JPS6335870B2 - - Google Patents

Description

Detailed Description of the Invention (Technical Field) The present invention provides pressure change characteristics appropriate for each purpose when the hydraulic pressure of a common friction element is selectively controlled for different purposes in an automatic transmission. The present invention relates to a pressure control device for.

(Prior art) In recent years, in order to increase the product value of automatic transmissions,
Creep prevention devices, shift shock prevention devices, etc. are increasingly being installed, but all of these devices tend to be configured to control the hydraulic pressure of hydraulically operated friction elements to a predetermined value through duty control of a solenoid valve. It is becoming. When a plurality of such devices are provided, there is a possibility that a common friction element (its operating oil pressure) may be controlled for different purposes.

That is, in the forward travel range of the automatic transmission, the third speed is selected by hydraulic operation of the front clutch, but this front clutch is also a friction element that is operated in the reverse travel range. Therefore, in order to prevent creep while the vehicle is stopped in the reverse travel range, the hydraulic pressure of the front clutch must be adjusted to, for example, X in Figure 6.
It is necessary to control the front clutch transmission torque to obtain the front clutch transmission torque range shown in Figure 6, and in order to prevent shift shock when shifting from 2 to 3 in the forward travel range, the front clutch hydraulic pressure must be controlled, for example, at Y in Figure 6. It is necessary to perform control to obtain the front clutch transmission torque range shown in the figure. Therefore, these anti-creep and anti-shift shock devices each control a common friction element (front clutch) for different purposes.

On the other hand, since these two types of devices share a common control target and do not operate at the same time for the purpose of control, it is customary to have a common hardware configuration, which creeps the duty control of the solenoid valve. It was common practice to use the same drive frequency and drive voltage regardless of whether it was prevention or prevention of shift shock.

However, in this case, the change characteristics of the front flatch transmission torque with respect to changes in the duty ratio are fixed to, for example, those shown by solid lines or dotted lines in FIG. On the other hand, there is a demand for this characteristic to be moderate in the creep prevention control range X and the shift shock prevention control range Y.
Otherwise, the amount of deviation in the front clutch transmission torque with respect to the deviation in the duty ratio will increase, resulting in control hunting not only in the case of feedback control but also in the case of feedback control. However, in the case of the solid line characteristic in FIG. 6, although the characteristic becomes gentle in the control range Y, it is insufficiently gentle in the control range X, and in the case of the dotted line characteristic in FIG.
Although the characteristics become gentle in the control range X,
As mentioned above, when the characteristics become steep in the control range Y, it is impossible to have a fixed characteristic and to make the characteristics in different control ranges sufficiently gentle.

In order to solve this problem, it is conceivable to install the hardware of each device individually instead of making them common, but this would be very disadvantageous in terms of cost and installation space.

(Objective of the Invention) The present invention attempts to solve the above-mentioned problems by changing the above-mentioned characteristics to have just the right slope in each control range for each purpose of control. In an automatic transmission equipped with a pressure control device that uses common hardware to control pressure to a predetermined value required for each control purpose, the above characteristics in each hydraulic range have just the right slope for each control purpose. It is an object of the present invention to enable the characteristics to be changed so that stable control that does not cause hunting can be obtained for both types of control.

(Structure of the Invention) For this purpose, the present invention, as conceptually shown in FIG. The automatic transmission is provided in an automatic transmission configured to individually achieve the following, and the operating hydraulic pressure of the friction element is set to a predetermined voltage during the energization time width within the set cycle determined by the energization time width determining means in any of the oil pressure ranges. In the pressure control device, the pressure is set to a predetermined value by an electromagnetic valve that is activated by the application of pressure, and includes a hydraulic pressure region determining means for determining the hydraulic pressure region to be controlled, and a hydraulic pressure region discriminating means for determining the hydraulic pressure region to be controlled, and the setting period or the hydraulic pressure region determined by the means. The present invention is characterized in that it is provided with a pressure change characteristic changing means for changing a predetermined voltage to change a pressure change characteristic with respect to a proportion of the energization time width occupied in a set cycle.

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

Figure 2 shows the progress 3 of applying the pressure control device of the present invention.
The power transmission portion of a lock-up automatic transmission with one speed and one reverse speed is schematically shown.

The power transmission part in Figure 2 is the prime mover (engine).
a crankshaft 4, a lockup torque converter 1 with a lockup mechanism 17, an input shaft 7, a front clutch 104,
Rear clutch 105, second brake 10
6, low reverse brake 107, one-way brake 108, intermediate shaft 109, first planetary gear group 110, second planetary gear group 111, output shaft (transmission output shaft) 112, first governor valve 113, first It is composed of two governor valves 114 and an oil pump 13. The torque converter 1 consists of a pump impeller 3, a turbine impeller 8, and a stator impeller 9. The pump impeller 3 is driven by a crankshaft 4 and rotates the torque converter hydraulic oil contained therein to the input shaft 7. Torque is applied to the turbine wheel 8 fixed to the The torque is further transmitted by the input shaft 7 to the transmission gear train. The stator wheel 9 is placed on a fixed sleeve 12 via a one-way clutch 10. The one-way clutch 10 has a structure that allows the stator wheel 9 to rotate in the same direction as the crankshaft 4, that is, in the direction of the arrow (hereinafter referred to as normal rotation), but does not allow rotation in the opposite direction (hereinafter referred to as reverse rotation). ing. The first planetary gear group 110 includes an internal gear 117 fixed to the intermediate shaft 109 and a sun gear 11 fixed to the hollow conduction shaft 118.
9. It can rotate and revolve simultaneously while meshing with the internal gear 117 and the sun gear 119.
The planetary gear 120 is fixed to the output shaft 112 and consists of more than one small gear.
Consisting of a planetary gear support 121 that supports
The second planetary gear group 111 is the output shaft 1
12, a sun gear 123 fixed to the hollow conduction shaft 118, and a planet consisting of two or more small gears that can rotate and revolve simultaneously while meshing with each of the internal gear 122 and the sun gear 123. It is composed of a gear 124 and a planetary gear support 125 that supports the planetary gear 124.
The front clutch 104 connects the input shaft 7 driven by the turbine impeller 8 and the hollow transmission shaft 118 which rotates together with both sun gears 119 and 123 via a drum 126, and the rear clutch 105 intermediate shaft 109
It serves to connect the input shaft 7 and the internal gear 117 of the first planetary gear group 110 via the input shaft 7. The second brake 106 fixes both sun gears 119 and 123 by winding and tightening a drum 126 fixed to a hollow conductive shaft 118, and the low reverse brake 107 fixes both sun gears 119 and 123 to the planetary gear support of the second planetary gear group 111. It works to fix 125. The one-way brake 108 has a structure that allows the planetary gear support 125 to rotate in the normal direction, but not in the reverse direction. The first governor valve 113 and the second governor valve 114 are connected to the output shaft 11.
2 and generates governor pressure according to vehicle speed.

Next, a description will be given of the power transmission train when the gear shift lever is set to the D (forward automatic shifting) position.

In this case, only the rear clutch 105, which is the forward input clutch, is engaged first. Power from the engine via the torque converter 1 is transmitted from the input shaft 7 to the internal gear 117 of the first planetary gear group 110 through the rear clutch 105. The internal gear 117 rotates the planetary gear 120 in the normal direction. Therefore, the sun gear 119 rotates in reverse, and the planet gears 124 of the second planet gear group 111 rotate in the normal direction to reverse the sun gear 123 of the second planet gear group 111, which rotates integrally with the sun gear 119. One-way brake 108 prevents sun gear 123 from reversing planetary gear support 125 and acts as a forward reaction brake. Therefore, the internal gear 122 of the second planetary gear group 111 rotates normally. Therefore, the output shaft 112, which rotates integrally with the internal gear 122, also rotates in the normal direction, and the reduction ratio of the first forward speed is obtained. In this state, the vehicle speed increases and the second
When the brake 106 is engaged, the engine is moved from the imbutshaft 7 to the rear clutch 1 as in the first gear.
05 is transmitted to the internal gear 117. The second brake 106 fixes the drum 126, prevents rotation of the sun gear 119, and acts as a forward reaction brake. For this reason, the planetary gear 120 revolves around the stationary sun gear 119 while rotating, and therefore the planetary gear support 121 and the output shaft 112 integrated therewith
Although it is decelerated, it rotates normally at a faster speed than in the case of the first speed, and the reduction ratio of the second forward speed is obtained.
When the vehicle speed increases further and the second brake 106 is released and the front clutch 104 is engaged, the power transmitted to the input shaft 7 is
One side passes through the rear clutch 105 to the internal gear 11.
7 and the other is transmitted to the sun gear 119 via the front clutch 104. Therefore, the internal gear 117 and the sun gear 119 are interlocked, and together with the planetary gear support 121 and the output shaft 112, they all rotate normally at the same rotational speed to obtain the third speed. In this case, the input clutches are the front clutch 104 and the rear clutch 105, and since no torque is increased by the planetary gear, neither reaction brake is activated.

Next, the power transmission train when the speed selection rod is set to the R (reverse travel) position will be explained.

In this case, the front clutch 104 and the low
Reverse brake 107 is engaged. The power from the engine passes through the torque converter 1 and is transferred from the input shaft 7 to the front clutch 104.
Sun gears 119, 123 through drum 126
be guided by. At this time, since the rear planet carrier 125 is fixed by the low reverse brake 107, the sun gears 119, 1
The internal gear 122 is decelerated and reversed by the forward rotation of 23, and a backward reduction ratio is obtained from the output shaft 112 that rotates integrally with this internal gear.

FIG. 3 shows the hydraulic system of the speed change control device related to the automatic transmission, together with the device of the present invention, including the oil pump 13, line pressure regulating valve 128, pressure increase valve 129, torque converter 1, and speed selection valve 13.
0, first governor valve 113, second governor valve 11
4, 1-2 shaft valve 131, 2-3 shaft valve 132, throttle pressure reducing valve 133, cut-down valve 134, second lock valve 135, 2-3
Timing valve 136, solenoid down shift valve 137, throttle back up valve 13
8, vacuum throttle valve 139, vacuum diaphragm 140, front clutch 10
4, consisting of a rear clutch 105, a second brake 106, a servo 141, a low reverse brake 107, and a hydraulic circuit network. The oil pump 13 is driven by the prime mover via the crankshaft 4 and the pump wheel 3 of the torque converter 1, and is always in the reservoir 142 during engine operation.
The oil from which harmful dust has been removed is sucked up through a strainer 143 and sent to a line pressure circuit 144.

The oil is regulated to a predetermined pressure by a line pressure regulating valve 128 and sent to the torque converter 1 and speed selection valve 130 as working oil pressure. The line pressure regulating valve 128 consists of a spool 172 and a spring 173. In addition to the spring 173, the throttle pressure of the circuit 165 and the line pressure of the circuit 156 act on the spool 172 through the spool 174 of the pressure increase valve 129. The resulting force opposes line pressure acting above spool 172 through orifice 175 from circuit 144 and pressure acting from circuit 176. The operating oil pressure of the torque converter 1 is supplied from the circuit 144 to the line pressure regulating valve 128.
The oil introduced into the circuit 145 via the hydraulic oil inflow passage 50 flows into the torque converter 1 and is then removed via the hydraulic oil outflow passage 51 and the pressure holding valve 146. is maintained. If the pressure exceeds a certain pressure, the pressure retention valve 14
6 is opened and oil is further sent from circuit 147 to the rear lubrication section of the power transmission mechanism. When this lubricating oil pressure is too high, the relief valve 148 opens and the pressure is lowered. On the other hand, lubricating oil is supplied from the circuit 145 to the front lubricating section of the power transmission mechanism by opening the front lubricating valve 149. The speed selection valve 130 is a manually operated fluid direction switching valve, and is composed of a spool 150, which is connected to a speed selection rod (not shown) via a linkage, and the spool 150 is actuated by each speed selection operation to control the line. This is for switching the pressure feeding passage of the pressure circuit 144.
The state shown in FIG. 3 is the N (neutral) position, where the line pressure circuit 144 is connected to ports d and e.
is open to The first governor valve 113 and the second
1-2 shaft valves 131 and 2 due to the governor pressure generated in the governor valve 114 during forward travel.
-3 It operates the shift valve 132 to perform an automatic gear change action and also controls the line pressure, and the speed selection valve 1
30 is in each position D, and, the oil pressure is transferred from the line pressure circuit 144 to the port C of the speed selection valve 130.
When the vehicle is running, the governor pressure regulated by the second governor valve 114 is sent to the circuit 157 and introduced into the first governor valve 113, and when the vehicle speed reaches a certain speed, the governor pressure is regulated by the second governor valve 114. The spool 177 of the governor valve 113 moves and the circuit 1
57 is connected to circuit 158 to generate governor pressure, and from circuit 158 governor pressure is transferred to 1-2 shift valve 13.
It is counterbalanced by respective springs acting on each end face of the 1, 2-3 shift valve 132 and the cut-down valve 134, urging each of these valves to the right. In addition, the 1-2 shift valve 131 and the second brake are connected in the middle of the hydraulic circuit from the port C of the speed selection valve 130 through the circuit 153, the circuit 161, and the circuit 162 to the engagement side hydraulic chamber 169 of the servo 141 that tightens the second brake 106. A lock valve 135 is provided separately, and a circuit 152 extending from port b of the speed selection valve 130 to the second lock valve 135 is provided.

Therefore, when the speed selection rod is set to the D position, the spool 150 of the speed selection valve 130 moves and the line pressure circuit 1
44 leads to ports a, b, and c. The hydraulic pressure passes through the circuit 151 from port a, and part of it acts on the lower part of the second lock valve 135, and the spring 17
The spool 178, which is pressed upward by the spool 9, is lowered by the hydraulic pressure acting from the port b through the circuit 152, so that the conductive circuits 161 and 162 are not cut off, and a part of the spool 178 is pressed upward by the orifice 166. From circuit 167 to 2-
3 shift valve 132, and from port c, circuit 1
53 and reaches the second governor valve 114, rear clutch 105 and 1-2 shift valve 131, and the transmission is in the first forward speed. In this state, when the vehicle speed reaches a certain speed, the spool 160 of the 1-2 shift valve 131, which is pressed to the right by the spring 159, moves to the left due to the governor pressure of the circuit 158, shifting from the first forward speed to the second forward speed. Automatic gear shifting is performed, and the circuit 153 and circuit 161 are brought into conduction, and the hydraulic pressure reaches the engagement-side hydraulic chamber 169 of the servo 141 via the second lock valve 135, and the second brake 106 is engaged. forward 2nd
Be in a state of speed. In this case, the 1-2 shift valve 13
1 is compact, the spool 160 moves to the left at the required speed without increasing the speed at the shift point, and an automatic shift action from forward first speed to second speed is performed. When the vehicle speed increases further and reaches a certain speed, the governor pressure of the circuit 158 overcomes the spring 163 and becomes 2-
3 Press the spool 164 of the shift valve 132 to the left to connect the circuit 167 and the circuit 168 so that the hydraulic pressure is partially transferred from the circuit 168 to the release side hydraulic chamber 1 of the servo 141.
70, the second brake 106 is released, a portion reaches and engages the front clutch 104, and the transmission is in third forward gear.

Furthermore, when the driver depresses the accelerator pedal until the throttle opening is close to full throttle while driving in the D position, desiring a large acceleration force, the kick-down switch is turned on and the solenoid down shift valve 137 is turned on. A down shift solenoid 137a disposed opposite to the down shift solenoid 137a is energized by energization. As a result, the spool 190 of the solenoid down shift valve 137 is pushed downward by the spring 191 from its locked position upward in FIG. At this time, the kickdown circuit 180 that was connected to the circuit 154 is connected to the line pressure circuit 144, and the line pressure is applied to the circuit 14.
4,180 and 1-2 shift valve 131 and 2
−3 is supplied to the shift valve 132 so as to oppose the governor pressure. If you are running in 3rd gear at this time,
First, the spool 164 of the 2-3 shift valve 132 is forcibly pushed from the leftward position to the rightward position by the line pressure against the governor pressure, and the spool 164 of the 2-3 shift valve 132 is forced to move from the 3rd gear to the 2nd gear within a certain vehicle speed limit. A forced downshift is performed to provide sufficient acceleration. By the way, when the above-mentioned kickdown is performed while running in the second gear, the load is large at this time and the speed is low, so the governor pressure is also low, so the line pressure led to the circuit 180 is transferred to the 1-2 shift valve 131. The spool 160 is also moved from the leftward position to the right against the governor pressure. Therefore, in this case, a forced downshift from second speed to first speed is performed, and a stronger acceleration force corresponding to a large load can be obtained.

When the speed selection rod is set to the (second forward speed fixed) position, the spool 150 of the speed selection valve 130 moves and the line pressure circuit 144 communicates with ports b, c, and d.
Oil pressure reaches the same locations from ports b and c as in case D, engaging the rear clutch 105, while the lower part of the second lock valve 135 is not receiving oil pressure in this case and the spool 178 circuit. Since the areas of the lands above and below the part that opens to 152 and applies hydraulic pressure are larger at the bottom, the spool 178 of the second lock valve 135 is pushed down against the force of the spring 179, and the circuits 152 and 16
2 becomes conductive, the hydraulic pressure reaches the engagement-side hydraulic chamber 169 of the servo 141, engages the second brake 106, and the transmission enters the second forward speed state. From port d, hydraulic pressure passes through circuit 154 to solenoid down.
Shift valve 137 and throttle back up valve 138 are reached. There is a disconnection between port a of the speed selection valve 130 and the line pressure circuit 144, and the hydraulic pressure has not reached the 2-3 shift valve 132 from the circuit 151, so the second brake 106 is released and the front clutch 104 is released. The engagement is not performed and the transmission is not in the third forward speed state, and the second lock valve 135 works in conjunction with the speed selection valve 130 to fix the transmission in the second forward speed state. . When the speed selection rod is set to the (first forward speed fixed) position, the line pressure circuit 144 is connected to ports c, d and e.
Leads to. Hydraulic pressure reaches the same location from ports c and d as in , engaging the rear clutch 105, and from port e via circuit 155 through the 1-2 shift valve 131 and partially from circuit 171 to the low
The low reverse brake 10 reaches the reverse brake 107 and acts as a forward reaction brake.
7, the transmission is in the first forward speed state, and a portion reaches the left side of the 1-2 shift valve 131 and the spring 1
59 to press the spool 160 to the right, and the first forward speed is fixed.

In addition, as shown in Fig. 2, the torque converter 1
A lockup mechanism 17 is provided inside the lockup mechanism 17, which is controlled by a lockup control device 100 comprising a lockup control valve 30 and a lockup solenoid 31 shown in FIG. 3, and these lockup mechanism 17 and lockup control device 100 are well known. However, since it is not related to the present invention, a description thereof will be omitted.

In the present invention, in the automatic transmission, R
Pressure control device 200 configured as a creep prevention device that prevents creep while the vehicle is stopped in the range, and a shift shock prevention device that prevents shift shock caused by 2→3 shift up in D range.
are provided as shown in Figure 3. In the R range, as is clear from the above, starting in reverse becomes possible by supplying the line pressure P _L from the oil passage 168 to the front clutch 104 as the front clutch pressure P _F/C , and in the D range, As is clear from the above description, the 2 to 3 shift is also achieved by supplying the line pressure P _L from the oil passage 168 to the front clutch 104 as the front clutch pressure P _F/C . Therefore, the above-mentioned creep prevention and shift shock prevention can be achieved by controlling the hydraulic pressure P _F/C supplied to the front clutch 104 from the oil passage 158 in a predetermined manner, and for these purposes, a pressure control device is used. Reference numeral 200 is provided in relation to the oil passage 168 to perform the above pressure control.

As clearly shown in FIG. 4, the pressure control device 200 includes a pressure control valve 201 inserted into the oil passage 168, and this valve supplies part of the line pressure P _L to the drain circuit 20.
2, the front clutch pressure P _F/C shall be adjusted. For this purpose, the pressure control valve 201 includes a spool 201a, one end of which is connected to the chamber 201b.
In addition, the other end is configured to face the chamber 201c, respectively,
The front clutch pressure P _F/C is introduced into the chamber 201b via a branch path 203, and the line pressure _PL is introduced into the chamber 201c via a branch path 204. An orifice 205 is provided in the branch passage 204, and the chamber 201c is also communicated with the drain circuit 202 through an orifice 206 having a larger opening area.A solenoid valve 207 is provided opposite to the orifice 206, and the solenoid valve is used to open the chamber 201c. The control pressure inside shall be set to P _S.

The electromagnetic valve 207 has a plunger 207b elastically supported by a spring 207a in the open position shown in the lower half of the figure.
The chamber 201c is cut off from the drain circuit 202 when the plunger is electromagnetically attracted to the valve closing position shown in the upper half of the figure by the energization of the solenoid 207c.

The solenoid valve 207 (solenoid 207c) is duty-controlled so as to be energized during the pulse width (on time) of the pulse signal from the computer 208 as shown in FIGS. 5a and 5b. As shown in FIG. 5a, when the duty (%) is small, the communication time between the chamber 201c and the drain circuit 202 is long, and therefore the control pressure P _S in the chamber 201c is as shown by the solid line or dotted line in FIG. It decreases as the duty decreases, and finally reaches the lowest value determined by the difference in the opening area of the orifices 205 and 206. On the other hand, when the duty (%) is large as shown in FIG. 5b, the chamber 201c and the drain circuit 20
The communication time with No. 2 is short, so the control pressure P _S is
As shown by the solid line or dotted line in the figure, it increases as the duty increases, and finally the source pressure, that is, the line pressure P _L
is set to the maximum value equal to .

On the other hand, as the control pressure P _S increases, the pressure control valve 201 moves the pool 201a to the upper half position in FIG. 4, and does not remove the line pressure P _L from the drain circuit 202. This reduces front clutch pressure.
P _F/C rises in response to replenishment of line pressure P _L , and this front clutch pressure P _F/C pushes spool 201a back to the lower half position in FIG. 4 in chamber 201b. At this time, the front clutch pressure P _F/C is supplemented by the line pressure P _L , but the amount is reduced and at the same time it is removed from the drain circuit 202. As a result, the front clutch pressure P _F/C decreases and this As a result, the force for pushing back the spool 201a is reduced, and the spool 201a again strokes toward the upper half position in FIG. 4. By repeating this action, the pressure control valve 201 adjusts the front clutch pressure P _F/C to the control pressure.
Keep P equal to _S.

As described above, the control pressure P _S changes as shown in Fig. 6 depending on the duty (%) of the solenoid valve solenoid 207c, so the front clutch pressure P _F/C
If the control pressure change characteristic is as shown in the solid line or dotted line in FIG. 6, the front clutch transmission torque will also have the change characteristic as shown in the solid line or dotted line in FIG.

By the way, the control pressure (P _S ) characteristics and, therefore, the front clutch pressure (P _F/C ) characteristics with respect to the duty (%) of the solenoid valve solenoid 107c are
When the driving frequency is low, it becomes as shown by the solid line in FIG. 6, and when this driving frequency is high, it becomes as shown by the dotted line in the same figure. The reason for this will be explained below. As shown in FIG. 12, the electromagnetic valve 207 cannot avoid dead time t ₁ and delay time t ₂ in response to energization, and as shown in FIG. 12 a, b, and c, the driving frequency When the driving frequency is high, as shown in d, e, and f in the figure, these are more affected by the dead time _t1 and the delay time _t2 than when the driving frequency is low.
In other words, as is clear from the comparisons of a and d, b and e, and c and f in FIG. The width becomes smaller and the solenoid valve 207 closes to the orifice 2 in the duty region.
The extent to which it is not possible to fully close 06 is greater when the drive frequency is high than when it is low. Therefore, even with the same duty ratio, the amount of drain from the orifice 206 (the shaded area in FIG. 12 corresponds to this amount of drain) is greater when the drive frequency is high than when it is low, and the control pressure P _S When the driving frequency is high, the control pressure P S changes as shown by the dotted line in FIG. 6, and when the driving frequency is low, as shown by the solid line in the same figure, the control pressure P _S is generally higher than the dotted line characteristic.

The computer 208 for controlling the duty of the solenoid valve 207 is driven by the power supply +V as shown in FIG _.
Signal S _P from pressure sensor 209 that detects engine throttle opening TH, Signal S TH from throttle opening sensor 210 that detects engine throttle opening _TH , Engine rotation speed
The signal S _ir from the engine rotation speed sensor 211 that detects N _E , the output rotation speed of the torque converter 1 (the rotation speed of the turbine blade wheel 8 that transmits rotation to the input shaft 7), the torque converter output rotation speed that detects N _T A transmission output rotation speed sensor 2 that detects the signal S _tr from the sensor 212 and the transmission output rotation speed (rotation speed of the output shaft 112) N _p
13, a signal S _S from a shift switch 214 that detects the gear position (gear position), presence or absence of shifting, and type of shifting of the automatic transmission, _and an inhibitor switch 2 that detects the selected range of the automatic transmission.
15, a signal S _T from the idle switch _{216 that detects the release of the accelerator pedal, and a signal S C from} the temperature sensor 217 that detects the hydraulic oil temperature T of the automatic transmission. 2
07 (solenoid 207c). In addition, the shift switch 214
For example, as shown in Japanese Unexamined Patent Publication No. 56-127856, it is possible to use one constructed by being incorporated into shift valves 131 and 132.

The control computer 208 includes, for example, a microprocessor unit (MPU) 24 including random access memory (RAM) as shown in FIG.
A read-only memory (ROM) 25, an input/output interface circuit (I/O) 26, and an analog . Digital (A/D) transmission 27 and waveform shaping circuit 2
8 and an amplifier 29, and executes the control programs shown in FIGS. 8 to 11.

FIG. 8 shows the main routine. When the engine ignition switch is turned on at block 70, the computer 208 starts operating, and at the next block 71, the initial values of the MPU 24 and I/O 26 are set (initialized). It is done. Control then passes to block 72 where MPU 24
is after converting the throttle opening signal _STH from the throttle opening sensor 210 into a digital signal by the A/D converter 27 (however, in this example, the period from throttle fully closed to fully open is divided into 8 parts and the digital signal is converted into a digital signal. At the same time as reading the throttle opening TH, the idle signal S _I from the idle switch 216 is read. In the next block 73, the MPU 24 converts the signal S _P from the pressure sensor 209 to the A/D converter 2.
7, it is converted into a digital signal and then read through the I/O 26 to read the front clutch pressure P _F/C .

Control then passes to block 74 where the MPU
24 is a signal from the engine rotation speed sensor 211
Based on S _ir , the interrupt routine shown in FIG. 9a is executed as follows to calculate the engine rotational speed N _E. The sensor 211 detects the ignition signal of the engine and emits a signal S _ir as shown in FIG. 9b, and this signal is noise-removed by the waveform shaping circuit 28 and becomes the ignition signal as shown in FIG. 9b. A rectangular wave signal S _ir ' rises with each input. Then, the MPU 24 receives the signal
Each time S _{ir '} rises, the interrupt routine shown in _FIG . The signal period T _E is measured from the time difference, and the MPU 24 can calculate the engine rotation speed N _E from this period T _E . Control then proceeds to block 42, which returns to the main routine of FIG.

In the next block 75 in FIG.
After converting the signal S _C from 17 into a digital signal by the A/D converter 27, it is read through the I/O 26,
Read the automatic transmission hydraulic oil temperature T. In the next block 76, the engine speed change ΔN E during one calculation cycle is calculated from the engine speed N _E and the previous engine speed N _E (OLD) by _{ΔN E =} N _E (OLD) -
Determine by calculating N _E. Since the calculation cycle is constant, this engine rotational speed change ΔN _E can be regarded as the time change rate of the engine rotational speed. In the next block 77, the MPU 24 receives the signal S _tr from the sensor 212.
Based on this, the interrupt routine shown in FIG. 10a is executed as follows to calculate the output rotational speed N _T of the torque converter 1. The sensor 212 is connected to the input shaft 7
A sine waveform generator is mounted around the sine waveform generator and outputs a signal _Str shown in FIG. The waveform shaper converts the signal into a rectangular wave signal S _tr ' shown in FIG. 10b. Then _, the MPU ₂₄ starts the interrupt routine shown in FIG.
In the next block 51, the signal period T _T is measured from the time difference with the previous signal S _tr ', and the MPU 24 can calculate the output rotational speed N _T of the torque converter 1 based on this period. The control then goes to block 52.
The program then returns to the main routine shown in FIG.

In the next block 78 in FIG. 8, the MPU 24 calculates the torque converter output rotation speed N _T determined as described above.
is divided into an appropriate number of blocks, quantized, and then the next block 79
Then, the MPU 24 calculates the output rotation speed N _p of the automatic transmission based on the signal S _pr from the sensor 213. Sensor 213 is similar to sensor 212 and is attached to output shaft 112. Therefore, the transmission output rotation speed N _p is also determined by the MPU 24 executing an interrupt routine similar to that shown in FIG. 10a.
It can be calculated in the same way as the torque converter output rotational speed N _T was determined in block 77.

In the next block 80, it is determined whether the shift determiner _S1 is set to 1 or not. In this example, this shift determiner is designed to reduce shift shock 2→
3. This indicates that an upshift is in progress; S ₁ =1 indicates that the shift is in progress; S ₁ =0 indicates that another shift is in progress or not. If S ₁ =1 during a 2->3 shift up, the control returns to block 72 and repeats each of the execution blocks described above, but if S≠1 and the shift up is not in progress, the control proceeds to blocks 81 to 83 in sequence. Block 81 reads the shift signal S _S from the shift switch 214, and block 82 reads the shift signal S S from the shift switch 214.
The selection range of the automatic transmission is read from the signal S _L , and in block 83 it is determined whether there is a 2→3 shift command from the shift signal S _S , that is, the 2→3 shift valve 132 selects the third speed. It is determined whether or not the shift-up position has been shifted to the shift-up position.

If there is no 2→3 shift command, block 7
If there is a 2→3 shift command, control proceeds to block 84 where S ₁ is set to 1 to indicate that a 2→3 shift is in progress. The next block 85 determines the timing for starting the 2→3 shift prevention function in this example, that is, when the front clutch 104
The change in engine speed (ΔN _E ) _p during one calculation cycle is read from the ROM 25 in order to determine when the engine starts to engage (speed change start time). The reason why the shift start timing can be determined based on this (ΔN _E ) _p is that when the front clutch 104 starts to engage and the 2nd to 3rd shift starts, the engine speed N _E changes suddenly, and after the shift command, the engine speed changes. When the rotational speed change ΔN _E becomes equal to or greater than (ΔN _E ) _p , it can be considered that the shift starts. Note that after execution of block 85, control returns to block 72, and the above-described loop is repeated.

FIG. 11 shows an interrupt routine for carrying out the anti-freep control and the 2→3 shift prevention control while the vehicle is stopped in the R range, which is the objective of this example, based on the execution results of the main routine shown in FIG. 8. The routine is repeatedly executed in block 300 by an interrupt signal input from a timer (not shown) at set time intervals ΔT _ns , and is executed by the creep prevention control block group 301, the shift shock prevention control block group 302, and the block groups of these blocks. It is roughly divided into a common duty output block group 303 that outputs calculation results.

First, in block 304, it is determined whether the selected range of the automatic transmission is the reverse (R) range. If it is not in the R range, it is determined in block 305 whether it is in the neutral (N) or parking (P) range, and if it is in the N or P range, it is determined in block 305.
In step 06, the torque converter input/output rotational speed difference ΔN' during the range, that is, in the neutral state where the automatic transmission is unable to transmit power, is determined by calculating ΔN'=N _{E -NT} . Then, in the next block 307, the output duty (Duty) is set to 0.
%, and block 308 returns control to the main routine of FIG. At this output duty of 0%, the computer 208 operates the solenoid valve 207 (solenoid 20
7c) to keep the orifice 206 open to keep the control pressure P _S and therefore the front clutch pressure P _F/C at a minimum value as seen in FIG. Therefore, the front clutch 104 is kept inactive and does not transmit power, allowing the vehicle to be parked in this N or P range.

When the block 304 determines that the range is R, the block 309 is selected, and the drive frequency of the solenoid valve 207 is set to the control pressure (P _S ) as shown by the dotted line in FIG.
Switch to a high frequency to obtain the desired characteristics and front clutch pressure (P _F/C ) characteristics. Control then passes to block 310 where the transmission output speed is
It is determined whether N _p (vehicle speed) is greater than or equal to the minute setting value N _lin . If so, since the vehicle is running normally in the R range, the control proceeds to block 311 and then returns to the main routine of FIG. 8 in block 308. In block 311, the output duty (Duty)
is set to 100%, so that the solenoid valve solenoid 207c continues to be energized via the amplifier 29, and the drain orifice 206 is normally closed. Thus the control pressure P _S
As is clear from Fig. 6, the front clutch pressure P _F/C is set to the maximum value, and the front clutch pressure 10
4 is kept in a fully engaged state, all of the engine power can be transmitted to the output shaft 112 by this front clutch, allowing the vehicle to travel normally backwards.

If N _p ≧ N _lin , that is, if the vehicle is stopped in the R range, block 310 is set to block 3.
12 is selected, and it is determined from the idle signal S _I whether or not the engine is idling with the idle switch 216 turned on. If the vehicle is idling, there is no intention to start in reverse, so from block 313, creep prevention is performed while the vehicle is stopped in the R range as follows. That is, in block 313, the upper limit value (control target value) ΔN _ref required to prevent the torque converter input/output rotational speed difference from creeping is read from the ROM 25. In the next block 314, the correction amount α corresponding to the neutral torque converter input/output rotational speed difference ΔN' obtained in block 306 is read from the ROM 25, and in the next block 314', the correction amount α is added to the control target value ΔN _ref. The corrected control target value ΔN′ _ref is determined by addition. In the next block 315, the torque converter input/output rotational speed difference ΔN _real is calculated as N _E −N _T
In the next block 316, the deviation ΔN ₁ of the actual rotational difference ΔN _real from the target value ΔN′ _ref is calculated as ΔN ₁
Calculate by calculating = ΔN _real −ΔN′ _ref .

In the next block 317, it is determined whether the deviation _ΔN1 is positive or not. If ΔN ₁ > 0, ΔN _real > ΔN′ _ref
Therefore, proceed to block 318 and select Duty (NEW).
= Duty (DLD) + K ₁ · ΔN ₁ is calculated in the direction of decreasing the output duty. If ΔN ₁ > 0, then ΔN _real <ΔN' _ref , so proceed to block 319 and set Duty (NEW) = Duty (OLD). -K ₁ ·ΔN ₁ is calculated in the direction of increasing the output duty. here
Duty (NEW) is the duty that should be updated.
(OLD) indicates the previous duty, and K ₁ indicates the control constant.

Therefore, when ΔN ₁ >0, as the duty increases, the control pressure P _S and the front clutch pressure P _F/C increase, as is clear from FIG. 6, and the front clutch 10
4, thereby reducing the torque converter input/output rotational speed difference ΔN _real to the target value ΔN′ _ref
When ΔN ₁ is not > 0, the control pressure P _S and front clutch pressure P _F/C decrease as shown in FIG.
As a result, the torque converter input/output rotational speed difference ΔN _real can be increased and brought closer to the target value ΔN′ _ref .

From block 318 or 319, control proceeds to block 320, where the launch control start determiner _S3 is reset to zero. It is assumed that this determiner is set to 1 after start control, which will be described later, is started.

After that, the control passes through blocks 321 and 322, reaches block 308, and returns to the main routine of FIG.
(NEW) and set Duty in block 322.
By setting (NEW) as the output duty, the creep prevention control is executed to bring the torque converter input/output rotational speed difference ΔN _real closer to the target value ΔN′ _ref as described above. By repeating such creep prevention control, the front clutch 104 is controlled so that the torque converter input/output rotational speed difference ΔN _real is maintained at a predetermined value (control target value) ΔN' _ref necessary for creep prevention, and the automatic transmission Front clutch 10 creep in R range
This can be prevented by keeping the sliding connection state of 4.

By the way, during such creep prevention, block 3
Since the solenoid valve driving frequency is switched to a high frequency in 09, the control pressure (P _S ) characteristics and front clutch pressure (P _F/C ) characteristics with respect to duty are as shown by the dotted lines in Fig. 6, respectively, under creep prevention control. It becomes gentle in the range X, and the control can be stably performed during the creep prevention without causing hunting.

Next, when the driver depresses the accelerator pedal in order to start the vehicle, the idle switch 21
6 is turned OFF, the corresponding signal S _I
Based on this, block 312 selects block 323, and reverse start control is executed as follows.

First, in block 323, it is determined whether the start control start determiner _S3 is 0, that is, whether creep prevention control was performed previously. If so, control proceeds to block 324, where the solenoid valve drive frequency is set to a control pressure (P _S ) as shown by the solid line in FIG.
Switch to a low frequency to obtain the characteristics and front clutch pressure (P _F/C ) characteristics. Next block 3
25, the first target torque converter output rotation speed N _ref that should occur at the start of the relevant start is stored in the ROM 25.
Read from. In the next block 326, the torque converter output rotational speed target change rate for rear start is determined.
dNT/dt is read from the ROM 25, and the target rate of change is set to be as large as possible without causing a start shock. Next, in block 327, the determiner _S3 is set to 1 to indicate that the start control has started.

Therefore, after that, block 323 becomes block 3.
28, and in this block 2
Target torque converter output rotation speed N _ref after the first time
(NEW) is calculated using the following formula.

N _ref (NEW) = N _ref (OLD) − ΔN _ns・(dN _T /dt) Here, N _ref (NEW) is the target torque converter output rotation speed that should be updated this time, and N _ref (OLD) is the previous (previously If it is the first time, N _ref ) is the target torque converter output rotation speed, and ΔT _ns respectively indicate the time interval at which the interrupt routine of FIG. 11 is executed. Thus, the target torque converter output rotation speed decreases from N _ref by ΔT _ns ·(dN _T /dt) rotations every ΔT _ns .

In block 329, block 325 (first time)
Or the torque converter output rotation speed N _T for the target value N _ref found in 328 (second time onwards)
The deviation ΔN ₂ of ΔN 2 is calculated by ΔN ₂ =N _T −N _ref , and it is determined in the next block 330 whether or not the result is positive. If ΔN ₂ >0, since N _T >N _ref , Duty (NEW) is set in block 331.
=Duty(OLD)+K ₂・ΔN ₂ Calculation is performed in the direction of increasing the output duty, and if ΔN ₂ >0,
Since N _T <N _ref , in block 332, a calculation is performed in the direction of decreasing the output duty: Duty (NEW) = Duty (OLD) - K ₂ ·ΔN ₂ . In addition,
K ₂ is a control constant.

Therefore, when ΔN ₂ >0, as the duty increases, the control pressure P _S and the front clutch pressure P _F/C increase as shown in FIG.
By increasing the fastening force of 04, it is possible to lower the torque converter output rotational speed N _T and bring it closer to the target value N _ref . Also, when ΔN ₂ > 0, the duty is reduced and the control is performed as shown in Fig. 6. The pressure P _S and the front clutch pressure P _F/C decrease to weaken the engagement force of the front clutch 104, thereby increasing the torque converter output rotational speed N _T to approach the target value N _ref .

This control is accomplished by transitioning from block 331 or 332 to block 321, 322,
By repeating such start control, the torque converter output rotational speed N _T is changed (decreased) in accordance with the target rate of change dN _T /dt, and the start shock is smoothly avoided and unnecessary front clutch 1 is applied.
The vehicle can be quickly started backwards without causing slippage.

In addition, the creep prevention mentioned above is performed by the front clutch 1.
04 is achieved with the clutch still in the slidingly engaged state, there is no loss stroke when the front clutch 104 transitions to the fully engaged state at the time of the start, and there is no time delay in this clutch engagement, and the engine does not start revving. The rotational inertia does not cause a starting shock.

By the way, during this start control, the solenoid valve drive frequency is switched to a low frequency in block 324, so the control pressure (P _S ) and front clutch pressure (P _F/C ) characteristics with respect to duty are shown in FIG. 6, respectively. As shown by the solid line, the start control is gentle throughout the area Z where the control should be performed particularly smoothly, and the start control can be performed stably in this area Z without hunting.

Note that in block 305 in FIG.
If it is determined that it is not in the range, it will move forward (D,
or) Since it is a range, block the control 3
The program proceeds to step 33, where 2->3 shift shock prevention control, which will be described next, is executed. For this control, first, in block 333, the solenoid valve driving frequency is switched to a low frequency so as to obtain the control pressure (P _S ) characteristics and front clutch pressure (P _F/C ) characteristics shown by the solid line in FIG. . In the next block 334, depending on whether the shift determiner _S1 is 1 or not, it is determined whether or not the present example is in the middle of a 2->3 shift, which is targeted for shift shock prevention. If so, control passes to block 335;
Gear ratio G (1 in this case) corresponding to the gear position (3rd gear in this case) determined from the shift signal S _S
is read from POM25, and this and block 79
The torque converter output rotation speed G×N _p when the front clutch 104 is completely engaged is calculated by multiplying it by the transmission output rotation speed N ₀ obtained in block 77, and this is calculated as the actual torque converter output rotation speed obtained in block 77. By subtracting it from N _T , that is, by calculating N _T -G·N _p , a shift end determination factor (completion of engagement of the front clutch 104) F is obtained.

If the front clutch 104 is firmly engaged after the shift has been completed, F=0, but if the front clutch 104 is not yet fully engaged during the shift, F≠O. Furthermore, in an upshift in which the gear ratio G changes from large to small during a gear change, F>O, and in a downshift in which the gear ratio G changes from small to large, F>O.

In block 336, it is determined whether the shift end determiner F is positive or negative, and the process proceeds to block 337 only when F>O, that is, an upshift.
Since this is a 3-speed shift, block 3 is only possible during this shift.
37 is selected. When this shift is completed, F=O
However, due to a slight difference in the detection time of the torque converter output rotation speed N _T and the transmission output rotation speed N _P and a calculation error, there is a possibility that F=O does not completely hold. Therefore, in block 337, the absolute value |F| of the shift end determiner F is set to the set minute value ε ₁
If S 1 =0, it is determined that the shift has been completed, and control proceeds to block 338, where S ₁ =0.
Reset to . On the other hand, if |F|>ε ₁ , that is, if the gear is being changed, control proceeds to block 339.

At block 339, front clutch 104
It is determined whether or not the shift start determiner S ₂ is 1, which is 1 when the 2→3 shift has already been started by the start of engagement, and is 0 when this shift has not yet started. If it is determined that S ₂ ≠ 1, block 340
It is determined whether or not a new shift from 2 to 3 has started. In making this determination, the change in engine speed ΔN _E obtained at block 76 in FIG.
exceeds the change in engine rotational speed (ΔN _E ) _p determined at block 85 to start the shift, so it is determined that the 2->3 shift has started due to the start of engagement of the front clutch 104.

When gear shifting has started, block 340 selects block 341, reads the front clutch pressure P ₁ (P _F/C read in block 73) at this time, and
In the next block 342, the front clutch pressure (front clutch 1
Correction value P _p for the above pressure P ₁ of the fastening force of 04
is read from ROM25. This correction value P _p is stored in the ROM as table data of the type of shift (in this case, 2→3 shift) determined by reading the shift signal S _S and the throttle opening TH read as described above.
25 and read out using the table lookup method. Note that in this example, the correction value P _p
is fixed at the read value, but it is also possible to gradually increase it as the shift progresses as long as a shift shock does not occur.

In the next block 343, the target value P _ain of the front clutch pressure that should be maintained to prevent gear shift shock _{is determined.}
=P ₁ +P ₀ , and then in block 344 the shift start determiner S ₂ is set to S ₂ =1. Next, control proceeds to block 345, but once S ₂ =1 is set, block 339 selects block 345, so blocks 340 to 344 are skipped and not selected until the next 2->3 shift.

In block 345, the difference ΔP between the actual value P _F/C of the front clutch pressure and the target pressure P _ain is calculated, and in the next block 346, it is determined whether ΔP>0 or not, and ΔP>
In the case of 0, that is, in the case of P _F/C > _Pain , the control proceeds to block 347, where a calculation is performed in the direction of duty reduction as Duty (NEW) = Duty (OLD) - K・ΔP, and the calculation result is Duty (NEW). in the next block 321
Replace with Duty (OLD). And block 3
22, the output duty is set to Duty (NEW), and the solenoid valve 207 is output via the amplifier 29 in FIG.
(Solenoid 207C). In addition, the Duty
(NEW) is the output duty that should be newly updated.
Duty (OLD) is the current output duty, and the feedback coefficient K is a constant value. Of course, it is also possible to make the feedback coefficient K a function of the pressure difference ΔN. As shown in Fig. 6, the front clutch pressure P _F/C becomes lower as the output duty decreases, so P _F/C > P _ain is corrected to bring the front clutch pressure P _F/C to the target pressure P _ain . You can approach it.

On the other hand, if the determination result of block 346 is ΔP≠0, that is, if P _F/C < _Pain , control proceeds from block 346 to block 348. block 34
In 8, the calculation in the direction of increasing the output duty is performed as Duty (NEW) = Duty (OLD) + K・ΔP, and the calculation result Duty (NEW) is applied to the next block 321,
322, and is output to the solenoid valve 207 as in the case of ΔP>0 described above. In this case as well, as shown in Fig. 6, the front clutch pressure P F/C increases as the output duty increases, and the front clutch pressure P _F/C is increased by correcting P F _{/ C} < P _ain . Target pressure P _ain
can be approached.

Thus, the shift start determiner _S2 is set to 1 from 2 to 3.
The front clutch pressure P _F/C remains at the target value from the start of the shift until the end of the shift when the shift determiner _S1 is set to 0.
The torque transmission capacity of the front clutch 104 can be kept _constant . Therefore, 2
→When shifting to 3rd gear, the combined rotational force of the engine's inertia and driving force is transmitted to the transmission output shaft at a constant rate, causing the engine rotational speed N _E to gradually change with respect to the transmission output rotational speed N _p . At the same time, the transmission output torque can be made to have no large peak torque, and a 2→3 shift shock can be prevented.

In determining the target value P _ain , the front clutch pressure P ₁ at the start of gear shifting is used as a reference, and the front clutch pressure P ₁ required to reduce the engine speed N In order to obtain the torque transmission capacity of the clutch 104, the target value P _ain was set as the value obtained by correcting P ₁ by the correction value P _p , so even if there are product variations or changes over time in the front clutch 104, these will not be affected. Therefore, the target value can always be set to the optimum value without being affected, and shift shock can be reliably prevented without causing excessive slippage of the front clutch (early wear and power loss).

By the way, since the solenoid valve drive frequency is switched to a low frequency in block 333 during the above-mentioned shift shock prevention, the control pressure (P _S ) characteristic and front clutch pressure (P _F/C ) characteristic with respect to duty are respectively 6th. As shown by the solid line in the figure, the shift shock prevention control range Y is gradual;
While the shift shock is being prevented, the control can be performed stably without causing hunting.

In addition, when block 334 in FIG. 11 determines that S ₁ ≠ 1, that is, when it determines that there is no 2→3 shift command, and when block 336 determines that F<0, that is, F<0, the shift is a downshift shift. If it is determined that there is, the above-mentioned shift shock prevention control is not performed, so the control proceeds to block 349, similar to after execution of block 338, and the shift start determiner _S2 is set to 0.
control then passes to block 350. For block 350, the output duty is 100.
%, but this output duty of 100% is calculated by changing the front clutch pressure P _F/C to the line pressure as shown in Figure 6.
The P _L itself is used, and the operation control of the front clutch 104 is left to the shift control hydraulic circuit shown in FIG. Execution of block 350 is executed by block 34.
This is also performed after the 2→3 shift command and before the start of the 2→3 shift when it is determined that 0 is not ΔN _E >(ΔN _E ) _p .

In the above example, the control pressure (P _S ) characteristic and front clutch pressure (P _F/C ) characteristic were changed to those shown by the solid line or dotted line in Fig. 6 by switching the solenoid valve driving frequency. This change can also be achieved by switching the solenoid valve drive voltage. That is, the solenoid valve 207 attracts the plunger 207b and closes the orifice 206 by the electromagnetic force generated each time the solenoid 207c is energized, and the electromagnetic force and the driving voltage are in a proportional relationship. That is, the lower the driving voltage is, the weaker the electromagnetic force is, and therefore the dead time t ₁ and delay time t ₂ are larger in FIG. 13b, where the driving voltage is low, than in FIG. 13a, where the driving voltage is high. Therefore, as is clear from the comparison of Figure 13a and b, the drain flow rate indicated by diagonal lines increases as the drive voltage decreases even at the same duty ratio, and the control pressure
When the drive voltage is high, P _S changes with the characteristics shown by the solid line in FIG. 14, but when the drive voltage is low, it changes with the characteristics shown by the dotted line in the same figure. Although not shown in FIG. 14, the front clutch pressure (P _F/C ) characteristic also exhibits a corresponding change as the drive voltage is switched.

From this point of view, the content of execution in block 309 in FIG.
Even if the content of execution is changed to one in which the electromagnetic valve drive voltage is changed to be higher, the same effects as in the above-mentioned example can be obtained.

(Effects of the Invention) Thus, as described above, the pressure control device of the present invention controls the hydraulic pressure of the common friction elements of automatic transmissions to the required value for each control purpose. Since the pressure change characteristics are changed to obtain just the right characteristics, stable control without hunting can be obtained for any control purpose, and the commercial value of the automatic transmission can be greatly increased.

[Brief explanation of the drawing]

FIG. 1 is a conceptual diagram of the pressure control device of the present invention, FIG. 2 is a skeleton diagram showing the power transmission section of an automatic transmission equipped with the device of the present invention, and FIG. 3 is a shift control hydraulic circuit diagram of the automatic transmission. Fig. 4 is a system diagram of the device of the present invention, Figs. 5a and 5b are time charts showing changes in duty output by the computer of the device, respectively, and Fig. 6 shows changes in control pressure and front clutch pressure with respect to duty. A characteristic diagram showing the change situation using the driving frequency as a parameter, Fig. 7 is a computer block diagram, Fig. 8, Fig. 9 a,
Figures 10a and 11 are flowcharts of the control program executed by the computer, respectively.
Figures b and 10b are explanatory diagrams of the waveforms before and after waveform shaping of the engine rotational speed signal and the torque converter output rotational speed signal, respectively. Time chart showing the degree of opening change of the solenoid valve for the same duty when the duty is low.
Figures 3a and 3b are time charts showing a comparison of the degree of opening change of the solenoid valve for the same duty when the driving voltage of the solenoid valve is high and low, respectively, No. 14.
The figure is a characteristic diagram showing how the control pressure changes with respect to the duty using the drive voltage as a parameter. 24...Microprocessor unit, 25...
...Read-only memory, 26...I/O interface circuit, 27...A/D converter, 28...Waveform shaping circuit, 29...Amplifier, 104...Front clutch (common friction element), 168... Front clutch pressure circuit, 200... Pressure control device of the present invention, 201... Pressure control valve, 202... Drain circuit, 203, 204... Branch path, 205, 206
... Orifice, 207 ... Solenoid valve, 208 ...
Control computer, 209...Pressure sensor, 2
10... Throttle opening sensor, 211... Engine rotation speed sensor, 212... Torque converter output rotation speed sensor, 213... Transmission output rotation speed sensor, 214... Shift switch, 215...
Inhibitor switch, 216... Idle switch, 217... Temperature sensor.

Claims (1)

[Scope of Claims] 1. An automatic transmission provided in which the hydraulic pressure of a selected common friction element is controlled in different hydraulic ranges according to a plurality of pressure control purposes to individually achieve the pressure control purposes. , the working oil pressure of the friction element is set to a predetermined value in any of the hydraulic ranges by a solenoid valve that is operated by applying a predetermined voltage during the energization time width within a set cycle determined by the energization time width determining means; In the pressure control device, the hydraulic pressure region determining means determines the hydraulic pressure region to be controlled, and the set period or the predetermined voltage is changed according to the hydraulic pressure region determined by the means, and the energization time width occupied in the set period is determined. 1. A pressure control device comprising pressure change characteristic changing means for changing pressure change characteristics with respect to a ratio of .