JP2007064464A - Gear shift control device of automatic transmission for vehicle - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To properly perform the second gear shift motion while inhibiting delay in response in the change of hydraulic pressure by properly switching a hydraulic pressure command value for controlling engagement/disengagement of a frictional engagement device, in multiple gear shift to transfer to the second gear shift motion based on determination of the second gear shift during the first gear shift motion. <P>SOLUTION: As a correction hydraulic pressure ΔP is determined on the basis of pre-output gradient Φ1 when an initial value of the hydraulic oil command value SPB3 is corrected after multiple gear shift output (time t<SB>2</SB>), in switching a third brake B3 from engagement control to release control in multiple down-shift of 4→3→2, actual delay in response in change of hydraulic pressure P<SB>B3</SB>in switching from engagement control to release control can be properly compensated. Thus, down-shift of 3→2 as a second gear shift can be quickly and properly performed. <P>COPYRIGHT: (C)2007,JPO&INPIT

Description

  According to the present invention, a hydraulic pressure command value for controlling the hydraulic pressure of the friction engagement device is changed to the first shift at the time of the multiple shift in which the second shift determination is made during the first shift operation and the second shift operation is executed. The present invention relates to an improvement of a shift control device for switching from a command value for operation to a command value for a second shift operation.

In a vehicle automatic transmission in which the gear stage is switched by a friction engagement device and the hydraulic pressure of hydraulic oil supplied to the friction engagement device is controlled by a hydraulic control device according to a hydraulic pressure command value, the friction engagement The second shift determination is made during the first shift operation for switching the engagement / disengagement state of the device, and the hydraulic pressure command value for the hydraulic control device is changed to the first shift at the time of multiple shift to shift to the second shift operation. Switching from the command value for the operation to the command value for the second speed change operation is performed. FIG. 9 shows that the third brake B3 engaged in the 4 → 3 downshift (first shift) in the vehicle automatic transmission 14 shown in FIGS. A shift chart of 3 → 2 downshift (second shift) is made before full engagement, and the time chart when the 3 → 2 downshift is performed is 4 → 3 downshift. The hydraulic pressure command value SPB3 of the third brake B3 that is combined and released by a 3 → 2 downshift is a value for releasing from an increasing pattern for engagement before and after the 3 → 2 downshift command (time t ₂ ). Switch to decrease pattern.

However, since the actual oil pressure P _B3 of the third brake B3 changes later than the change in the oil pressure command value SPB3, the oil pressure command value SPB3 is suddenly changed from the increase pattern to the decrease pattern as shown by the solid line in FIG. _However , the hydraulic pressure P _B3 (one-dot chain line) continues to increase for a while after that, and only after the delay time TT has elapsed from the 3 → 2 downshift command (time t ₂ ), it is lower than the hydraulic pressure at the time of the 3 → 2 downshift command. Become. Such a response delay of the change in hydraulic pressure is caused by the response of a hydraulic control device such as a linear solenoid valve that performs hydraulic control according to the hydraulic pressure command value SPB3, the inertia of the hydraulic oil due to the hydraulic change, the flow resistance of the hydraulic oil, and the like. When the change pattern of the hydraulic pressure command value SPB3 decreases from increase or reverses from decrease to increase, in other words, it is remarkable in the case of engagement → release or release → engagement of the friction engagement device. However, even in the case of increase from increase (engagement of friction engagement device → engagement) or decrease from decrease (release of friction engagement device → release), the same occurs if the change gradient of hydraulic pressure is different. If there is a response delay in the oil pressure change in this way, there is a risk that the shift time will become longer due to a time loss, or that the oil pressure will increase or decrease more than necessary, resulting in energy loss or a shift shock. is there.

On the other hand, Patent Document 1 proposes that the hydraulic pressure command value for the friction engagement device is changed stepwise from the command value for the first shift operation to the command value for the second shift operation. In this way, the change of the oil pressure at the time of switching is accelerated, and it is expected to reduce the response delay of the oil pressure change.
JP 2000-135938 A

  However, even in the technique described in Patent Document 1, since the change gradient of the oil pressure immediately before the oil pressure command value is switched to the command value for the second shift operation is not taken into account, the response delay of the oil pressure change at the time of switching is not considered. It cannot always be compensated appropriately. That is, the response delay of the hydraulic pressure change due to the switching of the hydraulic pressure command value is caused by the response of the hydraulic control device, the inertia of the hydraulic oil due to the hydraulic pressure change, the flow resistance of the hydraulic oil, etc. It also depends on the change gradient. For this reason, the second speed change operation is not always appropriate, for example, the speed change time of the second speed change operation becomes longer, or the oil pressure rises or falls more than necessary, causing energy loss or a gear change shock. Sometimes it was not done.

  The present invention has been made against the background of the above circumstances, and the object of the present invention is at the time of multiple shifts in which the second shift determination is made during the first shift operation and the shift to the second shift operation is performed. Thus, the hydraulic pressure command value for controlling the engagement and release of the friction engagement device is appropriately switched to suppress the response delay of the hydraulic pressure change so that the second shift operation is appropriately performed.

  In order to achieve such an object, the first invention includes (a) a hydraulic control device that controls hydraulic pressure of hydraulic oil supplied to a friction engagement device that switches a gear stage of an automatic transmission according to a hydraulic pressure command value. (B) The second shift determination is made during the first shift operation for switching the engagement / disengagement state of the friction engagement device by the hydraulic control device, and at the time of the multiple shift to shift to the second shift operation, And a hydraulic pressure command means for switching a hydraulic pressure command value for the hydraulic pressure control device from a command value for the first gear shift operation to a command value for the second gear shift operation. (C) When the hydraulic pressure command means switches the hydraulic pressure command value to the command value for the second shift operation, an initial value is determined based on a change gradient of the hydraulic pressure command value immediately before switching. Equipped with value calculation means I am characterized in.

  According to a second aspect of the present invention, in the shift control device for an automatic transmission for a vehicle according to the first aspect, the switching initial value calculation means takes into account the viscosity of the hydraulic oil in addition to the change gradient of the hydraulic pressure command value immediately before the switching. The initial value is determined.

  According to a third aspect of the present invention, in the shift control apparatus for an automatic transmission for a vehicle according to the first aspect, the switching initial value calculation means considers the oil temperature of the hydraulic oil in addition to the change gradient of the hydraulic pressure command value immediately before the switching. And determining the initial value.

  According to a fourth aspect of the present invention, in the shift control device for an automatic transmission for a vehicle according to any one of the first to third aspects, the switching initial value calculation means includes the change of the hydraulic pressure command value immediately before the switching. The initial value is determined in consideration of the change gradient of the command value for the second speed change operation.

  In such a shift control device for an automatic transmission for a vehicle, when the hydraulic pressure command value is switched to the command value for the second shift operation, the initial value is determined based on the change gradient of the hydraulic pressure command value immediately before switching. Therefore, it is possible to appropriately compensate for the response delay of the hydraulic pressure change at the time of switching according to the change gradient of the hydraulic pressure before the switching. In other words, the response delay of the actual oil pressure change accompanying the change of the change gradient of the oil pressure command value is caused by the response of the oil pressure control device, the inertia of the operating oil due to the oil pressure change, the flow resistance of the operating oil, etc. It depends on the change gradient of the previous hydraulic pressure, and also on the change gradient of the hydraulic pressure command value. By determining the initial value based on the change gradient of the hydraulic pressure command value immediately before switching, such actual hydraulic pressure change The response delay is properly compensated.

  As a result, the second shift operation can be performed quickly and appropriately, the shift time becomes longer due to the response delay of the hydraulic pressure change, the hydraulic pressure rises or falls more than necessary, and energy loss occurs or the shift shock occurs. Is prevented from occurring.

  In addition, the responsiveness of the hydraulic control device and the flow resistance of the hydraulic oil change depending on the viscosity of the hydraulic oil, which may affect the response of the hydraulic pressure change. In the second invention, the initial value is determined in consideration of the viscosity of the hydraulic oil. Therefore, the response delay of the hydraulic pressure change due to the viscosity of the hydraulic oil is appropriately compensated, and the second speed change operation is more appropriately performed.

  Further, the viscosity or the like changes depending on the oil temperature of the hydraulic oil, and the responsiveness of the hydraulic control device and the flow resistance of the hydraulic oil change, which also affects the responsiveness of the hydraulic pressure change. Since the initial value is determined in consideration of the temperature, the response delay of the hydraulic pressure change due to the hydraulic oil temperature is appropriately compensated, and the second speed change operation is performed more appropriately.

  In the fourth aspect of the invention, the initial value is determined in consideration of the change gradient of the command value for the second shift operation, so that the actual hydraulic pressure quickly corresponds to the command value for the second shift operation. Thus, the second speed change operation is performed more quickly and appropriately.

  The present invention is preferably applied to, for example, a planetary gear type automatic transmission in which a plurality of gear stages are established in accordance with operating states of a plurality of clutches and brakes. Also in the transmission, when the second shift determination is made during the first shift operation and the multiple shift is performed, the same friction engagement device is continuously controlled before and after the multiple shift. Can be applied as well.

  The friction engagement device is a single plate type or multi-plate type clutch or brake engaged with a hydraulic actuator such as a hydraulic cylinder, a belt type brake, etc. The hydraulic control device for controlling the hydraulic pressure has a magnitude of solenoid current. A solenoid valve such as a linear solenoid valve that continuously changes the hydraulic pressure in response to the above or an ON-OFF solenoid valve that continuously changes the hydraulic pressure by duty control of the solenoid current is preferably used. Also, a large flow type that directly controls the hydraulic pressure for the hydraulic actuator of the friction engagement device may be used, or a hydraulic control may be performed via a control valve that can be operated using the output hydraulic pressure of such a solenoid valve as a signal pressure.

  Multiple shift is a continuous upshift of 2 → 3 → 4, a continuous downshift of 4 → 3 → 2, etc., a return shift of 2 → 3 → 2, 4 → 3 → 4, etc. Various modes are possible, such as those with the addition of. As for the engagement release control of the friction engagement device at the time of this multiple shift, the present invention is particularly preferably applied in the case of engagement → release or release → engagement where the change in hydraulic pressure is large. In this case, it can also be applied to the case of release → release.

  Even if the multiple shift is automatically performed according to a predetermined shift condition such as a shift map due to ON / OFF of the accelerator (change in the output request operation) or a change in the vehicle speed, the driver's manual operation is performed by operating the shift lever or the like. It may be performed according to a speed change operation. The first shift may be an automatic shift and the second shift may be a manual shift.

  The command value for the first speed change operation and the command value for the second speed change operation are generally set so as to change the oil pressure in a predetermined change pattern determined according to the operation state such as the type of speed change and the oil temperature. In general, a gradual increase portion and a gradual decrease portion that change the hydraulic pressure at a constant change gradient (change rate) are provided to prevent shift shocks and the like. The gradient of the gradually increasing portion and gradually decreasing portion may be set to a predetermined value in advance according to the type of shift, etc., but the operating state such as oil temperature and input torque (throttle valve opening etc.) is used as a parameter. It may be determined.

  The hydraulic pressure command value before and after the multiple shift changes, for example, from the gradually decreasing portion to the gradually increasing portion or from the gradually increasing portion to the gradually decreasing portion, but it is not always necessary to directly change between them, and the multiple shift timing Various modes are possible, for example, a constant hydraulic pressure command value may be maintained for a predetermined time at the time of switching by a change pattern or the like. Further, the second shift operation in the case of multiple shifts may be different from the shift operation in the case of a normal single shift, and at least the initial value of the hydraulic pressure command value for the second shift operation is the main value. It may be determined by the switching initial value calculation means of the invention.

  The switching initial value calculation means for determining the initial value of the hydraulic pressure command value when shifting to the second speed change operation, for example, a correction hydraulic pressure (hydraulic step amount) to be increased or decreased at once with reference to the hydraulic pressure command value at that time, Although it is configured to be obtained from a predetermined map, an arithmetic expression, or the like, various modes are possible such as obtaining the initial value of the hydraulic pressure command value itself from the map or the like.

  In determining the initial value, in addition to the change gradient of the hydraulic pressure command value immediately before switching, the viscosity of the hydraulic oil, the oil temperature of the hydraulic oil, or the change gradient of the command value for the second speed change operation may be considered. Although desirable, it is also possible to determine the initial value in consideration of further parameters (physical quantities and the like) such as the actual hydraulic pressure value and its change gradient.

Hereinafter, embodiments of the present invention will be described in detail with reference to the drawings.
FIG. 1 is a skeleton diagram of a horizontally-mounted vehicle drive device such as an FF (front engine / front drive) vehicle. An output of an engine 10 constituted by an internal combustion engine such as a gasoline engine includes a torque converter 12, Via an automatic transmission 14, a differential gear device (not shown) is transmitted to drive wheels (front wheels). The engine 10 is a power source for vehicle travel, and the torque converter 12 is a fluid coupling.

  The automatic transmission 14 includes a first transmission unit 22 mainly composed of a single pinion type first planetary gear unit 20, a single pinion type second planetary gear unit 26, and a double pinion type third planetary gear unit. The second transmission unit 30, which is mainly composed of 28, is provided on the coaxial line, and the rotation of the input shaft 32 is shifted and output from the output gear 34. The input shaft 32 corresponds to the input member, and in this embodiment is the turbine shaft of the torque converter 12, and the output gear 34 corresponds to the output member, and rotates the left and right drive wheels via the differential gear device. To drive. The automatic transmission 14 is substantially symmetrical with respect to the center line, and the lower half of the center line is omitted in FIG.

  The first planetary gear unit 20 constituting the first transmission unit 22 includes three rotation elements, that is, a sun gear S1, a carrier CA1, and a ring gear R1, and the sun gear S1 is connected to the input shaft 32 to be rotationally driven. At the same time, the ring gear R1 is fixed to the case 36 through the third brake B3 so as not to rotate, whereby the carrier CA1 is decelerated and rotated with respect to the input shaft 32 as an intermediate output member. Further, the second planetary gear device 26 and the third planetary gear device 28 constituting the second transmission unit 30 are partially connected to each other to constitute four rotating elements RM1 to RM4. Specifically, the first rotating element RM1 is constituted by the sun gear S3 of the third planetary gear device 28, and the ring gear R2 of the second planetary gear device 26 and the ring gear R3 of the third planetary gear device 28 are connected to each other to perform the second rotation. The element RM2 is configured, and the carrier CA2 of the second planetary gear unit 26 and the carrier CA3 of the third planetary gear unit 28 are coupled to each other to configure a third rotating element RM3. A four-rotation element RM4 is configured. In the second planetary gear device 26 and the third planetary gear device 28, the carriers CA2 and CA3 are constituted by a common member, the ring gears R2 and R3 are constituted by a common member, and the second The pinion gear of the planetary gear device 26 is a Ravigneaux type planetary gear train that also serves as the second pinion gear of the third planetary gear device 28.

  The first rotating element RM1 (sun gear S3) is selectively coupled to the case 36 by the first brake B1 and stopped rotating, and the second rotating element RM2 (ring gears R2, R3) is selectively selected by the second brake B2. The fourth rotation element RM4 (sun gear S2) is selectively connected to the input shaft 32 via the first clutch C1, and the second rotation element RM2 (ring gears R2, R3) is connected to the case 36 and stopped. Is selectively coupled to the input shaft 32 via the second clutch C2, and the first rotating element RM1 (sun gear S3) is integrally coupled to the carrier CA1 of the first planetary gear device 20 as an intermediate output member, The third rotation element RM3 (carriers CA2, CA3) is integrally connected to the output gear 34 to output rotation.

The clutches C1, C2 and the brakes B1, B2, B3 (hereinafter simply referred to as the clutch C and the brake B unless otherwise distinguished) are hydraulic friction members that are engaged and controlled by a hydraulic actuator such as a multi-plate clutch or a band brake. 2 is engaged and released as shown in FIG. 2 by switching the hydraulic circuit by excitation or non-excitation of the linear solenoid valves SL1 to SL5 of the hydraulic control circuit 98 (see FIG. 3) or a manual valve (not shown). The state is switched, and the six forward gears and the reverse one gear are established according to the operation position (position) of the shift lever 72 (see FIG. 3). “1st” to “6th” in FIG. 2 mean the first to sixth gears for forward travel, and “Rev” is the reverse gear for the gear ratio (= input shaft rotation). (Speed N _IN / output shaft rotational speed N _OUT ) is appropriately determined by the gear ratios ρ1, ρ2, and ρ3 of the first planetary gear device 20, the second planetary gear device 26, and the third planetary gear device 28. “◯” in FIG. 2 means engagement, and a blank means release.

  The shift lever 72 is, for example, in accordance with the shift pattern shown in FIG. 4, a parking position “P”, a reverse travel position “R”, a neutral position “N”, a forward travel position “D”, “4”, “3”, “2” ”And“ L ”, and in the“ P ”and“ N ”positions, neutral is established to cut off the power transmission. However, in the“ P ”position, a mechanical parking mechanism (not shown) is used to mechanically The rotation of the drive wheel is prevented. In the “D” position, an automatic transmission mode is established in which all forward gears “1st” to “6th” are used to automatically change gears according to a predetermined shift map as shown in FIG. , “4”, “3”, “2”, “L” positions, 4 ranges for shifting at the 4th speed gear stage “4th” or less, 3 ranges for shifting at the 3rd speed gear stage “3rd” or less, Two ranges for shifting at the second speed gear stage “2nd” or less and an L range for fixing to the first speed gear stage “1st” are established. Accordingly, when the shift lever 72 is operated to these “4”, “3”, “2”, and “L” positions, the automatic transmission 14 is moved to the fourth gear stage “4th”, the third gear stage “ 3rd ",...

FIG. 3 is a block diagram for explaining a control system provided in the vehicle for controlling the engine 10 and the automatic transmission 14 of FIG. 1, and the operation amount (accelerator opening) Acc of the accelerator pedal 50 is an accelerator operation. It is detected by the quantity sensor 51. The accelerator pedal 50 is largely depressed according to the driver's requested output amount, and corresponds to an accelerator operation member, and the accelerator operation amount Acc corresponds to the requested output amount. In addition, an electronic throttle valve 56 whose opening degree θ _TH is changed by a throttle actuator 54 is provided in the intake pipe of the engine 10. In addition, an intake air amount sensor 60 for detecting an intake air quantity Q of the engine rotational speed sensor 58, the engine 10 for detecting the rotational speed NE of the engine 10, the intake for detecting the temperature T _A of intake air The air temperature sensor 62, the throttle valve 64 with an idle switch for detecting the fully closed state (idle state) of the electronic throttle valve 56 and the opening degree θ _TH, and the rotational speed (output shaft) of the output gear 34 corresponding to the vehicle speed V a vehicle speed sensor 66 for detecting the corresponding) N _OUT of the rotational speed, the cooling water temperature sensor 68 for detecting the cooling water temperature T _W of the engine 10, a brake switch 70 for detecting the presence or absence of foot brake operation, the shift lever 72 Lever position sensor 74 for detecting the lever position (operation position) _PSH of the turbine, and detecting the turbine rotation speed NT Are provided with a turbine rotation speed sensor 76, an AT oil temperature sensor 78 for detecting an AT oil temperature T _OIL that is the temperature of the hydraulic oil in the hydraulic control circuit 98, an ignition switch 82, and the like. , Engine speed NE, intake air amount Q, intake air temperature T _A , throttle valve opening θ _TH , vehicle speed V (output shaft rotation speed N _OUT ), engine coolant temperature T _W , presence / absence of brake operation, shift lever 72 A signal representing the lever position P _SH , turbine rotational speed NT, AT oil temperature T _OIL , operation position of the ignition switch 82, etc. is supplied to the electronic control unit 90. The turbine rotational speed NT is the same as the rotational speed of the input shaft 32 (input shaft rotational speed N _IN ) that is an input member.

The electronic control unit 90 includes a so-called microcomputer having a CPU, a RAM, a ROM, an input / output interface, and the like. The CPU uses a temporary storage function of the RAM, and signals according to a program stored in the ROM in advance. By performing the processing, output control of the engine 10, shift control of the automatic transmission 14, and the like are executed, and the engine control and the shift control are divided as necessary. As for the output control of the engine 10, in addition to controlling the opening and closing of the electronic throttle valve 56 by the throttle actuator 54, the fuel injection valve 92 is controlled for controlling the fuel injection amount, and the ignition device 94 such as an igniter is controlled for controlling the ignition timing. Control. Control of the electronic throttle valve 56, for example, drives the throttle actuator 54 based on the actual accelerator operation amount Acc from the relationship shown in FIG. 5, the accelerator operation amount Acc increases the throttle valve opening theta _TH enough to increase. Further, when the engine 10 is started, cranking is performed by a starter (electric motor) 96.

As for the shift control of the automatic transmission 14, for example, the gear to be shifted of the automatic transmission 14 based on the actual throttle valve opening θ _TH and the vehicle speed V from the previously stored shift diagram (shift map) shown in FIG. The gear is determined, that is, the shift determination from the current gear to the shift destination gear is executed, the shift output for starting the shift operation to the determined gear is executed, and the shift such as a change in driving force is executed. The excitation states of the linear solenoid valves SL1 to SL5 of the hydraulic control circuit 98 are continuously changed so that a shock does not occur and the durability of the friction material is not impaired. The solid line in FIG. 6 is an upshift line, and the broken line is a downshift line so that the gear can be switched to a low gear side with a large gear ratio as the vehicle speed V decreases or the throttle valve opening _θTH increases. In the figure, “1” to “6” mean the first speed gear stage “1st” to the sixth speed gear stage “6th”.

FIG. 7 is a main part of the hydraulic control circuit 98, and the hydraulic oil pressure-fed from the oil pump 40 is regulated by the relief-type first pressure regulating valve 100 to become the first line pressure PL1. The oil pump 40 is, for example, a mechanical pump that is rotationally driven by the engine 10. The first pressure regulating valve 100 regulates the first line pressure PL1 according to the turbine torque T _T, that is, the input torque T _IN of the automatic transmission 14, or the throttle valve opening θ _TH that is a substitute value thereof. The one-line pressure PL1 is supplied to the manual valve 104 that is interlocked with the shift lever 72. Then, when the shift lever 72 is operated to the forward drive position such as "D" position is a forward position pressure P _D of the same size from the manual valve 104 and the first line pressure PL1 is the linear solenoid valve SL1~SL5 Supplied. The linear solenoid valves SL1 to SL5 are disposed corresponding to the clutches C1 and C2 and the brakes B1 to B3, respectively, and the excitation state is controlled according to the solenoid current supplied from the electronic control unit 90. Thus, the hydraulic pressures P _C1 , P _C2 , P _B1 , P _B2 , P _B3 are controlled independently. Accordingly, any one of the first speed gear stage “1st” to the sixth speed gear stage “6th” can be established alternatively. The linear solenoid valves SL1 to SL5 correspond to hydraulic control devices that control the hydraulic pressures P _C1 , P _C2 , P _B1 , P _B2 , and P _B3 of the clutch C and the brake B that are friction engagement devices.

  FIG. 8 is a block diagram for explaining functions related to shift control provided in the electronic control unit 90, and includes shift determination means 120, hydraulic pressure command means 130, solenoid current conversion means 140, and solenoid current output means 142. Yes. The hydraulic pressure command means 130 includes a hydraulic pressure change pattern setting means 132, a hydraulic pressure command value output means 134, and a switching initial value correction means 136.

The shift determining means 120 determines the gear stage to be shifted based on the actual throttle valve opening θ _TH and the vehicle speed V according to the shift map shown in FIG. That is, when the shift lever 72 is operated to the “D” position, all the forward gears “1st” to “6th” are used to determine the gear to be shifted, but at the “4” position. The gear to be shifted is determined from the forward gears below the fourth gear stage “4th”, and the gear to be shifted from the forward gears below the third gear stage “3rd” is determined at the “3” position. In the “2” position, the gear to be shifted is determined from the forward gears below the second gear “2nd”. In the “1” position, the first gear is fixed at “1st”. To do.

The hydraulic pressure change pattern setting unit 132, when performing a shift according to the determination of the shift determination unit 120, determines the hydraulic pressure command value according to the type of shift, the AT oil temperature T _OIL , or the operating state such as single shift or multiple shift. The hydraulic pressure command value output means 134 outputs a hydraulic pressure command value in accordance with the change pattern. The first half of the time chart of FIG. 9, that is, the portion between time t ₁ and t ₂ , is the portion that performs the 4 → 3 downshift operation according to the shift determination, that is, releases the second clutch C2 and releases the third brake B3. In the engaged portion, “SPC2” is a hydraulic pressure command value for the linear solenoid valve SL2 that directly controls the hydraulic pressure P _C2 of the second clutch C2, and the hydraulic pressure P has a predetermined change gradient in order to prevent a shift shock due to sudden release. A gradual reduction part PDWN for lowering _C2 is provided. “SPB3” is a hydraulic pressure command value for the linear solenoid valve SL5 that directly controls the hydraulic pressure P _B3 of the third brake B3, and increases the hydraulic pressure P _B3 with a predetermined gradient to prevent a shift shock due to sudden engagement. The gradual increase part PUP is provided. The actual engagement pressures P _C2 and P _B3 change with a delay from these hydraulic command values SPC2 and SPB3. “SPC1” in FIG. 8 is a hydraulic pressure command value for the linear solenoid valve SL1 that directly controls the hydraulic pressure P _C1 of the first clutch C1, and “SPB1” is a linear solenoid valve that directly controls the hydraulic pressure P _B1 of the first brake B1. “SPB2” is a hydraulic pressure command value for the linear solenoid valve SL4 that directly controls the hydraulic pressure P _B2 of the second brake B2.

FIG. 9 shows a 3 → 2 downshift determination during the 4 → 3 downshift operation, that is, before the third brake B3 is completely engaged, for example, by depressing the accelerator pedal 50 according to the shift map of FIG. This is a case of multiple shifts in which the 3 → 2 downshift operation is immediately executed. That is, the time t ₂ in FIG. 9 is a 3 → 2 downshift time multiple shift determination is made of, together with the third brake B3 is released immediately, the first brake B1 are engaged at a predetermined timing . The hydraulic pressure command value SPB3 of the third brake B3 is immediately reduced at a predetermined change gradient according to a change pattern at the time of multiple shifts set by the hydraulic pressure change pattern setting means 132, here, a decrease pattern for release. The hydraulic pressure command value SPB1 for one brake B1 is changed according to the change pattern set by the hydraulic pressure change pattern setting means 132. In the multiple downshift of 4 → 3 → 2 shown in FIG. 9, the 4 → 3 downshift is the first shift, and the 3 → 2 downshift is the second shift. Further, the third brake B3, which is continuously engaged and disengaged across the multiple shift output (time t ₂ ), corresponds to the friction engagement device according to claim 1, and in this embodiment, the multiple shift is performed. Along with this, the engagement control is switched to the release control.

In the time chart of FIG. 9, the scales “2nd”, “3rd”, and “4th” on the vertical axis in the column of the turbine rotational speed NT are the synchronous rotational speeds of these gear stages, and the vehicle speed, that is, the output shaft rotational speed. It is obtained by multiplying N _OUT and the gear ratio of each gear stage. When the turbine rotational speed NT matches the synchronous rotational speed, it means that the gear stage is established. If it is located in the middle of these synchronous rotational speeds, it means that the gear is being shifted.

Returning to FIG. 8, the hydraulic pressure command value SPC1 and the like output from the hydraulic pressure command value output means 134 are supplied to the solenoid current conversion means 140, and according to a conversion map determined in advance based on the operating characteristics of the linear solenoid valves SL1 to SL5. Converted to solenoid current. The solenoid current is output from the solenoid current output means 142 to the linear solenoid valves SL1 to SL5, so that the hydraulic pressures P _C1 , P _C2 , P _B1 , P _B2 , P _B3 of the clutch C and the brake B are changed. Control is performed according to the hydraulic pressure command values SPC1, SPC2, SPB1, SPB2, and SPB3, respectively.

On the other hand, the switching initial value correcting means 136 is for correcting the initial value of the hydraulic pressure command value at which the change pattern is switched in accordance with the multiple shift at the time of the multiple shift as shown in FIG. That is, in the case of the multiple downshift of 4 → 3 → 2 shown in FIG. 9, if the hydraulic pressure command value SPB3 is continuously changed from the increase pattern to the decrease pattern as shown by the solid line in FIG. Since the oil pressure P _B3 changes later than the change in the oil pressure command value SPB3, it is finally lower than the oil pressure at the time of the 3 → 2 downshift command after the delay time TT has elapsed from the 3 → 2 downshift command (time t ₂ ). Become. In order to suppress such a response delay of the oil pressure change, in this embodiment, as shown in FIG. 10, the initial value of the oil pressure command value SPB3 at the time of the 3 → 2 downshift operation is decreased stepwise by the correction oil pressure ΔP. The actual hydraulic pressure P _B3 is quickly reduced with the 3 → 2 shift output, which suppresses a delay in response to the change in the hydraulic pressure P _B3 and an unnecessarily increase in the second shift. The 3 → 2 downshift is performed quickly and appropriately. FIG. 10 is an enlarged view of changes in the hydraulic pressure command value SPB3 before and after the multiple shift output (time t ₂ ) and the actual hydraulic pressure P _B3 , and the thin line indicates the hydraulic pressure command value SPB3 continuously as shown in FIG. It is a case where it is changed to.

The switching initial value correcting means 136 corresponds to switching initial value calculating means. In this embodiment, as shown in FIGS. 11 (a) to (f), at least the pre-output gradient, that is, the hydraulic pressure command value SPB3 immediately before the multiple shift output. The correction hydraulic pressure ΔP is calculated according to a predetermined map using the change gradient Φ1 (see FIG. 10) as a parameter. FIG. 11A shows a case where the corrected hydraulic pressure ΔP is obtained using only the pre-output gradient Φ1 as a parameter. FIG. 11B shows a case where the corrected hydraulic pressure ΔP is obtained using the pre-output gradient Φ1 and the viscosity of the hydraulic oil (ATF viscosity) as parameters. The viscosity of the hydraulic oil is obtained from, for example, a map or an arithmetic expression that is defined using an AT oil temperature T _OIL that is the oil temperature of the hydraulic oil as a parameter. FIG. 11C shows a case where the corrected hydraulic pressure ΔP is obtained using the pre-output gradient Φ1 and the AT oil temperature T _OIL as parameters. (D) in FIG. 11 shows the corrected hydraulic pressure ΔP using the pre-output gradient Φ1 and the post-output gradient, that is, the change gradient Φ2 (see FIG. 10) of the change pattern at the time of multiple shifts set by the hydraulic pressure change pattern setting means 132 as parameters. It is a case to ask. This output after gradient Φ2, at time t ₂ the multiple shift is outputted is set to AT oil temperature T _OIL, a hydraulic command value SPB3 like at that time as a parameter by the hydraulic change pattern setting means 132. FIG. 11E shows a case where the corrected hydraulic pressure ΔP is obtained using the pre-output gradient Φ1, the hydraulic oil viscosity (ATF viscosity), and the post-output gradient Φ2 as parameters. FIG. 11 (f) shows a case where the corrected hydraulic pressure ΔP is obtained using the pre-output gradient Φ1, the AT oil temperature T _OIL , and the post-output gradient Φ2 as parameters.

As described above, in the shift control apparatus for the automatic transmission for a vehicle according to the present embodiment, when the third brake B3 is switched from the engagement control to the release control by the multiple downshift of 4 → 3 → 2, the multiple shift output ( When correcting the initial value of the hydraulic pressure command value SPB3 after time t ₂ ), the correction hydraulic pressure ΔP is obtained based on the pre-output gradient Φ1 that is the change gradient of the hydraulic pressure command value SPB3 immediately before switching, so that it is released from the engagement control. It is possible to appropriately compensate for a response delay of a change in the actual hydraulic pressure P _{B3 at} the time of switching to control. That is, the response delay of the actual change of the hydraulic pressure P _{B3 due to} the change of the change gradient of the hydraulic pressure command value SPB3 is caused by the response of the linear solenoid valve SL5, the inertia of the hydraulic oil due to the hydraulic pressure change, the flow resistance of the hydraulic oil, and the like. Therefore, it also depends on the change gradient of the hydraulic pressure P _B3 immediately before switching, and also on the change gradient of the hydraulic pressure command value SPB3. By obtaining the corrected hydraulic pressure ΔP based on the pre-output gradient Φ1, The response delay of the actual change of the hydraulic pressure P _B3 is appropriately compensated.

As a result, the 3 → 2 downshift, which is the second shift, is performed quickly and appropriately, the shift time becomes longer due to a delay in response to the change in hydraulic pressure, or the hydraulic pressure P _B3 increases more than necessary, resulting in energy loss. Or the occurrence of a shift shock is suppressed.

Further, the response of the linear solenoid valve SL5 and the flow resistance of the hydraulic oil change depending on the viscosity of the hydraulic oil, which affects the response of the change in the hydraulic pressure P _B3 . However, the correction of (b) and (e) in FIG. In the hydraulic pressure map, the corrected hydraulic pressure ΔP is obtained in consideration of the viscosity of the hydraulic oil, so that the response delay of the change in the hydraulic pressure P _B3 due to the viscosity of the hydraulic oil is appropriately compensated for, which is the second shift 3 → The 2 downshift will be performed more appropriately.

Further, the viscosity or the like changes depending on the oil temperature of the hydraulic oil, that is, the AT oil temperature T _OIL , and the responsiveness of the linear solenoid valve SL5 and the flow resistance of the hydraulic oil change, and the responsiveness of the change in the hydraulic pressure P _B3 is affected. but the correction oil pressure map (c) and (f) of FIG. 11, since the correction oil pressure ΔP is calculated by considering the oil temperature i.e. the aT oil temperature T _oIL of the working oil, hydraulic P _B3 by the working oil temperature The response delay of the change in the value is appropriately compensated, and the 3 → 2 downshift as the second shift is more appropriately performed.

Further, in the corrected hydraulic pressure maps (d) to (f) in FIG. 11, the corrected hydraulic pressure is considered in consideration of the post-output gradient Φ2 that is the change gradient of the hydraulic pressure command value SPB3 at the time of the second shift 3 → 2 downshift. Since ΔP is obtained, the actual oil pressure P _B3 is quickly set to the oil pressure command value SPB3 for the 3 → 2 downshift, and the 3 → 2 downshift is performed more quickly and appropriately. .

  In the above embodiment, the case where the third brake B3 is switched from the engagement control to the disengagement control in the multiple downshift of 4 → 3 → 2 has been described, but the multiple up of 2 → 3 → 4 or 4 → 5 → 6 is described. When the third brake B3 is switched from engagement control to disengagement control by shift or multiple downshift of 6 → 5 → 4, the first brake is switched by multiple downshift of 3 → 2 → 1 or multiple upshift of 1 → 2 → 3. Similar change characteristics when switching B1 from engagement control to disengagement control, or when switching the third brake B3 from engagement control to disengagement control by a return shift such as 2 → 3 → 2, 4 → 3 → 4, etc. It becomes.

FIG. 12 shows a case where the third brake B3 is switched from the release control to the engagement control in the multiple downshift of 5 → 4 → 3. In this case, the hydraulic pressure command value SPB3 is switched from the decrease pattern to the increase pattern. If the initial value of the hydraulic pressure command value SPB3 after the shift output (time t ₂ ) is increased by the correction hydraulic pressure ΔP obtained using a map similar to FIG. 11, a response delay or necessary change in the hydraulic pressure P _B3 is required. The above reduction is suppressed. As a result, the 4 → 3 downshift which is the second speed change is performed quickly and appropriately, and the same effect as in the above embodiment can be obtained. 12A is a diagram corresponding to the time chart of FIG. 9, and FIG. 12B is a diagram corresponding to FIG. When the third brake B3 is switched from the release control to the engagement control by a return shift such as 3 → 4 → 3, 3 → 2 → 3, etc., the first brake B1 is released by a multiple jump shift of 6 → 4 → 2. In the case of switching from the engagement control to the engagement control, the same change characteristic is obtained.

FIG. 13 shows a case where the first brake B1 is continuously released in the multiple downshift of 6 → 5 → 4. In this case, the change gradient of the decrease pattern of the hydraulic pressure command value SPB1 is switched from large to small. If the initial value of the hydraulic pressure command value SPB1 after the shift output (time t ₂ ) is increased by the corrected hydraulic pressure ΔP obtained using a map similar to that in FIG. 11, a response delay or necessary change in the hydraulic pressure P _B1 is required. The above reduction is suppressed. As a result, the 5 → 4 downshift as the second shift is performed quickly and appropriately, and the same effect as in the above-described embodiment can be obtained. 13A is a diagram corresponding to the time chart of FIG. 9, and FIG. 13B is a diagram corresponding to FIG. When the change gradient of the decrease pattern of the hydraulic command value SPB1 is switched from small to large, the initial value of the hydraulic command value SPB1 after the multiple shift output (time t ₂ ) is decreased by the corrected hydraulic pressure ΔP. It ’s fine. Also, when the second clutch C2 is continuously released and controlled by multiple downshifts of 4 → 3 → 2, and when the first brake B1 is continuously released and controlled by multiple upshifts of 2 → 3 → 4, etc. Similar change characteristics are obtained.

FIG. 14 shows the case where the first clutch C1 is continuously controlled in the multiple downshift of 5 → 4 → 3. In this case, the change gradient of the increase pattern of the hydraulic command value SPC1 is switched from large to small. If the initial value of the hydraulic pressure command value SPC1 after the multiple shift output (time t ₂ ) is decreased by the corrected hydraulic pressure ΔP obtained using the same map as in FIG. 11, the response delay of the change in the hydraulic pressure P _C1 Increases more than necessary are suppressed. As a result, the 4 → 3 downshift which is the second speed change is performed quickly and appropriately, and the same effect as in the above embodiment can be obtained. 14A is a diagram corresponding to the time chart of FIG. 9, and FIG. 14B is a diagram corresponding to FIG. When the change gradient of the increase pattern of the hydraulic command value SPC1 is switched from small to large, the initial value of the hydraulic command value SPC1 after the multiple shift output (time t ₂ ) is increased by the correction hydraulic pressure ΔP. It ’s fine. Further, when the second clutch C2 is continuously controlled by multiple upshift of 3 → 4 → 5, when the first clutch C1 is continuously controlled by multiple jump shift of 6 → 2 → 4, Etc. have similar change characteristics.

  As mentioned above, although the Example of this invention was described in detail based on drawing, these are one Embodiment to the last, This invention is implemented in the aspect which added the various change and improvement based on the knowledge of those skilled in the art. be able to.

1 is a schematic diagram of a vehicle drive device to which the present invention is applied. FIG. 2 is a diagram illustrating engagement and disengagement states of clutches and brakes for establishing each gear stage of the automatic transmission of FIG. 1. It is a figure explaining the input-output signal of the electronic control apparatus provided in the vehicle of the Example of FIG. It is a figure which shows an example of the shift pattern of the shift lever of FIG. Is a diagram showing an example of a relationship between the accelerator operation amount Acc and the throttle valve opening theta _TH used in the throttle control performed by the electronic control unit of FIG. It is a figure which shows an example of the shift map (map) used by the shift control of the automatic transmission performed by the electronic controller of FIG. FIG. 4 is a circuit diagram illustrating a configuration of a portion related to shift control of the automatic transmission in the hydraulic control circuit of FIG. 3. It is a block diagram explaining the function regarding the shift control of the automatic transmission performed by the electronic controller of FIG. It is an example of the time chart explaining the change of the related oil pressure command values SPB3, SPB1, SPC2 and the turbine rotation speed NT when the multiple downshift of 4 → 3 → 2 is performed. FIG. 10 is an enlarged view showing an example of changes in the hydraulic pressure command value SPB3 of the third brake B3 and the actual hydraulic pressure P _B3 continuously controlled by the multiple downshift of FIG. 9. It is a figure explaining some aspects of the map when the switch initial value correction | amendment means of FIG. 8 calculates | requires correction oil pressure (DELTA) P. (a) is an example of a time chart showing a change in the hydraulic pressure command value SPB3 of the third brake B3 continuously controlled when multiple downshifts of 5 → 4 → 3 are performed, and (b) is a second time chart. is an enlarged view showing an example of a change in oil pressure command value SPB3 and actual hydraulic pressure P _B3 before and after shifting the output of a shift 4 → 3 downshift. (a) is an example of a time chart showing a change in the hydraulic pressure command value SPB1 of the first brake B1 that is continuously controlled when a multiple downshift of 6 → 5 → 4 is performed, and (b) is a second time chart. is an enlarged view showing an example of a change in oil pressure command value SPB1 and actual hydraulic pressure P _B1 before and after shifting the output of a shift 5 → 4 downshift. (a) is an example of a time chart showing a change in the hydraulic pressure command value SPC1 of the first clutch C1 that is continuously controlled when a multiple downshift of 5 → 4 → 3 is performed, and (b) is a second time chart. It is a figure which expands and shows change of oil pressure command value SPC1 before and after shifting output of 4 → 3 downshift which is shifting, and actual oil pressure P _C1 . FIG. 10 is a diagram illustrating a case where the hydraulic pressure command value SPB3 of the third brake B3 continuously controlled by the multiple downshift of FIG. 9 is continuously changed from an increasing pattern to a decreasing pattern together with an actual change of the hydraulic pressure P _B3 . .

Explanation of symbols

14: Automatic transmission 90: Electronic control device 130: Hydraulic pressure command means 136: Switching initial value correction means (switching initial value calculation means) C1, C2: Clutch (friction engagement device) B1 to B3: Brake (friction engagement device) SL1 to SL5: Linear solenoid valve (hydraulic control device) SPC1, SPC2, SPB1, SPB2, SPB3: Hydraulic pressure command value Φ1: Pre-output gradient Φ2: Post-output gradient T _OIL : AT oil temperature (hydraulic oil temperature)

Claims (4)

A hydraulic control device for controlling the hydraulic pressure of hydraulic oil supplied to the friction engagement device that switches the gear stage of the automatic transmission according to a hydraulic pressure command value;
A second shift determination is made during the first shift operation for switching the engagement / release state of the friction engagement device by the hydraulic control device, and the hydraulic control device at the time of multiple shift to shift to the second shift operation. A hydraulic pressure command means for switching a hydraulic pressure command value for the first gear shift operation from a command value for the first gear shift operation to a command value for the second gear shift operation;
In a shift control device for a vehicle automatic transmission having
The hydraulic pressure command means includes a switching initial value calculation means for determining an initial value based on a change gradient of the hydraulic pressure command value immediately before switching when switching the hydraulic pressure command value to a command value for the second speed change operation. A shift control apparatus for an automatic transmission for a vehicle, comprising: 2. The vehicle automatic according to claim 1, wherein the switching initial value calculation means determines the initial value in consideration of the viscosity of the hydraulic oil in addition to the change gradient of the hydraulic pressure command value immediately before the switching. A transmission control device for a transmission. The vehicle initial value according to claim 1, wherein the switching initial value calculation means determines the initial value in consideration of an oil temperature of the hydraulic oil in addition to a change gradient of a hydraulic pressure command value immediately before the switching. A shift control device for an automatic transmission. The switching initial value calculating means determines the initial value in consideration of the change gradient of the command value for the second shift operation in addition to the change gradient of the hydraulic pressure command value immediately before the switching. The shift control apparatus of the automatic transmission for vehicles according to any one of claims 1 to 3.

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