JP6190858B2 - Hydraulic circuit control device - Google Patents

Description

  The present invention relates to a control device for a hydraulic circuit that controls the hydraulic pressure supplied to a continuously variable transmission.

  2. Description of the Related Art Conventionally, a hydraulic control device used in a transmission, which operates a first oil pump that supplies low hydraulic pressure such as cooling and lubrication of transmission components, and a hydraulic operation section that operates according to high hydraulic pressure In order to achieve this, a hydraulic control device including a second oil pump that supplies high hydraulic pressure is known (see, for example, Patent Document 1). In the hydraulic control device of Patent Document 1, the first oil pump is driven by an internal combustion engine, and the second oil pump is driven by an electric motor.

  In this hydraulic control device, since the two oil pumps are selectively used depending on the application (ie, cooling and lubrication, or operating at a high pressure), the driving force of the oil pump of the internal combustion engine is reduced, and the internal combustion engine is correspondingly reduced. The driving force output from the engine is reduced.

JP 2001-74130 A

  However, in the hydraulic control apparatus as described in Patent Document 1, the second oil pump is configured to be able to output the maximum possible power among the powers necessary for operating the hydraulic operation unit. It is necessary to use a relatively large electric motor as the electric motor for driving the second oil pump. For this reason, even when the load acting on the electric motor is small, the electric motor becomes large and the electric resistance increases, resulting in poor energy efficiency. In addition, not only when using an electric motor as a drive source for driving the oil pump, but also when using another drive source, it is the same that the energy efficiency deteriorates as the second oil pump becomes larger. .

  The objective of this invention is providing the control apparatus of the hydraulic circuit which can drive an oil pump efficiently.

In order to achieve the above object, the invention described in claim 1
A supplied part (for example, supplied part 3 in an embodiment described later) to which a low hydraulic pressure is supplied, and a hydraulic operating part (for example, in an embodiment described later) to which a high hydraulic pressure higher than the low hydraulic pressure is supplied. Control device (for example, management ECU 125 in the embodiment described later) of a hydraulic circuit (for example, hydraulic circuit 1 in the embodiment described later) for supplying hydraulic pressure to
The hydraulic operating part is
An input-side pulley (for example, an input-side pulley Dr in an embodiment described later) and an output-side pulley (for example, an output-side pulley Dn in an embodiment described later) whose width can be changed by supplying the high hydraulic pressure. A continuously variable transmission (for example, a continuously variable transmission T in an embodiment described later) having a continuously variable gear ratio;
The hydraulic circuit is
A mechanical oil pump (for example, a large-capacity oil pump Pb in an embodiment described later) driven by a drive source (for example, an internal combustion engine ENG in an embodiment described later) that outputs a driving force for traveling of the vehicle; ,
An oil pump that is driven by an electric motor (for example, an electric motor MOT in an embodiment described later) and has a smaller capacity than the mechanical oil pump, and further pressurizes the hydraulic pressure supplied from the mechanical oil pump to operate the hydraulic pressure An electric oil pump (for example, a small-capacity oil pump Ps in an embodiment described later) to be supplied to the unit,
A first flow path for supplying hydraulic pressure supplied from the mechanical oil pump to the electric oil pump (for example, a first flow path L1 in an embodiment described later);
A second flow path for supplying the hydraulic pressure supplied from the electric oil pump to the hydraulic operating section (for example, a second flow path L2 in an embodiment described later);
A third flow path (for example, a third flow path L3 in an embodiment described later) for supplying the hydraulic pressure supplied from the mechanical oil pump to the hydraulic operation unit without passing through the electric oil pump;
A line pressure adjusting valve (for example, a ninth pressure control valve 30 in an embodiment described later) provided between the mechanical oil pump and the first flow path;
A first shift control valve that receives the line pressure of the second flow path or the third flow path and performs control to supply / discharge hydraulic oil to / from the input pulley (for example, a sixth pressure control in an embodiment described later) Valve 16),
A second shift control valve that receives the line pressure in the second flow path or the third flow path and performs control to supply and discharge hydraulic oil to and from the output pulley (for example, a seventh pressure in an embodiment described later) Control valve 17);
A first solenoid valve that creates a first pilot pressure and controls the operation of the first shift control valve by acting on the first shift control valve (for example, a fourth pressure control valve 14 in an embodiment described later);
A second solenoid valve that creates a second pilot pressure and controls the operation of the second shift control valve by acting on the second shift control valve (for example, a fifth pressure control valve 15 in an embodiment described later); With
The line pressure adjusting valve includes a first spool (for example, a first spool 31 in an embodiment described later) and a second spool (for example, a second spool 32 in an embodiment described later) inside,
The second spool is urged toward the first spool by a first elastic member (for example, a first elastic member 33 in an embodiment described later),
The first spool is separated from the second spool by a second elastic member (for example, a second elastic member 34 in an embodiment described later) disposed between the first spool and the second spool. Energized,
The line pressure adjusting valve includes a first port (for example, a first port 30a in an embodiment described later) to which hydraulic pressure from the mechanical oil pump is supplied,
A second port that is always in communication with the first port and connected to the first flow path (for example, a second port 30b in an embodiment described later);
A third port that is provided on the side farther from the second spool than the second port, and is connected to the supplied part via a lubricating oil passage (for example, a twelfth oil passage R12 in an embodiment described later). (For example, the third port 30c in the embodiment described later),
A fourth port (for example, a fourth port 30d in an embodiment described later) provided on the side farther from the second spool than the third port and supplied with the line pressure;
A fifth port to which the first pilot pressure is supplied (for example, a fifth port 30e in an embodiment described later);
A sixth port to which the second pilot pressure is supplied (for example, a sixth port 30f in an embodiment described later),
The first spool is biased toward the second spool by the line pressure supplied from the fourth port, so that the first port, the second port, and the third port are Configured to communicate,
The first spool is urged toward the side away from the second spool by a relief pressure which is the higher of the first pilot pressure and the second pilot pressure. And the communication between the second port and the third port can be cut off,
The control device controls the relief pressure so that the first port, the second port, and the third port do not communicate with each other when idling is stopped when the mechanical oil pump is stopped. It is.

The invention according to claim 2 is the invention according to claim 1,
When the control device controls the relief pressure so that the first port, the second port, and the third port do not communicate with each other, one of the first pilot pressure and the second pilot pressure is selected. The first solenoid valve or the second solenoid valve is controlled so that the higher pilot pressure is further increased.

In the invention according to claim 3, in the invention according to claim 1 or 2,
The first solenoid valve and the second solenoid valve are normally open solenoid valves,
The control device performs control so that the energization amount to the solenoid valve that outputs the higher pilot pressure of the first pilot pressure and the second pilot pressure becomes zero.

The invention according to claim 4 is the invention according to any one of claims 1 to 3,
The control device prevents the first port, the second port, and the third port from communicating after the drive of the drive source is stopped and the rotation of the mechanical oil pump is completely stopped. The relief pressure is controlled.

The invention according to claim 5 is the invention according to any one of claims 1 to 4,
The control device calculates the torque of the electric oil pump based on the hydraulic pressure applied to the output pulley when the operation of the drive source is stopped and the vehicle is stopped.

The invention according to claim 6 is the invention according to claim 5,
The control device includes a current supplied to the electric motor calculated based on a rotational speed of the electric oil pump determined according to a target line pressure and an oil temperature of the hydraulic oil, and a rotational speed of the electric oil pump. And a fluid friction torque coefficient for calculating the torque of the electric oil pump based on the oil temperature of the hydraulic oil and the oil pressure to the output pulley.

  According to the first aspect of the present invention, the hydraulic pressure output from the large-capacity mechanical oil pump is supplied to the small-capacity electric oil pump. For this reason, the electric oil pump only needs to increase the pressure by a deficiency relative to the hydraulic pressure output from the mechanical oil pump, and the pressure that the electric oil pump should apply to the oil is reduced compared to the conventional one. For this reason, the energy consumption in an electric oil pump can be reduced.

  Also, when a large amount of power is required to drive the oil pump, such as when supplying a large amount of oil to the hydraulic actuator, a high hydraulic pressure is not transferred from the mechanical oil pump to the hydraulic actuator without using an electric oil pump. In some cases, the total amount of power for driving each oil pump may be less when the oil is directly supplied than when the electric oil pump is used.

  In such a case, the operation of the electric oil pump is stopped, and the hydraulic pressure output from the mechanical oil pump is supplied to the hydraulic operating portion via the third flow path, so that the electric oil pump is required. The maximum power that can be output can be reduced. For this reason, a relatively small device can be used as the electric oil pump, and as a result, energy efficiency when driving the electric oil pump can be improved.

  The control device also controls the relief pressure so that the first port, the second port, and the third port do not communicate with each other when idling is stopped when the mechanical oil pump is stopped. As a result, the path from the supplied portion to the electric oil pump is interrupted, so that it is possible to prevent air from flowing backward through the path and mixing into the electric oil pump when idling is stopped. Further, since a check valve for preventing air from entering the electric oil pump is not required, the hydraulic circuit can be reduced in size and cost.

  According to the invention of claim 2, by controlling the higher pilot pressure of the first pilot pressure and the second pilot pressure to be higher, while preventing air from being mixed into the electric oil pump, It is possible to prevent the gear ratio of the continuously variable transmission from changing. That is, since the mechanical oil pump is stopped when idling is stopped, the line pressure is supplied by the electric oil pump. At this time, even if the higher one of the first pilot pressure and the second pilot pressure is further increased, the line pressure does not increase any further, and the input pulley and the output pulley of the continuously variable transmission are not increased. Since each hydraulic pressure of the hydraulic oil does not change, the gear ratio of the continuously variable transmission does not change. For this reason, the merchantability of a vehicle equipped with a continuously variable transmission is not affected.

  According to the invention of claim 3, since the first solenoid valve or the second solenoid valve is a normally open type, the solenoid valve that outputs the higher pilot pressure of the first pilot pressure and the second pilot pressure is provided. The power consumption can be reduced while increasing the pilot pressure.

  According to the fourth aspect of the present invention, after the drive of the drive source is stopped and the rotation of the mechanical oil pump is completely stopped, the line pressure adjusting valve is connected between the first port, the second port, and the third port. By completely closing the path, the path from the supplied part to the electric oil pump is interrupted. For this reason, it is possible to prevent the line pressure from unintentionally rising due to the passage being completely closed while the mechanical oil pump is rotating and the oil being not supplied to the supplied portion.

  According to the fifth and sixth aspects of the present invention, the driving torque of the electric oil pump is calculated from an equation including a fluid friction torque coefficient as a parameter, and therefore the driving torque can be calculated with high accuracy. In order to derive the fluid friction torque coefficient, for example, a pulley pressure sensor provided for controlling the pulley pressure with high accuracy can be used. Therefore, the driving torque of the electric oil pump can be increased without providing a dedicated sensor. It can be controlled accurately.

1 is a block diagram illustrating an internal configuration of a vehicle equipped with a hydraulic circuit control device according to an embodiment of the present invention. BRIEF DESCRIPTION OF THE DRAWINGS It is a figure which shows the outline | summary of the hydraulic circuit of one Embodiment of this invention, (a) is a figure which shows the case where both a large capacity oil pump and a small capacity oil pump drive, (b) is a large capacity oil pump It is a figure which shows the case where only this is driven. It is a figure explaining the flow volume and power of a hydraulic circuit. It is a figure which shows the detailed structure of a hydraulic circuit. It is a figure which shows the flow of the oil of the 1st spool in the 9th pressure control valve of the hydraulic circuit at the time of the action | operation of a large capacity | capacitance oil pump and a small capacity | capacitance oil pump. FIG. 12 is a diagram showing the movement of the first spool and the flow of oil when the control shown in FIG. 11 is not performed in the ninth pressure control valve of the hydraulic circuit when the large-capacity oil pump is stopped and the small-capacity oil pump is activated. . It is a figure which shows the action | operation of the hydraulic circuit at the time of normal driving | running | working. It is a figure which shows the action | operation of the hydraulic circuit at the time of sudden shifting. It is a figure which shows the action | operation of the hydraulic circuit at the time of coasting down. It is a figure which shows the action | operation of the hydraulic circuit at the time of a stop. It is a graph which shows an example of each time-dependent change of the rotation speed, pilot pressure, and line pressure of an internal combustion engine when the action | operation of an internal combustion engine stops. FIG. 13 is a diagram showing the movement of the first spool and the flow of oil when the control shown in FIG. 11 is performed in the ninth pressure control valve of the hydraulic circuit when the large-capacity oil pump is stopped and when the small-capacity oil pump is activated. is there. It is a flowchart at the time of setting a fluid friction torque coefficient. It is a figure which shows the timing of the test cycle of a small capacity | capacitance oil pump when a vehicle stops from normal driving | running | working. FIG. 5 is a block diagram showing a function of calculating a discharge pressure of a small capacity oil pump based on a pulley pressure request value of a continuously variable transmission and calculating a drive torque of the small capacity oil pump based on the discharge pressure and the like. It is a flowchart which shows the subroutine of the process of step S117 shown in FIG.

  Hereinafter, embodiments of the present invention will be described with reference to the drawings.

  FIG. 1 is a block diagram showing an internal configuration of a vehicle equipped with a hydraulic circuit control device according to an embodiment of the present invention. The vehicle shown in FIG. 1 includes an internal combustion engine ENG, a clutch C, a continuously variable transmission T, a hydraulic circuit 1, a vehicle speed sensor 121, a rotation speed sensor 123, and a management ECU (MG ECU) 125. In FIG. 1, dotted arrows indicate value data, alternate long and short dashed arrows indicate control signals including instruction contents, and solid arrows indicate hydraulic pressure.

  The internal combustion engine ENG outputs a driving force for the vehicle to travel. The output of the internal combustion engine ENG is transmitted to the drive wheels 129 via the clutch C, the continuously variable transmission T, and the drive shaft 127. The output of the internal combustion engine ENG is also used to drive a large-capacity oil pump Pb, which will be described later, included in the hydraulic circuit 1.

  The clutch C connects and disconnects the transmission path of the driving force from the internal combustion engine ENG to the driving wheel 129 according to the hydraulic pressure of the hydraulic oil supplied from the hydraulic circuit 1. The continuously variable transmission T is a continuously variable transmission (CVT) capable of continuously adjusting the gear ratio in accordance with the hydraulic pressure of the hydraulic oil supplied from the hydraulic circuit 1. The hydraulic circuit 1 supplies a predetermined hydraulic pressure to the clutch C and the continuously variable transmission T via hydraulic oil in accordance with control from the management ECU 125. Details of the hydraulic circuit 1 will be described later.

  The vehicle speed sensor 121 detects the traveling speed (vehicle speed VP) of the vehicle. A signal indicating the vehicle speed VP detected by the vehicle speed sensor 121 is sent to the management ECU 125. The rotational speed sensor 123 detects the rotational speed Ne of the internal combustion engine ENG. A signal indicating the rotational speed Ne detected by the rotational speed sensor 123 is sent to the management ECU 125.

  The management ECU 125 controls the internal combustion engine ENG and the hydraulic circuit 1 based on the vehicle speed VP, the rotational speed Ne of the internal combustion engine ENG, the accelerator pedal opening (AP opening), the brake pedal pressing force (BRK pressing force), and the like. Do. Details of the management ECU 125 will be described later.

  Hereinafter, the hydraulic circuit 1 will be described in detail.

(1. Outline of hydraulic circuit)
The outline of the hydraulic circuit 1 will be described with reference to FIG.

  The hydraulic circuit 1 is supplied with hydraulic pressure to a hydraulic operating portion 2 (high pressure system, for example, the clutch C and the continuously variable transmission T shown in FIG. 1) to which a relatively high hydraulic pressure (high hydraulic pressure) should be supplied. If the hydraulic pressure is sufficiently low (low hydraulic pressure), the hydraulic pressure is supplied to a sufficiently supplied portion 3 (low pressure system, for example, an operation member that requires lubrication or cooling with oil, or a lock-up clutch of a torque converter that operates at low pressure). A supply circuit is configured. The hydraulic circuit 1 includes a large capacity oil pump Pb driven by the internal combustion engine ENG and a small capacity oil pump Ps driven by the electric motor MOT.

  The large-capacity oil pump Pb pumps oil from an oil tank (not shown) and applies pressure to output low oil pressure to the low-pressure supply part 3 and oil pressure to the hydraulic operation part 2 as well. It is a pump. The small capacity oil pump Ps is an oil pump having a capacity smaller than that of the large capacity oil pump Pb. The small-capacity oil pump Ps outputs a hydraulic pressure that operates the hydraulic operation unit 2. In addition, the small-capacity oil pump Ps further pressurizes the supplied hydraulic pressure and supplies it to the hydraulic operation unit 2.

  The hydraulic circuit 1 includes a first flow path L1, a second flow path L2, and a third flow path L3 as main flow paths among the flow paths through which oil flows. The first flow path L1 connects the large capacity oil pump Pb and the small capacity oil pump Ps. The second flow path L2 connects the small capacity oil pump Ps and the hydraulic operation unit 2. The third flow path L3 is a flow path that connects the large-capacity oil pump Pb and the hydraulic operation unit 2 without using the small-capacity oil pump Ps. The supplied part 3 has a flow path connected to the large-capacity oil pump Pb.

  With the configuration described above, the hydraulic pressure output from the large-capacity oil pump Pb is supplied to the small-capacity oil pump Ps. For this reason, the pressure that the small-capacity oil pump Ps should apply to the oil decreases. For this reason, energy consumption when driving the small capacity oil pump Ps can be reduced.

Specifically, the torque τ (Nm) required to drive the small capacity oil pump Ps is
τ = ΔP · V / 2π (1)
Given in.

  Here, ΔP (MPa) is the pressure that the small-capacity oil pump Ps pressurizes, and V (cc / rev) is the theoretical displacement volume (discharge amount per pump rotation) of the small-capacity oil pump Ps. Further, π is the circumference ratio.

  Here, when the hydraulic pressure pressurized by the large-capacity oil pump Pb is represented by P_pb and the hydraulic pressure to be supplied to the hydraulic operation unit 2 is represented by P_line, the torque τ required to drive the small-capacity oil pump Ps is: “(P_line−P_pb) · V / 2π”. That is, the torque τ for driving the small-capacity oil pump Ps can be reduced by “P_pb · V / 2π” as compared with the case where the small-capacity oil pump Ps directly pressurizes the oil pumped up from the oil tank to a high hydraulic pressure. Therefore, energy consumption when driving the small capacity oil pump Ps can be reduced.

  Here, for example, in a case where only one oil pump is used and appropriate hydraulic pressure is supplied to both the high-pressure system hydraulic operation unit and the low-pressure system supply unit, the oil pressure is supplied to the hydraulic operation unit. It is necessary to configure the oil pump and its drive source so as to be able to supply the maximum value of the hydraulic pressure and the maximum amount of the flow rate of the oil to be supplied to the hydraulic operation part and the supplied part.

  However, in general, a high-pressure hydraulic operation unit requires high oil pressure for its operation, but the flow rate of supplied oil may be small in many cases. On the other hand, there are cases where the low-pressure supply part needs to supply oil at a large flow rate for lubrication or cooling. At this time, if there is only one oil pump, a large flow rate is supplied to the low-pressure supply part at a high hydraulic pressure, so excessive work is performed and the energy consumption of the oil pump increases.

  On the other hand, by using two oil pumps Pb and Ps having a large capacity and a small capacity as in the present embodiment, when the hydraulic pressure is supplied to the low-pressure supply part 3, the small capacity oil pump Ps is not used. In addition, if the large-capacity oil pump Pb pressurizes the oil pumped up from the oil tank so as to have a low hydraulic pressure, a large amount of oil can be supplied at a low hydraulic pressure. At this time, the large-capacity oil pump Pb does not need to make the oil have a high hydraulic pressure, so that the energy consumption can be reduced.

  Also, when supplying hydraulic pressure to the high-pressure hydraulic operating section 2, the large-capacity oil pump Pb pressurizes the oil pumped up from the oil tank to a low hydraulic pressure, and then further increases with the small-capacity oil pump Ps. High hydraulic pressure can be supplied by pressing. At this time, the electric motor MOT that drives the small-capacity oil pump Ps for further pressurization can be configured to be relatively small, so that the energy efficiency of the electric motor MOT can be improved.

(1-1. Relationship between flow rate and work rate)
With reference to FIG. 3, the flow rate of oil supplied to the hydraulic operation unit 2 (hereinafter referred to as “high pressure flow rate”; horizontal axis) L and the power Pw (vertical axis) of the hydraulic circuit 1 will be described. FIG. 3 shows the change in the work rate (hereinafter referred to as “1 pump work rate”) Pwb of the hydraulic circuit 1 with respect to the change in the supplied high pressure flow rate L when only the large capacity oil pump Pb is driven, and the large capacity oil. It shows the change in the work rate (hereinafter referred to as “2 pump work rate”) Pws of the hydraulic circuit 1 with respect to the change in the supplied high pressure flow rate when both the pump Pb and the small capacity oil pump Ps are driven. .

  When driving the large-capacity oil pump Pb and the small-capacity oil pump Ps, the large-capacity oil pump Pb supplies a low hydraulic pressure to the supplied part 3 and the small-capacity oil pump Ps, and the small-capacity oil pump Ps High oil pressure is supplied to the hydraulic operation unit 2 by further pressurizing the oil supplied from the pump Pb.

  When the high pressure flow rate L is the predetermined flow rate α, the 1 pump work rate Pwb and the 2 pump work rate Pws are equal. When the high pressure flow rate L is smaller than the predetermined flow rate α, the 2 pump work rate Pws is smaller than the 1 pump work rate Pwb. This is because, when the flow rate is small, when supplying only to the high-pressure hydraulic operating unit 2 and the low-pressure supplied unit 3 with only the large-capacity oil pump Pb, the low-pressure supplied unit 3 is also high-pressure. This is because a large flow rate is supplied at a high rate, so that excessive work is performed and energy consumption is increased.

  Further, when the high-pressure flow rate L is larger than the predetermined flow rate α (for example, when the width of the pulley needs to be changed instantaneously for sudden gear shifting and the flow rate supplied to the hydraulic operation unit 2 increases greatly), 2 One pump power Pwb is smaller than pump power Pws. This is because when the small-capacity oil pump Ps tries to supply a large amount of oil, a large load acts on the small-capacity oil pump Ps and the motor MOT that drives the small-capacity oil pump Ps, and the power loss of the motor MOT increases. is there.

  Therefore, in the hydraulic circuit 1 of the present embodiment, when the high pressure flow rate L is smaller than the predetermined flow rate α, as shown in FIG. 2A, a large capacity oil that operates using the driving force of the internal combustion engine ENG. The low-pressure oil supplied from the pump Pb is pressurized to a high pressure by the small-capacity oil pump Ps operated by the electric motor MOT via the first flow path L1. The oil pressurized to a high pressure by the small-capacity oil pump Ps is supplied to the hydraulic operation unit 2 via the second flow path L2.

  Further, when the high-pressure flow rate L is higher than the predetermined flow rate α, the hydraulic circuit 1 stops the operation of the small-capacity oil pump Ps by the electric motor MOT, as shown in FIG. 2B, and the large-capacity oil pump Pb. Only the oil is pressurized to a high pressure, and the hydraulic pressure is supplied to the hydraulic operation section 2 via the third flow path L3.

  Thus, by selecting the operation or stop of the small-capacity oil pump Ps according to the high pressure flow rate L, the energy consumption of the entire hydraulic circuit 1 can be optimized. Furthermore, the small-capacity oil pump Ps and the electric motor MOT that drives the small-capacity oil pump Ps can be configured so that the maximum high-pressure flow rate L that can be supplied by the small-capacity oil pump Ps is at least a predetermined flow rate α or less. At this time, the small-capacity oil pump Ps and the electric motor MOT that drives the small-capacity oil pump Ps can be configured to be relatively small, so that energy consumption can be reduced. Thus, an energy efficient hydraulic circuit can be provided.

(2. Detailed configuration of hydraulic circuit)
Next, a detailed configuration of the hydraulic circuit 1 described with reference to FIG. 2 will be described with reference to FIG.

  The hydraulic circuit 1 is used for a belt-type or chain-type continuously variable transmission T (so-called friction drive) including a torque converter.

  The continuously variable transmission T includes a pair of input-side pulleys Dr, a pair of output-side pulleys Dn, and a belt or chain (not shown) that can transmit power between the input-side pulley Dr and the output-side pulley Dn. Prepare.

  The pair of input-side pulleys Dr comprises a pulley (movable pulley) movable along an input shaft (not shown) of the continuously variable transmission T and a fixed pulley (fixed-side pulley). In accordance with the supply of oil, the side pressure of the movable pulley of the input side pulley Dr changes, and the width in the axial direction of the input shaft of the input side pulley Dr changes. In this way, by adjusting the supplied oil, the clamping pressure of the belt between the pair of input-side pulleys Dr is adjusted.

  The pair of output pulleys Dn includes a pulley (movable pulley) movable along an output shaft (not shown) of the continuously variable transmission T and a fixed pulley (fixed pulley). As the oil is supplied, the lateral pressure of the movable pulley of the output pulley Dn changes, and the width of the output shaft of the output pulley Dn in the axial direction changes. In this way, by adjusting the supplied oil, the clamping force of the belt between the pair of output side pulleys Dn is adjusted.

  Here, in the input side pulley Dr and the output side pulley Dn, the side pressure means that the movable side input side pulley Dr and the output side pulley Dn are moved along the axial direction of the input shaft and the output shaft. It refers to the pressure that pushes toward Dr and the output pulley Dn. As the side pressure increases and the pinching pressure increases, the wrapping radius of the belt in the input side pulley Dr or the output side pulley Dn increases. The transmission ratio of the continuously variable transmission T is adjusted steplessly by controlling the hydraulic pressure supplied to the input side pulley Dr and the output side pulley Dn (that is, controlling the side pressure or the clamping pressure).

  2 corresponds to the input-side pulley Dr, the output-side pulley Dn, and the clutch C that operates with high hydraulic pressure, as shown in FIG.

  Referring to FIG. 4, the hydraulic circuit 1 includes the first to eighth pressure control valves 11 to 18, a ninth oil pressure pump Pb and a small capacity oil pump Ps shown in FIG. A pressure control valve 30, first to sixteenth oil passages R <b> 1 to R <b> 16, and a direction control valve 21 are provided.

  The first, third, fourth, and fifth four pressure control valves 11, 13, 14, and 15 can arbitrarily change the hydraulic pressure according to the current supplied to the linear solenoid instructed from the management ECU 125. It is a pressure control valve. Note that the fourth and fifth pressure control valves 14 and 15 are so-called normal open types in which a primary side port (not shown) and a secondary side port (not shown) communicate with each other when power is not supplied to the linear solenoid. It is configured as a valve. Therefore, if the energization amount to the linear solenoid is 0, the valve is opened and the hydraulic pressure reaches the maximum value. On the other hand, the first and third pressure control valves 11 and 13 are of a so-called normal close type in which a primary port (not shown) and a secondary port (not shown) communicate with each other in a state where current is supplied to the linear solenoid. It is configured as a valve. Therefore, if the energization amount to the linear solenoid is zero, the valve is closed and the hydraulic pressure becomes zero. The sixth and seventh pressure control valves 16 and 17 are pilot operation type pressure control valves, and are pressure control valves capable of arbitrarily changing the hydraulic pressure by changing the pilot pressure supplied from the outside. The second pressure control valve 12 reduces the hydraulic pressure supplied from the input side so as to become a predetermined hydraulic pressure, and the eighth pressure control valve 18 supplies the hydraulic pressure when the hydraulic pressure supplied from the input side is greater than or equal to a predetermined pressure. To do.

The direction control valve 21 includes a first port 21a, a second port 21b, a third port 21c, a fourth port 21d, a fifth port 21e, and a sixth port 21f. Further, the direction control valve 21 has a seventh port 21g to which hydraulic pressure is supplied as a pilot pressure. The direction control valve 21 communicates with the first port 21a, the second port 21b, and the third port 21c, and with the fourth port 21d, the fifth port 21e, and the Switch communication with 6 port 21f. The hydraulic pressure input to the seventh port 21g is equal to the pilot pressure _PDRC of the sixth pressure control valve 16 output from the fourth pressure control valve 14.

  Specifically, when the hydraulic pressure input to the seventh port 21g is lower than a predetermined pressure, the direction control valve 21 causes the first port 21a and the third port 21c to communicate with each other, and the second port 21b and the third port 21 The communication with the port 21c is released, the fourth port 21d and the sixth port 21f are connected, and the communication between the fifth port 21e and the sixth port 21f is released.

  Further, when the hydraulic pressure input to the seventh port 21g is higher than the predetermined pressure, the direction control valve 21 causes the second port 21b and the third port 21c to communicate with each other, and the first port 21a and the third port 21c. The communication between the fourth port 21e and the sixth port 21f is released, and the communication between the fourth port 21d and the sixth port 21f is released. Although details are omitted, the directional control valve 21 adjusts the hydraulic pressure by the first, third, fourth, and fifth pressure control valves 11, 13, 14, and 15 driven by the linear solenoid when the electrical system is abnormal. Even in the case where the vehicle cannot be operated, a constant hydraulic pressure is supplied to the input-side pulley Dr and the output-side pulley Dn of the continuously variable transmission T so that the vehicle can continue running.

  The hydraulic pressure supplied to the first oil passage R1 from the large-capacity oil pump Pb driven by the internal combustion engine ENG is regulated by the ninth pressure control valve 30, and then supplied to the second oil passage R2. The second oil passage R2 branches into a third oil passage R3 and a fourth oil passage R4. The third oil passage R3 is connected to a small capacity oil pump Ps (an oil pump driven by the electric motor MOT).

  The small-capacity oil pump Ps further pressurizes the hydraulic pressure supplied from the third oil passage R3 and outputs it to the fifth oil passage R5. The small-capacity oil pump Ps is configured to pressurize the oil pumped from the oil tank 40 via the second check valve 42 and output it to the fifth oil path R5.

  A first check valve 41 is provided in the middle of the fourth oil passage R4 that bypasses the small capacity oil pump Ps. The fourth oil passage R4 is connected to the fifth oil passage R5. The first check valve 41 allows oil to flow from the connection point between the fourth oil passage R4 and the second oil passage R2 to the connection point between the fourth oil passage R4 and the fifth oil passage R5. It is provided to prevent the oil from flowing in the direction opposite to the direction.

  The fourth oil passage R4 and the fifth oil passage R5 are connected to the sixth oil passage R6 and the seventh oil passage R7.

  The sixth oil passage R6 branches into a tenth oil passage R10 and an eleventh oil passage R11. The hydraulic pressure (line pressure) supplied to the sixth oil passage R6 is supplied to the fourth port 30d of the ninth pressure control valve 30 via the sixteenth oil passage R16. The tenth oil passage R <b> 10 is provided with a first pressure control valve 11, and the eleventh oil passage R <b> 11 is connected to the second pressure control valve 12. The first pressure control valve 11 supplies hydraulic pressure to the supplied part 3 through the tenth oil passage R10 when electric power is supplied by the linear solenoid. The relief pressure of the first pressure control valve 11 can be arbitrarily changed with a linear solenoid. The relief pressure when the current supplied to the linear solenoid is zero is set to be equal to or higher than the maximum hydraulic pressure (line pressure) of the tenth oil passage R10. When the current supplied to the linear solenoid is zero, the relief pressure is set via the tenth oil passage R10. It is configured not to supply hydraulic pressure to the supplied part 3. The second pressure control valve 12 reduces the hydraulic pressure supplied from the sixth oil passage R6 to the eleventh oil passage R11 so as to be a predetermined pressure. The second pressure control valve 12 supplies the reduced hydraulic pressure to the third pressure control valve 13, the eighth pressure control valve 18, the fourth port 21 d of the direction control valve 21, the fourth pressure control valve 14, and the fifth pressure control valve 15. Supply to each of the.

  The third pressure control valve 13 reduces the hydraulic pressure of the thirteenth oil passage R13 in accordance with the supply current to the linear solenoid according to the instruction from the management ECU 125, and outputs it to the first port 21a of the direction control valve 21.

  The eighth pressure control valve 18 outputs to the fifth port 21e of the direction control valve 21 when the hydraulic pressure of the fourteenth oil passage R14 is equal to or higher than a predetermined hydraulic pressure.

The fourth pressure control valve 14 reduces the hydraulic pressure supplied from the second pressure control valve 12 via the fifteenth oil passage R15 so as to become the pilot pressure _PDRC of the sixth pressure control valve 16, and Output to the control valve 16. The fourth pressure control valve 14 is also connected to the fifth port 30e of the ninth pressure control valve 30 via the eighth oil passage R8, and is also connected to the seventh port 21g of the direction control valve 21. For this reason, the pilot pressure _PDRC of the sixth pressure control valve 16 output from the fourth pressure control valve 14 is also output to the fifth port 30e of the ninth pressure control valve 30 and the seventh port 21g of the direction control valve 21. Is done.

The fifth pressure control valve 15 reduces the hydraulic pressure supplied from the second pressure control valve 12 via the fifteenth oil passage R15 so as to become the pilot pressure _PDNC of the seventh pressure control valve 17, and Output to the control valve 17. The fifth pressure control valve 15 is also connected to the sixth port 30f of the ninth pressure control valve 30 via the ninth oil passage R9. Therefore, the pilot pressure _PDNC of the seventh pressure control valve 17 output from the fifth pressure control valve 15 is also output to the sixth port 30 f of the ninth pressure control valve 30.

  The seventh oil passage R <b> 7 is connected to the sixth pressure control valve 16 and the seventh pressure control valve 17. The sixth pressure control valve 16 reduces the hydraulic pressure supplied from the seventh oil passage R7 to a predetermined pressure corresponding to the pilot pressure supplied from the fourth pressure control valve 14, and the input side pulley of the continuously variable transmission T. Supply to Dr. The seventh pressure control valve 17 reduces the hydraulic pressure supplied from the seventh oil passage R7 to a predetermined pressure corresponding to the pilot pressure supplied from the fifth pressure control valve 15, and the output side pulley of the continuously variable transmission T. Supply to Dn.

  Here, the first flow path L1 shown in FIG. 2 corresponds to the first oil path R1, the second oil path R2, and the third oil path R3 shown in FIG. Further, the second flow path L2 shown in FIG. 2 corresponds to the fifth oil path R5 and the seventh oil path R7 shown in FIG. Further, the third flow path L3 shown in FIG. 2 corresponds to the first oil path R1, the second oil path R2, the fourth oil path R4, and the seventh oil path R7 shown in FIG.

  The ninth pressure control valve 30 includes a first spool 31 and a second spool 32 inside. The second spool 32 is biased toward the first spool 31 (left side in FIG. 4) by the first elastic member 33. Further, the first spool 31 is urged toward the side away from the second spool 32 (left side in FIG. 4) by the second elastic member 34 disposed between the first spool 31 and the second spool 32.

  The ninth pressure control valve 30 includes first to seventh seven ports 30a to 30g. The first port 30a is connected to the first oil passage R1, and is supplied with hydraulic pressure from the large-capacity oil pump Pb. The second port 30b is provided at the same position in the axial direction as the first port 30a, always communicates with the first port 30a, and is connected to the second oil passage R2. The third port 30c is provided on the side farther from the second spool 32 than the second port 30b, and is connected to the supplied portion 3 via a twelfth oil passage R12 which is a low oil pressure lubricating oil passage.

  The fourth port 30d is provided on the side farther from the second spool 32 than the third port 30c, and is supplied with hydraulic pressure (line pressure) supplied to the sixth oil passage R6 (and the seventh oil passage R7). . An annular groove is provided around the first spool 31 in a portion corresponding to the fourth port 30d, and the first spool 31 resists the urging force of the second elastic member 34 by the hydraulic pressure supplied from the fourth port 30d. Then, a force in a direction approaching the second spool 32 (right direction in FIG. 4) is generated.

The fifth port 30e is provided between the first spool 31 and the second spool 32. The fifth port 30e has a pilot pressure P for the sixth pressure control valve 16 output from the fourth pressure control valve 14. _DRC is supplied. The sixth port 30f is provided on the side of the second spool 32 away from the first spool 31, and the sixth port 30f has a pilot pressure for the seventh pressure control valve 17 output from the fifth pressure control valve 15. P _DNC is supplied.

  The seventh port 30g is provided on the side of the first spool 31 away from the second spool 32, and the seventh port 30g is supplied with the hydraulic pressure output from the sixth port 21f of the direction control valve 21. The hydraulic pressure is also supplied to the sixth pressure control valve 16.

In the ninth pressure control valve 30, either higher pilot pressure of the pilot pressure P _DNC output from the pilot pressure P _DRC and the fifth pressure control valve 15 which is output from the fourth pressure control valve 14 is first spool It acts as a force for moving 31 in the direction of separating from the second spool 32 (the left direction in FIG. 4). That is, when the pilot pressure _PDRC output from the fourth pressure control valve 14 is higher, the first spool 31 moves in a direction away from the second spool 32 (left direction in FIG. 4). On the other hand, when the pilot pressure _PDNC output from the fifth pressure control valve 15 is higher, the second spool 32 approaches the first spool 31 against the urging force of the second elastic member 34 (see FIG. 4). The first spool 31 moves in the direction away from the second spool 32 (the left direction in FIG. 4) by the pressing force from the second spool 32.

  Here, in order to appropriately adjust the gear ratio of the continuously variable transmission T, at least the higher hydraulic pressure among the hydraulic pressures required for the input-side pulley Dr and the output-side pulley Dn, that is, the line pressure P_line is the sixth oil passage. It must be supplied to R6 and the seventh oil passage R7. According to the hydraulic circuit 1 of FIG. 2, since the hydraulic pressure supplied to the sixth oil passage R6 and the seventh oil passage R7 is supplied to the fourth port 30d, it is supplied to the sixth oil passage R6 and the seventh oil passage R7. When the hydraulic pressure is changed and the first spool 31 moves, the flow rate of oil discharged from the third port 30c changes. Thereby, the hydraulic pressure supplied to the sixth oil passage R6 and the seventh oil passage R7 is maintained at a desired line pressure P_line.

  Specifically, when the large-capacity oil pump Pb is supplying the line pressure P_line, the small-capacity oil pump Ps is driven to further increase the hydraulic pressure supplied from the large-capacity oil pump Pb by ΔP. Then, since the line pressure P_line increases, as shown in FIG. 5, the first spool 31 moves in a direction approaching the second spool 32 against the urging force of the second elastic member 34 (right direction in FIG. 5). To do. When the first spool 31 moves in the direction approaching the second spool 32 (the right direction in FIG. 5), the path between the third port 30c and the first port 30a and the second port 30b expands, and the large-capacity oil pump Part of the hydraulic pressure supplied from the Pb to the ninth pressure control valve 30 is relieved from the third port 30c to the supplied portion 3 via the twelfth oil passage R12. As a result, the hydraulic pressure P_pb supplied from the large-capacity oil pump Pb to the small-capacity oil pump Ps via the ninth pressure control valve 30 decreases to “P_line−ΔP”. Although details are omitted, similarly, when the operation of the small-capacity oil pump Ps is stopped from this state and ΔP = 0, there is a gap between the third port 30c and the first port 30a and the second port 30b. And the hydraulic pressure P_pb supplied by the large-capacity oil pump Pb automatically rises again to become the line pressure P_line.

  In the hydraulic circuit 1 configured as described above, the line pressure P_line is adjusted by enlarging / reducing the path between the third port 30c, the first port 30a, and the second port 30b. If the internal combustion engine ENG is driven, as shown in FIG. 5, oil flows from the first port 30a to the second port 30b and the third port 30c, but the operation of the internal combustion engine ENG is stopped as when idling is stopped. Then, as shown in FIG. 6, air flows backward from the supplied part 3 due to the negative pressure of the small-capacity oil pump Ps, and the supplied part 3 → the twelfth oil passage R12 → the third port 30c → the second port. 30b → Air may enter the small-capacity oil pump Ps via the third oil passage R3. A method for preventing the air from entering the small-capacity oil pump Ps will be described in detail later in (4. Control of hydraulic circuit when operation of internal combustion engine is stopped).

(3. Operation of hydraulic circuit)
Next, the operation of the hydraulic circuit 1 will be described for each state of the vehicle in which the hydraulic circuit 1 is mounted (“normal driving”, “sudden shifting”, “coasting down”, and “stopped”). .

(3-1. During normal driving)
With reference to FIG. 7, the operation of the hydraulic circuit 1 during normal travel will be described. During normal travel, the internal combustion engine ENG is operating, and the large-capacity oil pump Pb is driven by the internal combustion engine ENG. The small-capacity oil pump Ps pressurizes oil supplied from the large-capacity oil pump Pb via the ninth pressure control valve 30, and outputs a high hydraulic pressure to the fifth oil passage R5. The first spool 31 of the ninth pressure control valve 30 is controlled so that the first port 30a, the second port 30b, and the third port 30c communicate with each other, and the pressure control valve 11 is not energized. Yes. For this reason, during normal driving, “oil tank 40 → large capacity oil pump Pb → first oil path R1 → 9th pressure control valve 30 → second oil path R2 → third oil path R3 → small capacity oil pump Ps → first High oil pressure is supplied through a route of 5 oil passage R5 → sixth oil passage R6, seventh oil passage R7, and “oil tank 40 → large capacity oil pump Pb → first oil passage R1 → 9th pressure control valve 30 → The low hydraulic pressure is supplied to the supplied part 3 through the route of the twelfth oil passage R12.

  Here, the hydraulic pressure supplied to the sixth oil passage R6 and the seventh oil passage R7, that is, the line pressure is represented by P_line, and the small oil pressure is reduced from the large capacity oil pump Pb through the first oil passage R1 to the third oil passage R3. The hydraulic pressure supplied to the displacement oil pump Ps is represented by P_pb. At this time, the pressure ΔP that is applied by the small-capacity oil pump Ps is “P_line−P_pb”, and the torque τ for driving the small-capacity oil pump Ps is expressed by “(P_line−P_pb) · V / 2π ". Accordingly, the torque τ for driving the small-capacity oil pump Ps can be reduced as compared with the case where the oil that has not been pressurized is pressurized by the small-capacity oil pump Ps.

(3-2. Sudden shift)
With reference to FIG. 8, the operation of the hydraulic circuit 1 at the time of sudden shift will be described. At the time of sudden shift, it is necessary to change the gear ratio of the continuously variable transmission T abruptly, the flow rate of oil supplied to either the input side pulley Dr or the output side pulley Dn is very large, and the high pressure flow rate L is More than the predetermined flow rate α. Accordingly, the operation of the small capacity oil pump Ps is stopped, and only the large capacity oil pump Pb is operated. The first spool 31 of the ninth pressure control valve 30 is controlled so that the first port 30a, the second port 30b, and the third port 30c communicate with each other, and the first pressure control valve 11 is not energized. Has been. For this reason, at the time of sudden shift, the input side pulley Dr or the route of “oil tank → large capacity oil pump Pb → first oil passage R1 → second oil passage R2 → fourth oil passage R4 → seventh oil passage R7” A high oil pressure is supplied to the output pulley Dn, and the low oil pressure is supplied through a route of “oil tank 40 → large capacity oil pump Pb → first oil passage R1 → 9th pressure control valve 30 → 12th oil passage R12”. 3 is supplied. Thus, at the time of sudden shift, the operation of the small capacity oil pump Ps is stopped and only the large capacity oil pump Pb is operated, so that the energy efficiency of the hydraulic circuit 1 can be improved.

(3-3. During coasting down)
With reference to FIG. 9, the operation of the hydraulic circuit 1 at the time of coasting down when idling is stopped will be described. At the time of coasting down when idling is stopped, the drive of the internal combustion engine ENG is stopped and the clutch C is disengaged, so the rotational speed Ne becomes zero. For this reason, the large-capacity oil pump Pb does not operate. Therefore, the small capacity oil pump Ps pressurizes the oil pumped from the oil tank 40 and outputs a high oil pressure to the fifth oil passage R5. At this time, as will be described in detail in (4. Control of Hydraulic Circuit when Operation of Internal Combustion Engine is Stopped), the third port 30c of the ninth pressure control valve 30 is connected between the first port 30a and the second port 30b. The path is completely closed. Further, the pressure control valve 11 is energized, and the hydraulic pressure is supplied to the supplied part 3 through the tenth oil passage R10. Therefore, at the time of coasting down, “oil tank 40 → second check valve 42 → third oil passage R3 → small capacity oil pump Ps → fifth oil passage R5 → sixth oil passage R6, seventh oil passage R7” The high oil pressure is supplied through the path “oil tank 40 → second check valve 42 → third oil path R3 → small capacity oil pump Ps → fifth oil path R5 → sixth oil path R6 → tenth oil path R10”. The low hydraulic pressure is supplied to the supplied part 3 through the path of “(first pressure control valve 11)”.

(3-4. When stopped)
With reference to FIG. 10, the operation of the hydraulic circuit 1 when the vehicle stops idling will be described. When the vehicle stops idling, the operation of the internal combustion engine ENG stops and the rotational speed Ne is 0, so the large-capacity oil pump Pb does not operate. Therefore, the small capacity oil pump Ps pressurizes the oil pumped from the oil tank 40 and outputs a high oil pressure to the fifth oil passage R5. At this time, as will be described in detail in (4. Control of hydraulic circuit when operation of internal combustion engine is stopped), the third port 30c of the ninth pressure control valve 30 is connected between the first port 30a and the second port 30b. Is completely closed. Further, when the vehicle stops, the rotating element of the supplied portion 3 also stops and does not require lubrication, so the pressure control valve 11 is not energized. For this reason, when the vehicle stops, the route of “oil tank 40 → second check valve 42 → third oil passage R3 → small capacity oil pump Ps → fifth oil passage R5 → sixth oil passage R6, seventh oil passage R7”. The high hydraulic pressure is supplied to the supplied portion 3, and the hydraulic pressure is not supplied to the supplied portion 3.

(4. Control of hydraulic circuit when operation of internal combustion engine is stopped)
Next, the control of the hydraulic circuit 1 when the operation of the internal combustion engine ENG stops (idling stop) will be described.

FIG. 11 is a graph showing an example of changes over time of the rotational speed Ne, the pilot pressures P _DRC and P _DNC and the line pressure P_line of the internal combustion engine ENG when the operation of the internal combustion engine ENG stops. When the driving of the internal combustion engine ENG is stopped in the vehicle that is traveling by the driving force of the internal combustion engine ENG, for example, because the AP opening becomes 0, as shown in FIG. 11, the management ECU 125 rotates the rotational speed Ne of the internal combustion engine ENG. as any higher pilot pressure of the pilot pressure P _DRC and the pilot pressure P _DNC is further gradually increases with the decrease in the control of the linear solenoid of the fourth pressure control valve 14 or the fifth pressure control valve 15 To do. Since the fourth pressure control valve 14 and the fifth pressure control valve 15 are configured as normally open type valves, the pilot pressure P _DRC or the pilot pressure P _DNC is reduced by reducing the energization amount to the linear solenoid. Maximum.

In the example shown in FIG. 11, since towards the pilot pressure _{P DNC} is high, management ECU125 controls the fifth pressure control valve 15 in accordance with the decrease in the rotational speed Ne of the internal combustion engine ENG to the seventh pressure control valve 17 The pilot pressure _PDNC is controlled to gradually increase. When the pilot pressure P _{DNC of} the seventh pressure control valve 17 is increased, the pilot pressure P _DNC is also output to the sixth port 30f of the ninth pressure control valve 30 via the ninth oil passage R9, and therefore, as shown in FIG. As described above, the second spool 32 of the ninth pressure control valve 30 moves in the direction approaching the first spool 31 against the urging force of the second elastic member 34 (the left direction in FIG. 12) and moves to the first spool 31. At the end, the first spool 31 moves in the direction away from the second spool 32 (the left direction in FIG. 12) by the pressing force from the second spool 32. In this state, since the path between the third port 30c of the ninth pressure control valve 30 and the first port 30a and the second port 30b is completely closed, the path from the supplied portion 3 to the small capacity oil pump Ps Is cut off and air does not enter the small capacity oil pump Ps.

The following describes why controlled to be further increased either higher pilot pressure of the pilot pressure P _DRC and the pilot pressure P _DNC. As described above, when the driving of the internal combustion engine ENG is stopped, the hydraulic pressure supplied to the sixth oil passage R6 and the seventh oil passage R7, that is, the line pressure, is supplied by the small-capacity oil pump Ps. In other words, the line pressure is determined by the output of the small capacity oil pump Ps. Therefore, even if the higher pilot pressure is further increased, the line pressure is not affected, and the hydraulic pressures of the hydraulic oil supplied to the input-side pulley Dr and the output-side pulley Dn do not change. Therefore, according to this control, it is possible to prevent air from entering the small-capacity oil pump Ps without changing the gear ratio of the continuously variable transmission T.

  If the lower pilot pressure is higher than the higher pilot pressure, air can be prevented from entering the small-capacity oil pump Ps, but it is determined by the output of the small-capacity oil pump Ps. Although the line pressure does not change, the hydraulic pressure of the hydraulic oil supplied to the input-side pulley Dr or the output-side pulley Dn changes, so that the gear ratio of the continuously variable transmission T changes. Therefore, when the change of the transmission ratio of the continuously variable transmission T is allowed, the lower pilot pressure is set higher than the higher pilot pressure to prevent air from entering the small capacity oil pump Ps. Also good.

In addition, the management ECU 125 stops the drive of the internal combustion engine ENG and the energization amount to the linear solenoid of the pressure control valve that supplies the higher pilot pressure after the rotational speed Ne of the internal combustion engine ENG becomes 0 is reduced to 0. It is preferable to control so that. That is, in the example shown in FIG. 11, since towards the pilot pressure _{P DNC} is high, management ECU125 is after the rotation speed Ne of the internal combustion engine ENG becomes 0, controls the linear solenoid of the fifth pressure control valve 15 Then, the pilot pressure _PDNC of the seventh pressure control valve 17 is controlled to gradually increase. As a result, while the large-capacity oil pump Pb driven by the internal combustion engine ENG is still rotating, the path to the supplied part 3 is completely closed, and the line pressure is not intended by not supplying oil. It can be prevented from rising.

(5. Learning control by hydraulic circuit)
Next, control capable of improving the prediction accuracy of the driving torque of the small capacity oil pump Ps will be described.

  There is an individual difference in the small-capacity oil pump Ps, but due to the individual difference, a difference occurs between the target hydraulic pressure and the hydraulic pressure actually supplied by the small-capacity oil pump Ps. When the drive distribution of the internal combustion engine deviates from the ideal value due to this difference, the system efficiency is lowered. Therefore, it is necessary to improve the prediction accuracy of the drive torque of the small capacity oil pump Ps.

  FIG. 13 is a flowchart for setting the fluid friction torque coefficient. As shown in FIG. 13, the management ECU 125 determines whether or not there is an idling stop request (step S101). If there is an idling stop request, the management ECU 125 proceeds to step S103. In step S103, the management ECU 125 shifts to the stop mode. Subsequently, the management ECU 125 executes a test cycle of the small capacity oil pump Ps (step S105). Next, the management ECU 125 stores the output value of the sensor (step S107). Next, the management ECU 125 determines whether or not the test cycle of the small-capacity oil pump Ps has ended (step S109). If completed, the process proceeds to step S117, and if not completed, the process returns to step S107.

  In step S113, the management ECU 125 shifts to the normal travel mode. Subsequently, the management ECU 125 determines whether learning of the small-capacity oil pump Ps is being executed (step S115). If it is being executed, the process proceeds to step S117, and if it is not being executed, the process is terminated. In step S117, the management ECU 125 calculates fluid friction torque coefficients Cr and Cf. Next, the management ECU 125 determines whether the error determination result of the learning result is normal or abnormal (step S119). If normal, the process proceeds to step S121, and if abnormal, the process ends. In step S121, the management ECU 125 updates the fluid friction torque coefficients Cr and Cf.

For the determination of the error determination result performed in step S119, the following indices (a) to (d) are used.
(A) less than the number of samples is defined value (b) the coefficient of determination ^{(R 2} value) is less than the prescribed value (c) calculation result NaN: a (Not a Number-Number) (d) calculation result _C f, C Out of upper / lower limit range of _r = can be defined by upper / lower limit of tolerance variation Note that the coefficient of determination (R ² value) is expressed by the following equation (2).

FIG. 14 is a diagram illustrating the timing of the test cycle of the small-capacity oil pump Ps when the vehicle stops from normal travel. As shown in FIG. 14, immediately after the vehicle stops, by direct pressure P _H and low-capacity oil pump Ps according to the high-pressure side pulley of the continuously variable transmission T is, from the hydraulic P _{H =} fifth pressure control valve 15 The output pilot pressure _PDNC for the seventh pressure control valve 17 is obtained. The test cycle of the small capacity oil pump Ps starts from this state.

なお、Ｄｐは、設計値として得られる小容量オイルポンプＰｓの理論吐出量である。また、流体摩擦トルク係数Ｃｒ，Ｃｆも設計値として得られる。また、作動油の絶対粘度μは、作動油の温度によって異なる値である。また、回転数ωは、小容量オイルポンプＰｓに内蔵されたホール素子から得られる。 FIG. 15 shows a function of calculating the discharge pressure of the small capacity oil pump Ps based on the pulley pressure request value of the continuously variable transmission T and calculating the drive torque of the small capacity oil pump Ps based on the discharge pressure and the like. It is a block diagram. As shown in FIG. 15, the management ECU 125 calculates the discharge pressure Δp of the small-capacity oil pump Ps based on the pulley pressure request value of the continuously variable transmission T, and then calculates the discharge pressure Δp, the absolute viscosity μ of the hydraulic oil, and Based on the rotational speed ω of the small-capacity oil pump Ps and the like, the drive torque Ta of the small-capacity oil pump Ps is calculated using the following equation (3).

Dp is a theoretical discharge amount of the small capacity oil pump Ps obtained as a design value. Further, the fluid friction torque coefficients Cr and Cf are also obtained as design values. The absolute viscosity μ of the hydraulic oil is a value that varies depending on the temperature of the hydraulic oil. Further, the rotational speed ω is obtained from a Hall element built in the small-capacity oil pump Ps.

  In the present embodiment, the prediction accuracy of the driving torque of the small capacity oil pump Ps can be improved by the second term and the third term on the right side of the equation (3).

式（４）を「Ｙ＝ＣｆＸ＋Ｃｒ」と表すと、ＣｆとＣｒは線形結合であり、ＸとＹは実験結果より既知である。このため、「Ｙ＝ＣｆＸ＋Ｃｒ」によって表される線形近似式の傾き（Ｃｆ）と切片（Ｃｒ）を最小二乗法により求めることで、流体摩擦トルク係数Ｃｒ，Ｃｆが得られる。 The fluid friction torque coefficients Cr and Cf included in the second and third terms on the right side of Equation (3) are obtained as follows. Equation (3) can be transformed into Equation (4).

When Expression (4) is expressed as “Y = CfX + Cr”, Cf and Cr are linear combinations, and X and Y are known from experimental results. For this reason, the fluid friction torque coefficients Cr and Cf are obtained by obtaining the slope (Cf) and intercept (Cr) of the linear approximation represented by “Y = CfX + Cr” by the least square method.

（添え字ｉは、標本を表す。） FIG. 16 is a flowchart showing a subroutine of the process of step S117 shown in FIG. As shown in FIG. 16, the management ECU 125 acquires parameters necessary for obtaining the above X and Y from a sensor or the like included in the hydraulic circuit 1 (step S201). Since X and Y can be expressed as follows, the parameters acquired in step S201 are the current I flowing through the armature of the motor MOT that drives the small-capacity oil pump Ps and the small-capacity oil pump Ps. The discharge pressure Δp, the hydraulic oil temperature _TATF, and the rotation speed ω of the small capacity oil pump Ps.

(The subscript i represents a sample.)

Next, management ECU125 converts the parameter I obtained in step S201 in the drive torque Ta of the small-capacity oil pump Ps, it converts the parameters _{T ATF} absolutely viscosity of the hydraulic oil mu (step S203). Next, the management ECU 125 calculates the sample values Xi and Yi, and obtains the slope (Cf) and intercept (Cr) of the linear approximation obtained from the sample values Xi and Yi by the least square method, thereby obtaining the fluid friction torque coefficient. Cr and Cf are calculated (step S205). The fluid friction torque coefficients Cr and Cf obtained by the least square method are expressed as follows.

  Thus, the driving torque Ta of the small capacity oil pump Ps is calculated from the above equation (3) including the fluid friction torque coefficients Cr and Cf as parameters, and the fluid friction torque coefficients Cr and Cf are obtained from the sample values Xi and Yi. It is obtained by obtaining the slope (Cf) and intercept (Cr) of the obtained linear approximation equation by the least square method. Therefore, the prediction accuracy of the driving torque of the small capacity oil pump Ps can be improved. As a result, the error between the target oil pressure and the oil pressure actually supplied by the small capacity oil pump Ps is reduced, and the system efficiency can be improved. In order to perform the above control, for example, a pulley pressure sensor provided for controlling the pulley pressure with high accuracy can be used. Therefore, the driving torque of the small-capacity oil pump Ps can be increased without providing a dedicated sensor. It can be controlled with high accuracy. As a result, the number of sensors can be minimized.

  As described above, in the present embodiment, the hydraulic pressure output from the large capacity oil pump Pb driven by the internal combustion engine ENG is supplied to the small capacity oil pump Ps driven by the electric motor MOT. For this reason, the small-capacity oil pump Ps only needs to increase the pressure by a deficiency relative to the hydraulic pressure output from the large-capacity oil pump Pb. Decrease. For this reason, the energy consumption in the small capacity oil pump Ps can be reduced.

  Further, when a large amount of power is required for driving the oil pump, such as when supplying a large amount of oil to the hydraulic operation unit 2, the hydraulic operation is performed from the large capacity oil pump Pb without using the small capacity oil pump Ps. When the high hydraulic pressure is directly supplied to the section 2, the total power for driving each oil pump may be smaller than when the small-capacity oil pump Ps is used.

  In such a case, the operation of the small-capacity oil pump Ps is stopped, and the hydraulic pressure output from the large-capacity oil pump Pb is supplied to the hydraulic operation unit 2 via the third flow path L3. It is possible to reduce the maximum output power required for the oil pump Ps. For this reason, a relatively small device can be used as the small-capacity oil pump Ps, and as a result, energy efficiency when driving the small-capacity oil pump Ps can be improved.

  Further, the management ECU 125 controls the relief pressure so that the first port 30a and the second port 30b do not communicate with the third port 30c when idling is stopped when the large-capacity oil pump Pb is stopped. As a result, the path from the supplied portion 3 to the small-capacity oil pump Ps is blocked, so that it is possible to prevent air from flowing backward through the path and mixing into the small-capacity oil pump Ps when idling is stopped. Furthermore, since a check valve for preventing air from entering the small-capacity oil pump Ps is unnecessary, the hydraulic circuit 1 can be reduced in size and cost.

In addition, by controlling the pilot pressure P _DRC or pilot pressure P _DNC , whichever is higher, to prevent the air from entering the small-capacity oil pump Ps, the continuously variable transmission T Can be prevented from changing. That is, since the large capacity oil pump Pb is stopped when idling is stopped, the line pressure is supplied by the small capacity oil pump Ps. At this time, be further increased either higher pilot pressure of the pilot pressure P _DRC and the pilot pressure P _DNC, the line pressure is not increased more, the continuously variable transmission T input pulley Dr and the output-side pulley Since each hydraulic pressure of the hydraulic oil to Dn does not change, the gear ratio of the continuously variable transmission T does not change. For this reason, the merchantability of the vehicle equipped with the continuously variable transmission T is not affected.

Moreover, since the fourth pressure control valve 14 or the fifth pressure control valve 15 is a normally open type, the power supply amount to a valve for outputting one higher pilot pressure of the pilot pressure P _DRC and the pilot pressure P _DNC By setting it to 0, power consumption can be reduced while increasing the pilot pressure.

  Further, after the driving of the internal combustion engine ENG is stopped and the rotation of the large-capacity oil pump Pb is completely stopped, the first port 30a, the second port 30b, and the third port 30c of the ninth pressure control valve 30 are stopped. Is completely closed, the path from the supplied portion 3 to the small-capacity oil pump Ps is blocked. For this reason, it is possible to prevent the line pressure from unintentionally rising due to the passage being completely closed while the large-capacity oil pump Pb is rotating and the oil being not supplied to the supplied portion 3. .

  Further, since the drive torque of the small capacity oil pump Ps is calculated from an equation including the fluid friction torque coefficients Cr and Cf as parameters, the drive torque can be calculated with high accuracy. In order to derive the fluid friction torque coefficients Cr and Cf, for example, a pulley pressure sensor provided for controlling the pulley pressure with high accuracy can be used. Therefore, the small-capacity oil pump Ps is provided without providing a dedicated sensor. Can be controlled with high accuracy.

  In addition, this invention is not limited to embodiment mentioned above, A deformation | transformation, improvement, etc. are possible suitably.

DESCRIPTION OF SYMBOLS 1 Hydraulic circuit 2 Hydraulic action part 3 Supply part 11 1st pressure control valve 12 2nd pressure control valve 13 3rd pressure control valve 14 4th pressure control valve 15 5th pressure control valve 16 6th pressure control valve 17 7th Pressure control valve 18 Eighth pressure control valve 21 Direction control valve 30 Ninth pressure control valve 30a 1st port 30b 2nd port 30c 3rd port 30d 4th port 30e 5th port 30f 6th port 30g 7th port 31 1st Spool 32 Second spool 33 First elastic member 34 Second elastic member 40 Oil tank 121 Vehicle speed sensor 123 Rotational speed sensor 125 Management ECU (MG ECU)
127 Drive shaft 129 Drive wheel C Clutch Dn Output side pulley Dr Input side pulley ENG Internal combustion engine L1 First flow path L2 Second flow path L3 Third flow path Pb Large capacity oil pump Ps Small capacity oil pump R1 First oil path R2 2nd oil path R3 3rd oil path R4 4th oil path R5 5th oil path R6 6th oil path R7 7th oil path R8 8th oil path R9 9th oil path R10 10th oil path R11 11th oil path R12 12th oil path R13 13th oil path R14 14th oil path R15 15th oil path R16 16th oil path T continuously variable transmission

Claims (6)

A control device for a hydraulic circuit for supplying hydraulic pressure to a supplied portion to which low hydraulic pressure is supplied and a hydraulic operating portion to which high hydraulic pressure higher than the low hydraulic pressure is supplied,
The hydraulic operating part is
A continuously variable transmission having an input-side pulley and an output-side pulley capable of changing a width by being supplied with the high hydraulic pressure, and capable of adjusting a gear ratio steplessly;
The hydraulic circuit is
A mechanical oil pump driven by a driving source that outputs a driving force for the vehicle to travel;
An oil pump driven by an electric motor and having a smaller capacity than the mechanical oil pump, further pressurizing the hydraulic pressure supplied from the mechanical oil pump and supplying the hydraulic pressure to the hydraulic operating unit;
A first flow path for supplying hydraulic pressure supplied from the mechanical oil pump to the electric oil pump;
A second flow path for supplying the hydraulic pressure supplied from the electric oil pump to the hydraulic operating unit;
A third flow path for supplying the hydraulic pressure supplied from the mechanical oil pump to the hydraulic actuator without passing through the electric oil pump;
A line pressure regulating valve provided between the mechanical oil pump and the first flow path;
A first shift control valve that receives the line pressure of the second flow path or the third flow path and performs control to supply and discharge hydraulic oil to and from the input pulley;
A second shift control valve that receives the line pressure of the second flow path or the third flow path and performs control to supply and discharge hydraulic oil to and from the output pulley;
A first solenoid valve that creates a first pilot pressure and controls the operation of the first shift control valve by acting on the first shift control valve;
A second electromagnetic valve that creates a second pilot pressure and controls the operation of the second shift control valve by acting on the second shift control valve;
The line pressure regulating valve includes a first spool and a second spool inside,
The second spool is biased toward the first spool by a first elastic member;
The first spool is urged toward the side away from the second spool by a second elastic member disposed between the first spool and the second spool,
The line pressure adjusting valve includes a first port to which hydraulic pressure from the mechanical oil pump is supplied;
A second port always in communication with the first port and connected to the first flow path;
A third port provided on a side farther from the second spool than the second port and connected to the supplied portion via a lubricating oil path;
A fourth port which is provided on a side farther from the second spool than the third port and to which the line pressure is supplied;
A fifth port to which the first pilot pressure is supplied;
A sixth port to which the second pilot pressure is supplied,
The first spool is biased toward the second spool by the line pressure supplied from the fourth port, so that the first port, the second port, and the third port are Configured to communicate,
The first spool is urged toward the side away from the second spool by a relief pressure which is the higher of the first pilot pressure and the second pilot pressure. And the communication between the second port and the third port can be cut off,
The control device controls the relief pressure so that the first port, the second port, and the third port do not communicate with each other when idling is stopped when the mechanical oil pump is stopped. . The hydraulic circuit control device according to claim 1,
When the control device controls the relief pressure so that the first port, the second port, and the third port do not communicate with each other, one of the first pilot pressure and the second pilot pressure is selected. A control device for a hydraulic circuit, which controls the first solenoid valve or the second solenoid valve so that a higher pilot pressure is further increased. A control device for a hydraulic circuit according to claim 1 or 2,
The first solenoid valve and the second solenoid valve are normally open solenoid valves,
The said control apparatus is a control apparatus of the hydraulic circuit which controls so that the energization amount to the solenoid valve which outputs the higher pilot pressure of the said 1st pilot pressure or the said 2nd pilot pressure may become zero. The hydraulic circuit control device according to any one of claims 1 to 3,
The control device prevents the first port, the second port, and the third port from communicating after the drive of the drive source is stopped and the rotation of the mechanical oil pump is completely stopped. A control device of a hydraulic circuit for controlling the relief pressure. The control device for a hydraulic circuit according to any one of claims 1 to 4,
The control device calculates a torque of the electric oil pump based on a hydraulic pressure to the output-side pulley when the operation of the drive source is stopped and the vehicle is stopped. . A control device for a hydraulic circuit according to claim 5,
The control device includes a current supplied to the electric motor calculated based on a rotational speed of the electric oil pump determined according to a target line pressure and an oil temperature of the hydraulic oil, and a rotational speed of the electric oil pump. And a hydraulic circuit control device for deriving a fluid friction torque coefficient for calculating the torque of the electric oil pump based on the oil temperature of the hydraulic oil and the oil pressure to the output pulley.

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