Hydraulic pressure control apparatus for an automatic transmission

A hydraulic pressure control apparatus for an automatic transmission including a microprocessor-based control unit for controlling hydraulic pressure supplied to hydraulic servos of friction engagement elements. The control apparatus calculates a target hydraulic pressure P.sub.TA for a condition at the start of the inertia phase in accordance with the input torque, and a gradient based on the target hydraulic pressure and a predetermined time t.sub.TA. A first up-sweep of hydraulic pressure is performed by the apparatus with the gradient. A relatively gradual gradient .delta.P.sub.TA is set based on a target rotation change rate for the input rotational speed to provide a predetermined change amount when the hydraulic pressure becomes the target hydraulic pressure P.sub.TA. The control apparatus performs a second up-sweep with the gradient .delta.P.sub.TA. When the rotational speed change .DELTA.N of the input rotation becomes a rotation change start-determining rotational speed dN.sub.S, the control apparatus feedback-controls the hydraulic pressure with a predetermined gradient while referring to the input rotation change. Further, the control apparatus detects the target shift start time and the rotational speed change rate at the target shift start time to correct the target hydraulic pressure P.sub.TA, the gradient .delta.P.sub.TA of the second sweep section, and the target shift start time t.sub.aim.

BACKGROUND OF THE INVENTION 
1. Field of the Invention 
The present invention relates to a hydraulic pressure control apparatus for 
an automatic transmission of a vehicle and, more particularly, to an 
apparatus that controls hydraulic pressure supplied to hydraulic servos 
for changing transmission paths of an automatic shift mechanism. 
2. Related Art 
An apparatus for reducing transmission shift shock by controlling hydraulic 
pressure is disclosed in Japanese patent application laying-open No. 
SHO-63-270971. The apparatus disclosed therein calculates a turbine 
(input) torque occurring immediately before the output of a shift signal 
using a turbine torque estimating means and turbine torque correcting 
means. Based on the calculated turbine torque and the gear ratios before 
and after the shift, the apparatus estimates a turbine torque that will be 
caused at the gear ratio after the shift. Using this estimate, the 
apparatus controls a fluid pressure regulating means for regulating the 
fluid pressure on hydraulic servos so that the turbine torque will 
smoothly change. 
Although the above-mentioned control apparatus may detect a turbine torque 
with a high precision and estimate an accurate turbine torque occurring 
after a shift, the apparatus cannot calculate a precise hydraulic pressure 
from the estimated turbine torque due to variations in the .mu. 
characteristics of the friction members, piston stroke errors, and the 
like. Accordingly, if the acting hydraulic pressure calculated by the 
apparatus is high relative to a proper hydraulic pressure, a drop of the 
rotational change during the beginning period of the inertia phase becomes 
large and causes a shift shock. If the acting hydraulic pressure is low 
relative to the proper hydraulic pressure, the start of the inertia phase 
is delayed. This causes a prolonged shift process followed by a sharp 
change in rotational speed, also producing shift shock. 
SUMMARY OF THE INVENTION 
Accordingly, it is an object of the present invention to provide a control 
apparatus of an automatic transmission that reduces shift shock by 
supplying controlled hydraulic pressure gradients to the friction 
engagement elements of the transmission. 
This and other objects of the invention are achieved by a hydraulic 
pressure control apparatus for controlling an automatic transmission of a 
vehicle including an automatic shift mechanism, changeover valves, 
pressure regulators, and a control unit. The automatic shift mechanism 
changes a transmission input shaft rotational speed over transmission 
paths by engaging and disengaging a plurality of friction engagement 
elements through operation of hydraulic servos. Hydraulic pressure to the 
hydraulic servos is regulated by the pressure regulators, preferably 
linear solenoids. The changeover valves (or shift valves) supply the 
regulated hydraulic pressure to, or discharge it from, the hydraulic 
servos to engage and disengage the friction engagement elements. 
To control hydraulic pressure to the servos, the microprocessor-based 
control unit generates feedback-controlled electrical outputs to the 
pressure regulators. The control unit includes: input torque calculation 
means, target hydraulic pressure calculation means, and hydraulic pressure 
control means. 
The input torque calculation means calculates an input torque of the input 
shaft based on a running condition of the vehicle. The input torque, thus 
calculated, is supplied to the target hydraulic pressure calculation means 
for calculating a target hydraulic pressure, for a condition immediately 
before the input rotational speed starts to change. The hydraulic pressure 
control means outputs a signal to the pressure regulators to provide a 
first sweep section of hydraulic fluid and to change the hydraulic 
pressure to the target hydraulic pressure with a predetermined first 
gradient. The control means then initiates a second sweep section of 
hydraulic fluid where the hydraulic pressure is changed from the target 
hydraulic pressure with a second gradient that is less than the 
predetermined first gradient. 
Preferably, the first sweep section is provided for a predetermined time 
established in consideration of a hydraulic pressure response delay. The 
predetermined first gradient is established on the basis of the 
predetermined time and the target hydraulic pressure. The second sweep 
section is established on the basis of a target rotation change rate that 
occurs when the input rotational speed changes by a predetermined amount. 
The second sweep section extends until the rotational speed change of the 
input rotation becomes a rotation change start-determining rotational 
speed. 
In an inertia phase, the hydraulic pressure control apparatus uses as a 
target value a rotation change rate of the input rotational speed. The 
hydraulic pressure control means is set so that the target value gradually 
changes during beginning and ending periods of said inertia phase. The 
hydraulic pressure control apparatus measures the time of the second sweep 
section, and corrects the target hydraulic pressure in accordance with the 
measured time. 
At the end of the second sweep section the hydraulic pressure control 
apparatus detects a rotational speed change rate of the input rotational 
speed, and corrects the gradient of the second sweep section in accordance 
with the detected change rate. The duration of the second sweep section is 
then compared with the predetermined time established in consideration of 
a hydraulic pressure response delay. The control apparatus corrects the 
target hydraulic pressure and the gradient of the second sweep section 
based on the comparison. 
Also, the control apparatus supplies hydraulic pressure to the hydraulic 
servos so that the hydraulic servos will complete a piston stroke before 
the first sweep section starts. The apparatus detects a rotational speed 
change rate of the input rotational speed and a time of the second sweep 
section at end of the second sweep section. The piston stroke time of the 
hydraulic servos is corrected in accordance with the detected rotational 
speed change rate and the detected time. 
The friction engagement elements that are engaged and disengaged may 
include two friction engagement elements that are simultaneously operated. 
The pressure to a first hydraulic servo of a first one of the friction 
engagement elements is controlled by the control unit as described above. 
The hydraulic servo for the second one of the two friction engagement 
elements is controlled on the basis of a hydraulic pressure calculated 
using a predetermined relational equation that depends on the hydraulic 
pressure of the first hydraulic servo. The relational equation is related 
to a predetermined coefficient established in accordance with a tie-up 
degree of the first and second friction engagement elements.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT 
Preferred embodiments of the present invention will now be described with 
reference to the accompanying drawings. As shown in FIG. 1, the present 
invention is useful in connection with an automatic transmission 
(automatic shift apparatus) which includes an changeover valves 36, 
hydraulic servos 37, and an automatic shift mechanism 31. The changeover 
valves 36 supply controlled hydraulic pressure to the hydraulic servos 37 
for engaging and disengaging a plurality of friction engagement elements 
33 of the automatic switch mechanism 31. The input shaft 34 of the 
automatic shift mechanism is connected to the output shaft of a torque 
converter (not shown), and the output shaft is connected to the vehicle 
drive wheels 35. Responsive to the controlled operation of the friction 
engagement elements, the automatic shift mechanism selects a transmission 
path through planetary gears 32. 
According to the present invention, control of the automatic transmission 
is achieved through an engine control unit 1 (ECU). The control unit 1 
includes a microcomputer for performing system control calculations using 
preset constants and input signals from an engine speed sensor 2, a 
throttle opening sensor 3, a transmission input shaft rotational speed 
(turbine rotational speed) sensor 5, a vehicle speed (=transmission output 
shaft rotational speed) sensor 6, and an oil temperature sensor 7. Outputs 
of the control unit are provided to linear solenoid valves SLS, SLU. 
Means 1a for calculating an input torque based on signals from the engine 
speed sensor 2, the throttle opening sensor 3 and the vehicle speed sensor 
6, are included within the control unit. The output of the input torque 
calculation means 1a is provided as an input to means 1b for calculating a 
target hydraulic pressure, in accordance with the calculated input torque, 
for a condition immediately before the start of the transmission inertia 
phase. Outputs of the input torque calculations means 1a and the target 
hydraulic pressure calculation means 1b are provided to hydraulic pressure 
control means 1c along with a signal from the transmission input shaft 
rotational speed sensor 5. The hydraulic pressure control means outputs 
signals to pressure regulators, preferably linear solenoid valves SLS, 
SLU, to achieve predetermined changing of hydraulic pressure using a first 
sweep section and a second sweep section. 
As shown in FIG. 2, hydraulic circuit according to the invention preferably 
includes the two linear solenoid valves SLS, SLU and a plurality of 
hydraulic servos. The hydraulic servos 9, 10 engage and disengage a 
plurality of friction engagement elements (clutches and brakes) to achieve 
various vehicle speeds. For example, four or five forward speeds and one 
rear speed, may be achieved by changing over the transmission paths 
through a planetary gear unit of the automatic shift mechanism. 
Input ports a.sub.1, a.sub.2 of the linear solenoid valves SLS and SLU are 
supplied with solenoid modulator pressure. According to the output of the 
hydraulic pressure control means 1c (FIG. 1), the linear solenoid valves 
SLS, SLU supply control pressure from their output ports b.sub.1, b.sub.2 
to control hydraulic chambers 11a, 12a of pressure control valves 11, 12. 
Input ports 11b, 12b of the pressure control valves 11, 12 are supplied 
with line pressure. The pressure regulated by the control pressure from 
the solenoid valves is supplied from the output ports 11c, 12c to the 
hydraulic servos 9, 10 via changeover (or shift) valves 13, 15, 
respectively. 
The hydraulic circuit diagram of FIG. 2 is merely for illustration of the 
basic concept thereof, and the hydraulic servos 9, 10 and the changeover 
valves 13, 15 are shown for illustrative purposes. As is known, an actual 
automatic shift mechanism is provided with many hydraulic servos and many 
changeover valves for switching the hydraulic pressure to the hydraulic 
servos. In each hydraulic servo, as shown in the hydraulic servo 10, a 
piston 19 is fitted in a cylinder 16 by an oil-tight seal 17. The piston 
19 is moved against the force of a spring 21 to contact outer friction 
plates 22 with inner friction members 23 in accordance with the regulated 
pressure applied to a hydraulic chamber 20 by the control valve 12. 
Although the friction plates and members are shown in the form of a clutch 
in FIG. 2, it should be understood that a brake may be constructed and 
operated in a similar manner. 
With reference to FIGS. 3 and 4, the release-side control for a 
transmission upshift will be now be described. Time measurement for the 
upshift starts (S1) with time t=0 when the control unit 1 outputs an 
upshift signal (START) based on the signals from the throttle opening 
sensor 3 and the vehicle speed sensor 6. The control unit 1 then 
calculates and outputs a predetermined signal to a linear solenoid valve 
SLS or SLU so that the hydraulic pressure to an engagement-side hydraulic 
servo (the engagement-side hydraulic pressure) P.sub.A becomes a 
predetermined pressure P.sub.S1 (S2). 
The predetermined pressure P.sub.S1 has been established to a value 
required to fill the hydraulic chamber 20 of the engagement-side hydraulic 
servo. The predetermined pressure P.sub.S1, is maintained for a 
predetermined time t.sub.SA. When the predetermined time t.sub.SA elapses 
(S3), the engagement-side hydraulic pressure P.sub.A sweeps down with a 
predetermined gradient (P.sub.S1 -P.sub.S2)/t.sub.SB !. When the 
engagement-side hydraulic pressure P.sub.A becomes a predetermined low 
pressure P.sub.S2, (S5), the down-sweep is stopped, and the pressure is 
held at the predetermined low pressure P.sub.S2 (S6). 
The predetermined low pressure P.sub.S2 has been established by the control 
unit so that it remains at the piston stroke pressure, or above, and 
inhibits rotation changes on the input shaft in any conditions. The 
predetermined low pressure P.sub.S2 is maintained until the measured time 
t indicates the elapse of the predetermined time t.sub.SE (S7). 
The control unit 1 then calculates an engagement-side hydraulic pressure 
P.sub.TA for a condition immediately before the rotation change of the 
input rotational speed N.sub.T starts (immediately before the inertia 
phase starts). The pressure P.sub.TA is calculated on the basis of a 
predetermined function P.sub.TA =F.sub.PTA (T.sub.T)! that changes with 
the input torque T.sub.T as indicated in FIG. 5(a) (S8). The target 
engagement-side hydraulic pressure P.sub.TA for a condition immediately 
before the start of the inertia phase is calculated by first calculating 
an engagement-side torque T.sub.A (=1/a(T.sub.T), where a is a torque 
dividing rate). The engagement-side torque T.sub.A is then used in an 
equation to determine P.sub.TA : 
EQU P.sub.TA =(T.sub.A /A.sub.A)+B.sub.A +dP.sub.TA 
where B.sub.A is a piston stroke (=spring load), A.sub.A is (friction 
plate effective radius).times.(piston area).times.(the number of friction 
plates).times.(coefficient of friction)!, and dP.sub.TA is a hydraulic 
pressure corresponding to a hydraulic pressure response delay. 
Based on the engagement hydraulic pressure P.sub.TA calculated 
corresponding to the input torque T.sub.T, for a condition immediately 
before the start of the inertia phase, the control unit 1 calculates a 
gradient from a predetermined time t.sub.TA (P.sub.TA -P.sub.SI)/t.sub.TA 
!, and sweeps up the engagement-side hydraulic pressure with the 
calculated gradient (S9). This up-sweep with the relatively small gradient 
increases the engagement torque so that the hydraulic pressure rises to a 
level occurring immediately before the input rotational speed starts to 
change, i.e. to the calculated target engagement hydraulic pressure 
P.sub.TA (S10). This state immediately before the upshift is a torque 
state where the output shaft torque temporarily falls sharply. 
The input torque T.sub.T (=turbine torque) is determined as indicated in 
FIGS. 6(a) and 6(b) by finding an engine torque corresponding to the 
throttle opening and the engine speed based on a map with linear 
interpolation, calculating a speed ratio from the input and output 
rotational speeds of the transmission, determining a torque ratio 
corresponding to the speed ratio based on a map, and multiplying the 
engine torque by the torque ratio. 
When the target engagement hydraulic pressure P.sub.TA is reached, i.e. 
when the hydraulic pressure change is considered to enter the inertia 
phase in which the rotation change of the input shaft rotational speed 
starts, the control unit 1 calculates a change .delta.P.sub.TA. This 
change is calculated using a function .delta.P.sub.TA =f.delta.P.sub.TA 
(.omega.a')! corresponding to the target rotation change rate 
(d.omega.a/dt, expressed as .omega.a') which is used as a target at the 
start of rotation change of the input shaft rotational speed N.sub.T, as 
indicated in FIG. 5(b) (S11). More specifically, the hydraulic pressure 
change .delta.P.sub.TA is calculated by an equation: 
EQU .delta.P.sub.TA =I(.omega.a)/k(t.sub.aim) 
where k is a constant, t.sub.aim is a target shift start time, .omega.a' is 
a target rotation change rate, and I is an inertia amount. The control 
unit 1 then sweeps up the hydraulic unit with a gradient corresponding to 
the hydraulic pressure change .delta.P.sub.TA (S12). This second up-sweep 
is continued until the rotation change .DELTA.N from the input shaft 
rotational speed N.sub.TS at the start of the rotation change reaches the 
shift start-determining rotational speed dN.sub.S (S13). 
The target shift start time t.sub.aim is set as a function of the input 
shaft rotational speed N.sub.T as indicated in FIG. 5(c). The shift 
start-determining rotational speed dN.sub.S is a minimum rotational speed 
that provides for detection of a rotational speed change as indicated in 
FIG. 5(d), and is dependent on the detection precision of the input shaft 
rotational speed sensor 5. Since the rotation detecting precision 
deteriorates in a low speed range, the rotational speed for detection 
needs to be sufficiently high. As the shift start determining rotational 
speed dN.sub.S, increases, the target shift start time t.sub.aim also 
increases as indicated in FIG. 5(e). 
The engagement-side hydraulic pressure change .delta.P.sub.I is 
feedback-controlled with a change .DELTA.N of the rotational speed based 
on detection of the input shaft rotational speed sensor 5. The control 
unit 1 then sweeps up the hydraulic pressure with a gradient of 
.delta.P.sub.I (S14). The up-sweep with .delta.P.sub.I is continued until 
.alpha..sub.1 % (e.g 70%) of the rotation change .DELTA.N over the 
completion of the shift is reached (S15), i.e. until 
(.DELTA.N.times.100)/N.sub.TS (g.sub.i -g.sub.i+1) becomes .alpha.%; where 
N.sub.TS is the input shaft rotational speed, .DELTA.N is the rotational 
change, g.sub.i is the gear ratio before the shift, and g.sub.i+1 is the 
gear ratio after the shift. 
When .alpha..sub.1 % of the rotation change is exceeded, another hydraulic 
pressure change .delta.P.sub.L is set by feedback control based on a 
smooth input shaft rotational speed change .DELTA.N. The control unit 1 
then sweeps up the hydraulic pressure with a gradient of .delta.P.sub.L 
(S16). Generally, the hydraulic pressure change .delta.P.sub.L provides a 
slightly less gradient than the hydraulic pressure change .delta.P.sub.I. 
The upsweep with the hydraulic pressure change .delta.P.sub.L is continued 
until .alpha..sub.2 % (e.g. 90%) of the rotational speed change up to 
point where the completion of the shift is nearing (S17). The target 
upsweep shift time t.sub.I with .delta.P.sub.I and .delta.P.sub.L is set 
on the basis of a plurality of throttle opening-vehicle speed maps 
corresponding to different oil temperatures as shown in FIG. 5(f). 
When the target shift time t.sub.I elapses, the measured time t.sub.F is 
set (S18). This state approximately corresponds to a state occurring after 
the end of the inertia phase. A relatively sharp hydraulic pressure change 
.delta.P.sub.F is then established. The control unit 1 sweeps up the 
hydraulic pressure sharply with the hydraulic pressure change 
.delta.P.sub.F (S19). When a time t.sub.FE set sufficiently for a rise to 
the engagement pressure elapses following the measured time t.sub.F (S20), 
the hydraulic pressure control at the engagement side is completed. 
The manner of setting the values .delta.P.sub.TA, .delta.P.sub.I, and 
.delta.P.sub.L will now be described with reference to FIGS. 7(a) and 
7(b). The target rotation change rate .omega.a' for the second upsweep 
based on .delta.P.sub.TA is calculated on the basis of the target shift 
start time t.sub.aim in accordance with the relation between time t and 
the time-differentiation (gradient) .omega.' of the input shaft rotational 
speed as indicated in FIG. 7(a). The gradual upsweep based on 
.delta.P.sub.I is roughly fixed to the target rotation change rate 
.omega.a'. This fixed state continues for a.sub.1 % (e.g. 70%) of the 
target shift time t.sub.I. Then, during the more gradual upsweep based on 
.delta.P.sub.L, the rotation change rate .omega.' gradually decreases from 
the target rotation change rate .omega.a'. This decreasing state 
(.omega.t') continues for a.sub.2 % (e.g. 30%) of the target shift time 
t.sub.I. 
The aforementioned proportions of .delta.P.sub.I and .delta.P.sub.L are 
expressed as proportions of a.sub.1, a.sub.2, and (a.sub.1 +a.sub.2 =1) of 
the target shift time t.sub.I. The flow chart of FIG. 4, however, 
expresses the proportions of .delta.P.sub.I and .delta.P.sub.L by 
proportions a.sub.1, a.sub.2 of the rotational speed change .DELTA.N. The 
two expressions mean substantially the same where the proportions to the 
entire shift are set. 
The changes of input shaft rotational speed N.sub.T in the shift beginning 
period (t.sub.aim) and the shift ending period (a.sub.2 .times.t.sub.I) 
become smooth as indicated in FIG. 7(b) on the basis of the changes of the 
rotation change rate .omega.' indicated in FIG. 7(a). The shift shock is 
thus reduced. The values of .delta.P.sub.TA, .delta.P.sub.I, and 
.delta.P.sub.L are set so that the rotation change rate .omega.' becomes 
as described above. 
More specifically, since N.sub.T S (g.sub.i -g.sub.i+1)/(t.sub.aim 
+t.sub.I), .omega.a'.times.(1/2)t.sub.aim +a.sub.1 .times.t.sub.I 
+a.sub.2 .times.t.sub.I .times.(1/2)!=N.sub.TS (g.sub.i 
-g.sub.i+1)/(t.sub.aim +t.sub.I), therefore, .omega.a'=N.sub.TS (g.sub.i 
-g.sub.i+1)/(1/2)t.sub.aim +a.sub.1 .times.t.sub.I +a.sub.2 
.times.t.sub.I .times.(1/2)!(t.sub.aim +t.sub.I). Thus, the target 
rotation change rate is calculated from the target shift start time 
t.sub.aim and the target shift time t.sub.I. If t=t.sub.aim to t.sub.aim 
+a.sub.1 .times.t.sub.I, then .omega.t'=.omega.a. If t=t.sub.aim +a.sub.1 
.times.t.sub.I to t.sub.aim +a.sub.1 .times.t.sub.I +a.sub.2 
.times.t.sub.I, then .omega.t'=.omega.a'-(.omega.a'/a.sub.2 
.times.t.sub.I).times.(t-t.sub.aim -a.sub.1 .times.t.sub.I). 
The control of the release-side hydraulic pressure P.sub.B for an upshift 
as described above will be described with reference to FIGS. 3 to 8. 
Although FIG. 3 indicates the engagement-release simultaneous control, 
that is, so-called "clutch-to-clutch" control, what is indicated therein 
also holds for control at the engagement-side alone. 
Referring particularly to the flow chart of FIG. 8, the control unit 1 
outputs a shift instruction so that time measurement for the release-side 
hydraulic pressure control starts simultaneously with that for the 
engagement-side hydraulic pressure control (S21). A high pressure P.sub.W 
provided by the engaging pressure is supplied as the release-side 
hydraulic pressure P.sub.B (S22). The supply of the high pressure P.sub.W 
is held for a time t.sub.SE until the first up-sweep starts (S23). 
The control unit 1 then calculates a release-side torque T.sub.B ' using 
the engagement-side hydraulic pressure P.sub.A and a function T.sub.B 
'=f.sub.TB (P.sub.A, T.sub.T)! of input torque T, (S24) as shown in FIG. 
9(a). The margin rates S.sub.1U and S.sub.2U, are then considered (T.sub.B 
=S.sub.1U .times.T.sub.B '+S.sub.2U), and the release-side torque T.sub.B 
is calculated (S25). A release-side hydraulic pressure P.sub.B is 
calculated from the release-side torque T.sub.B P.sub.B =f.sub.PB 
(T.sub.B)! (S26). 
More specifically, the torque T.sub.A divided to the engagement-side 
friction engaging elements is calculated as T.sub.A =A.sub.A +P.sub.A 
+B.sub.A !; where A.sub.A is the (effective radius).times.(piston 
area).times.(number of plates).times.(coefficient of friction), and 
B.sub.B is piston stroke pressure. Using the torque T.sub.A, the torque 
T.sub.B ' divided to the release-side friction elements is calculated as 
T.sub.B '=(1/b)T.sub.T -(a/b)T.sub.A !; where b is the release-side 
torque division, a is the engagement-side torque division, and T.sub.T is 
input shaft torque. Using the margin rates (tie-up degrees) S.sub.1U, 
S.sub.2U the control unit 1 sets a tie-up degree with respect to the 
engagement-side friction elements, with consideration of the drive wheels, 
and then calculates a release-side torque T.sub.B as T.sub.B =S.sub.1U 
.times.T.sub.B '+S.sub.2U !. 
The margin rates S.sub.1U, S.sub.2U are established by a plurality of 
throttle opening-vehicle speed maps selectively used corresponding to 
different oil temperatures as shown in FIG. 9(b), so that they agree with 
the driver's taste. The margin rates are normally set within ranges of 
S.sub.1U &gt;1.0, S.sub.2U &gt;0.0. Furthermore, the control unit 1 calculates 
from the release-side torque T.sub.B and the margin rates a release-side 
hydraulic pressure P.sub.B as P.sub.B =(T.sub.B /A.sub.B)+B.sub.B !; 
where A.sub.B is (the release-side friction element effective 
radius).times.(piston area).times.(number of plates).times.(coefficient of 
friction), and B.sub.B is the release-side piston stroke pressure. 
Since the down-sweep with the thus-calculated release-side hydraulic 
pressure P.sub.B is dependent on the engagement-side hydraulic pressure 
P.sub.A, the down-sweep has a two-gradient slope that bends at the start 
of the inertia phase (t.sub.TA) at which the input shaft rotational speed 
starts to change. Thus, the release-side hydraulic pressure P.sub.B 
includes a relatively steep down-sweep corresponding to the first up-sweep 
at the engagement side, and a relatively gradual down-sweep corresponding 
to the second up-sweep at the engagement side. Similarly, the down-sweep 
at the release side continues until the input shaft rotation change 
.DELTA.N becomes the predetermined rotation change start-determining 
rotational speed dN.sub.S (S27). The control unit 1 then sets a 
release-side hydraulic pressure change .delta.P.sub.E, and performs a 
down-sweep with the set hydraulic pressure change (S28). The down-seep 
continues until the release-side hydraulic pressure P.sub.B becomes 0 
(S29). The hydraulic pressure control at the release side is thus 
completed. 
Turning to FIGS. 10 and 11, the control of the release-side hydraulic 
pressure P.sub.A during a downshift will now be described. For a 
downshift, the release-side hydraulic pressure is the main object of the 
control while the engagement-side hydraulic pressure is controlled 
depending on the release-side hydraulic pressure. This is in contrast to 
the control for an upshift where the engagement-side hydraulic pressure is 
the main object of control while the release-side hydraulic pressure is 
controlled depending on the engagement-side hydraulic pressure, as 
described above. 
As shown in FIG. 11, the control unit 1 first outputs a downshift 
instruction so that the time measurement starts (S30). The release-side 
hydraulic pressure P.sub.A is a predetermined engaging pressure P.sub.W 
(S31). This state of the hydraulic pressure is continued for a 
predetermined time t.sub.sE, considering a hydraulic pressure rise time 
(t.sub.SA +t.sub.SB) (S32). The control unit 1 then calculates a 
release-side hydraulic pressure P.sub.TA occurring when the input shaft 
rotational speed starts to change (when the inertia phase starts) from the 
function P.sub.TA =fP.sub.TA (T.sub.T)! of the input indicated in FIG. 
5(a) (S33). More specifically, the release-side torque T.sub.A is 
calculated as: T.sub.A =(1/a)T.sub.T -(b/a){S.sub.2D /(1+S.sub.1D }; where 
S.sub.1D, S.sub.2D, are margin rates for downshift. 
From the torque T.sub.A, a target hydraulic pressure P.sub.TA is calculated 
as P.sub.TA =(T.sub.A /A.sub.A)+B.sub.A +dP.sub.TA !. The margin rates 
S.sub.1D, S.sub.2D are selected from a map as indicated in FIG. 8(b). They 
are normally set within ranges of S.sub.1D &lt;1.0, S.sub.2D &gt;0.0. The 
control unit 1 then determines a gradient to the target hydraulic pressure 
P.sub.TA based on a predetermined time t.sub.TA as (P.sub.TA 
-P.sub.W)/t.sub.TA !, and performs a first down-sweep with the gradient 
(S34). The first down-sweep is a relatively steep down-sweep, and 
continues until the release-side hydraulic pressure P.sub.A becomes the 
target hydraulic pressure P.sub.TA immediately before the start of the 
inertia phase (S35). 
The control unit 1 then calculates a release-side hydraulic pressure change 
.delta.P.sub.TA based on the function .delta.P.sub.TA =f.delta..sub.PTA 
(.omega.a')!, as indicated in FIG. 5(b) (S36). More specifically, the 
hydraulic pressure change .delta.P.sub.TA is calculated as 
.delta.P.sub.TA =(I/k) (.omega.a/t.sub.aim (1+S.sub.1D))!. The control 
unit 1 then performs a second down-sweep with the gradient of the 
hydraulic pressure change .delta.P.sub.T (S37). The down-sweep continues 
from the input shaft rotational speed N.sub.TS., occurring before the 
shift starts, to the shift start determining rotational speed dN.sub.S, at 
which a rotation change .DELTA.N is detected with a predetermined 
precision (S38). The second down-sweep continues until the target shift 
start time t.sub.aim has a more gradual gradient than the first 
down-sweep. 
Then the control unit 1 performs the down-sweep with the gradient of a 
predetermined hydraulic change .delta.P.sub.I, by feedback control while 
detecting a rotational speed change .DELTA.N based on the detection by the 
input shaft rotational seed sensor. The down-sweep is continued until the 
rotational speed change reaches .alpha..sub.1 % of the total rotational 
speed change up to the completion of the shift (S40). After that, a 
down-sweep with a more gradual gradient of the hydraulic pressure change 
.delta.P.sub.L is performed by similar feedback control (S41). This 
down-sweep is continued until .alpha..sub.2 % of the total rotational 
speed change is reached. 
After the completion of the shift up to .alpha..sub.2 % the control unit 1 
sets a hydraulic pressure change .delta.P.sub.F having a relatively steep 
gradient, and performs a down-sweep with the gradient (S43). When the 
release-side hydraulic pressure P.sub.B becomes 0, the release-side 
hydraulic pressure control for the downshift is completed (S44). 
The control of the engagement-side hydraulic pressure P.sub.B for a 
downshift will now be described with reference to FIGS. 10 and 12. The 
initial control steps S51-S57 are the same as steps S1-S7 of the 
engagement-side hydraulic pressure control during an upshift, and will not 
be described again. 
When the control enters the torque phase after the initial control steps, 
the control unit 1 calculates an engagement-side torque T.sub.B ' based on 
the release-side hydraulic pressure P.sub.A and a function T.sub.B 
'=f.sub.TB (P.sub.A, T.sub.T)! of input torque (S58), as indicated in FIG. 
9(a). Then, considering the margin rates, the control unit 1 calculates an 
engagement-side torque T.sub.B as T.sub.B =S.sub.1D .times.T.sub.B 
'+S.sub.2D ! (S59). From the engagement-side torque T.sub.B, an 
engagement-side hydraulic pressure P.sub.B is calculated as P.sub.B 
=f.sub.PB (T.sub.B)! (S60). 
More specifically, the engagement-side hydraulic pressure P.sub.B is 
calculated as T.sub.A =A.sub.A .times.P.sub.A +B.sub.A !.fwdarw.T.sub.B 
'=(1/b)T.sub.T -(a/b)T.sub.A !.fwdarw.T.sub.B =S.sub.1D .times.T.sub.B 
'+S.sub.2D !.fwdarw.P.sub.B =(T.sub.B /A.sub.B)+B.sub.B ! as in the 
calculation of the release-side hydraulic pressure for an upshift. Since 
the hydraulic pressure P.sub.B is dependent on the release-side hydraulic 
pressures based on the first down-sweep, the second down-sweep, and the 
down-sweep with .delta.P.sub.I, the hydraulic pressure P.sub.B, undergoes 
the first up-sweep with a relatively steep gradient, the second up-sweep 
with a relatively gradual gradient, and the third up-sweep with a more 
gradual gradient corresponding to .delta.P.sub.I. The hydraulic pressure 
P.sub.B continues up to .alpha..sub.1 % of the total input shaft 
rotational speed change (S61). 
The changeover of engagement of the friction engagement elements is nearly 
completed when .alpha..sub.1 % (e.g 70%) of the total input shaft 
rotational speed change is reached. For the remaining control process, the 
control unit 1 calculates an engagement-side torque T.sub.B T.sub.B 
=f.sub.TB (P.sub.A, T.sub.T)! (S62), and performs control based on the 
engagement-side hydraulic pressure P.sub.B =f.sub.PB (T.sub.B)! 
determined from the engagement side torque T.sub.B without involving 
margin rates (S63). More specifically, the engagement-side hydraulic 
pressure P.sub.B that does not involves margin rates (tie-up degrees) is 
calculated as T.sub.A =A.sub.A .times.P.sub.A +B.sub.B !.fwdarw.T.sub.B 
=(1/b)(T.sub.T)-(a/b)T.sub.A !.fwdarw.P.sub.B =(T.sub.B /A.sub.B)+B.sub.B 
!. Control is performed based on the engagement-side hydraulic pressure 
P.sub.B dependent on the release-side hydraulic pressure, until 
.alpha..sub.2 % (e.g. 90%) of the total input shaft rotational speed 
change is reached (S64). 
The time t.sub.F at the completion of .alpha..sub.2 % is stored (S65). 
Then, the up-sweep with a relatively steep hydraulic pressure change 
.delta.P.sub.F is performed (S65). When a predetermined time t.sub.FE 
corresponding to the engagement of a one-way clutch elapses after the 
target set time t.sub.I, which is set based on a map as indicated in FIG. 
5(f), for example, the engagement-side hydraulic pressure control for the 
downshift is completed (S67). 
Turning now to FIGS. 13 and 14, if a permissible range of 
0.9.times.t.sub.aim to 1.1.times.t.sub.aim is set with respect to the 
target shift start time t.sub.aim, and a permissible range of 0.9 
.omega.a' to 1.1 .omega.a' is set with respect to the target rotation 
change rate .omega.a', as shown in FIG. 13(a), there exist patterns 1 to 6 
outside the permissible range (the shadowed central section). As shown in 
FIG. 13(b), if the target hydraulic pressure P.sub.TA for the time of 
start of the input shaft rotation change (start of the inertia phase) is 
corrected to P.sub.TA =P.sub.TA +.DELTA.P.sub.0 !, the gradient 
.delta.P.sub.TA of the second up-sweep (or down-sweep) is corrected to 
.delta.P.sub.TA =.delta.P.sub.TA +.delta..DELTA.P.sub.0 !, and the target 
shift start time t.sub.aim is corrected to t.sub.aim =t.sub.aim 
+.DELTA.t.sub.0 !, then the values .DELTA.P.sub.0, .delta..DELTA.P.sub.0, 
and .DELTA.t.sub.0 are corrected as indicated in patterns 1 to 6. 
If the engaging force occurring with respect to the hydraulic pressure is 
made too large or too small by variations in the piston stroke, the return 
spring load, the friction coefficients and the like, the control unit 1 
corrects the hydraulic pressure P.sub.TA for a condition immediately 
before the start of the second sweep (.DELTA.P.sub.0). More specifically, 
if the rotation change rate .omega.s' is within the permissible range (0.9 
.omega.a'.ltoreq..omega.s'.ltoreq.1.1 .omega.a'), and the time before the 
start of shift is short (t.sub.S &lt;0.9 t.sub.aim), then the target 
hydraulic pressure P.sub.TA is corrected to a reduced level (P.sub.TA 
=P.sub.TA -.DELTA.P.sub.0), as shown in pattern 1. If the rotation change 
rate .omega.s' is within the permissible range and the time before the 
start of shift is long (t.sub.S &gt;1.1 t.sub.aim), then the target hydraulic 
pressure P.sub.TA is corrected to an increased level (P.sub.TA =P.sub.TA 
+.DELTA.P.sub.0), as shown in pattern 2. The time t.sub.S is calculated as 
t.sub.S =t.sub.TE -t.sub.TS !; where t.sub.TS is the time at which the 
hydraulic pressure P.sub.A becomes the target hydraulic pressure P.sub.TA, 
and t.sub.TE is the time at which the rotation change .DELTA.N becomes 
detectable (.DELTA.N.gtoreq.dN.sub.S). From the equation P.sub.TA 
=(T.sub.A /A.sub.A)+B.sub.A +dP.sub.TA, corrected hydraulic pressures at 
the engagement-side and release-side friction engagement elements are 
calculated. Then an average of number n of such calculated values is 
calculated to determine a corrected hydraulic pressure. 
If the rotation change is made too large or too small with respect to the 
hydraulic pressure change (increase or decrease) by variations in the 
inertia (I) amount, the friction coefficients and the like, the control 
unit 1 corrects the sweep gradient .delta.P.sub.TA 
(.DELTA..delta.P.sub.0). More specifically, after determining the rotation 
change rate .omega.s' occurring when the input shaft rotational speed 
change becomes the rotation change start-determining rotational speed, if 
1.0 .omega.a'&lt;.omega.s' and t.sub.S .ltoreq.1.1 t.sub.aim, then the sweep 
gradient is corrected to a more gradual gradient (.delta.P.sub.TA 
=.delta.P.sub.TA -.DELTA..delta.P.sub.0), as shown in pattern 3. If 0.9 
.omega.a'&gt;.omega.s and t.sub.S &gt;0.9 t.sub.aim, then the sweep gradient is 
corrected to a steeper gradient (.delta.P.sub.TA =.delta.P.sub.TA 
+.DELTA..delta.P.sub.TA), as shown in pattern 4. The inertia amount I is 
corrected using the aforementioned equation .delta.P.sub.TA =(I/A.sub.A 
*a)*(.omega.a'/t.sub.aim)!. An average of number n of such values is 
calculated to determine a corrected value. 
If both the rotation change rate .omega.s' and the target time t.sub.S 
become outside the permissible ranges because of a too large or too small 
hydraulic pressure response delay, the control unit 1 corrects the 
aforementioned sweep initial pressure P.sub.TA, sweep gradient 
.delta.P.sub.TA, and target shift start time t.sub.aim (.DELTA.P.sub.0, 
.DELTA..delta.P.sub.0, .DELTA.t.sub.0). If 1.1 .omega.a'&lt;.omega.s and 
t.sub.S &gt;1.1 t.sub.aim, the control unit 1 corrects the initial pressure 
P.sub.TA to an increased level, the sweep gradient .delta.P.sub.TA to a 
reduced value, and the target time t.sub.aim to an increased time 
(P.sub.TA =P.sub.TA +.DELTA.P.sub.0, .delta.P.sub.TA =.delta.P.sub.TA 
-.DELTA..delta.P.sub.0, t.sub.aim =t.sub.aim +.DELTA.t.sub.0), as shown in 
pattern 5. If 0.9 .omega.a'&gt;.omega.s' and t.sub.s &lt;0.9 t.sub.aim, the 
control unit 1 corrects the initial pressure P.sub.TA to an increased 
level, the sweep gradient .delta.P.sub.TA to an increased value, and the 
target time t.sub.aim to a shortened time (P.sub.TA =P.sub.TA 
-.DELTA.P.sub.0, .delta.P.sub.TA =.delta.P.sub.TA +.DELTA..delta.P.sub.0, 
t.sub.aim =t.sub.aim -.DELTA.t.sub.0), as shown in pattern 6. 
In patterns 5 and 6, based on the hydraulic pressure response delay, 
control is performed considering a hydraulic pressure response delay as 
well as the target time t.sub.aim. If the hydraulic pressure response 
delay is not considered, repeated calculation of target times t.sub.aim 
will only fail to achieve a target time within a range of the shadowed 
central section. If the signal value from the control unit 1 has risen to 
the initial pressure P.sub.TA and an up-sweep with a sweep gradient Of 
.delta.P.sub.TA has been instructed, as indicated in FIG. 13(a), then the 
control unit 1 corrects the initial pressure P.sub.TA to an increased 
level (+.DELTA.P.sub.0) if a hydraulic pressure response delay (d.sub.PTA) 
is large and causes a slow rise of the actual hydraulic pressure. 
Correspondingly, the control unit 1 corrects the sweep gradient 
.delta.P.sub.TA to a reduced value (-.DELTA..delta.P.sub.0) and corrects 
the target time t.sub.aim to an increased time (.DELTA.t.sub.0). The 
control unit 1 then determines a corrected value based on an average of 
number n of values as described above. 
If the piston stroke does not reach a predetermined value or it exceeds the 
value before the first sweep starts (i.e. if the conditions are outside 
what are covered by the table of FIG. 13(a)) correction is performed by 
determining the first fill time t.sub.SE. If .omega.s'&gt;.omega..sub.max and 
(t.sub.TE -t.sub.TS).gtoreq.t.sub.Smax, where t.sub.TS is time at which 
the hydraulic pressure P.sub.A becomes the sweep initial pressure P.sub.TA 
(P.sub.A =P.sub.TA), and t.sub.TE is time at which the input shaft 
rotational speed .DELTA.N becomes the change start-determining rotational 
speed dN.sub.S (.DELTA.N.gtoreq.dN.sub.S), and .omega.s' is the rotation 
change rate at the time t.sub.TE, the correction is performed in such a 
direction that the time t.sub.SE up to the start of the first sweep will 
increase (t.sub.SE =t.sub.SE +.DELTA.t.sub.SE0). If t.sub.TE 
.ltoreq.t.sub.SE, the correction is performed in such a direction that the 
time t.sub.SE will decrease (t.sub.SE =t.sub.SE -.DELTA.t.sub.SE0). In the 
aforementioned equations, W'.sub.max is a predetermined maximum rotation 
change rate, and t.sub.Smax is a predetermined maximum time for the second 
sweep. 
There is thus provided, according to the present invention, a hydraulic 
pressure control apparatus for an automatic transmission wherein the 
hydraulic pressure shifts from the first sweep section to the second sweep 
section with a relatively small gradient as the inertia phase starts. The 
control apparatus quickly starts a shift and smoothes the torque 
fluctuation during the shift, thus preventing a shift shock or a prolonged 
shift process, that would otherwise be caused by an excessively high or 
low hydraulic pressure. 
The gradient of the first sweep section is established on the basis of a 
predetermined time set in consideration of a hydraulic pressure response 
delay. The control apparatus is able to reduce the errors caused by the 
response delay of hydraulic pressure and control the second sweep section 
that follows. Since the gradient of the second sweep section is set using 
as a target value the rotation change rate occurring when the input 
rotational speed changes by a predetermined amount, the control apparatus 
improves the follow-up of the actual rotation change rate to the target 
value. Precise hydraulic pressure control is thus achieved without causing 
shift shocks. 
Also, since the second sweep section continues until the actual rotational 
speed change becomes the rotation change start determining rotational 
speed, the control apparatus is able to always detect the rotational speed 
with a high precision regardless of input rotational speeds. Precise 
feedback correction control is thus achieved. 
Since the rotational speed changes in the beginning and ending periods of 
the inertia phase are smoothed, and since variations in the time of the 
second sweep section are corrected by determining a target hydraulic 
pressure of the first sweep section, the control apparatus reduces shift 
shocks and improves the shift feel. Shift shocks are also reduced by the 
control apparatus by correcting variations in the rotation change rate 
during the second sweep section through correction of the gradient of the 
first sweep section. Moreover, since the predetermined time is established 
in consideration of the hydraulic pressure response delay, the control 
apparatus performs appropriate corrections even if the hydraulic pressure 
response delay is too large or too small. 
According to the present invention the control apparatus is also able to 
determine and control the piston stroke time, even if the piston stroke 
achieved before the start of the first sweep section is shorter or longer 
than a predetermined stroke. A hydraulic pressure response delay and shift 
shocks at the start of the inertia phase are thus prevented despite an 
excessively short or long piston stroke. 
Also, since a second one of the hydraulic servos for the friction 
engagement elements is controlled depending on the hydraulic pressure 
control of the first one of the hydraulic servos, the control apparatus 
simplifies the hydraulic pressure control by simultaneously switching over 
the friction engagement elements (i.e. by performing "clutch-to-clutch" 
changeover). This reduces the required memory capacity of the control 
along with the associated cost thereof. Furthermore, since the relation 
between the engagement-side and release-side hydraulic pressure servos is 
maintained in a predetermined state, the control apparatus prevents shift 
shocks that would otherwise be caused by engine speed rise or tie-up 
resulting from improper independent setting of hydraulic pressures. Also, 
since the tie-up degree can be changed merely by changing the 
predetermined coefficient, the control apparatus facilitates the 
calibration, increases the freedom in varying the predetermined 
coefficient, and enables appropriate control that agrees with the driver's 
tastes (for example, appropriate feel caused by throttle changes with a 
fixed input torque). 
The embodiments which have been described herein are but some of the 
several which utilize this invention and are set forth here by way of 
illustration but not of limitation. It is obvious that many other 
embodiments which will be readily apparent to those skilled in the art may 
be made without departing materially from the spirit and scope of this 
invention.