JP4849928B2 - Shift control device for automatic transmission - Google Patents

Description

The present invention relates to a shift control device for an automatic transmission, and more particularly to a shift control device that performs a power-on downshift.

In general, an automatic transmission automatically determines a gear position from a shift map according to driving conditions such as a vehicle speed and a throttle opening, and shifts by supplying or discharging hydraulic pressure to an engagement element. For example, when the accelerator pedal is depressed while driving at the third speed, the operating point determined by the vehicle speed and the throttle opening at that time crosses the 3-2 downshift line of the shift map, so that a shift command to the second speed is issued. The second gear is shifted by engaging one engagement element and releasing another engagement element. Performing a downshift with the throttle opening (accelerator opening) opened to some extent is called a power-on downshift.

FIG. 7 is an example of a conventional shift control method, and shows the turbine rotation speed, the command current of the engagement side engagement element, and the release side engagement element in the power-on downshift from the third speed to the second speed. The hydraulic pressures of the engagement side engagement element and the release side engagement element are indicated by broken lines. The instruction current refers to a current input to a linear solenoid valve for controlling the hydraulic pressure of the engagement element.
When a power-on downshift gear shift command is issued at time t1, the engagement-side engagement element is loosened until time t2, and then a standby pressure is output to the engagement-side engagement element at time t2. Here, the standby pressure is a hydraulic pressure that is in a standby state without having a torque capacity so as not to affect the feedback control of the disengagement side engagement element. In this example, the standby pressure is changed stepwise at times t2 to t3 and t3 to t4, but may be constant. On the other hand, if the hydraulic pressure of the disengagement side engagement element is reduced to the initial release pressure at time t1, and it is detected at time t3 that the turbine rotational speed has increased by a constant value n1 from the rotational speed at the third speed (out of synchronization detection). Then, the hydraulic pressure of the disengagement-side engagement element is feedback-controlled so that the turbine rotational speed has a predetermined time change rate. The engagement hydraulic pressure is output to the engagement side engagement element at time t4. The engagement hydraulic pressure is a hydraulic pressure that generates a predetermined torque capacity in the engagement-side engagement element, and is maintained at a constant pressure. As for the output timing t4 of the engagement hydraulic pressure, the time t6 at which the second-stage turbine rotation speed is reached is estimated from the turbine rotation speed during the shift (for example, time t3) and the time change rate thereof, and the engagement side from that time t6 is estimated. It is obtained by subtracting the hydraulic response time Tr of the engaging element. The hydraulic response time Tr is substantially equal to the time during which the engagement hydraulic pressure of the engagement side engagement element rises to the hydraulic pressure at the completion of engagement after the command current is output to the engagement side solenoid valve. Thereafter, when it is detected (synchronized prediction) that the engine speed has increased to a value lower than the turbine speed at the second speed by a predetermined value n2 at time t5, feedback control of the disengagement side engagement element is terminated and the hydraulic pressure is reduced at a predetermined gradient. In addition to performing the termination process, the termination process of increasing the hydraulic pressure of the engagement side engagement element with a predetermined gradient is performed to enter the second speed state.

FIG. 7 is an example of a power-on downshift at medium and high vehicle speeds. The time from the out-of-synchronization detection (t3) to the synchronization prediction (t5) is relatively long, and from time t6 when the turbine speed of the second speed stage is reached. This is the case where the output timing t4 of the engagement hydraulic pressure calculated by subtracting the hydraulic response time Tr is after the loss of synchronization detection (t3). Therefore, if the engagement hydraulic pressure is output at this timing t4, the gear can be shifted without generating engine rotation or a shift shock. However, in a power-on downshift at a low vehicle speed, the difference in turbine speed between the high speed stage and the low speed stage is small, and the turbine speed reaches the speed of the low speed stage in a short time. In such a case, the output timing t4 of the engagement hydraulic pressure may not be obtained.

FIG. 8 is an example of a power-on downshift at a low vehicle speed, and the same reference numerals are given to the times corresponding to FIG.
In this case as well, when an out-of-synchronization of the turbine rotational speed is detected at time t3, a time t6 to reach the second speed turbine rotational speed is estimated from the turbine rotational speed at that time t3 and its time change rate, and the time The output timing t4 of the engagement hydraulic pressure is obtained by subtracting the hydraulic response time Tr from t6. However, the output timing t4 obtained by calculation is a time before the current time (time t3) and is meaningless time. For this reason, the engagement hydraulic pressure can be output only at time t3 at the earliest, and the shift shock may increase or the engine rotation may increase due to the delay in engagement of the engagement side engagement element.

In Patent Document 1, even when the speed of increase of the input rotation speed is different during power-on downshift with clutch changeover, the application hydraulic pressure is applied at an appropriate timing to prevent shift shock and engine rotation from rising. A speed change control method has been proposed.
However, also in the shift control method described in Patent Document 1, the arrival time to the low speed stage is estimated from the turbine rotation speed during the shift and the time change rate thereof, and the engagement-side engagement element is estimated from the estimated arrival time. Since the timing at which the hydraulic response time is subtracted is used as the output timing of the engagement hydraulic pressure, the output timing of the engagement hydraulic pressure is delayed at low vehicle speeds where the change range of the turbine rotation speed is small, as in the above-described conventional technology, There is a possibility that the engine will blow up and shift shocks may occur.

In Patent Document 2, the output start timing of the engagement hydraulic pressure, the backlash time of the engagement-side engagement element, the turbine rotation speed during the shift, the target rotation speed at the completion of the shift, and the current rotation speed during the shift are disclosed. A shift control device that calculates a change rate or a set target change rate is disclosed.
However, even in this transmission control device, the case where the output timing of the engagement hydraulic pressure is outside the change zone of the turbine rotation speed is not assumed, and the above-described problem cannot be solved.

In Patent Document 3, a downshift is performed by controlling the release pressure of the release side engagement element exclusively at high vehicle speed and high torque, and the release side engagement element and the engagement side engagement element at low vehicle speed or low torque. A technique is disclosed in which a downshift is performed by setting the torque sharing ratio to 1: 1 and simultaneously executing the hydraulic pressures of the engagement elements on both the release side and the engagement side in parallel.
However, the method of controlling the turbine speed by sharing the torque between the engagement elements on both the release side and the engagement side increases the influence factors and the control is complicated, which causes a problem in accuracy. .
JP 2005-337410 A JP-A-5-141513 JP-A-9-296862

The object of the present invention is to automatically suppress the delay in the output timing of the engagement hydraulic pressure and suppress the occurrence of engine revolution and the occurrence of a shift shock even in a power-on downshift at a low vehicle speed with a small change in input rotation speed. An object of the present invention is to provide a transmission control device for a transmission.

To achieve the above object, the present invention outputs an engagement hydraulic pressure to an engagement-side engagement element after a power-on downshift gear shift command, and the input rotation speed changes to a low-speed rotation speed at a predetermined time change rate. The hydraulic pressure of the disengagement-side engagement element is adjusted so that the speed approaches, and when the input rotation speed approaches the rotation speed of the low speed stage, the disengagement-side engagement element is released and the engagement-side engagement element is fastened to change the speed. In the shift control device of the automatic transmission to be completed, the arrival time to the low speed stage is estimated from the actual input rotation speed and the rate of change over time, and the hydraulic response time of the engagement side engagement element is estimated from the estimated arrival time. The first timing calculation means for calculating the output timing of the engagement hydraulic pressure of the engagement side engagement element by subtracting the value, the input rotation speed after the shift at the time of the shift command is predicted, and the predicted input rotation speed and the input rotation speed Time to reach the low speed stage from the target time change rate Predicted, a second timing calculating means for calculating an output timing of the engagement oil pressure of the engagement side engagement element by subtracting the hydraulic pressure response time of the engagement side engagement element from the expected arrival time, the actual input rotational When the timing calculated by the second timing calculation means arrives before the number starts to change toward the low-speed rotation speed, the engagement hydraulic pressure is output at this timing, and the actual input rotation speed If the timing calculated by the second timing calculation means does not arrive before the start of change toward the rotational speed of the low speed stage, the engagement hydraulic pressure is increased at the timing calculated by the first timing calculation means. Provided is a shift control device for an automatic transmission comprising an engagement hydraulic pressure output means for outputting .

In the present invention, the output timing of the engagement hydraulic pressure of the engagement side engagement element is calculated by two methods. First, the arrival time to the low speed stage is estimated from the actual input rotation speed and its rate of change over time, and the hydraulic response time of the engagement-side engagement element is subtracted from this estimated arrival time. The output timing of the engagement hydraulic pressure of the engagement element is calculated. This timing is called actual output timing. Secondly, the input rotation speed after the shift is predicted at the time of the shift command, the arrival time to the low speed stage is predicted from the predicted input rotation speed and the target time change rate of the input rotation speed, and the engagement is performed from the predicted arrival time. The output timing of the engagement hydraulic pressure of the engagement side engagement element is calculated by subtracting the hydraulic response time of the side engagement element. This is called estimated output timing. Here, it is determined whether the input rotational speed is before or after the change starts toward the rotational speed of the low speed stage (before or after the detection of out-of-synchronization). The engagement hydraulic pressure of the engagement side engagement element is output at the actual output timing if the engagement hydraulic pressure of the engagement side engagement element is output and out of synchronization is detected. That is, after detection of loss of synchronization, the actual output timing is always set as the output timing of the engagement hydraulic pressure regardless of the actual output timing and the estimated output timing.

In power-on downshifts at medium and high vehicle speeds, the range of change in the input rotational speed is large, so that both the actual output timing and the estimated output timing are often detected after loss of synchronization. In that case, as in the conventional case (see FIG. 7), the arrival time to the low speed stage is estimated from the actual input rotation speed and the rate of change over time, and the hydraulic response of the engagement side engaging element is estimated from this estimated arrival time. What is necessary is just to make the time (actual output timing) which deducted time into the output timing of the engagement hydraulic pressure of an engagement side engagement element. At medium and high vehicle speeds, no engagement delay occurs even if the engagement hydraulic pressure is output at the actual output timing as in the conventional case. On the other hand, in the power-on downshift at a low vehicle speed, since the change range of the input rotation speed is small, it cannot be obtained even if the actual output timing is calculated. Therefore, the input rotation speed after the shift is predicted at the time of the shift command, the arrival time to the low speed stage is predicted from the predicted input rotation speed and the target time change rate of the input rotation speed, and the engagement side engagement is determined from the predicted arrival time. The time point (estimated output timing) after subtracting the hydraulic response time of the combined element is set as the output timing of the engagement hydraulic pressure of the engagement side engaging element. Therefore, the engagement delay of the engagement side engagement element at the time of low vehicle speed can be eliminated, and shift shock and engine rotation can be prevented from being blown up.
In the present invention, the output timing of the engagement hydraulic pressure of the engagement side engagement element is not determined by the vehicle speed or the torque, but only one of the actual output timing and the estimated output timing is selected before and after the detection of loss of synchronization. Therefore, control is very simple and stable control can be performed.

The increasing speed (time change rate) of the input rotation speed during the power-on downshift varies depending on the vehicle load (input torque), the vehicle speed, and the like. For this reason, it is necessary to set in advance the input rotation speed after the shift predicted at the time of the shift command and the target time change rate of the input rotation speed in accordance with the load and the vehicle speed. In particular, if the target time change rate is learned and corrected using the time change rate of the actual input rotational speed, the target time change rate with little error due to individual variation of the vehicle or change with time can be obtained.

As described above, according to the present invention, the output timing of the engagement hydraulic pressure of the engagement side engagement element is calculated by two methods, and the estimated output timing is set as the output timing of the engagement hydraulic pressure before detection of loss of synchronization. After detection of loss of synchronization, the actual output timing is set as the output timing of the engagement hydraulic pressure. In other words, when the range of change in the input rotation speed is large, such as the power-on downshift at medium and high vehicle speeds, the output timing of the engagement hydraulic pressure is determined from the actual input rotation speed and the rate of change over time. When the change range of the input rotation speed is small, such as on-down shift, the input rotation speed after the shift is predicted at the time of the shift command, and the engagement hydraulic pressure is output from the predicted input rotation speed and the target time change rate of the input rotation speed. The timing is determined. Therefore, even in a power-on downshift at a low vehicle speed where the change range of the input rotational speed is small, it is possible to suppress a delay in the output timing of the engagement hydraulic pressure, and to prevent the engine rotation from blowing up and the occurrence of a shift shock.

Hereinafter, preferred embodiments of the present invention will be described with reference to examples.

FIG. 1 shows an example of a vehicle system equipped with an automatic transmission according to the present invention.
The output of the engine 1 is transmitted to the transmission mechanism 4 via the torque converter 3 of the automatic transmission 2, and the transmission mechanism 4 is connected to wheels (not shown) via the output shaft 5. The automatic transmission 2 includes an oil pump 6 that is driven by the engine 1 via the torque converter 3, and the discharge pressure of the oil pump 6 is sent to the hydraulic control device 7. The hydraulic control device 7 includes first to third solenoid valves 21 to 23 for speed change control. By controlling these solenoid valves 21 to 23 with the AT controller 20, various mechanisms built in the speed change mechanism 4 are provided. The hydraulic pressure of the combined element is controlled according to the traveling state. As the solenoid valves 21 to 23, not only linear solenoid valves but also duty solenoid valves can be used.
Note that the number of solenoid valves 21 to 23 for shift control is not limited to three, and solenoid valves having other functions such as a lockup clutch control and a line pressure control may be added.

Signals such as engine speed, throttle opening, turbine speed (input speed), vehicle speed, and shift position are input to the AT controller 20. The memory of the AT controller 20 stores a shift map, a power on / off determination value map, a shift control program, and the like. From the input signal, it is determined whether the power is on or off based on the determination map, and the gear position is determined based on the shift map based on the vehicle speed and the throttle opening at that time, and the solenoid valves 21 to 23 are controlled. Note that signals other than those described above may be input to the AT controller 20.

FIG. 2 shows an example of the speed change mechanism 4.
The speed change mechanism 4 includes an input shaft 10 to which engine power is transmitted via a torque converter 3, three clutches C1 to C3 and two brakes B1 and B2, which are friction engagement elements, a one-way clutch F, a Ravigneaux type planet, A gear mechanism 11 and a differential device 14 are provided.

The forward sun gear 11a and the input shaft 10 of the planetary gear mechanism 11 are connected via a C1 clutch, and the rear sun gear 11b and the input shaft 10 are connected via a C2 clutch. The carrier 11c is connected to the center shaft 15, and the center shaft 15 is connected to the input shaft 10 via a C3 clutch. The carrier 11c is connected to the transmission case 16 via a B2 brake and a one-way clutch F that allows only forward rotation (engine rotation direction) of the carrier 11c. The carrier 11c supports two types of pinion gears 11d and 11e, the forward sun gear 11a meshes with a long pinion 11d having a long axial length, and the rear sun gear 11b meshes with the long pinion 11d via a short pinion 11e having a short axial length. . A ring gear 11f that meshes only with the long pinion 11d is coupled to the output gear 12. The output gear 12 is connected to the differential device 14 via the intermediate shaft 13.

FIG. 3 shows the operation of the clutches C1, C2, C3, the brakes B1, B2 and the one-way clutch F constituting the transmission mechanism 4, and realizes four forward speeds and one reverse speed. The C1 clutch is engaged only in the R range, the C2 clutch is engaged in the 1st to 3rd gears in the D range, the C3 clutch is engaged in the 3rd and 4th gears in the D range, and the B1 brake is engaged in the 2nd and 4th gears in the D range. The B2 brake is engaged in the R range and the L range (not shown).

FIG. 3 also shows operating states of the first to third solenoid valves (SOL1 to SOL3) 21 to 23. This operation table represents an operation in a steady state, where ◯ represents an energized state and x represents a non-energized state. In this embodiment, the first solenoid valve 21 is for B1 brake control, the second solenoid valve 22 is for C2 clutch control, and the third solenoid valve 23 is for both C3 clutch control and B2 brake control. Yes.
In this embodiment, the first solenoid valve 21 is normally closed, and the second and third solenoid valves 22 and 23 are normally open. However, the present invention is not limited to this.

FIG. 4 shows a method for controlling the power-on downshift from the third speed to the second speed at medium and high vehicle speeds. At the time of downshift from the 3rd speed to the 2nd speed, the engaging side engaging element is the B1 brake, and the releasing side engaging element is the C3 clutch. FIG. 4 shows the turbine speed, the instruction current of the release side solenoid valve 23, and the instruction current of the engagement side solenoid valve 21, but the release side solenoid valve 23 is normally closed (normally open type and current). Values are shown as reversed).

In FIG. 4, the same reference numerals are assigned to the same times as in FIG. In this control method, after detecting the out-of-synchronization at time t3, the time t6 to reach the second-stage turbine speed is estimated from the turbine speed and the time change rate at time t3, and from that time t6 By subtracting the hydraulic response time Tr of the B1 brake, the output timing (referred to as actual output timing) t4 of the engagement hydraulic pressure of the B1 brake is calculated.

On the other hand, simultaneously with the shift command t1 of the power-on downshift, the turbine speed after the shift is predicted, and the second speed stage is determined from the predicted input speed and the target time change rate of the turbine speed (shown by a broken line in FIG. 4). The arrival time t7 is predicted. By subtracting the hydraulic response time Tr of the B1 brake from the predicted arrival time t7, an output timing (referred to as an estimated output timing) t8 of the engagement hydraulic pressure of the B1 brake is calculated.

As described above, the two output timings t4 and t8 are calculated. Since these output timings t4 and t8 are both after the out-of-synchronization detection t3, the actual output timing t4 is determined as the output timing of the engagement hydraulic pressure, and the engagement hydraulic pressure is output to the B1 brake at time t4.
In the case of the control method of FIG. 4, since the actual output timing t4 and the estimated output timing t8 are both after the out-of-synchronization detection t3, the estimated output timing t8 using the target time change rate (shown by a broken line in FIG. 4). Is not effectively used, resulting in the same result as the conventional control method shown in FIG.

FIG. 5 shows a control method of the power-on downshift from the third speed to the second speed at a low vehicle speed.
In this case, after detecting the out-of-synchronization at time t3, the time t6 at which the second-stage turbine speed is reached is estimated from the turbine speed and the time change rate at time t3. Even if the actual output timing t4 is calculated by subtracting the hydraulic response time Tr of the B1 brake, the calculated actual output timing t4 is already meaningless because it is already in the past.

On the other hand, simultaneously with the shift command t1 for the power-on downshift, the turbine rotational speed after the shift is predicted, and from this predicted rotational speed and the target time change rate of the turbine rotational speed (indicated by a broken line in FIG. 5) to the second speed stage. The estimated output timing t8 is calculated by subtracting the hydraulic response time Tr of the B1 brake from the predicted arrival time t7. The estimated output timing t8 is a time point before the loss of synchronization detection t3. At this timing t8, the engagement hydraulic pressure is output to the B1 brake. In this way, in the power-on downshift at a low vehicle speed where the variation range of the turbine rotation speed is small, the engagement hydraulic pressure can be output before the out-of-synchronization t3 is detected. It is possible to suppress the occurrence of spills and shift shocks.

FIG. 6 shows an example of the shift control flow in the power-on downshift from the third speed to the second speed. Here, the control of the engagement hydraulic pressure of the B1 brake will be described in particular.
First, it is determined whether or not a power-on shift command from the third speed to the second speed has been issued (step S1). When a shift command is issued, initial control is performed (step S2). Initial control refers to a series of controls such as backpacking, standby pressure output, and out-of-synchronization detection.
Immediately after or in parallel with the initial control, it is determined whether or not it is before out-of-synchronization detection (step S3). If it is determined that it is before out-of-synchronization detection, the processing from step S4 is executed. If it is determined after detection of out-of-synchronization is detected, step S4 is not entered and the processing from step S9 is executed. carry out. As described above, when it is determined that the synchronization has been detected, the actual output timing is set as the engagement hydraulic pressure output timing regardless of the estimated output timing and the actual output timing. This is because after the loss of synchronization is detected, the actual output timing reflects the actual control state rather than the estimated output timing, so that accurate control can be performed.
In step S4, the turbine speed of the second speed stage is predicted and calculated from the output speed (vehicle speed) at the time of the shift command, the rate of change of the target turbine speed is calculated (step S5), and the second speed stage is reached. Time (estimated time) is calculated (step S6). The predicted second gear arrival time is the time t7 in FIGS.
Next, the predicted time obtained by calculation is compared with the hydraulic response time Tr of the B1 brake (step S7). The hydraulic response time Tr is substantially equal to the time for which the engagement hydraulic pressure of the B1 brake rises to the hydraulic pressure at the completion of engagement after the command signal is output to the engagement side solenoid valve 21. If the predicted time is less than or equal to the hydraulic response time Tr, the B1 brake engagement hydraulic pressure is output (step S8). This means that the engagement hydraulic pressure is output at the estimated output timing t8 in FIG.
On the other hand, if the predicted time is longer than the hydraulic response time Tr, the actual rate of change of the turbine speed is calculated (step S9), and the time to reach the second gear (estimated time) is calculated (step S10). Here, the estimated arrival time of the second gear is the time t6 in FIGS.
Next, the estimated time obtained by calculation is compared with the hydraulic response time Tr of the B1 brake (step S11). When the estimated time is longer than the hydraulic pressure response time Tr, the steps after step S3 are repeated, and when the estimated time becomes equal to or shorter than the hydraulic pressure response time Tr, the B1 brake engagement hydraulic pressure is output (step S8). This means that the engagement hydraulic pressure is output at the actual output timing t4 in FIG.

Of the series of steps shown in FIG. 6, steps S4 to S7 correspond to the second timing calculation means of the present invention, and steps S9 to S11 correspond to the first timing calculation means of the present invention. As described above, the predicted time and the estimated time are calculated and compared with the hydraulic pressure response time Tr, and when either time becomes equal to the hydraulic pressure response time Tr, the engagement hydraulic pressure is output. When the estimated output timing comes before detection of out-of-synchronization, the engagement hydraulic pressure is output at the estimated output timing. Therefore, the engagement delay of the B1 brake can be eliminated even in the power-on downshift at the low speed. In addition, when the estimated output timing and actual output timing arrive after detection of out-of-synchronization, the engagement hydraulic pressure is output at the actual output timing, so the power-on downshift at medium and high speeds is changed to the actual turbine speed change. It can be controlled in principle.

In the above description, it has been described that the actual output timing (estimated time) is calculated from the turbine rotation speed at the time when the out-of-synchronization is detected and the time change rate thereof, but actually, the actual output timing ( Estimated time) can be calculated. For example, since the turbine rotation speed is almost constant before detection of out-of-synchronization, the estimated time is close to infinity, but in that case, the estimated time is always longer than the hydraulic response time, so there is no particular problem in processing. Because. In FIG. 6, steps S9 to S11 (first timing calculation means) are performed after steps S4 to S7 (second timing calculation means), but the actual output timing can be calculated immediately after the shift command as described above. For example, steps S9 to S11 may be performed before steps S4 to S7.

In the above embodiment, the hydraulic response time Tr has been described as being constant, but it may be changed depending on the ATF oil temperature or the like. That is, since the viscosity increases when the oil temperature is low, the hydraulic pressure response time Tr tends to be longer than when the oil temperature is high. Therefore, when the oil temperature is low, by setting the oil pressure response time Tr longer, a power-on downshift without a shock can be performed regardless of the oil temperature change.

4 and 5, an example is described in which the target time change rate of the turbine speed (shown by a broken line) is smaller than the time change rate of the actual turbine speed (shown by a solid line). It may be larger than the actual time change rate. It is desirable that the target time change rate be close to the actual turbine speed change rate by learning control or the like.

In the above embodiment, the shift control during the power-on downshift from the 3rd speed to the 2nd speed has been described as an example. However, if the power-on downshift is accompanied by the clutch change, for example, from the other speed stage, for example, the 4th speed The same applies to shifting to the third speed.
4 and 5 show examples in which the engagement hydraulic pressure of the engagement side engagement element B1 is increased stepwise and then held at a constant pressure, but may be increased with a predetermined time gradient.

It is a system diagram carrying the automatic transmission for vehicles in the present invention. FIG. 2 is a skeleton diagram of a transmission mechanism of the automatic transmission of FIG. 1. 3 is an operation table of each friction engagement element and solenoid valve of the speed change mechanism shown in FIG. 2. It is a time change figure of the turbine speed of the power-on downshift from the 3rd speed to the 2nd speed at the time of medium and high vehicles of the present invention, the indication current of the release side solenoid valve, and the indication current of the engagement side solenoid valve. It is a time change figure of the turbine speed of the power-on downshift from the 3rd speed to the 2nd speed at the time of low vehicle speed of the present invention, the indication current of the release side solenoid valve, and the indication current of the engagement side solenoid valve. It is a flowchart figure of an example of the speed-change control method which concerns on this invention. It is a time change figure of the turbine speed of the power-on downshift from the 3rd speed to the 2nd speed at the time of the conventional middle and high vehicle speed, the indication current of the release side solenoid valve, and the indication current of the engagement side solenoid valve. It is a time change figure of the turbine speed of the power-on downshift from the 3rd speed to the 2nd speed at the time of the conventional low vehicle speed, the indication current of the release side solenoid valve, and the indication current of the engagement side solenoid valve.

Explanation of symbols

B1 engagement side engagement element C3 release side engagement element 20 AT controller 21 engagement side solenoid valve 23 release side solenoid valve

Claims (1)

After the power-on downshift command, the engagement hydraulic pressure is output to the engagement-side engagement element, and the input-side rotation speed approaches the low-speed rotation speed at a predetermined time change rate. In the shift control device for an automatic transmission that releases the disengagement-side engagement element and fastens the engagement-side engagement element when the input rotation speed approaches the rotation speed of the low speed stage and completes the shift. ,
Estimate the arrival time to the low speed stage from the actual input rotation speed and the rate of change over time, and subtract the hydraulic response time of the engagement side engagement element from the estimated arrival time to engage the engagement side engagement element. First timing calculating means for calculating the output timing of the combined hydraulic pressure;
The input rotation speed after the shift is predicted at the time of the shift command, the arrival time to the low speed stage is predicted from the predicted input rotation speed and the target time change rate of the input rotation speed, and the engagement side engagement element is calculated from the predicted arrival time. Second timing calculation means for calculating the output timing of the engagement hydraulic pressure of the engagement side engagement element by subtracting the hydraulic response time of
When the timing calculated by the second timing calculation means arrives before the actual input rotation speed starts to change toward the low speed rotation speed, the engagement hydraulic pressure is output at this timing, If the timing calculated by the second timing calculating means does not arrive before the input rotational speed of the engine starts changing toward the low speed, the timing calculated by the first timing calculating means A shift control device for an automatic transmission , comprising: an engagement hydraulic pressure output means for outputting an engagement hydraulic pressure .

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