JP2001227606A - Belt-type continuously variable transmission for automobile having motor-driven oil pump - Google Patents

Abstract

PROBLEM TO BE SOLVED: To reduce power loss of an oil pump in a belt-type continuously variable transmission and especially to reduce power loss of the oil pump at the time of stopping. SOLUTION: This continuously variable transmission having a primary pulley and a secondary pulley having a belt stretched between the primary and secondary pulleys is provided a high pressure pump 46 driven by a motor 49, and a low pressure pump 47 driven by an engine. Operating fluid from the high pressure pump 46 is pressure-regulated and supplied to a primary oil chamber 18 and a secondary oil chamber 21, and pressure-regulated by a clutch pressure regulating valve 57 to be supplied as clutch pressure to a clutch oil chamber 32a of a forward-backward switching mechanism, a brake oil chamber 45a, and an apply chamber 9a of a lockup clutch of a torque converter. The operating fluid from the low pressure pump 47 is supplied through a lubricating pressure passage 55 to a lubricating part and a release chamber 9b of the lockup clutch.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a belt type continuously variable transmission for an automobile having an electric oil pump whose pulley groove width is adjusted by an electric oil pump.

[0002]

2. Description of the Related Art A belt-type continuously variable transmission (CV) for automobiles.
As T), a metal pulley is provided between a primary pulley having a variable pulley groove width provided on a primary shaft on a driving side and a secondary pulley having a variable pulley groove width provided on a secondary shaft on a driven side, that is, a driven side. There is a type in which a drive belt is stretched, the pulley diameters of a primary pulley and a secondary pulley are changed by hydraulic pressure, and the rotational speed of a secondary shaft is changed steplessly.

[0003] In the belt type CVT, a forward clutch and a reverse rotation between the primary shaft and the crankshaft for transmitting rotation of the crankshaft to the primary shaft in a forward direction in order to transmit the output of the engine to the primary shaft. A forward / reverse switching mechanism having a reverse brake transmitting in the direction is provided. In a belt type CVT provided with a torque converter having a lock-up clutch, hydraulic oil is supplied to each of a lock-up apply chamber and a lock-up release chamber.

In the belt type CVT, power transmission and gear ratio control are performed using hydraulic oil from an oil pump driven by an engine, but a plurality of ranks of hydraulic pressures having different hydraulic pressure levels are used. . In other words, the pulley operating system for changing the groove width of the pulley supplies a high pressure of 2 to 5 MPa in order to generate the required belt clamping force with a cylinder as small as possible. The medium pressure of about 1 MPa is supplied to the power transmission element, and the low pressure of 0.3 to 0.5 MPa is supplied to the cooling and lubrication parts of the torque converter.

[0005] In a normal CVT, an operating oil pressure is generated by one oil pump driven by an engine. When one pump is used, the highest pressure (normally, the hydraulic pressure supplied to the hydraulic circuit) is used. Pulley operating pressure) is generated by a pump. The driving torque _Top of the oil pump is given by: _Top = PD / 2π (P: discharge pressure, D:
(The discharge amount per one rotation of the pump), and the higher the discharge pressure, the larger the _Top . The pump loss other than the pulley operation system is used after being reduced to an appropriate value in the hydraulic circuit, and all the reduced pressure is consumed as a loss such as an increase in oil temperature. In particular, in the structure in which the oil pump is driven by the engine, the discharge amount D is set so that the required flow rate can be secured at the minimum engine speed. And invalid power loss occurs.

The first prior example (Japanese Patent Laid-Open No. 3-213772) is as follows.
Disclosed is a CVT provided with an oil pump having two discharge ports. In this case, when the engine speed is high, hydraulic oil from one discharge port is returned to the suction port to reduce pump loss in a high rotation range. I try to make it.

The second prior example (Japanese Patent Laid-Open No. 10-89445)
A stepped automatic transmission (AT) in which an oil pump is driven by an electric motor is disclosed. In this case, the rotation speed of the pump is made independent of the engine rotation speed, thereby reducing the engine speed at low engine rotation speed. Insufficient flow and power loss at high speeds are eliminated.

[0008]

[Problems to be Solved by the Invention] Reduction of idle speed In order to further improve fuel efficiency, there is a tendency to reduce the idle speed when the vehicle is stopped at a traffic light. When the minimum engine speed decreases, it becomes necessary to increase the capacity of the oil pump in order to secure the function of the CVT. In this case, in the first prior example, even when the discharge amount is switched to half, the surplus amount of the discharge amount increases, and the power loss increases. In the case of the second prior art, the pump speed can be set independently of the engine speed, so that no inconvenience occurs even when the idle speed decreases.

[0009] 2. Stopping the Engine When the Vehicle Stops There is a case where the engine rotation is selectively stopped to suppress fuel consumption when the vehicle stops. In the case of the first preceding example, when the rotation of the engine stops, the operation of the hydraulic system also stops. During the stop, the hydraulic pressure of each part leaks from the cylinder and the hydraulic oil escapes due to gravity drainage. During this time, the hydraulic pressure is not stable. For example,
When an idle stop is performed while waiting for a traffic light, it is required that the traffic light changes and the vehicle can start immediately, so that the time until the hydraulic pressure becomes normal needs to be very short. in addition,
In order to avoid any discomfort in the time lag from the restart of the engine until the vehicle can be started, the normal N range
The goal is to complete the entire operation in about the select operation time for the range. However, the above requirements cannot be satisfied by an engine-driven oil pump.

In the case of the second prior art, if the pump continues to be operated even while the engine is stopped, the problem of operation delay does not occur. However, in order to ensure the select performance, a predetermined discharge amount must be ensured even when the vehicle is stopped. Since the generator does not operate while the engine is stopped, the battery load increases. Normally, in vehicles with an idle stop function, the system automatically restarts the engine when the charged amount decreases.If the pump load is excessive, the engine restarts frequently and achieves the predetermined purpose. It may not be possible.

3. Problems when applying an electric oil pump to a CVT A conventional CVT consumes more pump power than a normal stepped automatic transmission (AT). This is because the CVT requires the same working flow rate as the AT and a higher pump discharge pressure than the AT. Power consumption L for driving the oil pump is represented by the following equation.

L = P × Q / 612 (where L = power consumption (KW), P = generated oil pressure (kg / cm ² ),
Q = discharge rate (l / min) Therefore, simply replacing the oil pump in the conventional system as disclosed in the second prior art with a motor drive causes excessive power consumed by the pump in the standby state and causes the battery to charge. There is a high possibility that the idle stop cannot be maintained due to a shortage.

As described above, in the prior art, the CVT is used to reduce the idling of the engine or to stop the engine.
Function cannot be improved.

An object of the present invention is to reduce power loss of an oil pump in a belt-type continuously variable transmission, and in particular, to reduce power loss of an oil pump when the vehicle is stopped.

[0015]

SUMMARY OF THE INVENTION A belt type continuously variable transmission for an automobile having an electric oil pump according to the present invention is provided with a primary pulley having a variable pulley groove width mounted on a primary shaft and a secondary pulley mounted on a secondary shaft. A low-pressure pump driven by an engine and a high-pressure pump driven by an electric motor, comprising a belt-type continuously variable transmission for a vehicle having a secondary pulley having a variable pulley groove width over which a belt is stretched between a primary pulley and a secondary pulley. And a motor control means for controlling the electric motor according to the running state of the vehicle.

A belt type continuously variable transmission for an automobile having an electric oil pump according to the present invention is provided between a primary pulley having a variable pulley groove width mounted on a primary shaft and a primary pulley mounted on a secondary shaft. A belt-type continuously variable transmission for an automobile, comprising: a secondary pulley having a variable pulley groove width around which a belt is wound, and a forward / reverse switching mechanism for transmitting rotation of the engine to the primary shaft by switching between forward rotation and reverse rotation. A low-pressure pump driven by the engine to supply hydraulic oil to the lubrication unit; a primary oil chamber driven by an electric motor to change the groove width of the primary pulley; and a secondary oil to change the groove width of the secondary pulley. A high-pressure pump for supplying hydraulic oil to the chamber and the forward / reverse switching mechanism; And having a motor control means for controlling the electric motor.

A belt type continuously variable transmission for an automobile having an electric oil pump according to the present invention is provided between a primary pulley having a variable pulley groove width mounted on a primary shaft and a primary pulley mounted on a secondary shaft. For a vehicle having a secondary pulley with a variable pulley groove width around which a belt is wound, a forward / reverse switching mechanism for transmitting the rotation of the engine to the primary shaft by switching between a forward rotation direction and a reverse rotation direction, and a torque converter with a lock-up clutch. A belt-type continuously variable transmission, wherein the low-pressure pump is driven by the engine and supplies hydraulic oil to a release chamber and a lubrication unit of a lock-up clutch, and is driven by an electric motor to change a groove width of the primary pulley. Primary oil chamber, secondary oil chamber that changes the groove width of the secondary pulley Apply chamber of the lock-up clutch, and a high-pressure pump for supplying hydraulic oil to the forward-reverse switching mechanism, and having a motor control means for controlling said electric motor in accordance with a running state of the vehicle.

In the belt-type continuously variable transmission for an automobile having the electric oil pump according to the present invention, the electric motor is a shunt-type DC motor, and the motor control means controls a driving voltage of the motor, an armature voltage, and the like. The electric motor is an induction motor or a permanent magnet type synchronous motor, and the motor control means controls the driving voltage and frequency of the motor. An accumulator for accumulating clutch pressure is provided in a clutch pressure path for supplying hydraulic oil to the forward / reverse switching mechanism. In addition, a check valve is provided in a communication path that connects the discharge port of the low-pressure pump and the suction port of the high-pressure pump, so that the discharge pressure of the low-pressure pump flows into the suction port of the high-pressure pump depending on the discharge pressure of the high pressure pump. It is characterized by the following.

[0019]

Embodiments of the present invention will be described below in detail with reference to the drawings.

FIG. 1 is a schematic diagram showing an example of a drive system of a belt-type continuously variable transmission, that is, a CVT. A crankshaft 1 driven by an engine (not shown) includes a front cover of a pump side case 3 of a torque converter 2. 3a is directly connected to the pump side case 3 via a drive plate 4.
A turbine runner 5 disposed opposite to a pump impeller 3 b provided therein is directly connected to a turbine shaft 6.
A stator 7 is arranged between the pump impeller 3b and the turbine runner 5, and the stator 7 is supported by a one-way clutch 8a attached to a stator support shaft 8. A lock-up clutch 9 is directly connected to the turbine shaft 6 so as to be movable between an engagement position for engaging the drive plate 4 and an open position away from the drive plate 4. Engine power is transmitted via the torque converter 2 or the lock-up clutch 9. The power is transmitted to the turbine shaft 6.

One side of the lock-up clutch 9 is a supply chamber, that is, an apply chamber 9a, and the other side is an open chamber, that is, a release chamber 9b. The hydraulic pressure supplied to the release chamber 9b is circulated through the apply chamber 9a. Thus, the torque converter 2 enters an operating state. On the other hand, by supplying hydraulic pressure to the apply chamber 9a and reducing hydraulic pressure in the release chamber 9b, the lock-up clutch 9 is engaged with the front cover 3a to enter a lock-up state. By adjusting the pressure in the release chamber 9b, slip pressure control is performed so that the lock-up clutch 9 is slid.

The turbine shaft 6 is transmitted to an input shaft of a continuously variable transmission 12, that is, a primary shaft 13 via a forward / reverse switching mechanism 11. A primary pulley 14 is provided on the primary shaft 13, and the primary pulley 14 is a fixed pulley 14 a fixed to the primary shaft 13.
And a movable pulley 14b opposed to the primary shaft 13 so as to be freely slidable in the axial direction by a ball spline or the like. The cone surface distance of the pulley, that is, the pulley groove width is variable. . An output shaft arranged in parallel with the primary shaft 13, that is, a secondary shaft 15 is provided with a secondary pulley 16.
6 is a fixed pulley 16a fixed to the secondary shaft 15
And a movable pulley 16b opposed to the secondary shaft 15 so as to be slidable in the axial direction on the secondary shaft 15 similarly to the movable pulley 14b, and the groove width of the pulley is variable. The entire drive system is incorporated in a case 10 such as a transmission case.

A drive belt 17 is stretched between the primary pulley 14 and the secondary pulley 16, and by changing the groove width of both pulleys 14, 16,
By changing the ratio of the winding diameter to each of the pulleys 14 and 16, the rotational speed of the secondary shaft 15 is continuously variable.

In order to change the groove width of the primary pulley 14, a primary oil chamber 18 is provided between the primary pulley 14 and the movable pulley 14b.
Is mounted on the primary shaft 13, and a plunger 22 forming a secondary oil chamber 21 with the movable pulley 16 b is mounted on the secondary shaft 15 in order to change the groove width of the secondary pulley 16. The plunger 22 forms a cylinder on the secondary side.

The secondary shaft 15 is connected to an intermediate shaft 25 via gears 23 and 24. A gear 26 attached to the intermediate shaft 25 meshes with a final gear 28 of a differential device 27, and the differential device 27
Axles 29a, 29b connected to the wheels 31a, 31
b is attached. In the case of a front wheel drive vehicle, the wheels 31a and 31b are front wheels.

The forward / reverse switching mechanism 11 has a clutch cylinder 32 provided on a forward clutch drum fixed to the turbine shaft 6 and a clutch hub 33 fixed to the primary shaft 13. Is provided with a multi-plate type forward clutch 34. This forward clutch 34
Is incorporated in the clutch cylinder 32. Accordingly, when the hydraulic pressure is supplied to the clutch oil chamber 32 a in the clutch cylinder 32 and the forward clutch 34 is connected, the rotation of the turbine shaft 6 is transmitted to the primary shaft 13 via the clutch hub 33 and the primary shaft 13 It rotates in the same forward direction as the turbine shaft 6.

A sun gear 36 is fixed to the primary shaft 13, and a ring gear 37 is rotatably provided inside the case 10 outside the sun gear 36. Carrier 3 attached to a forward clutch drum unit having a clutch cylinder 32
A pair of planetary pinion gears 41 and 42 that mesh with each other are rotatably mounted on 8, and one planetary pinion gear 41 is meshed with the sun gear 36, and the other planetary pinion gear 42 is meshed with the internal teeth of the ring gear 37. These gears constitute a double pinion type planetary gear. The planetary pinion gears 41 and 42 are shown one by one in FIG. 1 for convenience of drawing, but mesh with each other in pairs, and a plurality of pairs are provided.

A multi-plate reverse brake 43 is provided between the ring gear 37 and the case 10. A hydraulic piston 44 for operating the reverse brake 43 is provided.
Are incorporated in a brake cylinder 45 formed in the case 10. Accordingly, when the hydraulic pressure is supplied to the oil chamber 45a in the brake cylinder 45 to bring the reverse brake 43 into a braking state while the forward clutch 34 is released, the ring gear 37 is fixed to the case 10. The rotation of the carrier 38 together with the turbine shaft 6 causes the sun gear 36 and the primary shaft 1 via the paired planetary pinion gears 41 and 42.
3 rotates in the reverse rotation direction opposite to the turbine shaft 6. The forward clutch 34 and the reverse brake 43 are friction engagement elements of the forward / reverse switching mechanism 11.

FIG. 2 is a circuit diagram showing a hydraulic circuit of the CVT. An electric oil pump or high-pressure pump 46 driven by an electric motor 49 and a low-pressure pump driven by an engine via a pump side case 3 by a crankshaft. 4
The hydraulic fluid is supplied to each of the above-mentioned oil chambers by using the hydraulic fluid 7 as a hydraulic pressure source.

Hydraulic oil in an oil pan 48 flows into the high-pressure pump 46 via an oil strainer 48a. The discharge port of the high-pressure pump 46 has a secondary pressure line, that is, a line pressure line 50, which allows the secondary pulley 16 to move. It is connected to the secondary oil chamber 21 for operating the pulley 16b and to the secondary pressure port of the secondary pressure adjusting valve 51. This secondary pressure regulating valve 51
Thereby, the secondary pressure supplied to the secondary oil chamber 21 is adjusted to a predetermined pressure and controlled to a value corresponding to the transmission capacity required for the drive belt 17. That is, when the engine output is large, such as when climbing a hill or suddenly accelerating, the secondary pressure is raised to prevent the drive belt 17 from slipping. When the engine output is small, the secondary pressure is reduced, and the loss and transmission efficiency of the high-pressure pump 46 are reduced. Is improved. The drain port of the secondary pressure regulating valve 51 is connected to the suction port of the high-pressure pump 46 by a return line 59.

The line pressure passage 50 is connected to the input port of the primary pressure regulating valve 52, and the output port of the primary pressure regulating valve 52 operates the movable pulley 14 b of the primary pulley 14 via the primary pressure passage 53. Connected to the primary oil chamber 18. The primary pressure is adjusted by the primary pressure adjusting valve 52 to a value corresponding to the target gear ratio, the vehicle speed, and the like, and the groove speed of the primary pulley 14 is changed to control the vehicle speed.

The line pressure line 50 is connected to the input port of the clutch pressure regulating valve 57, and the output port is connected to the clutch pressure line 5
8 are connected. Hydraulic oil is supplied to the clutch oil chamber 32a of the forward clutch 34 and the brake oil chamber 45a of the reverse brake 43 of the forward / reverse switching mechanism 11 via the clutch pressure path 58. When the external pilot pressure is supplied to the
The oil pressure in 8 is set to a low pressure, and when the supply of the external pilot pressure is stopped, the pressure is set to a higher pressure than when it was supplied. Further, the clutch pressure hydraulic oil is supplied to the apply chamber 9a of the lock-up clutch 9 of the torque converter 2 via the clutch pressure path 58.

A low-pressure pump 47 driven by an engine
The hydraulic oil in the oil pan 48 is supplied to the oil strainer 4
8b, and the low-pressure pump 47
A lubricating pressure passage 55 is connected to the discharge port of the hydraulic motor. Through this lubricating pressure passage 55, hydraulic oil is supplied to the lubricating portion of the forward / reverse switching mechanism 11, the lubricating portion of the drive belt 17, and the balance chamber of the secondary shaft. It is supplied to the release chamber 9b of the lock-up clutch 9 while being supplied.
The lubricating pressure in the lubricating pressure path 55 is regulated by a lubricating pressure adjusting valve 54 having a relief valve structure, and the hydraulic oil to be relieved is supplied to a low pressure pump 4 by a return path 56 connected to a drain port.
7 is returned to the inlet. As a result, the amount of hydraulic oil sucked through the oil strainer 48a can be reduced, and the torque loss of the pump drive due to the oil flow resistance of the strainer 48b can be reduced.

An application pressure path 61 connected to the application chamber 9a of the torque converter 2, a release pressure path 62 connected to the release chamber 9b, and a switching pressure for brake connected to a brake oil chamber 45a for operating the reverse brake 43 Road 6
3, and a switch valve 65 for controlling connection between the clutch switching pressure passage 64 connected to the clutch oil chamber 32a for operating the forward clutch 34, the lubrication pressure passage 55 and the clutch pressure passage 58, and the like. Is provided.

The switch valve 65 has four portions each having a three-port switching valve structure, and as shown in FIG. 2, the F & R mode when the external pilot pressure is not applied, that is, the vehicle speed becomes lower than a predetermined value. The lock-up clutch 9 operates in two positions: an open position of the lock-up clutch 9 in the state, and an engagement position of the lock-up clutch 9 in a state in which the external pilot pressure is applied.

In the open position, the lubrication pressure path 55 and the release pressure path 62 are in communication with each other by the switch valve 65, and the cooling path 67 provided with the oil cooler 66 and the apply pressure path 61 are in communication. . As a result, the hydraulic circuit operates the torque converter 2 to enter a mode in which the hydraulic control of the forward / reverse switching mechanism 11 can be performed, that is, the F & R mode.
b, is discharged from the apply chamber 9a, and is returned to the oil pan via the oil cooler 66.

On the other hand, at the engagement position, the clutch pressure passage 58 and the apply pressure passage 61 are in communication with each other, and the working oil set to the clutch pressure is supplied to the apply chamber 9a.
At this time, the slip pressure passage 68 connected to the clutch pressure passage 58 is communicated with the release pressure passage 62. The slip pressure control valve 71 is provided in the slip pressure passage 68, and the slip pressure control valve 71 controls the slip pressure supplied to the slip pressure passage 68 in accordance with the external pilot pressure supplied to the external pilot chamber. , The pressure is controlled to an arbitrary pressure in the range from the same pressure as the clutch pressure to zero pressure. Therefore, when the slip pressure becomes 0, the lock-up clutch 9 is engaged to enter the lock-up mode, and when the slip pressure becomes equal to the clutch pressure, the lock-up clutch 9 is released. By appropriately controlling the slip pressure, slip control of the lock-up clutch 9 for controlling the rotation difference of the lock-up clutch 9 to be constant can be performed. When the switch valve 65 is in the engaged position, the lubricating pressure passage 55
Communicates with the cooling passage 67 through the switch valve 65 to cool the hydraulic oil.

In order to supply the external pilot pressure to the slip pressure regulating valve 71, a pilot pressure passage 7 is provided between the pilot port of the slip pressure regulating valve 71 and the clutch pressure passage 58.
The pilot pressure path 72 is provided with a pilot pressure adjusting valve 73 for controlling the pilot pressure.

A manual valve 75 and a reverse signal valve 76 which are respectively linked to the control lever, that is, a select lever 74 for switching the driving mode provided in the vehicle cabin are connected to each other.
Reference numeral 76 operates at five positions corresponding to the P (parking) range, R (reverse) range, N (neutral) range, D (drive) range, and Ds (sports drive) range set by the select lever 74.

A pilot pressure passage 77 for connecting the clutch pressure passage 58 to an external pilot chamber of the switch valve 65 through a reverse signal valve 76 is provided with a three-port solenoid type switching valve 78. When the solenoid 78a of the switching valve 78 is energized, the switch valve 65 becomes the lock-up control position, that is, the engagement position of the lock-up clutch, and when the energization to the solenoid 78a is turned off, FIG.
The F & R mode position is set as shown in FIG. The pilot pressure passage 77 is connected to an external pilot chamber of the clutch pressure adjusting valve 57 as shown by a broken line, and is connected to the reverse signal valve 7.
When 6 is set to any of the N position, the D position and the Ds position, the clutch pressure is supplied to the external pilot chamber of the clutch pressure adjusting valve 57, and the clutch pressure is set to a low pressure. On the other hand, when the reverse signal valve 76 is set to the P position and the R position other than those described above, the hydraulic pressure is not supplied to the external pilot chamber of the clutch pressure adjusting valve 57, and the clutch pressure becomes higher than the above. Is set.

A common switching pressure passage 79 is provided between the switch valve 65 and the manual valve 75. The switching pressure passage 79 communicates with the slip pressure passage 68 when the switch valve 65 is in the F & R mode position. When the switch valve 65 reaches the lock-up control position, it communicates with the clutch pressure passage 58. When the manual valve 75 is set to one of the D range and the Ds range by operating the select lever 74, the switching pressure path 79 is in communication with the clutch switching pressure path 64 via the manual valve 75, and is switched to the R range. When it is set, it is in a state of communication with the brake switching pressure path 63.

The two oil pans 4 shown in FIG.
8 can be separate or shared. When the idling stop is not performed, one strainer can be shared without using both strainers 48a and 48b, and the return paths 56 and 59 can be returned to a common strainer, thereby reducing the number of parts. Can be. However, when performing idle stop, the low-pressure pump 47 driven by the engine is stopped, so that the negative pressure generated by the high-pressure pump 46 driven by the motor 49 causes the clearance between the low-pressure pump 47 and the lubrication pressure adjusting valve 54 to be increased. If a common strainer is used, it is necessary to manage the clearance of these components because air may be sucked from the air.

FIG. 3 is a block diagram showing a motor control circuit. A target line pressure is calculated by an engine torque estimating section 81, a torque converter output torque estimating section 82, a necessary line pressure calculating section 83 and a correcting section 84, and the target line pressure is calculated. The signal of the line pressure value is sent to the valve driver 85, and the operation of the secondary pressure regulating valve 51 is controlled.

The rotation speed of the high-pressure pump 46 driven by the electric motor 49 is controlled in several operation modes in order to avoid electrical loss and inertia loss due to frequent rotation speed control. The operation mode is the operation mode setting section 86
The operation mode setting section 86 includes a throttle opening signal θ corresponding to the opening of a throttle valve that operates according to an accelerator pedal, and a range set by operating the select lever 74. , A secondary rotation speed signal Ns, and a brake switch signal SWb for detecting the operation of the brake. Target setting section 87 to which a signal from operation mode setting section 86 is sent.
Then, the rotation speed of the motor 49 and the generated maximum torque are set according to the operation mode.

The target setting section 87 receives a signal from a pump driving torque estimating section 88 for estimating a pump driving torque value based on the target line pressure signal from the correcting section 84 and the hydraulic oil temperature signal Ttm. By monitoring whether or not the torque generated by the motor 49 satisfies the required value of the pump 46 by this signal, for example, a line pressure exceeding the operation mode assumption due to some failure is generated. In such a case, the motor 49
Will be corrected. The pump drive torque estimating unit 88 also performs control such as reducing the motor starting voltage to reduce the load on the electric system at the time of restart after warming up.

The motor driver 89 drives the motor 49 in accordance with the rotation speed and the generated maximum torque signal set by the target setting section 87. The circuit of the motor driver 89 differs depending on the type of the motor 49 used.

In order to avoid a belt slip or the like due to a malfunction of the motor 49, a target line pressure and a secondary pressure signal Ps are sent to the abnormality detecting section 90. If the actual line pressure shortage exceeds the set value, the abnormality detection unit 90
Outputs an abnormal signal to the transmission control unit TCU. Further, the abnormality detection unit 90 outputs a signal from the motor driver 89 based on signals of the drive voltage and the current value of the motor 49 (the drive frequency in the case of an AC drive motor).
The operation state of the motor 49 is monitored, and abnormality is detected according to the situation. The TCU informs the driver of the abnormality and can prevent the drive system from being damaged by forcibly disengaging the clutch.

FIG. 4 is a flowchart showing an algorithm for setting the operation mode in the operation mode setting section 86.

Starting mode (M1): If it is detected in step S1 that the ignition has changed from off to on, the starting mode (M1) is set in step S2, and a voltage sufficient for starting is temporarily applied to the motor. 49.

Stall mode (M2): In step S3, it is determined that the secondary rotation speed is equal to or less than the set value, and step S4
If it is determined that the throttle opening is equal to or greater than the set value in step S5, the stall mode (M2) is set in step S5, and the high-pressure pump 46 is controlled to a state in which a low discharge amount but high pressure can be generated.

Idle mode (M3): If it is determined in step S6 that the motor is in the range other than the N range and the P range, that is, the D range or the R range, the mode is set to the idle mode (M3) in step S7, and the motor 49 It is controlled to the set minimum required number of revolutions and torque.

Idle stop mode (M4): N in step S6
If it is determined to be in the range or P range,
In step S8, the idling stop mode (M4) is set, and the number of revolutions is set higher than in the idling mode by the operating flow rate when the range is selected from the N range to the D range or the R range.

Sudden braking mode (M5): When the deceleration of the secondary rotation (obtained by differentiating the secondary rotation speed signal) is equal to or larger than the set value in step S9 (when the brake switch signal SWb is on as a backup of the rotation sensor). Is set to the sudden braking mode (M5) in step S10,
The line pressure and the number of revolutions that can cope with the sudden downshift of the pulley ratio are maintained.

Normal driving mode (M6): If it is determined in step S9 that the deceleration is equal to or less than the set value, the flow proceeds to step S1.
In step 1, the normal running mode (M6) is set, and the motor 49 is controlled so that the number of rotations required for a normal pulley ratio change and the line pressure required for the maximum torque of the engine can be generated.

FIG. 5 is an operation characteristic diagram schematically showing the relationship between the maximum possible line pressure and the motor rotation speed in each of the above-described operation modes (M1) to (M6).
9 are shown, and the idle mode and the idle stop mode in which the engine speed is low are set so that the electric load is minimized.

FIG. 6 shows a motor 4 for driving the high-pressure pump 46.
Reference numeral 9 denotes a motor control circuit in the case of a DC motor. The DC motor is generally preferable in the present invention because it has high torque generation efficiency and easy speed control. It is necessary to pay attention to the characteristics, especially the installation environment of the motor.

As shown in FIG. 6, the voltage of the battery 91 is temporarily converted to a high voltage V1 by the variable booster 92 in order to supply a current required for the rated output of the motor 49.
The target voltage of the variable booster 92 is set by the target voltage signal from the target setting section 87 shown in FIG. 3, and the voltage V1 is supplied to the shunting section 93 of the DC motor to generate the field magnetic flux φ. On the other hand, the armature voltage V is adjusted by PWM (pulse width modulation) control of the switching element 94, and the target value duty ratio of PWM is set by the target setting unit 87 in FIG. Reference numeral 95 indicates a freewheel diode.

The characteristics of a shunt DC motor are generally expressed by the following equation.

[0059] generated torque _{T = K 1 Ia φ ... (} 1) rotational speed _{n = K 2 {V- (Ia} r + e b)} / φ ... (2) where, K _1, K ₂ are proportionality constants, la is Armature current, r
The coil resistance of the armature, e _b is the voltage drop due to the brush.

FIG. 7 is a graph showing the characteristics of the equations (1) and (2). At a constant PWM rate D, a change in the load torque appears in the armature current Ia. As the load torque increases, the armature current Ia increases, and the rotational speed n decreases according to equation (2). Therefore, as shown by the lines D1 to D3 in FIG. 7, the rotation speed decreases with respect to the load torque. However, when the supply voltage V1 is increased, both the rotation speed Ia and the magnetic flux φ increase. The load torque can be increased without changing.

According to the above characteristics, the target pump output can be obtained by changing the voltage V1 according to the generated oil pressure and performing PWM control on the target rotation speed. Here, the control of the voltage V1 can be omitted in the case of a small motor. However, when a high-output motor is used, it is preferable to control the switching element to a required value for the purpose of reducing the durability and noise generation of the switching element.

FIG. 8 is a circuit diagram showing a motor control circuit in the case where an induction motor is used as the motor 49. The induction motor has a simple structure and high reliability, but is larger than a DC motor and requires AC. Therefore, an AC generator, that is, an inverter, is required for driving with a battery.

Synchronous speed n = 120 f / P (3) Generated torque T∝ (V / f) ² (4) where f is the drive frequency, P is the number of motor poles, and V is the drive voltage.

As shown in FIG. 8, the motor 49 is controlled by an inverter 96 that converts DC power of the battery 91 into AC. Although FIG. 7 uses a three-phase motor, a two-phase motor may be used. The inverter 96 includes a target setting unit 87
Gives the drive frequency f and the target value of the voltage V. According to the characteristic formula (4) of the induction motor, the drive voltage can be reduced for the same torque load at the time of low rotation.

FIGS. 9A and 9B are diagrams showing characteristics of the induction motor. As shown in these figures, the target of the voltage V is set so that the operation efficiency of the induction motor is high. Determine the value. Specifically, as shown in FIG. 9A, the motor rotates with a slip relative to the synchronous rotation speed. The slip ratio is determined by the ratio between the load torque and the rated torque determined by the equation (4). In a normal range, as the load increases, the slip ratio increases. On the other hand, the efficiency increases as the torque approaches the rated torque as shown in FIG. Therefore, it is desirable from the viewpoint of efficiency that the drive voltage V is controlled to be a necessary value with respect to the pump drive torque. However, in order to simplify the control, a constant voltage is set for each mode shown in FIG. It is also possible to control so that

It is also possible to use a synchronous motor as the motor 49, and the motor control circuit when the synchronous motor is used is the same as that for the induction motor shown in FIG. As a type of the synchronous motor, a permanent magnet type synchronous motor having a permanent magnet built in the rotor has no commutator, which is advantageous in terms of reliability. Both the rotation characteristics and the torque characteristics are expressed by the same equations (3) and (4) as the induction motor.

FIG. 10 is a diagram showing the operating characteristics of the synchronous motor. The difference between the characteristics is that, as shown in FIG. In this respect, highly stable rotation control can be performed on the induction motor with slippage. If the load torque increases and exceeds the escape torque, synchronization is lost and the rotation speed cannot be maintained. The characteristics of load torque and efficiency are also shown in FIG.
As shown in (B), it is the same as the induction motor, and it is desirable to adjust the drive voltage so as to operate at an appropriate load factor.

FIG. 11 is a comparison diagram showing the relationship between the discharge pressure and the discharge flow rate of the oil pump in the case where the oil pump is driven only by the engine, the case where the oil pump is driven only by the electric motor, and the case of the present invention. The hatched thin-line frames indicate the respective operation modes of the present invention shown in FIG. 4, the thick-line frames indicate the case of driving only by the engine, and the broken-line frames indicate the case of driving only by the motor. Incidentally, the power consumed is shown by a constant power line H ₁ to H ₅ by the pump drive.

Here, the driving force of the pump is calculated by the following equation.

[0070] When driving the pump by the engine or _{motor, L = (Q 1 + Q} 2 + Q 3) × P S / 612 ... (5) when used in conjunction with a motor drive and an engine _{drive, L = (Q 1 + Q} 2 ) × P _S / 612 + Q ₃ × P _Lub / 612 (6) In these equations, Q ₁ is a high-pressure (pulley) system supply flow, Q ₂ is a medium-pressure (clutch) system supply flow, and Q ₃ is a low-pressure ( Lubrication) system supply flow rate, _PS is line pressure, P
_Lub is the lubrication pressure.

(Normal running) The steady operating point in the normal running mode of the present invention is the same as that shown in FIG.
It becomes the position shown by, its power is represented by H _2.

The required oil amount is the same in any of the driving methods. However, in the case of the engine driving, the discharge flow is determined by the engine speed, so that the oil is discharged unnecessarily. It will operate in the region shown, causing excessive power loss. The power line H ₂ occupies only a part of the low-speed rotation region in the engine driving region,
In most normal driving states, power exceeding the power consumption of the present invention is consumed for driving the oil pump.

On the other hand, the power consumption when the pump is driven only by the motor is indicated by H _2.
Power consumption than by the present invention correspondingly to generate a low-voltage supply flow rate Q ₃ at the line pressure P _S as shown by 6m becomes high.

(Stall State) In the stalled state, a high line pressure is required. Therefore, when the engine is driven, a large power loss is generated as indicated by reference numeral M2e. On the other hand, in the present invention, since it is sufficient to generate the minimum necessary high-pressure flow rate, it is possible to suppress the power loss indicated by reference numeral M2.

On the other hand, when driving only with a motor,
(5) controls the rotational speed of the motor according to type, but rather than the case of the engine driven by generating the minimum required flow rate can be reduced power consumption, low-voltage supply flow rate Q ₃ for lubrication It is generated at an unnecessarily high line pressure P _S , and as shown by reference numeral M2m, generates more power loss than in the present invention.

(At the time of sudden braking) At the time of sudden braking, a large flow rate is required in order to quickly inject hydraulic oil into the cylinder for operating the secondary pulley and to shift to a low gear. Normally, this mode is most short about one second generation time, there is no practical problem even if insufficient low-voltage supply flow rate Q _3. Therefore, in any of the driving methods, the same power consumption can be obtained if the flow rates Q ₁ and Q ₂ are set so as to be minimum. In FIG. 11, the case of driving the engine is indicated by reference numeral M5e, the case of driving only the motor is indicated by reference numeral M5m, and the case of the present invention is indicated by reference numeral M5.

(Idling state) When the vehicle is in the idling state in the traveling range, the operating flow rate of the piston is not generated due to the operation from the N range to the D range and from the N range to the R range. Become. Even in this mode, in order to generate a flow rate of Q ₃ system, if the engine drive (M3e) and the case of driving the motor only (M3M), a number of power consumption than the present invention at M3 generation I do. As a result, the present invention is also advantageous for improving fuel efficiency in a driving state including a signal waiting mode such as the 10 mode or the 15 mode.

(Idle Stop State) In the idle stop state, it is necessary to secure a flow rate necessary for starting immediately when the engine is restarted. An engine-driven pump cannot support this mode. In the case of driving the motor only (M4M), the power consumption for generating the actuation flow rate Q ₃ of the torque converter extra needed. In contrast, it is sufficient to ensure the operation flow Q ₂ as shown in M4 In the present invention, power consumption can be reduced, less power consumption during idle stop, thereby maintaining the idle stop state longer Can be.

According to the present invention, the flow rate of the hydraulic oil in the torque converter system is generated by the low pressure pump driven by the engine, and the supply of the hydraulic oil is started when the rotation of the engine is returned.

[0080] On the other hand, when driving the pump only in motor, the operation flow of the torque converter as a method other than the constantly generating as described above, constantly Q ₂
It is conceivable that the rotation speed is rapidly increased in synchronism with the cranking of the rotation return of the engine.However, in actuality, the power consumption of the starter motor increases during cranking. It is difficult to secure enough power to jump the numbers.

FIG. 12 shows another embodiment C of the present invention.
FIG. 3 is a circuit diagram showing a VT hydraulic circuit, in which members common to those of the hydraulic circuit shown in FIG. 2 are denoted by the same reference numerals, and description thereof is omitted.

In this case, a pressure accumulator 100 is provided in the clutch pressure passage 58, and the pressure is stored in the N range from the D range.
When the range is switched to the range, the capacity of the hydraulic oil required for operating the clutch and the brake of the forward / reverse switching mechanism is secured. The forward / reverse switching mechanism is required to operate in a short time in order to prevent a delay in starting the vehicle, and the operating volume of the piston is small, but a large amount of instantaneous flow is required. By ensuring the working volume of the piston with the accumulator 100, the pump discharge amount during idling can be reduced.

FIG. 13A is a sectional view showing a specific example of the accumulator 100. A piston 103 is slidably mounted in an axial direction in a housing 102 having a pressure accumulating chamber 101 communicating with a clutch pressure passage 58. And the piston 10
3, the spring force of the compression coil spring 104 is urged. A drain port 105 is formed in the housing 102 so as to communicate with the spring chamber, and the spring force of the coil spring 104 is set so as to balance the clutch pressure. That is,
Assuming that the clutch pressure is P, the pressure receiving area of the piston 103 is A, and the spring constant is k, the spring constant k is set so that P · A = F ₀ + kx.

Generally, the reverse brake pressure is set higher than the forward clutch pressure to reduce the number of brake friction materials and reduce the friction loss caused by the reverse brake during forward travel. Even in the case shown in FIG. 2, the clutch pressure is set to be high when the vehicle is reversing by the action of the reverse signal pressure. Accumulator 100 shown in FIG.
Since there is only one balance point, for example, if it is set optimally at the forward clutch pressure, which is frequently used, there is no pressure accumulating effect at the high reverse brake pressure.

FIG. 13B shows an accumulator 100a corresponding to a change in the clutch pressure depending on the range. The accumulator 101a communicates with the clutch pressure passage 58, and a pilot pressure passage for guiding the pilot pressure, that is, the reverse signal pressure. Two pressure accumulating chambers 101b communicating with 77 are provided in the housing 102, and a reverse signal pressure from the pilot pressure passage 77 is applied to the accumulating chamber 101b in a range other than the P range and the R range. Therefore, the accumulator 100a shown in FIG.
The balance pressure of the accumulator can be lower in the D range and higher in the R range than in the accumulator 100 shown in FIG.

FIG. 14 is a circuit diagram showing a hydraulic circuit of a CVT according to still another embodiment of the present invention. Components common to those of the hydraulic circuit shown in FIG. Is omitted.

In this case, the motor-driven oil pump 4
Communication path 110 between the return path 59 to the lubrication line 6 and the lubrication pressure path 55
And a check valve 111 is provided in the communication passage 110 to allow the flow of the hydraulic oil in a direction in which the lubricating pressure flows into the recirculation passage 59 according to the magnitude relationship between the line pressure and the lubricating pressure.
A one-way valve 112 that allows only the flow from the oil strainer 48a to the high-pressure pump 46 is provided between the return line 59 and the oil strainer 48a.

FIG. 15 is a cross-sectional view showing the check valve 111. The check valve 111 has a valve body 11b on which a poppet valve body 111a is slidably mounted. The poppet valve body 111a is connected via a communication passage 110. The communication between the input port communicating with the lubrication pressure path 55 and the output port communicating with the recirculation path 59 is opened and closed. A spring force in the closing direction is applied to the poppet valve body 111a by the coil spring 111c, and is applied in the closing direction by the oil pressure of the pressurizing chamber to which the line pressure is supplied, and the lubricating pressure is a direction in which the poppet valve body 111a opens. Act on.

Assuming that the area of the input port is A ₁ , the pressure receiving area of the poppet valve body 111 by the pressurizing chamber is A ₂ , the line pressure is P ₅₀ , the lubricating pressure is P ₅₅ , and the spring force is F _SP , The conditions for balancing the opening and closing forces acting on the poppet valve body 111a are as follows.

P ₅₀ · A ₂ + F _SP = P ₅₅ · A ₁ Therefore, when the value on the right side increases, the check valve 11
1 is open, and when the left side becomes large, the check valve will close.

Normally, since the line pressure is higher than the lubrication pressure, the check valve 111 is closed. However, if the electric motor 49 cannot be started at a low temperature or when the dust is caught, no line pressure is generated. Pressure check valve 11
1 is opened and flows into the return path 59. At this time, check valve 1
Since the valve 12 is closed, lubricating pressure acts on the recirculation path 59, and a reverse pressure difference is generated in the oil pump 46 to generate torque as a hydraulic motor. This torque is the electric motor 4
9, the reliability of the start-up failure of the electric motor 49 is enhanced.

The failure in starting the motor 49 is detected by the abnormality detection unit 90 shown in FIG. 3. For example, in the case of a failure in starting the pump in a cold start, the difference between the target value of the line pressure and the actually measured value is determined. Is generated and the starting current of the motor becomes excessively large, and processing according to the business situation in which the occurrence occurs is performed. If the motor 49 does not start even by cranking, a warning is issued and the drive system is controlled to be neutral.
Further, the power supply to the motor 49 can be cut off to prevent burnout due to continuous power supply to the motor.

The present invention is not limited to the above-described embodiment, and it goes without saying that various modifications can be made without departing from the scope of the invention. For example, the drive system of the belt-type continuously variable transmission is not limited to the case shown in FIG. 1, and the present invention can be applied to various types such as a type having no torque converter 2.

[0094]

According to the present invention, the hydraulic pressure for controlling the operation of the belt-type continuously variable transmission is generated by the low-pressure pump driven by the engine and the high-pressure pump driven by the electric motor. By generating the hydraulic oil on the side by an electric motor, pump loss can be reduced.

The pump power can be reduced in any of the normal running state, the stall state, and the rapid braking mode. Fuel efficiency can be improved by reducing power loss.

The flow rate of the pump in the idling state can be reduced, and the power loss can be reduced. When the engine is restarted from the idling stop state, the flow rate of the hydraulic oil required for starting can be immediately secured.

[Brief description of the drawings]

FIG. 1 is a schematic diagram illustrating an example of a drive system of a belt-type continuously variable transmission.

FIG. 2 is a circuit diagram showing a hydraulic circuit of the continuously variable transmission according to one embodiment of the present invention.

FIG. 3 is a block diagram illustrating a motor control circuit.

FIG. 4 is a flowchart showing an operation mode of the high-pressure pump.

FIG. 5 is a characteristic diagram showing a relationship between a motor rotation speed and a pump discharge pressure in an operation mode of the high-pressure pump.

FIG. 6 is a circuit diagram showing a control circuit of the DC motor.

FIG. 7 is a diagram showing operating characteristics of a DC motor.

FIG. 8 is a circuit diagram showing a control circuit of an induction motor or a synchronous motor.

FIGS. 9A and 9B are diagrams showing operating characteristics of an induction motor.

FIGS. 10A and 10B are diagrams showing operating characteristics of a synchronous motor.

FIG. 11 is a comparison diagram showing a relationship between a discharge pressure and a discharge flow rate of the oil pump in a case where the oil pump is driven only by the engine, a case where the oil pump is driven only by the motor, and a case of the present invention.

FIG. 12 is a circuit diagram showing a hydraulic circuit of a continuously variable transmission according to another embodiment of the present invention.

FIGS. 13A and 13B are cross-sectional views each showing an accumulator.

FIG. 14 is a circuit diagram showing a hydraulic circuit of a continuously variable transmission according to still another embodiment of the present invention.

FIG. 15 is a sectional view showing the check valve shown in FIG. 14;

[Explanation of symbols]

 Reference Signs List 10 case 11 forward / reverse switching mechanism 12 continuously variable transmission 14 primary pulley 16 secondary pulley 17 belt 18 primary oil chamber 21 secondary oil chamber 32 clutch cylinder 32a clutch oil chamber 34 forward clutch 45 brake cylinder 46 high-pressure pump 47 low-pressure pump 49 motor Reference Signs List 50 line pressure path 51 secondary pressure regulating valve 52 primary pressure regulating valve 54 lubricating pressure regulating valve 55 lubricating pressure path 100 accumulator 111 one-way valve (check valve)

 ──────────────────────────────────────────────────続 き Continued on the front page F term (reference) 3J050 AA03 BA03 BB12 CB01 CD05 DA02 3J052 AA11 CA21 FB02 FB22 FB25 FB29 GC23 GC34 GC42 GC43 GC44 GC64 GC72 GC73 HA02 HA11 HA18 KA01 LA01 3J063 AA02 AB23 AC04 BB50 CC12 A22 AC23 BA04 CA02 CA03 CA08 CA32 DA52 DB08 DB18 DB32 EA71 FB90 GA01

Claims (7)

[Claims] 1. An automobile having a variable pulley groove width primary pulley mounted on a primary shaft and a variable pulley groove width variable pulley mounted on a secondary shaft and having a belt stretched between the primary pulleys. Belt type continuously variable transmission, comprising: a low-pressure pump driven by an engine; a high-pressure pump driven by an electric motor; and motor control means for controlling the electric motor in accordance with a running state of a vehicle. A belt type continuously variable transmission for automobiles having an electric oil pump. 2. A primary pulley having a variable pulley groove width mounted on a primary shaft, a secondary pulley having a pulley groove width variable mounted on a secondary shaft and having a belt stretched between the primary pulley and an engine. An automotive belt-type continuously variable transmission having a forward / reverse switching mechanism for transmitting rotation to the primary shaft by switching between forward rotation and reverse rotation, and is driven by the engine to supply hydraulic oil to a lubrication unit. A low-pressure pump, a primary oil chamber driven by an electric motor to change the groove width of the primary pulley, a secondary oil chamber to change the groove width of the secondary pulley, and a high-pressure pump for supplying hydraulic oil to the forward / reverse switching mechanism And a motor control means for controlling the electric motor according to the running state of the vehicle. A belt type continuously variable transmission for automobiles having an electric oil pump. 3. A primary pulley having a variable pulley groove width mounted on a primary shaft, a secondary pulley having a variable pulley groove width mounted on a secondary shaft and having a belt stretched between said primary pulleys, A belt-type continuously variable transmission for an automobile having a forward / reverse switching mechanism that switches and transmits rotation to the primary shaft in a forward rotation direction and a reverse rotation direction, and a torque converter with a lock-up clutch, which is driven by the engine, A low-pressure pump for supplying hydraulic oil to a release chamber and a lubrication unit of a lock-up clutch; a primary oil chamber driven by an electric motor to change a groove width of the primary pulley; and a secondary oil chamber to change a groove width of the secondary pulley. The lock-up clutch apply chamber and the forward / reverse switching device Automotive belt-type continuously variable transmission having an electric oil pump, characterized in that it comprises a high pressure pump, and a motor control means for controlling said electric motor in accordance with the running state of the vehicle to supply hydraulic fluid to. 4. The motor-driven belt-type continuously variable transmission having the electric oil pump according to claim 1, wherein the electric motor is a shunt-type DC motor, and A belt-type continuously variable transmission for an automobile having an electric oil pump, wherein the means controls a motor driving voltage and an armature voltage. 5. An automobile belt-type continuously variable transmission having an electric oil pump according to claim 1, wherein the electric motor is an induction motor or a permanent magnet type synchronous motor, A motor-driven belt-type continuously variable transmission having an electric oil pump, wherein the motor control means controls a drive voltage and a frequency of a motor. 6. A belt type continuously variable transmission for an automobile having the electric oil pump according to claim 2, wherein a clutch pressure is supplied to a clutch pressure path for supplying hydraulic oil to the forward / reverse switching mechanism. An automotive belt-type continuously variable transmission having an electric oil pump, comprising an accumulator for accumulating pressure. 7. An automobile belt-type continuously variable transmission having the electric oil pump according to claim 2, wherein a discharge port of the low-pressure pump communicates with a suction port of the high-pressure pump. A belt type continuously variable transmission for an automobile having an electric oil pump, characterized in that a check valve is provided in the communication passage for allowing the discharge pressure of the low pressure pump to flow into the suction port of the high pressure pump depending on the discharge pressure of the high pressure pump.