JP2005351484A - Automatic transmission controller for vehicle - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide an automatic transmission controller for a vehicle capable of satisfactorily controlling oil pressure immediately after change of an engaging condition of a friction engaging means to suppress unevenness of gear shift shock. <P>SOLUTION: Control duty dn is increased (steps 107 to 113) along a curve for increasing change rate gradually by monotonous increase expressed by dn = k Mn + d0 after performing rapid filling treatment (step 101), computation of an initial value d0 (step 103), and computation of coefficient k (step 105). Even when oil pressure different from a design value for an oil pressure command value is outputted, oil pressure change time in a period from start to end of torque phase does not become uneven. Consequently, unevenness of gear shift shock in torque phase can be suppressed even if there is individual difference. <P>COPYRIGHT: (C)2006,JPO&NCIPI

Description

  The present invention relates to an automatic transmission control device for a vehicle, and more particularly, to an automatic transmission control device for a vehicle that executes control for reducing a shock at the initial stage of gear shifting.

  It is well known that in an automatic transmission for a vehicle, it is necessary to avoid a shock at the end of a shift due to a rapid shift and a damage to a frictional engagement element due to the excessive shift. Yes.

  For this reason, conventionally, as a technique for reducing the shock at the end of a shift, the clutch hydraulic pressure is increased at a constant gradient, and then the clutch hydraulic pressure is kept constant in the second half of the inertia phase period (see, for example, Patent Document 1). In other words, the clutch hydraulic pressure is initially increased at a large increase rate, and the increase rate is decreased during the inertia phase period (see, for example, Patent Document 2).

  Also, in a system with an accumulator, the line pressure is gradually increased by a gain according to the throttle opening at high loads, thereby increasing the clutch oil pressure increase gradient and shortening the shift period. There is also known a technique for reducing the shift shock by keeping the pressure constant to reduce the rising gradient of the clutch hydraulic pressure (Patent Document 3).

Furthermore, a technique for optimizing the starting hydraulic pressure for feedback control in the inertia phase period by learning control is also known (Patent Document 4).
JP-A 63-28359 JP-A-6-341524 JP-A-1-266353 JP-A-60-227046 JP-A-4-157258

  All these conventional technologies focus on the behavior in the inertia phase and control the inertia phase start hydraulic pressure and the inertia phase end hydraulic pressure. It was only raising.

  For this reason, for example, in a system using an accumulator, as shown in FIG. 25, when the line pressure appears higher than the appropriate value, the clutch hydraulic pressure increases gradually, resulting in a short torque phase. It ends in time, generating a steep torque change (see the dotted line in the figure) compared to a torque change at an appropriate line pressure (see the dotted line in the figure), causing the occupant to feel a shock. In addition, since the rising gradient of the clutch oil pressure during the period from the end of the torque phase to the start of feedback control of the line pressure based on the input shaft speed is large, the amount of torque change during this period is large, which also shocks the occupant. feel it. Such shock in the initial period of gear shift becomes more conspicuous as the technology for reducing shock at the end of gear shift progresses.

  On the contrary, when the line pressure appears lower than the appropriate value, the clutch hydraulic pressure increases with a gentler slope than when it is appropriate, and the start of the torque phase is delayed and the end of the torque phase is also delayed as shown by the dashed line in the figure. Further, the start of the subsequent feedback control is greatly delayed. As a result, the torque gradient near the torque phase is too gentle, and the driver feels a strong sense of deceleration. If the line pressure is significantly lower than the appropriate value, the gear shifting may not be completed even if the accumulator piston moves by a full stroke, and a sudden engagement may occur due to a subsequent rapid increase in pressure. When such a sudden engagement occurs, a large shock occurs in addition to the above-mentioned deceleration feeling.

  Such a shift in line pressure is caused by individual differences of various parts constituting the automatic transmission. Further, the torque phase start hydraulic pressure and torque phase end hydraulic pressure at the time of shifting also change due to individual differences including changes in the output state of the engine, changes with time of the friction material of the friction engagement elements, and the like. The conventional control cannot cope with the problem that the shock generated at the initial stage of the shift varies due to individual differences or aging.

  In addition, it is difficult to perform feedback control such as the inertia phase period with respect to such shock variations. This is because the transmission input shaft rotational speed sensor cannot sufficiently detect the shift behavior near the torque phase. Of course, if the output shaft torque is directly detected using a torque sensor, feedback control for the shift behavior near the torque phase is also possible. However, the additional mounting of the torque sensor leads to a significant cost increase and is not realistic.

  By the way, in Patent Document 5, in order to prevent the start of the inertia phase from being delayed too much, when the start of the inertia phase is not detected before the guard timer counts up, the line pressure that has been kept constant until then is gradually increased. The problem is solved to some extent when the line pressure is too low relative to the appropriate value.

  However, the technique described in Patent Document 5 has a problem that it cannot be handled at all when the line pressure appears high. Therefore, the present invention aims to reduce the variation in shock at the initial stage of shifting and to reduce the uncomfortable feeling at the time of shifting due to individual differences and the like, and in particular, to be able to cope sufficiently even when the line pressure appears high. For the purpose.

  In order to achieve the above object, the invention according to claim 1 is directed to an input shaft, an output shaft, friction engagement means capable of adjusting an engagement state, and a speed change according to the engagement state of the friction engagement means. A vehicular automatic transmission control device for controlling a vehicular automatic transmission having transmission means for transmitting the rotational force of the input shaft to the output shaft in a ratio, and applying pressure oil according to an instructed control amount By supplying the engagement state adjusting means for adjusting the engagement state of the friction engagement means, and the change over time of the control amount gradually increases with a monotonous increase or gradually changes with a monotonous decrease. The gist of the automatic transmission control device for a vehicle is provided with control amount calculation means for calculating the control amount so as to follow a curve in which the rate decreases.

  In the present invention configured as described above, the control amount calculation means is configured so that the change over time of the control amount is along a curve in which the rate of change gradually increases with a monotone increase or gradually decreases with a monotone decrease. The control amount is calculated, and the engagement state adjusting unit adjusts the engagement state of the friction engagement unit by supplying pressure oil according to the control amount.

  Here, when the control amount increases monotonously, the hydraulic pressure can be controlled as follows. For example, if the engagement state of the friction engagement means changes after the control amount becomes considerably large due to a characteristic change such as the line pressure, the control amount change rate (= the slope of the curve) immediately after that increases. Conversely, if the engagement state changes while the control amount is small, the control amount change rate immediately after that becomes small. Also, when the control amount monotonously decreases, if the engagement state changes while the control amount is still large, the control amount change rate immediately after that is large, and the engagement state changes after the control amount decreases. The control rate change rate immediately after that becomes small.

  The hydraulic pressure at which the engagement state of the friction engagement means changes is substantially constant, and then the control amount change rate becomes a value corresponding to the control amount when the engagement state changes. For this reason, in this invention, the hydraulic pressure immediately after the engagement state of a friction engagement means changes can be controlled favorably, and, thereby, a shift shock can be suppressed favorably.

  The invention described in claim 2 is characterized in that, in addition to the configuration described in claim 1, learning correction means for learning and correcting the curve according to the past shift state is further provided. Here, examples of the speed change state include the time required for the speed change and the total control amount of the feedback control in the case of a device that also uses the feedback control. For this reason, in the present invention, it is possible to calculate the control amount using a curve adapted to a change over time of a control system (for example, a solenoid valve) of the automatic transmission, in addition to the effect of the invention according to claim 1. Thus, there is an effect that the shift shock can be suppressed more satisfactorily for a long time.

  According to a third aspect of the invention, in addition to the configuration of the first or second aspect, the slope of the curve is corrected in accordance with the hydraulic pressure supplied to the engagement state adjusting means and the input torque. A correction means is provided. When the hydraulic pressure (for example, line pressure) supplied to the engagement state adjusting means is high, it is desirable to reduce the control amount accordingly in order to suppress torque fluctuation and shift shock. Further, when the input torque is large, even if torque fluctuation occurs to some extent, it is rarely felt as a shift shock. For this reason, the control amount can be increased or decreased relatively rapidly to shorten the shift time.

  In the present invention, the correction is made in accordance with the hydraulic pressure supplied to the engagement state adjusting means and the input torque, so in addition to the effect of the invention according to claim 1 or 2, the shift shock is further suppressed. At the same time, there is an effect that the shift time can be shortened.

  According to a fourth aspect of the present invention, in addition to the configuration according to any one of the first to third aspects, the initial value of the control amount depends on characteristics of the friction engagement means and the engagement state adjustment means. It is characterized by being set. If the initial value of the control amount is set incorrectly, the hydraulic pressure applied to the frictional engagement means will decrease once at the beginning of shifting, then increase, or increase once and then decrease. There is. However, if the initial value of the control amount is set in accordance with the characteristics of the friction engagement means and the engagement state adjustment means, such a situation can be satisfactorily prevented.

  For this reason, in the invention according to claim 4, in addition to the effect of the invention according to any one of claims 1 to 3, the irregular behavior of the hydraulic pressure is prevented, and the shift shock is more effectively suppressed. The effect of being able to occur. In particular, the invention according to claim 5, which is made for the purpose of preventing a variation in shift shock when the friction engagement element is switched from the non-engaged state to the engaged state, such as upshifting, includes the output shaft and the input. A transmission gear provided with a plurality of friction engagement elements which are provided between the shaft and engaged by hydraulic pressure, and a planetary gear mechanism in which the restraint state of the rotation element is determined by the engagement state of the plurality of friction engagement elements A mechanism, an engagement state switching means for selecting a friction engagement element that switches from the non-engagement state to the engagement state in response to the shift of the gear position among the plurality of friction engagement elements, and the engagement state switching Automatic transmission for vehicle comprising hydraulic control means for controlling the hydraulic pressure applied to the friction engagement element selected by the means, and control command output means for outputting a control command value for hydraulic control to the hydraulic control means In the control device The hydraulic control is performed so that the control command output means gradually increases the hydraulic pressure applied to the friction engagement element in the stage before the start of the inertia phase at the initial stage of the gear shift operation, and the increase rate increases with time. A shift initial hydraulic pressure control means for outputting while increasing the control command value for the means is provided.

  According to the automatic transmission control apparatus of the fifth aspect, the hydraulic pressure applied to the friction engagement element that is switched from the non-engaged state to the engaged state increases with time, and the rate of increase of the hydraulic pressure also increases with time. It grows with. Therefore, even if the actual oil pressure appears lower than the control command, the rate of increase of the oil pressure increases with time, so the torque phase start time is not significantly delayed, and the torque phase starts and ends. The period until is not too long.

  Therefore, according to the control device of the fifth aspect, when a certain friction engagement element is changed from the non-engagement state to the engagement state, the shock occurrence state at the initial stage of the shift is the actual clutch hydraulic pressure level with respect to the hydraulic pressure command value. Will no longer vary. In this case, as described in claim 6, the shift initial hydraulic pressure control means may be configured as means for outputting a control command value expressed by a higher-order function of second order or higher with respect to time. .

  In addition, as described in claim 7, in the automatic transmission control device for a vehicle according to claim 5 or 6, the shift initial hydraulic pressure control means is provided after the increase rate of the control command value reaches a predetermined value. May be configured as means for outputting a control command in which the increasing rate of the control command value is fixed to the predetermined value. With such a configuration, it is possible to eliminate a problem that the torque phase period becomes too short when the hydraulic pressure is extremely low.

  The object of the present invention can also be achieved by the following apparatus. That is, as described in claim 8, a plurality of frictional engagement elements provided between the output shaft and the input shaft and engaged by hydraulic pressure, and a rotation according to the engagement state of the plurality of frictional engagement elements. A transmission gear mechanism including a planetary gear mechanism in which a restraint state of the element is determined, and a friction engagement element that switches from the non-engagement state to the engagement state in response to switching of the gear position among the plurality of friction engagement elements Engagement state switching means for selecting the oil pressure, hydraulic pressure control means for controlling the hydraulic pressure applied to the friction engagement element selected by the engagement state switching means, and a control command value for controlling the hydraulic pressure to the hydraulic pressure control means In the vehicular automatic transmission control device including the control command output means for outputting the hydraulic pressure, the control command output means gradually applies the hydraulic pressure applied to the friction engagement element in the stage before the start of the inertia phase at the initial stage of the gear shift operation. Increase to In addition, the control command value for the hydraulic pressure control means increases along the first gradient within a predetermined time and then increases beyond the first gradient so that the increase rate becomes larger after the predetermined time elapses. A frictional engagement element is also engaged from a non-engaged state by a vehicular automatic transmission control device characterized by comprising an initial gear change hydraulic pressure control means that outputs while increasing along a second gradient having a large rate. Variations in shift shocks when switching to a state can be prevented.

  In claims 5 and 6, as explained as the reason for adopting claim 7, there is a possibility that the shock may increase when the actual hydraulic pressure is extremely low with respect to the command value. In the described device, since the first half is the first gradient and the second half is the second gradient, the increase rate of the hydraulic pressure does not increase at an accelerated rate in the second half. There is an effect that the problem of shock occurrence when the actual hydraulic pressure is extremely low with respect to the command value can be avoided.

  In the description of each device of claims 5 to 8, it has been described that the clutch hydraulic pressure with respect to the command value is higher than the appropriate value, and the case where the clutch oil pressure is low. Even when the clutch hydraulic pressure is an appropriate value, the following effects are exhibited.

  For example, when the engine torque is large during upshifting, a higher pressure is required as the torque phase starting oil pressure. Conversely, when the engine torque is low, a lower pressure is sufficient as the torque phase starting oil pressure. This is because even if the value is appropriate, the clutch hydraulic pressure is relatively higher or lower depending on the operating condition of the engine.

  More specifically, for example, if the upshift design condition is when the air conditioner is on, and the upshift condition is met while the air conditioner is off, the upshift is performed under a situation where the engine torque is higher than the design condition. This is the same situation as when the clutch oil pressure is relatively high. On the other hand, when the upshift design condition is that the air conditioner is off, if the upshift is performed while the air conditioner is on, the situation is the same as when the clutch hydraulic pressure is relatively low.

  According to each apparatus of Claims 5-8, it should just be considered that the above-mentioned effect is exhibited with respect to such relative clutch hydraulic pressure levels. Therefore, according to each of the devices described in claims 5 to 8, when the actual shift is performed with respect to the engine torque higher than the engine torque designed as the shift condition, or with a lower engine torque. Even when it is implemented, the engine torque adopted in the design can be accommodated in the same shock as when the upshift was implemented, and the effect of preventing variation in shift shock due to driving conditions is also demonstrated. You can do it.

  In addition, the present invention is also effective in the case of so-called manual shift in which a driver operates a shift lever to perform a shift regardless of a predetermined shift line. This is because recently, some vehicles equipped with an automatic transmission can be shifted like a manual gear. When a manual shift is selected in such a vehicle, an upshift may occur before the original shift line, or conversely, an upshift may occur after passing through the original shift line. → Even if there are two shifts, the relationship between the acceleration state of the vehicle and the engine torque or the transmission input shaft torque is not constant. For this reason, an appropriate clutch oil pressure cannot be set in advance. However, if control is performed to change the clutch oil pressure as in the present invention, an appropriate value can be reached in the course of oil pressure change, and any shift operation caused by manual shift is prevented from becoming excessive. it can.

  In addition, as described in claim 9, in the automatic transmission control device for a vehicle according to any one of claims 5 to 8, the shift initial control means has an initial value of the control command lower than an appropriate value. It is more preferable to provide a lower initial value setting means for setting a lower initial value that gives the engagement hydraulic pressure.

  The apparatus according to claim 9 increases particularly when the clutch oil pressure is higher than the design value with respect to the oil pressure command value, and when the clutch oil pressure with respect to the oil pressure command is the design value but the engine torque is lower than the design condition. Even if the shift is about to be performed, the initial value is set so as to give a lower engagement hydraulic pressure. This is because it is possible to avoid a situation where the torque phase suddenly advances.

  Furthermore, as described in claim 10, in the automatic transmission control device for a vehicle according to any one of claims 5 to 9, the control command output unit is configured to perform the operation before the operation of the shift initial hydraulic control unit. It is preferable to provide quick filling control means for rapidly filling the hydraulic fluid to the friction engagement element that is switched from the non-engaged state to the engaged state. This is because the oil passage to the friction engagement element in the non-engagement state is in a state in which the hydraulic oil is released, so that the state in which the hydraulic oil is filled is realized quickly so that the shift is not delayed too much. It is. In particular, in the configuration using the accumulator as described in claim 13 described later, it is extremely effective for ending the shift during the period in which the piston of the accumulator is stroked. In order to cope with the hydraulic pressure deviation in advance, it becomes even more effective when combined with control using a low hydraulic pressure command as an initial value.

  In addition, as described in claim 11, in the automatic transmission control device for a vehicle according to any one of claims 5 to 10, the hydraulic control means directly controls the pressure oil applied to the friction engagement element. The present invention can also be applied as it is to a system that employs so-called direct clutch control, such as a control means, and as described in claim 12, the hydraulic control means is an oil passage for supplying pressure oil to the friction engagement element. The present invention can also be applied to a system which is a line pressure control means for adjusting the line pressure.

  In addition, as described in Claim 13, in the system which employ | adopted line pressure control, it is good to be comprised so that the said line pressure may be applied to the said friction engagement element via an accumulator. This is because by interposing the accumulator, the change in the engagement hydraulic pressure of the friction engagement element is buffered and the smooth shift is compensated.

  Further, as described in claim 14, in the automatic transmission control device for a vehicle according to any one of claims 5 to 7, the shift initial hydraulic control command means determines the control command based on a shift result. By providing learning correction means for learning and correcting the calculation conditions, it is possible not only to eliminate variations in the progress of the torque phase, but also to prevent the time from the shift command to the start of the torque phase from varying. Is preferred.

  As described in claim 15, the learning correction means includes a time measuring means for measuring the time from the start of the increase of the initial hydraulic pressure to the start of the inertia phase which is substantially the end of the torque phase, and a reference for setting a predetermined reference time. The time setting means and an initial oil pressure changing means for changing the initial oil pressure at the start of the increase according to the difference between the time measured by the time measuring means and the reference time can be provided.

  More specifically, as described in claim 16, the initial hydraulic pressure changing means may be a means for increasing the initial hydraulic pressure as the measured time is longer than the reference time. As described above, the initial hydraulic pressure changing means may be a means for increasing the initial hydraulic pressure by an amount proportional to the difference between the measurement time and the reference time. This is because when it is difficult to enter the torque phase, the torque phase start time can be brought close to an appropriate value by increasing the increase start hydraulic pressure from the next time.

  Further, as described in claim 18, the learning correction means includes a time measuring means for measuring a time from the start of an increase in initial hydraulic pressure to the start of an inertia phase, a reference time setting means for setting a predetermined reference time, An increase rate changing means for changing the increase rate of the control command value by the initial hydraulic pressure control means according to the difference between the measurement time by the time measuring means and the reference time may be provided.

  More specifically, as described in claim 19, the increase rate changing means may be a means for increasing the increase rate as the measurement time is longer than the reference time. As described above, the initial hydraulic pressure changing means may be a means for increasing the increase rate by an amount proportional to the difference between the measurement time and the reference time. This is because it is possible to enter the torque phase earlier from the next time in a system that cannot easily enter the torque phase even by increasing the increase rate instead of the hydraulic pressure at the start of the increase.

  Further, as described in claim 21, in the automatic transmission control device for a vehicle according to claim 8, wherein the pressure increase control is performed by the first gradient and the second gradient, the shift initial hydraulic control command means However, it is preferable to provide learning correction means for learning and correcting the calculation condition of the control command based on the result of the shift.

  In this case, as described in claim 22, in the automatic transmission control device for a vehicle according to claim 21, the learning correction unit includes a time measuring unit that measures a time from the start of the increase of the initial hydraulic pressure to the start of the inertia phase. A reference time setting means for setting a predetermined reference time; and a first slope changing means for changing the first slope according to a difference between the time measured by the time measuring means and the reference time. Good.

  More specifically, as described in claim 23, the first gradient changing unit may be a unit that increases the first gradient as the measurement time is longer than the reference time. As described in Item 24, the first gradient changing means may be configured as means for increasing the first gradient by an amount proportional to a difference between the measurement time and the reference time.

  On the other hand, as described in claim 25, in the automatic transmission control device for a vehicle according to claim 21, the learning correction means includes a time measuring means for measuring a time from the start of the initial hydraulic pressure increase to the start of the inertia phase, Reference time setting means for setting a predetermined reference time, and second slope changing means for changing the second slope according to the difference between the time measured by the time measuring means and the reference time may be provided. .

  More specifically, as described in claim 26, the second gradient changing means may be a means for increasing the second gradient as the measurement time is longer than the reference time. As described in Item 27, the second gradient changing means may be means for increasing the change amount of the second gradient by an amount proportional to a difference between the measurement time and the reference time.

  Further, as described in claim 28, in the automatic transmission control device for a vehicle according to claim 21, the learning correction means includes a time measuring means for measuring a time from the start of the increase of the initial hydraulic pressure to the start of the inertia phase, Reference time setting means for setting a predetermined reference time, and gradient switching timing changing means for changing the predetermined time according to the difference between the time measured by the time measuring means and the reference time may be provided.

  More specifically, as described in claim 29, the gradient switching timing changing means may be means for shortening the predetermined time as the measured time is longer than the reference time. As described above, the slope switching timing changing means may be means for shortening the predetermined time by an amount proportional to the difference between the measurement time and the reference time.

  When learning correction is performed with the apparatus according to claim 8, since it is difficult to start the torque phase, it is possible to cope with the problem by increasing the hydraulic pressure increase rate by increasing the first gradient or the second gradient. is there. This is also because the same action and effect can be achieved by advancing the timing for switching the gradient.

[First Embodiment]
Next, embodiments of the present invention will be described with reference to the drawings. FIG. 1 is a schematic diagram showing the configuration of a vehicle automatic transmission control device to which the present invention is applied. This apparatus controls an automatic transmission 1 for an automobile. First, the automatic transmission 1 is configured as follows. In FIG. 1, only the configuration related to the upshift from the third speed to the fourth speed of the D range in the automatic transmission 1 is shown.

  As shown in the figure, the automatic transmission 1 includes a planetary gear unit 3 as a transmission unit and a clutch 5 as a friction engagement unit. An input shaft 7 to which a rotational force is input from an engine (not shown) via a torque converter is connected to a carrier 31 of the planetary gear device 3 so as to be integrally rotatable. A pinion gear 33 is rotatably connected to the tip of the carrier 31, and the pinion gear 33 is disposed between the sun gear 35 and the ring gear 37 of the planetary gear device 3. The sun gear 35 is connected to the clutch disk 51 of the clutch 5 so as to be rotatable together, and is connected to the carrier 31 via a one-way clutch 39 so as not to rotate at a higher speed than the input shaft 7. Further, the ring gear 37 is connected to a driving wheel (not shown) so as to be rotatable integrally with an output shaft 9 that transmits a rotational force via another transmission gear or the like.

  A clutch plate 53 that compresses the clutch disk 51 is fixed to the housing 11 of the automatic transmission 1. Further, the clutch 5 is provided with a hydraulic cylinder 55 that applies a soot pressure to the clutch plate 53 so as to press the clutch disc 51, and this hydraulic cylinder 55 is provided with an electromagnetic valve 57 as an engagement state adjusting means. Hydraulic pressure corresponding to the valve opening duty is supplied. More specifically, the electromagnetic valve 57 is installed at a branch position of a pressure chamber of the hydraulic cylinder 55, a line pressure supply path for supplying line pressure, and a drain path in the line pressure supply path, and duty control is performed. By controlling the communication state between the drain path and the line pressure supply path and the pressure chamber of the hydraulic cylinder 55, the hydraulic pressure of the pressure chamber of the hydraulic cylinder 55 is directly controlled, so-called direct control.

  When the automatic transmission 1 is set to the third speed of the D range, no hydraulic pressure is applied to the hydraulic cylinder 55, and the clutch disc 51 and the clutch plate 53 are not engaged. At this time, when the input shaft 7 rotates in the Tt direction, the pinion gear 33 revolves in the same direction. Then, a force to rotate in the direction opposite to the revolution direction (Tp) acts on the pinion gear 33, but this rotation is blocked by the one-way clutch 39. For this reason, the pinion gear 33, the sun gear 35, the ring gear 37, and the output shaft 9 are integrally rotated at the same speed in the same direction (To) as the input shaft 7. That is, the gear ratio according to the illustrated configuration of the automatic transmission 1 is 1.

  Next, when the upshift of the automatic transmission 1 is commanded, the hydraulic pressure is applied to the hydraulic cylinder 55, and the clutch disc 51 and the clutch plate 53 are engaged. Then, the rotation of the sun gear 35 is prevented. For this reason, the pinion gear 33 revolves in the revolving direction (the reverse direction of Tp) while revolving in the Tt direction, and the ring gear 37 and the output shaft 9 are in the same direction (To) as the input shaft 7 at a faster speed than the input shaft 7. Rotate. That is, the gear ratio according to the illustrated configuration of the automatic transmission 1 is greater than 1.

  The automatic transmission control device for a vehicle is provided with an electronic control circuit (ECU) 63 that controls opening and closing of the electromagnetic valve 57 via a drive circuit 61. The electronic control circuit 63 includes an input shaft rotational speed sensor 65 that detects the rotational speed Nt of the input shaft 7, a throttle opening sensor 66 that detects the throttle opening TA, an engine rotational speed sensor 67 that detects the engine rotational speed Ne, Detection signals from various sensors such as an oil temperature sensor 69 for detecting the oil temperature of the torque converter and a vehicle speed sensor 70 for detecting the vehicle speed Vo are input.

  This electronic control circuit 63 is constituted by a well-known microcomputer centering on CPU, ROM, and RAM. When the D range is shifted up from the 3rd speed to the 4th speed, the electromagnetic valve 57 is controlled based on the detection signals of the various sensors. An upshift control process for opening and closing is executed.

  Next, the upshift control process will be described with reference to the flowchart of FIG. The electronic control circuit 63 also executes a well-known shift control process for determining the gear position of the automatic transmission 1 based on the throttle opening degree TA, the vehicle speed Vo, and the like as a separate routine. When an upshift from the third speed to the fourth speed is instructed, this routine is executed.

  As shown in the figure, when the process is started, step 101 is executed. Here, the duty of the solenoid valve is set to 100% for a predetermined time set in advance, and the hydraulic cylinder 55 is rapidly filled with pressure oil until the piston is movable. When this rapid filling control is completed, the routine proceeds to steps 103 and 105, and based on the throttle opening degree TA, using the maps of FIGS. 3A and 3B, an initial value d0 and a coefficient of the control duty dn described later. k is calculated. The map of FIG. 3A used for calculating the initial value d0 is set based on the capacity of the clutch 5 and the characteristics of the electromagnetic valve 57. 3 (A) and 3 (B), this map uses a map for shifting from the 3rd speed to the 4th speed indicated by a one-dot chain line. The map for shifting from the second speed to the third speed is also illustrated by solid lines and wavy lines.

  In the following step 107, the control duty dn is calculated by the following equation. Then, drive control for opening the solenoid valve 57 with the control duty dn is executed by another routine (not shown).

なお、Ｍは制御開始時に０に設定される変数で、最初にこのステップへ移行したときはｄｎ＝ｄ０となる。また、ｎは予め設定された２以上の整数である。続いて、ステップ１１１にてＭの値を１加算した後ステップ１１３へ移行し、フィードバック（Ｆ／Ｂ）条件が成立しているか否かを判断する。

Note that M is a variable that is set to 0 at the start of control, and dn = d0 when the process first proceeds to this step. N is an integer greater than or equal to 2 set in advance. Subsequently, in step 111, the value of M is incremented by 1, and then the process proceeds to step 113 to determine whether or not a feedback (F / B) condition is satisfied.

  Here, the feedback condition is a condition that indicates that the feedback control of the control duty dn is possible when the clutch 5 starts to be engaged (that is, the inertia phase is started). It can be defined so as to be established when the shaft rotational speed Nt decreases by a predetermined amount. When the feedback condition is not satisfied (step 113: NO), the process proceeds to step 107 again and the above-described processing is repeated. That is, the control duty dn is increased along the nth order curve until the feedback condition is satisfied.

  When the feedback condition is satisfied (step 103: YES), the process proceeds to step 115 to execute feedback control. Here, the control duty dn is feedback-controlled so that the change rate ΔNt of the input rotational speed Nt approaches a predetermined target value. In the following step 117, it is determined whether or not the shift is completed, that is, whether or not the clutch 5 is completely engaged, and the process of step 115 is repeated until the shift is completed.

  When the shift is completed and an affirmative determination is made in step 117, the routine proceeds to step 119, where the initial value d0 and the correction amount of the coefficient k are calculated. Here, based on the time required for the entire shift and the total control amount of the feedback control (step 115), the correction amount is calculated so that the time and the control amount can be further reduced. Then, from the next processing, in steps 103 and 105, a value obtained by adding the correction amount to the value read from FIGS. 3A and 3B is calculated as the initial value d0 and the coefficient k. For example, in the next processing, the initial value d0 ′ corrected by the following equation is used as the initial value d0.

続くステップ１２１では、制御デューティｄｎを比較的急峻に１００％まで上昇させる終了制御を行って、一旦処理を終了する。このように構成された本車両用自動変速機制御装置では、図４に例示するように、変速開始後、急速充填制御が終了して（時点ａ）からフィードバック制御が可能となる（時点ｃ）までの間、ｎ次の曲線に沿って制御デューティｄｎを増加させている。このため、その間（時点ａ〜ｃ）、制御デューティｄｎは、単調増加で徐々に変化率が増加する曲線に沿って変化することになり、次のようにライン圧のばらつきや電磁弁５７の特性のばらつきの影響を除去することができる。

In the following step 121, end control for increasing the control duty dn to 100% relatively steeply is performed, and the process is temporarily ended. In the vehicular automatic transmission control apparatus configured as described above, as illustrated in FIG. 4, after the start of the shift, the quick filling control is ended (time point a) and feedback control is possible (time point c). In the meantime, the control duty dn is increased along the nth-order curve. Therefore, during that time (time points a to c), the control duty dn changes along a curve in which the rate of change gradually increases with a monotonous increase, and the line pressure variation and the characteristics of the electromagnetic valve 57 are as follows. It is possible to eliminate the influence of the variation of the.

  For example, assuming that the inertia phase starts while the line pressure is high and the control duty dn is small (time point b), the control duty dn until the feedback condition is satisfied (time points b to c) after the inertia phase starts. The rate of change (= the slope of the curve) becomes smaller. Conversely, if the inertia phase starts after the line pressure is low and the control duty dn becomes considerably large, the rate of change of the control duty dn from when the inertia phase starts until the feedback condition is satisfied increases.

  That is, when the inertia phase starts (time point b), the hydraulic pressure applied to the clutch 5 is substantially constant, and the control duty dn change rate at that time is a value corresponding to the control duty dn at the start of the inertia phase. Become. For this reason, the change of the hydraulic pressure applied to the clutch 5 between the time points b to c can be made almost constant regardless of the line pressure, the characteristic change of the electromagnetic valve 57, and the like.

  For example, the higher the line pressure, the greater the influence of the change in the control duty dn on the hydraulic pressure applied to the clutch, but as described above, the higher the line pressure, the smaller the change rate of the control duty dn. Therefore, the fluctuation ΔT of the output torque To can be reduced and the shift shock can be satisfactorily suppressed.

  In FIG. 4, a case where the control duty dn is linearly increased until the inertia phase starts is indicated by a wavy line as a comparative example. In this case, the change in the line pressure and the electromagnetic valve characteristic is directly reflected in the hydraulic pressure applied to the clutch 5 from the start of the inertia phase (time point b ′) to when the feedback control is possible (time point c ′). Is done. Therefore, as illustrated in FIG. 4, when the line pressure is high, the fluctuation ΔT of the output torque To increases, and the shift shock cannot be suppressed.

  FIG. 5 illustrates a correspondence relationship between the line pressure PL and the variation ΔT of the output torque To. In this embodiment illustrated by ● in FIG. 5, ΔT hardly changes even if the line pressure PL varies. On the other hand, in the comparative example illustrated by ◯, the change in the line pressure PL is reflected as it is in ΔT, and the shift shock cannot be prevented satisfactorily. In the comparative example, ΔT also decreases when the line pressure PL decreases.

  In the above embodiment, the initial value d0 and the coefficient k are learned and corrected in accordance with the time required for shifting and the total control amount of feedback control (step 119). For this reason, assuming that the curve is adapted to changes over time of the electromagnetic valve 57 and the like, the shift shock can be suppressed more satisfactorily over a long period of time.

  Further, in the above embodiment, the map for calculating the initial value d0 is set based on the capacity of the clutch 5 and the characteristics of the electromagnetic valve 57. If the initial value d0 is set incorrectly, the hydraulic pressure of the clutch 5 may once decrease and then increase at the beginning of shifting, but the initial value d0 may be based on the capacity of the clutch 5 and the characteristics of the electromagnetic valve 57. If this is set, the occurrence of such a situation can be satisfactorily prevented. Therefore, the shift shock can be more effectively suppressed.

  In the above embodiment, steps 107 and 111 correspond to the control amount calculation means, and step 119 corresponds to the learning correction means. The present invention is not limited to the above-described embodiment, and can be implemented in various forms without departing from the gist of the present invention.

  For example, in the above embodiment, the control duty dn is increased along the nth order curve. However, as long as the rate of change gradually increases with a monotonic increase such as an exponential function, the control duty dn is increased along various curves. Can be increased. Also in this case, substantially the same operation and effect as the above embodiment can be obtained.

  In the above embodiment, the coefficient k is set based on the throttle opening degree TA. However, the coefficient k is set based on the line pressure PL and the input torque Tt using a map as illustrated in FIG. May be. When the line pressure PL is high, it is desirable to reduce the control duty dn accordingly in order to suppress torque fluctuations and shift shocks. Further, when the input torque Tt is large, even if torque fluctuation occurs to some extent, it is rarely felt as a shift shock. For this reason, the control duty dn can be increased or decreased relatively abruptly to shorten the shift time. Therefore, if the coefficient k is set according to the map of FIG. 6, the shift shock can be suppressed well and the shift time can be shortened.

  This type of map is preferably prepared for each speed change mode (for example, 1st speed → 2nd speed, 2nd speed → 3rd speed,...). In this case, the input torque Tt is determined by the engine speed. Based on the outputs of the sensor 67 and the oil temperature sensor 69, it can be estimated as follows. That is, first, the capacity coefficient C of the torque converter is calculated from the oil temperature of the torque converter detected through the oil temperature sensor 69. Subsequently, based on a value obtained by squaring the engine speed Ne detected through the engine speed sensor 67, the input torque Tt is estimated by the following equation.

但し、Ｋは所定係数である。なお、図６のマップに基づいて係数ｋを算出する処理が傾斜補正手段に相当する処理である。更に、上記各実施の形態では、シフトアップ時の変速制御について説明したが、本発明は、シフトダウン時の変速制御にも適用することができる。図７は、本発明をシフトダウンに適用した場合の効果を例示するタイムチャートである。なお、車両用自動変速機制御装置の構成は図１に示したものと同様であるとする。

However, K is a predetermined coefficient. Note that the process of calculating the coefficient k based on the map of FIG. 6 is a process corresponding to the inclination correcting means. Further, in each of the above embodiments, the shift control at the time of upshifting has been described, but the present invention can also be applied to the shift control at the time of downshifting. FIG. 7 is a time chart illustrating the effect when the present invention is applied to downshifting. It is assumed that the configuration of the vehicle automatic transmission control device is the same as that shown in FIG.

In this case, the control duty dn is gradually decreased along a curve in which the rate of change gradually decreases with a monotonous decrease. Then, the hydraulic pressure applied to the clutch 5 decreases, and when the engagement starts to be released, the input rotational speed Nt increases rapidly. When the line pressure PL is high, the engagement starts to be disengaged after the control duty dn becomes considerably small, and the control duty change rate (= the slope of the curve) at that time is small. On the contrary, when the line pressure PL is low, the engagement starts to be disengaged while the control duty dn is large, so that the control duty change rate at that time is large. Since the hydraulic pressure when the engagement starts to be disengaged is substantially constant, and the subsequent control duty change rate becomes a value corresponding to the line pressure PL, the hydraulic pressure after the engagement starts to be disengaged can be satisfactorily controlled. it can. Therefore, the fluctuation ΔT of the output torque To is also reduced, and the shift shock can be suppressed satisfactorily. Similarly,
A change ΔT can be reduced even with respect to a characteristic change of the electromagnetic valve 57 or the like, and a shift shock can be suppressed.

  In FIG. 7, the case where the control duty dn is linearly decreased is indicated by a wavy line as a comparative example. In this case, similarly to the above-described shift-up, the change in the line pressure PL or the like is directly reflected in the change in hydraulic pressure after the engagement starts to be disengaged, and the fluctuation ΔT of the output torque To increases.

  Furthermore, the present invention can also be applied to shift control for shifting up by disengaging the clutch and shift control for shifting down by engaging the clutch. can get.

[Second Embodiment]
FIG. 8 shows the overall configuration of the system according to the second embodiment.

  This system includes an electronic control engine 101 connected to a drive wheel 104 via an automatic transmission 102 and a differential gear 103. The electronic control engine 101 includes an engine control computer 105, which includes an engine rotation sensor 106 that detects the engine speed, a vehicle speed sensor 107 that detects a vehicle speed (the output shaft speed of the automatic transmission), and an engine throttle. Signals of a throttle sensor 108 that detects the opening and an intake air amount sensor 109 that detects the intake air amount are input.

  The engine control computer 105 determines the fuel injection amount based on the input information, issues a command to the engine 101, and supplies an ignition signal to the engine 101 (not shown). In response to this command, a fuel supply device and an ignition device (not shown) are operated, and fuel is supplied and combusted in accordance with the rotation of the engine 101 to operate the engine.

  The automatic transmission 102 includes a torque converter 110 and a transmission gear mechanism 111. As shown in FIG. 9, the transmission gear mechanism 111 incorporates various friction engagement elements such as a clutch and a brake, and is appropriately supplied with hydraulic pressure from a control valve 115 driven based on a command from the transmission control computer 114. It is a known one that realizes a shift by operating the hydraulic pressure to various friction engagement elements.

  As shown in FIG. 8, the control valve 115 includes shift control solenoids 115a and 115b that switch the path for supplying the hydraulic pressure at each shift stage in response to a command from the shift control computer 114, and a line pressure that controls the hydraulic pressure. A control solenoid 116 is provided. In this embodiment, the two shift control solenoids 115a and 115b are used. However, the number of solenoids may be increased in accordance with the number of shift stages and the internal configuration of the control valve 115. Further, a solenoid for adjusting the timing for rapid filling and discharging of the hydraulic oil at the time of shifting transition may be added. Further, the line pressure control solenoid 116 will be described as a duty solenoid in the present embodiment, but other means such as a linear solenoid may be used as long as it has a mechanism capable of changing the hydraulic pressure. The duty solenoid or linear solenoid is provided in the part where the hydraulic pressure generated by a hydraulic pump (not shown) is introduced into the pressure regulator for adjusting the line pressure, and the line pressure is controlled by adjusting the drain ratio. It has been done.

  Further, the control valve 115 incorporates an accumulator 120 as shown in FIG. The accumulator 120 is provided in an oil passage 121 that supplies hydraulic oil from the control valve 115 to the clutch CL in the transmission gear mechanism 110, and is attached to a downstream position of the orifice 122 and the check valve 123 in the oil passage 121. Yes. The accumulator 120 includes a piston 124 fitted in a stepped hole so as to be movable in the axial direction, and a coil spring 125 that presses the piston 124 upward in the drawing. A hydraulic pressure with the line pressure PL as the original pressure is always applied to the back surface of the large diameter portion of the piston 124 via the oil passage 126. The spring pressure of the coil spring 125 is set to be smaller than the line pressure PL supplied from the oil passage 126.

  This accumulator 120 raises the clutch hydraulic pressure PC with a constant gradient from the time when the oil passage 121 is filled with the hydraulic oil when the control valve 115 is opened and, for example, a constant line pressure PL is supplied to the clutch CL. Demonstrate the effect. Incorporating such an accumulator 120 in the control valve 115 is a conventional practice. However, conventionally, the line pressure before the start of feedback control at the time of shifting is controlled to be constant, but this embodiment is characterized in that line pressure control as described later is performed before the start of feedback control. That is, the accumulator 120 similar to the conventional one is used, but the feature of this embodiment is that the rising state of the clutch hydraulic pressure before starting the feedback becomes a curve as described later.

  Note that the shift control computer 114 that performs such line pressure control and the like is configured by a microcomputer including a CPU, a ROM, a RAM, and an I / O device (not shown). As shown in FIG. 8, the shift control computer 114 receives signals from the input shaft rotation sensor 117, the vehicle speed sensor 107, and the throttle sensor 108 for measuring the number of rotations of the input shaft 112, and the gear should be shifted. Judgment of whether or not the condition is satisfied and setting of the line pressure is performed.

  The engine control computer 105 and the shift control computer 114 are connected by a communication line 118 so that control information and commands can be communicated bidirectionally. The communication line 118 may use a multiple communication mechanism such as a LAN (Local Area Network), or may be a wiring that connects an input / output port of a control computer for each necessary communication.

  Next, the control contents executed by the shift control computer 114 in the present embodiment will be described. This control content is for a case where an upshift is performed. First, as shown in FIG. 11, the shift control computer 114 reads the throttle opening degree θacc, vehicle speed (= output shaft rotational speed) Nos, and input shaft rotational speed Nt necessary for shift determination (step 210). Then, it is determined whether or not there is a shift based on the shift diagram (step 220).

  As the shift diagram, the shift control computer 114 stores in advance a map using the vehicle speed and the throttle opening as parameters. As shown in FIG. 14, this map shows a shift (upshift) from the n-th speed (n = 1, 2, 3) to the n + 1-th speed and the m-th speed ( The shift from m = 2, 3, 4) to the (m-1) th speed (downshift) is configured to use different judgment lines as shown by a solid line in the case of upshift and by a broken line in the case of downshift. Has been.

  Using this shift diagram, the shift control computer 114 uses a line connecting the relational position of the vehicle speed-throttle opening in the previous calculation and the relational position in the current calculation to cross the solid or broken transmission line. If it is determined that there is a gear change. If it is determined that there is a shift, the ON / OFF states of the shift control solenoids 115a and 115b are switched so as to correspond to the new shift stage determined from the shift diagram (step 230).

  The ON / OFF states of the shift control solenoids 115a and 115b are controlled as shown in Table 1 below, for example. This relationship is also stored in advance in the shift control computer 114.

次に、タイマｔ１の値を初期化する（ステップ２４０）。このタイマｔ１は、ステップ２３０のソレノイド切換によって新たに係合されることとなる摩擦係合要素に通じる油路に対して、作動油を急速に充填するための制御に用いられる。こうしてタイマｔ１の初期化が完了すると、次に、ソレノイドバルブ１１６に対して最高油圧を達成するための駆動デューティを出力する（ステップ２５０）。これは、ライン圧ＰＬを最高油圧として、新たな油路に対して急速に作動油を充填するためである。そして、タイマｔ１の値をインクリメントし（ステップ２６０）、このタイマｔ１の値を充填制御期間として定めた所定値ＴＦと比較する（ステップ２７０）。比較の結果、ｔ１＜ＴＦ（即ち、ステップ２７０＝ＮＯ）と判定された場合には、ステップ２５０に戻り、最高油圧による充填制御を続行する。一方、ｔ１≧ＴＦ（即ち、ステップ２７０＝ＹＥＳ）と判定された場合には、ステップ２８０の初期油圧増圧制御ルーチンに進む。

Next, the timer t1 is initialized (step 240). This timer t1 is used for control for rapidly filling the oil passage leading to the friction engagement element to be newly engaged by the solenoid switching in step 230. When the initialization of the timer t1 is completed in this way, the drive duty for achieving the maximum hydraulic pressure is then output to the solenoid valve 116 (step 250). This is because the hydraulic oil is rapidly filled into the new oil passage with the line pressure PL as the maximum oil pressure. Then, the value of the timer t1 is incremented (step 260), and the value of the timer t1 is compared with a predetermined value TF determined as the filling control period (step 270). As a result of the comparison, if it is determined that t1 <TF (that is, step 270 = NO), the process returns to step 250, and the filling control with the maximum hydraulic pressure is continued. On the other hand, if it is determined that t1 ≧ TF (ie, step 270 = YES), the routine proceeds to the initial hydraulic pressure increase control routine of step 280.

  Here, if the condition determination loop of the above steps 250 to 270 is executed every 16 msec, for example, if the value of TF is set to 6, the line pressure PL is the maximum hydraulic pressure for rapid filling. Is output for 16 × 6 = 96 msec.

  In this embodiment, the hydraulic pressure for rapid filling is set to the maximum hydraulic pressure that can be output. However, the hydraulic pressure may be the maximum hydraulic pressure, and if the hydraulic pressure is higher than the starting hydraulic pressure of the initial hydraulic control described below, the mechanical qualities It may be an appropriate value determined by. Further, the predetermined time is determined by the size of the hydraulic pressure set at the time of filling and the mechanical properties, and can be set in a range of about 50 to 200 msec.

  The hydraulic pressure and time for the above filling control may be selected so that the clutch hydraulic pressure increases due to the charging command, but does not enter the torque phase at that hydraulic pressure. At that time, it is necessary to consider the response delay of the clutch hydraulic pressure with respect to the line pressure command, and an appropriate hydraulic pressure value and time can be obtained experimentally prior to control design.

  In this filling control, since the initial hydraulic pressure control to be described below is an increase from a low hydraulic pressure, it takes a lot of time to “fill” the hydraulic circuit connected to the friction engagement element with the pressurized oil. This is for preventing and shortening the time until the start of the torque phase. In particular, the structure having an accumulator in the hydraulic circuit used in this embodiment has a great effect in preventing so-called off-shelf shifting, in which shifting (inertia phase) does not end before the end of the accumulator stroke. Further, even in the configuration without an accumulator, there is an effect of shortening the time from the shift command to the start of the torque phase (and thus the shift time).

  When the filling control is completed in this way, the process proceeds to step 280, where the initial hydraulic pressure increase control is executed. This initial oil pressure increase control is executed in the procedure as shown in FIG. First, the timer t is initialized (step 2810). Subsequently, the shift start hydraulic pressure Pi is set as the line pressure PL (step 2820).

  The shift start hydraulic pressure Pi is obtained from a map using the throttle opening as a parameter. As shown in FIG. 15, two types of maps are stored in advance in the shift control computer 114, ie, for shifts indicated by a one-dot chain line and for shifts indicated by a solid line.

  Here, the present embodiment is characterized in that a lower line pressure is used as a shift start line pressure than the dotted line map Pfitt that has been conventionally set for shifting. The conventional shift map Pfitt shown in FIG. 15 is determined on the premise that the system performs as designed. Therefore, in the present embodiment, the shift start hydraulic pressure Pi is set with a line pressure lower than the design condition. As described above, one characteristic of the present embodiment is that the line pressure lower than the design condition is used as the shift start hydraulic pressure.

  Next, in step 2830, the control value of the line pressure PL at this stage is calculated based on the following equation (4). The larger the control value, the smaller the drain ratio in the duty solenoid 116. Therefore, in other words as the drain ratio, the control command for the duty solenoid 116 is calculated so that the drain ratio of the pump discharge oil in the pressure regulator is gradually decreased and the decrease ratio increases with time.

ここで、ａは変速機のメカ的特性によって決まる定数である。その設定の目安としては、図１５に点線で示した変速時最適油圧マップから得られるＰｆｉｔｔと、イナーシャ相開始までに要する設計上の時間をＴｆｉｔｔとしたとき、ａ×Ｔｆｉｔｔ²＋Ｐｉが、Ｐｆｉｔｔよりやや大きめの値となるように設定するとよい。即ち、変速機が設計通りの性能を発揮している場合に、大きな不具合が生じない様に定めるのである。なお、Ｐｉは、前述の様に図１５のマップにより求めた変速開始油圧である。また、Δｐｉは、後述する学習制御の結果としての補正項である。出荷時にはΔｐｉ＝０に設定されている。

Here, a is a constant determined by the mechanical characteristics of the transmission. As a guideline of the setting, when Pfit obtained from the optimum hydraulic pressure map at the time of shifting shown in FIG. 15 and the design time required until the start of the inertia phase is Tfit, a × Tfit ² + Pi is obtained from Pfit It should be set to a slightly larger value. That is, when the transmission is performing as designed, it is determined so as not to cause a major problem. Pi is the shift start hydraulic pressure obtained from the map of FIG. 15 as described above. Δpi is a correction term as a result of learning control described later. At the time of shipment, Δpi = 0 is set.

  Therefore, when this step 2830 is executed for the first time, t = 0 and Δpi = 0, so that PL = Pi. In subsequent steps 2840 and 2850, the drive duty of the solenoid 116 is determined based on the calculation result of the above equation (4) and output. This driving duty is determined using a map as shown in FIG. The map of FIG. 16 is also stored in advance in the shift control computer 114.

  Thereafter, the process proceeds to step 2860, where the start of inertia phase is determined based on the following equation (5).

ここで、Ｎｏｓは車速センサ１０７の出力、ｇは変速前のギア比、Ｎｔは入力軸回転数、ΔＮｔは所定の定数で例えば５０ｒｐｍといった数値が設定される。
このステップ２８６０の判定の結果、まだイナーシャ相が開始していないと判定された場合は、ステップ２８７０に進んで、タイマｔをインクリメントし、さらにステップ２８３０に戻って増圧制御を継続する。なお、このステップ２８３０〜２８６０のループ演算は、一定の周期（例えば１６ｍｓｅｃ）で繰り返し実行される。この結果、イナーシャ相が開始されるまでの間、ライン圧ＰＬは（４）式に従って、２次曲線的に増圧されていくことになる。

Here, Nos is an output of the vehicle speed sensor 107, g is a gear ratio before shifting, Nt is an input shaft rotational speed, ΔNt is a predetermined constant, and a numerical value such as 50 rpm is set.
As a result of the determination in step 2860, if it is determined that the inertia phase has not yet started, the process proceeds to step 2870, the timer t is incremented, and the process returns to step 2830 to continue the pressure increase control. Note that the loop calculation in steps 2830 to 2860 is repeatedly executed at a constant period (for example, 16 msec). As a result, the line pressure PL is increased in a quadratic curve according to the equation (4) until the inertia phase is started.

  On the other hand, if it is determined in step 2860 that the shift has proceeded to the inertia phase, the process proceeds to step 2880, where the value of timer t is stored in the memory as the initial hydraulic pressure control period Ti, and the initial hydraulic pressure increase control routine is exited. . The initial hydraulic pressure control period Ti is used for learning processing of an initial hydraulic pressure boosting control parameter described later.

  When the initial hydraulic pressure increase control is thus completed, the input shaft rotational speed feedback control is subsequently executed as shown in the main routine of FIG. 11 (step 290). In this feedback control, the time point at which the input shaft rotational speed Nt changes from rising to lowering is detected, and the line pressure is feedback controlled so that the input shaft rotational speed Nt decreases at a predetermined gradient according to the design conditions. Since this feedback control itself is the same as that conventionally performed, detailed description is omitted. This feedback control is repeatedly executed, for example, every 16 msec until it is determined in step 300 that the shift has been completed.

  In the case of the upshift described here, the end of step 300 is determined to increase again when the direction of change in the input shaft rotational speed is reversed, that is, when the rotational speed has decreased due to the progress of shifting. What is necessary is just to detect the time of turning to. Alternatively, when the difference between the value obtained by multiplying the vehicle speed by the reciprocal of the gear ratio of the stage in which the speed change of the speed change gear mechanism 111 is proceeding and the input shaft rotation speed is equal to or less than a predetermined value (for example, 50 rpm), the shift end is determined. May be.

  If it is determined in step 300 that the shift has been completed, the process proceeds to step 310 where the initial hydraulic pressure increase control parameter learning process is performed. This learning process is executed in the procedure as shown in FIG. First, in step 3110, the difference between the initial hydraulic control period Ti and the reference time TB is calculated and set as the time difference to. Next, in step 3120, the start hydraulic pressure correction amount Δpi is calculated from the time difference to. The calculation method of the start hydraulic pressure correction amount Δpi is performed using a map as shown in FIG. 13 (b) or (c). A characteristic of these maps is that when the time difference to is small, there is a dead zone in which the start hydraulic pressure correction amount Δpi is 0 (zero).

  The reason is to prevent the learning result from becoming unstable due to an erroneous detection of the torque phase required time Ti, an excessive reaction of the learning process due to a temporary change in the shift state, or the like. Since the initial hydraulic pressure increase control originally can absorb the variation of the shift shock with respect to some hydraulic pressure deviation, it is not necessary to adjust the hydraulic pressure deviation strictly by the learning process, and thus such processing can be performed.

  The reference time TB defines a time during which the pressure increase control can be completed when the hydraulic pressure is properly output and the shift proceeds as designed. (4) This corresponds to Tfit described as a guideline for setting a in equation (a). After the learning process is executed, when it becomes a state where the upshift is to be executed again, in steps 2820 and 2830, the starting hydraulic pressure is set to Pi + Δpi reflecting the learning result instead of Pi. .

  The above is the control when it is determined that the speed is changed in step 220. However, when the speed is not determined as the speed change, as shown in FIG. 16 is set (step 320), the duty value is calculated using the map of FIG. 16 (step 330), and this duty value is output to the solenoid 116 (step 340).

  Next, the function and effect of the present embodiment configured as described above will be described based on the timing chart of FIG. If it is determined in step 220 that the upshift condition has been met, step 230 is executed and a shift command is output (time T0). Then, by executing steps 250 to 270, the duty for achieving the maximum hydraulic pressure is output as the line pressure command (time T0 to T1). Thus, the line pressure PL is maintained at the maximum hydraulic pressure during the filling period TF. By this filling control, the hydraulic oil is rapidly filled in the oil passage, and the clutch hydraulic pressure is rapidly increased.

  When the filling period TF elapses, the duty is changed to a duty corresponding to the lower shift start hydraulic pressure Pi, and the line pressure PL is once reduced to Pi and then subjected to secondary control by initial hydraulic pressure increase control. It rises while drawing a curve (time T1-T4). During this time, the clutch hydraulic pressure also decreases once and then increases while drawing a quadratic curve by the action of the accumulator 120. As the initial hydraulic pressure increase control proceeds, the shift control enters the torque phase (time T2), eventually ends the torque phase and starts the inertia phase (time T3). When the input shaft rotation speed Nt starts to decrease at time T4, the control shifts to feedback control.

  Thus, according to the present embodiment, the clutch hydraulic pressure increases according to the quadratic curve until the input shaft feedback control is executed after the rapid filling. Here, the increase of the clutch hydraulic pressure becomes a curve with a larger differential coefficient at the same time as the line pressure is higher.

As a result, when the clutch hydraulic pressure is appropriate, when the clutch hydraulic pressure is high, and when the clutch hydraulic pressure is low, as shown in FIG. (Time period) t _Tr is not so different.

However, as shown in FIG. 18 (b) as a comparative example, if the torque phase period is controlled only by the action of the accumulator with the line pressure kept constant, the torque phase is maintained in a tight gradient state in a transmission with a high line pressure. Whereas the required period t _Tr is determined (in the case of the dotted line in the figure), the torque phase required period t _Tr is determined in the state of a gentle slope in the transmission with a low line pressure (in the case of the dotted line in the figure), and the torque phase is required. The period t _Tr varies greatly depending on the line pressure.

Thus, according to the present embodiment, the torque phase required period t _Tr is substantially constant regardless of whether the line pressure is high or low, and the change state of the output shaft torque is not affected by the high or low line pressure. Then, by changing the shift start hydraulic pressure from Pi to Pi + ΔPi by learning control, the torque phase start timing can be brought closer to the design condition.

  That is, as can be seen from FIG. 18A, the torque phase end time (inertia phase start time) becomes later as the line pressure is shifted to a lower side. Therefore, if the initial hydraulic pressure is increased by an amount proportional to the time exceeding the reference time TB, the inertia phase can be entered in a time close to the target reference time.

  By the way, the correction of the shift start hydraulic pressure is sufficiently effective even if it is always corrected by Δpi regardless of the throttle opening, but it is more preferable to correct the learning correction amount by the throttle opening as follows. In other words, if the inter-throttle degree at the time of the previous shift for which the start hydraulic pressure correction amount is obtained is TVO1, and the current throttle opening is TVO2, the correction amount may be calculated by the following equation (6) (see FIG. 15). .

この（６）式を用いることにより、スロットル開度に対する開始油圧Ｐｉの特性を反映した学習補正を行うことができるのである。なお、変速開始油圧を学習補正するする方法については、図１９に示す様に、時間差ｔｏに基づいて、時間差ｔｏが大きいほど増圧の２次関数の２次の係数（（４）式のａ）を大きくする様に補正量Δａを学習するようにしてもよい（ステップ３１１１，３１２１）。

By using this equation (6), it is possible to perform learning correction reflecting the characteristics of the starting oil pressure Pi with respect to the throttle opening. As for the method of learning and correcting the shift start hydraulic pressure, as shown in FIG. 19, based on the time difference to, the larger the time difference to, the higher the quadratic coefficient of the quadratic function of pressure increase ( ) May be learned so as to increase (steps 3111 and 3121).

  In this case, the line pressure PL may be calculated using the following equation (7) instead of the equation (4) used in step 2830.

この場合も、図１９（ｂ）または（ｃ）に示す様に、Δａは、所定の不感帯を持ったマップから求めることが望ましい。また、この場合にも上述したのと同様のスロツトル開度に対する補正を加えると一層よい。以上述べたような制御演算によって、変速の開始にあたって適切な油圧制御を実行し、またフィードバック制御を行って、個体差や条件変動があっても安定でばらつきの少ない変速ショックの制御が出来る。

Also in this case, as shown in FIG. 19B or 19C, it is desirable to obtain Δa from a map having a predetermined dead zone. Also in this case, it is better to add a correction to the throttle opening similar to that described above. By the control calculation as described above, appropriate hydraulic control is executed at the start of shifting, and feedback control is performed, so that shifting shock can be controlled stably and with little variation even if there are individual differences and condition fluctuations.

  Thus, according to the present embodiment, even if the line pressure deviates from the design value due to individual differences of the transmission, aging or other factors, the torque phase can be started in a state close to the design conditions, The torque phase required period can be controlled.

[Third Embodiment]
Next, a third embodiment will be described. In the third embodiment, the contents of step 280 in the second embodiment are changed as shown in FIG.

  First, the operation targeted in this embodiment will be described with reference to FIG. First, as in the second embodiment, a control period for rapid filling is provided immediately after the shift command is output. Then, as the subsequent initial hydraulic pressure increase control, first, a hydraulic pressure command with a gentle pressure increase gradient (first gradient) is output. At this time, the clutch hydraulic pressure gradually increases due to the action of the accumulator 120.

  The hydraulic pressure command for increasing the pressure according to the first gradient is output for a predetermined period T1, and thereafter, the hydraulic pressure command is output at a gradient (second gradient) that is tighter than the first gradient. . The gradient switching timing is set to come before the torque phase end time when the appropriate clutch oil pressure indicated by the solid line in FIG. 20 is output. This is to accelerate the pressure increase when the pressure is shifted to the low hydraulic pressure side.

With this setting, when the clutch oil pressure deviates to the higher side as shown by a broken line, also when shifted to the low pressure side is in each hydraulic control as indicated by a chain line, the time t _Tr of the torque phase in FIG. 20 Thus, unlike the case of the constant line pressure control shown in FIG.

  Next, the flowchart of FIG. 21 will be described. First, in step 410, the timer t is initialized. Subsequently, at step 420, the initial value of the line pressure PL is set to the pressure increase control start oil pressure Pi indicated by the one-dot chain line in FIG. Next, in step 430, the line pressure PL is calculated according to the following equation (8).

ここで、δｐ１が上述した第１の増圧勾配である。ステップ４４０、４５０で、この計算結果に基づくデューティ出力を行う。さらにその後、ステップ４６０に進んで、前述した（５）式に基づいてイナーシャ相の開始判定を行う。この判定で、まだ変速がイナーシャ相にまで進行していないと判定されたときは、ステップ４７０に進む。一方、ステップ４６０の判定で、入力軸１２の回転数が変化を始めてイナーシャ相まで変速が進行したと判定されたときには、ステップ５４０に進む。

Here, δp1 is the first pressure increase gradient described above. In steps 440 and 450, duty output based on the calculation result is performed. Thereafter, the process proceeds to step 460, where the start of inertia phase is determined based on the above-described equation (5). If it is determined in this determination that the shift has not yet advanced to the inertia phase, the process proceeds to step 470. On the other hand, if it is determined in step 460 that the rotational speed of the input shaft 12 has started to change and the shift has proceeded to the inertia phase, the process proceeds to step 540.

  In step 470, the value of the timer t is compared with the first gradient setting period T1. When the timer t is less than T1, the routine proceeds to step 480, where the timer t is incremented, and the routine returns to step 430 to continue the pressure increase control at the first gradient. This loop calculation is repeatedly executed at a constant cycle (for example, 16 msec).

  If the output period of the first gradient has passed the set period T1, but has not yet reached the start of the inertia phase (step 460 = NO, step 470 = YES), the process proceeds to step 490. In step 490, the line pressure PL is calculated according to the following equation (9).

ここで、δｐ２が上述した第２の増圧勾配であり、この値はδＰ１より大きく設定される。その後、ステップ５００，５１０において、この計算結果に基づくデューティ出力を行う。

Here, δp2 is the above-described second pressure increase gradient, and this value is set to be larger than δP1. Thereafter, in steps 500 and 510, duty output based on the calculation result is performed.

  Further, the routine proceeds to step 520, where the start of inertia phase is determined based on the equation (5). If it is determined in this determination that the speed change has not yet progressed to the inertia phase, the routine proceeds to step 530 where the timer t is incremented and the routine returns to step 490. On the other hand, when the start of the inertia phase is detected, the process proceeds to step 540. In step 540, the value of the timer t is stored in the memory Ti, and the control in step 280 is terminated. The memory Ti is used for the learning process of the initial hydraulic pressure increase control parameter, as in the second embodiment.

  In the learning process, the start oil pressure Pi, the first pressure increase gradient δp1 or the second pressure increase gradient δp2 is increased as the to is increased in accordance with the difference to between the time Ti until the inertia phase start detection and the predetermined time TB. Either a method of increasing or a method of shortening the first gradient period T1 may be adopted.

  The method of increasing pressure by the combination of linear functions as described above has the following advantages over the method of increasing pressure by a quadratic function as in the second embodiment or a function of higher order. First, when the pressure increase start oil pressure is extremely low, such as when the friction coefficient of the friction engagement element is abnormally low due to individual differences or changes over time, or when the engine torque is large, the pressure increase control is terminated. It takes time.

  At that time, when the pressure is increased by a high-order function, although the torque phase can be entered in a relatively short time from a low oil pressure, the increase in oil pressure per unit time becomes large, so the end of the torque phase proceeds abruptly. There is a risk that. By combining this with a linear function and increasing the gradient in the second half, the time to enter the torque phase is shortened in the same way as the high-order function, and the pressure increasing gradient after entering the torque phase is Since it is determined by the second gradient, it does not become as steep as the higher order function, and the problem that the torque phase ends rapidly does not occur.

  However, except for the case where there is such an extreme deviation, the torque phase disclosure time when the pressure is increased by a curve such as a quadratic function as in the second embodiment is lower when the line pressure is lower. This is advantageous.

[Fourth Embodiment]
Still another embodiment will be described.

  This fourth embodiment combines the advantages of both the second and third embodiments. As shown in FIG. 22, the fourth embodiment is initially a quadratic function as in the second embodiment. After the pressure is increased and the pressure increase gradient reaches a certain value, the pressure increase gradient is fixed and the pressure is increased in a linear function.

  That is, the pressure is increased by a quadratic function of time with the start of pressure increase, and when the pressure increase gradient reaches a predetermined value, the gradient is made constant (a linear function of a predetermined gradient). By configuring in this way, when the initial oil pressure is low, it has the same effect as the second embodiment for shortening the time to enter the torque phase, and further, by limiting the pressure increase gradient, Similar to the third embodiment, unnecessary sudden pressure increase from the start of the torque phase can be prevented, and both the advantages of the second and third embodiments are combined.

[Experimental Example 1]
Next, the results of experiments conducted on the so-called direct control system targeted by the first embodiment will be described.

  The experiment adopts a method of increasing the control duty for the direct control electromagnetic valve 57 according to a quadratic function of time after rapid filling as shown in FIG. When the line pressure PL is set to three types of 0.8 MPa, 0.7 MPa, and 0.6 MPa, the output shaft torque change is measured. Note that the speed change was 1 → 2 and the throttle opening TVO = 28%.

The result of this experiment is as shown in FIG. As is apparent from this result, even if the line pressure PL on the upstream side of the solenoid valve 57 varies due to individual differences of the transmission, the torque phase period t _Tr is substantially constant, and when the torque phase shifts to the inertia phase. It can be confirmed that the output shaft torque change ΔT is substantially constant.

In addition, as a comparative experiment, as shown in FIG. 23 (b), a case where control is performed so that the control duty after rapid filling is a constant value is also described. As can be seen from FIG. 23 (b), when the line pressure PL upstream of the solenoid valve 57 is 0.6 MPa, a shift cannot be achieved as a result. Further, when PL = 0.8 MPa, a large torque change occurred when shifting from the torque phase to the inertia phase. The torque phase period _tTr is also somewhat difficult to understand from the graph, but when PL = 0.8 MPa, it is shorter than when PL = 0.7 MPa, and there is a tendency to suddenly enter the inertia phase. .

[Experiment 2]
Next, the results of experiments conducted on a system of a type in which an accumulator is provided in an oil passage targeted by the second embodiment and the clutch hydraulic pressure is controlled by line pressure control will be described.

  The experiment adopts a method in which the control duty for the line pressure control solenoid 116 is increased according to a quadratic function of time after rapid filling, and when the initial value of the line pressure is in accordance with the design condition and only +0.05 MPa from the design condition. This is a measurement of what kind of output shaft torque change occurs in the three cases of the high case and the low case of −0.05 Pa. Note that the speed change was 1 → 2 and the throttle opening TVO = 2/8.

The result of this experiment is as shown in FIG. As is apparent from this result, even if the line pressure PL varies from the design condition due to individual differences, the torque phase period t _Tr is substantially constant, and the output shaft torque change ΔT when the torque phase is shifted to the inertia phase is also substantially equal. It can be confirmed that it is constant.

In addition, also in this experimental example 2, as a comparative experiment, the experimental result when the control duty for the line pressure control solenoid 116 after the rapid filling is controlled to a constant value is also shown in FIG. As can be seen from FIG. 24B, it can be confirmed that the torque phase period t _Tr becomes longer in an individual (in the case of −0.05 MPa) where the line pressure appears lower.

[Summary]
Although several embodiments have been described above, in any case, the shift shock of the torque phase is reduced, and the shock variation due to individual differences and changes in the driving state is absorbed, and the shift is always performed with the same sense of shift. And the driver can be prevented from feeling uncomfortable.

  It should be noted that the present invention is not limited to the above-described embodiments, and it goes without saying that the present invention can be implemented in various modes as long as it does not depart from the gist of the invention.

1 is a schematic configuration diagram showing a system of a first embodiment. It is a schematic block diagram of the automatic transmission in 1st Embodiment. It is a flowchart which shows the content of the control processing by the transmission control computer in 1st Embodiment. 3 is a graph corresponding to a shift line map used in the first embodiment. It is a graph equivalent to the line pressure calculation map used in the first embodiment. 4 is a graph corresponding to a line pressure control duty calculation map used in the first embodiment. It is a flowchart which shows the content of the control processing by the engine control computer in 1st Embodiment. It is a schematic block diagram which shows the system of 2nd Embodiment. It is a schematic block diagram which shows the gear transmission mechanism in 2nd Embodiment. It is a schematic block diagram which shows the accumulator in 2nd Embodiment. It is a flowchart which shows the content of the control processing by the transmission control computer in 2nd Embodiment. It is a flowchart which shows the content of initial hydraulic pressure increase control among the content of the control processing by the speed-change control computer in 2nd Embodiment. It is a flowchart which shows the content of the learning process among the content of the control processing by the transmission control computer in 2nd Embodiment. It is a graph equivalent to the shift line map used in 2nd Embodiment. It is a graph equivalent to the line pressure calculation map used in the second embodiment. It is a graph equivalent to the line pressure control duty calculation map used in the second embodiment. It is a timing chart explaining the operation effect of a 2nd embodiment. It is a timing chart explaining the effect of 2nd Embodiment in contrast with a prior art. It is a flowchart which shows the deformation | transformation content of the learning process in 2nd Embodiment. It is a timing chart explaining the effect which 3rd Embodiment aims. It is a flowchart which shows the content of the control processing by the transmission control computer in 3rd Embodiment. It is a timing chart explaining the contents of clutch oil pressure increase control aimed at by the fourth embodiment. It is a graph about the experiment example regarding 1st Embodiment. It is a graph about the experiment example regarding 2nd Embodiment. It is a timing chart which shows the problem of a prior art.

Explanation of symbols

  DESCRIPTION OF SYMBOLS 1 ... Automatic transmission, 3 ... Planetary gear apparatus, 5 ... Clutch, 7 ... Input shaft, 9 ... Output shaft, 51 ... Clutch disk, 53 ... Clutch plate, 55 ... Hydraulic cylinder, 57 ... Solenoid valve, 63 ... Electronic control circuit, 101 ... Electronic control engine, 102 ... Automatic transmission, 103 ... Differential gear, 104 ... Drive Wheel 105, engine control computer 106 engine speed sensor 107 vehicle speed sensor 108 throttle sensor 109 intake air amount sensor 110 torque converter 111 .. Transmission gear mechanism, 112 ... Transmission input shaft, 113 ... Output shaft, 114 ... Shift control computer, 115 ... Control valve, 115a, 115b - shift control solenoid, 116 ... line pressure control solenoid, 117 ... input shaft rotation sensor, 118 ... communication line, 120 ... accumulator.

Claims (30)

An input shaft to which the rotational force from the engine is input, an output shaft that outputs the rotational force to the drive wheel, friction engagement means that can adjust the engagement state, and the engagement state of the friction engagement means A vehicular automatic transmission control apparatus for controlling a vehicular automatic transmission having transmission means for transmitting the rotational force of the input shaft to the output shaft at a transmission ratio,
An engagement state adjusting means for adjusting the engagement state of the friction engagement means by supplying pressure oil according to the instructed control amount;
A control amount calculation means for calculating the control amount so that the change over time of the control amount follows a curve in which the change rate gradually increases with a monotonic increase or gradually decreases with a monotone decrease;
An automatic transmission control device for a vehicle, comprising: 2. The automatic transmission control apparatus for a vehicle according to claim 1, further comprising learning correction means for learning and correcting the curve according to a past shift state. 3. An inclination correcting means for correcting the inclination of the curve according to the hydraulic pressure supplied to the engagement state adjusting means and the torque applied to the input shaft. Automatic transmission control device for vehicles. 4. The vehicle automatic transmission according to claim 1, wherein an initial value of the control amount is set in accordance with characteristics of the friction engagement means and the engagement state adjustment means. Machine control device. An input shaft for inputting rotational force from the engine;
An output shaft that outputs rotational force to the drive wheel side;
A plurality of friction engagement elements provided between the output shaft and the input shaft and engaged by hydraulic pressure, and a planetary state in which a restraint state of the rotation element is determined by an engagement state of the plurality of friction engagement elements A transmission gear mechanism comprising a gear mechanism;
Engagement state switching means for selecting a friction engagement element that switches from a non-engagement state to an engagement state in response to switching of the gear position among the plurality of friction engagement elements;
Hydraulic control means for controlling the hydraulic pressure applied to the friction engagement element selected by the engagement state switching means;
In a vehicle automatic transmission control device comprising a control command output means for outputting a control command value for hydraulic control to the hydraulic control means,
The control command output means is
In the stage before the start of the inertia phase at the initial stage of the gear stage switching operation, the hydraulic pressure applied to the friction engagement element is gradually increased, and the control command value for the hydraulic control means is increased so that the increase rate increases with time. An automatic transmission control device for a vehicle, comprising: a gear shift initial hydraulic pressure control means that outputs the gear while outputting. The automatic transmission control device for a vehicle according to claim 5,
An automatic transmission control apparatus for a vehicle, wherein the shift initial hydraulic pressure control means is configured as means for outputting a control command value expressed by a higher-order function of second order or higher with respect to time. The automatic transmission control device for a vehicle according to claim 5 or 6,
The shift initial hydraulic control means is configured as means for outputting a control command in which the increase rate of the control command value is fixed to the predetermined value after the increase rate of the control command value reaches a predetermined value. A vehicle automatic transmission control device characterized by the above. An input shaft for inputting rotational force from the engine;
An output shaft that outputs rotational force to the drive wheel side;
A plurality of friction engagement elements provided between the output shaft and the input shaft and engaged by hydraulic pressure, and a planetary state in which a restraint state of the rotation element is determined by an engagement state of the plurality of friction engagement elements A transmission gear mechanism comprising a gear mechanism;
Engagement state switching means for selecting a friction engagement element that switches from a non-engagement state to an engagement state in response to switching of the gear position among the plurality of friction engagement elements;
Hydraulic control means for controlling the hydraulic pressure applied to the friction engagement element selected by the engagement state switching means;
In a vehicle automatic transmission control device comprising a control command output means for outputting a control command value for hydraulic control to the hydraulic control means,
The control command output means is
A control command for the hydraulic pressure control means so that the hydraulic pressure applied to the friction engagement element is gradually increased in the stage before the start of the inertia phase at the initial stage of the gear shift operation, and the increase rate is increased after a predetermined time has elapsed. Shift initial hydraulic control means is provided for increasing the value along a first gradient within a predetermined time and then increasing the value along a second gradient having a larger rate of increase than the first gradient. An automatic transmission control device for a vehicle characterized by the above. In the automatic transmission control device for vehicles according to any one of claims 5 to 8,
The vehicle shift initial control means includes a low initial value setting means for setting a low initial value that gives a lower engagement hydraulic pressure than an appropriate value as an initial value of the control command. Automatic transmission control device. In the automatic transmission control device for vehicles according to any one of claims 5 to 9,
Prior to the operation of the shift initial hydraulic pressure control means, the control command output means rapidly fills the friction engagement element that is switched from the non-engaged state to the engaged state with a quick filling control means. An automatic transmission control device for a vehicle, comprising: In the vehicle automatic transmission control device according to any one of claims 5 to 10,
The automatic transmission control apparatus for a vehicle, wherein the hydraulic control means is direct control means for directly controlling pressure oil applied to the friction engagement element. In the vehicle automatic transmission control device according to any one of claims 5 to 10,
The automatic transmission control apparatus for a vehicle, wherein the hydraulic control means is a line pressure control means for adjusting a line pressure of an oil passage for supplying pressure oil to the friction engagement element. The automatic transmission control device for a vehicle according to claim 12,
An automatic transmission control apparatus for a vehicle, wherein the line pressure is configured to be applied to the friction engagement element via an accumulator. In the automatic transmission control device for vehicles according to any one of claims 5 to 7,
The automatic transmission control apparatus for a vehicle, wherein the shift initial hydraulic control command means includes learning correction means for learning and correcting a calculation condition of the control command based on a result of shift. The automatic transmission control device for a vehicle according to claim 14,
The learning correction means is
A time measuring means for measuring the time from the start of the initial hydraulic pressure increase to the start of the inertia phase;
A reference time setting means for setting a predetermined reference time;
An automatic transmission control device for a vehicle, comprising: an initial hydraulic pressure change means for changing an initial hydraulic pressure at the start of increase according to a difference between a measurement time by the time measuring means and the reference time. The automatic transmission control device for a vehicle according to claim 15,
The automatic transmission control device for a vehicle, wherein the initial hydraulic pressure change means is a means for increasing the initial hydraulic pressure as the measured time is longer than the reference time. The automatic transmission control device for a vehicle according to claim 16,
The automatic transmission control apparatus for a vehicle, wherein the initial hydraulic pressure changing means is a means for increasing the initial hydraulic pressure by an amount proportional to a difference between the measurement time and the reference time. The automatic transmission control device for a vehicle according to claim 14,
The learning correction means is
A time measuring means for measuring the time from the start of the initial hydraulic pressure increase to the start of the inertia phase;
A reference time setting means for setting a predetermined reference time;
An automatic transmission for a vehicle, comprising: an increase rate changing means for changing an increase rate of a control command value by the initial hydraulic pressure control means in accordance with a difference between a measurement time by the time measuring means and the reference time. Machine control device. The automatic transmission control device for a vehicle according to claim 18,
The automatic transmission control apparatus for a vehicle, wherein the increase rate changing means is a means for increasing the increase rate as the measurement time is longer than the reference time. The automatic transmission control device for a vehicle according to claim 19,
The automatic transmission control apparatus for a vehicle, wherein the initial hydraulic pressure change means is a means for increasing the increase rate by an amount proportional to a difference between the measurement time and the reference time. The automatic transmission control device for a vehicle according to claim 8,
The automatic transmission control apparatus for a vehicle, wherein the shift initial hydraulic control command means includes learning correction means for learning and correcting a calculation condition of the control command based on a result of shift. The automatic transmission control device for a vehicle according to claim 21,
The learning correction means is
A time measuring means for measuring the time from the start of the initial hydraulic pressure increase to the start of the inertia phase;
A reference time setting means for setting a predetermined reference time;
An automatic transmission control apparatus for a vehicle, comprising: a first gradient changing unit that changes the first gradient according to a difference between a measurement time measured by the time measuring unit and the reference time. The automatic transmission control device for a vehicle according to claim 22,
The first gradient changing means is
The automatic transmission control apparatus for a vehicle, wherein the first gradient is increased as the measurement time is longer than the reference time. The automatic transmission control device for a vehicle according to claim 23,
The first gradient changing means is
An automatic transmission control apparatus for a vehicle, characterized in that the first gradient is increased by an amount proportional to a difference between the measurement time and the reference time. The automatic transmission control device for a vehicle according to claim 21,
The learning correction means is
A time measuring means for measuring the time from the start of the initial hydraulic pressure increase to the start of the inertia phase;
A reference time setting means for setting a predetermined reference time;
An automatic transmission control apparatus for a vehicle, comprising: a second gradient changing unit that changes the second gradient in accordance with a difference between a time measured by the time measuring unit and the reference time. The automatic transmission control device for a vehicle according to claim 25,
The second gradient changing means is
The vehicle automatic transmission control device according to claim 1, wherein the second gradient is increased as the measurement time is longer than the reference time. The automatic transmission control device for a vehicle according to claim 26,
The second gradient changing means is
An automatic transmission control device for a vehicle, characterized in that it is means for increasing the amount of change of the second gradient by an amount proportional to the difference between the measurement time and the reference time. The automatic transmission control device for a vehicle according to claim 21,
The learning correction means is
A time measuring means for measuring the time from the start of the initial hydraulic pressure increase to the start of the inertia phase;
A reference time setting means for setting a predetermined reference time;
An automatic transmission control device for a vehicle, comprising: a gradient switching timing changing means for changing the predetermined time in accordance with a difference between a time measured by the time measuring means and the reference time. The automatic transmission control device for a vehicle according to claim 28,
The gradient switching time changing means is
The automatic transmission control device for a vehicle, wherein the predetermined time is shortened as the measurement time is longer than the reference time. The automatic transmission control device for a vehicle according to claim 29,
The gradient switching time changing means is
An automatic transmission control device for a vehicle, characterized in that the predetermined time is shortened by an amount proportional to a difference between the measurement time and the reference time.

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