JP6897175B2 - Continuously variable transmission - Google Patents

Description

The present invention relates to a continuously variable transmission.

Conventionally, a hydrostatic continuously variable transmission (HST) used for traveling a vehicle has been known (see, for example, Patent Document 1). The HST described in Patent Document 1 is a closed circuit in which a hydraulic pump driven by an engine and a hydraulic motor driven by the hydraulic pump generated by the hydraulic pump are connected by a closed circuit.

JP-A-2015-52321

However, the HST described in Patent Document 1 has a problem that a large-capacity hydraulic pump and a hydraulic motor are required to drive the vehicle at high speed.

An object of the present invention is to provide a continuously variable transmission capable of driving a vehicle at high speed without using a large-capacity hydraulic pump and a hydraulic motor.

The stepless transmission according to the present invention comprises a variable capacitance type first hydraulic pump motor that is rotationally driven by a drive source, the first hydraulic pump motor, and a closed circuit, and is delivered from the first hydraulic pump motor. A variable-capacity type hydraulic motor that is rotationally driven by the pressure oil, is connected to an oil passage that connects the first hydraulic pump motor and the hydraulic motor, and is connected to the hydraulic motor. and second hydraulic pump motor, provided in a portion between the second hydraulic pump motor and the hydraulic motor in the oil passage, Bei give a, and a switching valve capable of switching the direction of the pressure oil flow, the first 2 The maximum capacity of the hydraulic pump motor is equal to the capacity of the hydraulic motor.

According to the continuously variable transmission according to the present invention, the vehicle can be driven at high speed without using a large-capacity hydraulic pump and a hydraulic motor.

A skeleton diagram showing an overall configuration of a vehicle equipped with a continuously variable transmission according to the present invention. Schematic diagram of a cylinder in a variable displacement pump motor The figure which shows the process of sucking hydraulic oil from a low pressure side and discharging it to a high pressure side. The figure which shows the process of sucking hydraulic oil from a low pressure side and discharging it to a high pressure side. The figure which shows the process of sucking hydraulic oil from a low pressure side and discharging it to a high pressure side. The figure which shows the process of sucking hydraulic oil from a low pressure side and discharging it to a high pressure side. The figure which shows the process of sucking hydraulic oil from a low pressure side and discharging it to a high pressure side. The figure which shows the process of sucking hydraulic oil from a low pressure side and discharging it to a high pressure side. The figure which shows the process which the hydraulic oil flows in from a high pressure side and flows out to a low pressure side. The figure which shows the process which the hydraulic oil flows in from a high pressure side and flows out to a low pressure side. The figure which shows the process which the hydraulic oil flows in from a high pressure side and flows out to a low pressure side. The figure which shows the process which the hydraulic oil flows in from a high pressure side and flows out to a low pressure side. The figure which shows the process which the hydraulic oil flows in from a high pressure side and flows out to a low pressure side. The figure which shows the process which the hydraulic oil flows in from a high pressure side and flows out to a low pressure side. The figure which shows the process which the hydraulic oil flows in from a high pressure side and flows out to a low pressure side. Skeleton diagram showing low torque start Skeleton diagram showing high torque start Skeleton diagram showing high speed driving Skeleton diagram showing deceleration state at high speed Skeleton diagram showing deceleration state at low speed Skeleton diagram showing backward movement

Hereinafter, embodiments of the present invention will be described in detail with reference to the drawings. The embodiment described below is an example, and the present invention is not limited to this embodiment.

First, the overall configuration of the vehicle 1 according to the present embodiment will be described with reference to FIG.

The vehicle 1 includes a drive source 10, a damper 20, and a continuously variable transmission 30. The drive wheels are connected to the output side of the continuously variable transmission 30 via a propeller shaft, a differential, and a drive shaft (not shown) so that power can be transmitted.

The drive source 10 is, for example, a diesel engine. The drive source 10 may be a gasoline engine, an electric motor, or the like. In this embodiment, the drive source 10 will be described as a diesel engine. The input side member 21 of the damper 20 is connected to the output shaft 11 of the drive source 10.

The damper 20 includes an input side member 21, an output side member 22, and an elastic connecting member 23 that elastically connects the input side member 21 and the output side member 22. The output side member 22 of the damper 20 is connected to the input shaft 30a of the continuously variable transmission 30. The damper 20 has a function of suppressing the rotational vibration generated by the drive source 10 from being transmitted to the continuously variable transmission 30. Since the structure of the damper 20 is the same as that of a general damper, detailed description thereof will be omitted.

The continuously variable transmission 30 comprises a closed circuit in which the first pump motor 31, the second pump motor 32, the third pump motor 33, and the switching valve 34 are connected to each other by an oil passage. In FIG. 1, some components (charge pump, charge oil passage, relief oil passage, bypass oil passage, etc.) in the closed circuit are omitted.

The first pump motor 31 has a high-pressure side port 31a that opens in the high-pressure side oil passage 35 and a low-pressure side port 31b that opens in the low-pressure side oil passage 36. The first pump motor 31 is a variable capacity type pump motor capable of changing the capacity from zero capacity to the maximum capacity Q1. The first pump motor 31 corresponds to the "first hydraulic pump motor" of the present invention.

Here, the configuration of the first pump motor 31 will be described in detail with reference to FIG. FIG. 2 depicts one cylinder 40 in the first pump motor 31. The cylinder 40 corresponds to the "hydraulic oil chamber" of the present invention.

The cylinder 40 is connected to the low pressure side oil passage 36 via the low pressure side check valve 41 and is connected to the high pressure side oil passage 35 via the high pressure side check valve 42. The cylinder 40 is provided with a piston 43.

The vicinity of the lower end of the piston 43 is connected to the crankshaft 45 via a link mechanism 44. The link mechanism 44 converts the rotational movement of the crankshaft 45 into the vertical movement of the piston 43. On the contrary, the link mechanism 44 converts the vertical movement of the piston 43 into the rotational movement of the crankshaft 45.

The crankshaft 45 is connected to the input shaft 30a and is rotationally driven by the drive source 10. Then, the piston 43 moves up and down in the cylinder 40 based on the rotational movement of the crankshaft 45, so that the volume in the cylinder 40 changes periodically.

The low-pressure check valve 41 includes a first valve seat 41a, a first valve body 41b, and a first spring 41c that presses the first valve body 41b in a direction away from the first valve seat 41a. Has. The low pressure side check valve 41 is, for example, a poppet valve. The low-pressure check valve 41 corresponds to the "first check valve" of the present invention.

Further, the low-pressure side check valve 41 is configured to be able to close the valve at a desired timing. Specifically, the low-pressure check valve 41 is provided with a first solenoid 41d that seats the first valve body 41b on the first valve seat 41a by a control signal from the control device 50. The first solenoid 41d corresponds to the "first actuator" of the present invention.

The first solenoid 41d seats the first valve body 41b on the first valve seat 41a only while the power is being supplied from the control device 50. The rotation angle of the crankshaft 45 detected by the crank angle detection sensor 45a is input to the control device 50. The control device 50 controls the first solenoid 41d based on the input rotation angle of the crankshaft 45 to seat the first valve body 41b on the first valve seat 41a.

The high-pressure check valve 42 is a second spring 42c that urges the second valve seat 42a, the second valve body 42b, and the second valve body 42b in the direction of being seated on the second valve seat 42a. And have. The high-pressure check valve 42 is, for example, a poppet valve. The high-pressure side check valve 42 corresponds to the "second check valve" of the present invention.

Further, the high-pressure side check valve 42 opens at a predetermined timing. Further, the high-pressure side check valve 42 is configured to be able to maintain the valve open state. Specifically, the high-pressure check valve 42 is provided with a second solenoid 42d that holds the separated state of the second valve body 42b separated from the second valve seat 42a. The second solenoid 42d is controlled by a control signal from the control device 50. The second solenoid 42d corresponds to the "second actuator" of the present invention.

The second solenoid 42d separates the second valve body 42b from the second valve seat 42a only while the power is being supplied from the control device 50. The control device 50 controls the second solenoid 42d based on the rotation angle of the crankshaft 45 to separate the second valve body 42b from the second valve seat 42a.

Next, with reference to FIGS. 3A to 3F, a process of sucking the hydraulic oil from the low pressure side oil passage 36 into the cylinder 40 and discharging the hydraulic oil to the high pressure side oil passage 35 will be described.

In the suction stroke (FIGS. 3A to 3C) in which the piston 43 descends from the top dead center to the bottom dead center as the crankshaft 45 rotates, power is not supplied from the control device 50 to the first solenoid 41d. The first valve body 41b is in a state of being separated from the first valve seat 41a. In this way, in the suction stroke, the low-pressure side check valve 41 is opened, and the hydraulic oil in the low-pressure side oil passage 36 is introduced into the cylinder 40. Further, in the suction stroke, the check valve 42 on the high pressure side is closed.

The control device 50 does not supply power to the first solenoid 41d during a predetermined period in the discharge stroke in which the piston 43 rises from bottom dead center to top dead center. As a result, the first valve body 41b is maintained in a state of being separated from the first valve seat 41a.

When the first valve body 41b is separated from the first valve seat 41a and the low pressure side check valve 41 is opened, the high pressure side check valve 42 maintains the closed state as shown in FIG. 3D. , The hydraulic oil in the cylinder 40 does not flow into the high pressure side oil passage 35.

When electric power is supplied to the first solenoid 41d in the middle of the discharge stroke, the low-pressure side check valve 41 closes, the pressure in the cylinder 40 rises, and the high-pressure side check valve 42 opens. Then, the hydraulic oil in the cylinder 40 is led out to the high pressure side oil passage 35 (FIGS. 3E to 3F). At this time, the thrust of the second solenoid 42d is not always necessary to maintain the valve open state of the high-pressure side check valve 42, but the thrust of the second solenoid 42d causes the second valve body 42b to have a thrust. The load of the acting second spring 42c may be canceled.

By adjusting the timing of closing the low-pressure check valve 41 in this way, the capacity of the first pump motor 31 can be adjusted from zero capacity to maximum capacity Q1.

Next, with reference to FIGS. 4A to 4G, a process in which hydraulic oil flows into the cylinder 40 from the high pressure side oil passage 35 and flows out to the low pressure side oil passage 36 will be described.

FIG. 4A shows the state in which the piston 43 is rising and immediately before reaching top dead center. At this time, no power is supplied from the control device 50 to the first solenoid 41d, and the low-voltage side check valve 41 is opened by the first spring 41c.

In this state, the control device 50 supplies electric power to the first solenoid 41d. As a result, the low-pressure side check valve 41 is closed, and in the subsequent ascending stroke of the piston 43, the high-pressure side check valve 42 is opened, as shown in FIG. 4B. Further, the control device 50 also supplies electric power to the second solenoid 42d immediately after the start of supplying electric power to the first solenoid 41d.

As a result, when the piston 43 reaches top dead center, the high-pressure check valve 42 maintains the valve open state (FIG. 4C). Subsequently, the hydraulic oil in the high-pressure side oil passage 35 flows into the cylinder 40 and pushes down the piston 43 (FIG. 4D).

Immediately before the piston 43 reaches bottom dead center, the control device 50 stops supplying power to the first solenoid 41d. As a result, the low-pressure side check valve 41 is opened, and in the subsequent downward stroke of the piston 43, the high-pressure side check valve 42 is closed, as shown in FIG. 4E. Further, the control device 50 stops the supply of electric power to the second solenoid 42d immediately after the supply of electric power to the first solenoid 41d is stopped. As a result, when the piston 43 reaches the bottom dead center, the low-pressure side check valve 41 is opened and the high-pressure side check valve 42 is closed (FIG. 4F). From here, when the piston 43 rises toward the top dead center due to the rotational inertia of the crankshaft 45, the hydraulic oil in the cylinder 40 flows out to the low pressure side oil passage 36 (FIG. 4G).

In the process in which the piston 43 rises from the bottom dead center to the top dead center, power is supplied to the first solenoid 41d to advance the timing of closing the low-voltage side check valve 41. It is possible to prevent the hydraulic oil from flowing out to the low pressure side oil passage 36. By adjusting the timing of closing the low-pressure check valve 41 in this way, the capacity of the first pump motor 31 can be adjusted from zero capacity to maximum capacity Q1.

Returning to FIG. 1, the second pump motor 32 has a high-pressure side port 32a that opens in the high-pressure side oil passage 35 and a low-pressure side port 32b that opens in the low-pressure side oil passage 36. The second pump motor 32 is a variable capacity type pump motor capable of changing the capacity from zero capacity to the maximum capacity Q2. In this embodiment, Q1 <Q2.

Since the configuration of the second pump motor 32 is the same as the configuration of the first pump motor 31, detailed description thereof will be omitted. Further, the second pump motor 32 is connected to the third pump motor 33 by a shaft 32c so as to rotate integrally with the third pump motor 33. The second pump motor 32 corresponds to the "second hydraulic pump motor" of the present invention.

The third pump motor 33 is a fixed capacity type hydraulic pump motor. The third pump motor 33 has a port 33a that opens in the oil passage 37 and a port 33b that opens in the oil passage 38 so that the hydraulic oil that has flowed in from one port flows out from the other port. It is configured.

The third pump motor 33 is, for example, an axial piston type pump motor. The direction of rotation of the third pump motor 33 depends on the port into which the hydraulic oil flows. For example, when the hydraulic oil flows in from the oil passage 37 and flows out to the oil passage 38, the third pump motor 33 rotates in the forward direction. Further, when the hydraulic oil flows in from the oil passage 38 and flows out to the oil passage 37, the third pump motor 33 rotates in the reverse direction.

In other words, when the third pump motor 33 rotates in the forward direction, the hydraulic oil flows from the oil passage 37 toward the oil passage 38. Further, when the third pump motor 33 rotates in the reverse direction, the hydraulic oil flows from the oil passage 38 toward the oil passage 37.

The capacity of the third pump motor 33 is Q2, which is equal to the maximum capacity of the second pump motor 32. The relationship between the flow direction of the hydraulic oil and the rotation direction of the motor is not limited to the above-mentioned relationship, and may be the opposite of the above-mentioned relationship. The third pump motor 33 corresponds to the "hydraulic motor" of the present invention.

The switching valve 34 is a two-position switching valve. The switching valve 34 is electrically connected to the control device 50, and switches the flow direction of the hydraulic oil based on the control command from the control device 50. When power is not supplied from the control device 50, the switching valve 34 is in the first state of communicating the high-pressure side oil passage 35 and the oil passage 37 and communicating the low-pressure side oil passage 36 and the oil passage 38. .. Further, when power is supplied from the control device 50, the switching valve 34 is in a second state in which the high-pressure side oil passage 35 and the oil passage 38 are communicated with each other and the low-pressure side oil passage 36 and the oil passage 37 are communicated with each other. Become.

Next, with reference to FIGS. 5 to 10, the operating states of the first pump motor 31, the second pump motor 32, and the third pump motor 33 and the switching position of the switching valve 34 in each running state will be described.

First, a low torque start will be described with reference to FIG. The low torque start is a case where the torque required at the time of starting is small, such as when starting on a flat road with a small load capacity.

In this case, the capacity of the second pump motor 32 is set to zero. Further, the switching valve 34 is in the first state. Further, the capacity of the first pump motor 31 is controlled from zero capacity to maximum capacity.

As a result, as shown in FIG. 5, the hydraulic oil discharged from the high-pressure side port 31a of the first pump motor 31 to the high-pressure side oil passage 35 flows into the port 33a of the third pump motor 33 and flows into the third pump motor 33. The 33 is rotated forward. Then, the hydraulic oil flowing out from the port 33b of the third pump motor 33 is sent to the low pressure side port 31b of the first pump motor 31.

As the capacity of the first pump motor 31 is gradually increased, the speed ratio (rotation speed of the third pump motor 33 / rotation speed of the drive source 10) increases. The maximum speed ratio at low torque start is (Q1 / Q2). In low torque start, it can be regarded as a combination of a variable capacity pump whose capacity is variable from zero capacity to maximum capacity Q1 and a fixed capacity motor with capacity Q2.

Next, a high torque start will be described with reference to FIG. The high torque start is a case where the vehicle is started in a high load state, a case where the vehicle is started on an uphill road, and the like, and is a case where the torque required at the time of starting is large.

In a high load state and an uphill road, the torque required at the time of starting may exceed the torque that can be generated when only the third pump motor 33 is used as a motor. In such a case, by operating the second pump motor 32 as a motor, high torque can be generated and the vehicle 1 can be started.

At the time of high torque start, the second pump motor 32 is controlled so that the hydraulic oil flows into the cylinder 40 from the high pressure side oil passage 35 and flows out to the low pressure side oil passage 36. At that time, the capacity of the second pump motor 32 is adjusted to an arbitrary capacity Q3 from zero capacity to the maximum capacity Q2 according to the required torque.

Further, the switching valve 34 is in the first state. Further, the capacity of the first pump motor 31 is controlled from zero capacity to the maximum capacity as in the case of low torque start.

As a result, as shown in FIG. 6, the hydraulic oil discharged from the high-pressure side port 31a of the first pump motor 31 to the high-pressure side oil passage 35 flows into the port 33a of the third pump motor 33 and flows into the third pump motor. While rotating 33 in the forward direction, it flows into the high-pressure side port 32a of the second pump motor 32 to rotate the second pump motor 32. Since the second pump motor 32 is connected to the third pump motor 33 via the shaft 32c, the rotation direction of the second pump motor 32 is forward rotation.

Then, the hydraulic oil that has flowed out from the low-pressure side port 32b of the second pump motor 32 and the port 33b of the third pump motor 33 is sent to the low-pressure side port 31b of the first pump motor 31.

As the capacity of the first pump motor 31 is gradually increased, the speed ratio increases. In high torque start, it can be regarded as a combination of a variable capacity pump whose capacity is variable from zero capacity to maximum capacity Q1 and a fixed capacity motor with capacity (Q2 + Q3).

Therefore, a higher torque can be generated as compared with a low torque start. In high torque start, higher torque can be generated in a predetermined capacity of the first pump motor 31 as compared with low torque start, but the speed ratio is lower than in low torque start.

When the capacity of the first pump motor 31 is increased to the maximum capacity Q1 and the speed ratio becomes (Q1 / (Q2 + Q3)) in a high torque start, the capacity of the second pump motor 32 is increased in order to further increase the speed ratio. The control to reduce from Q3 to zero is executed.

As the capacity of the second pump motor 32 is gradually reduced from Q3 to zero, the speed ratio increases. In this case, it can be regarded as a combination of a fixed capacity pump having a capacity of Q1 and a variable capacity type motor having a variable capacity from (Q2 + Q3) to Q2.

Therefore, the speed ratio can be increased to the maximum speed ratio (Q1 / Q2) at low torque start.

Next, high-speed running will be described with reference to FIG. 7. As described above, the speed ratio can be increased from zero to (Q1 / Q2) for both low torque start and high torque start. Then, in the present embodiment, the second pump motor 32 is controlled as follows in order to further increase the speed ratio.

The second pump motor 32 is controlled so as to suck the hydraulic oil from the low pressure side oil passage 36 and discharge the hydraulic oil to the high pressure side oil passage 35 after the capacity is set to zero. As a result, the sum of the suction amounts of the hydraulic oils of the second pump motor 32 and the third pump motor 33 becomes smaller, and the speed ratio can be further increased.

Further, in the present embodiment, the maximum capacity Q2 of the second pump motor 32 is equal to the capacity of the third pump motor 33. Therefore, by setting the capacity of the second pump motor 32 to the maximum capacity Q2, the sum of the suction amounts of the hydraulic oils of the second pump motor 32 and the third pump motor 33 can be set to zero.

Therefore, the second pump motor 32 and the third pump motor 33 can be rotated to the mechanical limit speeds of the second pump motor 32 and the third pump motor 33 regardless of the capacity of the first pump motor 31.

Next, a case where deceleration is performed by the engine brake will be described with reference to FIGS. 8 and 9. When the driving force generated by the drive source 10 is reduced due to the fact that the accelerator pedal is no longer depressed, the third pump motor 33 is rotated from the wheel side.

In this case, in the present embodiment, the following controls are executed in order to generate the engine braking force. First, the control device 50 supplies electric power to the switching valve 34, switches the switching valve 34, and communicates the oil passage 38 and the high-pressure side oil passage 35.

Further, the low-pressure side check valve 41 and the high-pressure side check valve 42 of the first pump motor 31 allow hydraulic oil to flow into the cylinder 40 from the high-pressure side oil passage 35 by the control device 50 to lower the piston 43. When the piston 43 is raised, the hydraulic oil is controlled to flow out to the low pressure side oil passage 36.

As a result, the first pump motor 31 is generated with a torque for rotating the drive source 10. On the other hand, the engine brake acts due to the internal resistance of the drive source 10.

The engine braking force depends on the number of revolutions of the drive source 10. Therefore, the control device 50 controls the capacity of the first pump motor 31 so that the rotation speed of the drive source 10 is such that a desired engine braking force can be obtained.

Further, for example, when decelerating during high-speed running, all the energy of the hydraulic oil discharged by the third pump motor 33 cannot be converted into the rotation of the first pump motor 31. Therefore, the second pump motor 32 is controlled as follows.

That is, as shown in FIG. 8, the low-pressure side check valve 41 and the high-pressure side check valve 42 of the second pump motor 32 allow hydraulic oil to flow into the cylinder 40 from the high-pressure side oil passage 35 and lower the piston 43. It is controlled so that the hydraulic oil flows out to the low pressure side oil passage 36 when the piston 43 is raised.

As a result, a torque for rotating the second pump motor 32, that is, a torque for accelerating the vehicle 1 is generated. Therefore, over-rotation of the first pump motor 31 can be prevented.

On the other hand, for example, when decelerating during low-speed traveling, the hydraulic oil discharged from the third pump motor 33 may be insufficient in order to obtain a desired engine braking force. In this case, the second pump motor 32 is controlled so as to suck the hydraulic oil from the low pressure side oil passage 36 and discharge the hydraulic oil to the high pressure side oil passage 35. By doing so, the flow rate of the hydraulic oil flowing into the first pump motor 31 can be satisfied.

Next, the reverse traveling will be described with reference to FIG. When traveling in reverse, the control device 50 supplies electric power to the switching valve 34, switches the switching valve 34, and communicates the high-pressure side oil passage 35 and the oil passage 38.

As a result, the hydraulic oil discharged from the first pump motor 31 to the high-pressure side oil passage 35 flows into the port 33b of the third pump motor 33 through the oil passage 38, and reversely rotates the third pump motor 33. Then, the hydraulic oil flowing out from the port 33a of the third pump motor 33 is sent to the low pressure side port 31b of the first pump motor 31.

At this time, the second pump motor 32 is controlled in the same manner as when moving forward, according to the torque and speed ratio required by the vehicle 1.

As described above, according to the stepless transmission according to the present embodiment, a variable displacement type first hydraulic pump motor that is rotationally driven by a drive source, the first hydraulic pump motor, and a closed circuit are configured. A fixed-capacity hydraulic motor that is rotationally driven by the pressure oil delivered from the first hydraulic pump motor, and an oil passage that connects the first hydraulic pump motor and the hydraulic motor are connected to the hydraulic motor. A variable displacement type second hydraulic pump motor connected to the above, and a portion of the oil passage between the hydraulic motor and the second hydraulic pump motor, which can switch the direction of the pressure oil flow. It is equipped with a valve.

As a result, the second hydraulic pump motor is rotated by the rotational force of the hydraulic motor, and the hydraulic oil discharged from the second hydraulic pump motor is sucked into the hydraulic motor. The sum of the inhalation amount can be reduced.

Therefore, a high speed ratio can be obtained without using a large-capacity hydraulic pump and a hydraulic motor, and the vehicle can be driven at high speed.

Further, according to the present embodiment, by sucking the hydraulic oil discharged from the first hydraulic pump motor into the second hydraulic pump motor, the second hydraulic pump motor is connected to each other by the hydraulic oil discharged from the first hydraulic pump motor. The hydraulic pump motor and the hydraulic motor can be driven to rotate.

As a result, high torque can be obtained when the vehicle starts without using a large-capacity hydraulic pump and hydraulic motor. Therefore, it is possible to start the vehicle in a highly loaded state. In addition, it is possible to start the vehicle on an uphill road.

Further, according to the present embodiment, the engine brake can be actuated by sucking the hydraulic oil discharged from the hydraulic motor into the first hydraulic pump motor during deceleration. In this case, over-rotation of the first hydraulic pump motor can be prevented by sucking the hydraulic oil discharged from the hydraulic motor into the second hydraulic pump motor as well. Further, the engine brake can be appropriately applied by sucking the hydraulic oil discharged from the hydraulic motor and the second hydraulic pump motor into the first hydraulic pump motor.

Further, in the present embodiment, the second hydraulic pump motor has a low-pressure side check valve and a high-pressure side check valve. The effect of this will be briefly described with reference to a comparative example.

When the second hydraulic pump motor is composed of a variable displacement pump motor having a valve structure using a clearance seal, such as a swash plate pump motor and a sloping shaft pump motor, when the second hydraulic pump motor is not used as a pump or a motor. However, hydraulic oil always leaks. Therefore, efficiency may deteriorate.

On the other hand, in the present embodiment, since the second hydraulic pump motor has the above-described configuration, the hydraulic oil does not leak even when the second hydraulic pump motor is not used as a pump or a motor. Therefore, efficiency can be improved.

In the above-described embodiment, the capacity of the first and second hydraulic pump motors is adjusted by adjusting the valve opening and closing timings in one stroke of one cylinder. And the method of adjusting the capacity of the second hydraulic pump motor is not limited to this.

As described above, the first and second hydraulic pump motors include a plurality of cylinders. Therefore, for some of the cylinders, keep the low-pressure check valve always open and the high-pressure check valve always closed, and do not use some of the cylinders as a pump or motor. By doing so, the discharge amount of hydraulic oil per unit time can be adjusted.

Specifically, for example, assuming that the number of cylinders of the hydraulic pump motor is 12, by not using 6 cylinders, the amount of hydraulic oil discharged per unit time can be adjusted by using all the cylinders. It can be 50% of. Further, for example, by not using nine cylinders, the discharge amount of hydraulic oil per unit time can be set to 25% of the case where all cylinders are used.

Further, in the above-described embodiment, the maximum capacity of the second hydraulic pump motor and the capacity of the hydraulic motor are equalized, but the maximum capacity of the second hydraulic pump motor and the capacity of the hydraulic motor are not limited to this. , May be different.

Further, in the above-described embodiment, the hydraulic motor is an axial piston type hydraulic motor, but the present invention is not limited to this. Specifically, for example, the hydraulic motor may be a hydraulic motor on the radial piston side, or may be another type of hydraulic motor.

The continuously variable transmission of the present invention is suitably used for a vehicle such as a truck that requires high torque at the time of starting and travels at high speed.

1 Vehicle 10 Drive source 11 Output shaft 20 Damper 21 Input side member 22 Output side member 23 Elastic connecting member 30 Stepless transmission 30a Input shaft 31 First pump motor 31a High pressure side port 31b Low pressure side port 32 Second pump motor 32a High pressure Side port 32b Low pressure side port 32c Shaft 33 Third pump motor 33a Port 33b Port 34 Switching valve 35 High pressure side oil passage 36 Low pressure side oil passage 37 Oil passage 38 Oil passage 40 Cylinder 41 Low pressure side check valve 41a First valve seat 41b 1st valve body 41c 1st spring 41d 1st solenoid 42 High pressure side check valve 42a 2nd valve seat 42b 2nd valve body 42c 2nd spring 42d 2nd solenoid 43 piston 44 link mechanism 45 Crankshaft 45a Crank angle detection sensor 50 Control device

Claims (3)

A variable displacement type first hydraulic pump motor that is rotationally driven by a drive source,
A fixed-capacity hydraulic motor that constitutes a closed circuit with the first hydraulic pump motor and is rotationally driven by the pressure oil delivered from the first hydraulic pump motor.
A variable displacement type second hydraulic pump motor connected to the oil passage connecting the first hydraulic pump motor and the hydraulic motor and also connected to the hydraulic motor.
Said provided portions between the hydraulic motor and the second hydraulic pump motor, Bei give a, and a switching valve capable of switching the direction of the pressure oil flow in the oil passage,
The maximum capacity of the second hydraulic pump motor is equal to the capacity of the hydraulic motor.
Continuously variable transmission. The first and second hydraulic pump motors
The hydraulic oil chamber and
A piston whose volume in the hydraulic oil chamber can be changed,
A first check valve that opens when the pressure oil flows from the oil passage toward the hydraulic oil chamber.
A first actuator capable of controlling the opening and closing of the first check valve,
A second check valve that opens when the pressure oil flows from the hydraulic oil chamber toward the oil passage, and
Continuously variable transmission according to claim 1, and a second actuator capable of controlling the opening and closing of the second check valve. The hydraulic motor is an axial piston type hydraulic motor.
The continuously variable transmission according to any one of claims 1 and 2.

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