JP6836972B2 - Hydraulic control device for automatic transmission - Google Patents

Description

The present invention relates to an automatic transmission hydraulic control device, and more particularly to an automatic transmission hydraulic control device capable of switching the flow rate of hydraulic oil to a hydraulic circuit in at least two stages.

Conventionally, there is a hydraulic control device for an automatic transmission that reduces the load of an oil pump driven by an internal combustion engine (engine) to improve fuel economy (fuel efficiency) (see, for example, Patent Document 1). Patent Document 1 has a line pressure control valve capable of switching the pressure (line pressure) for supplying hydraulic oil to a main hydraulic circuit having a friction fastening element into two stages of low pressure and high pressure, and the low pressure can be reduced as needed. Control to increase. When the line pressure becomes low, the load of the oil pump that supplies the hydraulic oil to the line pressure control valve is reduced as compared with the case where the line pressure is high. Then, the load on the engine that drives the oil pump is also reduced, and fuel efficiency is improved.

Here, in order to output the required line pressure in the main hydraulic circuit, the flow rate of the hydraulic oil supplied from the oil pump to the main hydraulic circuit must be sufficient. However, if more hydraulic oil than necessary is constantly supplied to the high-pressure main hydraulic circuit, the load on the oil pump will increase. For this reason, conventionally, the hydraulic oil supplied from the oil pump to two hydraulic circuits composed of a high-pressure main hydraulic circuit having a clutch or the like and a low-pressure sub-hydraulic circuit composed of a lubrication system or the like. When the flow rate is estimated and it is determined that more flow rate than necessary is supplied to the main hydraulic circuit, a part of the hydraulic oil discharged from the oil pump is switched to flow to the sub-hydraulic circuit. As a result, the flow rate switching control was performed to reduce the flow rate from the oil pump to the high-pressure main hydraulic circuit and reduce the load on the oil pump. Then, the flow rate switching control was switched based on the estimated flow rate of the hydraulic oil from the oil pump to the main hydraulic circuit, and the estimated flow rate was calculated while the oil pump was operating.

However, in calculating the estimated flow rate, conventionally, the estimated flow rate of hydraulic oil discharged from the oil pump is determined based on the oil pump, which has a relatively low discharge capacity, in consideration of individual differences and durability deterioration of the oil pump. Was. In this case, even if the oil pump actually has a high discharge capacity, the flow rate switching control must be performed using the estimated flow rate of the oil pump having a low discharge capacity. Assuming an oil pump with a low discharge capacity, estimating the supply flow rate of the oil pump will supply more flow rate to the main hydraulic circuit than is actually required. As a result, the load on the oil pump becomes large, and there is a risk that fuel efficiency may deteriorate.

Japanese Unexamined Patent Publication No. 2002-089679

The present invention has been made in view of the above points, and an object of the present invention is to provide a hydraulic control device for an automatic transmission capable of reducing the load on an oil pump and improving fuel efficiency.

In order to solve the above problems, the flood control device for an automatic transmission according to the present invention is an automatic transmission having a friction engaging element (24c) that transmits power by engagement and transmitting power at a predetermined gear ratio. , The oil pressure of the automatic transmission including the oil pressure supply mechanism (46) for supplying the oil pressure for engaging the friction engagement element (24c) and the control means (91) for controlling the oil pressure supply mechanism (46). In the control device, the hydraulic supply mechanism (46) requires a main hydraulic circuit (47) having a friction engaging element (24c) and requiring a relatively high hydraulic pressure, and a relatively low hydraulic pressure. A sub-hydraulic circuit (48), a hydraulic supply source (46a1, 46a2) driven by an internal combustion engine (10) to supply hydraulic oil to the main hydraulic circuit (47) and the sub-hydraulic circuit (48), and a hydraulic supply source (46a1). , 46a2), the first state (lubricating mode) in which the flow rate supplied from the main hydraulic circuit (47) is relatively small, and the flow rate supplied from the hydraulic supply sources (46a1, 46a2) to the main hydraulic circuit (47) The flow rate adjusting means (46 g) that can be switched between the relatively large second state (high pressure mode) and at least two stages, and the flow rate of hydraulic oil to the main hydraulic circuit (47) in the first state (lubricating mode). An estimated supply flow rate (Qs), which is an estimated value, and an estimated consumption flow rate (Qc), which is an estimated value of the flow rate of hydraulic oil consumed in the main hydraulic circuit (47) in the first state (lubricating mode). The control means (91) includes a flow rate estimation means (92) for estimating and a slip detection means (93) that outputs a slip detection signal (SI) when the slip of the friction engaging element (24c) is detected. When the estimated supply flow rate (Qs) is lower than the estimated consumption flow rate (Qc), the first state (lubrication mode) is switched to the second state (high pressure mode), and the estimated supply flow rate (Qs) becomes the estimated consumption flow rate (Qc). ) Is exceeded, the flow rate switching control for switching from the second state (high pressure mode) to the first state (lubricating mode) is performed, and the slip detection signal (SI) is transmitted from the slip detection means (93) during the flow switching control. If it is not output, the first state increase correction that increases the first state (lubricating mode) by performing at least one of the correction that increases the estimated supply flow rate (Qs) and the correction that decreases the estimated consumption flow rate (Qc). It is characterized by performing.

In this way, slip detection of the friction engaging element is performed at the time of flow rate switching control, and when the slip detection signal is not output, sufficient operation for engagement of the friction engaging element is performed at the time of flow rate switching control. The control means determines that the flow rate of oil can be supplied. Therefore, the control means performs the first state increase correction. Then, the time for the control means to select the first state can be increased, and the flow rate supplied from the hydraulic supply source to the high-pressure main hydraulic circuit is reduced, so that the hydraulic supply source and the internal combustion engine driving the hydraulic supply source are affected. Such a load can be reduced. As a result, fuel efficiency can be improved.

Further, in the hydraulic control device of the automatic transmission, the control means (91) reduces the estimated supply flow rate (Qs) when the slip detection signal (SI) is output from the slip detection means (93). Alternatively, the first state reduction correction for reducing the first state (lubrication mode) may be performed by performing at least one of the corrections for increasing the estimated consumption flow rate (Qc).

When the slip detection signal is output in this way, the control means determines that a sufficient flow rate of hydraulic oil cannot be supplied for the engagement of the friction engaging element. Therefore, the control means performs the first state reduction correction. Then, the time for the control means to select the first state is reduced, and the time for selecting the second state is increased, so that a sufficient flow rate of the hydraulic oil is supplied to the main hydraulic circuit.

Further, in the hydraulic control device of the automatic transmission, the control means (91) corrects the first state increase at the time of flow rate switching control when the slip detection signal (SI) is not output from the slip detection means (93). Then, when the slip detection signal (SI) is output from the slip detection means (93), the first state decrease correction may be performed and the subsequent first state increase correction may be stopped.

As described above, before the slip detection signal is output, the first state increases each time the first state increase correction is performed, so that the flow rate supplied to the main hydraulic circuit can be reduced. Further, when the slip detection signal is output, the first state reduction correction can be performed to return to the estimated flow rate immediately before the slip detection signal is output, that is, the estimated supply flow rate or the estimated consumption flow rate. The estimated flow rate can be considered as a value that does not detect slippage and is most in line with the actual flow rate. Here, by stopping the subsequent first state increase correction, the estimated flow rate is not changed, and by performing the flow rate switching control based on the estimated flow rate, the flow rate switching control close to the actual flow rate for the main hydraulic circuit is performed. Is possible. Since the first state reduction correction is not stopped even after the slip detection signal is output once, the characteristics of the hydraulic supply source change due to continuous use of the hydraulic supply source, and the slip detection signal is output again. By performing the first state reduction correction again, it is possible to control the flow rate switching close to the actual flow rate.

Further, in the hydraulic control device of the automatic transmission, the main hydraulic circuit (47) is a first hydraulic oil supply unit and a first hydraulic oil supply unit to which hydraulic oil is supplied from the hydraulic supply sources (46a1, 46a2). The second hydraulic oil supply unit may be provided with a second hydraulic oil supply unit to which the surplus hydraulic oil is supplied, and the second hydraulic oil supply unit may supply the oil pressure to the friction engaging element (24c).

In this way, when the hydraulic oil is supplied to the friction engaging element after the first hydraulic oil supply unit, the friction engaging element in the second hydraulic oil supply unit to which the hydraulic oil is supplied later However, the flow rate of the hydraulic oil is more likely to be insufficient than that of the first hydraulic oil supply unit. Here, by detecting the slip of the friction engaging element and determining whether or not the flow rate of the hydraulic oil is insufficient, it is possible to prevent the flow rate of the hydraulic oil from being insufficient in the first hydraulic oil supply unit. Further, by detecting the second hydraulic oil supply unit where the flow rate shortage is likely to occur, the flow rate shortage can be detected quickly, so that the accuracy of the balance calculation of the hydraulic oil flow rate can be improved.

Further, in the hydraulic control device of the automatic transmission, the friction engaging element (24c) may be a lockup clutch of a torque converter.

As described above, assuming that the friction engaging element (24c) is a lockup clutch of the torque converter, when the first hydraulic oil supply unit is set to, for example, the main transmission mechanism of the automatic transmission, the main transmission mechanism is used. It is possible to prevent insufficient flow of hydraulic oil. Further, by detecting the slip of the lockup clutch of the torque converter, which tends to cause a flow rate shortage, the flow rate shortage can be detected quickly, so that the accuracy of the hydraulic oil flow rate balance calculation can be improved.

The reference numerals in parentheses above indicate the reference numerals of the corresponding components of the embodiments described later as an example of the present invention.

According to the hydraulic control device for an automatic transmission according to the present invention, the load on the oil pump can be reduced and fuel efficiency can be improved.

It is a figure which shows the whole configuration example of the vehicle which includes the hydraulic control device of the automatic transmission which concerns on this embodiment. It is a hydraulic circuit diagram of a hydraulic supply mechanism. It is a block diagram of a hydraulic control device. It is a flowchart about the determination method of the estimated flow rate. It is a flowchart about estimated flow rate correction. It is a figure which shows the example which performed the increase correction of the estimated supply flow rate at the time of switching from a lubrication mode to a high pressure mode. It is a figure which shows the example which performed the reduction correction of the estimated supply flow rate when the slip detection signal was detected. It is a figure which shows the example which performed the increase correction of the estimated supply flow rate at the time of switching from a high pressure mode to a lubrication mode. It is a figure which shows the example which performed the reduction correction of the estimated consumption flow rate at the time of switching from a lubrication mode to a high pressure mode.

Hereinafter, embodiments of the present invention will be described in detail with reference to the accompanying drawings. FIG. 1 is a diagram showing an overall configuration example of a vehicle including a hydraulic control device for an automatic transmission according to the present embodiment. The vehicle shown in the figure includes an engine 10 (internal combustion engine) as a drive source, a torque converter 24, and a continuously variable transmission 26 (CVT) that shifts and outputs rotation due to the driving force of the engine 10. , A forward / backward switching device 28 is provided. The forward / reverse switching device 28 includes a forward clutch 28a provided for connecting / disconnecting the transmission of the driving force of the engine 10 to the continuously variable transmission 26. The vehicle also includes an engine controller 66 and a shift controller 90, which are control devices for controlling the engine 10, the continuously variable transmission 26, and the forward / backward switching device 28.

The throttle valve (not shown) arranged in the intake system of the engine 10 is a DBW mechanism 16 composed of an actuator such as an electric motor whose mechanical connection with the accelerator pedal arranged on the floor surface of the driver's seat of the vehicle is cut off. It is connected to (Drive By Wire mechanism) and opened and closed by the DBW mechanism 16.

The intake air metered by the throttle valve flows through the intake chamber (not shown) and mixes with the fuel injected from the injector 20 near the intake port of each cylinder to form an air-fuel mixture, and the intake valve (figure). When the valve (not shown) is opened, it flows into the combustion chamber (not shown) of the cylinder. The air-fuel mixture is ignited and burned in the combustion chamber, drives the piston to rotate the crankshaft 22, and then becomes exhaust gas and is discharged to the outside of the engine 10.

The crankshaft 22 of the engine 10 is connected to the pump impeller 24a of the torque converter 24, while the turbine runner 24b, which is arranged to face the pump impeller 24a and receives the fluid (hydraulic oil), is connected to the main shaft MS (input shaft). To. As a result, the rotation of the crankshaft 22 is input to the torque converter 24. Further, the torque converter 24 has a lockup clutch 24c (friction engagement element).

Further, the rotation of the crankshaft 22 is input to the continuously variable transmission 26 via the torque converter 24. The continuously variable transmission 26 includes a drive pulley 26a arranged on the main shaft MS, more accurately, the outer peripheral side shaft thereof, and a counter shaft CS (output shaft) parallel to the main shaft MS, more accurately, the counter shaft CS. It is composed of a driven pulley 26b arranged on the outer peripheral side shaft of the above, and an endless flexible member, for example, a metal belt 26c, which is hung between them.

The drive pulley 26a is divided into a fixed pulley half body 26a1 which is arranged so that it cannot rotate relative to the outer peripheral side shaft of the main shaft MS and cannot move in the axial direction, and a fixed pulley half body 26a1 which cannot rotate relative to the outer peripheral side shaft of the main shaft MS. On the other hand, it is composed of a movable pulley half body 26a2 that can move relative to the axial direction. The driven pulley 26b has an axial direction with respect to a fixed pulley half body 26b1 which is arranged on the outer peripheral side shaft of the counter shaft CS so as not to rotate in the axial direction and a fixed pulley half body 26b1 which cannot rotate relative to the counter shaft CS and cannot move in the axial direction. It consists of a movable pulley half body 26b2 that can move relative to each other.

The continuously variable transmission 26 is connected to the engine 10 via the forward / backward switching device 28. The forward / backward switching device 28 includes a forward clutch 28a that enables the vehicle to travel in the forward direction, a reverse brake clutch 28b that enables the vehicle to travel in the reverse direction, and a planetary gear mechanism 28c arranged between them. The continuously variable transmission 26 is connected to the engine 10 via a forward clutch 28a.

In the planetary gear mechanism 28c, the sun gear 28c1 is fixed to the main shaft MS, and the ring gear 28c2 is fixed to the fixed pulley half body 26a1 of the drive pulley 26a via the forward clutch 28a. A pinion 28c3 is arranged between the sun gear 28c1 and the ring gear 28c2. The pinion 28c3 meshes with the sun gear 28c1 and is integrally formed with the carrier 28c4. The carrier 28c4 is locked by the reverse brake clutch 28b when it is activated.

The rotation of the counter shaft CS is transmitted from the secondary shaft SS (intermediate shaft) to the drive wheels 12 via gears. That is, the rotation of the counter shaft CS is transmitted to the secondary shaft SS via the gears 30a and 30b, and the rotation is transmitted from the differential 32 to the drive shaft 34 via the gear 30c, and finally the left and right drive wheels 12 (right side). Only shown).

Disc brakes 36 are arranged in the vicinity of the driving wheels 12 (front wheels) and the driven wheels (rear wheels) (not shown). A brake pedal 40 and an accelerator pedal 56 are arranged on the floor surface of the driver's seat of the vehicle. A brake switch 40a is provided near the brake pedal 40. The brake switch 40a outputs an on signal in response to the operation of the driver's brake pedal 40. Further, an accelerator opening sensor 56a is provided in the vicinity of the accelerator pedal 56. The accelerator opening sensor 56a outputs a signal proportional to the accelerator opening corresponding to the driver's accelerator pedal operation amount.

In the forward / backward switching device 28, the driver operates the range selector 44 provided in the driver's seat of the vehicle to switch between the forward clutch 28a and the reverse brake clutch 28b, for example, one of the ranges such as P, R, N, and D. It is done by selecting. The range selection by the operation of the range selector 44 by the driver is transmitted to the manual valve of the hydraulic supply mechanism 46. A range selector switch 44a is provided in the vicinity of the range selector 44. The range selector switch 44a outputs a signal corresponding to a range such as P, R, N, D selected by the driver.

When, for example, the D, S, L range is selected via the range selector 44, the spool of the manual valve moves accordingly, and the hydraulic oil (flood) is discharged from the piston chamber of the reverse brake clutch 28b, while Flood is supplied to the piston chamber of the forward clutch 28a to engage the forward clutch 28a.

When the forward clutch 28a is engaged, all gears rotate integrally with the main shaft MS, and the drive pulley 26a is driven in the same direction (forward direction) as the main shaft MS. Therefore, the vehicle travels in the forward direction.

When the R range is selected, hydraulic oil is discharged from the piston chamber of the forward clutch 28a, while oil is supplied to the piston chamber of the reverse brake clutch 28b to operate the reverse brake clutch 28b.

When the P or N range is selected, hydraulic oil is discharged from both piston chambers, both the forward clutch 28a and the reverse brake clutch 28b are released, power transmission via the forward / backward switching device 28 is cut off, and the engine The power transmission between the 10 and the drive pulley 26a of the continuously variable transmission 26 is cut off.

A crank angle sensor 50 is provided at an appropriate position such as near the camshaft (not shown) of the engine 10. The crank angle sensor 50 outputs a signal indicating the engine speed NE for each predetermined crank angle position of the piston. Further, an absolute pressure sensor 52 is provided at an appropriate position downstream of the throttle valve in the intake system. The absolute pressure sensor 52 outputs a signal proportional to the absolute pressure (engine load) PBA in the intake pipe. The actuator of the DBW mechanism 16 is provided with a throttle opening sensor 54.

The output of the crank angle sensor 50 and the like is sent to the engine controller 66. The engine controller 66 includes a microcomputer, determines a target throttle opening degree based on the sensor outputs to control the operation of the DBW mechanism 16, and determines a fuel injection amount to drive the injector 20.

The NT sensor 70 is provided on the main shaft MS. The NT sensor 70 outputs a pulse signal indicating the rotation speed of the turbine runner 24b, specifically, the rotation speed NT of the main shaft MS, and more specifically, the input shaft rotation speed of the forward clutch 28a.

An NDR sensor 72 is provided in the vicinity of the drive pulley 26a of the continuously variable transmission 26. The NDR sensor 72 outputs a pulse signal corresponding to the rotation speed NDR of the drive pulley 26a, in other words, the output shaft rotation speed of the forward clutch 28a.

An NDN sensor 74 is provided in the vicinity of the driven pulley 26b. The NDN sensor 74 outputs a pulse signal indicating the rotation speed NDN of the driven pulley 26b, that is, the rotation speed of the counter shaft CS. A V sensor 76 is provided in the vicinity of the gear 30b of the secondary shaft SS. The V sensor 76 outputs a pulse signal indicating the vehicle speed V through the rotation speed of the secondary shaft SS. In the oil pressure supply mechanism 46, an oil pressure sensor 82 that is arranged in a predetermined oil passage and measures the oil pressure and an oil temperature sensor 84 that measures the oil temperature are arranged.

The outputs of the various sensors described above, including the outputs of other sensors (not shown), are sent to the shift controller 90. The shift controller 90 also includes a microcomputer including a CPU, ROM, RAM, I / O, etc., and is configured to be freely communicative with the engine controller 66.

FIG. 2 is a hydraulic circuit diagram of the hydraulic supply mechanism 46. The hydraulic supply mechanism 46 of the present embodiment includes a main hydraulic circuit 47 and a sub hydraulic circuit 48. The main hydraulic circuit 47 includes a PH control valve 46c described later, and is a hydraulic circuit for controlling each part of the continuously variable transmission 26, the forward / backward switching device 28, and the torque converter 24. The sub-hydraulic circuit 48 is a hydraulic circuit having a lubrication system 46j described later. Since the main hydraulic circuit 47 has many components that require hydraulic control, the oil pressure of the main hydraulic circuit 47 is relatively high. On the other hand, the oil pressure of the sub-hydraulic circuit 48 is relatively low.

Further, in the main hydraulic circuit 47, the hydraulic oil supplied via the PH control valve 46c is supplied first to the first hydraulic oil supply unit and the second hydraulic oil to be supplied after the first hydraulic oil supply unit. It consists of a supply unit. In the present embodiment, the first hydraulic oil supply unit is a hydraulic circuit configured for the continuously variable transmission 26 and the forward / backward switching device 28, and the second hydraulic oil supply unit is for the torque converter 24. It is a configured hydraulic circuit.

The hydraulic supply mechanism 46 is provided with a first oil pump 46a1 and a second oil pump 46a2. The rotor of the first oil pump 46a1 and the rotor of the second oil pump 46a2 are arranged on the same axis as the rotation axis of the engine 10. Therefore, the first oil pump 46a1 and the second oil pump 46a2 are driven by the rotation of the engine 10.

The first oil pump 46a1 pumps the hydraulic oil stored in the reservoir 46b below the CVT case (not shown) and pumps the hydraulic oil into the oil passage 46d connected to the PH control valve 46c. The second oil pump 46a2 pumps hydraulic oil from the reservoir 46b and pumps the hydraulic oil to the oil passage 46e connected to the pump switching valve 46g (flow rate adjusting means).

A PH control valve 46c is connected to the oil passage 46d. The PH control valve 46c adjusts the discharge pressure (primary pressure) of the first oil pump 46a1 and the discharge pressure applied from the second oil pump 46a2 as needed to the PH pressure (line pressure) to adjust the oil passage. Output to 46k.

The pump switching valve 46g has a spring 46g1 which is an urging member at one end of the spool of the pump switching valve 46g. The pump switching valve 46g is urged to the left in the figure by the spring 46g1.

The output of the pump switching valve 46g is connected to the oil passage 46f connected to the oil passage 46d on the one hand and to the oil passage 46h on the other hand, and is connected to the lubrication system 46j from there via the lubrication control valve 46i. Lubrication. The lubrication system 46j is a general term for components or members that require lubrication.

Then, in the oil pressure supply mechanism 46, the mode (state) can be switched to at least two stages of the lubrication mode (first state) and the high pressure mode (second state) by switching the pump switching valve 46g. The lubrication mode is a mode in which the flow rate of the hydraulic oil supplied to the main hydraulic circuit 47 is relatively small, and the high pressure mode is a mode in which the flow rate of the hydraulic oil supplied to the main hydraulic circuit 47 is relatively large.

The switching operation of the pump switching valve 46g at the time of switching between the lubrication mode and the high pressure mode will be described. In the lubrication mode, the pump switching valve 46g supplies the hydraulic oil supplied from the second oil pump 46a2 to the oil passage 46h on the lubrication side. Therefore, the hydraulic oil is supplied to the PH control valve 46c only from the first oil pump 46a1. On the other hand, in the high pressure mode, the pump switching valve 46g supplies the hydraulic oil supplied from the second oil pump 46a2 to the oil passage 46f on the high pressure side. Therefore, hydraulic oil is supplied to the PH control valve 46c from the first oil pump 46a1 and the second oil pump 46a2. In this way, by switching the oil passage in which the pump switching valve 46g supplies the hydraulic oil, the flow rate of the hydraulic oil supplied to the PH control valve 46c is relatively large compared to the first flow rate Q1 which is relatively small. It switches to either the flow rate Q2.

The oil passage 46k is connected to the piston chamber 26a21 of the movable pulley half body 26a2 of the drive pulley 26a via the DR control valve 46m1. Further, the oil passage 46k is connected to the piston chamber 26b21 of the movable pulley half body 26b2 of the driven pulley 26b via the DN control valve 46m2.

In the first linear solenoid valve 46m11 and the second linear solenoid valve 46m21, the pilot pressure adjusted by using the CR pressure sent from the oil passage 46p as the original pressure is applied to one end of the spools of the DR control valve 46m1 and the DN control valve 46m2. Supply to.

The DR control valve 46m1 is regulated using the PH pressure as the original pressure, and is supplied to the piston chamber 26a21 of the movable pulley half body 26a2 of the drive pulley 26a. The DN control valve 46m2 is regulated using the PH pressure as the original pressure, and is supplied to the piston chamber 26b21 of the movable pulley half body 26b2 of the driven pulley 26b. In this way, the drive pulley side pressure and the driven pulley side pressure are generated.

As a result, in the continuously variable transmission 26, a pulley side pressure for moving the movable pulley half body 26a2 and the movable pulley half body 26b2 in the axial direction is generated, and the pulley widths of the drive pulley 26a and the driven pulley 26b change. As a result, the winding radius of the belt 26c changes, and the gear ratio that transmits the output of the engine 10 to the drive wheels 12 can be changed steplessly.

The oil passage 46k, on the other hand, is connected to the CR valve 46o via the oil passage 46n. The CR valve 46o reduces the PH pressure adjusted by the PH control valve 46c to the CR pressure (clutch reducing pressure (control pressure)) and discharges it to the oil passage 46p. The output pressure (CR pressure) of the CR valve 46o discharged to the oil passage 46p is input to the third linear solenoid valve 46q, where the pressure is adjusted to an appropriate level according to the excitation of the solenoid.

The oil pressure adjusted by the third linear solenoid valve 46q is input from the input port 46r1 of the backup valve 46r provided for backup at the time of failure, and is output from the output port 46r2. Then, it is connected to the piston chamber 28a1 of the forward clutch 28a of the forward / backward switching device 28 or the piston chamber 28b1 of the reverse brake clutch 28b via the manual valve 46s.

The manual valve 46s applies the output pressure adjusted by the third linear solenoid valve 46q according to the output signal of the range selector 44 operated by the driver to the piston chamber 28a1 of the forward clutch 28a or the piston chamber of the reverse brake clutch 28b. Connect to 28b1. This allows the vehicle to travel forward or backward.

Further, the discharge pressure of the PH control valve 46c is sent to the TC control valve 46u via the oil passage 46t as the torque converter original pressure. The output pressure of the TC control valve 46u is sent to the piston chamber of the lockup clutch 24c of the torque converter 24, and the discharge pressure is sent to the lubrication system 46j.

The structure of the flood control device will be described with reference to FIG. FIG. 3 is a block diagram of the hydraulic control device. The hydraulic control device of the present embodiment has a configuration in which the shift controller 90 includes at least a control unit 91 (control means), a flow rate estimation means 92, and a slip detecting means 93. The flow rate estimation means 92 is used to correct the estimated supply flow rate Qs and the estimated consumption flow rate Qc by the storage unit 92a and the control unit 91 in which information such as a map for obtaining the estimated supply flow rate Qs and the estimated consumption flow rate Qc, which will be described later, is stored. The correction unit 92b of the above is further provided.

The control unit 91 controls the pump switching valve 46g of the oil supply mechanism 46 based on the results obtained by the flow rate estimating means 92 and the slip detecting means 93. Specifically, the pump switching valve 46g is switched to adjust the flow rate supplied to the PH control valve 46c to the first flow rate Q1 or the second flow rate Q2.

The flow rate estimation means 92 operates based on the rotation speed of the first oil pump 46a1 directly connected to the engine 10, the detected values of the oil pressure sensor 82 and the oil temperature sensor 84, and the map related to the flow rate stored in the storage unit 92a. Estimate the flow rate of oil. The estimated flow rate (estimated value) of the hydraulic oil estimated by the flow rate estimation means 92 includes an estimated supply flow rate Qs supplied from the first oil pump 46a1 and an estimated consumption flow rate Qc consumed by the main hydraulic circuit 47. ..

The estimated supply flow rate Qs in the present embodiment is an estimate of the flow rate of the hydraulic oil supplied from the first oil pump 46a1 of the two oil pumps. However, it is not limited to this. For example, even when the hydraulic supply source is configured by one oil pump, when the hydraulic oil is supplied from the oil pump to the main hydraulic circuit 47, the supply flow rate is relative to the relatively small first flow rate. When it is possible to switch to two stages with the second flow rate, which is often the case, the first flow rate may be estimated.

The map stored in the storage unit 92a is composed of, for example, the oil temperature of the hydraulic oil, the rotation speed of the oil pump, the line pressure, etc., and is a reference map of the supply flow rate, the oil temperature and the line pressure of the hydraulic oil, etc. There is a map that is configured and serves as a reference for the flow rate of hydraulic oil consumed, a map that serves as a reference for the flow rate of hydraulic oil consumed by the continuously variable transmission 26, and the like. The specific structure of the map is not limited to this.

The correction unit 92b corrects the value of the map based on the result of the flow rate balance calculation performed by comparing the estimated supply flow rate Qs and the estimated consumption flow rate Qc performed by the control unit 91. For example, when it is determined that the estimated supply flow rate Qs0 obtained from the reference map Ms0 on the supply side is different from the actual supply flow rate, the value obtained from the map Ms0 is multiplied by a predetermined correction coefficient C1 to obtain C1 × Ms0. The obtained map Ms1 is used as a reference map for obtaining the next estimated supply flow rate Qs. Similarly, when it is determined that the estimated consumption flow rate Qc0 obtained from a certain consumption-side reference map Mc0 is different from the actual consumption flow rate, a predetermined correction coefficient C2 is added to the value obtained from the consumption-side reference map Mc0. , And the map Mc1 obtained from C2 × Mc0 is used as a reference map for obtaining the next estimated consumption flow rate Qc. The correction method does not necessarily have to be performed by multiplying the value obtained from the map by the correction coefficient, and may be performed by adding or subtracting a predetermined correction value to the value obtained from the map. Further, the correction may be performed by multiplying the value obtained from a part of the map by the correction coefficient and further adding or subtracting a predetermined correction value.

The slip detecting means 93 detects the slip ratio ETR (LC engagement rate) of the lockup clutch 24c of the torque converter 24. Specifically, ETR (%) is obtained by measuring the ratio between the detected rotation speed of the crank angle sensor 50 and the detected rotation speed of the NT sensor 70. If the ETR is not 100%, the control unit 91 determines that the slip detection signal SI has been output from the slip detection means 93.

Using FIG. 4, the control unit 91 determines the estimated flow rate (estimated supply flow rate Qs and estimated consumption flow rate Qc), determines the flow rate to be supplied to the PH control valve 46c based on the estimated flow rate, and determines the pump switching valve 46g. The switching timing will be described. FIG. 4 is a flowchart relating to a method for determining the estimated flow rate.

As shown in FIG. 4, first, the control unit 91 determines the estimated supply flow rate Qs and the estimated consumption flow rate Qc based on the values obtained from the above-mentioned various sensors and the map stored in the storage unit 92a of the flow rate estimation means 92. Calculate (step S1).

Then, the control unit 91 compares the estimated supply flow rate Qs with the estimated consumption flow rate Qc (step S2). Here, when the estimated supply flow rate Qs exceeds the estimated consumption flow rate Qc, it is determined that the flow rate supplied to the main hydraulic circuit 47 may be the relatively small first flow rate Q1 (step S3). On the other hand, when the estimated supply flow rate Qs does not exceed (or may be lower than) the estimated consumption flow rate Qc, it is determined that the flow rate supplied to the main hydraulic circuit 47 should supply a relatively large second flow rate Q2. (Step S4).

Next, the control unit 91 determines whether or not the pump switching valve 46g needs to be switched (step S5). Specifically, when the flow rate supplied to the PH control valve 46c changes from the first flow rate Q1 to the second flow rate Q2, or when the flow rate supplied to the PH control valve 46c changes from the second flow rate Q2 to the first flow rate Q1. Corresponds to this.

Here, when the flow rate switching is necessary, the control unit 91 performs control for correcting the estimated flow rate, which will be described later (step S10), and switches the pump switching valve 46g (step S6). On the other hand, when the flow rate switching is not necessary, the control unit 91 does not switch the pump switching valve 46g.

The correction of the estimated flow rate will be described with reference to FIG. FIG. 5 is a flowchart relating to the estimated flow rate correction. Step S10 in FIG. 5 is the same as step S10 shown in FIG.

In the control for correcting the estimated flow rate (step S10), the control unit 91 confirms whether or not the condition for correcting the estimated flow rate is satisfied (step S11). If the condition for correction is not satisfied, the control is terminated without performing anything in particular. On the other hand, when the condition is satisfied, the control unit 91 confirms whether or not the slip detection signal SI is output from the slip detection means 93 (step S12).

Next, when the slip detection signal SI is not output from the slip detection means 93, the control unit 91 proceeds to correct the estimated flow rate (step S13). That is, at the time when the flow rate switching control of the pump switching valve 46g is performed, the magnitude of the estimated supply flow rate Qs and the estimated consumption flow rate Qc is usually reversed. Therefore, the estimated supply flow rate Qs and the estimated consumption flow rate Qc at the time of the flow rate switching control match. At this point, if there is no slippage of the lockup clutch 24c, it can be determined that there was actually a sufficient supply flow rate for engaging the lockup clutch 24c, and the estimated supply flow rate Qs is the estimated consumption flow rate. It can be estimated that it was larger than Qc.

Therefore, in step S13, the control unit 91 increases the estimated supply flow rate Qs used at the time of switching the next pump switching valve 46g. Alternatively, it is determined that the actual consumption flow rate is less than the estimated consumption flow rate Qc used at the time of switching the pump switching valve 46g, and the estimated consumption flow rate Qc used at the time of switching the next pump switching valve 46g is reduced. It should be noted that these controls are not necessarily limited to performing only one of them, and may perform both a correction for increasing the estimated supply flow rate Qs and a correction for decreasing the estimated consumption flow rate Qc. The correction for increasing the time of the lubrication mode (first state) by performing at least one of the correction for increasing the estimated supply flow rate Qs and the correction for decreasing the estimated consumption flow rate Qc is called the first state increase correction.

On the other hand, when the slip detection signal SI is output from the slip detection means 93, the control unit 91 returns the correction of the estimated flow rate to the state at least one step before (step S14). That is, when the lockup clutch 24c slips, it can be determined that the supply flow rate is not sufficient to engage the lockup clutch 24c, and the estimated supply flow rate Qs is the estimated consumption flow rate Qc. It can be judged that it was smaller than.

Therefore, in step S14, the control unit 91 makes a correction for reducing the estimated supply flow rate Qs or a correction for increasing the estimated consumption flow rate Qc. It should be noted that these controls are not necessarily limited to performing only one of them, and may perform both a correction for decreasing the estimated supply flow rate Qs and a correction for increasing the estimated consumption flow rate Qc. The correction for reducing the time of the lubrication mode (first state) by performing at least one of the correction for reducing the estimated supply flow rate Qs and the correction for increasing the estimated consumption flow rate Qc is called the first state reduction correction.

The slip detection signal SI from the slip detection means 93 is always transmitted to the control unit 91. Therefore, the time point when the slip detection signal SI of the lockup clutch 24c is output is not necessarily limited to the time point of the flow rate switching control.

The control unit 91 interrupts the estimated flow rate correction after step S14 (step S15). Therefore, the estimated supply flow rate Qs and the estimated consumption flow rate Qc are corrected values at least one step before the time when the slip detection signal SI is obtained from the slip detection means 93.

Next, the effect of the above estimated flow rate correction will be described by exemplifying a specific case. FIG. 6 is a diagram showing an example in which the increase correction of the estimated supply flow rate Qs is performed when switching from the lubrication mode to the high pressure mode. In the example of FIG. 6, the time point at which the estimated flow rate correction is performed is the time point of the flow rate switching control for switching from the lubrication mode to the high pressure mode, only the estimated supply flow rate Qs is corrected, and the estimated supply flow rate Qs is fixed.

First, before the time point Ta1 of the first flow rate switching control, the estimated supply flow rate Qs is calculated using the reference map Ms0. Since the engagement rate (LC engagement rate) of the lockup clutch 24c is 100% at the time Ta1 of the first flow rate switching control, the slip detection signal SI is not output. In this case, a correction is made to increase the estimated supply flow rate Qs. Specifically, the estimated supply flow rate Qs after the time Ta1 of the first flow rate switching control is calculated based on the map Ms1 in which the estimated supply flow rate Qs is larger than the map Ms0. Further, even at the time point Ta2 of the second flow rate switching control, since the LC fastening rate is 100%, the correction for increasing the estimated supply flow rate Qs is performed. Specifically, correction is performed based on the map Ms2 so that the estimated supply flow rate Qs is further increased.

As a result, after the flow rate switching control, the transition time from the lubrication mode to the high pressure mode is delayed, and the transition time from the high pressure mode to the lubrication mode is delayed as compared with the case where the estimated supply flow rate is continuously calculated based on the map Ms0. It's getting faster. Therefore, the lubrication mode time becomes longer, and the flow rate for supplying hydraulic oil to the high-pressure main hydraulic circuit 47 decreases, so that the load on the engine 10 rotating coaxially with the second oil pump 46a2 decreases, and fuel efficiency improves.

FIG. 7 is a diagram showing an example in which reduction correction of the estimated supply flow rate Qs is performed when the slip detection signal SI is detected. In the example of FIG. 7, the time point at which the estimated flow rate correction is performed is the time point of the flow rate switching control for switching from the lubrication mode to the high pressure mode, only the estimated supply flow rate Qs is corrected, and the estimated consumption flow rate Qc is not corrected.

First, at the time point Tb1 of the flow rate switching control, a correction for increasing the estimated supply flow rate Qs is performed, and the estimated supply flow rate Qs is calculated based on the map Ms2. Here, when the slip detection signal SI is output from the slip detection means 93, the correction of the estimated flow rate is returned to the state at least one step before. That is, after the time point Tb2 when the slip detection signal SI is output, the estimated supply flow rate Qs is calculated based on the map Ms1 such that the estimated supply flow rate Qs is smaller than the map Ms2. That is, the estimated supply flow rate Qs is calculated based on the map Ms1 used in the correction one step before the map Ms2.

As a result, when the estimated supply flow rate Qs is calculated based on the map Ms1, it can be grasped that the calculation of the estimated supply flow rate Qs based on the map Ms1 in which the slip detection signal SI is not output is optimal.

FIG. 8 is a diagram showing an example in which the increase correction of the estimated supply flow rate Qs is performed when switching from the high pressure mode to the lubrication mode. In the example of FIG. 8, the time point at which the estimated flow rate correction is performed is the time point of the flow rate switching control for switching from the high pressure mode to the lubrication mode, only the estimated supply flow rate Qs is corrected, and the estimated consumption flow rate Qc is not corrected.

First, before the time point Tc1 of the first flow rate switching control, the estimated supply flow rate Qs is calculated using the reference map Ms0. At the time point Tc1 of the first flow rate switching control, since the LC fastening rate is 100%, the slip detection signal SI is not output. In this case, a correction is made to increase the estimated supply flow rate Qs. Specifically, the estimated supply flow rate Qs after the time point Tc1 of the first flow rate switching control is calculated based on the map Ms1 in which the estimated supply flow rate Qs is larger than the map Ms0. Further, even at the time point Tc2 of the second flow rate switching control, since the LC fastening rate is 100%, the correction for increasing the estimated supply flow rate Qs is performed. Specifically, correction is performed based on the map Ms2 so that the estimated supply flow rate Qs is further increased.

As a result, after the flow rate switching control, the transition time from the lubrication mode to the high pressure mode is delayed, and the transition time from the high pressure mode to the lubrication mode is delayed as compared with the case where the estimated supply flow rate is continuously calculated based on the map Ms0. It's getting faster. Therefore, the time of the lubrication mode becomes long, and the load of the engine 10 is reduced as described above, so that the fuel consumption is improved.

FIG. 9 is a diagram showing an example in which the decrease correction of the estimated consumption flow rate Qc is performed when switching from the lubrication mode to the high pressure mode. In the example of FIG. 9, the time point at which the estimated flow rate correction is performed is the time point of the flow rate switching control for switching from the lubrication mode to the high pressure mode, and the estimated supply flow rate Qs is not corrected and only the estimated consumption flow rate Qc is corrected.

First, before the time point Td1 of the first flow rate switching control, the estimated consumption flow rate Qc is calculated using the reference map Mc0. At the time point Td1 of the first flow rate switching control, since the LC fastening rate is 100%, the slip detection signal SI is not output. In this case, correction is performed to reduce the estimated consumption flow rate Qc. Specifically, the estimated consumption flow rate Qc after the time point Td1 of the first flow rate switching control is calculated based on the map Mc1 such that the estimated consumption flow rate Qc is smaller than the map Mc0. Further, even at the time point Td2 of the second flow rate switching control, since the LC fastening rate is 100%, the correction for reducing the estimated consumption flow rate Qc is performed. Specifically, the correction is performed based on the map Mc2 so that the estimated consumption flow rate Qc is further reduced.

As a result, after the flow rate switching control, the transition time from the lubrication mode to the high pressure mode is delayed, and the transition time from the high pressure mode to the lubrication mode is delayed as compared with the case where the estimated consumption flow rate is continuously calculated based on the map Mc0. It's getting faster. Therefore, the time of the lubrication mode becomes long, and the load of the engine 10 is reduced as described above, so that the fuel consumption is improved.

As described above, according to the hydraulic control device of the automatic transmission of the present embodiment, the slip of the lockup clutch 24c is detected at the time of the flow rate switching control, and when the slip detection signal SI is not output, the flow rate switching control is performed. At this time, the control unit 91 determines that a sufficient flow rate of hydraulic oil has been supplied for the engagement of the lockup clutch 24c. Therefore, the control unit 91 performs the first state increase correction. Then, the time for the control unit 91 to select the lubrication mode can be increased, and the flow rate of supplying hydraulic oil from the second oil pump 46a2 to the high-pressure main hydraulic circuit 47 is reduced, so that the second oil pump 46a2 and The load applied to the engine 10 that drives this can be reduced. As a result, fuel efficiency can be improved.

Further, according to the present embodiment, when the slip detection signal SI is output, the control unit 91 determines that a sufficient flow rate of hydraulic oil cannot be supplied for engaging the lockup clutch 24c. Therefore, the control unit 91 performs the first state reduction correction. Then, the time for the control unit 91 to select the lubrication mode is reduced and the time for selecting the high pressure mode is increased, so that a sufficient flow rate of the hydraulic oil is supplied to the main hydraulic circuit 47.

Further, according to the present embodiment, before the slip detection signal SI is output, the lubrication mode increases each time the first state increase correction is performed, so that the flow rate supplied to the main hydraulic circuit 47 can be reduced. .. Further, when the slip detection signal SI is output, the first state reduction correction is performed to return to the estimated flow rate immediately before the slip detection signal SI is output, that is, the estimated supply flow rate Qs or the estimated consumption flow rate Qc. Can be done. The estimated flow rate can be considered as a value that does not detect slippage and is most in line with the actual flow rate. Here, by canceling the subsequent first state increase correction, the estimated flow rate is not changed, and by performing the flow rate switching control based on the estimated flow rate, the flow rate switching close to the actual flow rate with respect to the main hydraulic circuit 47 is performed. Control is possible. Since the first state reduction correction is not stopped even after the slip detection signal SI is output once, the characteristics of the hydraulic supply source change due to the continuous use of the hydraulic supply source, and the slip detection signal SI is output again. In that case, the flow rate switching control close to the actual flow rate can be performed by performing the first state reduction correction again.

Further, according to the present embodiment, the main hydraulic circuit 47 has a first hydraulic oil supply unit that supplies hydraulic oil to the continuously variable transmission 26 and the forward / backward switching device 28, and a surplus portion of the first hydraulic oil supply unit. It is composed of a second hydraulic oil supply unit to which hydraulic oil is supplied. The second hydraulic oil supply unit has a torque converter 24. In this case, the torque converter 24 in the second hydraulic oil supply unit to which the hydraulic oil is supplied later is more likely to have a shortage of the hydraulic oil flow rate than the first hydraulic oil supply unit. Here, by detecting the slippage of the lockup clutch 24c of the torque converter 24 and determining whether or not the flow rate of the hydraulic oil is insufficient, it is possible to prevent the flow rate of the hydraulic oil from being insufficient in the first hydraulic oil supply unit. When the first hydraulic oil supply unit is set to the pulley or the like which is the main transmission mechanism of the continuously variable transmission 26, it is possible to prevent the flow rate of the hydraulic oil of the main transmission mechanism from being insufficient. Further, by detecting the slip of the lockup clutch 24c in the second hydraulic oil supply unit where the flow rate shortage is likely to occur, the flow rate shortage can be detected quickly, so that the accuracy of the hydraulic oil flow rate balance calculation can be improved. ..

Although the embodiments of the present invention have been described above, the present invention is not limited to the above-described embodiments, and the present invention is not limited to the above-described embodiments. It can be transformed.

10 ... Engine 24 ... Torque converter 24c ... Lockup clutch (friction engagement element)
26 ... Continuously variable transmission 26a ... Drive pulley 26b ... Driven pulley 28 ... Forward / backward switching device 28a ... Forward clutch 28b ... Reverse brake clutch 28c ... Planetary gear mechanism 46 ... Hydraulic supply mechanism 46a1 ... First oil pump (hydraulic supply source)
46a2 ... Second oil pump (hydraulic supply source)
46c ... PH control valve 46g ... Pump switching valve (flow rate adjusting means)
46j ... Lubrication system 46u ... TC control valve 47 ... Main hydraulic circuit 48 ... Sub hydraulic circuit 50 ... Crank angle sensor 70 ... NT sensor 82 ... Hydraulic sensor 84 ... Oil temperature sensor 90 ... Shift controller 91 ... Control unit (control means)
92 ... Flow rate estimation means 92a ... Storage unit 92b ... Correction unit 93 ... Slip detection means Q1 ... First flow rate Q2 ... Second flow rate Qc ... Estimated consumption flow rate Qs ... Estimated supply flow rate SI ... Slip detection signal

Claims (5)

An automatic transmission having a frictional engagement element that transmits power by engagement and transmitting power at a predetermined gear ratio, a hydraulic supply mechanism that supplies hydraulic pressure for engaging the frictional engagement element, and the above. An automatic transmission hydraulic control device including a control means for controlling a hydraulic supply mechanism.
The hydraulic supply mechanism is
The main hydraulic circuit having the friction engaging element and requiring a relatively high hydraulic pressure,
A sub-hydraulic circuit that requires relatively low flood control,
A hydraulic supply source driven by an internal combustion engine to supply hydraulic oil to the main hydraulic circuit and the sub-hydraulic circuit,
At least a first state in which the flow rate supplied from the hydraulic supply source to the main hydraulic circuit is relatively small and a second state in which the flow rate supplied from the hydraulic supply source to the main hydraulic circuit is relatively large. Flow rate adjustment means that can be switched in two stages,
Estimated supply flow rate, which is an estimated value of the flow rate of hydraulic oil to the main hydraulic circuit in the first state, and estimated consumption, which is an estimated value of the flow rate of hydraulic oil consumed by the main hydraulic circuit in the first state. A flow rate estimation means for estimating the flow rate and
A slip detecting means for outputting a slip detection signal when the slip of the friction engaging element is detected is provided.
The control means
When the estimated supply flow rate is lower than the estimated consumption flow rate, the first state is switched to the second state, and when the estimated supply flow rate exceeds the estimated consumption flow rate, the second state is changed to the first state. Performs flow rate switching control to switch to the state,
When the slip detection signal is not output from the slip detection means during the flow rate switching control, the first state increase correction is at least one of the correction for increasing the estimated supply flow rate and the correction for decreasing the estimated consumption flow rate. A hydraulic control device for an automatic transmission, which is characterized by performing. The control means
When the slip detection signal is output from the slip detection means, the first state reduction correction, which is at least one of the correction for reducing the estimated supply flow rate and the correction for increasing the estimated consumption flow rate, is performed. The hydraulic control device for an automatic transmission according to claim 1. The control means
When the slip detection signal is not output from the slip detection means, the first state increase correction is performed at the time of the flow rate switching control.
The automatic transmission according to claim 2, wherein when the slip detection signal is output from the slip detecting means, the first state decrease correction is performed and the subsequent first state increase correction is stopped. Hydraulic control device. The main hydraulic circuit
The first hydraulic oil supply unit to which the hydraulic oil from the hydraulic supply source is supplied, and
A second hydraulic oil supply unit to which a surplus hydraulic oil of the first hydraulic oil supply unit is supplied is provided.
The hydraulic control device for an automatic transmission according to any one of claims 1 to 3, wherein the second hydraulic oil supply unit supplies hydraulic pressure to the friction engaging element. The hydraulic control device for an automatic transmission according to claim 4, wherein the friction engaging element is a lockup clutch of a torque converter.