JPH0379589B2 - - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION Field of Industrial Application The present invention relates to a control device for a continuously variable transmission for a vehicle (hereinafter, "continuously variable transmission" will be referred to as "CVT").

BACKGROUND TECHNOLOGY In vehicles equipped with a continuously variable transmission in which the gear ratio can be changed steplessly, the amount of accelerator pedal operation or intake throttle control is usually determined based on a predetermined relationship that takes into account fuel economy or drivability. A control target value is determined based on an acceleration operation amount such as the opening degree, and the gear ratio is adjusted so that the control target value matches the engine rotational speed or the gear ratio. When adjusting the gear ratio, a first solenoid valve with two-position operation is used to control the direction of change in the gear ratio of the continuously variable transmission, and a second solenoid valve is operated to control the speed of change in the gear ratio of the continuously variable transmission. and a position-operated second solenoid valve.
In order to control the changing direction and changing speed, the first electromagnetic valve is operated to change the gear ratio in the shift direction in which the control deviation is eliminated, and when the control deviation exceeds a predetermined value, the speed changing ratio is changed. The second electromagnetic valve is operated so that the change speed of the gear ratio becomes rapid, and the change speed of the gear ratio becomes slow when the change speed falls below the predetermined value, thereby improving control responsiveness. For example, the patent application filed earlier by the applicant in 1982-
This is the control device for a continuously variable transmission for a vehicle described in No. 203130 (Japanese Unexamined Patent Publication No. 60-95262).

Problems to be Solved by the Invention By the way, in the control device for a continuously variable transmission for a vehicle as described above, a certain response delay occurs when the speed ratio change speed is changed by the second electromagnetic valve of the speed change control means. For this reason, at a predetermined speed ratio change speed, the speed ratio change width within the response delay time is constant, whereas the controlled variable, for example, the input side rotation speed of a continuously variable transmission (=engine rotation speed) Since the rate of change changes in relation to the intake throttle opening, the amount of change in the input side rotational speed within the response delay time increases in accordance with the intake throttle opening at that time. However, in conventional control devices, the predetermined reference value used when determining the switching of the gear ratio change speed based on the magnitude of the control change is constant, as described above. Therefore, if the predetermined judgment reference value is determined to be applicable to a relatively small intake throttle opening, the input rotation speed will greatly exceed the control target value when the intake throttle opening becomes large. If a phenomenon (overshoot) occurs and, conversely, the above-mentioned predetermined judgment reference value is determined to be adapted to a relatively large intake throttle opening, when the intake throttle opening becomes small, the above-mentioned input side rotational speed will decrease. This method has the disadvantage that there is a delay in the approach of the control target value.

The present invention has been made against the background of the above circumstances, and its purpose is to prevent a decrease in the control accuracy of the gear ratio due to a response delay when switching the gear ratio changing speed of a continuously variable transmission. An object of the present invention is to provide a control device for a continuously variable transmission for a vehicle.

Means for Solving the Problems To achieve the object, the gist of the present invention is to provide (a) a two-position operating first solenoid valve for controlling the direction of change in the gear ratio of a continuously variable transmission; (b) a speed change control means having a second solenoid valve with two-position operation for controlling the speed ratio change speed of the continuously variable transmission;
Target input side rotation speed determining means for determining a target input side rotation speed based on the actual throttle opening;
(c) control deviation calculation means for calculating a control deviation that is the difference between the target input side rotation speed and the actual input side rotation speed; The second solenoid valve operates the first solenoid valve so that the speed ratio change speed becomes rapid when the control deviation exceeds a predetermined value, and becomes gentle when the control deviation falls below the predetermined value. and (e) predetermined value calculation means for determining the predetermined value to be a large value in accordance with the throttle opening.

Operation and Effects of the Invention With this arrangement, the predetermined value used when switching the speed ratio change speed in the gear ratio control means is determined to be a large value in accordance with the throttle opening degree by the predetermined value calculation means. Therefore, even if the range of change in the ratio control amount (input side rotational speed) within the operation delay time of the speed change control means changes in accordance with the acceleration operation amount, the control accuracy of the controlled amount does not decrease. For example, when the throttle opening degree is operated to a large extent, the predetermined value is increased, so that the predetermined value is quickly lowered and the gear ratio changing speed is made gentle, and the overshoot phenomenon is suitably eliminated. In addition, when the throttle opening is operated small, the predetermined value is reduced, so there is a delay in falling below the predetermined value, and a delay in the speed of change in the gear ratio being gradual, eliminating control changes. This effectively eliminates the phenomenon of delays.

Example Embodiments of the present invention will be described with reference to the drawings.

In FIG. 2, the CVT 1 includes a pair of input sheaves 2a, 2b, a pair of output sheaves 4a, 4b,
It also includes a belt 6 that is hung over the sheaves on the input side and the output side to transmit engine power. One input side sheave 2a is provided on the input shaft 8 so as to be movable in the axial direction and fixed in the rotational direction, and the other input side sheave 2b is fixed to the input shaft 8. Further, one output side sheave 4a is fixed to the output shaft 10,
The other output side sheave 4b is provided on the output shaft 10 so as to be movable in the axial direction but fixed in the rotational direction. The opposing surfaces of the input side sheaves 2a, 2b and the opposing surfaces of the output side sheaves 4a, 4b are tapered radially outward to increase the mutual distance, and the cross section of the belt 6 has an isosceles trapezoidal shape. It is formed. The pressing force of the output side sheaves 4a, 4b is limited to the minimum value that can avoid belt 6 slippage and ensure power transmission, and the pressing force of the input side sheaves 2a, 2b is determined by the gear ratio γ of the CVT 1 (=input shaft 8 rotation speed
Nin/rotational speed Nout of the output shaft 10) is determined.
The fluid coupling 12 includes a pump 16 connected to the crankshaft 14 of the engine, and a turbine 18 rotated by oil from the pump 16 and fixed to the input shaft 8. The lockup clutch 22 controls the connection between the crankshaft 14 and the input shaft 8, and the tamper 24 absorbs shock and engine torque fluctuations as the lockup clutch is switched from a disengaged state to an engaged state. When the vehicle speed or engine speed exceeds a predetermined value, the lock-up clutch 22 is held in an engaged state to avoid loss of power transmission due to oil in the fluid coupling 12. The oil pump 26 rotates integrally with the pump 16 and pumps oil to the CVT via a hydraulic control device.
1. Send to fluid coupling 12, etc. The counter shaft 28 is
Provided parallel to the output shaft 10 of the CVT 1,
It has two gears 30 and 32. Output shaft 10
The engine power is transmitted from a gear 34 coaxial with the output shaft 10 to a differential gear 36 via gears 30, 32 on a counter shaft 28, and further from the differential gear 36 to left and right axles 38, 40. is sent to the drive wheels. The sub-transmission 42 is the output shaft 10 of the CVT 1
coaxially provided with respect to the The sub-transmission 42 includes a Lavigneau type compound planetary gear set 43, and this planetary gear set 43 includes a first sun gear 44, a second sun gear 44,
46, a first planetary gear 48 meshing with the first sun gear 44;
and a second planetary gear 50 that meshes with the second sun gear 46, a ring gear 52 that meshes with the first planetary gear 48, and a carrier 54 that rotatably directs the first and second planetary gears 48, 50. The second sun gear 46 is connected to a shaft 64 that is integral with the output shaft 10 of the CVT 1 as an input part of the auxiliary transmission 42, and
4 is connected to gear 34. The high speed clutch 56 controls the connection between the shaft 64 and the first sun gear 44, and the low speed brake 58 controls the connection between the shaft 64 and the first sun gear 44.
The reverse brake 60 controls the fixation of the ring gear 52.

FIG. 3 shows the operating state of each frictional engagement element of the sub-transmission 42 and the reduction ratio in each range. ○ means engaged state, × means released state, ρ1
and ρ2 are defined from the following equation.

ρ1=Zs1/Zr ρ2=Zs2/Zr where Zs1 is the number of teeth of the first sun gear 44, Zs
2 is the number of teeth of the second sun gear 46, and Zr is the number of teeth of the ring gear 52. In other words, in the low gears of the L and D ranges, the first sun gear 44 is fixed by the low gear brake 58, so the reduction ratio is 1+ρ1/ρ2.
In the high speed gears of the L and D ranges, the high speed clutch 56 is engaged and the planetary gear unit 43 rotates as a unit, thereby transmitting the engine power at a reduction ratio of 1. , R range, the ring gear 52 is fixed by the reverse brake 60, so engine power is transmitted in reverse rotation with a reduction ratio of 1-1/ρ2.

4 to 6 are detailed views of the hydraulic control device. The oil pump 26 pressurizes the oil sucked in through the strainer 72 and supplies it to the line pressure oil path 74. The throttle valve 76 generates a throttle pressure Pth related to the intake throttle opening θ at the output port 78. The spool 77 of the throttle valve 76 receives an acting force that increases as the throttle opening θ increases from the throttle cam 79 and a throttle pressure Pth as a feedback pressure from the control port 81 oppositely. Controls connection with output port 78. The manual valve 80 has its axial position controlled in relation to the L (low), D (drive), N (neutral), R (reverse), and P (parking) ranges of the shift lever, and is connected to the line hydraulic path 74. The first line pressure Pl1 is sent to port 83 in the R range, to port 85 in the L range, and to port 8 in the D range.
5 and 87, respectively.

The relief valve 89 has a function as a safety valve that releases oil from the line pressure oil passage 74 when the first line pressure Pl1 of the line pressure oil passage 74 exceeds a predetermined value.

The secondary hydraulic oil passage 82 is connected to the line pressure oil passage 74 via an orifice 84 and a port 85 from which excess oil of the primary regulator valve 198 is discharged, and the secondary pressure regulator valve 86 is The control chamber 90 is connected to the secondary hydraulic oil passage 82 via the hydraulic oil passage 88, and the connection between the secondary hydraulic oil passage 82 and the port 94 is controlled in relation to the oil pressure in the control chamber 90 and the load of the spring 92. control to maintain the secondary hydraulic pressure Pz of the secondary hydraulic oil passage 82 at a predetermined value. The lubricating oil passage 95 is connected to the secondary hydraulic oil passage 82 via a port 94 or an orifice 97. Lockup control valve 96 selectively connects secondary hydraulic fluid line 82 to the engagement and release sides of lockup clutch 22 in parallel with fluid clutch 12. The solenoid valve 100 controls the connection between the control chamber 102 of the lock-up control valve 96 and the drain 104,
When the solenoid valve 100 is off (de-energized), the secondary hydraulic oil passage 8 is routed to the release side of the lock-up clutch 22.
The secondary hydraulic pressure Pz from 2 is transmitted and the engine power is transmitted through the fluid clutch 12, and the solenoid valve 100
is on (energized), the secondary hydraulic pressure Pz from the secondary hydraulic oil passage 82 is supplied to the engagement side of the lock-up clutch 98 and the oil cooler 106, and the engine power is transmitted via the lock-up clutch 98. Ru. Cooler bypass valve 107 controls cooler pressure.

The gear ratio control device 108 includes first and second spool valves 110, 112, and first and second electromagnetic valves 114, 116. First solenoid valve 1
14 is off, the first spool valve 110
The spool of is pressed toward the spring 118 by the secondary hydraulic pressure Pz of the chamber 117, and the first line pressure Pl1 of the port 119 is applied to the second spool valve 112 through the port 120 of the first spool valve 110. port 122, port 124 and drain 12
The connection with 6 has been severed. First solenoid valve 114
is on, the hydraulic pressure in the chamber 117 is discharged through the drain 128 of the first solenoid valve 114, and the first
The spool of the spool valve 110 is pushed toward the chamber 117 by the spring 118, and there is no line pressure Pl at the port 120, and the port 124 is connected to the drain 126.
connected to. Further, during the period when the second solenoid valve 116 is off, the spool of the second spool valve 112 is pressed toward the spring 130 by the secondary hydraulic pressure Pz of the chamber 128, and the connection between the ports 122 and 132 is cut off. , port 134 is connected to port 136. Ports 132 and 134 are oil passages 138
It is connected to the input side hydraulic cylinder of CVT1 via. During the period when the second solenoid valve 116 is on, the hydraulic pressure in the chamber 128 is discharged from the drain 139 of the second solenoid valve 116, and the spool of the second spool valve 112 is pressed toward the chamber 128 by the spring 130. port 122 is connected to port 132;
The connection between ports 134 and 136 is severed. Port 136 is connected to port 1 via oil passage 142.
It is connected to 24. Orifice 140 is the second
When the solenoid valve 116 is turned off, a small amount of oil is guided from the port 122 to the port 132. Therefore, the first
The second solenoid valve 114 is off and the second solenoid valve 116 is off.
During the period when is on, oil is quickly supplied to the input hydraulic cylinder of the CVT 1, and the gear ratio γ decreases. During the period when the first solenoid valve 114 is off and the second solenoid valve 116 is off, oil is supplied to the input hydraulic cylinder of the CVT 1 through the orifice 14.
0, and the gear ratio r of the CVT 1 gradually decreases. When the first solenoid valve 114 is on and the second solenoid valve 116 is on, the CVT1
Oil is not supplied to or discharged from the input hydraulic cylinder, and the CVT 1 is held constant at the gear ratio γ. During the period when the first solenoid valve 114 is on and the second solenoid valve 116 is off, the oil in the input hydraulic cylinder is further discharged from the drain 126, so the gear ratio γ of the CVT 1 rapidly increases.

The gear ratio detection valve 146 is shown in detail in FIG. Sleeves 148, 150 are coaxially disposed within valve bore 152 and are axially secured by snap spring 154. A rod 156 passes through the end of the sleeve 148 and has a spring seat 158 secured thereto. Another rod 160 is coupled at both ends to the input movable sheave 2a and the rod 156, respectively, and moves the rod 156 in the axial direction by a displacement equal to the axial displacement of the input movable sheave 2a. The spool 162 has lands 164 and 166 and is fitted within the sleeve 150 so as to be movable in the axial direction. Land 164 is land 164 and 1
The land 166 has a passage 172 that communicates a space 168 between the land 166 and the oil chamber 170 to the oil chamber 170.
The opening area of the port 174 of the sleeve 150 is controlled. Port 174 is connected to drain 176 through a space around the outer circumference of sleeve 148 .
Oil chamber 170 is output port 1 that generates control pressure Pc
78, and the output port 178 is connected to the orifice 18.
0 to the line pressure oil passage 74.
A spring 182 is provided between the spring receiver 158 and the sleeve 150 and biases the rod 156 in a direction to push the rod 156 out of the sleeve 148.
8 and the flange 186 of the spool 162 to bias the spool 162 toward the oil chamber 170. As the amount of displacement of the input side movable sheave 2a of the CVT 1 with respect to the input side fixed sheave 2b increases, the speed ratio γ increases. As the rod 156 is pushed out of the sleeve 148 due to an increase in the amount of displacement of the input side movable sheave 2a, the spring 1 moves toward the oil chamber 170.
The biasing force exerted on the spool 162 by 84 is reduced.
As a result, the spool 162 moves toward the rod 156, and the land 166 increases the opening area of the port 174 to increase the oil discharge flow rate, so that the gear ratio pressure Pγ of the output port 178 decreases. Since the gear ratio pressure Pγ is generated by controlling the amount of hydraulic medium discharged from the output port 178, the upper limit is defined as the line pressure Pl. The broken lines in FIGS. 8 and 9 illustrate two relationships between the gear ratio pressure Pγ and the gear ratio γ. First line pressure as described below
Pl1 decreases as the gear ratio γ decreases, but the gear ratio γ1 at which the gear ratio pressure Pγ becomes equal to the line pressure Pl1
(This gear ratio γ1 is a function of the throttle pressure Pth and therefore the engine torque Te.)
In a gear ratio range beyond this range, Pγ=Pl1. Note that in FIGS. 8 and 9, the two-dot chain line represents the ideal value of the first line pressure Pl1, and T1>T2.

The cut-off valve 190 has a chamber 194 communicating with the control chamber 102 of the lock-up control valve 96 via an oil passage 192, and a spool 1 that moves in relation to the oil pressure in the chamber 194 and the spring force of a spring 195.
96, and when the solenoid valve 100 is off, that is, when the lock-up clutch 22 is in the released state (when changing gears in the auxiliary transmission 42, the lock-up clutch is closed to absorb the shock of the power transmission system). 22 is in the open state), and is in the closed state to prevent the gear ratio pressure Pγ from being transmitted to the primary regulator valve 198.

The primary regulator valve 198 as a first line pressure generating means has a port 200 supplied with the throttle pressure Pth, a port 202 supplied with the gear ratio pressure Pγ, and a port 204 connected to the line pressure oil path 74. , a port 206 connected to the suction side of the oil pump 26, and a port 210 supplied with the first line pressure Pl1 via an orifice 208, which move in the axial direction to connect the ports 204 and 206. Spool 212 to control, spool 2 in response to throttle pressure Pth
12 toward port 202
4, and a spring 216 that biases the spool 212 toward the port 202. Spool 21
The area of the two lands from the bottom of 2 is A1, A2, the area of the land of the spool 214 that receives the throttle pressure Pth is A3, and the acting force of the spring 216 is W1.
By defining each of these, the following formula holds true.

Cut-off valve 190 opens and port 202
When the gear ratio pressure Pγ is at
If the gear ratio pressure Pγ does not arrive at In FIG. 9, they are indicated by solid lines and dashed-dotted lines, respectively.

The sub-primary regulator valve 220 as a second line pressure generating means receives the first line pressure Pl1 from the input port 222, which is led from the port 85 of the manual valve 80, and the second
The output port 224 generates the second line pressure Pl2, the port 226 receives the gear ratio pressure Pγ, the port 230 receives the second line pressure Pl2 as feedback pressure via the orifice 228, the input port 222 and the output port 224. spool 232 that controls the connection of
A spool 236 that receives Pth and urges the spool 232 toward the port 226, and a spool 23
2 toward port 226. If the areas of the bottom two lands of the spool 232 are defined as B1 and B2, the area of the land of the spool 236 that receives the throttle pressure Pth is defined as B3, and the acting force of the spring 238 W2, the following equation holds true.

Pl2=(B3・Pth+W2−B1・Pγ) /(B2−B1) ...(3) Figure 10 shows the relationship between the second line pressure Pl2 generated by the sub-primary regulator valve 220 and its ideal value. It shows.

The shift valve 250 has an input port 252 to which the second line pressure Pl2 is introduced in the D and L ranges, output ports 254 and 256, and an orifice 258, and a discharge oil passage 26 that terminates at a drain 260.
1, a control port 264 that is supplied with the first line pressure Pl1 from the port 87 of the manual valve 80 in the D range, other control ports 266, 268, a drain 270,
It has a spool 272 and a spring 274 that biases the spool 272 toward the port 248.
Secondary hydraulic pressure Pz is guided to the control ports 266, 268 via an orifice 276, and the control ports 266, 268
The oil pressure at 268 is controlled by a solenoid valve 278.
The area of the two lands from the bottom of the spool 272 is S
1, S2, and S1<S2. Further, whether the solenoid valve 278 is turned on or off is controlled in relation to the driving parameters of the vehicle, and when the solenoid valve 278 is turned on, oil is discharged from the drain 280.

When spool 272 is in the spring 274 position, input port 252 is connected to output port 254 and output port 256 is connected to port 262. Accordingly, the second line pressure Pl2 is supplied from the output port 254 to the accumulator 282 having the piston 281 and the high speed clutch 56, and the sub-transmission 42 is set to the high speed.

When the spool 272 is in the port 268 position, the input port 252 is connected to the output port 256 and the output port 254 is connected to the drain 270. Therefore, the second output from output port 256
The line pressure Pl2 is supplied to the low speed accumulator 58, and the auxiliary transmission 42 is set to the low speed.

In the case of the L range, the first line pressure Pl1 is not introduced to the control port 264, so the solenoid valve 27
8 is turned off, the spool 272 is first moved toward the spring 274 by the secondary hydraulic pressure Pz acting on the land with area S2, and then by the secondary hydraulic pressure Pz acting on the land with area S1, but the solenoid valve 278
is turned on, the oil pressure in the control ports 266, 268 decreases, so the spool 272
to port 268. That is, L
In the range, the auxiliary transmission 42 can be switched between a high speed gear and a low speed gear by turning the solenoid valve 278 on and off.

In the D range, the first line pressure Pl1 is introduced to the control port 264, so once the spool 272 is at the position on the spring 274 side, the first line pressure Pl1 from the control port 264 acts on the land with an area S2. Regardless of whether the solenoid valve 278 is turned on or off after that, the spool 272 is held in the position on the spring 274 side, and therefore the sub-transmission 42 is held in the high speed gear.

The shift timing valve 290 has a control port 292 that communicates with the high speed clutch 56, and a spool 294 whose axial position is controlled by the hydraulic pressure of the control port 292, and is used during upshifting from a low speed to a high speed. Clutch 5 for high speed gear
6 and the flow rate of oil discharged from the low speed brake 58.

The solenoid valves 100, 114, 116, and 278 are guided by the secondary hydraulic pressure Pz from the secondary hydraulic oil passage 82, and control the discharge of the secondary hydraulic pressure Pz. In the hydraulic control device disclosed in Japanese Patent Application No. 59-12071, the solenoid valve is guided by the throttle pressure Pth. Therefore, in the conventional device, the spring force and solenoid suction force of the solenoid valve must be set to cope with the maximum throttle pressure, which has the disadvantage of increasing the size of the solenoid valve. There are problems in that the responsiveness of the spool of the spool valve deteriorates and the setting of the spring force acting on the spool becomes complicated. In this embodiment, by using the secondary hydraulic pressure Pz, these problems are solved and the degree of freedom in design is improved.

FIG. 11 is a control block diagram. The electronic control device 310 controls the intake throttle opening θ, the vehicle speed V,
Parameters such as the input side rotational speed Nin of the CVT 1, the engine cooling water temperature Tw, and the shift position are received as input signals, and the solenoid valves 100, 114, 116, and 278 of the hydraulic control device 312 are controlled via the amplification stage 314. do.

FIG. 12 shows a routine for selecting the control mode of the CVT 1. First, the electronic control unit 310 is initialized (step 318), and then data necessary for the following control, such as input from various sensors, is read (step 320). In this way, the system performs diagnostics to check whether various devices are operating normally (step 322), mutual control with the engine computer that controls engine ignition timing, fuel injection amount, etc. (step 324), and lock-up clutch 22. lock-up control (step 326) that controls engagement and release of the transmission, and transmission section control (step 326) that controls the transmission section including the CVT 1 and the auxiliary transmission 42.
328). For transmission control
CVT1 shift control (step 330) and sub-transmission 4
2. Switching between high speed and low speed (step 332)
including. Finally, the control values calculated as a result of these controls are output (step 334). The present invention relates to the shift control of the CVT 1 in step 330.

Figure 13 shows the target input side rotation speed Nino and CVT1
This shows the relationship with shift control. In other words, the target input side rotation speed Nino is the actual input side rotation speed.
Compared with Nin, NinoNin+ΔN1, Nin+ΔN
2Nino<Nin+ΔN1, NinNino<Nin+
ΔN2, Nin−ΔN3＜Nino＜Nin, Nin−ΔN4
＜NinoNin−ΔN3, and NinoNin−ΔN
4 (where ΔN1, ΔN2, ΔN3, and ΔN4 are positive predetermined values), and depending on each case, a rapid downshift, an intermediate speed downshift, a slow downshift, a slow upshift, or Medium speed upshifts and rapid upshifts are performed.

FIG. 14 shows the relationship between the control signals of the first and second electromagnetic valves 114 and 116 when divided in FIG. 13. In the case of NinΔNino, the first
The solenoid valve 114 of the first spool valve 110 is turned on, that is, the port 124 of the first spool valve 110 is connected to the drain 126, and oil is discharged from the input hydraulic cylinder of the CVT 1 through the drain 126, causing a downshift. When Nin>Nino, the first solenoid valve 114 is turned off, that is, the port 119 of the first spool valve 110 is connected to the port 120, oil is supplied to the input hydraulic cylinder of the CVT 1, and an upshift occurs. . NinoNin
+ΔN1, the second solenoid valve 116 is turned off, and the second spool valve 112
4 is connected to port 136, which allows CVT
The cross-sectional area of the oil discharge passage from the input hydraulic cylinder 1 increases, causing a rapid downshift. In the case of NinNin+ΔN2, the second solenoid valve 116 is turned on and the connection between the ports 134 and 136 is cut off.As a result, the flow path cross-sectional area of the oil discharge passage from the input side hydraulic cylinder of the CVT 1 becomes zero, and the cylinder Alternatively, a slow downshift may occur due only to oil leaking from the valve. Nin
When +ΔN2Nino<Nin+ΔN1, the second solenoid valve 116 is duty-controlled (controlled by a pulse signal), and the flow cross-sectional area of the discharge oil path is changed by repeating between the increased state and 0 at the frequency of the duty control. The result is an intermediate magnitude, resulting in a downshift that is intermediate in speed between a fast downshift and a slow downshift. Nin−ΔN3<
If Nino<Nio, the second solenoid valve 116 is turned off, and the ports 122 and 132 of the second spool valve 112 are connected via the orifice 140, and oil is supplied to the input hydraulic cylinder of the CVT 1. The cross-sectional area of the oil passage is very small and slow upshift occurs. NinoNin−
In the case of ΔN4, the second solenoid valve 116 is turned on and the port 12 is turned on at the second spool valve 112.
2 and port 132 are connected, thereby
The cross-sectional area of the oil supply passage for supplying oil to the input hydraulic cylinder of the CVT 1 is large, and a rapid upshift occurs. When Nin−ΔN4<NinoNin−ΔN3, the second electromagnetic valve 116 is duty-controlled, and the cross-sectional area of the supply oil passage is repeatedly changed between an increased state and a very small state at the duty-controlled frequency. This results in an upshift of intermediate magnitude and speed intermediate between a rapid upshift and a slow upshift.

In the present invention, ΔN1 to ΔN4 are functions of the intake throttle opening degree θ.

FIG. 15 is a flowchart of the CVT 1 shift control routine. The target input rotation speed Nino is calculated from the intake throttle opening θ and the vehicle speed V (step 338), and the difference between the target input side rotation speed Nino and the actual input side rotation speed Nio is calculated as ΔNin.
(step 340). Compare ΔN with 0 (step 342), and if ΔNin0, that is, downshift is required, predetermined values ΔN1, ΔN2 are set.
is calculated from the intake throttle opening θ (step
344), and if ΔNin<0, that is, if an upshift is required, predetermined values ΔN3 and ΔN4 are calculated from the intake throttle opening θ (step 345).
Next, ΔNin and the predetermined values ΔN1, ΔN2, −ΔN4, −
Compare ΔN3 (steps 346, 347, 348,
350), and controls the first and second solenoid valves 114, 116 as shown in the case breakdown of FIG. 14 (steps 352 to 362).

Next, the effects of the above embodiment will be explained. Of Figures 16 to 19, Figure 16 and 1
Figure 9 shows the change in the input side rotational speed N _io of this embodiment, of which Figure 16 shows the case where the intake throttle opening is small (θ = θa);
Figure 9 shows the case where the intake throttle opening is large (θ=
θb, provided that θa<θb. ) is shown. Also,
FIG. 17 shows a case where the judgment reference value ΔN _1a having the same magnitude as the state shown in FIG. 16 is fixedly used as the judgment reference value ΔN ₁ and the intake throttle opening degree is increased (θ=θb). The figure shows a case where the judgment standard value ΔN _1b (however, ΔN _1b < ΔN _1a ), which is the same as the condition in FIG. ₁₉ , is used as the judgment standard value ΔN 1, and the intake throttle opening is made small (θ=
θa).

For example, when the gear ratio γ of the CVT 1 is downshifted in connection with an acceleration operation, the input side rotational speed N _io is increased, and N _1a ≦N _iop −
When ΔN _1a is satisfied, the rapid downshift region shown in FIG. 13 occurs, and then the actual input side rotational speed N _1a approaches the target input side rotational speed N _iop and the control deviation N _iop −N _1a increases. By becoming smaller
When N _1a ≦N _iop −ΔN _1a is satisfied, the downshift region of intermediate speed shown in FIG. 13 is reached.
At this time, there is an operation delay time T of the second spool valve 112 that functions as a flow control valve between the rapid downshift region and the intermediate speed downshift region, and within this response delay time T, The rapid downshift continues, and after the response delay time T has elapsed, an intermediate downshift begins. This response delay time T is constant regardless of the intake throttle opening θ, and the amount of change in the input side rotational speed N _io during the response delay time T increases as the intake throttle opening θ increases.
The larger the N _io and the target input side rotational speed N _iop are, the larger it becomes. In this embodiment, the judgment reference value ΔN ₁ is determined according to the intake throttle opening θ. For example, when the intake throttle opening θ is θ _a , ΔN _1a
When the intake throttle opening degree θ increases to θ _b , it becomes ΔN _1b , so as shown in FIGS. 16 and 19, the input side rotational speed N _io which is the controlled variable
can be quickly brought close to the target input side rotational speed N _iop , and control accuracy at the time of transient response of the input rotational speed N _io can be suitably obtained. Incidentally, in the conventional case where the judgment reference value ΔN ₁ is constant, an overshoot phenomenon occurs as shown in FIG. 17 _, or as shown in FIG. This may cause a delay in approaching the speed N _iop .

FIG. 1 is a functional block diagram of the present invention.

The input side rotation speed sensor 363 and the throttle opening sensor 364 each have an input side rotation speed Nin.
and intake throttle θ. The target input side rotational speed calculation means 365 calculates the intake throttle opening degree θ.
The target input side rotational speed Nino is calculated based on the following, and the difference calculating means 366 calculates Nino-Nin as the difference ΔNin. The predetermined value calculation means 368 calculates the predetermined values ΔN1 to ΔN4 from the intake throttle degree θ, and the case calculation means 370 detects which case in FIG. 14 corresponds to the difference ΔNin and the predetermined values ΔN1 to ΔN4, The first and second solenoid valves 114 and 116 are controlled depending on each case. The first solenoid valve 114 functions as a speed change direction (upshift or downshift) control means, and the second solenoid valve 116 functions as a speed change speed control means.

Although the present invention has been described with reference to embodiments, it will be apparent to those skilled in the art that the present invention is not limited thereto and can be modified and modified in various ways.

[Brief explanation of drawings]

Fig. 1 is a functional block diagram of the present invention, Fig. 2 is a functional block diagram of the present invention.
A skeleton diagram of the power transmission device with CVT, Figure 3 is a diagram showing the relationship between the range and the operating state of the friction engagement device, Figures 4 to 6 are detailed diagrams of the hydraulic control device, and Figure 7 is the gear ratio. Detailed view of the detection valve, Figures 8 and 9 are graphs showing the characteristics of the first line pressure,
FIG. 10 is a graph showing the characteristics of the second line pressure,
FIG. 11 is a control block diagram, FIG. 12 is a diagram showing a CVT control mode selection routine, and FIG. 13 is a diagram showing the relationship between target input side rotational speed and shift speed.
Fig. 14 is a chart showing the control of the first and second solenoid valves in each case, Fig. 15 is a flowchart of the CVT speed change control routine, and Figs. 16 to 19 are input side rotational speeds in various cases. FIG. 1...CVT, 114...first solenoid valve, 11
6...Second solenoid valve, 363...Input side rotational speed sensor, 364...Throttle opening sensor, 36
5...Target input side rotational speed calculation means, 366...
Difference detection means, 368... Predetermined value calculation means, 370
...A case detection means.

Claims (1)

[Scope of Claims] 1. A first solenoid valve that operates in two positions for an increasing speed change direction state and a decelerating speed change direction state to control the changing direction of the speed ratio of the continuously variable transmission, and the speed change ratio of the continuously variable transmission. A speed change control valve device having a second solenoid valve operating in two positions, one in which hydraulic oil is allowed to flow and the other in which it is inhibited, in order to control the speed of change, and a target input side rotation speed determined based on the actual acceleration operation amount. a control deviation calculation means for calculating a control deviation that is the difference between the target input side rotation speed and the actual input side rotation speed; The first electromagnetic valve is actuated to change the control deviation, and when the control deviation exceeds a predetermined value, the speed ratio change speed is set to be rapid as the flow permissible state, and when it falls below the predetermined value, the speed ratio change speed is set as the flow suppressed state. gear ratio control means for operating the second electromagnetic valve so that the speed ratio change speed becomes gentle; and calculating the predetermined value based on the acceleration operation amount so that control characteristics by the gear ratio control means are optimized. A control device for a continuously variable transmission for a vehicle, comprising: predetermined value calculation means for calculating a predetermined value.