JP2018150964A - Stepless gear change device - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a stepless gear change device which can make a vehicle travel at a high speed without using a large-capacity hydraulic pump motor.SOLUTION: A stepless gear change device comprises: a variable capacity type first hydraulic pump motor which is rotationally driven by a drive source; a fixed capacity type hydraulic motor which constitutes a closed circuit together with the first hydraulic pump motor, and is rotationally driven by pressure oil sent from the first hydraulic pump motor; a variable capacity type second hydraulic pump motor which is connected to an oil passage for connecting the first hydraulic pump motor and the hydraulic motor, and connected to the hydraulic motor; and a changeover valve which is arranged in a portion between the hydraulic motor in the oil passage and the second hydraulic pump motor, and can switch a direction of a flow of the pressure oil.SELECTED DRAWING: Figure 1

Description

  The present invention relates to a continuously variable transmission.

  BACKGROUND ART Conventionally, a hydrostatic continuously variable transmission (hydrostatic transmission: HST) used for driving a vehicle is 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 hydraulic pressure generated by the hydraulic pump are connected.

Japanese Patent Laying-Open No. 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 a high speed.

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

  The continuously variable transmission according to the present invention comprises a variable displacement first hydraulic pump motor that is rotationally driven by a drive source, a closed circuit with the first hydraulic pump motor, and is sent from the first hydraulic pump motor. A fixed-capacity hydraulic motor that is rotationally driven by the pressurized oil, and a variable-capacity type first motor that is connected to an oil passage that connects the first hydraulic pump motor and the hydraulic motor, and is connected to the hydraulic motor. 2 hydraulic pump motors, and a switching valve provided in a portion of the oil passage between the hydraulic motor and the second hydraulic pump motor and capable of switching the flow direction of the pressure oil.

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

FIG. 1 is a skeleton diagram showing the overall configuration of a vehicle equipped with a continuously variable transmission according to the present invention. Schematic diagram of cylinder in variable displacement pump motor Diagram showing the process of drawing hydraulic oil from the low pressure side and discharging it to the high pressure side Diagram showing the process of drawing hydraulic oil from the low pressure side and discharging it to the high pressure side Diagram showing the process of drawing hydraulic oil from the low pressure side and discharging it to the high pressure side Diagram showing the process of drawing hydraulic oil from the low pressure side and discharging it to the high pressure side Diagram showing the process of drawing hydraulic oil from the low pressure side and discharging it to the high pressure side Diagram showing the process of drawing hydraulic oil from the low pressure side and discharging it to the high pressure side The figure which shows the stroke where hydraulic oil flows in from the high pressure side and flows out to the low pressure side The figure which shows the stroke where hydraulic oil flows in from the high pressure side and flows out to the low pressure side The figure which shows the stroke where hydraulic oil flows in from the high pressure side and flows out to the low pressure side The figure which shows the stroke where hydraulic oil flows in from the high pressure side and flows out to the low pressure side The figure which shows the stroke where hydraulic oil flows in from the high pressure side and flows out to the low pressure side The figure which shows the stroke where hydraulic oil flows in from the high pressure side and flows out to the low pressure side The figure which shows the stroke where hydraulic oil flows in from the high pressure side and flows out to the low pressure side Skeleton diagram showing when starting low torque Skeleton diagram showing high torque start Skeleton diagram showing high-speed driving Skeleton diagram showing the deceleration state at high speed Skeleton diagram showing deceleration state during low-speed driving Skeleton diagram showing reverse travel

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

  First, an 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. 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 the present embodiment, description will be made assuming that the drive source 10 is a diesel engine. An input side member 21 of a 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 connection 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 30 a of the continuously variable transmission 30. The damper 20 has a function of suppressing the rotation 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 the structure of a general damper, detailed description is omitted.

  The continuously variable transmission 30 is configured by connecting a first pump motor 31, a second pump motor 32, a third pump motor 33, and a switching valve 34 to each other through an oil path to form a closed circuit. 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 31 a that opens to the high-pressure side oil passage 35 and a low-pressure side port 31 b that opens to the low-pressure side oil passage 36. The first pump motor 31 is a variable displacement 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. In FIG. 2, one cylinder 40 in the first pump motor 31 is depicted. 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 the link mechanism 44. The link mechanism 44 converts the rotational movement of the crankshaft 45 into the vertical movement of the piston 43. Conversely, 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 30 a 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 side 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 away from the first valve seat 41a. Have The low pressure side check valve 41 is, for example, a poppet valve. The low pressure side check valve 41 corresponds to the “first check valve” of the present invention.

  Moreover, the low pressure side check valve 41 is configured to be closed at a desired timing. Specifically, the low pressure side check valve 41 is provided with a first solenoid 41d for seating 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 electric power is 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 side check valve 42 includes a second valve seat 42a, a second valve body 42b, and a second spring 42c that urges the second valve body 42b in a direction to seat the second valve seat 42a on the second valve seat 42a. And have. The high pressure side 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 check valve 42 opens at a predetermined timing. Furthermore, the high pressure side check valve 42 is configured to be able to maintain a valve open state. Specifically, the high-pressure check valve 42 is provided with a second solenoid 42d that maintains the separated state of the second valve body 42b that is 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 electric power is 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 FIG. 3A to FIG. 3F, the process of sucking the hydraulic oil from the low pressure side oil passage 36 into the cylinder 40 and discharging it to the high pressure side oil passage 35 will be described.

  In the intake stroke (FIGS. 3A to 3C) in which the piston 43 descends from the top dead center to the bottom dead center with the rotation of the crankshaft 45, the first solenoid 41d is not supplied with power from the control device 50, The first valve body 41b is in a state of being separated from the first valve seat 41a. In this way, in the intake 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. In the intake stroke, the high-pressure check valve 42 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 the bottom dead center to the top dead center. Thereby, the 1st valve body 41b maintains the state spaced apart from the 1st valve seat 41a.

  In a state where 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 is kept closed 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 during the discharge stroke, the low pressure check valve 41 is closed and the pressure in the cylinder 40 is increased, and the high pressure check valve 42 is opened. Then, the hydraulic oil in the cylinder 40 is led to the high pressure side oil passage 35 (FIGS. 3E to 3F). At this time, the thrust of the second solenoid 42d is not necessarily required to maintain the open state of the high-pressure check valve 42, but the second valve element 42b is forced by the thrust of the second solenoid 42d. You may cancel the load of the 2nd spring 42c which acts.

  Thus, by adjusting the timing for closing the low pressure side check valve 41, the capacity of the first pump motor 31 can be adjusted from zero capacity to the maximum capacity Q1.

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

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

  In this state, the control device 50 supplies power to the first solenoid 41d. Thereby, the low pressure side check valve 41 is closed, and in the subsequent upward stroke of the piston 43, as shown in FIG. 4B, the high pressure side check valve 42 is opened. The control device 50 also supplies power to the second solenoid 42d immediately after the start of power supply to the first solenoid 41d.

  Thereby, in a state where the piston 43 has reached the 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 the bottom dead center, the control device 50 stops the power supply to the first solenoid 41d. Thereby, 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 power to the second solenoid 42d immediately after the supply of power to the first solenoid 41d is stopped. As a result, when the piston 43 has reached bottom dead center, the low pressure check valve 41 is opened and the high pressure check valve 42 is closed (FIG. 4F). From this point, 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 stroke in which the piston 43 rises from the bottom dead center to the top dead center, the electric power to the first solenoid 41d is supplied to advance the timing for closing the low pressure check valve 41. It is possible to prevent the hydraulic oil from flowing out to the low pressure side oil passage 36. Thus, by adjusting the timing for closing the low pressure side check valve 41, the capacity of the first pump motor 31 can be adjusted from zero capacity to the maximum capacity Q1.

  Returning to FIG. 1, the second pump motor 32 has a high-pressure side port 32 a that opens to the high-pressure side oil passage 35 and a low-pressure side port 32 b that opens to the low-pressure side oil passage 36. The second pump motor 32 is a variable capacity pump motor that can change the capacity from zero capacity to the maximum capacity Q2. In the present 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 is omitted. 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 displacement hydraulic pump motor. The third pump motor 33 has a port 33a that opens to the oil passage 37 and a port 33b that opens to the oil passage 38, so that the hydraulic oil flowing 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 rotation direction of the third pump motor 33 depends on the port through which hydraulic oil flows. For example, when hydraulic oil flows in from the oil passage 37 and flows out to the oil passage 38, the third pump motor 33 rotates forward. Further, when 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 forward, 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. In addition, the relationship between the flow direction of the hydraulic oil and the rotation direction of the motor is not limited to the above relationship, and may be a relationship opposite to the above 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 direction in which the hydraulic oil flows based on a control command from the control device 50. The switching valve 34 is in a first state in which the high pressure side oil passage 35 and the oil passage 37 are communicated and the low pressure side oil passage 36 and the oil passage 38 are communicated when electric power is not supplied from the control device 50. . In addition, 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 and the low pressure side oil passage 36 and the oil passage 37 are communicated. Become.

  Next, 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 traveling state will be described with reference to FIGS.

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

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

  Accordingly, 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 the third pump motor. 33 is rotated forward. The hydraulic fluid that has flowed out of the port 33 b of the third pump motor 33 is sent to the low-pressure side port 31 b of the first pump motor 31.

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

  Next, the 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, or the like, and a case where the torque required at the time of start is large.

  In a high loading 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, a high torque can be generated and the vehicle 1 can be started.

  When the high torque starts, 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 into the low pressure side oil passage 36. At this 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.

  Moreover, the switching valve 34 is set to the first state. Moreover, the capacity | capacitance of the 1st pump motor 31 is controlled from the zero capacity | capacitance to the maximum capacity | capacitance similarly to the case of the low torque start.

  Accordingly, 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 33 is rotated forward, it flows into the high-pressure side port 32 a of the second pump motor 32 and rotates the second pump motor 32. In addition, since the 2nd pump motor 32 is connected with the 3rd pump motor 33 via the axis | shaft 32c, the rotation direction of the 2nd pump motor 32 turns into normal rotation.

  The hydraulic fluid that has flowed out from the low pressure side port 32 b of the second pump motor 32 and the port 33 b of the third pump motor 33 is sent to the low pressure side port 31 b 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 starting, it can be considered as a combination of a variable displacement pump whose capacity is variable from zero capacity to maximum capacity Q1 and a fixed capacity motor with capacity (Q2 + Q3).

  For this reason, it is possible to generate a higher torque compared to the low torque start. In the high torque start, a higher torque can be generated in the predetermined capacity of the first pump motor 31 than in the low torque start, but the speed ratio is lower than in the low torque start.

  In high 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)), the capacity of the second pump motor 32 is increased to further increase the speed ratio. Control is performed to decrease from Q3 to zero.

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

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

  Next, high speed traveling will be described with reference to FIG. As described above, the speed ratio can be increased from zero to (Q1 / Q2) for both low torque start and high torque start. And in this embodiment, in order to raise a speed ratio further, the 2nd pump motor 32 is controlled as follows.

  The second pump motor 32 is controlled so as to draw 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 setting the capacity to zero. As a result, the sum of the amounts of hydraulic oil drawn by the second pump motor 32 and the third pump motor 33 is reduced, and the speed ratio can be further increased.

  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 hydraulic oil of the second pump motor 32 and the third pump motor 33 can be made zero.

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

  Next, with reference to FIG. 8 and FIG. 9, the case where deceleration is performed by engine braking will be described. When the driving force generated by the driving source 10 decreases due to the depression of the accelerator pedal or the like, the third pump motor 33 is rotated from the wheel side.

  In this case, in the present embodiment, the following control is executed to generate the engine braking force. First, the control device 50 supplies power to the switching valve 34, switches the switching valve 34, and causes the oil passage 38 and the high-pressure side oil passage 35 to communicate with each other.

  Further, the low pressure side check valve 41 and the high pressure side check valve 42 of the first pump motor 31 cause the control device 50 to cause the hydraulic oil to flow into the cylinder 40 from the high pressure side oil passage 35 and to lower the piston 43. Control is performed so that the hydraulic oil flows out to the low-pressure side oil passage 36 when the piston 43 is raised.

  As a result, torque is generated in the first pump motor 31 so as to rotate the drive source 10. On the other hand, the engine brake acts by the internal resistance of the drive source 10.

  The engine braking force depends on the rotational speed of the drive source 10. Therefore, the control apparatus 50 controls the capacity | capacitance of the 1st pump motor 31 so that it may become the rotation speed of the drive source 10 which can obtain desired engine braking force.

  Furthermore, for example, when deceleration is performed during high-speed traveling, all of the energy of the hydraulic oil discharged from the third pump motor 33 cannot be converted into 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 cause the hydraulic oil to flow into the cylinder 40 from the high pressure side oil passage 35 and lower the piston 43. The hydraulic oil is controlled to flow out to the low pressure side oil passage 36 when the piston 43 is raised.

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

  On the other hand, for example, when the vehicle is decelerated during low-speed traveling, the hydraulic oil discharged from the third pump motor 33 may be insufficient to obtain a desired engine braking force. In this case, the second pump motor 32 is controlled so as to suck 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, reverse travel will be described with reference to FIG. During reverse travel, the control device 50 supplies power to the switching valve 34, switches the switching valve 34, and causes the high-pressure side oil passage 35 and the oil passage 38 to communicate with each other.

  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 via the oil passage 38, and reversely rotates the third pump motor 33. The hydraulic fluid that has flowed out from the port 33 a of the third pump motor 33 is sent to the low-pressure side port 31 b of the first pump motor 31.

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

  As described above, according to the continuously variable transmission according to the present embodiment, the variable displacement type first hydraulic pump motor that is rotationally driven by the drive source, the first hydraulic pump motor, and the closed circuit are configured, A fixed displacement hydraulic motor that is rotationally driven by pressure oil delivered from the first hydraulic pump motor; and an oil passage that connects the first hydraulic pump motor and the hydraulic motor; and the hydraulic motor A variable displacement type second hydraulic pump motor connected to the hydraulic passage, and a switch provided in a portion of the oil passage between the hydraulic motor and the second hydraulic pump motor, and capable of switching a flow direction of the pressure oil. And a valve.

  Thus, 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, so that the hydraulic oil of the second hydraulic pump motor and the hydraulic motor is reduced. The sum of the amount of inhalation can be reduced.

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

  Further, according to the present embodiment, the second hydraulic pump motor discharges the hydraulic oil discharged from the first hydraulic pump motor, whereby the second hydraulic pump motor is discharged from the first hydraulic pump motor so that the second hydraulic oil is connected to each other. The hydraulic pump motor and the hydraulic motor can be rotationally driven.

  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, the vehicle can be started on the uphill road.

  Further, according to the present embodiment, the engine brake can be applied by causing the first hydraulic pump motor to suck the hydraulic oil discharged from the hydraulic motor during deceleration. In this case, the hydraulic oil discharged from the hydraulic motor is also sucked into the second hydraulic pump motor, thereby preventing the first hydraulic pump motor from over-rotating. In addition, by causing the first hydraulic pump motor to suck the hydraulic oil discharged from the hydraulic motor and the second hydraulic pump motor, the engine brake can be appropriately applied.

  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 constituted by a variable displacement pump motor having a valve structure using a clearance seal, such as a swash plate pump motor or a swash shaft pump motor, the second hydraulic pump motor is not used as a pump or a motor. However, hydraulic fluid always leaks. Therefore, the efficiency may be deteriorated.

  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 adjustment of the first and second hydraulic pump motors has been described by way of example by adjusting the valve opening and closing timings in one stroke of one cylinder. And the method of adjusting the capacity | capacitance of a 2nd 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 cylinders, keep the low-pressure check valve open and the high-pressure check valve always closed, and do not use the cylinder as a pump or motor. By doing so, the discharge amount of hydraulic oil per unit time can be adjusted.

  Specifically, for example, when the number of cylinders of the hydraulic pump motor is 12, when all cylinders are used, the discharge amount of hydraulic oil per unit time can be reduced by not using six cylinders. Of 50%. Further, for example, by not using nine cylinders, the discharge amount of hydraulic oil per unit time can be 25% when all cylinders are used.

  In the above-described embodiment, the maximum capacity of the second hydraulic pump motor is equal to the capacity of the hydraulic motor. However, the present invention is not limited to this, and the maximum capacity of the second hydraulic pump motor and the capacity of the hydraulic motor are , May be different.

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

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

DESCRIPTION OF SYMBOLS 1 Vehicle 10 Drive source 11 Output shaft 20 Damper 21 Input side member 22 Output side member 23 Elastic connection member 30 Continuously variable transmission 30a Input shaft 31 1st pump motor 31a High pressure side port 31b Low pressure side port 32 2nd 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 (4)

A variable displacement first hydraulic pump motor that is rotationally driven by a drive source;
A fixed displacement 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 second hydraulic pump motor connected to an oil passage connecting the first hydraulic pump motor and the hydraulic motor and coupled to the hydraulic motor;
A switching valve provided in a portion between the hydraulic motor and the second hydraulic pump motor in the oil passage, and capable of switching a flow direction of the pressure oil.
Continuously variable transmission. The maximum capacity of the second hydraulic pump motor is equal to the capacity of the hydraulic motor;
The continuously variable transmission according to claim 1. The first and second hydraulic pump motors are
A hydraulic oil chamber;
A piston capable of changing a volume in the hydraulic oil chamber;
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 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;
The continuously variable transmission according to claim 1, further comprising: a second actuator capable of controlling 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 to 3.

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