Control system for automatic transmission

A control system for an automatic transmission, which is capable of controlling the hydraulic pressures for frictional engagement units when a predetermined shift is performed such that a first frictional engagement unit is engaged and a second frictional engagement unit is released, being connected to a power source, the revolving speed of which is temporarily raised when a predetermined shift down is performed, the control unit for the automatic transmission for a vehicle being structured such that shift which is performed by engaging and releasing the two frictional engagement units is judged, whether the judged shift is shift down in which the revolving speed of the power source is temporarily raised or shift down in which the revolving speed of the power source is not raised is judged, and the contents of control of the hydraulic pressures for the frictional engagement units are changed between shift down in which the revolving speed of the power source is temporarily raised and shift down in which the revolving speed of the power source is not raised.

TECHNICAL FIELD 
The present invention relates to a control system for an automatic 
transmission, and more particularly to a system for controlling hydraulic 
pressure for use in so-called clutch-to-clutch shift or direct pressure 
control in accordance with control of a power source, such as an engine or 
a motor. 
BACKGROUND ART 
Since shift of the gear stage in an automatic transmission involves change 
in the rotations of a plurality of rotative elements, the inertia forces 
of the rotative elements must be absorbed to make the torque shift to be 
performed smoothly in order to prevent shift shock. Control for preventing 
the shift shock has been performed by controlling engaging pressure or 
releasing pressure for frictional engagement units, such as clutches and 
brakes, for performing the shift so as to absorb the inertia forces 
(energy) attributable to sliding of the frictional engagement units. 
Change in the rotations of the engine which occurs when shift is performed 
becomes different between a power-on state in which the accelerator pedal 
has been depressed and a power-off state contrary to the power-on state. 
Therefore, when clutch-to-clutch shift is performed in which the states of 
engagement of two frictional engagement units are simultaneously changed, 
the engaging pressure for the on-coming frictional engagement unit is made 
adaptable to the state of revolution of the engine when the shift is 
performed. 
When shift down, which is clutch-to-clutch shift, is performed, the 
engaging pressure for the on-coming frictional engagement unit is 
gradually raised (swept up) after rise in the engine revolving speed to 
the synchronized revolving speed at the gear stage set by the shift down 
because the engine revolving speed is attempted to be raised in the 
power-on state. In the power-off state, the engine revolving speed is 
undesirably lowered if the frictional engagement unit which has realized 
the gear stage is released. Therefore, the engaging pressure for the 
on-coming frictional engagement unit is enlarged in an early stage to 
raise the engine revolving speed to the synchronized revolving speed at 
the gear stage set by the shift down. That is, the engaging pressure for 
the frictional engagement unit is controlled to be adaptable to the 
tendency of the change in the revolving speed of the engine when the shift 
is performed. 
When the clutch-to-clutch shift is performed, learning control is performed 
such that the hydraulic pressure is corrected in accordance with the state 
of fuel injection in the engine and the tied-up state when the previous 
shift has been performed and the shift is performed with the corrected 
hydraulic pressure when the next shift is performed. 
That is, the clutch-to-clutch shift is performed such that the hydraulic 
pressure for at least one of the frictional engagement unit of the 
frictional engagement units for performing the shift is successively 
changed to correspond to the state of progress of the shift to prevent 
shock attributable to rapid change of the output torque. In this case, 
change in the revolving speed (the engine revolving speed) input to the 
automatic transmission is affected by the input torque, the friction 
coefficient of the frictional member or the change rate of the hydraulic 
pressure. Thus, there arises a possibility that fuel injection in the 
engine is undesirably performed excessively or a tied-up state occurs on 
the contrary. 
Therefore, the foregoing problems have been prevented by correcting the 
controlled value of the hydraulic pressure in accordance with the detected 
state when the shift has been performed and the next clutch-to-clutch 
shift is controlled in accordance with the corrected controlled value. 
Since the foregoing control is able to use the individual difference in 
the automatic transmission and the factor such as the change of the 
frictional engagement unit as the time lapses in the control of the shift, 
control of the shift suitable for each case can be performed. Therefore, 
shift shock occurring when the clutch-to-clutch shift is performed can be 
prevented more satisfactorily. 
Since it is preferable that rapid change in the rotations is prevented in 
order to prevent shift shock, a throttle valve of the engine has been 
electronically controlled to also control the engine revolving speed as 
well as the control the hydraulic pressure for the automatic transmission, 
in recent years. An example of the foregoing structure has been disclosed 
in Japanese Patent Laid-Open No. 5-231525 (JPA-5-231525). 
The invention disclosed as described above relates to a hydraulic pressure 
control when so-called synchronizing shift is performed in which the 
opening of the throttle is enlarged by detecting the shift down when the 
shift down has been performed in a state where the throttle valve is 
closed. Thus, the engine revolving speed is synchronized with the 
revolving speed at the gear stage after the shift and shift down is 
performed in the foregoing state. Moreover, hydraulic pressure control 
means for preventing or restraining rise in the line pressure occurring 
attributable to the temporary enlargement of the opening of the throttle 
when the synchronizing shift is performed is provided. Thus, shock 
occurring because of the rapid torque capacity of the frictional 
engagement unit when the shift down is performed is prevented. 
When the above-mentioned clutch-to-clutch shift or direct control of the 
pressure is performed, initial hydraulic pressure control has been 
performed in which the hydraulic pressure which is applied to the 
on-coming frictional engagement unit is temporarily raised simultaneously 
or immediately after the shifted output has been performed to reduce a 
so-called pack clearance so as to cause the frictional engagement unit to 
immediately be provided with a torque capacity when higher hydraulic 
pressure is applied. 
The initial hydraulic pressure control is a control which is capable of 
bringing the frictional engagement unit into a standby state in which the 
frictional engagement unit can immediately and substantially be engaged. 
That is, if insufficient control is performed such that the initial 
hydraulic pressure is too low, timing for the frictional engagement unit 
to substantially be engaged is delayed and thus the shift response 
deteriorates. If the initial hydraulic pressure is too high, the 
frictional engagement unit is undesirably provided with an excessively 
large torque capacity. As a result, there arises a risk that a next 
control of low-pressure standby cannot satisfactorily be performed. 
The above-mentioned synchronizing shift is performed at a down shift in a 
substantial power-off state when a manual selection by a driver is carried 
out, for example. If the shift down is performed by so-called 
clutch-to-clutch shift in which engagement/release states of two 
frictional engagement units are simultaneously changed, the engaging 
pressure for the off-going frictional engagement unit is relatively early 
swept up to raise the engine revolving speed to the synchronized revolving 
speed at the gear stage after the shift. That is, in accordance with the 
state of the revolution of the engine when the shift down has been judged, 
the control of the hydraulic pressure for the frictional engagement unit 
is judged and performed. 
On the other hand, the engine is controlled such that the revolving speed 
is raised in accordance with a fact that the shift down is the 
synchronizing shift. Since the control to raise the revolving speed is 
performed by temporarily opening the throttle valve, also the engine 
torque is simultaneously enlarged. 
In this case, the automatic transmission is controlled in accordance with 
the contents of control in the power-off state and the engaging pressure 
for the on-coming frictional engagement unit is raised when the shift is 
completed. However, since the throttle opening is enlarged after the shift 
has been started, a power-on state is undesirably realized. As a result, 
the control of the hydraulic pressure for the automatic transmission and 
the state of the operation of the engine do not coincide with each other. 
Thus, the revolving speed of the engine is undesirably raised when the 
shift is completed, thus raising a possibility that the shift shock takes 
place. 
The above-mentioned problem also arises when the foregoing learning control 
of the hydraulic pressure is performed. That is, the learned value of the 
hydraulic pressure includes the torque applied to the automatic 
transmission when the previous shift has been performed. Therefore, if the 
hydraulic pressure for the synchronizing shift is controlled in accordance 
with the learned value of the hydraulic pressure when an ordinary shift 
has been performed, there arises a possibility that the shift shock takes 
place because the states of the input torque are considerably different 
from each other. 
The foregoing problems also arise when ordinary clutch-to-clutch shift 
except for the synchronizing shift is performed. If the learned value 
includes data obtained when the synchronizing shift has been performed, 
the input torque at a gear shift in which the learned value is obtained 
differs from the input torque at the gear shift which must be controlled, 
and the learned value to be used in the gear shift is inadequate. Thus, 
there arises a possibility that excessive shift shock takes place. 
When the foregoing control of the initial hydraulic pressure is performed 
such that the ordinary shift and the synchronizing shift are controlled in 
the same manner, the difference between the operation state of the engine 
when the shift is performed and the input to the automatic transmission 
may cause an inadequate initial hydraulic pressure at the gear shift. 
Moreover, there arises a possibility that the following control of the 
hydraulic pressure during the shift is delayed. 
An object of the present invention is to prevent shift shock in so-called 
synchronizing. 
Another object of the present invention is to provide a control unit which 
is capable of properly controlling learning of the hydraulic pressure in a 
clutch-to-clutch shift. 
Another object of the present invention is to provide a control unit which 
is capable of adequately controlling the initial hydraulic pressure at a 
clutch-to-clutch shift or at a shift carried out by directly controlling 
the hydraulic pressure. 
DISCLOSURE OF THE INVENTION 
A control unit according to the present invention is arranged such that the 
contents of control of the hydraulic pressures of frictional engagement 
units are made to be different from those employed when an ordinary shift 
down operation is performed in a case where the shift down, in which a 
first frictional engagement unit is engaged and a second means is 
released, is shift of a type in which the revolving speed of a power 
source is temporarily raised to about the synchronized revolving speed 
after the shift has been performed, that is, in a case where synchronizing 
shift is performed. In case that the hydraulic pressure for the first 
frictional engagement unit is temporarily raised immediately after the 
shifted output has been performed, the hydraulic pressure is made to be 
higher than that for the ordinary shift or the time for which the raised 
pressure is maintained is elongated. Therefore, delay of the shift and 
deterioration in the durability occurring attributable to sliding of the 
frictional engagement unit can be prevented. 
When the engaging pressure for the first frictional engagement unit is 
gradually raised to perform the shift in case that shift down is performed 
such that the revolving speed of the power source is temporarily raised, 
the engaging pressure is early raised as compared with another shift down 
operation or the raising ratio is raised. Therefore, even if so-called 
synchronizing shift has been performed and thus the revolving speed of the 
power source has been raised, the hydraulic pressures for the frictional 
engagement units are made to be suitable to the input torque. As a result, 
shift shock and deterioration in the durability of the frictional 
engagement units can be prevented. 
When the shift down is performed by the clutch-to-clutch shift method in 
which the revolving speed of the power source is temporarily raised, the 
present invention is arranged such that the learning control of the 
hydraulic pressures for the frictional engagement units for performing the 
shift is performed in a manner different from that for another shift down 
operation. Therefore, the controlled value obtained from the learning 
control can be made to be adaptable to the input torque. As a result, 
shift shock and deterioration in the durability of the frictional 
engagement units can be prevented.

BEST MODE FOR CARRYING OUT THE INVENTION 
The present invention will now be described further specifically with 
reference to the drawings. Initially, the overall control system will now 
be described. FIG. 11 shows the control system showing an engine 1 and an 
automatic transmission 3. A signal corresponding to the depression of an 
acceleration pedal 20 is supplied to an engine electronic control unit 21. 
A suction duct of the engine 1 is provided with an electronic throttle 
valve 23 which is operated by a throttle actuator 22. In accordance with 
the depression of the acceleration pedal 20, a control signal is output 
from the engine electronic control unit 21 to the throttle actuator 22 so 
that the degree of opening is controlled in accordance with the controlled 
variable. 
There are disposed an engine revolving speed sensor 24 for detecting the 
revolving speed of the engine 1, an air flow meter 25 for detecting the 
quantity of inlet air, an inlet-air temperature sensor 26 for detecting 
the temperature of the inlet air, a throttle sensor 27 for detecting 
opening degree .theta. of the electronic throttle valve 23, a vehicle 
speed sensor 28 for detecting vehicle velocity V in accordance with the 
revolving speed of an output shaft 17 or the like, a cooling-water 
temperature sensor 29 for detecting the temperature of cooling water for 
the engine 1, a brake switch 30 for detecting the operation of a brake and 
an operation position sensor 32 for detecting the operated position of a 
shift lever 31. Signals representing engine revolving speed Ne, inlet-air 
temperature Tha, opening .theta. of the electronic throttle valve 23, 
vehicle velocity V, engine cooling water temperature THw, operation state 
BK of the brake and operated position Psh of the shift lever 31 are 
supplied from the foregoing sensors to the engine electronic control unit 
21 or the transmission electronic control unit 33. Note that the 
transmission electronic control unit 33 is supplied with signals 
representing the opening .theta. of the electronic throttle valve 23 and 
engine cooling water temperature THw and a signal representing the 
operated position Psh of the shift lever 31. 
Moreover, a signal representing turbine revolving speed N.sub.T is supplied 
from a turbine revolving speed sensor 34 for detecting the revolving speed 
of a turbine runner to the transmission electronic control unit 33. A 
signal representing a kick-down operation is supplied from a kick down 
switch 35 for detecting the operation of the acceleration pedal 20 to the 
maximum operation position to the transmission electronic control unit 33. 
Moreover, a sports-mode switch 39 which is manually operated to output a 
gear shift signal and a synchronizing shift switch 40 are connected to the 
transmission electronic control unit 33. An example of an apparatus of the 
foregoing has been disclosed in Japanese Patent Laid-Open No. 6-307527 and 
Japanese Patent Application No. 7-215892. 
The sports-mode switch is a switch for selecting a mode for shifting the 
transmission by a manual operation or a switch for outputting a 
transmission signal generated in the manual operation, the sports-mode 
switch being provided for a shift apparatus or an instrument panel (not 
shown). Structure of the foregoing type have been disclosed in, for 
example, Japanese Patent Laid-Open No. 6-307527, Japanese Patent Laid-Open 
No. 6-48216 and Japanese Patent Laid-Open No. 6-2761. The synchronizing 
shift switch is a switch for shifting down the transmission by one step, 
synchronizing shift switch being disposed at an arbitrary position, for 
example, in a central portion of a steering wheel (not shown). When the 
transmission is shifted down by operating the foregoing switches, a 
so-called synchronizing shift control is performed such that the 
electronic throttle valve 23 is opened by a degree greater than the 
depression of the acceleration pedal 20 in response to an output signal 
from the engine electronic control unit 21 and the engine revolving speed 
Ne is raised to the synchronized revolving speed for the shifted stage 
after the transmission has been shifted down. An example of the 
above-mentioned control has been disclosed in, for example, Japanese 
Patent Application No. 7-215892. 
The engine electronic control unit 21 is a so-called microcomputer having a 
central processing unit (CPU), a storage unit (RAM and ROM) and an 
input/output interface. The CPU uses a supplied signal in accordance with 
a program previously stored in the ROM while using a temporal storage 
function of the RAM to perform various engine control operations. For 
example, the CPU controls a fuel injection valve 36 for controlling the 
quantity of fuel to be injected, an igniter 37 for controlling the 
ignition timing, a bypass valve (not shown) for controlling the idling 
speed and all of throttle controls including the traction control by 
causing the throttle actuator 22 to control the electronic throttle valve 
23. 
Also the transmission electronic control unit 33 is a microcomputer 
similarly to the foregoing engine electronic control unit 21. The CPU 
processes a supplied signal in accordance with a program previously stored 
in the ROM by using the temporary storage function of the RAM. Moreover, 
the CPU operates solenoid valves or linear solenoid valves in a 
hydraulic-pressure control circuit 38. For example, the transmission 
electronic control unit 33 controls a linear solenoid valve S.sub.LT for 
generating output pressure PS.sub.LT corresponding to the opening of the 
electronic throttle valve 23, a linear solenoid valve S.sub.LN for 
controlling the back pressure for accumulators and the quantity of slip of 
the lock-up clutch. Moreover, the CPU operates a linear solenoid valve 
S.sub.LU for controlling the engaged pressure for a predetermined clutch 
or a brake during shift of the transmission in accordance with the 
progress of the shift of the transmission and to correspond to an applied 
torque. 
The transmission electronic control unit 33 judges the gear stage of the 
automatic transmission 3 and an engagement state of the lock-up clutch in 
accordance with the standard throttle opening .theta. (opening of the 
throttle obtained by converting the depression of the accelerator pedal 
with a predetermined non-linear characteristic), the vehicle velocity V 
and a shift map using the foregoing factors as parameters. To realize the 
judged gear stage and the state of engagement, the transmission electronic 
control unit 33 operates No. 1 to No. 3 solenoid valves S.sub.OL1, 
S.sub.OL2 and S.sub.OL3 of the hydraulic-pressure control circuit 38. When 
engine brake is generated, the transmission electronic control unit 33 
operates a No. 4 solenoid valve S.sub.OL4. 
The automatic transmission 3 according to this embodiment is structured to 
be capable of setting five forward and one reverse gear stages, as shown 
in a skeleton diagram shown in FIG. 12. That is, as shown in FIG. 12, the 
automatic transmission 3 is connected to the engine 1 through a torque 
converter 2. The torque converter 2 has a pump impeller 5 connected to a 
crank shaft 4, a turbine runner 7 connected to an input shaft 6 of the 
automatic transmission 3, a lock-up clutch 8 for establishing the direct 
connection between the pump impeller 5 and the turbine runner 7 and a 
stator 10, the one directional rotation of which is inhibited by a one-way 
clutch 9. 
The automatic transmission 3 has a sub-transmission section 11 for 
selecting a high gear stage or a low gear stage and a main-transmission 
section 12 which is capable of selecting gear stage from the reverse gear 
stage and the four forward gear stages. The sub-transmission section 11 
has; a planetary gear unit 13 which is composed of a sun gear S.sub.0, a 
ring gear R.sub.0 and a pinion P0 rotatively supported by a carrier 
K.sub.0 and meshed with the sun gear S.sub.0 and the ring gear R.sub.0 ; a 
clutch C.sub.0 and a one-way clutch F.sub.0 disposed between the sun gear 
S.sub.0 and the carrier K.sub.0 ; and a brake B.sub.0 disposed between the 
sun gear S.sub.0 and a housing 19. 
The main-transmission section 12 has; a first planetary gear unit 14 
composed of a sun gear S.sub.1, a ring gear R.sub.1 and a pinion P.sub.1 
rotatively supported by a carrier K.sub.1 and meshed with the sun gear 
S.sub.1 and the ring gear R.sub.1 ; a second planetary gear unit 15 
composed of a sun gear S.sub.2, a ring gear R.sub.2 and a pinion P.sub.2 
rotatively supported by a carrier K.sub.2 and meshed with the sun gear 
S.sub.2 and the ring gear R.sub.2 ; and a third planetary gear unit 16 
composed of a sun gear S.sub.3, a ring gear R.sub.3 and a pinion P.sub.3 
rotatively supported by a carrier K.sub.3 and meshed with the sun gear 
S.sub.3 and the ring gear R.sub.3. 
The sun gear S.sub.1 and the sun gear S.sub.2 are integrally connected to 
each other, the ring gear R.sub.1, the carrier K.sub.2 and the carrier 
K.sub.3 are integrally connected to one another. The carrier K.sub.3 is 
connected to the output shaft 17. The ring gear R.sub.2 is integrally 
connected to the sun gear S.sub.3. A first clutch C.sub.1 is disposed 
among the ring gear R.sub.2, the sun gear S.sub.3 and an intermediate 
shaft 18. A second clutch C.sub.2 is disposed among the sun gear S.sub.1, 
the sun gear S2 and the intermediate shaft 18. 
As brake means, a band type first brake B.sub.1 for stopping rotations of 
the sun gear S.sub.1 and the sun gear S.sub.2 is provided for the housing 
19. A first one-way clutch F.sub.1 and a second brake B.sub.2 are in 
series disposed among the sun gear S.sub.1, the sun gear S.sub.2 and the 
housing 19. The first one-way clutch F.sub.1 is arranged to be engaged 
when the sun gear S.sub.1 and the sun gear S.sub.2 are inversely rotated 
opposite to the direction of rotation of the input shaft 6. 
A third brake B.sub.3 is disposed between the carrier K.sub.1 and the 
housing 19. A fourth brake B.sub.4 and a second one-way clutch F.sub.2 are 
in parallel disposed between the ring gear R.sub.3 and the housing 19. The 
second one-way clutch F.sub.2 is arranged to be engaged when the ring gear 
R.sub.3 is rotated inversely. The clutches C.sub.0, C.sub.1 and C.sub.2 
and the brakes B.sub.0, B.sub.1, B.sub.2, B.sub.3 and B.sub.4 are 
hydraulic frictional engagement units having frictional elements which are 
engaged when hydraulic pressure is applied. 
The foregoing automatic transmission is able to set any one of five forward 
and reverse gear stages. The states of engagements and releases of each 
frictional engagement units for setting the gear stage are shown in an 
engagement operation table shown in FIG. 13. Referring to FIG. 13, mark 
.largecircle. indicates an engaged state and mark X indicates a released 
state. 
FIG. 14 shows the operated positions of the shift lever 31. Referring to 
FIG. 14, the shift lever 31 is supported by a supporting unit (not shown) 
which is capable of shifting the shift lever 31 to eight positions 
combining six positions in the longitudinal direction of the vehicle and 
two positions in the lateral direction of the vehicle. Letter P represents 
a parking range position, letter R represents a reverse range position, 
letter N represents a neutral range position, letter D represents a drive 
range position, numeral "4" represents a "4" range position for setting 
gear stages to the fourth speed, numeral "3" represents a "3" range 
position for setting gear stages to the third speed, numeral "2" 
represents a "2" range position to the second speed and letter L 
represents a low range position for inhibiting up-shift higher than the 
first speed. Note that the sports-mode switch 39 is disposed between the 
"2" range position and the low range position at a position more rearwards 
of the vehicle. 
As shown in FIG. 13, the above-mentioned automatic transmission 3 is 
arranged to perform the clutch-to-clutch shift between the second speed 
and the third speed so that both of the engaged states of the third brake 
B.sub.3 and the second brake B.sub.2 are switched. This shift must be 
controlled in such a manner that the frictional engagement units 
concerning the shift are brought to an underlap state or an overlap state 
in accordance with the power on/off state or the shift up/down state. 
Specifically, the hydraulic pressure for the second brake B.sub.2 must be 
controlled in accordance with the applied torque and the hydraulic 
pressure for the third brake B.sub.3 must be controlled in accordance with 
the progress of the shift. Accordingly, the hydraulic-pressure control 
circuit 38 includes a circuit shown in FIG. 15 in order to smoothly and 
quickly perform the shift. The structure of the circuit will be briefly 
described. 
Referring to FIG. 15, reference numeral 70 represents a 1-2 shift valve, 
reference numeral 71 represents a 2-3 shift valve and reference numeral 72 
represents a 3-4 shift valve. The states of communications of ports in 
each shift valves 70, 71 and 72 are as indicated below the shift valves 
70, 71 and 72. Note that the accompanying numerals represent the gear 
stages. Among the ports of the 2-3 shift valve 71, the third brake B.sub.3 
is, through an oil passage 75, connected to a brake port 74 which is 
communicated with an input port 73 at the first speed and the second 
speed. An orifice 76 is interposed in the oil passage. A damper valve 77 
is connected between the orifice 76 and the third brake B.sub.3. The 
damper valve 77 serves as a buffer by sucking a small quantity of oil 
pressure in case that a line pressure has been rapidly applied to the 
third brake B.sub.3. 
Reference numeral 78 represents a B-3 control valve. The B-3 control valve 
78 directly controls the engaging pressure for the third brake B.sub.3. 
The above-mentioned hydraulic pressure control is called as a direct 
pressure control. That is, the B-3 control valve 78 has a spool 79, a 
plunger 80 and a spring 81 disposed between the spool 79 and the plunger 
80. The oil passage 75 is connected to an input port 82 which is 
opened/closed by the spool 79. An output port 83 selectively communicated 
with the input port 82 is connected to the third brake B.sub.3. The output 
port 83 is connected to a feed-back port 84 formed at the leading end of 
the spool 79. On the other hand, a port 86 among the ports of the 2-3 
shift valve 71 which outputs a D-range pressure when the gear stage is not 
lower than the third speed is communicated through an oil passage 87 with 
a port 85 opened at a position at which the spring 81 is disposed. A 
linear solenoid valve S.sub.LU for the lock-up clutch is connected to a 
control port 88 formed in an end portion of the plunger 80. 
Therefore, the pressure level which is controlled by the B-3 control valve 
78 is set in accordance with the elastic force of the spring 81 and the 
hydraulic pressure applied to the port 85. The elastic force by the spring 
81 is enlarged in proportion to the level of the signal pressure which is 
supplied to the control port 88. 
Referring to FIG. 15, reference numeral 89 represents a 2-3 timing valve. 
The 2-3 timing valve 89 has a spool 90 having a small land and two lands 
each having a large diameter, a first plunger 91, a spring 92 disposed 
between the spool 90 and the first plunger 91, and a second plunger 93 
disposed opposite to the first plunger 91 in such a manner that the spool 
90 is interposed. An oil passage 95 is connected to a port 94 formed in 
the intermediate portion of the 2-3 timing valve 89. The oil passage 95 is 
connected to a port 96 among the ports of the 2-3 shift valve 71 which is 
communicated with the brake port 74 when the gear stage is not lower than 
the third speed. 
The oil passage 95 is branched at an intermediate position so as to be, and 
connected through an orifice to the port 97 opened between the 
small-diameter land and the large-diameter land. A port 98 selectively 
communicated with the port 94 formed at the intermediate position is 
connected to a solenoid relay valve 100 through an oil passage 99. The 
linear solenoid valve S.sub.LU for the lock-up clutch is connected to a 
port opened at an end of the first plunger 91. The second brake B.sub.2 is 
connected through an orifice to a port opened at an end of the second 
plunger 93. 
The oil passage 87 supplies/discharges hydraulic pressure to and from the 
second brake B.sub.2. A small-diameter orifice 101 and an orifice 102 
having check ball are interposed at an intermediate position of the oil 
passage 87. A large-diameter orifice 104 having a check ball which is 
opened when pressure is discharged from the second brake B.sub.2 is 
interposed in an oil passage 103 branched from the oil passage 87. The oil 
passage 103 is connected to an orifice control valve 105 to be described 
below. 
The orifice control valve 105 is a valve for controlling pressure discharge 
rate from the second brake B.sub.2. The second brake B.sub.2 is connected 
to a port 107 arranged to be opened/closed by a spool 106 of the orifice 
control valve 105. The oil passage 103 is connected to a port 108 formed 
below the port 107 when viewed in FIG. 15. A port 109 formed above the 
port 107 to which the second brake B2 is connected when viewed in FIG. 15 
is a port selectively communicated with a drain port. A port 111 of the 
B-3 control valve 78 is connected through an oil passage 110 to the port 
109. Note that the port 111 is a port which is selectively communicated 
with the output port 83 to which the third brake B.sub.3 is connected. 
A control port 112 among the ports of the orifice control valve 105 which 
is formed at an end opposite to the spring for pressing the spool 106 is 
connected through an oil passage 113 to a port 114 of the 3-4 shift valve 
72. The port 114 is a port which outputs the signal pressure of the third 
solenoid valve S.sub.OL3 when the gear stage is not higher than the third 
speed and which outputs the signal pressure of the fourth solenoid valve 
S.sub.OL4. An oil passage 115 branched from the oil passage 95 is 
connected to the orifice control valve 105 so that the oil passage 115 is 
selectively communicated with the drain port. 
A port 116 of ports of the 2-3 shift valve 71 which output the D-range 
pressure when the gear stage is not higher than the second speed is 
connected through an oil passage 118 to a port 117 of the 2-3 timing valve 
89 formed at a position at which the spring 92 is disposed. A port 119 of 
ports of the 3-4 shift valve 72 which is communicated to the oil passage 
87 when the gear stage is not higher than the third speed is connected 
through an oil passage 120 to the solenoid relay valve 100. 
Referring to FIG. 15, reference numeral 121 represents an accumulator for 
the second brake B.sub.2. A back pressure chamber of the accumulator 121 
is supplied with an accumulator control pressure regulated in accordance 
with the hydraulic pressure output from the linear solenoid valve 
S.sub.LN. Note that the accumulator control pressure is controlled in 
accordance with the applied torque to be raised in inverse proportion to 
the output pressure from the linear solenoid valve S.sub.LN. Therefore, 
the transitional hydraulic pressure for engaging/releasing the second 
brake B.sub.2 is shifted at higher levels in inverse proportion to the 
signal pressures of the linear solenoid valve S.sub.LN. By temporarily 
lowering the signal pressure of the linear solenoid valve S.sub.LN, the 
engaging pressure for the second brake B.sub.2 can temporarily be raised. 
Reference numeral 122 represents a C-0 exhaust valve 122, and reference 
numeral 123 represents an accumulator for clutch C.sub.0. The C-0 exhaust 
valve 122 acts to engage the clutch C.sub.0 to effect the engine braking 
at only the second speed in the second speed range. 
Therefore, in the above-mentioned hydraulic-pressure circuit, when the port 
111 of the B-3 control valve 78 is communicated with the drain, the 
engaging pressure for the third brake B.sub.3 can directly be regulated by 
the B-3 control valve 78. Moreover, the regulation level can be changed by 
the linear solenoid valve S.sub.LU. If the spool 106 of the orifice 
control valve 105 is positioned in the left-hand half portion of the 
drawing, the second brake B.sub.2 is communicated with the oil passage 103 
through the orifice control valve 105. Therefore, the pressure can be 
discharged through the large-diameter orifice 104. Thus, the draining rate 
from the second brake B.sub.2 can be controlled. 
The shift between the second speed and the third speed of the automatic 
transmission 3 is performed by the clutch-to-clutch shift such that both 
of the engaging/releasing states of the second brake B.sub.2 and the third 
brake B.sub.3 are simultaneously changed. When the third speed is shifted 
down to the second speed, the second brake B.sub.2 engaged at the third 
speed is gradually released in accordance with the input revolving speed 
so that the rotation is changed. When the input revolving speed is changed 
to the synchronized revolving speed of the second speed, the engaging 
pressure for the third brake B.sub.3 is rapidly raised at the moment at 
which the revolving speed reaches a predetermined revolving speed so that 
the second speed is realized. 
As described above, the third brake B.sub.3 must be immediately engaged in 
response to the change in the rotation occurring attributable to the 
progress of the shift. On the other hand, each of general frictional 
engagement units including the third brake B.sub.3 has slight clearance 
between the friction plates and between the friction plate and the piston 
of a hydraulic servo mechanism. Therefore, any torque capacity is not 
provided so far as the clearance exists. Accordingly, hydraulic pressure 
is rapidly supplied to the engaging side frictional engagement units 
simultaneously or immediately after the shifted output of the 
clutch-to-clutch to realize a state immediately before the engagement in 
which the torque capacity is substantially zero. That is, initial 
hydraulic pressure control is performed. Since the state in which the 
engagement is immediately established attributable to the further rise in 
the hydraulic pressure is different depending upon the output from the 
engine or the revolving speed applied to the automatic transmission, the 
above-mentioned control unit must perform the following control. 
The engaging pressure for each frictional engagement unit of the automatic 
transmission 3 is judged by the line pressure which is controlled in 
accordance with the throttle opening .theta. of the engine 1. For example, 
engaging pressure P.sub.B3 for the third brake B.sub.3 when the 
clutch-to-clutch is performed between the second speed and the third speed 
is controlled in accordance with the progress of the shift. For example, 
the third speed by engaging the third brake B.sub.3. The engaging pressure 
P.sub.B3 for the third brake B.sub.3 is quickly raised in order to raise 
the engine revolving speed toward the synchronized revolving speed for the 
second speed in a power-off state in which the electronic throttle valve 
23 is closed when the shift is judged. In a power-on state in which the 
electronic throttle valve 23 is opened, the engaging pressure is 
maintained at a low pressure, and then above-mentioned control, rapid 
change in the engine revolving speed is prevented at a moment before and 
after the revolving speed reaches the synchronized revolving speed for the 
second speed. Thus, shift shock is prevented. 
However, when the gear stage is shifted down by manually operating the 
sports-mode switch 39 or the synchronizing shift switch 40, the gear stage 
is shifted down in a state where the engine revolving speed of the engine 
1 to which the automatic transmission 3 is connected has been raised to 
about the synchronized revolving speed for the gear stage after the shift 
down if the power-off state has been realized in which the electronic 
throttle valve 23 is closed when the shift is judged. Since the 
above-mentioned state in the operation is different from either the 
power-on state or the power-off state when the gear stage is shifted down 
from the third speed to the second speed, the following control is 
performed. 
FIG. 1 is a flow chart showing shift down from the third speed to the 
second speed in three cases. An input signal is processed (step 1), and 
then the so-called clutch-to-clutch shift is performed such that the third 
speed is shifted down to the second speed (step 2). Therefore, step 2 
corresponds to the shift determining means according to the present 
invention. If a negative determination is performed in step 2, any control 
is not performed. The operation is returned. If an affirmative 
determination is performed, whether or not the present mode is the sports 
mode is judged (step 3). Steps 2 and 3 correspond to the shift detecting 
means. 
As described above, the sports mode is a shift mode in which the shift is 
performed in accordance with the operation of the switch. The switch is 
structured in such a manner that each gear stage position is provided for 
a shifting apparatus and a switch, which is switched on by a shift level, 
is provided for each gear stage position. Another structure is formed such 
that a sports mode state is set and, in this state, an up-shift switch or 
a down-shift switch is switched on by a shift lever. Another structure is 
formed such that an up/down switch is provided for a steering wheel or an 
instrument panel. Therefore, the judgment in step 3 may be performed by 
judging whether or not an output has been made from the foregoing switch. 
When a negative judgment is performed in step 3 because the shift down has 
taken place due to change in the running condition, whether or not the 
state is the power-on state is judged (step 4). That is, whether or not 
the electronic throttle valve 23 has been opened and the vehicle is being 
operated by the output from the engine 1 is judged. This judgment can be 
performed in accordance with the opening .theta. of the throttle. 
When an affirmative judgment is performed in step 4 because of the power-on 
state, shift is performed by controlling release of the second brake 
B.sub.2 and by controlling the engagement of the third brake B.sub.3 (step 
5). The foregoing shift is performed because the 2-3 shift valve 71 shown 
in FIG. 15 is switched, the linear solenoid valve S.sub.LU regulates the 
engaging pressure for the third brake B.sub.3 and thus the pressure is 
discharged from the second brake B.sub.2 in a state where the accumulator 
121. Since the engine 1 is in an operation state in this case, the engine 
revolving speed, that is, input revolving speed N.sub.CO is raised because 
the engaging pressure for the second brake B.sub.2 state to the third 
speed is lowered. Therefore, the engaging pressure B.sub.P3 for the third 
brake B.sub.3 for realizing the third speed is controlled in accordance 
with pattern I shown in FIG. 2 (step 6). 
The control pattern I will be briefly described. The duty ratio of the 
linear solenoid valve S.sub.LU for determining the regulation level for 
the third brake pressure P.sub.B3 is raised at time t.sub.2 elapsed, by 
predetermined time T.sub.1, from time t.sub.1, at which the determination 
of the shift from the third speed to the second speed has been established 
so that the initial hydraulic pressure control is performed to reduce the 
pack clearance. That is, the duty ratio is set to be D.sub.1 and this 
value is maintained for T.sub.2 seconds. Then, the duty ratio is 
maintained at small value D.sub.2 to time t.sub.4 at which the input 
revolving speed N.sub.CO is raised to the revolving speed which is lower 
than the synchronized revolving speed for the second speed by 
predetermined revolving speed .DELTA..alpha.. Thus, the third brake 
pressure P.sub.B3 is maintained at a low pressure. Then, the duty ratio of 
the linear solenoid valve S.sub.LU is gradually raised so that the third 
brake pressure P.sub.B3 is gradually raised (swept up). At time t.sub.5 at 
which the input revolving speed N.sub.CO has reached the synchronized 
revolving speed, the third brake B.sub.3 is completely engaged. 
As a result of the above-mentioned control, the third brake B.sub.3 
substantially starts engaging when the engine revolving speed Ne has been 
raised to the predetermined revolving speed. Thus, rapid change in the 
engine revolving speed Ne before and after the moment at which it reaches 
the synchronized revolving speed for the second speed after the shift down 
can be prevented. Thus, the shift shock can satisfactorily be prevented. 
When the gear stage is shifted down in the power-on state, a control is 
performed such that the engine torque is reduced by delaying the ignition 
timing or by temporarily raising the releasing pressure for the second 
brake B.sub.2 on the releasing side. 
If a negative judgment is performed in step 4 because of the power-off 
state, the gear stage is shifted down by releasing the second brake 
B.sub.2 and by engaging the third brake B.sub.3 in the power-off state 
(step 7). Moreover, the engaging pressure for the third brake B.sub.3 is 
controlled in accordance with pattern II shown in FIG. 3 (step 8). 
That is, the initial hydraulic pressure control is performed such that a 
state in which the duty ratio of the linear solenoid valve S.sub.LU is set 
to D.sub.1 is maintained for T.sub.2 seconds. After the end time t.sub.3, 
the duty ratio of the linear solenoid valve S.sub.LU is gradually raised 
at time t.sub.6 at which the input revolving speed N.sub.CO is 
considerably lower than the synchronized revolving speed for the second 
speed so that the engaging pressure P.sub.B3 for the third brake B.sub.3 
is gradually raised. After time t.sub.7 at which the third brake pressure 
P.sub.B3 has been sufficiently raised, the input revolving speed N.sub.CO 
reaches the synchronized revolving speed for the second speed. Therefore, 
when the gear stage is shifted down from the third speed to the second 
speed in the power-off state, the engaging pressure P.sub.B3 for the third 
brake B.sub.3 is early raised so that the revolving speed of the engine 1 
is raised by the torque applied from the output shaft to quickly perform 
the shift. Moreover, the input revolving speed N.sub.CO is smoothly 
changed to the synchronized revolving speed so that the shift shock is 
satisfactorily prevented. Since the third brake B.sub.3 is, in this case, 
maintained in the state in which it can immediately be accordance with the 
input revolving speed, undesirable shift shock does not take place. Since 
the state is the power-off state in this case, the delay control of the 
ignition timing and the control for temporarily raising the releasing 
pressure for the second brake B.sub.2 on the releasing side to reduce the 
output torque are not performed. 
If an affirmative judgment is performed in step 3 because of the shift down 
in the sports mode, synchronizing shift is performed (step 9). Therefore, 
step 3 corresponds to the synchronizing shift determining means according 
to the present invention. The synchronizing shift is shift control in 
which the engine revolving speed Ne is raised to the synchronized 
revolving speed in the realized gear stage after the shift down and the 
shift down is performed in this state. Therefore, the electronic throttle 
valve 23 is temporarily opened by the engine electronic control unit 21 to 
correspond to the release of the second brake B.sub.2 and the engagement 
of the third brake B.sub.3. 
Moreover, coordinated control of the opening .theta. of the throttle and 
the third brake pressure P.sub.B3 is performed (step 10). Then, the third 
brake pressure P.sub.B3 is controlled in accordance with pattern III shown 
in FIG. 4 (step 11). Note that steps 6, 8 and 11 correspond to the 
engagement control changing means according to the present invention. 
The coordinated control in step 10 is control in which the throttle opening 
.theta. and the third brake pressure P.sub.B3 are relatively changed in 
order to cause the engine revolving speed Ne to be changed smoothly with 
respect to the synchronized revolving speed for the second speed when the 
gear stage is shifted down to the second speed in a state where the 
opening .theta. of the throttle is enlarged. The control is performed in 
such a manner that either or both of the throttle opening .theta. and the 
third brake pressure P.sub.B3 are arbitrarily changed. Therefore, step 10 
corresponds to the coordinated control means according to the present 
invention. 
The specific example of the coordinated control will be now described. FIG. 
5 shows an example in which the throttle opening is changed in such a 
manner that the duty ratio of the linear solenoid valve S.sub.LU is raised 
in a stepped manner so that the engaging pressure P.sub.B3 is swept up. 
Simultaneously, the throttle opening .theta. is reduced so that the rate 
of the enlargement of the engine revolving speed Ne is gradually lowered 
as indicated by a continues line shown in FIG. 5 so as to be smoothly 
changed to the synchronized revolving speed for the second speed. As a 
result, torsion in the power transmission system and shock attributable to 
the torsion can be prevented. Note that change in the throttle opening 
.theta. is not required to be always performed simultaneously with 
sweeping up to the third brake pressure P.sub.B3. The change may be 
performed at arbitrary time before and after sweeping up. As described 
above, the change rate of the third brake pressure P.sub.B3 may be changed 
(lowered) by changing the step width of the duty ratio in addition to the 
change in the throttle opening .theta.. 
The pattern III will be briefly described. The duty ratio of the linear 
solenoid valve S.sub.LU in the initial hydraulic pressure control is set 
to be value D.sub.3 judged on the basis of value D.sub.1 in substantially 
an ordinary state and larger than the value D.sub.1 so that the hydraulic 
pressure which is supplied to the third brake B.sub.3 is raised. When the 
duty ratio D.sub.1 has been changed because of learning, also the duty 
ratio D.sub.3 may synchronously be changed at the same rate. Therefore, 
even if the input revolving speed is raised, the third brake B.sub.3 can 
quickly be set to a state immediately before the satisfactory engagement 
corresponding to the input revolving speed so as to be brought to a 
standby state in a low pressure state. Then, the duty ratio is lowered to 
predetermined value D.sub.2 so that the third brake pressure P.sub.B3 is 
brought to the standby state at a low pressure. Since the throttle opening 
has been enlarged, the duty ratio is raised in a stepped manner at time 
t.sub.8 at which the engine revolving speed Ne (the input revolving speed 
N.sub.CO) has reached revolving speed which is lower than the synchronized 
revolving speed for the second speed by predetermined revolving speed 
.DELTA..beta. (&gt;.DELTA..alpha.) so that the third brake pressure P.sub.B3 
is raised. The duty ratio is raised stepwise, the width of each of which 
is made to be larger than that in the power-on state. Therefore, the ratio 
of rise in the third brake pressure P.sub.B3 is raised as compared with 
that in the shift down in the power-on state. That is, step 10 corresponds 
to the initial hydraulic pressure control means according to the present 
invention. 
Therefore, even if the synchronizing shift is performed in which the engine 
revolving speed is raised during shift, the initial hydraulic pressure 
control is completed more quickly as compared with the ordinary state. 
Therefore, sweeping up of the third brake pressure P.sub.B3 is not 
delayed. As a result, even if sweeping up is early performed or even if a 
period in which standby at low pressure is not substantially performed, 
the shift control can satisfactorily be performed. 
The control which is performed in accordance with the pattern III is 
arranged in such a manner that the engaging pressure P.sub.B3 for the 
third brake B.sub.3 which is the on-coming frictional engagement unit is 
swept up earlier than the case of the power-on down shift in the case of 
the synchronizing shift in which the throttle opening is smaller than the 
power-on state or the sweeping up ratio is raised. Therefore, shift can 
quickly be performed and the input revolving speed N.sub.CO (the engine 
revolving speed) can smoothly be changed to the synchronized revolving 
speed for the second speed. As a result, shift shock can satisfactorily be 
prevented. 
Since the initial hydraulic pressure control is control for moving the 
friction plate of the frictional engagement unit to a state immediately 
before the engagement, time for which the initial hydraulic pressure is 
supplied may be elongated to correspond to the state in which the input 
revolving speed has been raised. An example of this case is indicated by a 
broken line shown in FIG. 4 in which a predetermined duty ratio D.sub.1 is 
maintained for T.sub.3 (&gt;T.sub.2) seconds. In this case, T.sub.3 is judged 
on the basis of T.sub.2. If T.sub.2 is changed due to learning or the 
like, T.sub.3 may be changed at a predetermined ratio corresponding to the 
change. 
When the hydraulic pressure for use in the initial hydraulic pressure 
control or the time for which the hydraulic pressure is maintained is 
changed to be adaptable to the synchronizing shift, the duty ratio D.sub.3 
and time T.sub.3 may be judged as shown in FIGS. 6 to 8. That is, the 
foregoing values are set to be the functions of the change ratio (Ne dots) 
of the engine revolving speed Ne and their coefficients k1 and k2 and 
constants a and b are corrected in accordance with the values learned 
(learning is performed after the shift has been performed) in the state of 
standby at low pressure. FIG. 8 shows general tendency of the constants a 
and b with respect to the learned value of the standby pressure at low 
pressure. That is, the initial hydraulic pressure control time is 
elongated or the hydraulic pressure is raised in proportion to the change 
ratio of the engine revolving speed. 
Since the correction of the initial hydraulic pressure control is performed 
when the synchronizing shift is performed, judgment step (step 3') for 
judging whether or not the shift is the shift down performed by the 
synchronizing shift switch may be substituted for step 3' shown in FIG. 1. 
When the gear stage is shifted down from the third speed to the second 
speed, the second brake pressure P.sub.B2 and third brake pressure 
P.sub.B3 are controlled in accordance with the learned values. That is, 
the third speed is set such that the second brake B.sub.2 is engaged. In 
response to the shift signal, discharge of pressure is started from the 
second brake B.sub.2. The second brake pressure P.sub.B2 is 
feedback-controlled in accordance with the engine revolving speed Ne when 
the back pressure for the accumulator 121 is controlled by the linear 
solenoid valve S.sub.LN. The control is continued from time t.sub.11 at 
which shifted output is performed to time t.sub.12 at which the engine 
revolving speed Ne is raised to the synchronized revolving speed for the 
second speed. Then learning control is performed in which the controlled 
value of the back pressure for the accumulator 121 with respect to the 
input torque at the control is judged in accordance with the variation in 
the feedback quantity. That is, when the gear stage is next shifted down 
from the third speed to the second speed with the input torque, the 
learned controlled value is employed to control the back pressure for the 
accumulator, that is, the controlled value for the releasing pressure 
P.sub.B2 for the second brake B.sub.2. Note that control of the foregoing 
type has been disclosed in, for example, Japanese Patent Laid-Open No. 
1-150050 and Japanese Patent Laid-Open No. 63-291738. 
On the other hand, initial hydraulic pressure control (quick up) of the 
third brake B.sub.3 is performed at time t.sub.11 to reduce the pack 
clearance. Thus, the third brake B.sub.3 is maintained at a low pressure 
(brought to a standby state) from time t.sub.13 at which the initial 
hydraulic pressure is ended to time t.sub.14 at which the engine revolving 
speed Ne reaches a predetermined revolving speed. Then, the duty ratio of 
the linear solenoid valve S.sub.LU is raised step-wise so that the 
engaging pressure P.sub.B3 for the third brake B.sub.3 is swept up and 
first applied to time t.sub.12 at which the engine revolving speed Ne has 
reached the synchronized revolving speed. Thus, the engaging pressure is 
rapidly raised. The low standby pressure during the foregoing control is 
learning-controlled. Control of the foregoing type has been disclosed in, 
for example, Japanese Patent Laid-Open No. 6-331016. 
The engine revolving speed Ne is, when the gear stage is shifted, changed 
in accordance with the engine torque and the engaging forces of the brakes 
B.sub.2 and B.sub.3. Therefore, the learned value is obtained for each 
throttle opening . However, if the above-mentioned synchronizing shift is 
performed, the electronic throttle valve 23 is opened when the shift has 
been started and thus the output from the engine is enlarged. Therefore, 
the engine torque is made to be different from that in the case of the 
shift down in which the control of the synchronizing shift is not 
performed. Thus, the previous learned value is not suitable. Accordingly, 
the control unit according to the present invention performs the learning 
control as follows. 
FIG. 10 is a flow chart of the shift down operation from the third speed to 
the second speed in which the operation is classified into three states. 
After an input signal is processed (step 20), shift down from the third 
speed to the second speed, which is a so-called clutch-to-clutch shift, is 
judged (step 21). If a negative judgment is performed in step 21, any 
special control is not performed and the operation is returned. If an 
affirmative judgment is performed, whether or not the mode is the sports 
mode is judged (step 22). 
If a negative judgment is performed in step 22 because of the shift down 
which is performed because the running condition has been changed, whether 
or not the state is the power-on state is judged (step 23). If an 
affirmative judgment is performed because of the power-on state, shift is, 
as described above, performed by controlling the release of the second 
brake B2 and by controlling the engagement of the third brake B3 (step 
24). 
Foregoing steps 20 to 24 are the same as steps 1 to 5 shown in FIG. 1. 
If the shift down is performed in the power-on state, control of the linear 
solenoid valve S.sub.LU and control of the back pressure for the 
accumulator 121 by the linear solenoid valve S.sub.LN are performed in 
accordance with the learned value as described above. During the shift, 
learning of the hydraulic pressure for each of the brakes B.sub.2 and 
B.sub.3 is performed and the learned value is stored (step 25). In this 
case, the learned values are stored in the form of controlled values or in 
the form of, for example, a map composed of corrected values of the 
controlled values. The learned values are stored and used as learned 
values for the shift down in the power-on state and for the shift mode 
which is not the synchronizing shift. 
If a negative judgment is performed in step 23 because of the power-off 
state, shift down in the power-off state is performed (step 26). Also in 
this case, the second brake pressure P.sub.B2 is controlled by 
feedback-controlling the back pressure for the accumulator 121 in 
accordance with the characteristic of the accumulator 121 so as to be 
changed as shown in FIG. 9. Step 26 above is the same as step 7 shown in 
FIG. 1. 
As described above, the control of the linear solenoid valve S.sub.LU and 
the control of the back pressure for the accumulator 121 by the linear 
solenoid valve SLN are performed in accordance with the learned values. 
During the shift, the hydraulic pressure for each of the brakes B.sub.2 
and B.sub.3 is learned so as to be stored as learned value (step 27). In 
this case, the learned values are formed in the form of controlled values 
or in the form of a map composed of corrected values of the controlled 
values. The learned values are stored and used as learned values for the 
shift down in the power-off state and for a shift mode which is not the 
synchronizing shift. That is, the learning control is performed 
individually from that in the power-on state and the synchronizing shift 
to be described later. 
If an affirmative judgment is performed in step 22 because of the shift 
down in the sports mode, the synchronizing shift is performed (step 28). 
Control in step 28 is the same as step 9 shown in FIG. 1. 
The hydraulic pressure for each of the brakes B.sub.2 ad B.sub.3 during the 
synchronizing shift is learning-controlled (step 29). In this case, the 
learned values are stored as controlled values or in the form of, for 
example, a map composed of corrected values of the controlled values. 
Since the shift is the synchronizing shift in which the engine revolving 
speed Ne is raised to the synchronized revolving speed, the learned values 
are stored and used individually from the shift in the power-on state or 
the power-off state. The learning control is performed individually from 
that in the power-on state and the power-off state. That is, step 29 
corresponds to the learning control changing means according to the 
present invention. 
Since the synchronizing shift is not limited to the shift down in the 
sports mode and is performed by switching the the synchronizing shift 
switch on, the judgment step (step 22') for judging whether or not the 
shift down is the shift down in the synchronizing shift may be substituted 
for step 22 shown in FIG. 10. 
Although the foregoing embodiment is structured such that the learning 
control of the hydraulic pressure when the synchronizing shift is 
performed is performed individually from the learning control of the 
hydraulic pressure in the power-on state or the power-off state, the 
present invention may be structured such that the learning control of the 
hydraulic pressure in the synchronizing shift is further changed in 
accordance with the degree of rise in the revolving speed of the power 
source, for example, the engine. In this case, a structure may be formed 
such that the change rate (rise rate) is detected from a detected value of 
the revolving speed of the power source after the determination of the 
synchronizing shift has been performed. Then, the learning control of the 
hydraulic pressure is individually performed for each of results of the 
detection. Specifically, the structure is formed into a control unit for 
an automatic transmission characterized in that the learning control of 
the hydraulic pressure is changed to correspond to the degree of rise in 
the revolving speed in a case of the shift in which the revolving speed of 
the power source is temporarily raised. Although the description of the 
above-mentioned embodiment has been performed about the shift down from 
the third speed to the second speed, the present invention is not limited 
to the above-mentioned embodiment. The present invention may be applied to 
an apparatus for controlling shift down to another gear stage or an 
apparatus for directly controlling the hydraulic pressure for the 
frictional engagement unit by a linear solenoid valve or the like. 
Therefore, the frictional engagement unit, the engaging pressure of which 
including the initial hydraulic pressure must be controlled, may be a 
frictional engagement unit except for the second and third brakes. The 
present invention is characterized by controlled contents peculiar to the 
synchronizing shift. Therefore, the controlled contents are not limited to 
the control pattern shown in FIG. 4 and the contents may arbitrarily be 
changed. The present invention may be embodied in an automatic 
transmission or its control unit having a gear train and a 
hydraulic-pressure circuit different from those shown in FIGS. 12 and 15. 
Note that the power source may be another power output unit, such as an 
electric motor, which is employed in place of the engine. 
The advantages of the present invention will synthetically be described. 
According to the present invention, when the revolving speed of a power 
source is temporarily raised even in a case of a shift down in a power-off 
state, the hydraulic pressure for the frictional engagement unit is 
controlled in a manner different from that in the case of the ordinary 
power-off shift down. Therefore, the hydraulic pressure for the frictional 
engagement unit is made to be adaptable to the input torque. As a result, 
adverse shift shock and deterioration in the durability of the frictional 
engagement unit can reliably be prevented. 
Since the present invention is structured such that the initial hydraulic 
pressure is controlled to correspond to temporal rise in the revolving 
speed of the power source in a case of a so-called clutch-to-clutch down 
shift, delay of the shift, undesirable shift shock and deterioration in 
the durability of the frictional engagement unit can be prevented. 
Moreover, the present invention, structured such that the hydraulic 
pressure for the frictional engagement unit at the final stage of the 
shifting operation is controlled to correspond to the rise in the 
revolving speed of the power source, is able to prevent shock occurring 
attributable to torsional vibrations of the power transmission system. In 
particular, a significant effect can be obtained in a case where the 
revolving speed of the power source and the hydraulic pressure for the 
frictional engagement unit are controlled in a coordinated manner. 
When the hydraulic pressure for the frictional engagement unit is 
learning-controlled, the present invention is structured in such a manner 
that different learning controls are performed between the case where the 
revolving speed of the power source is temporarily raised and the case 
where the same is not raised even if the power-down shift is performed. 
Therefore, the hydraulic pressure for the frictional engagement unit can 
furthermore appropriately be controlled with respect to the input torque. 
As a result, shift shock can be prevented and the durability of the 
frictional engagement unit can be improved.