JP5660869B2 - Control device for belt type continuously variable transmission - Google Patents

Description

The present invention relates to a control device for a belt-type continuously variable transmission, and more particularly to a device for suppressing shift control variation and pinching control variation.

Conventionally, a belt is wound between a primary pulley and a secondary pulley, and the amount of oil supplied to the oil chamber of the primary pulley is controlled by a ratio control valve (flow control valve), thereby controlling the pulley ratio and the secondary pulley. A belt type continuously variable transmission that controls belt clamping pressure by controlling the hydraulic pressure supplied to the oil chamber with a clamping pressure control valve (pressure control valve) is known.

Patent Document 1 discloses an example of a hydraulic control device for a belt-type continuously variable transmission as described above. FIG. 8 shows an outline of the hydraulic control apparatus, where 100 is an oil chamber of a primary pulley, and 101 is an oil chamber of a secondary pulley. The solenoid modulator valve 102 is a constant pressure valve that receives a clutch modulator pressure and outputs a constant solenoid modulator pressure. The solenoid modulator pressure is supplied to the upshift duty solenoid valve 103 and the downshift duty solenoid valve 104, respectively, and each solenoid valve 103, 104 generates a signal pressure corresponding to the duty ratio. The ratio control valve 105 receives the signal pressures of the solenoid valves 103 and 104 in the opposite direction, controls the flow rate of the line pressure in accordance with these signal pressures, and supplies hydraulic oil to the primary pulley 100. The solenoid modulator pressure is also input to one end of the pinching control valve 106 as a signal pressure. A solenoid pressure adjusted by a linear solenoid valve (not shown) is input to the other end of the clamping pressure control valve 106. The line pressure is controlled by the relative relationship between the solenoid pressure and the solenoid modulator pressure, and hydraulic oil is supplied to the secondary pulley 101. The hydraulic pressure of the secondary pulley 101 is detected by a hydraulic pressure sensor 107 for feedback control of the belt clamping pressure.

In the case of the hydraulic control device having the above-described configuration, there is a problem that the shift behavior of the vehicle differs even if the same accelerator operation is performed due to the characteristic variation of the shift solenoid valves 103 and 104. FIG. 9 is a diagram showing the relationship between the duty ratio and the consumed flow rate in the solenoid valves 103 and 104. The consumption flow rate is the total flow rate of oil flowing through the solenoid valve (the sum of the oil amount supplied to the ratio control valve 105 and the drained oil amount). As shown in the figure, the consumption flow rate increases as the duty ratio increases, reaches the maximum flow rate in the vicinity of 50%, then decreases as the duty ratio increases, and becomes 100% at 100%. Since the solenoid valves 103 and 104 have variations in manufacturing characteristics, variations occur in the speed change characteristics. The solid line in FIG. 9 indicates the characteristics of the shift center (standard characteristics), the broken line indicates the shift upper limit characteristics, and the alternate long and short dashed line indicates the shift lower limit characteristics. The three characteristics differ from the upper limit to the lower limit of the shift in the duty ratio when the hydraulic pressure rises and the peak duty ratio. The range from 0% to the duty ratio when the hydraulic pressure rises is the dead zone. Since the consumption flow rate is large even at a relatively small duty ratio at the upper limit of the shift, shift hunting is likely to occur. On the other hand, at the lower limit of shift, the consumption flow rate is small even at the same duty ratio, so that the shift response is deteriorated. Generally, since setting is made so that hunting does not occur in all the shift control from the upper limit to the lower limit of the shift, the shift response is likely to be sacrificed.

Variations in the characteristics of the solenoid valve for speed change affect not only the speed change characteristic but also the belt clamping pressure characteristic. The belt clamping pressure is determined by the clamping pressure control valve 106 based on the balance between the solenoid modulator pressure, the spring force, and the solenoid pressure of the linear solenoid valve. The solenoid modulator pressure decreases when the flow rate of the shifting solenoid valve is large. There is a nature to do. Therefore, when the shift solenoid valve has a shift upper limit characteristic, the solenoid modulator pressure input to one end of the clamping pressure control valve 106 is relatively lower than when the shift lower limit characteristic has a shift limitation characteristic. The hydraulic pressure balance of the valve 106 changes, and the hydraulic pressure of the secondary pulley 101 also changes. As a result, the followability to the target belt clamping pressure may be deteriorated.

Patent Document 2 has an upshift valve and a downshift valve for supplying and discharging hydraulic oil to and from a primary pulley. While holding the downshift valve in the shutoff state, the upshift valve is switched from the shutoff state to the communication state, and Learn the drive current when the shift valve switches to the communication state, and similarly switch the downshift valve from the cutoff state to the communication state while holding the upshift valve in the cutoff state, and the downshift valve switches to the communication state A method for learning the driving current at the time is disclosed.

In the case of Patent Document 2, since it is difficult to perform the learning as described above in the traveling state, it is actually performed when the vehicle in the N range is stopped (see paragraph 0033). However, if the transmission is such that the drive wheels and the transmission are separated in the N range, that is, a transmission having a clutch downstream from the transmission mechanism, a shift can be made and learning as described above is possible. However, this is not possible with a transmission that does not have a clutch downstream from the transmission mechanism. Furthermore, since the relationship among the solenoid valve for controlling the upshift valve / downshift valve, the solenoid modulator valve, and the clamping pressure control valve is not shown, the above-described variation in the characteristics of the shifting solenoid valve causes the belt clamping pressure. There is no problem of affecting the characteristics.

Patent Document 3 discloses a shift control valve that controls the flow rate of hydraulic oil supplied to a primary pulley, hydraulic pressure detection means that detects the pressure of hydraulic oil in the primary pulley, and a shift control valve based on the detection result of the hydraulic pressure of the primary pulley. There is disclosed a shift control device provided with means for diagnosing the flow rate characteristic and correction means for correcting the flow rate characteristic of the shift control valve. In this speed change control device, the characteristics of the speed change control valve are corrected by detecting the hydraulic pressure of the primary pulley by forcibly driving the speed change control valve when the pulley is stopped.

However, since this shift control apparatus directly detects and corrects the hydraulic pressure of the primary pulley that is the control target of the shift control valve, a special hydraulic sensor is required to detect the hydraulic pressure of the primary pulley. In a typical continuously variable transmission, the hydraulic oil supplied to the primary pulley shifts by controlling the flow rate, not its hydraulic pressure, so a hydraulic sensor that has only a function for correcting the characteristics of the shift control valve is installed. Providing it separately has the disadvantage of increasing costs. Furthermore, there is no suggestion that the variation of the solenoid valve for speed change affects the belt clamping characteristics.

JP 2002-181175 A JP 2007-263261 A Japanese Patent Laid-Open No. 10-311416

An object of the present invention is to provide a control device capable of reducing variations in belt clamping pressure characteristics in a belt-type continuously variable transmission that controls the flow rate of hydraulic oil supplied to a primary pulley using a duty solenoid valve for shifting. is there.
Another object of the present invention is to provide a control device that can also suppress variation in transmission characteristics.

In order to achieve the above object, the present invention is a belt-type continuously variable transmission that includes a primary pulley and a secondary pulley around which a belt is wound, and in which both pulleys are provided with oil chambers for operating a movable sheave. Te, a shift control valve for controlling the flow rate hydraulic oil supplied to the oil chamber of the primary pulley, and a constant pressure valve for pressurizing regulated to a constant pressure hydraulic pressure supplied from the hydraulic source, the constant pressure valve according to the duty ratio that is input An upshift solenoid valve and a downshift solenoid valve that regulates a constant pressure supplied from the pressure generator and generates a signal pressure for controlling the shift control valve, and a consumption flow rate changes according to the duty ratio. a solenoid valve which causes a change in a constant pressure of the constant pressure valve as the original pressure, a constant pressure from the constant pressure valve is supplied as the first signal pressure, face the first signal pressure A second pressure signal is supplied from the clamping pressure control solenoid valve, and a clamping pressure control valve that controls the hydraulic pressure supplied to the oil chamber of the secondary pulley by balancing both signal pressures; and an oil chamber of the secondary pulley In a belt-type continuously variable transmission equipped with a hydraulic pressure sensor for detecting hydraulic pressure, when idling stops, one of the upshift solenoid valve and the downshift solenoid valve is held in a shut-off state, and a duty is input to the other Solenoid valve driving means for changing the ratio from a shut-off state to an open state by changing the ratio at a constant time gradient, and the detected value of the hydraulic sensor when the duty ratio is changed at a constant time gradient, to the duty ratio The rising hydraulic pressure storage means for storing the hydraulic pressure increase of the secondary pulley, the stored increase and duty ratio Based on the relationship, to provide a control apparatus for a belt-type continuously variable transmission, characterized in that a, a first correcting means for correcting the command signal to the clamping pressure control solenoid valve.

There is a problem that the hydraulic characteristics vary due to variations in individual solenoid valves for upshifting and downshifting used for gear shifting, and the gear shifting characteristics also vary. Therefore, in the present invention, learning is performed when the idling is stopped. The reason for performing learning when the vehicle is idling is that the secondary pressure does not change due to factors other than the operation of the shift solenoid valve. Note that learning may be performed not only in the non-traveling range such as the P range but also in the traveling range such as the D range, but the signal pressure or line pressure of the linear solenoid valve input to the clamping pressure control valve changes, and the secondary pressure It is preferable to learn in the non-traveling range.

In this state, one of the shift solenoid valves is turned OFF (duty ratio 0%), and the other duty ratio is changed from 0% to 100% with a constant time gradient, that is, sweeped up. The reason for sweeping is to reduce the response delay of the output pressure of the constant pressure valve and the secondary pressure with respect to the change of the duty ratio. Since the flow rate of the solenoid valve for shifting increases as the duty ratio changes, the output pressure of the constant pressure valve that supplies the original pressure also changes. Since the output pressure of the constant pressure valve is also supplied as a signal pressure to the clamping pressure control valve, the hydraulic pressure of the secondary pulley, which is the output pressure of the clamping pressure control valve, also changes. Generally, when the output pressure of the constant pressure valve decreases, the hydraulic pressure of the secondary pulley increases. Since the hydraulic pressure of the secondary pulley can be detected by a hydraulic pressure sensor, the duty ratio and the increase amount of the secondary pulley hydraulic pressure are stored in correspondence.

Thereafter, the other shifting solenoid valve is turned OFF (duty ratio 0%), the duty ratio of one shifting solenoid valve is changed from 0% to 100% with a constant time gradient, and the hydraulic pressure of the secondary pulley at that time is changed. The increase is detected by a hydraulic sensor, and the increase and the duty ratio are stored in correspondence.

Thus, the amount of increase in the secondary pulley determined in correspondence with the duty ratio of each shifting solenoid valve is stored as a learning value, and the command signal to the clamping pressure control solenoid valve is corrected based on this learning value. As a result, the amount of clamping pressure that rises during operation of the shift solenoid valve (duty solenoid valve) can be canceled by learning correction, the feed-forward correction accuracy can be improved, the feedback gain can be reduced, and a stable belt Nipping pressure control can be performed. In other words, since the feedback control feeds back the difference between the actual clamping pressure and the target clamping pressure, if the difference is reduced by feedforward, the actual clamping pressure and the target clamping pressure do not deviate even when feedback is not activated. In other words, the feedback becomes smaller by feeding back the small difference. Increasing the feedback gain approaches the target accordingly, but on the other hand, convergence may deteriorate and hunting may occur. Therefore, if feedback is provided for the divergence reduced by feedforward, it is possible to control the actual clamping pressure with good stability and small divergence without setting the gain too large.

Further, a duty ratio storage means for storing a duty ratio at which the detected value detected by the hydraulic pressure sensor becomes a maximum value, and a difference between the stored duty ratio and a preset standard duty ratio, an upshift solenoid Second correction means for correcting the duty ratio to the valve and the downshift solenoid valve may be provided. That is, the duty ratio when the amount of increase in the hydraulic pressure of the secondary pulley is maximized is obtained, the difference between this duty ratio and a preset standard duty ratio is obtained, and the command duty ratio of the shift solenoid valve is calculated using the difference. Is corrected, the speed change characteristics can be made uniform even if there is individual variation in the speed change solenoid valve.

The oil pressure sensor used in the present invention can be shared with the oil pressure sensor used for feedback control of the oil pressure of the secondary pulley, that is, the belt clamping pressure, so that no special oil pressure sensor is required.

As described above, according to the present invention, variations in belt clamping pressure characteristics caused by variations in the characteristics of the solenoid valve for shifting can be learned and corrected without adding a special device, so that individual variations in the solenoid valve for shifting occur. Even if it exists, stable belt clamping pressure control can be implemented.

In addition, since the command duty ratio to the speed change solenoid valve can be corrected using the difference between the duty ratio at which the detected increase in belt clamping pressure becomes the maximum value and the standard duty ratio, individual variations in the speed change solenoid valve can be corrected. The variation in the speed change characteristics due to the shift can be suppressed, and the speed change feeling and the fuel efficiency can be improved.

Furthermore, it is not necessary to add a special sensor or the like, and an existing hydraulic circuit can be used as it is, so that it can be configured simply and inexpensively.

1 is a skeleton diagram showing a configuration of a vehicle equipped with a continuously variable transmission according to the present invention. It is a circuit diagram showing an example of a hydraulic control device of a continuously variable transmission according to the present invention. It is an enlarged view of the principal part of the hydraulic control apparatus shown in FIG. It is a figure which shows the change of the secondary pressure (belt clamping pressure) when sweeping the solenoid valve for transmission. FIG. 5 is a diagram illustrating a relationship between a learning value (a rise in belt clamping pressure) obtained from FIG. 4 and a duty ratio. It is a flowchart figure of the learning and correction method which concerns on 1st Embodiment of this invention. It is a flowchart figure of the learning and correction method which concerns on 2nd Embodiment of this invention. It is the schematic of an example of the hydraulic circuit of the conventional belt-type continuously variable transmission. It is a figure which shows the relationship between the duty ratio of the solenoid valve for transmission, and consumption flow volume.

FIG. 1 shows an example of the configuration of a vehicle equipped with a belt type continuously variable transmission according to the present invention. An output shaft 1 a of the engine 1 is connected to a drive shaft (output shaft) 32 via a continuously variable transmission 2. The continuously variable transmission 2 is provided with a torque converter 3, a transmission 4, a hydraulic control device 7, an oil pump 6 driven by the engine 1, and the like.

The continuously variable transmission 2 includes a forward / reverse switching device 8, a primary pulley 11, a secondary pulley 21, and a V that is wound between the pulleys and transmits the rotation to the primary shaft 10 by switching the rotation of the turbine shaft 5 of the torque converter 3 between forward and reverse. The transmission 4 includes a belt 15, a differential device 30 that transmits the power of the secondary shaft 20 to the drive shaft 32, and the like. The turbine shaft 5 and the primary shaft 10 are arranged on the same axis, and the secondary shaft 20 and the drive shaft 32 are arranged parallel to the turbine shaft 5 and non-coaxially. Therefore, the continuously variable transmission 2 has a three-axis configuration as a whole. The V belt 15 used here is a known compression drive type metal belt.

The forward / reverse switching device 8 includes a planetary gear mechanism 80, a reverse brake B1, and a direct coupling clutch C1. The reverse brake B1 and the direct coupling clutch C1 are wet multi-plate brakes and clutches, respectively. A sun gear 81 of the planetary gear mechanism 80 is connected to the turbine shaft 5 as an input member, and a ring gear 82 is connected to the primary shaft 10 as an output member. The planetary gear mechanism 80 is a single pinion system, the reverse brake B1 is provided between the carrier 84 supporting the pinion gear 83 and the transmission case, and the direct coupling clutch C1 is provided between the carrier 84 and the sun gear 81. When the direct clutch C1 is released and the reverse brake B1 is engaged, the rotation of the turbine shaft 5 is reversed, decelerated and transmitted to the primary shaft 10, and the drive shaft 32 passes through the secondary shaft 20 in the same direction as the engine rotation direction. Since it rotates, it will be in a forward running state. Conversely, when the reverse brake B1 is released and the direct clutch C1 is engaged, the carrier 84 and the sun gear 81 rotate together, so that the turbine shaft 5 and the primary shaft 10 are directly connected, and the drive shaft 32 passes through the secondary shaft 20. Rotates in the direction opposite to the engine rotation direction, so that the vehicle travels backward.

The primary pulley 11 includes a fixed sheave 11a integrally formed on the primary shaft 10, and a movable sheave 11b supported on the primary shaft 10 so as to be axially movable and integrally rotatable. A cylinder 12 fixed to the primary shaft 10 is provided behind the movable sheave 11 b, and an oil chamber 13 is formed between the movable sheave 11 b and the cylinder 12. Shift control is performed by controlling the flow rate of the hydraulic oil supplied to the oil chamber 13 with ratio control valves 76 and 77 described later.

The secondary pulley 21 includes a fixed sheave 21a formed integrally on the secondary shaft 20, and a movable sheave 21b supported on the secondary shaft 20 so as to be axially movable and integrally rotatable. A piston 22 fixed to the secondary shaft 20 is provided behind the movable sheave 21 b, and an oil chamber 23 is formed between the movable sheave 21 b and the piston 22. By controlling the hydraulic pressure (secondary pressure) supplied to the oil chamber 23, a belt clamping pressure necessary for torque transmission is applied. In addition, a bias spring 24 for providing an initial clamping pressure is disposed in the oil chamber 23. A hydraulic pressure sensor 108 (see FIG. 2) that detects the secondary pressure is provided in the supply oil passage in the vicinity of the oil chamber 23 of the secondary pulley 21.

One end portion of the secondary shaft 20 extends toward the engine side, and the output gear 27 is fixed to this end portion. The output gear 27 meshes with the ring gear 31 of the differential device 30, and power is transmitted from the differential device 30 to the drive shaft 32 extending left and right to drive the wheels.

The continuously variable transmission 2 is controlled by an electronic control unit 100 (see FIG. 1). The electronic control unit 100 includes an engine speed sensor 101, a vehicle speed (or secondary pulley speed) sensor 102, a throttle opening (or accelerator opening) sensor 103, a shift position sensor 104, a primary pulley speed sensor 105, a brake signal. Detection signals are input from the sensor 106, the hydraulic oil temperature sensor 107 of the CVT, and the hydraulic pressure sensor 108 that detects the secondary pressure. Of course, other signals may be input as the input signal. The pulley ratio can be detected based on detection signals from the primary pulley rotation speed sensor 105 and the vehicle speed sensor 102.

The electronic control device 100 controls a plurality of solenoid valves built in the hydraulic control device 7. The hydraulic control device 7 is connected to the oil pump 6, the oil chamber 13 of the primary pulley 11, the oil chamber 23 of the secondary pulley 21, the reverse brake B1, and the direct coupling clutch C1. The electronic control unit 100 determines the target primary rotational speed according to a shift map set in advance according to the vehicle speed and the throttle opening, and controls the solenoid valve in the hydraulic control unit 7 to thereby control the continuously variable transmission 2. The amount of oil supplied to the oil chamber 13 of the primary pulley 11 is controlled, and the primary rotational speed is feedback-controlled to the target value. Further, the belt transmission torque is obtained from the engine torque and the gear ratio, and the hydraulic pressure (secondary pressure) supplied to the oil chamber 23 of the secondary pulley 21 is set to the target value so that the minimum belt clamping pressure that does not cause belt slippage is obtained. And feedback control. At this time, the actual secondary pressure is detected by the hydraulic pressure sensor 108. The hydraulic control device 7 also has a function of controlling the hydraulic pressure supplied to the reverse brake B1 and the direct coupling clutch C1.

FIG. 2 is a hydraulic circuit diagram of an example of the hydraulic control device 7, and FIG. 3 is a hydraulic circuit diagram of the main part thereof. The regulator valve 71 is a valve that regulates the discharge pressure of the oil pump 6 to a predetermined line pressure P _L. The clutch modulator valve 72 is a pressure reducing valve that reduces the line pressure and outputs the original pressure of the linear solenoid valve SLS and the original clutch modulator pressure Pcm to the direct connection clutch C1 and the reverse brake B1. The garage shift valve 74 is a switching valve that switches the oil passage so that the supply pressure to the direct coupling clutch C1 and the reverse brake B1 can be transiently controlled during the garage shift. The manual valve 75 is a manually operated valve that is mechanically connected to the shift lever and switches the oil path to the direct clutch C1 and the reverse brake B1. Since the regulator valve 71, the clutch modulator valve 72, the garage shift valve 74, and the manual valve 75 described above are valves that are not directly related to the present invention, detailed description thereof is omitted. In FIG. 2, only the hydraulic circuit relating to the primary pulley 11, the secondary pulley 21, the reverse brake B1 and the direct coupling clutch C1 is shown, but the hydraulic circuit such as the lockup clutch 3a built in the torque converter 3 is omitted. Yes.

The upshift ratio control valve 76 and the downshift ratio control valve 77 are examples of the shift control valve of the present invention. As will be described later, the upshift signal pressure Pds1 output from the solenoid valves DS1 and DS2 and the downshift signal control valve 77 This is a flow control valve that adjusts the amount of hydraulic oil supplied to and discharged from the primary oil chamber 13 based on the relative relationship with the signal pressure Pds2. The ratio check valve 78 switches the hydraulic oil to the primary oil chamber 13 from the flow rate control to the pressure control for closing control, and presets the ratio between the hydraulic pressure in the primary oil chamber 13 and the hydraulic pressure in the secondary oil chamber 23. It is a valve to keep in the established relationship.

SLS is a linear solenoid valve that outputs solenoid pressure Psls for performing line pressure control, transient control of the reverse brake B1 and direct coupling clutch C1, and pressure control of the oil chamber 23 of the secondary pulley 21. DS1 is an upshift duty solenoid valve that generates an upshift signal pressure Pds1, and DS2 is a downshift duty solenoid valve that generates a downshift signal pressure Pds2. The solenoid valves DS1 and DS2 are used not only for shift control but also for closing control and learning control described later. In this embodiment, the linear solenoid valve SLS is a normally open linear solenoid valve, and the solenoid valves DS1 and DS2 are both normally closed duty solenoid valves.

The shift solenoid valves DS1 and DS2 are controlled as follows according to the traveling state.

In Table 1, ○ indicates an operating state, and × indicates a non-operating state. In addition, (circle) and x contain not only an ON state or an OFF state but a duty control state. The closing control for turning off both solenoid valves simultaneously is the closing control at the vehicle speed = 0 and the lowest state, and is performed to prevent belt slippage at the time of restart. On the other hand, the closing control for turning on both solenoid valves is performed during the garage shift.

The solenoid modulator valve 73 is a constant pressure valve that regulates the clutch modulator pressure Pcm and generates a constant solenoid modulator pressure Psm corresponding to the spring load. The solenoid modulator pressure Psm is a source pressure of the duty solenoid valves DS1 and DS2, and is supplied as a signal pressure to the garage shift valve 74 and the pinching control valve 79.

The upshift ratio control valve 76 changes the valve opening area according to the relative relationship between the upshift signal pressure Pds1 and the downshift signal pressure Pds2, and adjusts the amount of oil supplied to the oil chamber 13 of the primary pulley 11. It is a control valve. That is, as shown in FIG. 3, the upshift ratio control valve 76 includes a spool 76b biased in one direction by a spring 76a, and a signal pressure Pds2 is applied to a signal port 76c on one end side where the spring 76a is accommodated. Is entered. The signal pressure Pds1 is input to the signal port 76d on the other end side facing the spring load. Line pressure P _L is supplied to the intermediate input port 76 e, and the output port 76 f is connected to the oil chamber 13 of the primary pulley 11. A port 76h connected to a port 78h of a ratio check valve 78 described later is formed between the input port 76e and the drain port 76g, and a downshift ratio control is provided between the output port 76f and the signal port 76d. A port 76 i connected to the port 77 f of the valve 77 and the port 78 d of the ratio check valve 78 is formed. When the upshift solenoid valve DS1 is turned on (100%) and the downshift solenoid valve DS2 is turned off (0%), the spool 76b is switched to the left side in FIG. 3 and the oil chamber of the primary pulley 13 via the ports 76e and 76f. When the downshift solenoid valve DS2 is turned on (100%) and the upshift solenoid valve DS1 is turned off (0%), the spool 76b is switched to the right side of FIG. Is discharged through ports 76f and 76i.

The downshift ratio control valve 77 also changes the valve opening area according to the relative relationship between the upshift signal pressure Pds1 and the downshift signal pressure Pds2, and adjusts the amount of oil discharged from the oil chamber 13 of the primary pulley 11. It is a control valve. The downshift ratio control valve 77 includes a spool 77b biased in one direction by a spring 77a, and a signal pressure Pds1 is input to a signal port 77c on one end side in which the spring 77a is accommodated. The signal pressure Pds2 is inputted to the signal port 77d on the other end side facing the spring load. In the intermediate portion, a drain port 77e, a port 77f connected to the port 76i of the upshift ratio control valve 76, and a port 77g connected to the port 78f of the ratio check valve 78 are formed in this order. When the upshift solenoid valve DS1 is turned on and the downshift solenoid valve DS2 is turned off, the spool 77b is switched to the left side in FIG. 3, and the port 77f communicates with the port 77g. When the downshift solenoid valve DS2 is turned on and the upshift solenoid valve DS1 is turned off, the spool 77b is switched to the right side in FIG. 3, and the hydraulic oil of the primary pulley 13 is drained through the ports 77f and 77e.

The ratio check valve 78 is a valve for maintaining the primary pressure and the secondary pressure in a preset relationship by switching the oil pressure of the oil chamber 13 of the primary pulley 11 from the flow rate control to the pressure control during the closing control. It is. That is, the primary pressure is controlled so that the ratio between the primary pressure and the secondary pressure has a predetermined relationship, and the valve maintains the predetermined speed ratio. The ratio check valve 78 includes a spool 78b biased in one direction by a spring 78a, and the hydraulic pressure of the secondary pulley oil chamber 23 is input to a signal port 78c on one end side in which the spring 78a is accommodated. The oil pressure of the primary oil chamber 13 is input to the signal port 78d on the other end side facing the spring load via the ports 76f and 76i of the upshift ratio control valve 76. Line pressure P _L is supplied to the input port 78e, and the output port 78f is connected to the port 77g of the downshift ratio control valve 77. Further, a port 78h connected to the port 76h of the upshift ratio control valve 76 is formed between the output port 78f and the drain port 78g. The supply oil passage connecting the ratio check valve 78 and the primary oil chamber 13 passes through the upshift ratio control valve 76 and the downshift ratio control valve 77, and has a small diameter in the oil passage between the ports 78f and 77g. A simple orifice 90 is set. Due to the action of these orifices 90, a sudden speed change is prevented when switching to the closing control.

The clamping pressure control valve 79 is a valve for controlling the hydraulic pressure (secondary pressure) of the hydraulic oil chamber 23 of the secondary pulley 21. A spool 79g biased in one direction by a spring 79f is provided, and a constant pressure Psm is supplied from a solenoid modulator valve 73 to a signal port 79a on one end side facing the spring load. Line pressure P _L is supplied to the input port 79b, the output port 79c is connected to the hydraulic oil chamber 23 of the secondary pulley 21, and the secondary pressure is fed back to the port 79d. The solenoid pressure Psls is supplied from the linear solenoid valve SLS to the signal port 79e on the other end side in which the spring 79f is accommodated. Port 79h is a drain port. Therefore, the hydraulic pressure obtained by amplifying the solenoid pressure Psls input to the signal port 79e with a predetermined amplification degree can be supplied to the hydraulic oil chamber 23 of the secondary pulley 21 as a secondary pressure. The hydraulic pressure (secondary pressure) in the hydraulic oil chamber 23 is detected by the hydraulic pressure sensor 108, and the belt clamping pressure or the belt transmission torque can be obtained based on the detected hydraulic pressure.

-First embodiment-
Next, a first embodiment of the present invention will be described. As shown in FIG. 9, when the shift solenoid valve DS1 or DS2 is operated, the flow rate of consumption changes, so that the solenoid modulator pressure Psm, which is the original pressure, decreases. The solenoid modulator pressure Psm is also input to the signal port 79a of the pinching control valve 79. When the solenoid modulator pressure Psm decreases, the spool 79g (see FIG. 3) moves upward from the pressure adjustment position, and the secondary pulley 23 The secondary pressure supplied to is higher than usual. In addition, the relationship between the increase and the duty ratio varies depending on the individual solenoid valves DS1 / DS2. As a result, variations in the characteristics of the shift solenoid valve also affect the belt clamping pressure control.

Therefore, in the present invention, in order to know the degree of influence of the belt clamping pressure due to variations in the characteristics of the speed change solenoid valve, one duty ratio of the speed change solenoid valves DS1 and DS2 is set to 0 with a constant time gradient as shown in FIG. Gradually change (sweep) to 100%, and obtain the change in secondary pressure (belt clamping pressure) at that time. When the duty ratios of the solenoid valves DS1 and DS2 are swept, the consumption flow rate is maximized in the vicinity of about 50%, the amount of decrease in the solenoid modulator pressure Psm is maximized, and the secondary pressure is also increased accordingly. Among the changes in the clamping pressure in FIG. 4, the thick line indicates the upper limit of the shift and the thin line indicates the lower limit of the shift. FIG. 5 shows the amount of increase in the secondary pressure detected and stored by the hydraulic sensor 108, and the amount of increase obtained as a learning value corresponding to each duty ratio.

FIG. 5 shows an example of the learning value (increase), but it has characteristics that are substantially similar to the characteristics of the consumed flow rate shown in FIG. That is, it rises from a small duty ratio at the upper limit of the shift and rises later at the lower limit of the shift. The maximum value is obtained when the duty ratio is about 50%, and is 0 when the duty ratio is 100%. If the command signal to the linear solenoid valve is corrected based on the learning value, the variation in belt clamping pressure due to the variation in operation of the shifting solenoid valve DS1 or DS2 can be eliminated or suppressed.

Next, a procedure for learning and correcting variations in the secondary pressure (belt clamping pressure) due to the operation of the shifting solenoid valves DS1 / DS2 will be described below.

(1) With one solenoid valve DS1 / DS2 turned OFF (0%) in the idling state, the other solenoid valve DS2 / DS1 is slowly raised (sweep) from 0 to 100% with a constant time gradient. . The reason for sweeping is to reduce the response delay of the solenoid modulator pressure and the secondary pressure with respect to the change of the duty ratio. Also, the reason for the idling state in the stop state is learned by setting the solenoid pressure Psls supplied from the linear solenoid valve SLS to a constant state so that the secondary pressure does not change due to factors other than the operation of the shifting solenoid valve. Because. When the vehicle is idling, the vehicle speed and the throttle opening are equal to or less than a predetermined value near 0, and the engine speed is near the idle speed. In addition, it is desirable to perform sweeping not only in the idling state where the vehicle is stopped but also in the stable state of the vehicle (transmission oil temperature, engine water temperature, battery voltage, engine speed, etc. are stable). It is desirable that learning be performed in a stable state of the vehicle. This is because learning in an unstable state may promote variations in belt clamping characteristics.

(2) When the duty ratio of the solenoid valve is swept to 0 to 100%, the consumed flow rate increases, so that the solenoid modulator pressure Psm decreases according to the operation amount. The solenoid modulator pressure Psm is also used as a signal pressure for the garage shift valve 74 and the clamping pressure control valve 79 in addition to the shifting solenoid valves DS1 / DS2, but both are only supplied to the closed signal port. Therefore, it is considered that the solenoid modulator pressure Psm does not decrease due to reasons other than the operation of the shifting solenoid valves DS1 / DS2. Since the solenoid modulator pressure Psm is input as a signal pressure to the port 79a of the clamping pressure control valve 79, a decrease in the solenoid modulator pressure Psm shifts the pressure regulation position of the clamping pressure control valve 79, and the secondary pressure as shown in FIG. (Belt clamping pressure) changes. The change in the secondary pressure can be detected by the hydraulic pressure sensor 108, and the increase in the secondary pressure (belt clamping pressure) with respect to the duty ratio to the solenoid valve is stored as a learned value as shown in FIG.

(3) Thereafter, using the stored learning value, the command signal to the linear solenoid valve SLS that controls the belt clamping pressure is corrected. For example, if the increase in belt clamping pressure at the duty ratio Da is Δp and the target command signal (target current value) to the linear solenoid valve SLS is Ar, the corrected command signal Ac at the duty ratio Da is
Ac = Ar−Δp × k (k: coefficient)
Can be expressed as The correction method is not limited to the above formula. If the command signal to the linear solenoid valve SLS is corrected using this correction value, the solenoid pressure Psls decreases by an amount corresponding to the correction amount, and the pressure control state of the pinching control valve 79 is maintained. Therefore, highly accurate belt clamping pressure control that eliminates or suppresses the influence of the shift solenoid valve DS1 or DS2 can be performed.

A specific learning correction method will be described with reference to the flowchart shown in FIG. First, it is determined whether or not the vehicle is idling (step S1). Specifically, it is a non-traveling range such as the P range, and it is determined whether the vehicle is stopped and the engine is in a stable idling state. In this state, the continuously variable transmission is normally closed at the lowest level, and the linear solenoid pressure Psls is maintained at a constant pressure.

Next, the up-shift solenoid valve DS1 is turned off (step S2), and the instruction duty ratio to the down-shift solenoid valve DS2 is changed from 0% to 100% with a constant time gradient (step S3). That is, sweep output is performed. As a result, the valve opening of the solenoid valve DS2 increases, and an increase in the flow rate of the solenoid valve DS2 causes a decrease in the solenoid modulator pressure Psm.

Next, since the secondary pressure increases due to the decrease in the solenoid modulator pressure Psm, the increase is detected by the hydraulic sensor 108, and the increase is stored in the memory as a learning value for the duty ratio (step S4). Note that when the downshift solenoid valve DS2 is swept, even if the duty ratio is increased to 100%, only the hydraulic oil in the oil chamber 13 of the primary pulley is discharged and the low state is maintained. There is no.

After the learning correction is performed for the downshift solenoid valve DS2, the same is performed for the upshift solenoid valve DS1. That is, the downshift solenoid valve DS2 is turned off (step S5), and the duty ratio for the upshift solenoid valve DS1 is swept from 0% to 100% (step S6). Then, an increase in the secondary pressure during the sweep is detected by the hydraulic sensor 108, and the increase is stored in the memory as a learning value for the duty ratio (step S7). Finally, the upshift solenoid valve DS1 is turned OFF to complete the learning (step S8).

Here, the downshift solenoid valve DS2 is swept first, but the upshift solenoid valve DS1 may be swept first. Since the belt and pulley are both stopped when the vehicle is stopped, the belt cannot be shifted and shifted without applying a force greater than the friction force between the sheave and belt of the primary and secondary pulleys. Even if the duty ratio of the solenoid valve DS1 is increased to the maximum value, the speed is not shifted to the upshift side.

After the learning shown in FIG. 6A is completed, the belt clamping pressure value shown in FIG. 6B is corrected. That is, first, the target belt clamping pressure P is determined in accordance with the load torque, the gear ratio, and the like (step S9), and then the learned value ΔP corresponding to the input duty ratio to the gear solenoid valve is read from the memory (step S10). ). Next, the command clamping pressure (P−ΔP) corrected by the learning value ΔP is obtained (step S11), and the target command current I to the linear solenoid valve SLS corresponding to P−ΔP is determined and output (step S12). . As a result, it is possible to suppress variations in the belt clamping pressure caused by variations in the operation of the shifting solenoid valve.

-Second Embodiment-
In the first embodiment, as shown in FIG. 4, the increase in the secondary pressure when the solenoid valves DS1 and DS2 are swept up is learned corresponding to each duty ratio, but in the second embodiment, the maximum secondary pressure is learned. This is a method for learning and correcting the characteristic variation of the solenoid valves DS1 and DS2 (that is, the shift control variation) by detecting the duty ratio that causes the value.

For example, when the belt clamping pressure (secondary pressure) in FIG. 4 is detected, Dm is obtained as the duty ratio that provides the maximum value of the secondary pressure, the duty ratio Dm is compared with the standard duty ratio Dr, and the difference ΔD (= Dr -Dm) is stored in memory. At the time of an actual gear shift command, the difference duty ratio ΔD may be used to correct the entire command signal to the gear shift solenoid valve. For example, when the difference ΔD is a positive value, the characteristic of the shift solenoid valve is considered to be on the upper limit side of the shift, so if the command duty ratio to the shift solenoid valve is D, the corrected duty ratio D− By outputting ΔD (ΔD> 0), it is possible to suppress shift variation due to the shift solenoid valve. Conversely, when the difference ΔD is a negative value, the characteristics of the shift solenoid valve are considered to be on the lower limit side of the shift, so if the command duty ratio to the shift solenoid valve is D, the corrected duty ratio By outputting D−ΔD (ΔD <0), it is possible to suppress shift variation due to the shift solenoid valve.

FIGS. 7A and 7B are flowcharts of a learning correction method for shift control in the second embodiment. FIG. 7A shows the flow of learning. It is determined whether or not idling is stopped (step S20), one solenoid valve DS1 is turned off (step S21), and the other solenoid valve DS2 is swept (step S22). ). These steps S20-22 are the same as steps S1-3 of FIG. During the sweep, the duty ratio at which the secondary pressure becomes the maximum value is obtained, and the difference ΔD between this duty ratio and the standard value is stored in the memory (step S23).

Next, the other solenoid valve DS2 is turned OFF (step S24), and one solenoid valve DS1 is swept (step S25). This is also the same as steps S5 and S6 in FIG. During the sweep, the duty ratio at which the secondary pressure becomes the maximum value is obtained, and the difference ΔD between the duty ratio and the standard value is stored in the memory (step S26). Finally, the solenoid valve DS1 is turned off (step S27).

FIG. 7B shows a correction flow in actual upshift transmission control. First, the target input rotational speed is obtained based on the throttle opening, the vehicle speed, etc., and the target duty ratio D of the shift solenoid valve DS1 necessary for changing from the current rotational speed to the target input rotational speed is determined (step) S28). Next, the difference ΔD of the solenoid valve DS1 is read from the memory (step S29), and it is determined whether the difference is positive, negative, or near 0 (step S30). If the difference ΔD is a positive value, the solenoid valve DS1 has a shift upper limit characteristic, so that D−ΔD is output to the solenoid valve DS1 (step S31). If the difference ΔD is a negative value, Since the solenoid valve DS1 has a shift lower limit characteristic, D-ΔD is output to the solenoid valve DS1 (step S32). Further, if the difference ΔD is close to 0, the solenoid valve DS1 has the characteristics of the center of the shift, so the target duty ratio D is output as it is (step S33). FIG. 7B shows the correction for one shift solenoid valve DS1, but the other shift solenoid valve DS2 may be corrected similarly. In this way, the influence of the individual variation of the solenoid valve DS1 on the shift control can be suppressed.

The first embodiment and the second embodiment can be performed simultaneously. That is, in both embodiments, the amount of increase in the secondary pressure when the solenoid valve for shifting is swept is detected during learning, or the duty ratio when the secondary pressure reaches the maximum value is detected. This is because in the former, the belt clamping pressure is learned and corrected based on the relationship between the hydraulic pressure increase and the duty ratio, and in the latter, the shift solenoid valve itself is learned and corrected using the maximum duty ratio.

-Third embodiment-
In the second embodiment, the command signal to the shift solenoid valve is corrected using the difference ΔD (= Dr−Dm) between the duty ratio Dm that provides the maximum value of the secondary pressure and the standard duty ratio Dr. In the mode, the shift ratio upper limit, the shift center, and the shift lower limit are classified by comparing the duty ratio Dm that gives the maximum value of the secondary pressure with the standard duty ratio Dr.
When Dm <Dr-α: When shift upper limit Dm> Dr + α: When shift lower limit Dr-α ≦ Dm ≦ Dr + α: Shift center Here α is a discrimination margin, for example, duty ratio 1 What is necessary is just about 2%.

If it is determined that the characteristic of the target shift solenoid valve is any one of the shift upper limit, the shift center, and the shift lower limit, the actual command duty ratio may be corrected accordingly. For example, in the case of a shift solenoid valve at the upper limit of the shift, the command duty ratio is corrected to be lower by a predetermined amount, and conversely, in the case of the solenoid valve for shift at the lower limit of the shift, the command duty ratio is set to the predetermined amount. What is necessary is just to correct | amend so that it may become high.

In the above-described embodiment, two flow rate control valves, that is, the upshift ratio control valve 76 and the downshift ratio control valve 77 are used as the shift control valve. In addition, the signal pressures of the upshift solenoid valve DS1 and the downshift solenoid valve DS2 may be input to both ends in opposition to each other.

Further, although the linear solenoid valve SLS is used as the clamping pressure control solenoid valve, it is needless to say that a duty solenoid valve may be used.

1 Engine 2 Continuously Variable Transmission 7 Hydraulic Control Unit 11 Primary Pulley 13 Primary Oil Chamber 21 Secondary Pulley 23 Secondary Oil Chamber 73 Solenoid Modulator Valve 76 Upshift Ratio Control Valve 77 Downshift Ratio Control Valve 78 Ratio Check Valve 79 Nipping Pressure Control valve 108 Hydraulic sensor SLS Linear solenoid valve (solenoid valve for clamping control)
DS1 Solenoid valve for upshift DS2 Solenoid valve for downshift

Claims (2)

A belt-type continuously variable transmission comprising a primary pulley and a secondary pulley around which a belt is wound, and an oil chamber for operating a movable sheave on each of the pulleys, wherein the pulley is supplied to the oil chamber of the primary pulley. a shift control valve for controlling the flow rate hydraulic fluid that a constant pressure valve for pressurizing regulated to a constant pressure hydraulic pressure supplied from the hydraulic source, a constant pressure supplied from the constant pressure valve by regulating in accordance with the duty ratio that is input, the An upshift solenoid valve and a downshift solenoid valve that generate a signal pressure for controlling the speed change control valve, the flow rate of the change varies according to the duty ratio, and the constant pressure of the constant pressure valve is the original pressure. to a solenoid valve that causes a change, a constant pressure from the constant pressure valve is supplied as the first signal pressure, said first signal pressure opposed to pinching control solenoid valve A clamping pressure control valve that controls the hydraulic pressure supplied to the oil chamber of the secondary pulley by a balance between the two signal pressures, and a hydraulic pressure sensor that detects the hydraulic pressure of the oil chamber of the secondary pulley, In a belt-type continuously variable transmission equipped with
When idling stops, one of the upshift solenoid valve and the downshift solenoid valve is held in the shut-off state, and the duty ratio input to the other is changed with a constant time gradient to change from the shut-off state to the open state. Solenoid valve driving means;
From the detected value of the hydraulic pressure sensor when the duty ratio is changed at a constant time gradient, the rising hydraulic pressure storage means for storing the increase in the hydraulic pressure of the secondary pulley relative to the duty ratio;
A belt type continuously variable transmission comprising: a first correction unit that corrects a command signal to the clamping pressure control solenoid valve based on a relationship between the stored increase amount and a duty ratio. Control device. A duty ratio storage means for storing a duty ratio at which a detected value detected by the hydraulic sensor becomes a maximum value;
Second correction means for correcting a duty ratio to the upshift solenoid valve and the downshift solenoid valve using a difference between the stored duty ratio and a preset standard duty ratio is provided. The control device for a belt-type continuously variable transmission according to claim 1.

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