I (C4), I (B1) is determined according to the calculated target command pressure P. And outputs the calculated drive current I to each linear solenoid valve via the output I / F 8400. The drive current I is not output to each linear solenoid valve corresponding to the released frictional engagement element.
　ライン圧制御部８２４０は、必要油圧算出部８２２０で算出された必要油圧ＰＲよりもライン圧ＰＬが予め定められた値αだけ高くなるスロットル圧（プライマリレギュレータバルブ４００６のパイロット圧）を生成する駆動電流Ｉ（Ｔ）を、出力Ｉ／Ｆ８４００を経由してＳＬＴ４３００に出力する。 The line pressure control unit 8240 generates a drive current that generates a throttle pressure (a pilot pressure of the primary regulator valve 4006) at which the line pressure PL is higher than the required hydraulic pressure PR calculated by the required hydraulic pressure calculation unit 8220 by a predetermined value α. I (T) is output to the SLT 4300 via the output I / F 8400.
　なお、本実施例において、変速制御判断部８２１０と、必要油圧算出部８２２０と、係合油圧制御部８２３０と、ライン圧制御部８２４０とは、いずれも演算処理部８２００であるＣＰＵが記憶部８３００に記憶されたプログラムを実行することによって実現される、ソフトウェアとして機能するものとして説明するが、ハードウェアによって実現されるようにしてもよい。なお、このようなプログラムは記憶媒体に記録されて車両に搭載される。 In this embodiment, the shift control determination unit 8210, the required hydraulic pressure calculation unit 8220, the engagement hydraulic pressure control unit 8230, and the line pressure control unit 8240 are all calculated by the CPU that is the arithmetic processing unit 8200 by the storage unit 8300. Although it is assumed that the program functions as software that is realized by executing the program stored in the program, it may be realized by hardware. Such a program is recorded on a storage medium and mounted on the vehicle.
　図９を参照して、本実施例に係る制御装置であるＥＣＵ８０００で実行されるプログラムの制御構造について説明する。なお、このプログラムは、予め定められたサイクルタイムで繰り返し実行される。 Referring to FIG. 9, a control structure of a program executed by ECU 8000 which is the control device according to the present embodiment will be described. Note that this program is repeatedly executed at a predetermined cycle time.
　ステップ（以下、ステップをＳと略す）１００にて、ＥＣＵ８０００は、非変速中であるか否かを判断する。非変速中であると（Ｓ１００にてＹＥＳ）、処理はＳ１０２に移される。そうでないと（Ｓ１００にてＮＯ）、この処理は終了する。 In step (hereinafter, step is abbreviated as S) 100, ECU 8000 determines whether or not a non-shifting operation is being performed. If not (YES at S100), the process proceeds to S102. Otherwise (NO in S100), this process ends.
　Ｓ１０２にて、ＥＣＵ８０００は、係合状態に維持すべき摩擦係合要素ごとに、必要油圧ＰＲを算出する。ＥＣＵ８０００は、上述したように、入力トルク×トルク分担率×安全係数で算出された値を、各摩擦係合要素ごとの必要油圧ＰＲとして算出する。 In step S102, the ECU 8000 calculates a required hydraulic pressure PR for each friction engagement element to be maintained in the engaged state. As described above, the ECU 8000 calculates the value calculated by the input torque × the torque sharing ratio × the safety factor as the required hydraulic pressure PR for each friction engagement element.
　たとえば前進３速ギヤ段での走行時においては、Ｃ１クラッチ３３０１とＣ３クラッチ３３０３とをそれぞれ係合状態に維持する必要があるため、ＥＣＵ８０００は、Ｃ１クラッチ３３０１の必要油圧ＰＲ（Ｃ１）を入力トルク×（Ｃ１クラッチ３３０１のトルク分担率）×安全係数として算出し、Ｃ３クラッチ３３０３の必要油圧ＰＲ（Ｃ３）を入力トルク×（Ｃ３クラッチ３３０３のトルク分担率）×安全係数として算出する。 For example, when traveling in the forward third gear, the C1 clutch 3301 and the C3 clutch 3303 need to be maintained in an engaged state, so the ECU 8000 uses the required hydraulic pressure PR (C1) of the C1 clutch 3301 as an input torque. X (torque sharing ratio of C1 clutch 3301) x safety factor, and necessary hydraulic pressure PR (C3) of C3 clutch 3303 is calculated as input torque x (torque sharing ratio of C3 clutch 3303) x safety coefficient.
　Ｓ１０４にて、ＥＣＵ８０００は、ライン圧ＰＬが、Ｓ１０２の処理で算出された必要油圧ＰＲのうちの最大圧よりも予め定められた値αだけ高くなる駆動電流Ｉ（Ｔ）をＳＬＴ４３００に出力する。これにより、たとえば前進３速ギヤ段での走行時において、Ｃ１クラッチ３３０１の必要油圧ＰＲ（Ｃ１）が最大圧である（すなわち必要油圧ＰＲ（Ｃ３）よりも高い）場合、ライン圧ＰＬは、必要油圧ＰＲ（Ｃ１）＋αに制御される。 In S104, ECU 8000 outputs to SLT 4300 drive current I (T) in which line pressure PL is higher by a predetermined value α than the maximum pressure of required hydraulic pressure PR calculated in S102. As a result, for example, when traveling at the third forward gear, if the required hydraulic pressure PR (C1) of the C1 clutch 3301 is the maximum pressure (that is, higher than the required hydraulic pressure PR (C3)), the line pressure PL is required. The hydraulic pressure is controlled to PR (C1) + α.
　Ｓ１０６にて、ＥＣＵ８０００は、Ｓ１０２の処理で算出された必要油圧ＰＲのうちの最大圧を、係合状態に維持すべき各摩擦係合要素に対する目標指令圧Ｐとして算出する。たとえば、ＥＣＵ８０００は、前進３速ギヤ段での走行時において必要油圧ＰＲ（Ｃ１）が最大圧である場合、必要油圧ＰＲ（Ｃ１）を、目標指令圧Ｐ（Ｃ１）および目標指令圧Ｐ（Ｃ３）として算出する。 In S106, ECU 8000 calculates the maximum pressure of necessary hydraulic pressure PR calculated in the process of S102 as target command pressure P for each friction engagement element to be maintained in the engaged state. For example, when the required hydraulic pressure PR (C1) is the maximum pressure during traveling at the third forward gear, ECU 8000 uses the required hydraulic pressure PR (C1) as the target command pressure P (C1) and the target command pressure P (C3). ).
　Ｓ１０８にて、ＥＣＵ８０００は、算出された目標指令圧Ｐに基づいて、駆動電流Ｉを算出する。たとえば、ＥＣＵ８０００は、前進３速ギヤ段での走行時において、目標指令圧Ｐ（Ｃ１）に対応する駆動電流Ｉ（Ｃ１）および駆動電流Ｉ（Ｃ３）を、上述した図６の関係に基づいて算出する。 In S108, ECU 8000 calculates drive current I based on calculated target command pressure P. For example, ECU 8000 calculates drive current I (C1) and drive current I (C3) corresponding to target command pressure P (C1) based on the above-described relationship of FIG. 6 when traveling at the third forward gear. calculate.
　Ｓ１１０にて、ＥＣＵ８０００は、算出された各駆動電流Ｉを、各リニアソレノイドバルブに出力する。たとえば、ＥＣＵ８０００は、前進３速ギヤ段での走行時において、駆動電流Ｉ（Ｃ１）をＳＬ（１）４２１０に出力し、駆動電流Ｉ（Ｃ３）をＳＬ（３）４２３０に出力する。 In S110, ECU 8000 outputs each calculated drive current I to each linear solenoid valve. For example, ECU 8000 outputs drive current I (C1) to SL (1) 4210 and drive current I (C3) to SL (3) 4230 when traveling at the third forward gear.
　以上のような構造およびフローチャートに基づく本実施例に係る制御装置であるＥＣＵ８０００で制御される目標指令圧Ｐについて、図１０を参照しつつ説明する。 The target command pressure P controlled by the ECU 8000, which is the control device according to the present embodiment based on the structure and flowchart as described above, will be described with reference to FIG.
　前進３速ギヤ段で変速を行わずに走行している場合（Ｓ１００にてＹＥＳ）を想定する。この場合、Ｃ１クラッチ３３０１およびＣ３クラッチ３３０３をそれぞれ係合状態に維持する必要があるため、Ｃ１クラッチ３３０１の必要油圧ＰＲ（Ｃ１）とＣ３クラッチ３３０３の必要油圧ＰＲ（Ｃ３）とが算出される（Ｓ１０２）。 Suppose that the vehicle is traveling without shifting at the forward third gear (YES in S100). In this case, since it is necessary to maintain the C1 clutch 3301 and the C3 clutch 3303 in the engaged state, the required hydraulic pressure PR (C1) of the C1 clutch 3301 and the required hydraulic pressure PR (C3) of the C3 clutch 3303 are calculated ( S102).
　このとき、上述したように、必要油圧ＰＲは、各クラッチを係合状態に維持するために必要な最小圧力よりもやや高い値に算出されている。そのため、図１０に示すように、必要油圧ＰＲ（Ｃ１）および必要油圧ＰＲ（Ｃ３）は、ともに目標指令圧Ｐの最大圧Ｐ（ＭＡＸ）よりも小さい値となる。 At this time, as described above, the necessary hydraulic pressure PR is calculated to be a value slightly higher than the minimum pressure required to maintain each clutch in the engaged state. Therefore, as shown in FIG. 10, both the required oil pressure PR (C1) and the required oil pressure PR (C3) are smaller than the maximum pressure P (MAX) of the target command pressure P.
　必要油圧ＰＲ（Ｃ１）が必要油圧ＰＲ（Ｃ３）よりも高い場合、目標指令圧Ｐ（Ｃ１）および目標指令圧Ｐ（Ｃ３）は、ともに必要油圧ＰＲ（Ｃ１）となる（Ｓ１０６）。すなわち、図１０に示すように、目標指令圧Ｐ（Ｃ１）および目標指令圧Ｐ（Ｃ３）は、ともに最大圧Ｐ（ＭＡＸ）よりも小さな値に設定される。 When the required hydraulic pressure PR (C1) is higher than the required hydraulic pressure PR (C3), the target command pressure P (C1) and the target command pressure P (C3) are both the required hydraulic pressure PR (C1) (S106). That is, as shown in FIG. 10, both the target command pressure P (C1) and the target command pressure P (C3) are set to values smaller than the maximum pressure P (MAX).
　そのため、最大電流値Ｉ（ＭＡＸ）よりも小さい駆動電流Ｉ（Ｃ１）がＳＬ（１）４２１０に出力されるとともに、最大電流値Ｉ（ＭＡＸ）よりも小さい駆動電流Ｉ（Ｃ３）がＳＬ（３）４２３０に出力される（Ｓ１０８、Ｓ１１０）。 Therefore, a drive current I (C1) smaller than the maximum current value I (MAX) is output to SL (1) 4210, and a drive current I (C3) smaller than the maximum current value I (MAX) is SL (3). ) 4230 (S108, S110).
　これにより、目標指令圧Ｐ（Ｃ１）および目標指令圧Ｐ（Ｃ３）をそれぞれ最大圧Ｐ（ＭＡＸ）に設定する場合に比べて、ＳＬ（１）４２１０およびＳＬ（３）４２３０で消費される電力を低減することができる。このとき、Ｃ１クラッチ３３０１には、必要油圧ＰＲ（Ｃ１）が供給され、Ｃ３クラッチ３３０３には、必要油圧ＰＲ（Ｃ３）よりも高い油圧が供給される。そのため、Ｃ１クラッチ３３０１およびＣ３クラッチ３３０３がそれぞれ係合状態に維持される。 Thereby, compared with the case where target command pressure P (C1) and target command pressure P (C3) are each set to maximum pressure P (MAX), electric power consumed by SL (1) 4210 and SL (3) 4230 Can be reduced. At this time, the required hydraulic pressure PR (C1) is supplied to the C1 clutch 3301, and the higher hydraulic pressure than the required hydraulic pressure PR (C3) is supplied to the C3 clutch 3303. Therefore, C1 clutch 3301 and C3 clutch 3303 are each maintained in an engaged state.
　さらに、図１０に示すように、ライン圧ＰＬが必要油圧ＰＲ（Ｃ１）よりも予め定められた値αだけ高く制御される（Ｓ１０４）。そのため、ライン圧ＰＬが脈動したり指令に対してばらついたりして多少低下しても、目標指令圧Ｐ（１）および目標指令圧Ｐ（３）はライン圧ＰＬを超えにくくなり、駆動電流Ｉが最大電流値Ｉ（ＭＡＸ）になることが抑制される（図６および図７参照）。 Further, as shown in FIG. 10, the line pressure PL is controlled to be higher than the required hydraulic pressure PR (C1) by a predetermined value α (S104). Therefore, even if the line pressure PL pulsates or varies slightly with respect to the command, the target command pressure P (1) and the target command pressure P (3) are unlikely to exceed the line pressure PL, and the drive current I Is suppressed to the maximum current value I (MAX) (see FIGS. 6 and 7).
　そのため、Ｃ１クラッチ３３０１およびＣ３クラッチ３３０３を、非調圧状態ではなく調圧状態に維持することができる。これにより、その後の変速時のＣ１クラッチ３３０１およびＣ３クラッチ３３０３の制御応答性を向上させることができる。 Therefore, the C1 clutch 3301 and the C3 clutch 3303 can be maintained in the pressure regulation state instead of the non-pressure regulation state. As a result, the control responsiveness of the C1 clutch 3301 and the C3 clutch 3303 during the subsequent shift can be improved.
　すなわち、Ｃ１クラッチ３３０１およびＣ３クラッチ３３０３を係合状態に維持する場合においても、ＳＬ（１）４２１０およびＳＬ（３）４２３０の各スプールが調圧可能位置に制御されるため、その後の変速時において、スプールを調圧可能位置まで移動させるための時間（フルストローク状態においては、調圧面４２１６を図５の間隔Ａだけ移動させる時間）が不要となり、その分だけ変速時間を短縮することができる。 That is, even when the C1 clutch 3301 and the C3 clutch 3303 are maintained in the engaged state, the spools of the SL (1) 4210 and the SL (3) 4230 are controlled to the pressure adjustable positions. The time required to move the spool to the pressure adjustable position (in the full stroke state, the time required to move the pressure adjusting surface 4216 by the interval A in FIG. 5) is not required, and the shift time can be shortened accordingly.
　また、係合状態が維持されている場合においてもＳＬ（１）４２１０およびＳＬ（３）４２３０には、最大電流値Ｉ（ＭＡＸ）よりも小さい駆動電流Ｉ（Ｃ１）、Ｉ（Ｃ３）が出力されているため、その後の変速時において、駆動電流Ｉを最大電流値Ｉ（ＭＡＸ）から調圧可能な電流値（すなわちＩ（ＰＬ））まで低下させるための時間が不要となり、その分だけ変速時間を短縮することができる。 Even when the engaged state is maintained, SL (1) 4210 and SL (3) 4230 output drive currents I (C1) and I (C3) smaller than the maximum current value I (MAX). Therefore, at the time of the subsequent shift, there is no need for time to decrease the drive current I from the maximum current value I (MAX) to the current value that can be regulated (that is, I (PL)). Time can be shortened.
　以上のように、本実施例に係る制御装置によれば、ノーマルクローズ型のソレノイドバルブで摩擦係合要素を係合状態に維持する場合、各摩擦係合要素を係合状態に維持可能であってかつ目標指令圧の最大値Ｐ（ＭＡＸ）よりも小さい油圧に応じた駆動電流Ｉが各リニアソレノイドバルブに出力される。そのため、各リニアソレノイドバルブに供給される駆動電流Ｉは、最大電流値Ｉ（ＭＡＸ）よりも小さくなる。これにより、摩擦係合要素の係合状態を維持しつつ、係合状態を維持するためにリニアソレノイドバルブで消費される電力を低減することができる。 As described above, according to the control device of the present embodiment, when the friction engagement elements are maintained in the engaged state by the normally closed solenoid valve, each friction engagement element can be maintained in the engaged state. In addition, the drive current I corresponding to the hydraulic pressure smaller than the maximum value P (MAX) of the target command pressure is output to each linear solenoid valve. Therefore, the drive current I supplied to each linear solenoid valve is smaller than the maximum current value I (MAX). Thereby, the electric power consumed by the linear solenoid valve for maintaining the engaged state can be reduced while maintaining the engaged state of the friction engagement element.
　なお、本実施例においては、図９のＳ１０４の処理にて、ライン圧ＰＬを必要油圧ＰＲの最大圧よりも予め定められた値αだけ高い値に設定したが、ライン圧ＰＬを必要油圧ＰＲの最大圧そのものの値に設定するようにしてもよい。このようにしても、リニアソレノイドバルブで消費される電力を低減することができる。 In the present embodiment, the line pressure PL is set to a value higher than the maximum pressure of the required hydraulic pressure PR by a predetermined value α in the process of S104 in FIG. 9, but the line pressure PL is set to the required hydraulic pressure PR. The maximum pressure itself may be set to a value. Even in this case, the power consumed by the linear solenoid valve can be reduced.
　また、本実施例においては、非変速中において係合状態に維持すべき摩擦係合要素に対応するリニアソレノイドバルブの制御について説明したが、本発明は、非変速中であることに限定されず、たとえば、変速中においても係合状態に維持すべき摩擦係合要素（たとえば１速から５速までの間で行われる変速におけるＣ１クラッチ３３０１）に油圧を供給するソレノイドバルブの制御に、本発明を適用するようにしてもよい。 Further, in the present embodiment, the control of the linear solenoid valve corresponding to the friction engagement element that should be maintained in the engaged state during non-shifting has been described, but the present invention is not limited to being non-shifting. For example, the present invention is used to control a solenoid valve that supplies hydraulic pressure to a friction engagement element (for example, a C1 clutch 3301 in a shift performed between the first speed to the fifth speed) to be kept engaged even during a shift. May be applied.
　＜変形例＞
　上述の実施例に係るＥＣＵ８０００が実行するプログラムの制御構造を、図９のフローチャートに示す構造に代えて、図１１のフローチャートに示す構造に変更してもよい。 <Modification>
The control structure of the program executed by the ECU 8000 according to the above-described embodiment may be changed to the structure shown in the flowchart of FIG. 11 instead of the structure shown in the flowchart of FIG.
　図１１を参照して、本変形例に係る制御装置であるＥＣＵ８０００が実行するプログラムの制御構造について説明する。なお、図１１に示したフローチャートの中で、前述の図９に示したフローチャートと同じ処理については同じステップ番号を付してある。それらについての処理も同じである。したがって、それらについての詳細な説明はここでは繰返さない。 Referring to FIG. 11, a control structure of a program executed by ECU 8000 which is a control device according to this modification will be described. In the flowchart shown in FIG. 11, the same steps as those in the flowchart shown in FIG. 9 are given the same step numbers. The processing for them is the same. Therefore, detailed description thereof will not be repeated here.
　Ｓ２００にて、ＥＣＵ８０００は、係合状態に維持すべき摩擦係合要素に対応するリニアソレノイドバルブの各目標指令圧Ｐを、Ｓ１０２の処理で算出された必要油圧ＰＲに対応させてそれぞれ算出する。たとえば、ＥＣＵ８０００は、前進３速ギヤ段での走行時において、Ｓ１０２の処理で必要油圧ＰＲ（Ｃ１）および必要油圧ＰＲ（Ｃ３）が算出された場合、必要油圧ＰＲ（Ｃ１）を目標指令圧Ｐ（Ｃ１）とし、必要油圧ＰＲ（Ｃ３）を目標指令圧Ｐ（Ｃ３）とする。 In S200, ECU 8000 calculates each target command pressure P of the linear solenoid valve corresponding to the frictional engagement element to be maintained in the engaged state in correspondence with the required hydraulic pressure PR calculated in the process of S102. For example, when the required hydraulic pressure PR (C1) and the required hydraulic pressure PR (C3) are calculated in the process of S102 during traveling at the third forward gear, ECU 8000 uses the required hydraulic pressure PR (C1) as the target command pressure P. (C1), and the required oil pressure PR (C3) is set as the target command pressure P (C3).
　Ｓ２０２にて、ＥＣＵ８０００は、算出された目標指令圧Ｐごとに、各リニアソレノイドに対する駆動電流Ｉを算出する。たとえば、ＥＣＵ８０００は、前進３速ギヤ段での走行時において、目標指令圧Ｐ（Ｃ１）に対応する駆動電流Ｉ（Ｃ１）、目標指令圧Ｐ（Ｃ３）に対応する駆動電流Ｉ（Ｃ３）を、上述した図６の関係に基づいて算出する。 In S202, ECU 8000 calculates drive current I for each linear solenoid for each calculated target command pressure P. For example, ECU 8000 provides drive current I (C1) corresponding to target command pressure P (C1) and drive current I (C3) corresponding to target command pressure P (C3) during traveling at the third forward gear. , Based on the relationship of FIG. 6 described above.
　このようにすると、上述した実施例と同様に前進３速ギヤ段での走行時であって必要油圧ＰＲ（Ｃ１）が必要油圧ＰＲ（Ｃ３）よりも高い場合において、図１２に示すように、目標指令圧Ｐ（Ｃ３）を、必要油圧ＰＲ（Ｃ１）より低い必要油圧ＰＲ（Ｃ３）に設定することができる（Ｓ２００）。すなわち、上述した実施例のように目標指令圧Ｐ（Ｃ１）および目標指令圧Ｐ（Ｃ３）をともに必要油圧ＰＲ（Ｃ１）にする場合に比べて、駆動電流Ｉ（Ｃ３）をさらに低下することができる。そのため、ＳＬ（３）４２３０で消費される電力をさらに低減することができる。 In this way, as in the above-described embodiment, when traveling at the third forward gear stage and the required hydraulic pressure PR (C1) is higher than the required hydraulic pressure PR (C3), as shown in FIG. The target command pressure P (C3) can be set to a required oil pressure PR (C3) lower than the required oil pressure PR (C1) (S200). That is, the drive current I (C3) is further reduced as compared with the case where the target command pressure P (C1) and the target command pressure P (C3) are both set to the required hydraulic pressure PR (C1) as in the above-described embodiment. Can do. Therefore, the power consumed by SL (3) 4230 can be further reduced.
　今回開示された実施例はすべての点で例示であって制限的なものではないと考えられるべきである。本発明の範囲は上記した説明ではなくて請求の範囲によって示され、請求の範囲と均等の意味および範囲内でのすべての変更が含まれることが意図される。 The embodiment disclosed this time should be considered as illustrative in all points and not restrictive. The scope of the present invention is defined by the terms of the claims, rather than the description above, and is intended to include any modifications within the scope and meaning equivalent to the terms of the claims.

Claims (8)

  1. 　油圧式の変速用摩擦係合部と、駆動電流の増大に伴なって出力油圧を増大させて前記摩擦係合部に出力する調圧部とを備えた自動変速機（２０００）の制御装置であって、
    　前記摩擦係合部を係合した状態に維持する必要がある場合、前記摩擦係合部を係合した状態に維持可能な油圧であってかつ前記出力油圧の目標指令圧の最大値よりも小さい油圧を算出する算出部（８２２０）と、
    　前記算出部（８２２０）によって算出された油圧に応じた駆動電流を前記調圧部に出力して、前記調圧部を制御する調圧制御部（８２３０）とを含む、制御装置。     A control device for an automatic transmission (2000) comprising a hydraulic friction engaging portion for speed change and a pressure adjusting portion for increasing an output hydraulic pressure with an increase in driving current and outputting it to the friction engaging portion. There,
    When it is necessary to maintain the friction engagement portion in the engaged state, the hydraulic pressure can be maintained in the state in which the friction engagement portion is engaged, and is smaller than the maximum value of the target command pressure of the output hydraulic pressure. A calculation unit (8220) for calculating the hydraulic pressure;
    A control apparatus comprising: a pressure adjustment control unit (8230) that outputs a drive current corresponding to the hydraulic pressure calculated by the calculation unit (8220) to the pressure adjustment unit to control the pressure adjustment unit.
  2. 　前記摩擦係合部は、複数の摩擦係合要素（３３０１～３３０４、３３１１）を含み、前記調圧部は、前記複数の摩擦係合要素（３３０１～３３０４、３３１１）にそれぞれ対応する複数の調圧弁（４２１０、４２２０、４２３０、４２４０、４２５０）を含み、
    　前記算出部（８２２０）は、前記複数の摩擦係合要素（３３０１～３３０４、３３１１）のうち少なくとも２つ以上の摩擦係合要素を係合した状態に維持する必要がある場合、係合した状態に維持する必要がある前記２つ以上の摩擦係合要素ごとに前記油圧を算出し、
    　前記調圧制御部（８２３０）は、前記算出された油圧のうちの最大圧に応じた駆動電流を、係合した状態に維持する必要がある前記２つ以上の摩擦係合要素に対応する調圧弁に出力する、請求の範囲第１項に記載の制御装置。     The friction engagement portion includes a plurality of friction engagement elements (3301 to 3304, 3311), and the pressure adjustment portion includes a plurality of adjustments respectively corresponding to the plurality of friction engagement elements (3301 to 3304, 3311). Pressure valves (4210, 4220, 4230, 4240, 4250),
    The calculation unit (8220) is in an engaged state when it is necessary to maintain at least two friction engagement elements among the plurality of friction engagement elements (3301 to 3304, 3311) in an engaged state. Calculating the hydraulic pressure for each of the two or more friction engagement elements that need to be maintained at
    The pressure adjustment control unit (8230) adjusts the drive current corresponding to the maximum pressure of the calculated hydraulic pressures corresponding to the two or more friction engagement elements that need to be maintained in an engaged state. The control device according to claim 1, which outputs to the pressure valve.
  3. 　前記摩擦係合部は、複数の摩擦係合要素（３３０１～３３０４、３３１１）を含み、前記調圧部は、前記複数の摩擦係合要素（３３０１～３３０４、３３１１）にそれぞれ対応する複数の調圧弁（４２１０、４２２０、４２３０、４２４０、４２５０）を含み、
    　前記算出部（８２２０）は、前記複数の摩擦係合要素（３３０１～３３０４、３３１１）のうち少なくとも２つ以上の摩擦係合要素を係合した状態に維持する必要がある場合、係合した状態に維持する必要がある前記２つ以上の摩擦係合要素ごとに前記油圧を算出し、
    　前記調圧制御部（８２３０）は、前記算出された油圧に対応するそれぞれの駆動電流を、前記算出された油圧に対応するそれぞれの調圧弁に出力する、請求の範囲第１項に記載の制御装置。     The friction engagement portion includes a plurality of friction engagement elements (3301 to 3304, 3311), and the pressure adjustment portion includes a plurality of adjustments respectively corresponding to the plurality of friction engagement elements (3301 to 3304, 3311). Pressure valves (4210, 4220, 4230, 4240, 4250),
    The calculation unit (8220) is in an engaged state when it is necessary to maintain at least two friction engagement elements among the plurality of friction engagement elements (3301 to 3304, 3311) in an engaged state. Calculating the hydraulic pressure for each of the two or more friction engagement elements that need to be maintained at
    2. The control according to claim 1, wherein the pressure regulation control unit (8230) outputs each drive current corresponding to the calculated hydraulic pressure to each pressure regulating valve corresponding to the calculated hydraulic pressure. 3. apparatus.
  4. 　前記自動変速機（２０００）には、前記調圧部に入力される元圧を調圧する元圧調圧弁（４３００）が備えられ、
    　前記制御装置は、前記算出された油圧に応じた駆動電流が前記調圧部に前記調圧制御部（８２３０）によって出力された場合、前記元圧が前記算出された油圧よりも予め定められた値だけ大きくなるように、前記元圧調圧弁（４３００）を制御する元圧制御部（８２４０）をさらに含む、請求の範囲第１項に記載の制御装置。     The automatic transmission (2000) includes a source pressure regulating valve (4300) that regulates the source pressure input to the pressure regulating unit,
    In the control device, when a drive current corresponding to the calculated hydraulic pressure is output to the pressure regulating unit by the pressure regulation control unit (8230), the original pressure is determined in advance from the calculated hydraulic pressure. 2. The control device according to claim 1, further comprising an original pressure control unit (8240) that controls the original pressure regulating valve (4300) so as to increase by a value.
  5. 　油圧式の変速用摩擦係合部と、駆動電流の増大に伴なって出力油圧を増大させて前記摩擦係合部に出力する調圧部とを備えた自動変速機（２０００）を制御する制御ユニットが行なう制御方法であって、
    　前記摩擦係合部を係合した状態に維持する必要がある場合、前記摩擦係合部を係合した状態に維持可能な油圧であってかつ前記出力油圧の目標指令圧の最大値よりも小さい油圧を算出するステップと、
    　前記油圧を算出するステップで算出された油圧に応じた駆動電流を前記調圧部に出力して、前記調圧部を制御するステップとを含む、制御方法。     Control for controlling an automatic transmission (2000) including a hydraulic shift friction engagement portion and a pressure adjusting portion that increases an output hydraulic pressure as the drive current increases and outputs the pressure to the friction engagement portion. A control method performed by the unit,
    When it is necessary to maintain the friction engagement portion in the engaged state, the hydraulic pressure can be maintained in the state in which the friction engagement portion is engaged, and is smaller than the maximum value of the target command pressure of the output hydraulic pressure. Calculating the oil pressure;
    And a step of outputting a drive current corresponding to the hydraulic pressure calculated in the step of calculating the hydraulic pressure to the pressure regulating unit to control the pressure regulating unit.
  6. 　前記摩擦係合部は、複数の摩擦係合要素（３３０１～３３０４、３３１１）を含み、前記調圧部は、前記複数の摩擦係合要素（３３０１～３３０４、３３１１）にそれぞれ対応する複数の調圧弁（４２１０、４２２０、４２３０、４２４０、４２５０）を含み、
    　前記油圧を算出するステップは、前記複数の摩擦係合要素（３３０１～３３０４、３３１１）のうち少なくとも２つ以上の摩擦係合要素を係合した状態に維持する必要がある場合、係合した状態に維持する必要がある前記２つ以上の摩擦係合要素ごとに前記油圧を算出するステップを含み、
    　前記調圧部を制御するステップは、前記算出された油圧のうちの最大圧に応じた駆動電流を、前記係合した状態に維持する必要がある前記２つ以上の摩擦係合要素に対応する調圧弁に出力するステップを含む、請求の範囲第５項に記載の制御方法。     The friction engagement portion includes a plurality of friction engagement elements (3301 to 3304, 3311), and the pressure adjustment portion includes a plurality of adjustments respectively corresponding to the plurality of friction engagement elements (3301 to 3304, 3311). Pressure valves (4210, 4220, 4230, 4240, 4250),
    The step of calculating the hydraulic pressure is performed when at least two friction engagement elements among the plurality of friction engagement elements (3301 to 3304, 3311) need to be maintained in an engaged state. Calculating the hydraulic pressure for each of the two or more friction engagement elements that need to be maintained at
    The step of controlling the pressure adjusting unit corresponds to the two or more friction engagement elements that need to maintain a driving current corresponding to the maximum pressure of the calculated hydraulic pressure in the engaged state. The control method according to claim 5, comprising a step of outputting to the pressure regulating valve.
  7. 　前記摩擦係合部は、複数の摩擦係合要素（３３０１～３３０４、３３１１）を含み、前記調圧部は、前記複数の摩擦係合要素（３３０１～３３０４、３３１１）にそれぞれ対応する複数の調圧弁（４２１０、４２２０、４２３０、４２４０、４２５０）を含み、
    　前記油圧を算出するステップは、前記複数の摩擦係合要素（３３０１～３３０４、３３１１）のうち少なくとも２つ以上の摩擦係合要素を係合した状態に維持する必要がある場合、係合した状態に維持する必要がある前記２つ以上の摩擦係合要素ごとに前記油圧を算出するステップを含み、
    　前記調圧部を制御するステップは、前記算出された油圧に対応するそれぞれの駆動電流を、前記算出された油圧に対応するそれぞれの調圧弁に出力するステップを含む、請求の範囲第５項に記載の制御方法。     The friction engagement portion includes a plurality of friction engagement elements (3301 to 3304, 3311), and the pressure adjustment portion includes a plurality of adjustments respectively corresponding to the plurality of friction engagement elements (3301 to 3304, 3311). Pressure valves (4210, 4220, 4230, 4240, 4250),
    The step of calculating the hydraulic pressure is performed when at least two friction engagement elements among the plurality of friction engagement elements (3301 to 3304, 3311) need to be maintained in an engaged state. Calculating the hydraulic pressure for each of the two or more friction engagement elements that need to be maintained at
    The range of claim 5, wherein the step of controlling the pressure regulating unit includes a step of outputting each driving current corresponding to the calculated hydraulic pressure to each pressure regulating valve corresponding to the calculated hydraulic pressure. The control method described.
  8. 　前記自動変速機（２０００）には、前記調圧部に入力される元圧を調圧する元圧調圧弁（４３００）が備えられ、
    　前記制御方法は、前記油圧を算出するステップで算出された油圧に応じた駆動電流が前記調圧部を制御するステップにて前記調圧部に出力された場合、前記元圧が前記算出された油圧よりも予め定められた値だけ大きくなるように、前記元圧調圧弁（４３００）を制御するステップをさらに含む、請求の範囲第５項に記載の制御方法。     The automatic transmission (2000) includes a source pressure regulating valve (4300) that regulates the source pressure input to the pressure regulating unit,
    In the control method, when the drive current corresponding to the hydraulic pressure calculated in the step of calculating the hydraulic pressure is output to the pressure regulating unit in the step of controlling the pressure regulating unit, the original pressure is calculated. The control method according to claim 5, further comprising a step of controlling said original pressure regulating valve (4300) so as to be larger than a hydraulic pressure by a predetermined value.
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