Line pressure control system for continuously variable transmission accounting for input load changes caused by inertia during shifting

A control system for a CVT comprises a controller arranged to compute an engine revolution condition on the basis of an engine revolution speed indicative signal, to compute actual and estimated engine outputs on the basis of the engine revolution condition, to compute a rotational change condition of the drive system, to compute an actual input load change amount caused by the inertia by multiplying the actual rotational change condition detection value by a predetermined inertial constant, to compute an estimated input load change amount by multiplying the estimated rotational change condition detection value by a predetermined inertia coefficient, to select larger one of the actual input load change amount and the estimated input load change amount as the input load change amount for outputting a command signal, and to output the command signal to a line pressure duty valve on the basis of the engine output and the selected input load change amount so as to further accurately control a line pressure.

The contents of Application No. TOKUGANHEI 9-186558, with a filing date 
Jul. 11, 1997 in Japan, are hereby incorporated by reference. 
BACKGROUND OF THE INVENTION 
The present invention relates to improvements in a control system for a 
continuously variable transmission (CVT) for a vehicle, and more 
particularly to a control system for a belt type CVT. 
A Japanese Patent Provisional Publication No. 8-200461 discloses a typical 
belt type CVT which comprises driver and follower pulleys and a belt for 
intercoupling the pulleys. Each of the driver and follower pulleys has a 
variable groove at which the belt is held. A belt holding force for 
holding the belt by each pulley is varied basically to correspond with a 
line pressure which is produced from a fluid pressurized by a pump through 
an electrically controlled pressure control valve. The line pressure is 
increased according to the increase of an input load applied from an 
engine to the CVT. The input load is basically calculated from a throttle 
opening and an engine revolution speed of an engine. 
SUMMARY OF THE INVENTION 
However, a line pressure control system of such a conventional CVT has not 
been arranged to take account of an inertia caused in a drive system 
including the engine. That is, if the rotational condition of the drive 
system is changed, an inertia moment of the drive system is also changed, 
and therefore the change amount of the inertia moment will be applied to a 
CVT mechanism. When an upshift is executed, more particularly, when a 
transmission ratio of the CVT is steppingly decreased during the upshift, 
the output revolution speed of the engine is quickly and largely 
decreased. Therefore, the engine torque to be applied to the CVT mechanism 
as an input torque will be decreased. However, during this upshift the 
follower pulley is quickly accelerated, and the inertia moment of the 
follower pulley is increased due to the acceleration of the follower 
pulley. That is, the engine torque which is decreasing is actually 
consumed to increase the rotation speed of the drive system. Therefore, 
the conventional control system tends to determine that the input load to 
the CVT is decreased by the upshift. This determination may invite the 
shortage of the line pressure. 
It is therefore an object of the present invention to provide a control 
system which further correctly determines an input load to a CVT upon 
taking account of the inertia due to shifting in order to enable a further 
suitable control of the line pressure to be applied to pulleys of the CVT. 
A control system according to the present invention comprises a 
continuously variable transmission (CVT), a CVT pressure control valve, an 
engine revolution condition detecting means, an engine output calculating 
means, a rotational change condition calculating mean, an input load 
change amount calculating means, and a hydraulic pressure control means. 
The CVT comprises a driver pulley, a follower pulley and a belt 
intercoupling the pulleys and varies a transmission ratio. The CVT 
pressure control valve controls a pressure of a fluid to be supplied to 
the CVT. The engine revolution condition detecting means detects a 
revolution condition of an engine connected to the CVT. The engine output 
calculating means calculates an engine output on the basis of the engine 
revolution condition. The rotational change condition calculating means 
calculates a rotational change condition of a drive system including the 
engine and the CVT. The input load change amount calculating means 
calculates a change amount of an input load which is caused by an inertia 
and applied to the CVT, on the basis of the rotational change condition. 
The hydraulic pressure control means outputs a pressure control signal to 
the CVT pressure control valve on the basis of the engine output and the 
input load change amount.

DETAILED DESCRIPTION OF THE INVENTION 
Referring to FIGS. 1 to 10D, there is shown an embodiment of a control 
system for a continuously variable transmission (CVT) in accordance with 
the present invention. 
FIG. 1 shows an embodiment of the CVT and the control system thereof 
according to the present invention. A power transmission mechanism of the 
CVT is basically the same as that of a power transmission mechanism 
disclosed in a U.S. Pat. No. 5,697,866 except that a torque converter 12 
is employed instead of a fluid coupling. Therefore, the same parts and 
elements of the power transmission mechanism are designated by same 
reference numerals, and a brief explanation thereof will be given 
hereinafter. 
As shown in FIG. 1, an engine 10 is connected to a V-belt type CVT 
mechanism 29 through the torque converter 12 and a forward/reverse change 
over mechanism 15. An output shaft 28 of the CVT mechanism 29 is connected 
to a differential unit 56. These elements 10, 12, 29 and 56 constitute a 
so-called drive system. The drive system transmits rotation of an output 
shaft 10a of the engine 10 to right and left drive shaft 66 and 68 at a 
selected speed ratio in a selected rotational direction to drive a 
vehicle. 
A throttle valve 19 is disposed in an air intake passage 11 of the engine 
10 and is arranged to change a degree of its opening in response to a 
depression degree of an accelerator pedal (no numeral). A throttle opening 
sensor 303 is installed to the throttle valve 19 to detect a throttle 
opening TVO of the throttle valve 19 and to output a signal indicative of 
the throttle opening TVO. An engine revolution speed sensor 301 is 
installed on the output shaft 10a of the engine 10 to detect an engine 
revolution speed N.sub.E and to output a signal indicative of the engine 
revolution speed N.sub.E. This engine revolution speed sensor 301 
functions as an engine revolution speed change detecting means and a drive 
system rotational change condition detecting means (an actual rotational 
change condition detecting means). 
The engine 10 is connected to an engine control unit 200 by which various 
engine operation factors such as fuel injection amount and timing and 
ignition timing are controlled so as to put the engine 10 in an optimum 
operating condition according to the running condition of the vehicle and 
a driver's intent. 
The throttle opening TVO indicative signal outputted from the throttle 
opening sensor 303 also corresponds to the magnitude of the depression 
degree of the accelerator pedal. Although the engine revolution speed 
sensor 301 is arranged to detect the rotation speed of the output shaft 
10a of the engine 10 in this embodiment, it will be understood that it may 
be arranged to count the number of ignition pulses of the engine 10 as an 
engine revolution speed. 
The torque converter 12 of this drive system is a lockup torque converter 
of a known type. This lockup torque converter 12 comprises a pump impeller 
(input member) 12c, a turbine runner or turbine (output member) 12d, a 
stator 12e, and a lockup facing member (or lockup clutch) 12f for 
providing a direct mechanical drive by directly coupling the input and 
output members 12c and 12d. The lockup facing member 12f is connected with 
an output shaft (turbine shaft) 13. The lockup facing member 12d separates 
an apply side fluid chamber 12a and a release side fluid chamber 12b. The 
release chamber 12b is formed between a cover 12g of a torque converter 12 
and the lockup facing member 12f, and the apply chamber 12a is disposed on 
the opposite side of the lockup facing member 12f. When a fluid pressure 
is supplied into the apply chamber 12a, the lockup facing member 12f is 
pressed against the cover 12g, and the torque converter 12 is put in a 
lockup state in which the input member 12c and the output member 12d are 
directly connected. When the fluid is supplied sufficiently into the 
release chamber 12b, the lockup facing member 12f is disengaged from the 
cover 12g, and the torque converter 12 is held in a non-lockup state (or 
unlockup state). The fluid pressure supplied to the release chamber 12b is 
drained through the apply chamber 12a. 
An input rotation speed sensor 305 is installed on an output shaft 13 
(turbine output shaft) of the torque converter 12 as a drive system 
rotation condition change detecting means (estimated rotational change 
condition detecting means) in order to detect an input rotation speed of 
the CVT mechanism 29. 
When the vehicle is put in a normal running condition, the forward clutch 
40 is put in a full engagement state. Therefore, the rotation speed of the 
turbine output shaft 13 is used as an input rotation speed N.sub.Pri of 
the CVT mechanism 29. When a foot of a driver is released from the 
accelerator pedal, the forward/reverse change over mechanism 15 functions 
to control a creep running force by variably adjusting the engagement 
force of the forward clutch 40. The fluid supplied to the release chamber 
12b is drained through the apply chamber 12a to a reservoir 130, and a 
drained fluid of the fluid supplied to the apply chamber 12c is supplied 
from the release chamber 12b to other cooling and lubricating systems. 
That is, the selecting control between the lockup state and the unlockup 
state is executed by changing a flow direction of the fluid supplied to 
the lockup mechanism of the torque converter 12 as mentioned above without 
changing the fluid passage. 
The forward/reverse drive direction change over mechanism 15 comprises a 
planetary gear system 17, the forward clutch 40 and a reverse brake 50. 
The planetary gear system 17 comprises multistage pinion trains and a 
pinion carrier supporting these pinion trains. The pinion carrier is 
connected to the driver pulley 16 of the CVT mechanism 29 through the 
driver pulley shaft 14, and a sun gear is connected to the turbine 
rotation shaft 13. The pinion carrier is engageable with the turbine 
rotation shaft 13 by means of the forward clutch 40. A ring gear of the 
planetary gear system 17 is engageable with a stationary housing. 
The reverse brake 50 is disposed between a ring gear of the planetary gear 
system 17 and a stationary housing to hold the ring gear. When the forward 
clutch 40 is put in the engaged state by the supply of the fluid pressure 
to a fluid chamber 40a, the driver pulley shaft 14 and the turbine output 
shaft 13 are rotated in the same direction with equal speed through a 
pinion carrier. When the reverse brake 50 is engaged by the supply of the 
fluid pressure to the fluid chamber 50a, the driver pulley shaft 14 and 
the turbine output shaft 13 are rotated with equal speed but in the 
opposite direction. 
The V-belt CVT mechanism 29 comprises a driver pulley 16, a follower (or 
driven) pulley 26, and a V-belt 24 for transmitting power between the 
pulleys. The driver pulley 16 is mounted on the driver pulley shaft 14. 
The driver pulley 16 comprises an axially stationary fixed conical disk 
18, and an axially movable conical disk 22, which confront each other and 
define a V-shaped pulley groove therebetween for receiving the V-belt 24. 
The fixed disk 18 rotates as a unit with the driver shaft 14. By a fluid 
pressure in a driver pulley cylinder chamber 20, the movable disk 22 is 
axially movable. 
The follower pulley 26 is mounted on a follower pulley shaft 28. The 
follower pulley 26 comprises an axially stationary fixed conical disk 30, 
a follower pulley cylinder chamber 32, and an axially movable conical disk 
34. The fixed and movable disks 30 and 34 confront each other and define a 
V-shaped pulley groove for receiving the V belt 24. The fixed disk 30 
rotates as a unit with the follower shaft 28. The movable disk 34 is 
axially movable in dependence on a fluid pressure in the follower pulley 
cylinder chamber 32. 
The V-belt CVT mechanism 29 further comprises a stepping motor 108 
controlled by a transmission control unit 300. A pinion 108a is connected 
to a rotation shaft of the stepping motor 108 and is meshed with a rack 
182. An end of the rack 182 and the movable conical disk 22 are 
interconnected through a lever 178. By the operation of the stepping motor 
108 according to the drive signal D.sub.S/M from the transmission control 
unit 300, the movable conical disk 22 of the driver pulley 16 and the 
movable conical disk 34 of the follower pulley 26 are axially moved to 
vary the effective radius of the contact position of each pulley of the 
driver and follower pulleys 16 and 26 axially. By so doing, the CVT 
mechanism 29 can vary the speed ratio (transmission ratio or pulley ratio) 
between the driver pulley 16 and the follower pulley 26. 
The shift control system (pulley ratio varying control) is generally 
arranged to vary the pulley groove width of one of the driver and follower 
pulleys and allow the groove width of the other to be adjusted 
automatically. The arrangement is achieved by the belt of a push type 
which transmits the driving force mainly in the pushing direction. The 
push type V-belt 24 comprises a set of plates which are overlappingly 
arranged in the longitudinal direction or wound direction of the belt. 
A drive gear 46 fixed to the follower pulley shaft 28 is meshed with an 
idler gear 48 formed on an idler shaft 52. A pinion gear 54 formed on the 
idler shaft 52 is meshed with a final gear 44 with which right and left 
drive shafts 66 and 68 are interconnected through the differential unit 
56. A vehicle speed sensor 302 is installed on this final output shaft to 
detect a vehicle speed VSP and output a signal indicative of the vehicle 
speed VSP to the transmission control unit 300. 
Next, a construction of a hydraulic pressure control apparatus of the CVT 
will be discussed. The hydraulic pressure control apparatus comprises a 
pump 101 driven by the engine 10. The pump 101 draws a working fluid from 
a reservoir 130 and supplies the fluid to an actuator unit 100 while 
applying a sufficient pressure to the fluid. The construction of the 
actuator unit 100 is basically the same as that of the actuator unit 
disclosed in a U.S. Pat. No. 5,697,866. Therefore, the same parts and 
elements are designated by same reference numerals, and only a brief 
explanation thereof will be given hereinafter. 
A manual valve 104 is directly operated by a selector lever 103 so as to 
switchingly control a clutch pressure P.sub.CL to a cylinder chamber 40a 
of the forward clutch 40 and a brake pressure P.sub.BRK to a cylinder 
chamber 50a of the reverse brake 50. 
An inhibitor switch 304 installed on the selector lever 103 detects a shift 
position selected by the selector lever 103 and outputs a shift range 
signal S.sub.RANGE indicative of the selected shift position. More 
particularly, the shift range signal S.sub.RANGE includes signals 
corresponding to P, R, N, D, 2, L according to the actually selected shift 
position. 
A shift control valve 106 is controlled according to a relative 
displacement between the stepping motor 108 and the movable conical disk 
22 of the driver pulley 16, more particularly it is controlled according 
to the movement of the lever 178. That is, the shift control valve 106 
controls the fluid pressure (line pressure) P.sub.L(Pri) supplied to the 
driver pulley 16 according to a relative relationship between the required 
transmission ratio and a groove width of the driver pulley 16. 
A lockup control duty valve 128 controls a lockup mechanism of the torque 
converter 12 so as to put the torque converter 12 in one of a lockup state 
and an unlockup state. A transmission control unit 300 outputs a drive 
signal D.sub.L/U to the lockup control valve 128 to operate the lockup 
mechanism of the torque converter 12. More particularly, when the drive 
signal D.sub.L/U indicative a large duty ratio is outputted to the lockup 
control duty valve 128, the torque converter 12 is put in the lockup 
state. When the drive signal D.sub.L/U indicative of a small duty ratio is 
outputted to the lockup control valve 128, the torque converter 12 is put 
in the unlockup (non-lockup) state. 
A clutch engagement duty valve 129 controls an engagement force for one of 
the forward clutch 40 and the reverse brake 50 according to a drive signal 
D.sub.CL from the transmission control unit 300. When the drive signal 
D.sub.CL indicates a large duty ratio, one of the forward clutch 40 and 
the reverse brake 50 is engaged. When the drive signal D.sub.CL indicates 
a small duty ratio, it is disengaged. 
A line pressure control duty valve 120 controls the line pressure P.sub.L 
supplied mainly to the follower pulley 26 and partly to the driver pulley 
16 so as to hold the belt 24 by the pulleys 26, 16 according to a drive 
signal D.sub.PL from the transmission control unit 300. This duty valve 
120 is represented as a modifier duty valve in the above cited Patent 
document. The reason for this is the output pressure from the duty valve 
120 once functions as a pilot pressure of a pilot pressure control valve 
named as a pressure modifier valve. As a result, an output pressure from 
the pressure modifier valve functions as a pilot pressure of the line 
pressure control valve so as to control the line pressure P.sub.L produced 
in an upstream side of the line pressure control valve. That is, by 
controlling the duty ratio of the duty valve 120, the line pressure 
P.sub.L is indirectly controlled. 
In this embodiment, the objective line pressure P.sub.L(OR) is linearly 
increased according to the increase of the control signal to the line 
pressure control duty valve 120 or duty ratio D/T.sub.PL of the drive 
signal except for dead zone as shown in FIG. 2. More particularly, when 
the output pressure of the pressure modifier valve is increased, the base 
(original) pressure of the clutch pressure and the base pressure of the 
lockup pressure of the torque converter 12 are simultaneously increased 
although the gradients and intercepts thereof are different respectively 
from others. 
The transmission control unit 300 comprises a microcomputer 310 functioning 
as a control means, and four drive circuits 311 to 314. The microcomputer 
310 outputs control signals for controlling the CVT mechanism 29 and the 
actuator unit 100 by executing a process shown by a flowchart of FIG. 3. 
The first to fourth drive circuits 311 to 314 respectively convert the 
control signals outputted from the microcomputer 310 into drive signals 
for practically controlling the actuators such as the stepping motor 108 
and the respective duty valves 120, 128 and 129. 
The microcomputer 310 comprises an input interface circuit 310a including 
an A/D converter, a calculation processor 310b such as a microprocessor, a 
storage device 310c such as ROM and RAM, and an output interface circuit 
310d including a D/A converter. 
The microcomputer 310 executes various operations such as obtaining a 
rotation angle of the stepping motor 108 for setting the actual 
transmission ratio, outputting a pulse signal S.sub.S/M for achieving the 
rotation angle, obtaining the optimum line pressure for holding the belt 
24, calculating the duty ratio D/T.sub.PL of the line pressure control 
duty valve 120 for achieving the necessary line pressure P.sub.L, 
obtaining the fluid pressure (torque converter pressure) P.sub.T/C for 
controlling the lockup mechanism of the torque converter 12, calculating 
the duty ratio D/T.sub.L/U of the lockup control duty valve 128 for 
achieving the torque converter pressure P.sub.T/C, outputting a lockup 
control signal S.sub.L/U according to the lockup control duty ratio 
D/T.sub.L/U, obtaining the fluid pressure (clutch pressure) P.sub.CL which 
is optimum to execute a creep running of the vehicle under when the 
accelerator pedal is put in the off condition (a foot of an operator is 
released from the accelerator pedal), calculating a duty ratio D/T.sub.CL 
of the clutch engagement control duty valve 129 necessary for achieving 
the required clutch pressure P.sub.CL, outputting a clutch engagement 
control signal S.sub.CL according to the clutch pressure control duty 
ratio D/T.sub.CL, and so on. 
The first drive circuit 311 converts the pulse control signal S.sub.S/M 
into a drive signal D.sub.S/M and outputs it to the stepping motor 108. 
The second drive circuit 312 converts the pulse control signal S.sub.PL 
into a drive signal D.sub.PL and outputs it to the line pressure control 
duty valve 120. The third drive circuit 313 converts the pulse control 
signal S.sub.L/U into a drive signal D.sub.L/U and outputs it to the 
lockup control duty valve 128. The forth drive circuit 314 converts the 
pulse control signal S.sub.CL into a drive signal D.sub.CL and outputs it 
to the clutch engagement control duty valve 129. If the control signals 
according to the duty ratio and the pulse control signals satisfactorily 
indicate the desired duty ratios and number of pulses, the drive circuits 
311 to 314 simply execute the amplifications thereof. 
The engine control unit 200 also comprises a microcomputer which is 
interconnected with the microcomputer 310 of the transmission control unit 
300 such that the engine 10 and the CVT are controlled according to the 
vehicle running condition. 
The transmission control of the present embodiment will be discussed with 
reference to a calculating process shown by a flowchart of FIG. 3 which is 
executed by the microcomputer 310. This calculating process is basically 
executed when the D-range is selected and when the engine control unit 200 
requires no process to the transmission control unit 300. The detailed 
explanation of this shift control is disclosed in the U.S. Pat. No. 
5,697,866. 
Herein, a general flow of the shift control will be discussed. This 
calculating process is implemented as a timer interruption process at 
predetermined sampling time .DELTA.T such as 10 milliseconds (msec). 
Although the flowchart of FIG. 3 does not show a step for a communication 
with others, the transmission control unit 300 timely executes a process 
for reading programs, maps and data used in the processor 310b from the 
storage device 310c and properly updates data calculated at the processor 
310b and stores it in the storage device 310c. 
At a step S1, the microcomputer 310 reads the signals indicative of the 
vehicle speed V.sub.SP outputted from the vehicle speed sensor 302, the 
engine revolution speed N.sub.E outputted from the engine revolution 
sensor 301, the input rotation speed N.sub.Pri outputted from the input 
rotation speed sensor 305, a throttle opening TVO outputted from the 
throttle, opening sensor 303 and a shift range signal S.sub.RANGE from the 
inhibitor switch 304. 
At a step S2, the microcomputer 310 calculates a present transmission ratio 
C.sub.P according to the vehicle speed V.sub.SP and the input rotation 
speed N.sub.Pri by executing an independently provided process (sub 
routine). More particularly, the output rotation speed N.sub.Sec of the 
CVT mechanism 29 is obtained by dividing the vehicle speed V.sub.SP in 
proportion with the final output shaft rotation speed by a final reduction 
speed ratio n (N.sub.Sec =V.sub.SP /n), and the present transmission ratio 
C.sub.P is obtained by dividing the input rotation speed N.sub.Pri by the 
obtained output rotation speed N.sub.Sec (C.sub.P =N.sub.Pri /N.sub.Sec) 
At a step S3, the microcomputer 310 calculates the engine torque T.sub.E 
on the basis of the throttle opening TVO and the engine revolution speed 
N.sub.E by executing an independently provided process (sub routine) such 
as a retrieval of a control map. More particularly, the present engine 
torque T.sub.E is obtained from the output characteristic map shown in 
FIG. 4 according to the throttle opening TVO and the engine revolution 
speed N.sub.E. 
At a step S4, the microcomputer 310 executes a lockup control by jumping 
the routine to an independently provided process (subroutine). More 
particularly, a lockup speed V.sub.ON and an unlockup speed V.sub.OFF are 
determined by using a control map shown in FIG. 5 according to the vehicle 
speed V.sub.SP and the throttle opening TVO. Basically, when the vehicle 
speed V.sub.SP is greater than the lockup speed V.sub.ON, the 
microcomputer 310 generates and outputs the control signal S.sub.L/U 
including a lockup command for putting the torque converter 12 into the 
lockup state. When smaller than the unlockup speed V.sub.OFF, the 
microcomputer 310 generates and outputs the control signal S.sub.L/U 
including an unlockup command for putting the torque converter 12 into the 
unlockup state. In particular, in case that the condition of the torque 
converter 12 is changed from the unlockup state to the lockup state, when 
a difference between the engine rotation speed N.sub.E and the input 
rotation speed N.sub.Pri is greater than a predetermined value, that is, 
when a difference between the engine revolution speed N.sub.E and the 
rotation speed of the turbine of the torque converter 12 is greater than 
the predetermined value, a gain employed for increasing the duty ratio 
D/T.sub.L/U is increased according to the magnitude of the difference. 
When the difference is smaller than a predetermined value, that is, when 
the torque converter 12 tends to be put in the lockup state, the gain for 
increasing the duty ratio D/T.sub.L/U is decreased so as to buffer shift 
shocks caused by the transition to the full lockup state. 
At a step S5, the microcomputer 310 calculates a goal transmission ratio 
C.sub.D by executing an independently provided process (subroutine) such 
as the retrieval of a control map. The goal transmission ratio C.sub.D is 
the most ideal transmission ratio for achieving the present engine 
revolution speed N.sub.E from the vehicle speed and the throttle opening 
TVO. More particularly, as shown in FIG. 6, if a transmission ratio C, by 
which the vehicle speed V.sub.SP, the throttle opening TVO and the engine 
revolution speed N.sub.E are completely matched, is set, it is possible to 
ensure an acceleration according to the depression degree of the 
accelerator pedal (the throttle opening TVO) while satisfying the vehicle 
speed V.sub.SP and the engine revolution speed N.sub.E. Herein, if it is 
assumed that the map shown in FIG. 6 is the control map for setting the 
goal transmission ratio C.sub.D, a straight line crossing with an origin 
point and having a constant gradient is a constant transmission ratio. For 
example, a straight line having the largest gradient in the whole area of 
the shift pattern is the largest speed reduction ratio of the vehicle, 
that is, a maximum transmission ratio C.sub.LO. In reverse, a straight 
line having the smallest gradient in the whole area of the shift pattern 
is the smallest speed reduction ratio of the vehicle, that is, a minimum 
transmission ratio C.sub.DHi. Herein, if the 2-range is selected through 
the select lever 103, the shift control is implemented within an area from 
the maximum transmission ratio C.sub.LO to a 2-range minimum transmission 
ratio C.sub.2Hi. 
At a step S6, the microcomputer 310 calculates the objective transmission 
ratio C.sub.R according to an individually prepared process (subroutine). 
More particularly, when the goal transmission ratio C.sub.D is greater 
than the present transmission ratio C.sub.P, the objective transmission 
ratio C.sub.R is set to execute the downshift. When smaller than the 
present transmission ratio C.sub.P, the objective transmission ratio 
C.sub.R is set to execute the upshift. For example, the objective 
transmission ratio C.sub.R is set at a transmission ratio obtained at a 
moment elapsing a predetermined sampling time .DELTA.T from when the 
shifting is executed from the present transmission ratio C.sub.P by the 
highest shift speed dC.sub.R /dt or smallest time constant .tau.. Herein, 
if the throttle opening TVO is decreased from a nearly full open state, 
that is, when the depression degree of the accelerator pedal is decreased, 
the shift speed dC.sub.R /dt is little decreased or the time constant 
.tau. is little increased. Further, when the throttle opening TVO is 
further quickly decreased, such that the depression of the accelerator 
pedal is cancelled, the shift speed dC.sub.R /dt is further decreased or 
the time constant .tau. is further increased. That is to say, the 
objective transmission ratio C.sub.R is changed according to the closing 
change amount of the throttle valve 19. 
In this embodiment, the time constant .tau. is employed in order to set the 
objective transmission ratio C.sub.R, that is, to control the shift speed. 
Therefore, when the goal transmission ratio C.sub.D is set, the objective 
transmission ratio C.sub.R is determined as a curve which takes values 
gradually converging to the goal transmission ratio C.sub.D. 
At a step S7, the microcomputer 310 executes a clutch engagement control 
according to an individually provided process (subroutine). More 
particularly, as a basic manner, when the vehicle speed V.sub.SP is 
greater than a creep control threshold, the forward clutch 40 is engaged. 
When the vehicle speed V.sub.SP is smaller than the creep control 
threshold and when the throttle opening TVO is greater than a creep 
control full close threshold, the engagement of the forward clutch 40 is 
released. Such operations are executed by generating and outputting the 
corresponding signal S.sub.CL from the microcomputer 310 to the fourth 
drive circuit 314. When the vehicle speed V.sub.SP is smaller that the 
creep control threshold and when the throttle opening TVO is smaller than 
the full close threshold, the gain for changing the duty ratio D/T.sub.CL 
is changed in inverse proportion to the difference between the engine 
revolution speed N.sub.E and the input rotation speed N.sub.Pri (turbine 
output shaft rotation speed). By this arrangement, the engagement force of 
the clutch 40 is decreased if the vehicle tends to generate a creep 
running due to a road condition such as a up-slope road running condition, 
and the engagement force of the clutch 40 is increased if the vehicle does 
not tend to generate a creep running. 
At a step S8, the microcomputer 310 executes the control of the line 
pressure P.sub.L according to the process shown by a flowchart of FIG. 7. 
The detailed explanation of this line pressure control will be done later 
with reference to the flowchart of FIG. 7. 
At a step S9, the microcomputer 310 executes a transmission ratio control 
according to an individually prepared process (subroutine). More 
particularly, with respect to the objective transmission ratio C.sub.R, 
the total number of the pulses and the number of pulses per a unit time 
are determined. Then, the microcomputer 310 generates and outputs the 
pulse control signal S.sub.S/M satisfying the both numbers. After the 
execution of the step S9, the routine of this program returns to the main 
program. 
Next, the line pressure control executed at the step S8 of the flowchart of 
FIG. 3 will be discussed in detail with reference to the flowchart of FIG. 
7. 
At a step S801, the microcomputer 310 calculates a torque ratio t of the 
torque converter 12 from the speed ratio N.sub.E /N.sub.Pri by executing 
the individually provided calculation such as the retrieval of the control 
map. More particularly, the torque converter input and output speed ratio 
N.sub.E /N.sub.Pri is obtained by dividing the engine revolution speed 
N.sub.E by the input rotation speed N.sub.Pri equal to the turbine output 
rotation speed. The microcomputer 310 determines according to the torque 
converter input and output speed ratio N.sub.E /N.sub.Pri whether the 
torque converter 12 is put in the unlockup (converter) state or the lockup 
state. Further, the microcomputer 310 computes the torque ratio t 
according to the torque converter input and output speed ratio if the 
unlock state. 
At a step S802, the microcomputer 310 calculates a reference input torque 
T.sub.Prio by multiplying the torque ratio t by the engine torque T.sub.E. 
At a step S803, the microcomputer 310 reads a previous engine revolution 
speed N.sub.E(n-1) which is a value detected in the previous routine and 
stored in the storage device 310 while being updated. 
At a step S804, the microcomputer 310 calculates a rate dN.sub.E /dt of 
change in the engine revolution speed N.sub.E per time from the following 
equation (1). 
EQU dN.sub.E /dt=(N.sub.E(n) -N.sub.E(n-1)/.DELTA.T (1) 
where N.sub.E(n) is a present value of the engine revolution speed read at 
the step S1. 
At a step S805, the microcomputer 310 calculates a feedback inertia torque 
T.sub.INP by multiplying the opposite number of the rate dN.sub.E /dt by 
an inertia moment I of the drive system (hereinafter, calling as an 
inertia coefficient) as follows: 
EQU T.sub.INP =-(dN.sub.E /dt).multidot.I (2) 
At a step S806, the microcomputer 310 calculates an estimated shift amount 
.DELTA.C by subtracting the present transmission ratio C.sub.P from the 
objective transmission ratio C.sub.R (.DELTA.C=C.sub.R -C.sub.P). 
At a step S807, the microcomputer 310 calculates a rate dN.sub.Pri /dt of 
change in the input rotation speed N.sub.Pri from the following equation 
(3). 
EQU dN.sub.Pri /dt=(k.multidot.V.sub.SP .multidot..DELTA.C)/.DELTA.T(3) 
where k is a coefficient constituted by an inverse number of the final 
reduction speed ratio necessary for calculating the output rotation speed 
N.sub.Sec from the vehicle speed V.sub.SP. 
At a step S808, the microcomputer 310 calculates a feedforward inertia 
torque T.sub.INR by multiplying an opposite number of the rate (dN.sub.Pri 
/dt) of the input rotation speed by the inertia coefficient I as follows: 
EQU T.sub.INR =-(dN.sub.Pri /dt).multidot.I (4) 
At a step S809, the microcomputer 310 determines as to whether or not the 
feedback inertia torque T.sub.INP is greater than or equal to the 
feedforward inertia torque T.sub.INR. When the determination at the step 
S809 is affirmative, the routine proceeds to a step S810 wherein the 
feedback inertia torque T.sub.INP is selected as a representative inertia 
torque T.sub.IN (T.sub.IN =T.sub.INP). When the determination at the step 
S809 is negative, the routine proceeds to a step S811 wherein the 
feedforward inertia torque T.sub.INR is selected as the representative 
inertia torque T.sub.IN (T.sub.IN =T.sub.INR). 
After the execution of the step S810 or S811, the routine proceeds to a 
step S812 wherein the sum of the inertia torque T.sub.IN and the reference 
inertia torque T.sub.Pri is treated as the input toque T.sub.Pri 
(T.sub.Pri =T.sub.Pri0 +T.sub.IN) 
At a step S813, the microcomputer 310 calculates (computes) the reference 
line pressure P.sub.LO from the control map of FIG. 9 according to the 
input torque T.sub.Pri and the present transmission ratio C.sub.P. The 
control map of FIG. 9 defines the reference line pressure according to the 
input torque T.sub.Pri and the present transmission ratio C.sub.P. Since 
the line pressure P.sub.L functions to apply a load to a side surface of 
the belt 24 through the pulleys 16 and 26, it is preferable to decrease 
the line pressure P.sub.L in view of the durability and the energy 
efficiency. In contrast, since the belt 24 receives and transmits the 
torque, it is necessary to hold the belt 24 between the conical disks of 
each pulley 16, 26 so as not to lose the torque between the belt and 
pulleys. Therefore, it is necessary to increase the line pressure P.sub.L 
according to the increase of the input torque T.sub.Pri and/or the 
increase of the transmission ratio C.sub.P since the torque is increased 
according to the increase of the transmission ratio C.sub.O and/or the 
transmission ratio C.sub.P. The reference line pressure P.sub.LO is 
determined only from the transmission ratio C.sub.P and the input torque 
T.sub.Pri and is set far smaller than the value affecting the durability 
of the belt 24. 
At a step S814, the microcomputer 310 determines the reference line 
pressure P.sub.LO as the present value P.sub.LOR(n) of the objective line 
pressure. 
At a step S815, the microcomputer 310 calculates (computes) the line 
pressure control duty ratio D/T.sub.PL for achieving the objective line 
pressure P.sub.LOR from the control map of FIG. 2. 
At a step S816, the microcomputer 310 generates and outputs the line 
pressure control signal S.sub.PL corresponding to the line pressure 
control duty ratio D/T.sub.PL by executing an individually provided 
process (subroutine). 
At a step S817, the microcomputer 310 updates the previous value 
N.sub.E(n-1) by the present engine revolution speed N.sub.E(n) and stores 
the updated value N.sub.E as a previous value. Then, the routine returns 
to the step S9 of FIG. 3. Since the control map of the duty line pressure 
control duty ratio D/T.sub.PL may employ a known duty ratio control, the 
explanation thereof will be omitted herein. Further, the generation of the 
line pressure control signal S.sub.PL corresponding to the line pressure 
duty ratio D/T.sub.PL may employ a known PWM (Pulse Width Modulation) 
control, and therefore the explanation thereof will be omitted herein. 
The function of the shift control of the flowchart of FIG. 7 will be 
discussed in detail. 
By the execution of the steps S801 and S802, the reference input torque 
T.sub.Pri0, inputted to the CVT mechanism 29 is calculated by multiplying 
the engine torque T.sub.E by the torque ratio t. 
By the execution of the steps S803 to S805, the feedback inertia torque 
T.sub.INP is calculated. 
By the execution of the steps S806 to S808, the feedforward inertia torque 
T.sub.INR is calculated. 
As mentioned above, when the rotational condition of an object is changed, 
more accurately, when an acceleration is applied to the rotational 
direction of the object, an inertia moment is applied to the subject. 
Since it may be assumed that the inertia moment is an inertia torque, a 
product of the acceleration applied to the drive system (a rate of change 
in rotation speed per a unit time) and the inertia coefficient I of the 
drive system may be treated as an inertia torque inputted to the drive 
system (a change amount of the input load according to the change of the 
rotation). 
At the step S804, an acceleration applied to the drive system is obtained 
from the rate dN.sub.E /dt of the actual engine revolution speed N.sub.E. 
At the step S805, the feedback inertia torque T.sub.INP is calculated by 
multiplying the obtained acceleration by the inertia coefficient I. 
On the other hand, at the steps S806 and S807, an acceleration (angular 
acceleration) applied to the drive system is obtained from the rate 
dN.sub.Pri /dt of the input rotation speed N.sub.Pri achieved by the 
present objective transmission ratio. Next, at the step S808, the 
feedforward inertia torque T.sub.INR is obtained by multiplying the rate 
dN.sub.Pri /dt by the inertia coefficient I (T.sub.INR =dN.sub.Pri 
/dt.multidot.I). The words "feedback" and "feedforward" in this embodiment 
are used to mean an object obtained from the actual rotational change 
condition and an object obtained from a future rotational change 
condition. That is, the former word corresponds to "actual", and the 
latter word corresponds to "estimated". Therefore, the feedback inertia 
torque may be represented as an actual change amount of the input load, 
and the feedforward inertia torque may be represented as an estimated 
change amount of the input load. 
In this embodiment, the opposite values of the change speed dN.sub.E /dt of 
the engine revolution speed N.sub.E and the change speed dN.sub.Pri /dt of 
the input rotation speed N.sub.Pri are used to calculate the feedback 
inertia torque T.sub.INP and the feedforward inertia torque T.sub.INR by 
means of the equations (2) and (4), the reason for this is that the 
inertia torque caused during the upshift (the transmission ratio is 
decreased) is also taken account. That is, the inertia torque caused by 
the rotational change of the drive system has a direction according to the 
increase and decrease of the change amount of the input load. The increase 
of the rotational change of the drive system is mainly caused by the 
increase of the engine revolution speed N.sub.E and the upshift (the 
transmission ratio C is decreased). In case that the engine revolution 
speed N.sub.E is increased, the increased amount of the input torque 
caused by the increase of the engine revolution speed N.sub.E is already 
included in the engine torque T.sub.E and is relatively slowly changed. 
Therefore, the inertia torque functioning as the change amount of the 
input load to the CVT mechanism 29 is relatively small, and it is possible 
to adapt to the inertia torque by setting the gain of the control system 
at a relatively large value, or by superimposing the estimated maximum 
inertia torque as a drift amount. On the other hand, the inertia torque 
caused by the upshift (the transmission ratio is steppingly decreased) is 
not small and is unstable in magnitude. The detailed explanation of this 
inertia torque caused by the upshift will be discussed later. 
In case of such stepping upshift, the transmission ratio is decreased 
quickly (relatively), and therefore the input rotation speed N.sub.Pri and 
the engine revolution speed N.sub.E are decreased also quickly although 
the increase of the vehicle speed V.sub.SP is not so quick. That is, the 
magnitude of the inertia torque functioning as an input load change amount 
to the CVT mechanism 29 becomes large according to the magnitude of the 
change speed. It will be understood that this inertia torque is an inertia 
torque absorbed by the drive line. In contrast to the increase of the 
inertia torque, the control input such as the input rotation speed 
N.sub.Pri and the engine revolution speed N.sub.E is decreased. Therefore, 
it may be considered that the inertia torque T.sub.IN is in proportion 
with the input deceleration. In order to clarify the directions 
therebetween, the opposite values of the change speed dN.sub.E /dt of the 
engine revolution speed N.sub.E and the change speed dN.sub.Pri /dt of the 
input rotation speed N.sub.Pri are used to determine the inertia torque 
T.sub.IN. In case that a downshift is executed steppingly, the inertia 
torque of the drive system becomes small, and the engine revolution speed 
N.sub.E and the input rotation speed N.sub.Pri are increased. Therefore, 
it will be understood that the above-mentioned setting manner of the 
inertia torque is proper. 
At the steps S809 to S811, the larger one of the feedback inertia torque 
T.sub.INP and the feedforward inertia torque T.sub.INR is selected as the 
inertia torque T.sub.IN. At the steps S812 and S813, the reference line 
pressure P.sub.LO is determined in correspondence with the input torque 
T.sub.Pri which is the sum of the inertia torque T.sub.IN and the 
reference input torque T.sub.Pri0. Thereafter, at the steps S815 and S816, 
the line pressure control signal S.sub.PL is generated and outputted. 
With reference to time charts of FIGS. 10A to 10D, the manner of operation 
as to the determination of the inertia toque T.sub.IN will be discussed 
hereinafter. 
These time charts indicate a result of a simulation executed under the 
condition that goal transmission ratio C.sub.D shown in the control map of 
FIG. 6 was upshifted steppingly, that is, a condition that at moment to, a 
shifting was executed from 2-range to D-range when the torque converter 12 
was maintained at the lockup state and when the vehicle speed V.sub.SP and 
the throttle opening TVO were kept high and constant, as shown by 
alternate long and two short dashes line of FIG. 10C. In contrast, the 
present embodiment according to the present invention is arranged to 
determine the objective transmission ratio C.sub.R so as to gradually 
approach the goal transmission ratio C.sub.D by controlling the time 
constant .tau., as shown by a short dashes line of FIG. 10C. Although the 
shift control is executed so that the actual transmission ratio C.sub.P is 
adjusted to the objective transmission ratio C.sub.R, the actual 
transmission ratio C.sub.P is practically changed with a corresponding 
delay as shown by a continuous line of FIG. 10C. 
On the other hand, since the torque converter 12 is put in the lockup 
state, the engine revolution speed N.sub.E and the input rotation speed 
N.sub.Pri are equal with each other. The actual input rotation speed 
N.sub.Pri was increased by a constant gradient until the moment t.sub.01. 
Then, the gradient indicative of an increase rate of the input rotation 
speed N.sub.Pri became small from moment t.sub.02 in FIG. 10D. At the 
moment t.sub.03 the input rotation speed N.sub.Pri started decreasing. 
Thereafter, the actual input rotation speed N.sub.Pri continued 
decreasing. Just before moment t.sub.05 when the transmission ratio 
C.sub.P asymptotically approached the goal transmission ratio C.sub.D, the 
input rotation speed N.sub.Pri started increasing. Thereafter, the input 
rotation speed N.sub.Pri continued increasing. That is, the upshift of the 
actual transmission ratio C.sub.P was started at the moment t.sub.02 and 
terminated just before the moment t.sub.05. In contrast, a short dashes 
line of FIG. 10D shows an estimated input rotation speed N.sub.Pri(i) 
which is obtained by executing the estimation thereof upon being assumed 
that the objective transmission ratio C.sub.R is achieved simultaneously 
with the execution of the process of FIG. 3, that is, it is achieved in 
real time. Since it is assumed that the objective transmission ratio 
C.sub.R is achieved in real time with respect to the actual transmission 
ratio C.sub.P including the response delay, the estimated input rotation 
speed N.sub.Pri(i) starts decreasing prior to the actual input rotation 
speed N.sub.Pri and starts increasing prior to the actual input rotation 
speed N.sub.Pri. 
Herein, a consideration as to the actual input rotation speed N.sub.Pri 
will be given. Although the actual input rotation speed N.sub.Pri is 
represented as a result of the execution of the actual upshift, the 
inertia torque indicative of the change amount of the input load to the 
CVT mechanism 29 is increased during the shifting. That is, the shifting 
is executed while the inertia torque is absorbed by the drive line. As a 
result, the actual input rotation speed N.sub.Pri (=N.sub.E) is decreased. 
Therefore, the feedback inertia torque T.sub.INP obtained from the change 
speed (dN.sub.E /dt) will delay as compared with the timing taking account 
of the inertia torque. Therefore, if only the feedback inertia torque 
T.sub.INP is employed to calculate the line pressure P.sub.L, the shortage 
of the line pressure P.sub.L may be caused with respect to the total input 
load to the CVT mechanism 29. In order to avoid such a shortage of the 
line pressure P.sub.L, the feedforward inertia T.sub.INR is calculated 
from the change speed of the input rotation speed N.sub.Pri obtained from 
the estimated input rotation speed N.sub.Pri(i), which change speed is 
faster than the feedback inertia T.sub.INP in phase. Further, the larger 
one of the feedback inertia torque T.sub.INP and the feedforward inertia 
torque T.sub.INR is selected as a representative inertia torque T.sub.IN. 
That is, during the upshift, the negative gradients of the actual input 
rotation speed N.sub.Pri and the estimated input rotation speed 
N.sub.Pri(i) are compared, and the larger one of them is selected as the 
representative inertia torque T.sub.IN to avoid the shortage of the line 
pressure P.sub.L. 
As shown in FIGS. 10A to 10D, at the moment t.sub.03 when the actual 
shifting has just been started, the negative gradient of the change speed 
dN.sub.E /dt of the engine rotation speed is still small. But the negative 
gradient of the change speed dN.sub.Pri /dt of the input rotation speed 
N.sub.Pri has already been large. Therefore, at the moment t.sub.03 the 
feedforward inertia torque T.sub.INR is greater than the feedback inertia 
torque T.sub.INP, and the actually generated inertia torque may become 
large. Accordingly, the feedforward inertia torque T.sub.INR is selected 
as the representative inertia torque T.sub.IN. 
On the other hand, at the moment t.sub.04 when the actual transmission 
ratio C.sub.P is not converged, the actual transmission ratio C.sub.P 
starts convergence. Therefore, the negative gradient of the change speed 
dN.sub.Pri /dt becomes small, and the negative gradient of the engine 
speed change speed dN.sub.E /dt is yet large. Therefore, the feedback 
inertia torque T.sub.INP becomes greater than the feedforward inertia 
torque T.sub.IN. Even if the shifting is not converged (starts to 
converge), the actually generated inertia torque is still large. 
Therefore, the feedback inertia torque T.sub.INP is selected as the 
representative inertia torque T.sub.IN. Accordingly, the representative 
inertia torque T.sub.IN takes a sufficiently large value to set the line 
pressure P.sub.L at a sufficient value satisfactorily preventing the 
slippage of the belt. 
The line pressure duty valve 120 constitutes a control valve of the CVT 
mechanism 29 according to the present invention. The engine revolution 
speed sensor 301 and the step S1 shown in FIG. 3 constitutes an internal 
combustion engine rotation condition detecting means. The step S3 of FIG. 
3 and the steps S801 and S802 of FIG. 7 constitute an internal combustion 
engine output detecting means. The steps S803 and S804 of FIG. 7 
constitute the actual rotational change condition detecting means. The 
steps S5 and S6 of FIG. 3 and the steps S806 and S807 of FIG. 7 constitute 
an estimated rotational change condition detecting means. The actual 
rotational change condition detecting means and the estimated rotational 
change condition detecting means constitute a rotational change condition 
detecting means. The step S805 of FIG. 7 constitutes an actual input load 
change amount calculating means. The step S808 of FIG. 7 constitutes an 
estimated input load change amount calculating means. The actual input 
load change amount calculating means and the estimated input load change 
amount calculating means constitute the input load change amount 
calculating means. The steps S809 to S811 of FIG. 7 constitute a selecting 
means. The steps S812 to S816 constitute a hydraulic fluid pressure 
control means. 
With the thus arranged control system according to the present invention, 
the input load change amount generated by the inertia is detected by 
detecting the rotational condition of the drive system. Further, the 
working fluid pressure applied to the CVT mechanism is controlled at a 
predetermined pressure according to a command signal outputted from the 
control unit to the CVT pressure control valve on the basis of the input 
load change amount and the output of the engine. This effectively 
functions to compensate the input load hange amount which tends to be 
short when the transmission ratio is decreased. Therefore, the fluid 
pressure supplied to the pulleys of the CVT mechanism is controlled at a 
proper value. 
Further, since the control system is arranged to directly detect the 
rotational condition of the drive system and to obtain the actual 
rotational change condition from the detected rotational condition and to 
calculating the actual input load change amount generated by the inertia 
by multiplying the actual rotational change condition detection value with 
a predetermined inertia coefficient, it is possible to accurately obtain 
the total input load change amount caused by the inertia during a period 
from the start of an actual shift to the convergence of the actual shift. 
This improves the control performance of the control system. 
Additionally, by detecting the estimated rotational change condition 
achieved from the transmission ratio control condition and by calculating 
the estimated input load change amount caused by the inertia by 
multiplying the estimated rotational change condition indicative value 
with the predetermined inertia coefficient, it is possible to accurately 
obtain the total input load change amount caused by the inertia during a 
period from the start of the shift control to the start of the actual 
shift. This similarly improves the control performance of the control 
system. 
Furthermore, the actual rotational change condition is detected by directly 
detecting the rotational condition of the drive system, and an estimated 
rotational change condition achieved from the transmission ratio control 
condition is calculated. The actual input load change amount caused by the 
inertia is calculated by multiplying the actual rotational change 
condition detection value by a predetermined inertial constant, and the 
estimated input load change amount is calculated by multiplying the 
estimated rotational change condition detection value by a predetermined 
inertia coefficient. Larger one of the actual input load change amount and 
the estimated input load change amount is selected as an input load change 
amount so as to further increase the input load to the CVT mechanism. 
Although the preferred embodiment according to the present invention has 
been shown and described to employ a normal selector lever and an 
inhibitor switch, it will be understood that a selector lever having a 
manual valve and an inhibitor switch disclosed in a Japanese Patent 
Provisional Publication No. 2-125174 may be employed instead of the normal 
lever and switch. This selector lever having the manual valve is arranged 
to enable the execution of the intended shift command (upshift command, 
downshift command) under the D-range. 
Although the preferred embodiment has been shown and described such that 
the control unit thereof is constituted by a microcomputer, it will be 
understood that a combination of electronic circuits such as calculating 
circuits may be employed in the control unit instead of the microcomputer. 
While the preferred embodiment has been shown and described such that the 
control system according to the present invention is applied to a belt 
type continuously variable transmission, it is of course that the control 
system according to the present invention may be applied to a toroidal 
type continuously variable transmission.