JPS63263248A - Shift control device for automatic transmission - Google Patents

Abstract

PURPOSE:To obtain a compact and reliable engine torque detecting method by regarding an addition value of a product value obtained by multiplying a change rate of an engine speed by a designated value and a detected transmission torque of a driving force transmission as an engine torque detection value. CONSTITUTION:A transmission control unit TCU 16 computes and stores an engine speed and its change rate according to a detection signal of Ne sensor 14 or the like. Further the TCU 16 computes and stores transmission torque of a damper clutch 28 from various information on supply oil pressure, friction area and friction coefficient of the damper clutch 28 and piston pressure receiving area. Accordingly, in a driving force transmission device of a slip direct- coupled clutch, the transmission torque can be controlled from the outside by controlling the quantity of electricity of a solenoid valve 54 for regulating the supply pressure. Subsequently, the inertia moment (a fixed value for each engine) of an engine 10 is read out from a storage device, and the instantaneous value of engine torque can be computed by computing an engine speed change rate and transmission torque of the damper clutch 28.

Description

DETAILED DESCRIPTION OF THE INVENTION (Field of Industrial Application) The present invention relates to an engine torque detection method applied to hydraulic control of a speed change clutch of an automatic transmission for a vehicle.

(Prior art and its problems) The valve opening degree of the throttle valve and the vehicle speed are detected to detect the hydraulic pressure supplied to the speed change clutch during gear shifting of an electronically controlled automatic transmission system, and a predetermined amount of electricity is actuated based on these. There are known solenoid valves that are added to hydraulic control solenoid valves for adjustment. In such a conventional automatic transmission, the throttle valve opening degree and the detected value of the vehicle speed are not parameters that necessarily accurately represent the transmission torque input to the transmission, so it is necessary to ensure smooth and quick gear changes without shift sluggishness. I can't.

Furthermore, there is a known system that detects the rate of change in the rotational speed of the input shaft of the transmission during gear shifting, and feedback-controls the supply pressure to the engaging clutch or the disengaging clutch so that the change rate matches the target rate of change. . However, with this type of feedbank control, if the valve opening of the throttle valve changes suddenly during gear shifting, if the followability is poor, the rate of rotation change of the input shaft will be hunting, and the output torque will also be hunting accordingly. If you do so, you will not be able to shift gears smoothly.
Furthermore, if the supply pressure (initial value) to the clutch at the start of gear shifting is not appropriate, hunting is likely to occur in this case as well.

In order to eliminate the above-mentioned disadvantages, it is required to detect the instantaneous value of the input shaft torque of the transmission and use this for hydraulic control of the transmission clutch.

Conventionally, methods of detecting the shaft torque of a power transmission shaft using strain gauges or magnetostriction have been known, but these sensors are large and are susceptible to thermal effects on the detected values. Since the torque is large, a slip ring is required to detect the shaft torque of the rotating body, and there are problems with reliability and quality of the slip ring.

One possible method is to create a map of the torque generated by the engine according to the throttle valve opening and engine speed, and then calculate the torque value according to these detected values, but this method is difficult to deal with deterioration in engine performance. Difficult, engine temperature (
There is also the problem that it cannot respond to changes in engine water temperature. Further, in engines equipped with a supercharger such as a turbocharger, there is a problem in that due to a time lag during sudden acceleration, the generated torque cannot be accurately detected only by the valve opening of the throttle valve and the engine rotation speed.

Furthermore, it is possible to create a map of the generated torque according to the fuel injection amount and intake air amount and calculate the torque value according to these detected values, but if the friction loss of the crankshaft etc. changes, there will be an error in the calculated torque value. There is also a problem that the error is large depending on the engine temperature.

The present invention was made in order to solve such problems, and an object of the present invention is to provide an engine torque detection method that can accurately and reliably detect the transmitted torque of a power transmission system without using a large-scale detection device. do.

(Means for Solving the Problems) In order to achieve the above-mentioned object, the present invention provides a drive system in which the driving force of an engine is transmitted to wheels via a driving force transmission device whose transmitted torque can be detected. In the engine torque detection method, a rate of change in the engine speed is detected, and a product value obtained by multiplying the detected transmission torque of the driving force transmission device and the detected engine speed change rate by a predetermined value is provided. An engine torque detection method is provided, which is characterized in that the added value is added and the added value is detected as a resin torque.

(Function) The net engine torque obtained by subtracting the engine's friction loss from the average torque due to explosion of the internal combustion engine is:
The transmission torque of a driving force transmission device such as a torque converter,
It can be calculated as an added value of the product value obtained by multiplying the engine speed change rate by a predetermined value such as crankshaft rotational inertia, and is the driving force for fluid couplings such as torque converters, slip-controlled electromagnetic powder cranks, viscous cranks, etc. The transmission torque of a transmission device can be detected almost quantitatively from the rotational speed of the input shaft and output shaft. By controlling N (control parameter value), the transmitted torque can be controlled from the outside, and by detecting this electrical quantity, the transmitted torque can be detected approximately quantitatively. Therefore, by detecting the rate of change in engine speed and the transmission torque of the driving force transmission device, the instantaneous value of the engine torque can be calculated.

(Example) Hereinafter, an example of the present invention will be described in detail based on the drawings.

FIG. 1 shows a schematic configuration of an electronically controlled automatic transmission equipped with a torque converter for a vehicle implementing the method of the present invention, and an internal combustion engine 10 is, for example, a 6-cylinder engine,
A flywheel 11 is attached to the crankshaft 10a, and one end of a drive shaft 21 of a torque converter 20 serving as a driving force transmission device is directly connected to the crankshaft 10a via the flywheel 11. The torque converter 2o includes a casing 20a, a pump 23, a stator 24,
and a turbine 25, the pump 23 is connected to the drive shaft 2 via the input casing 22 of the torque converter 20.
1, and the stator 24 is connected to the casing 20a via a one-way clutch 24a.

Further, the turbine 25 is connected to an input shaft 30a of a gear transmission 30.

The torque converter 20 of this embodiment includes a slip type direct coupling clutch, for example, a damper clutch 28, which is interposed between the input casing 22 and the turbine 25, and even when engaged (directly coupled). Pump 23 of torque converter 20 while allowing appropriate slip.
and the turbine 25 are directly connected mechanically, and the slip amount of the damper clutch 28, that is, the damper clutch 28
The torque transmitted through the damper clutch hydraulic pressure control circuit 50 is externally controlled. The damper clutch hydraulic control circuit 50 has a damper clutch control pulse 7''.
52 and a damper clutch control solenoid pulp 54, the solenoid pulp 54 is a normally closed on-off valve, and its solenoid 54a is electrically connected to a transmission control unit (hereinafter referred to as rTcUJ) 16. The damper clutch control pulp 52 switches the oil path of the hydraulic oil supplied to the damper clutch 2B, and also switches the oil path of the hydraulic oil supplied to the damper clutch 2B.
8. Controls the hydraulic pressure applied to 8. That is, the damper clutch control pulp 52 is housed in a spool 52a and a left end chamber 52b facing the left end surface of the spool 52a in the drawing.
The left end chamber 52b is connected to a pilot oil passage 55 that communicates with a biloft hydraulic pressure source (not shown).

The pilot oil passage 55 has a branch passage 55 that passes to the drain side.
a is connected, and the solenoid pulp 54 is disposed in the middle of this branch path 55a, and the magnitude of the pilot oil pressure supplied to the left end chamber 52b is controlled by opening and closing the solenoid pulp 54. The pilot hydraulic pressure from the biloft hydraulic pressure source is also supplied to the right end chamber 52d, which the right end surface of the spool 52a faces.

When the pilot hydraulic pressure acts on the left end chamber 52b and the spool 52a of the damper clutch control pulp 52 moves to the right column limit position shown in the figure, the torque converter (T/C) 1 hydraulic pressure supplied to the torque converter 20 is applied to the oil path 56 and the control The pulp 52 is supplied to the hydraulic chamber formed between the input casing 22 and the damper clutch 28 via the oil passage 57, and the engagement of the damper clutch 28 is released. On the other hand, when the pilot hydraulic pressure is not supplied to the left end chamber 52b and the spool 52a moves to the extreme left position shown in the figure, line pressure from a hydraulic pump (not shown) is applied to the damper clutch via the oil path 58, the control pulp 52, and the oil path 59. The damper clutch 28 is supplied to a hydraulic chamber formed between the turbine 28 and the turbine 25, and the damper clutch 28 is frictionally engaged with the human-powered casing 22.

Damper clutch solenoid pulp 54 by TCU16
When the duty rate Dc is controlled, the spool 52a moves to a position where the resultant force of the pilot oil pressure acting on the left end chamber 52b and the spring force of the spring 52c is balanced with the pilot oil pressure acting on the right end chamber 52d, and corresponds to this movement position. Hydraulic pressure is supplied to the damper clutch 28, and the damper clutch 2
The transmission torque Tc at 8 is controlled to a required value.

The gear transmission 30 has, for example, a gear train with four forward speeds and one reverse speed. FIG. 2 is a partial configuration diagram of the gear transmission 30, in which a first drive gear 31 and a second drive gear 32 are rotatably loosely fitted to the input shaft 30a. Input shaft 3 between second drive gear 32
0a is a hydraulic clutch 33 as a friction engagement element for speed change.
and 34 are fixedly installed, and the drive gears 31 and 32 engage the clutches 33 and 34, respectively, thereby driving the input shaft 30.
Rotates together with a. An intermediate transmission shaft 35 is disposed parallel to the input shaft 30a, and this intermediate transmission shaft 35 is connected to a drive axle via a final reduction gear (not shown).

The intermediate transmission shaft 35 has a first drive gear 31 that meshes with the first drive gear 31 .
A driven gear 36 and a second driven gear 37 that mesh with the second driving gear 32 are fixedly provided.
When the first driving gear 31 is engaged, the rotation of the input shaft 30a is transmitted to the clutch 33, the first driving gear 31, the first driven gear 36, and the intermediate transmission shaft 35, and the rotation of the input shaft 30a is transmitted to the clutch 33, the first driving gear 31, the first driven gear , and the intermediate transmission shaft 35.
For example, the first speed) is achieved. When the clutch 33 is disengaged and the clutch 34 and the second driving gear 32 are engaged, the rotation of the input shaft 30a is caused by the clutch 34, the second driving gear 32, the second driven gear 37, and the intermediate transmission shaft. 35, and the second gear stage (for example, second gear) is achieved.

FIG. 3 shows a hydraulic circuit 40 that supplies hydraulic pressure to the hydraulic clutches 33 and 34 shown in FIG.
4, a second hydraulic control valve 46, a solenoid valve 47, and a solenoid valve 48. Spools 45 and 49 are slidably inserted into the bores 44a and 46a of the first and second hydraulic control valves 44 and 46, respectively.
．． Right end chambers 44g and 46g are formed, respectively, to which the right end surfaces of 49 face. Each right end chamber 44g, 46g has a spring 44b.
, 46b are accommodated, and the springs 44b and 46b are connected to the spool 4.
5.49 is pressed to the right side of the figure. The first and second hydraulic control valves 44.46 each have a spool 44.46.
Left end chambers 44h and 46h facing each left end surface are formed, respectively. These left end chambers 44h and 46h are the orifice 4.
It communicates with the drain side via 4t and 46i.

The solenoid valve 47 is a normally open three-way switching valve and has three ports 47c, 47d, and 47e.

The solenoid valve 47 includes a valve body 47a and a valve body 47a.
A spring 47b presses the port 47a toward the boat 47a to close the boat 47e, and when biased, moves the valve body 47a toward the boat 47c against the spring force of the spring 47b to close the port 47c.
It is composed of a solenoid 47f that closes the. on the other hand,
The solenoid valve 48 is a normally closed three-way switching valve, and has three boats 48C.

48d and 48e. The solenoid valve 48 includes a valve body 48a, a spring 48b that presses the valve body 48a toward the port 48c to close 480, and a spring 48b that presses the valve body 48a toward the port 48c to close the valve 480.
It is composed of a solenoid 481 that moves the valve body 48a toward the port 48e against the spring force b to close the port 48e. The solenoids 4H and 48f of the solenoid valves 47 and 48 are respectively connected to the output side of the TCU 16.

Oil passage 41 extending from the hydraulic pump (not shown) Each boat 44 of the first hydraulic control valve 44 and the second hydraulic control valve 46
One end of the oil passage 41a is connected to the boat 44d of the first hydraulic control valve 44;
A hydraulic clutch 33 is connected to the other end of 1a. Second
One end of an oil passage 41b is connected to the boat 46d of the hydraulic control valve 46, and the other end of the oil passage 41b is connected to a hydraulic clutch 34.A pilot oil passage 42 extending from the pilot oil pressure source (not shown) is connected to a first oil passage 41b. and second hydraulic control valve 4
4.Boat 4 connected to each left end chamber 44h, 46h of 46
4e and 46e, and also connected to respective ports 47c and 48c of solenoid valves 47 and 48. Each boat 47d, 48d of the solenoid valves 47 and 48 is a pilot oil passage 42a.

Boats 44f and 46 communicate with the respective right end chambers 44g and 46g of the first and second hydraulic control valves 44 and 46 via 42b.
f, respectively. Each port 47e, 48e of the solenoid valves 47, 48 communicates with the drain side.

The oil passage 41 transfers the working oil pressure (line pressure) to a predetermined pressure by a pressure regulating valve (not shown) to the first and second oil pressure control valves 44.
．． 46, and the pilot oil passage 42 supplies pilot oil pressure regulated to a predetermined pressure by a pressure regulating valve (not shown) or the like to first and second oil pressure control valves 44, 46 and solenoid valves 47.
48.

When the spool 45 of the first hydraulic control valve 44 moves to the left, the land 45a of the spool 45 that was blocking the port 44c opens the boat 44c, and the working oil pressure is transferred to the oil path 41 and the port 44.
c, the boat 44d2 is supplied to the clutch 33 through the oil passage 41a, and when the spool 45 moves to the right, the land 45a closes the port 44c, while the port 44d communicates with the drain port 44j, so that the oil pressure of the clutch 33 is on the drain side. be excluded. When the spool 49 of the second hydraulic control valve 46 moves to the left, the land 49a of the spool 49 that had been blocking the port 46C opens the port 46c, and the working oil pressure flows through the oil passage 41, bow) 46c, boat 46d, and oil passage 41b. When the spool 49 moves to the right, the land 49a closes the port 46c, while the boat 46d communicates with the drain port 46j, and the hydraulic pressure of the clutch 34 is removed to the drain side.

A ring gear Ila that meshes with the pinion 12a of the starter 12 is fitted on the outer periphery of the flywheel 11, and this ring gear Ila has a predetermined number of teeth (for example, 11
0), and an electromagnetic pickup 14 is attached opposite the ring gear lla.

Electromagnetic pickup (hereinafter referred to as rNe sensor)
Reference numeral 14 detects the engine rotation speed Ne of the engine IO, as will be described in detail later, and is electrically connected to the input side of the TCU 16.

On the input side of the TCU 16, there are a turbine rotation speed sensor (NL sensor) 15 that detects the rotation speed Nt of the turbine 25 of the torque converter 20, and a transfer drive gear rotation speed sensor (No. sensor) 17, a throttle valve opening sensor (θL sensor) that detects the valve opening θL of a throttle valve disposed in the middle of an intake passage (not shown) of the engine 10;
18. An oil temperature sensor 19 for detecting the temperature of hydraulic oil discharged from a hydraulic pump (not shown) is connected, and detection signals from each sensor are supplied to the TCU 16.

Hereinafter, the operation of the gear transmission configured as described above will be explained.

The TCU 16' has a built-in storage device such as a ROM or RAM (not shown), a central processing unit, an I10 interface, a counter, etc., and the TCU 16 performs shift hydraulic control as follows according to a program stored in the storage device.

The TCU 16 repeatedly executes the main program routine shown in FIG. 4 at a predetermined frequency, for example, at a frequency of 35 Hz. In this main program routine, first, in step SIO, reading and setting of various initial values, which will be described later, are executed.
Detection signals from the sensor 14, Nt sensor 15, No sensor 17, θL sensor 18, oil temperature sensor 19, etc. are read and stored (step 5ll), and the TCU 16 calculates parameter values necessary for speed change control from these detection signals. Calculate and store as follows.

First, the TCU 16 calculates the engine rotational speed Ne and the rate of change ωe of the engine rotational speed Ne from the detection signal of the Ne sensor 14 (step 512). Each time four teeth are detected, one pulse signal is generated and supplied to the TCU 16. The TCU 16 has one duty cycle, 28.6ss, as shown in FIG.
The time L p (see) required to detect the last nine pulses of the pulse signals from the Ne sensor 14 supplied during ec (3511z) is measured, and the engine rotation is calculated from the following equation (1). A number N e (rps) is calculated and stored in the storage device as the engine rotation speed (Ne)n of the current duty cycle.

N e -(9x4)+110 +tpx6゜=
216+(1lxtp) +++++・(
1) Then, from the engine rotation speed (Ne) −+ stored in the previous duty cycle and the engine rotation speed (Me) stored in the current duty cycle, the engine rotation speed change rate ωe (rad/see”) is calculated. is calculated and stored using the following formula 〇).

ate−”ΔNeX 2 z +60+T= (g/3
0T) XΔNe ・・・・・・(2) −
Here, ΔNe-(Ne)n-(Ne)at, ? '
(T, +↑2)/2, and T+ and Tt are the time between the end of the tp waiting time count and the time between the start of the count (see ).

The TCU 16 then proceeds to step 313 and calculates the engine net torque Te and the torque converter output shaft torque (
Hereinafter, this will be referred to as "turbine shaft torque") TL (kg
・-) is calculated.

Here, the relationship between the friction torque Tb of the clutch on the releasing side or the engaging side during shifting, the turbine shaft torque Tt, and the turbine rotation rate of change ωL during shifting is expressed by the following equation (A1).

Tb=a-TL+b・ωt ・−・−(A1) Here, a and b are shift patterns (types of speed change) such as upshifting from 1st to 2nd speed, downshifting from 4th to 3rd speed, etc. This is a constant determined by the moment of inertia of each rotating part. As can be seen from the above formula (AI), the friction torque Tb of the clutch, that is, the working oil pressure of the clutch 33, 34 is expressed as the turbine shaft torque Tt and the rate of change in turbine rotation during gear shifting ωt.
If determined based on this, it can be set without being affected by deterioration of engine performance or engine water temperature, etc., and the experimental formulas and data obtained based on this idea can be easily applied to different types of engines. In addition, when it is desired to feedback control the turbine rotation rate of change ωt to a target value regardless of changes in the turbine shaft torque T, instead of correcting the deviation of the turbine rotation rate of change ωt from the target value, By increasing or decreasing the friction torque Tb, that is, the working oil pressure of the clutches 33 and 34 by the amount of change in the torque Tt, it is possible to perform stable shift control with good followability without having to set a large correction gain for feedback control. become. Further, the turbine shaft torque Tt at the time when the friction torque of the engaging side clutch starts to be generated at the start of the shift is set to an appropriate value, and the friction torque Tb is set such that the target turbine rotational change rate ωt is obtained from the above equation (A1). If the oil pressure supplied to the clutch is set so that the friction torque of the engaging side clutch starts to be generated, a turbine rotational change rate ωL close to the target value can be obtained, and the shift feeling can be improved.

Therefore, the turbine shaft torque Tt is calculated by the following equation (4) using the engine net torque Te calculated by the following equation (3), and these calculated values are stored in the storage device.

Te=C−Ne” −+lt ・a+e+Tc −
・−・・・ (3) Tt=t (Te-Tc) 10Tc "t(C-Ne" + Summer, -(1) e
)+T c -...= (4) Here, Te is the net torque obtained by subtracting Frixilan loss, oil pump drive torque, etc. from the average torque due to the explosion of the engine 10, and C is the torque capacity coefficient, which is stored in the storage device. The speed ratio e between the turbine rotation speed Nt and the engine rotation speed Ne is determined from the torque converter characteristics stored in advance.
(-Nt/Me). Therefore, the speed ratio e is calculated by first calculating the speed ratio e from the turbine rotation speed Nt detected by the Nt sensor 15 and the engine rotation speed Ne detected by the Ne sensor 14 as described above. The torque capacity coefficient C is read from the storage device in accordance with e. Section 1 is the moment of inertia of the engine IO, which is a constant value set for each engine, and t is the torque ratio, which is also determined from the torque converter characteristic table stored in advance in the storage device, the turbine rotation speed Nt and the engine rotation speed. It is read out according to the speed ratio e (-Nt/Ne) with respect to Ne.

Tc is the transmission torque of the damper clutch 28, and in this type of slip type direct coupling clutch, the torque Tc is expressed by the following formula (5)
is given by

Tc-Pc -A1 r ・B -a 1 - Dc-b 1 -- (5
) Here, Pc is the hydraulic pressure supplied to the damper clutch 28, A is the piston pressure receiving area of the damper clutch 28, r is the friction radius of the damper clutch 28, and μ is the damper clutch 28.
is the coefficient of friction. Since the hydraulic pressure Pc supplied to the damper clutch 28 is proportional to the duty ratio Dc of the damper clutch solenoid valve 54, the above equation (5) is obtained. Note that al and bl are constants set according to the shift mode, and are valid only when the Tc value calculated by the above formula (5) is positive, and when it is negative, Tc = 0. Set.

The engine net torque Te and turbine shaft torque Tt calculated and stored in this way are the engine rotation speed Ne detected by the Ne sensor 14, the turbine rotation speed Nt detected by the Nt sensor 15, and the damper clutch solenoid valve 54.
Each instantaneous value can be calculated and determined approximately quantitatively by the duty rate Oc. Moreover, the above-mentioned arithmetic expression (3) and (
4), the engine output torque Te is rt
- Since the calculation includes the 06 term, it is hardly affected by the turbine rotation rate of change ωt or the friction torque Tb. Therefore, when the friction torque Tb, that is, the clutch supply pressure is adjusted in order to set the turbine rotation rate of change ωt to the target value, the turbine shaft torque Tt changes, which is a situation where they interfere with each other and become uncontrollable. may occur (not,
In particular, when the turbine shaft torque increases or decreases due to disturbances such as accelerator work during gear shifting, and the friction torque Tb is adjusted to compensate for this, the above-mentioned interference does not occur, so responsive gear shifting control is possible. It is advantageous to obtain.

Next, in step S14, the TCU 16 sets the valve opening θt of the throttle valve and the transfer drive gear rotation speed N.
o, the gear stage to be established in the gear transmission 1F30 is determined. Figure 6 shows the first gear (hereinafter referred to as "first gear") and the second gear which is one higher gear (hereinafter referred to as "second gear"). The solid line in the figure is the boundary line that separates the 1st gear area and the 2nd gear fiI area when shifting up from 1st gear to 2nd gear, and the broken line in the figure is the shift range from 2nd gear to 2nd gear. This is a boundary line that separates the first speed region and the second speed region when downshifting to the first speed. The TCU 16 determines the gear stage to be established from FIG. 6, and stores this in the storage device.

Power on/off' 1 Next, the TCUI 6 proceeds to step S15 and executes a power on/off determination routine. FIG. 7 shows a flowchart of the power on/off determination routine. First, in step 5151, a determination value Tto is set. This discrimination value Tto is calculated by the following equation (6).

Ttoxa2・ωto=2π・a2・Ni −・・
(6) Here, a2 and Ni are predetermined values that are set in advance according to the shift pattern, and are set to a negative value in the case of an upshift, and a positive value in the case of a downshift. 0 Next, the TCU 16 determines whether the turbine shaft torque Tt calculated in step S13 is larger than llTto (step 5152), and if the determination result is affirmative (Yes), performs a power-on shift. If the determination is negative (No), it is determined that the power-off shift is to be performed (step 5154).
, the TCU 16 stores the power on/off determination result in the storage device and returns to the main routine shown in FIG.

The above-mentioned power on/off determination method is based on the following idea. That is, -11 is the clutch friction torque Tb
In the above equation (^1) that gives the relationship between the turbine shaft torque Tt and the turbine rotation rate of change ωt during gear shifting, if the turbine shaft torque Tt is set to 0 and the turbine rotation rate of change ωL is set to the target value ωto, the above equation is satisfied. (6) is obtained and the above target value ωt is in a state where elements other than the clutch are not operating.

Power on/off is determined based on whether or not enough turbine shaft torque Tt is generated to obtain the desired value.

This eliminates the following disadvantages of the conventional method, compared to the conventional method in which the power on/off determination was simply determined based on the positive/negative of the engine output.

In other words, in a system that performs shift control using different logic for the power-on state and power-off state, (1) in the case of an upshift, if the engine output takes a slightly negative value, it is determined to be the power-off state; The side friction element (clutch) remains released and the gear shift is not completed. (2) On the other hand, in the case of a downshift, if the engine output takes a slightly positive value, it is determined that the power is on.
You will have to wait for the rotation of the input shaft of the transmission to increase automatically, and the coupling side friction element (clutch)
This eliminates the inconvenience that the gear shift is not completed because the gears are not connected.

In addition, it is necessary to perform power on/off determination as quickly as possible during lift foot upshifts or downshifts while depressing the accelerator pedal, but in the power on/off determination described above, ・Turbine shaft torque Tt is calculated using the above formula (4).
Since we are using the so-called virtual turbine shaft torque, which is obtained by multiplying the engine net torque by the torque ratio t, from equation (4), 1.・Fruit obtained by omitting the term ωe - bottle shaft torque Tt'(=t-CNe" 10T
The power on/off determination can be made more quickly than in the case of using c). That is, for example, during a lift foot upshift, if a decrease in engine output is sensed as early as possible and the releasing element (clutch) is quickly released, deceleration shock in low gears can be avoided. To explain this with reference to Fig. 26, when the accelerator pedal is released and shifts to upshift (Fig. 26 (a)), the actual turbine shaft torque Tt' changes to the broken line shown in Fig. 26 (b). On the other hand, the virtual turbine shaft torque Tt changes along the solid line shown in FIG. 26(b). Therefore, when using the virtual turbine shaft torque Tt, L shown in FIG. 26(b)
When the actual turbine shaft torque Tt'' is used at one point in time, the power-off state can be detected at each of the 12 points shown in FIG. 26 (b).As a result, when the virtual turbine shaft torque Tt is used In this case, the power-off determination can be made faster by Δ1 (=t2-tl) than when the actual turbine shaft torque T1 is used, and the release side element can be released that much faster, which reduces the drop in the output shaft torque ( (see the shaded area in FIG. 26(C)), and deceleration shock can be avoided.

Returning to FIG. 4, next, the TCU 16 determines whether the shift range to be established determined in step 314 has changed from the result determined in the previous duty cycle. If there is no change, the process returns to step 11, and steps 311 and subsequent steps are repeated again. On the other hand, if the shift pattern has changed, a shift signal corresponding to the shift pattern determined in steps S14 and S15 is output (step 517), and the process returns to step Sll.

Power-on upshift door, 1' Figures 8 to 12 are flowcharts showing the shift hydraulic pressure control procedure in the case of power-on upshift. The control procedure will be explained with reference to FIG. 13 as an example.

The TCU 16 first calculates the initial duty ratios DL11 and I)ut of the solenoid valves 47 and 48 using the following equations (8) and (9) based on the shift signal of the power-on upshift from the first speed to the second speed. (Step 320).

D11+=a4 HI T t l +c4
−=(8) Du*=a5 ・l T t I +c5
-(9) Here, TL is the turbine shaft torque value calculated and stored in step S13 of FIG. 4 for each duty cycle, a4+c4 and a5.c5.
is a constant applied when shifting up from first speed to second speed.

Next, the TCU 16 sets the duty rate DL11 of the normally open solenoid valve 47 to the initial duty rate DI11 set in step S20, outputs a signal to open and close the solenoid valve 47 at the duty rate, and releases the solenoid valve 47. The supply of initial oil pressure corresponding to the initial duty rate IL+ is started to the first speed clutch 33, which is a side frictional engagement element, and the piston (not shown) of the first speed clutch 33 is moved toward the position immediately before clutch slippage occurs. retreat (step S21)
, time t1 in FIG. 13(b)), on the other hand, the duty rate [)t4 of the normally closed solenoid valve 48 is set to 100χ,
At the duty rate Dt4, a signal is output to open and close the solenoid valve 48, and the piston of the second speed clutch 34, which is a friction engagement element on the coupling side, is moved to a position immediately before clutch engagement is started (piston backlash reduction position). Proceed (Figure 13 (
C), the initial pressure supply time Ts is set on the timer.
+ is set (step 522).

This timer may be a hard timer built in the TCU 16, or the above initial pressure supply time T can be set by executing a program.
The initial pressure supply time T□, which may be a so-called soft timer that measures □, is determined by the initial pressure supply time T□. This is a predetermined value that allows the piston No. 34 to be advanced to a predetermined position immediately before the start of engagement.

The TCU 16 waits for the elapse of a predetermined time t, that is, one data cycle (28.6 ssec in this embodiment) (step 523), and when the predetermined time LD elapses, the duty cycle set in the previous duty cycle is resumed. rate DL
A predetermined duty rate ΔD1 is added to 11 to obtain a new duty rate DLl, and a signal for opening and closing the solenoid valve 47 is outputted at this duty rate DL11 (step 324). The duty rate DL11 is set to a value that increases at a predetermined rate (for example, 4M, which increases at a rate of 4% per second). (see change), and TC
Step 5 U16 determines whether the initial pressure supply time Tit set in step S22 has elapsed.
25) If the time has not elapsed, the process returns to step S23° and steps S23 to S25 are repeatedly executed.

If the determination result in step S25 is affirmative, that is, if the initial pressure supply time T□ has elapsed and the second speed clutch 34 has advanced to the predetermined position immediately before engagement, the TCU 16 proceeds to step S27 in FIG. The duty rate Dt4 of the solenoid valve 48 is set to a -H predetermined value DtamIn, and a drive signal for opening and closing the solenoid valve 48 at this duty rate Dt4 is output (time point L2 in FIG. 13(C)). Predetermined value D
tallin is the duty rate at which the hydraulic pressure supplied to the second speed clutch 34 via the second hydraulic control valve 46 provides a holding pressure that neither increases nor decreases. Then, for a predetermined time 1. , that is, one duty cycle (step 32B), and when a predetermined period of time has elapsed, a predetermined duty rate ΔD1 is added to the duty rate DLII of the solenoid valve 47 set in the previous duty cycle. and set a new duty rate DLI, and set the duty rate Dtn of the solenoid valve 48 to a predetermined duty rate ΔD.
2 is added to obtain a new duty rate DR4, and each solenoid valve 47.2 is added with these duty rates DLR and Dt4.
Output a signal to open and close 48 (step 530)
, the predetermined duty rate ΔD2 to be added is the solenoid valve 48
Duty rate D! 4 is set to a value that increases at a predetermined rate (for example, a value that increases at a rate of 15% per second).
(See the change in duty rate DI from time point L2 to time point 13 in FIG. 13(C)).

Next, the process proceeds to step S32, and the TCU 16 calculates the actual slip rotation speed Nl using the following equation 0ω, and compares this with a predetermined determination value ΔN31 (for example, 10 rps+).

N sm = N t N tel −
-.,,.6 (Smile) Ntcl is the calculated turbine rotation speed at 1st speed, and is obtained as the product value obtained by multiplying the transfer drive gear rotation speed NO detected by the No sensor 17 by a predetermined number.

Compare the actual slip rotation speed N g m with the predetermined judgment value ΔN,□, and the actual slip rotation speed N*11 becomes the predetermined judgment value ΔN3□
When it is smaller (Ns*<ΔN, □), the TCU 16 returns to step 32B and performs steps S2B to S3.
Repeat step 2. As a result, the first speed clutch 33 on the disengagement side is gradually disengaged and released, while the second speed clutch 34 on the engagement side is held at six points t just before the engagement starts.
(The turbine rotation speed Nt is gradually moved from f2 to the engagement side, but engagement has not yet started. In such a state, the turbine rotation speed Nt is
As the first speed clutch 33 on the release side is released, the rotation speed is gradually increased (FIG. 13 (second half of the active control section A), that is, the control section A (shift signal output time tl
In the control period from 1 to t3 when it is detected that the actual slip rotational speed N□ has exceeded the predetermined determination value ΔN31), the first speed clutch 33 is engaged before the friction torque of the second speed clutch 34 is generated. By gradually releasing the actual slip rotation speed N□, the predetermined target slip rotation speed N8. Raise it once towards . When it is detected that the actual slip rotation speed NSR has become equal to or greater than the predetermined determination value ΔN□1 (Nsm≧ΔN, 1), the process proceeds to step S34 shown in FIG. 1θ.

In step S34, the duty rate D14 of the coupling side solenoid valve 48 is set to the initial duty rate D□ calculated in step S20, and a signal for opening and closing the solenoid valve 48 is outputted at the duty rate Dg4. A predetermined duty rate ΔD4 (for example, 2 to 6%) is subtracted from the duty rate [)tm of the release side solenoid valve 47 set in the cycle to obtain a new duty rate DLR, and this duty rate DLI is used as the initial value and the actual Hydraulic control for feedback control of the slip rotation speed N□ to the predetermined target slip rotation speed NsO is started (step 535), that is, the TCU 16 waits for l duty cycles to elapse in the subsequent step 336, and then
The duty ratio DL,l of the release side solenoid valve 47 is set as follows for each duty cycle, and a drive signal for opening and closing the release side solenoid valve 47 at the set duty ratio DLll is output (step 33B).

(DLm)n=(Di),+PI'e,,+
Kw+(ea −em−+)...””(10 here, e
, is the target slip rotation speed N for this duty cycle,
. and the deviation of the actual slip rotation speed NSA (evr −Ns*
-Ns++), ea-r' is the target slip rotation speed N of the previous duty cycle. and actual slip rotation speed N,
The deviation is 3. □+K11l is a proportional gain and a differential gain, each of which is set to a predetermined value. (Di
), is an integral term, and is calculated by the following equation (lla).

(Di), 1 = (Di)−−+ + rl −e
+t +Dx+”・・”(lla)(Di) −−+
is an integral term set in the previous duty cycle, and Kl+ is an integral gain, which is set to a predetermined value.

DH+ is the amount of change Δ in the turbine shaft torque when the engine torque Te changes due to accelerator work during gear shifting, etc.
It is a correction value of the turbine shaft torque that is set according to Tt, and first, the amount of change ΔTt of the turbine shaft torque is calculated,
Duty rate correction amount DI11 according to this amount of change ΔTt
is calculated using the following equation 〇21.

D,,=a6・ATt ・−・・−Q7J Here, ATt is ΔT t −(
T t) −−(T t) *−+ ・・・・
... It is calculated in Q31, but in the power-off range described later, ATt --(Tt), , +(Tt), -t
−・・・04), (Tt) and (Tt)
7-r is the turbine shaft torque in the current and previous duty cycles, respectively set in step S13 in FIG. 4 above. Further, a6 is a constant that is preset according to the shift pattern. As can be seen from equations (lla) and (12), the integral term (Di) includes the duty rate correction ID□ determined by the amount of change ΔTt in the turbine shaft torque, so the duty rate DL11 can be Changes in turbine shaft torque can be corrected without delay, and the above-mentioned integral gain during feedback control
It is no longer necessary to set the proportional gain and the differential gain to large values, allowing for good followability and stable control.

Next, TCU16 sets the actual slip rotation speed tJsl to a negative predetermined slip rotation speed ΔN, 1 (for example, −3 to −7rp+
s) Determine whether it is less than or equal to (step 54G).

If the result of this determination is negative, the TCU1 6 returns to step S36 and repeats steps S36 to 340 until the actual slip rotation speed N□ becomes equal to or less than the negative predetermined slip rotation speed ΔN□. As a result, the duty rate DLll of the solenoid valve 47 on the release side is determined by the actual slip rotation speed N□ and the target slip rotation speed N, as described above.
. Feedback control is performed so that the difference between the actual slip rotation speed Nsm and the target slip rotation speed f'Jse becomes smaller, while the duty rate Ot4 of the solenoid valve 48 on the coupling side is equal to the initial duty rate. 1) u
It is kept constant at 2. As a result, the hydraulic pressure corresponding to the initial duty ratio I)us of the solenoid valve 48 is supplied to the second speed clutch 34 via the second hydraulic control valve 46.
A piston (not shown) of the clutch 34 gradually moves to the engagement side, and the clutch 34 starts to engage. As the clutch 34 starts to engage, the turbine rotation speed Nt tries to decrease, but since the engine 10 is in the power-on state, the turbine rotation speed is reduced by setting the duty rate DLII of the disengagement side solenoid valve 47 to a larger value. A decrease in Nt is prevented. However, the engagement of the clutch 34 on the engagement side progresses, and the duty rate D of the solenoid valve 47 on the disengagement side increases.
Even though L is set to a larger value, if the engagement force of the locking side clutch 34 exceeds this, the turbine rotation speed N
t starts to decrease, and at the time point 4 shown in FIG. 13(a), the actual slip rotation speed Ns becomes a negative predetermined slip rotation speed Δ.
Becomes NS+ or lower. When it is detected that the actual slip rotation speed Nml has become equal to or less than the negative predetermined slip rotation speed ΔNS+ (the determination result in step S40 is affirmative), the process proceeds to step S42 shown in FIG. 11.

In this way, the control section B shown in FIG. 13 (from time t3 to L
Hydraulic control in the four-point control period) ends.

In the control section B, when it is detected that the actual slip rotation speed N□ has become equal to or less than the negative predetermined slip rotation speed ΔNsI, step S42 in FIG. 11 is executed. For example, if the actual slip rotation speed N□ becomes equal to or less than the negative predetermined slip rotation speed ΔNS+ in consecutive duty cycles,
If the hydraulic pressure control is detected twice, the hydraulic control in the control section B may be omitted and the process may directly proceed to step 342 in FIG. 11 to start the hydraulic control of the control head C.

Hydraulic control in control section C and the subsequent control section E is based on the duty ratio of the solenoid valve 48 on the coupling side, 4, being the difference between the turbine rotation rate of change ωt and a predetermined target turbine rotation rate of change ωto. Feedback control is performed to the minimum value, and the turbine rotation speed Nt is gradually decreased toward the second speed calculation turbine rotation speed Ntc2. TCU1
6 is the duty rate DL of the solenoid valve 47 on the release side.
I+ is set to a predetermined duty rate DL, sax, and a drive signal is output to open and close the solenoid valve 47 at the set duty rate DLII (step 542), and this predetermined duty rate DLm@aX is the first hydraulic control valve. 44 to the first speed clutch 33 at a constant pressure (holding pressure), and maintain the piston position of the first speed clutch 33 at the position at time 14 shown in Fig. 13 (mountain). Set it to a value that allows. Note that the duty rate DLfi of the solenoid valve 47 on the release side applies the holding pressure to the first speed clutch 33 until the gear shift is substantially completed (from time t4 to time L8 shown in FIG. 13(b)). The predetermined duty rate D LmIlaX is maintained.

Next, the TCU 16 waits for a predetermined time t to elapse (step 543), and proceeds to step S44. In step S44, the target turbine rotation rate of change ωt is determined.

is set by the following equation G.

ωto=a7 ・N o +b7 ...-0
51 Here, a7. b7 is set to a predetermined value (negative value) according to the control interval C~, and a7. The b7 value sets the target turbine rotation rate of change ωtO set by 20ω to a value in which the turbine rotation speed Nt gradually decreases in the control interval C in which feedback control has just started, and in the control interval following the control interval C. The rate of change is set to a value larger than the absolute value of the rate of change in the control section C to accelerate the rate of decline of the turbine rotation speed Nt, and in the control section E, where the engagement of the second speed clutch 34 is completed, the absolute value of the rate of change is set to a smaller value again. This value is set to a certain value to prevent gear shift stagnation (see the time change in the turbine rotational speed Nt in FIG. 13(a)).

Next, the TCU 16 sets the duty rate D ta of the coupling side solenoid valve 48 to the duty rate at time t4 when it is detected that the actual slip rotation speed N□ has become equal to or less than the negative predetermined slip rotation speed ΔN□ as an initial value. The calculation is set using the following formula 00, and a drive signal is output to open and close the solenoid valve 48 at the set duty rate DR4 (step 5
46).

(Di4)n=(Di)n+Krg HEa +Kex
(E a E a-+)"'lle Second, Ell calculates the deviation ([! , =ωto-ωt), and the actual turbine rotation rate of change ωL is the actual turbine rotation speed (Nt) in the current and previous duty cycles, and (Nt).
It is determined from *-r using the following formula 01.

(ωt), t −(Nt), −(Nt)*−t
...07) Also, El-1 is the deviation between the target turbine rotation rate of change ωLo and the actual turbine rotation rate of change ωt of the previous duty cycle. , , , are a proportional gain and a differential gain, each of which is set to a predetermined value. (Di)n is an integral term and is calculated by the following formula 〇[D.

(mouth i) n = (Di) a-++ to lt ・, E
a + DNI + D, lt--0111 (Dl
)--+ is an integral term set in the previous duty cycle, and x+g is an integral gain, which is set to a predetermined value.

DNI is the amount of change Δ in turbine shaft torque when engine torque Te changes due to accelerator work during gear shifting, etc.
This is a correction value for the turbine shaft torque that is set according to Tt, and is obtained from the same arithmetic expressions as Equations 021 to OIO above.

DNt is a correction duty rate when changing the target turbine rotation rate of change, which is applied only when the control interval changes from C to D and from D to E, and is expressed by the following formula 〇! 11 and Q
Determined from [existence].

D,! =α・Δω[0 ・・・・・・ a
9Δωto=(auto), 1(auto)a−
1++++Q@Here, (ωto) is the target turbine rotation rate of change that should be applied after the current DA cycle, and (ωto)m−+ is the target turbine rotation rate of change that was applied until the previous time. . is a constant set according to the shift pattern.

In this way, the integral term (Di)n of the duty rate I)ta calculated for each duty cycle is also the same as the duty rate DLI of the release side solenoid valve 47 calculated in the aforementioned control section B. It is corrected by the correction amount DNI, that is, by the change amount ΔTt of the turbine shaft torque, and furthermore, when the control interval is changed, it is corrected according to the change amount ΔωtO of the target turbine rotation rate of change. It can be corrected without delay for changes in the target turbine speed change rate, and there is no need to set the above-mentioned integral gain, proportional gain, and derivative gain to large values during feedback control, and the followability is good. Moreover, stable control without hunting is possible.

After calculating the duty rate Dt4 and outputting the drive signal in step 346, the TCU 16 proceeds to step 34B,
Calculated turbine rotation speed N t when turbine rotation speed Nu is 2nd speed
Predetermined rotational speed just above a2 (calculated turbine rotational speed N at 2nd speed)
ΔNtc2 from ta2 (e.g. 80 to 12 Orpm)
It is determined whether the rotational speed Nta20 has been reached. If the result of this determination is negative, the process returns to step S43 and steps S43 to 34B are repeatedly executed.

At the point when the control section C has just entered, the coupling side clutch 34 has just started to engage, and by reducing the turbine rotation speed Nt at the target turbine rotation change rate ωto described above, the engagement amount at the beginning is reduced. Shift shock is avoided. Then, when the turbine rotation speed Nt decelerates and reaches the rotation speed obtained by multiplying the transfer drive gear rotation speed No. by a predetermined coefficient (for example, 2.8XNo.), the TCU 16
It is determined that the vehicle has left control zone C and entered control zone C.
Target turbine rotation rate of change ωto in step S44
The absolute value of is changed to a larger value (at time t5 in FIG. 13(a)).

When the absolute value of the target turbine speed change rate ωto is changed to a larger value, the duty rate Dt4 of the coupling side solenoid valve 48 is set to a value larger than the value set in the control section C (L5 in FIG. 13(C)). (from time t6 to time t6), the turbine rotational speed Nt rapidly decreases at approximately the target turbine rotational change rate ωto. The larger the absolute value of the target turbine rotation rate of change ωto is set, the better the shift response will be.

Next, when the turbine rotation speed Nt further decelerates and reaches the rotation speed obtained by multiplying the transfer drive gear rotation speed No by a predetermined coefficient (for example, 2.2XNo), that is, the piston of the second speed clutch 34 gradually engages. When it moves near the completion position, it is determined that it has left the control zone and entered the control zone, and the absolute value of the target turbine rotation rate of change ωto set in step S44 is changed to the value set in the control zone. When the absolute value of the target turbine rotation rate of change ωto is changed to a smaller value (time point 6 in FIG. 13(a)), the duty rate Dza of the coupling side solenoid valve 48 becomes smaller in the control interval. The turbine rotation speed NL is set to a value smaller than the set value (from time t6 to time t7 in FIG. 13(C)), and the turbine rotation speed NL gradually decreases at approximately the target turbine rotation change rate ωto, and the release side The engagement of the clutch 33 is completely released, thereby avoiding a shift shock near the time when the engagement of the engaged clutch 34 is completed.

If the determination result in step S4B is affirmative, that is, the turbine rotation speed Nt is the calculated turbine rotation speed N ta at 2nd speed.
When the predetermined direct rotation speed Ntc20 of 2 is reached (Fig. 13 (C
), TCU16 sets the timer to a predetermined time T.
Step 5: Set SF (for example, 0.5 sec)
50), wait for the elapse of predetermined times T and F (step 551).
), by waiting for the elapse of the predetermined time T3F, it is possible to reliably complete the engagement of the engaging side clutch 34.

When the predetermined time TIF has elapsed and the determination result in step S51 becomes affirmative, the TCU16 closes the release side solenoid valve 4.
7 and the duty rate DLl+ of the coupling side solenoid valve 48
Both DNAs are set to 100χ, and the duty rate DLl+D! At 4, the drive signal to open and close the solenoid valves 47 and 48 is output (also at time 8 in Figures 13(b) and (C)), thus turning on the power from the first gear to the second gear. Shift hydraulic control is completed.

Power-on downshift BW' Figures 14 to 16 are flowcharts showing the shift hydraulic control procedure in the case of power-on downshift.
An example of a shift hydraulic control procedure for downshifting from speed to first speed will be described with reference to FIG. 17.

The TCU 16 first operates the solenoid valve 47 in response to a power-on downshift shift signal from 2nd speed to 1st speed.
And the initial duty rate D□ and ILx of 48 are expressed by the above formula (
Calculation is performed using the following equations (21) and (22), which are similar to (8) and (9) (step 560).

Da+ =a8 ・l T t I +c8
・=-(21) Da*=a9 ・I T L l +c
9-(22) Here, a8. c8 and a9. c.
9 is a constant applied when downshifting from second speed to first speed.

Next, the TCUl6 sets the duty rate Dt4 of the solenoid valve 48 on the release side to the initial duty rate D□ set in step S60, outputs a signal to open and close the solenoid valve 48 at the duty rate Dt4, and An initial duty rate D41 is applied to the second speed clutch 34, which is an engagement element.
starts supplying initial hydraulic pressure corresponding to 2nd speed clutch 3.
The piston (not shown) No. 4 is moved back toward the position just before clutch slippage occurs (step 362, the first
7), on the other hand, the duty ratio DLR of the solenoid valve 47 on the coupling side is set to 0%, and a signal is output to open and close the solenoid valve 47 at the duty ratio DLR, that is, it is normally open. The solenoid valve 47 is fully opened and the piston of the first speed clutch 33, which is the frictional engagement element on the coupling side, is moved toward the position W (piston play reduction position) just before clutch engagement starts (Fig. 17). (C) tl
o) and sets the initial pressure supply time Ts2 in the timer (step 364).
When the normally open solenoid valve 47 is driven at the duty rate Ox and the working hydraulic pressure is supplied to the coupling side clutch 33,
The piston of the clutch 33 can be advanced to a predetermined position immediately before engagement starts.

In step 566, TCUl 6 determines whether the initial pressure supply time Ts2 set in step S64 has elapsed.
), if it has not yet passed, this initial pressure supply time Ts2
Iteratively executes step 366 and waits until .

If the determination result in step S66 is affirmative, that is, when the initial pressure supply time Ts2 has elapsed and the first speed clutch 33 has advanced to the predetermined position immediately before engagement, the TCU16 proceeds to step 368 in FIG. The duty ratio of the side solenoid valve 47 is set to a predetermined value DL, s+, which provides the holding pressure.
ax, and outputs a drive signal to open and close the solenoid valve 47 at this duty rate Dll (Fig. 17(C)
Note that the duty rate DLR of the solenoid valve 47 on the coupling side is maintained until the turbine rotational speed Nt reaches the calculated turbine rotational speed Ntel at 1st speed (Fig. 17 (
(from time tl1 to time tls shown in a)), a predetermined duty rate DI for applying the holding pressure to the first speed clutch 33.
Retained in IIIaX.

On the other hand, the piston of the clutch 34 on the disengagement side gradually moves to the disengaged side, and the friction torque of the clutch 34 is reduced, so that the turbine rotational speed NL gradually starts to increase. Then, the TCU16 determines whether or not the turbine rotation speed Nt has increased beyond a first predetermined determination value (for example, 1.5
If it does not exceed No, the determination in step 370 is repeated until the number exceeds No, and the process waits.

When the turbine rotation speed Nt exceeds the rotation speed 1.5XNo (
At time t12 in FIG. 17(a), the shift hydraulic control in control section A shown in FIG. After waiting for the elapsed time, hydraulic control is started to increase the turbine rotational speed N't toward the first speed calculation turbine rotational speed Ntd while adjusting the turbine rotational change rate ωt by feedbank control. That is, the hydraulic control in the control section B and the following control sections C and D is performed at the duty rate D! of the solenoid valve 48 on the release side. , is feedback-controlled to a value that minimizes the difference between the turbine rotation rate of change ωL and a predetermined target turbine rotation rate of change ωto, and the turbine rotation speed Nt is gradually increased toward the first speed calculation turbine rotation speed Ntel. It is.

First, in step S72, the TCU 16 sets the target turbine rotation rate of change ωto using the following equation (23).

ωto=alO・N o +blO−-(23) Here, alO, blo are set to predetermined values (positive values) according to the control section B-D, and the alo, blo values are calculated using the formula (
23) Target turbine rotation rate of change ωto set by
is set to a value in which the turbine rotation speed Nu gradually increases in control section B, which is just after feedback control is started, and control section B
In the control section C that follows, the rate of increase in the turbine rotation speed Nt is accelerated by setting the rate of change to a value larger than the rate of change in the control period B, and in the control section where the turbine rotation speed Nu approaches the calculated turbine rotation speed Ntel at 1st gear. , is set to a value that prevents the turbine rotational speed Nt from blowing up by setting it again to a small rate of change (turbine rotational speed Nt in Fig. 17(a)
).

Next, in step A, the TCU 16 sets the duty rate Dffi4 of the release side solenoid valve 48 so that the turbine rotational speed Nt is the rotational speed 1.
The duty rate at time t12, which exceeds 5
outputting a drive signal to open and close 8 (step 574);
Note that the integral gain in the above equations Oe and 08) is 18,
The proportional gain KP and the differential gain KD are each set to predetermined values that are optimal for the shift pattern in the power-on shift.

The TCU 16 has a duty rate O in step S74.
After calculating tg and outputting the drive signal, the process proceeds to step 376, where it is determined whether the turbine rotational speed Nt has reached the calculated turbine rotational speed Ntcl at the first speed. If the result of this determination is negative, the process returns to step S71 and steps 371 to S76 are repeatedly executed.

At the time when the control section B has just entered, the disengagement side clutch 34 has just started disengaging, and by increasing the turbine rotation speed Nt at the target turbine rotation speed change rate ωto described above, the turbine rotation speed Nt is increased. Blow-up is avoided. Then, the TCU 16 has a turbine rotation speed Nt
When the transfer drive gear rotation speed No. increases and reaches the rotation speed obtained by multiplying the transfer drive gear rotation speed No. by a predetermined coefficient (for example, 1.7 , the target turbine rotation rate of change ωto is changed to a larger value (see Fig. 17(a)).
) as of 113).

When the target turbine speed change rate ωto is changed to a larger value, the duty rate Dt4 of the release side solenoid valve 48 is set to a value smaller than the value set in the control 9n section B (FIG. 17), and from time 113 onwards tlJ. time),
Turbine rotation speed Nt is approximately the target turbine rotation rate of change ωto
will rise rapidly. The larger the target turbine rotation rate of change ωto is set, the better the shift responsiveness will be.

Next, when the turbine rotation speed Nt further increases and reaches the rotation speed obtained by multiplying the transfer drive gear rotation speed No by a predetermined coefficient (for example, 2.4XNo), that is, the engagement of the second speed clutch is gradually released. When the turbine rotation speed Nt approaches the 1st speed calculation turbine rotation speed N tel,
It is determined that the vehicle has left control zone C and entered control zone C.
The target turbine rotation rate of change ωto set in step S72 is changed to a smaller value than the value set in the control section C (time point 14 in FIG. 17(a)), and the target turbine rotation rate of change ωto is changed to a smaller value. When changing to , the duty ratio of the release side solenoid valve 48, 4, is set to a value larger than the value set in the control section C (from time tld to time tls in Fig. 17, etc.), and the turbine rotation speed Nt is approximately The turbine rotation speed will increase slowly at the target turbine rotation rate of change ωto, and a large overshoot of the turbine rotation speed Nt beyond the first speed calculation turbine rotation speed Ntel will be avoided.

The determination result in step S76 becomes affirmative, and it is detected that the turbine rotational speed Nt has reached the calculated turbine rotational speed Ntel at 1st speed (at time tls in Fig. 17(a))
, after completing the hydraulic control in the control section E, the hydraulic control in the control section E is started.
'Start B. Hydraulic control in this control section E is performed at an actual slip rotation speed of N8. The duty ratio of the solenoid valve 48 on the releasing side is feedback-controlled so as to minimize the deviation between the target slip rotation speed N3° (for example, 20 rpm+), and during this time, the first speed clutch 33 on the engaging side is engaged. It is controlled so that the pressure is gradually strengthened. That is, T.C.
In step 37B, U16 sets the duty rate of the solenoid valve 47 on the coupling side to an initial duty rate smaller than the duty rate DLRllmml set in step 560. and outputs a drive signal to open and close the solenoid valve 47 at the duty rate DLI (the first
(LlS point in Figure 7 (C)), this causes the first
The piston of the speed clutch 33 gradually begins to move toward the engagement side.

Next, in step S79, the TCU 16 performs a predetermined period of time 1. After waiting for the elapse of , the duty ratio DI4 of the release side solenoid valve 48 is determined by the following equations (24) and (2), which are similar to the above equations (11) and (lla), for each duty cycle.
4a) and outputs a drive signal to open and close the solenoid valve 48 at this duty rate DR4 (step 3).
80).

(Ota)n=(DI)a+□・ea+
Ke+(6a-e++-1)...(24)(Di
)n = (Di)a-++Kz He,l+D,,
...-(24a) Here, (Di)a-+ is the integral term set in the previous duty cycle,
As the initial value, the duty rate set immediately before time t15 when it is detected that the turbine rotation speed Nt exceeds the first speed calculation turbine rotation speed Ntd is used. ni,,ni2
□KSI is an integral gain, a proportional gain, and a differential gain, each of which is set to a predetermined value that is optimal for the power-on downshift. e7 is the target slip rotation speed N for the current DA cycle. and the deviation of the actual slip rotation speed NS1 (e a = Ns*-Ns++), e a-+
is the target slip rotation speed N of the previous duty cycle.

And the actual slip rotation speed N! , l is the deviation.

[)H+ is a correction value for the turbine shaft torque that is set according to the amount of change ΔTt in the turbine shaft torque when the engine torque Te changes due to accelerator work during gear shifting, etc., and this value is determined by the above-mentioned calculation formula 〇 Calculation is performed on the 21- side.

Next, in steps 382 to 85, the TCU 16 determines that the absolute value of the actual slip rotation speed N, l is smaller than the predetermined slip rotation speed (for example, 5 rpm) for two consecutive times.
It is determined whether or not it has been detected over the life cycle. That is, in step 382, it is determined whether the absolute value of the actual slip rotation speed N□ is smaller than the predetermined slip rotation speed (5rp-), and as long as the result of this determination is negative, the TCU
16 resets the flag FLG value to 0 (step 5
83), the process returns to step S79 and steps S79 to S82 are repeatedly executed. The friction torque of the clutch 33 on the engagement side is small, and in response to the increase in friction torque, the reduction amount (disengagement amount) of the friction torque of the clutch 34 is increased by the feedback control, and the engine 10 in the power-on state While the torque that tries to raise the turbine rotation speed Nt is superior, the turbine rotation speed Nt is calculated as the target slip rotation speed N from the turbine rotation speed N tel in the first gear. However, as the friction torque of the clutch 33 increases, the turbine rotation speed Nt gradually decreases, and the determination result in step S82 becomes affirmative, and step 384 is executed.

In step S84, it is determined whether the flag FLG value is equal to the value l. When the turbine rotational speed Nt decreases and it is determined to be affirmative for the first time in step S82, the determination result in step 384 becomes negative, and in such a case, the flag FLG value is set to the value 1 in step S85 and the step is repeated. Returning to S79, steps S79 & B and S80 are executed. Then, in step S82, when it is again determined that the absolute value of the actual slip rotation speed N□ is smaller than the predetermined slip rotation speed (5 rpa), that is, the absolute value of the actual slip rotation speed N□ is twice the predetermined slip rotation speed. If it is detected that the number is smaller than the number (time point 116 in FIG. 17(a)), the determination result in step S84 becomes affirmative, and the hydraulic control in the control area E ends and the process proceeds to step S.
87 will be executed.

In step 387, the TCU 16 sets the duty ratio DLII of the solenoid valves 47 and 48 on the coupling side and release side.
and I) z4 are both set to 0%, and the TCU 16 outputs no drive signal to the solenoid valves 47 and 48. Thus, the second gear clutch 34 is released and the first
After the engagement of the speed clutch 33 is completed, the shift hydraulic control for the power-on downshift from the second gear to the first gear is completed.

Power Off Shift It Suppression' Figures 18 to 20 are flowcharts showing the gear shift hydraulic control procedure in the case of a power off shift.
An example of a shift hydraulic control procedure for shifting up from a high speed to a second speed will be described with reference to FIG. 21.

The TCU 16 first calculates the initial duty rate [)uz of the solenoid valve 48 on the coupling side using the same calculation formula as the above formula (9) based on the shift signal of the power off shift from the first speed to the second speed. (Step 590).

Next, the TCU 16 determines the duty rate of the solenoid valve 47 on the release side. A predetermined duty rate DL that provides the holding pressure is
ITaa×, and outputs a signal to open and close the solenoid valve 47 at this duty rate DLR, so that the clutch completely slips on the piston (not shown) of the first speed clutch 33, which is the disengagement side frictional engagement element, and is engaged. In other words, when the engine 10 is in the power-off operation mode, the clutch is moved backward toward the standby position where the engine 10 can be immediately restarted (step S92, time t21 in FIG. 21(b)). Even if the clutch 33 is released from the clutch immediately after the output of the shift signal, there is no need to worry about the turbine rotational speed Nt rising; on the contrary, if the clutch 33 is not released as soon as possible, there is a risk that shift stagnation will occur. On the other hand, the duty rate Dza of the solenoid valve 48 on the coupling side is set to 100χ, and the duty rate Dt,
A signal for driving the solenoid valve 48 to open and close, that is, a driving signal for fully opening the solenoid valve 48, is output, and the engagement of the clutch is started for the piston of the second speed clutch 34, which is the engagement side frictional engagement element. Proceed toward the immediately preceding position (piston backlash reduction position) 1 (at time t21 in Figure 21 (C))
At the same time, the initial pressure supply time Tsl is set in a timer (step 393).

Then, the TCU 16 determines whether the initial pressure supply time Tsl set in step S93 has elapsed (step 595), and if it has not elapsed yet, repeats step S95 until the initial pressure supply time Tsl has elapsed.

If the determination result in step S95 is affirmative, that is, when the initial pressure supply time Tsl has elapsed and the second speed clutch 34 has advanced to the predetermined position immediately before engagement, the TCU 16 proceeds to step 396, and the coupling side solenoid valve 48 The duty rate D24 of Otm is set to the initial duty rate D0 calculated in step S90, and a valve opening drive signal for opening and closing the solenoid valve 48 at the duty rate Otm is output (21st
(time point t22 in Figure (C)).

Then, wait for the elapse of a predetermined time t, that is, the elapse of one duty cycle (step 398), and when the predetermined time t elapses, the duty rate I)z4 of the solenoid valve 48 set in the previous duty cycle is A predetermined duty rate ΔD5 is added to obtain a new duty rate I)xa, and a signal for opening and closing the solenoid valve 48 is outputted at this new duty rate D24 (step 599).The predetermined duty rate ΔD5 to be added is The duty rate D24 of the valve 48 is set to a predetermined speed (for example, duty rate D
(See the change in duty rate D14 from time t22 to time t23 in FIG. 21(C)).
.

Next, the process proceeds to step S100, and the TCU 16 calculates the actual slip rotation speed N□ using the above formula 0ω and converts it into a negative predetermined discrimination value ΔN□! (e.g. -8 to -12 rpm)
Compare with.

Actual slip rotation speed N! ll is compared with a predetermined judgment value ΔN, □, and the actual slip rotation speed Lm is a predetermined judgment value ΔN, □.

When it is larger than (N□>ΔN□t), TC:U16 returns to step 39B, and steps 39B to S
100 is repeated to gradually increase the duty rate D14 of the solenoid 48. As a result, the engaged clutch 34 starts to engage, and the friction torque of the clutch 34 gradually increases. Then, the turbine rotation speed Nt gradually decreases, the determination result in step 5100 becomes affirmative, and the TCU 16 proceeds to step 5102 shown in FIG. Start.

Hydraulic control in control section B and subsequent control sections C and D sets the duty rate Dt& of the solenoid valve 48 on the coupling side so that the difference between the turbine rotation rate of change ωL and a predetermined target turbine rotation rate of change ωto is the minimum. Feedback control is performed to a value such that the turbine rotation speed Nt is calculated at 2nd speed.
The amount is gradually decreased toward tc2.

First, in step 5102, the TCU 16 waits for the elapse of l duty cycles (elapse of the predetermined time t), and then changes the target turbine rotation rate of change ωto into the control interval B-
D is set to a predetermined value stored in advance. Target turbine rotation rate of change ωt set in each control section B-D
o is set to a value at which the turbine rotational speed Nt gradually decreases in control section B, which is just after feedback control is started, and to a value larger than the absolute value of the rate of change in control section B in control section C that follows control section B. to accelerate the rate of decline of the turbine rotational speed Nt, engagement of the second speed clutch 34 is almost completed, and the turbine rotational speed QNt approaches the calculated turbine rotational speed Ntc2 in the second speed (in the control section E, the rate of change again increases). The absolute value is set to a small value to prevent shift shock (see the time change in the turbine rotational speed Nt in FIG. 21(a)).

Next, the TCU 16 sets the duty rate of the coupling side solenoid valve 48 to a predetermined slip rotation speed ΔN3 where the actual slip rotation speed Ns1 is negative! (for example, −8 to 12 rpm) or less, the duty rate at time t23 is set as the initial value, and the above calculation formula 06) and (Il+3
A drive signal is set to open and close the solenoid valve 48 at the set duty rate own (step 5106). Note that the integral gain applied to the above calculation formulas 0ω and 0ω is 18, the proportional gain KPt, and the differential gain. and □ are respectively set to predetermined values that are optimal for the shift pattern of the power off shift.

The TCU 16 sets the duty rate D at step 5106.
After the calculation at t4 and the output of the drive signal, step 3107
Then, the turbine rotation speed Nt decreases to a predetermined rotation speed Ntc20 (2
From the speed calculation turbine rotation speed N tc2, ΔNtc2 (
For example, it is determined whether the rotational speed has reached a higher rotational speed (for example, 80 to 120 rpm). If the result of this determination is negative, the process returns to step 5102, and step 5102
Steps 3107 to 3107 are repeatedly executed.

At the time when the control section B has just entered, the coupling side clutch 34 has just started engagement, and by reducing the turbine rotation speed NL at the target turbine rotation speed change rate ωto described above, the shift at the start of engagement is performed. Shock is avoided. Then, when the turbine rotation speed Nt decelerates and reaches the rotation speed obtained by multiplying the transfer drive gear rotation speed NO by a predetermined coefficient (for example, 2.8XNo), the TCU 16
It is determined that the vehicle has left control zone B and entered control zone C.
In step 5104, the absolute value of the target turbine rotation rate of change ωto is changed to a value larger than the value applied to control section C (at time t24 in FIG. 21(a)).

When the absolute value of the target turbine rotation rate of change ωto is changed to a larger value, the duty rate Dt4 of the coupling side solenoid valve 48 is set to a value larger than the value set in the control section B (t24 in FIG. 21(C)). (from time point to time point L25), the turbine rotation speed Nt rapidly decreases at the target turbine rotation rate of change ωto, which is set to approximately this large value. Note that the larger the absolute value of the target turbine rotation rate of change ωto is set, the better the shift responsiveness will be.

Next, the turbine rotation speed Nt is further reduced to a product value (
For example, when reaching 2.2 The absolute value of the target turbine rotation rate of change ωto set in step 5104 is changed to a smaller value than the value set in the control section C (at time t25 in FIG. 21(a)). is changed to a smaller value, the duty ratio [)ta of the coupling side solenoid valves 4 and 8 is set to a smaller value than the value set in the control section C (from time t25 to time t26 in Fig. 21(C)). ), the turbine rotational speed NL will slowly decrease at approximately the target turbine rotational change rate ωto, and the turbine rotational speed Nt near the point where engagement of the coupling side clutch 34 is completed is equal to the calculated turbine rotational speed N tc2 at 2nd speed. This results in a smooth transition and avoids shift shock.

If the determination result in step 5107 is affirmative, that is,
Calculated turbine rotation speed N t when turbine rotation speed Nu is 2nd speed
When the predetermined direct rotation speed N tc20 of c2 is reached (time t26 in FIG. 21(C)), TC [J16 sets a predetermined time TSF (for example, 0.5 sec) in the timer (Step 5109), and this predetermined time Wait for the progress of T and F (
Step 5110) By waiting for the predetermined times T and F to elapse, engagement of the engaging clutch 34 can be completed reliably.

When the predetermined time T'sr has elapsed and the determination result in step 5110 becomes affirmative, the process proceeds to step 5112 and the TCU
16 is the duty rate DLII Dta of the release side solenoid valve 47 and the connection side solenoid valve 48, both of which are 100.
χ, and outputs a drive signal to open and close the solenoid valve 47.48 at the duty rate DLl+DMA (second
1(b) and (C) at t27), thus the first
The shift hydraulic control of the power off shift from the gear to the second gear is completed.

Pressure 1 during power-off downshift FIGS. 22 to 24 are flowcharts showing the gear shift hydraulic control procedure in the case of power-off downshift.
An example of a shift hydraulic control procedure for downshifting from 1st gear to 1st gear will be described with reference to FIG. 25.

'l'cU16 first changes the initial duty rate D□ of the solenoid valves 47 and 48 in response to a shift signal for a power-off downshift from 2nd speed to 1st speed. is calculated using the above equations (21) and (22) (step 3
114), a8. applied in equations (21) and (22). c8 and a9. c9 is set to a predetermined value that is optimal for power-off downshifting from second gear to first gear.

Next, the TCU 16 sets the duty rate [1z4 of the solenoid valve 48 on the release side to the initial duty rate Ddl set in step 5114, outputs a signal to open and close the solenoid valve 48 at the duty rate D24, and reduces the friction on the release side. The piston (not shown) of the second speed clutch 34, which is an engagement element, is retreated toward the position just before clutch slippage occurs (step 5115, point 131 of the 251st mountain), while the solenoid valve 47 on the coupling side The duty rate D0 is set to 100χ, and a signal is output to open and close the solenoid valve 47 at the duty rate DL11, and the piston of the first speed clutch 33, which is the coupling side frictional engagement element, starts to engage the clutch. At the same time, the initial pressure supply time Ts2 is set in the timer (step 5116).

The TCU 16 waits for the elapse of a predetermined time t, that is, one duty cycle (28,6 m5ec) (step 5118), and when the predetermined time L++ elapses, the duty rate set in the previous duty cycle is set.

A new duty rate Dz4 is obtained by subtracting a predetermined duty rate ΔD6 from
0), the predetermined duty rate ΔD6 calculated by M is set to a value at which the duty rate of the solenoid valve 48 decreases at a predetermined speed (for example, a value at which it decreases at a rate of 8 to 12% per second). (See the change in duty rate DZ4 from time t31 to time 133 in Figure 0)), and T
Col6 determines whether or not the initial pressure supply time Ts2 set in step 5116 has elapsed (step 5122), and if it has not elapsed yet, it determines in step 51
18 and repeats steps 3118 to 5122. As a result, the duty ratio I)x4 of the solenoid 48 gradually decreases, and the clutch 3 on the releasing side
4 gradually moves toward the bond release start position.

If the determination result in step 5122 is affirmative, that is, when the initial pressure supply time Ts2 has elapsed and the first speed clutch 33 has advanced to the predetermined position immediately before starting engagement, TCol6 proceeds to the 23rd step 5124, and the solenoid The initial duty rate calculated in step 5114 is the duty rate DL11 of the valve 47. and this duty rate D
At L11, a drive signal for opening and closing the solenoid valve 47 is output (at time 132 in FIG. 25(C)), whereby the piston of the engagement side clutch 33 continues to gradually move toward the engagement start position. Note that the solenoid valve 47 duty rate DLR remains constant until it enters a control section C (25th
(L34 point in figure (C)), the initial duty rate Dot
is maintained.

Then, TCol 6 waits for the elapse of a predetermined time t, i.e. 1
Wait for the duty cycle to pass (step 5t25)
, after the predetermined time t, the step 5120
In the same manner as above, the calculation of the new duty rate and the output of the valve opening drive signal are continued (step 3126
), and then proceeds to step S128, where TCol6
calculates the actual slip rotation speed N□ using the following equation (25) and sets it as a negative predetermined discrimination value ΔN□□ (for example, -8 to -1
2rpeg) Seven comparisons.

N*m-NtNtc2...
・” (25) Here, N tc2 is the calculated turbine rotation speed at 2nd speed, and the transfer drive gear rotation speed NO
It is obtained as the product value obtained by multiplying by a predetermined number.

A predetermined judgment value ΔN, where the actual slip rotation speed N5Il is negative.

When it is larger than (N, 〉ΔN, .), TCol6 returns to step 5125 and performs steps 5125 to 5.
12B-t-run repeatedly. As a result, the disengaged second speed clutch 34 is gradually disengaged and released.

At this time, if engagement of the first speed clutch 33 on the coupling side has not yet started, the turbine rotation speed Nt gradually decreases (control interval A in FIG. 25(a) (shift signal output time t31). The actual slip rotation speed NI is determined by the predetermined judgment value Δ
The second half of the control period (up to the time t33 when it is detected that the actual slip rotation speed is below Ns□), and the actual slip rotation speed N
8. If it is detected that has become equal to or less than the predetermined determination value Δr'Jsmt (N□≦ΔN□3), the process proceeds to step 5130.

In step 5130, TCol6 adds a predetermined duty rate ΔD7 (for example, 2 to 6
%) and set the duty rate 014 which is larger by the deniity rate ΔD7, and set this duty rate D□ as the initial value, and set the actual slip rotation speed N□ and the predetermined target slip rotation speed N□ (for example, −2 Orpm). Deviation e, (-Ns
tNs++) is started. That is, when the engagement of the engagement clutch 33 has not yet started, if the duty rate D□ of the disengagement clutch 34 is set to a smaller value, the turbine rotational speed Nt will tend to decrease due to a decrease in friction torque. On the other hand, duty rate D! If 4 is set to a larger value, the turbine rotational speed NL tends to rise due to an increase in friction torque, so it is possible to maintain the turbine rotational speed Nt at a predetermined rotational speed by feedback control of the duty ratio Dsa.

Therefore, after waiting for l duty cycles to elapse in step 3132, TCol6 sets the duty rate Dt4 of the release side solenoid valve 48 every duty cycle using the above-mentioned formula (24) (step 3134).
, Incidentally, the integral gain Kll, the proportional gain □, and the differential gain Kal applied to the arithmetic expression are each set to predetermined values that are optimal for power-off downshift.

Next, the TCU 16 sets the actual slip rotation speed N8. It is determined whether or not the rotational speed is greater than or equal to a predetermined slip rotational speed ΔN5z (for example, 3 to 8 rpm) (step 5135).

If this determination result is negative, the TCU 16 returns to step 5132, and the actual slip rotation speed N□ is equal to the predetermined slip rotation speed ΔN! Steps 5132 to 5135 are repeatedly executed until the above is reached. As a result, the duty ratio DR4 of the solenoid valve 48 on the release side is determined by the actual slip rotation speed N□ and the target slip rotation speed N□, as described above.
In other words, the actual slip rotation speed N
, is the target slip rotation speed N1. On the other hand, the duty rate DLI of the solenoid valve 47 on the coupling side is the initial duty rate. is kept constant.

As a result, the initial duty rate D4□ of the solenoid valve 47
The working oil pressure corresponding to 1 is supplied to the first speed clutch 33 via the first oil pressure control valve 44, engagement of the clutch 33 is started, and the piston (not shown) gradually moves to the engagement completion position. Due to the movement of the piston of the clutch 33, the turbine rotational speed Nt starts to increase. This turbine rotation speed NL
The duty rate [)za of the solenoid valve 48 is set to a smaller value so as to cancel out the increase in the duty rate I)z.
The value of a gradually decreases. Despite setting the duty rate Dt4 of the disengagement side solenoid valve 48 to a smaller value, the turbine rotational speed NL increases due to the increase in the engagement force of the engagement side clutch 33, as shown in FIG. 25(a). t34
At this point, the actual slip rotation speed 1'lsm becomes the predetermined slip rotation speed ΔN,! That's all. When the TCU 16 detects that the actual slip rotation speed Ns becomes equal to or higher than the predetermined slip rotation speed ΔN0 (the determination result in step 5135 is affirmative), the process proceeds to step 5136 shown in FIG. Control section B (from time t33 to t34
Hydraulic control in the control section between points in time) ends.

In the control section B, when it is detected that the actual slip rotation speed Nl has become equal to or higher than the predetermined slip rotation speed ΔN3, step 5136 in FIG. 24 is executed. If it is detected, for example, twice in consecutive duty cycles that the actual slip rotation speed N
The process may proceed to step 5136 in FIG. 4 and start hydraulic control in control area C.

Hydraulic control in control section C and subsequent control sections E is such that the duty rate DLI of the solenoid valve 47 on the coupling side is set such that the difference between the turbine rotation rate of change ωL and a predetermined target turbine rotation rate of change ωto is the minimum. Feedback control is performed to a value such that the turbine rotation speed Nt is calculated at 1st speed as the turbine rotation speed N
The value is gradually increased toward tel.

First of all, TCU16 is Step S! 36, the duty rate Dt4 of the release side solenoid 48 is set to a predetermined duty rate D1.36 that provides the holding pressure. s+in to supply holding pressure to the second speed clutch 34, and then, after waiting for the elapse of a predetermined time Lll (step 513B), the predetermined value prestored in the storage device is set in the control interval C-. E, and use this as the target turbine rotation rate of change ωto
(Step 5139), the read target turbine rotation rate of change ω0 is set as the turbine rotation '
The number Nt is set to a small value that gradually decreases, and in the control section following the control section C, it is set to a value larger than the rate of change of the control section C to accelerate the rate of decline of the turbine rotation speed Nt and maintain the first speed clutch 33. In the control section E where the shift is completed, the change rate is again set to a small value to prevent shift shock (second
(See the time change in the turbine rotational speed Nt in Fig. 5(a)).

Next, the TCU 16 adjusts the duty rate DLII of the coupling side solenoid valve 47 to a value at time t34 when it is detected that the actual slip rotation speed N□ has become equal to or higher than the predetermined slip rotation speed ΔN0.
The duty rate at Oat, that is, the initial duty rate Oat
is calculated and set using the above-mentioned calculation formulas 00 and (18) and the following similar formulas (26) and (26a) as initial values, and a drive signal to open and close the solenoid valve 47 at the set duty rate DLll is output (step 5140 ).

(Lm)n=(Di), ++[PI −Ea +[
Il+(Ha −Ha−+)−′・(26)(Di)n
= (Di)m-++ ++ HEn + ro □,
10on--(26a) Here, (Di), is the integral term set in the previous duty cycle, and Kl3
．． KPI and K□ are integral gain, proportional gain, and differential gain, and eEll is set to a predetermined value that is optimal for the power-off downshift. eEll is the target turbine for the current duty cycle set in step 5139. Deviation between rotational change rate ωto and actual turbine rotational change iωt (ETl = (11) LO-ωt), Ea-
+ is the deviation between the target turbine rotation rate of change ωto and the actual turbine rotation rate of change ωt of the previous duty cycle.

DNI is the amount of change Δ in the turbine shaft torque when the engine torque Te changes due to accelerator work during gear shifting, etc.
This is a correction value of the turbine shaft torque set according to Tt, and this value is calculated using the above-mentioned calculation formula (121 to 0).

DNt is a correction duty rate at the time of changing the target turbine rotation rate of change, which is applied only when the control section changes from C to D and from D to E, and is calculated by the above-mentioned calculation formula (19
) and (20). Note that the coefficient α in equation (19) is set to an optimal value for the power-off downshift shift pattern.

The TCU 16 sets the duty rate D at step 5140.
After calculating LI+ and outputting the drive signal, step 514
2, the turbine rotation speed Nt is set to a predetermined rotation speed (for example, 80 to 120r
It is determined whether the rotational speed NtclO has been reached, which is lower by pm). If the result of this determination is negative, the process returns to step 5138 and steps 8138 to 5142 are repeatedly executed.

At the time when the control section C has just entered, the coupling side clutch 33 has just started engagement, and by increasing the turbine rotation speed Nt at the above-mentioned target turbine rotation change rate ωto, the shift at the start of engagement is performed. Shock is avoided. Then, when the turbine rotation speed Nt increases and reaches the rotation speed obtained by multiplying the transfer drive gear rotation speed No. by a predetermined coefficient (for example, 1.7XNo), the TCU 16
It is determined that the vehicle has left control zone C and entered control zone C.
In step 5139, the target turbine rotation rate of change ωto is changed to a larger value (t in FIG. 25(a)).
(as of 35).

When the target turbine rotation rate of change ωto is changed to a larger value, the duty rate D□ of the coupling side solenoid valve 47 becomes 1.
It is set to a value smaller than the value set in the IIIJ control section c5 (from time t35 to time L36 in FIG. 25(C)), and the turbine rotation speed Nt rapidly increases at approximately the target turbine rotation change rate ωto. It turns out. The larger the target turbine rotation rate of change ωto is set, the better the shift responsiveness will be.

Next, when the turbine rotation speed Nt further increases and reaches the rotation speed obtained by multiplying the transfer drive gear rotation speed No by a predetermined coefficient (for example, 2.4XNo), that is, the piston of the first speed clutch 33 gradually becomes When the turbine rotation speed Nt approaches the 1st speed calculation turbine rotation speed Ntel, it leaves the control section and #
It is determined that the control interval E has entered, and the target turbine rotation rate of change ωto set in step 5139 is changed to a value smaller than the value set in the control interval (25th
(at time t36 in Figure (a)), target turbine rotation rate of change ωt
.

When changing to a smaller value, the coupling side solenoid valve 47
The duty rate D0 is set to a value larger than the value set in the control interval (from time t36 to time t37 in FIG. 25(C)), and the turbine rotation speed Nt is gradually increased at approximately the target turbine rotation rate of change ωto. As a result, the shift shock that occurs near the time when engagement of the engagement side clutch 33 is completed is avoided.

If the determination result in step 5142 is affirmative, that is,
Calculated turbine rotation speed Nte when turbine rotation speed Nt is 1st speed
When the rotational speed NtclG is lower than l by a predetermined rotational speed (80=120rpm) (13 in Fig. 25(C)),
7), the TCU 16 immediately calculates the duty ratio D of the release side solenoid valve 48 and the connection side solenoid valve 47! 4+
Both DLRs are set to 0%, and the duty rate D84
゜DLI+ outputs a drive signal to open and close solenoid valves 48 and 47 (Fig. 25 Φ and time t37 in (C))
.

In this way, the shift hydraulic control for the power-off downshift from the second gear to the first gear is completed.

In the above-mentioned embodiment, for the sake of simplifying the explanation, the first gear
Although only the hydraulic pressure control element i during a shift between gears has been explained, the hydraulic control procedure during a shift between other gears, such as a shift between the second and third gears, will also be explained in the same way. Of course it is possible.

Furthermore, although a hydraulic clutch has been described as an example of a frictional engagement element for shifting of an automatic transmission, the frictional engagement element for shifting is not limited thereto, and may be a brake for shifting.

In the above embodiment, the engine torque detection method of the present invention was applied to an automatic transmission device equipped with a torque converter for a vehicle. The transmission torque is not limited to those that can be approximately quantitatively determined from the rotational speed of the input and output shafts, such as slip control electromagnetic powder clutches and viscous clutches, or the transmission torque can be externally controlled and the transmission torque It goes without saying that the engine torque detection method of the present invention can be applied to, for example, a slip-type direct coupling clutch such as the damper clutch 28 of the embodiment, in which a corresponding control parameter value can be detected.

Further, in the above-described embodiments, an explanation was given of an application applied to hydraulic control of a speed change clutch of an automatic transmission, but the present invention is not limited thereto, and is applicable to hydraulic control of a slip-type direct coupling clutch such as the damper clutch 28. , current control of the electromagnetic powder clutch, engine torque control that adjusts the fuel supply amount to minimize the deviation between the engine torque and the target value set from the vehicle speed and accelerator pedal opening, and when the tire slip acceleration is large. According to the detected engine torque, when the engine torque is larger than a predetermined value, the engine torque reduction rate is set to a large value, and when it is smaller than the predetermined value, the engine torque reduction rate is set to a small value to prevent hunting. It is applicable to controls, etc.

(Effects of the Invention) As described in detail above, according to the engine torque detection method of the present invention, in the engine torque detection method of a drive system in which the transmission torque is set at a detectable drive speed, the change rate of the engine rotation speed is is detected, and the product value obtained by multiplying the detected transmission torque of the driving force transmission device and the detected engine rotational speed change rate by a predetermined value is calculated.

Since the added value is detected as the engine torque, there is no need to specially provide large-scale devices such as strain gauges and slip rings. This provides an excellent effect in that calculation and detection can be carried out accurately and reliably.

[Brief explanation of the drawing]

The drawings show one embodiment of the present invention, and FIG. 1 is a schematic diagram of an automatic transmission equipped with a torque converter in which the method of the present invention is implemented, and FIG. 2 is a diagram showing the gear transmission 30 shown in FIG. 1.
3 is a hydraulic circuit diagram showing a part of the internal structure of the hydraulic circuit 40 shown in FIG. 1, and FIG. 4 is a transmission control unit shown in FIG. 1. FIG. 5 is a flowchart of the main program routine showing the oil pressure control procedure during gear shifting executed by the (TCU) 16. Timing chart showing the occurrence situation, Part 6
The figure is a shift map diagram showing the gear range divided by the throttle valve opening and the transfer drive gear rotation speed, Figure 7 is a flowchart of the power on/off determination routine, and Figures 8 to 12 are power on/off determination routines. a flowchart showing the hydraulic control procedure executed during a shift;
Figure 13 shows the turbine rotation speed Nt and transfer drive gear rotation speed No. during power-on upshift.
14 to 16 are flowcharts showing the hydraulic control procedure executed at the time of power-on downshift, and FIG. Timing charts showing changes in the turbine rotational speed N [and transfer drive gear rotational speed No.] and changes in the duty ratios of the release-side and coupling-side solenoid valves during power-on downshift, FIGS. 18 to 20, FIG. 21 is a flowchart showing the hydraulic control procedure executed during a power-off shift, and shows temporal changes in the turbine rotation speed Nt and transfer drive gear rotation speed No, as well as the release side and connection side solenoid valves during a power-off shift. Timing charts showing changes in duty rate, Figures 22 to 24 are flowcharts showing hydraulic control procedures executed during power-off downshift, and Figure 25 shows turbine rotational speed Nt and transfer drive during power-off downshift. Gear rotation speed N
A timing chart showing changes in O over time and changes in the duty rates of the release side and connection side solenoid valves. Fig. 26 shows the changes over time in the valve opening of the throttle valve, the turbine shaft torque, and the output shaft torque during lift foot upshifting. 2 is a timing chart for explaining. IO...Internal combustion engine, lla...Ring gear, 1
4...Ne sensor, 15...Nt sensor, 16...
・Transmission control unit) (TCU
), 17...No sensor, 19...oil temperature sensor, 2
0... Torque converter, 21... Drive shaft, 23.
... Pump, 25... Turbine, 28... Damper clutch, 30... Gear transmission, 30a... Turbine shaft (input shaft), 31... First drive gear, 32...
-Second drive gear, 33.34...Hydraulic clutch (speed change clutch), 35...Intermediate transmission shaft, 41...Oil passage, 42...Pilot oil passage, 44...First oil passage Hydraulic control valve, 46... Second hydraulic control valve, 47... Normally open solenoid valve, 48... Normally closed solenoid valve, 50.
. . . Damper clutch hydraulic control circuit, 52 . . . Damper clutch control valve, 54 . . . Damper clutch control solenoid pulp. Applicant - Mitsubishi Motors Corporation Agent Patent Attorney Kanji Nagato Figure 2 Figure 6 Q Postal Number N□ (rpm) Figure 4 Figure 5 Figure 7 Figure 8 Figure 11 Figure 13 Figure 15 Fig. 16 Fig. 17 Fig. 1 Time t;

Claims (3)

[Claims] (1) In an engine torque detection method for a drive system in which the driving force of an engine is transmitted to wheels via a driving force transmission device capable of detecting transmitted torque, the rate of change in the rotational speed of the engine is detected, and the driving force is detected. An engine torque detection method characterized by adding the detected transmission torque of the transmission device and a product value obtained by multiplying the detected engine rotational speed change rate by a predetermined value, and detecting the added value as the engine torque. (2) Engine torque detection according to claim 1, wherein the transmitted torque of the driving force transmitting device is detected using the rotation speeds of the input shaft and output shaft of the driving force transmitting device as parameters. Method. (3) The transmission torque of the driving force transmission device can be controlled from the outside, and the transmission torque is detected by detecting a control parameter value corresponding to the transmission torque. engine torque detection method.