JPH04259675A - Hydraulic device with swash plate plunger - Google Patents

Abstract

PURPOSE:To make graduate oil pressure change in a cylinder hole mating with each plunger, and suppress change in the moment around a swash plate inclining shaft and also the thrust synthesized load. CONSTITUTION:A distribution valve plate 7 is provided with an influx port 15 and an efflux port 14, and a communication port 13a is formed between the influx and efflux port 15, 14. These ports 15, 14 are arranged so that the oil pressures in this communication port and cylinder hole leading to it meet the conditions theta1=theta2 and theta3=180 deg., where theta1 is the rotational angle of cylinder block corresponding to the division in which boosting is made from the pressure in the port on low pressure side to the pressure in the port on high pressure side, theta2 is the rotational angle of cylinder block corresponding to the division in which decompression is made from the pressure in the port on high pressure side to the pressure in the port on low pressure side, and theta3 is the rotational angle from commencement of boosting to starting decompression.

Description

[Detailed description of the invention]

0001

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a swash plate plunger type hydraulic device such as a swash plate plunger type hydraulic pump, motor, etc.

0002

[Prior Art] Such a swash plate plunger type hydraulic pump and motor are known, for example, from Japanese Patent Application Laid-Open No. 61-118566.
There is a device disclosed in the publication. In such devices, it is common to use an odd number of plungers so that the discharge and suction phases of each plunger are shifted to reduce fluctuations in flow rate and torque. Note that a hydraulic continuously variable transmission may be constructed by combining a swash plate plunger type hydraulic pump and a motor, and the same applies to such a case.

[0003] Also, the plunger is in the discharge stroke (contraction stroke).
When the plunger moves to the suction stroke (expansion stroke), a sudden change in hydraulic pressure occurs in the cylinder hole into which the plunger is slid, and this is transmitted as an excitation force to the plunger, swash plate, casing, etc., and the above-mentioned hydraulic system and this It has been known for some time that this is one of the causes of noise from hydraulic continuously variable transmissions. For this reason, in order to moderate the above oil pressure change, a pre-compression and pre-expansion section may be provided between the above two strokes, or a throttle passage (V-shaped groove, cutout hole, regulating valve, etc.) may be provided. There have been many proposals to date. (For example, Utility Model Application Publication No. 63-96372, Japanese Patent Application Publication No. 2-129461, etc.)

0004

[Problem to be Solved by the Invention] However, even if the oil pressure changes are made gentler as described above, the oil pressure changes in the cylinder holes corresponding to each plunger are only made gentler, and the thrust load acting on all the plungers is reduced. It has been found that there is a problem in that the composite value of (thrust composite load) fluctuates, and this thrust composite load acts as an excitation force, making it difficult to sufficiently reduce noise.

[0005] Fluctuations in the thrust composite load will be explained with reference to FIG. 21. This figure shows the thrust loads F1-F9 that act on each plunger according to the rotation of the cylinder block, the composite value of these thrust forces, and
That is, the thrust composite load Ft is shown. Although time t is shown on the horizontal axis, since the rotation angle of the cylinder block changes with time, the rotation angle may also be shown on the horizontal axis. As can be seen from this figure, the thrust load of each plunger changes smoothly with pressure increase and pressure decrease sections, but the combined thrust load Ft, which is a combination of these sections, fluctuates as shown.

On the other hand, if the swash plate plunger type hydraulic pump or motor is a variable displacement type and has a tilting support shaft, or if it is a fixed displacement type but has a support shaft similar to this tilting support shaft, each Even if the hydraulic pressure change in the cylinder hole corresponding to the plunger is made gentler, it is not possible to sufficiently suppress the fluctuation of the moment around the above-mentioned spindle, which is thought to be a factor in the excitation force, and the noise from these pumps or motors increases. It has also been found that there is a problem in that it is difficult to sufficiently reduce the

The present invention has been developed in view of the above-mentioned problems, and is designed to moderate changes in oil pressure (pressure increase and decrease changes) in the cylinder hole corresponding to each plunger, and to reduce fluctuations in thrust composite load and swash plate. It is an object of the present invention to provide a swash plate plunger type hydraulic device configured such that fluctuations in the moment applied around the support shaft of the swash plate can be suppressed to a small level.

[0008]

[Means for Solving the Problems] In order to achieve such an object, in the present invention, in a cylinder block formed by sliding an even number of plungers in an annular arrangement surrounding a rotating shaft, a side surface opposite to a swash plate is provided. A swash plate plunger type hydraulic device is constructed by disposing a distribution valve plate in sliding contact with the side surface of the cylinder block, and an even number of connection ports communicating with each cylinder hole are arranged on the side surface of the opposite side of the cylinder block. are arranged on a predetermined circumference centered on the rotation axis, and the distribution valve plate is connected to the cylinder hole in which the plunger in the expansion stroke is slid through the connection port according to the rotation of the cylinder block. and an outflow port that communicates with the cylinder hole in which the plunger in the contraction stroke is slid, and the connecting port is located between the inflow port and the outflow port in response to rotation of the cylinder block, A rotation angle θ1 of the cylinder block corresponding to a section in which the hydraulic pressure in this connection port and the cylinder hole communicating therewith is increased from the hydraulic pressure in the low-pressure side port to the hydraulic pressure in the high-pressure side port;
The rotation angle θ2 of the cylinder block corresponding to the section where the pressure is reduced from the hydraulic pressure in the high pressure side port to the hydraulic pressure in the low pressure side port, and the rotation angle θ3 of the cylinder block from the start of pressure increase to the start of pressure reduction are: θ1=θ2, and θ3
The inflow and outflow ports are formed so that the angle is 180°. In particular, θ1=θ2=360°/Z×k However, Z:
The number of the plungers is an even value k=1, 2, 3...
・(Integer) It is preferable to form the inflow and outflow ports so that

[0009]

DESCRIPTION OF THE PREFERRED EMBODIMENTS Preferred embodiments of the present invention will be described below with reference to the drawings. FIG. 1 shows an example in which the present invention is applied to a hydraulic pump. An input shaft 2 is rotatably supported within a casing 1 of the pump via a bearing 3, and a cylinder block 4 is slidably spline-coupled to the input shaft 2. Cylinder block 4 is rotatably supported by casing 1 via bearing 5. Further, in the casing 1, a swash plate 6 is installed at one end of the cylinder block 4 (left side in the figure), and a distribution valve plate 7 is installed at the other end (right side).

The swash plate 6 has an annular shape surrounding the input shaft 2, and is mounted in an annular swash plate holder 8 that is tiltably supported by the casing 1 via a trunnion shaft (tilting support shaft) 8a. It is being Therefore, the swash plate 6 is tilted about the trunnion shaft 8a together with the swash plate holder 8, and the tilt angle with respect to the axis of rotation of the cylinder block 4 can be arbitrarily changed. The distribution valve plate 7 is fixed to the casing 1 and supports the tip of the input shaft 2 passing through the cylinder block 4 through a bearing 9 at its center. Opposing surfaces 7f, 4 between the distribution valve plate 7 and the cylinder block 4
A spring 10 is disposed between the input shaft 2 and the cylinder block 4 to urge the cylinder block 4 toward the distribution valve plate 7 in order to bring the cylinder block 4 into sliding contact.

The cylinder block 4 has ten cylinder holes 11, 11, . . . arranged at equal intervals around its rotation axis and parallel to the rotation axis.
A plunger 12 is slidably fitted inside each of them. Each plunger 12 defines an oil chamber 13 in the corresponding cylinder hole 11, and has ten connection ports 13a connected to each oil chamber 13.
is formed in the cylinder block 4 so as to open at the opposing surface 4f of the cylinder block 4, as shown in FIG. Note that the openings are formed on the same circumference and lined up at equal intervals. On the other hand, as shown in FIG. 3, the distribution valve plate 7 has the connection port 13a on one half of the opposing surface 7f.
This half has one discharge port (outflow port) 14 that communicates with the opposing port, and the opposite half has one suction port (inflow port) that communicates with the connecting port 13a that faces this half. ) 15, and these suction and discharge ports 14 and 15 are connected to discharge and suction passages 14a and 15a, respectively. A shoe 16 is swingably connected to the tip of each plunger 12, and each shoe 16 is in sliding contact with the swash plate 6. Note that in order to ensure this sliding contact, the retainer plate 17 is attached to the shoe 16.
is attached so as to be pressed toward the swash plate 6.

In the hydraulic pump configured as described above, when the input shaft 2 is rotated counterclockwise when viewed from the left side in FIG. 1, the cylinder block 4 is also rotated counterclockwise together with the input shaft 2. As a result, for example, bottom dead center B.
D. C. The shoe 16 connected to the tip of the plunger 12, which is located at the most expanded state, is connected to the tilted swash plate 6.
The shoe 16 and plunger 12 are pushed up by the swash plate 6 and moved in the direction of contraction. As a result, the oil chamber 1 defined by this plunger 12
3 is reduced in size, and the hydraulic oil in the oil chamber 13 is pressurized and discharged to the discharge port 14 of the distribution valve plate 7. This plunger 1
2 is top dead center T. D. C. When the shoe 16 is rotated to the maximum size, the discharge stroke is completed.
move in the direction of expansion. This allows suction port 15
Hydraulic oil is sucked into the oil chamber 13 from the inside.

As shown in FIG. 3, there is a gap between the discharge port 14 and the suction port 15 formed in the distribution valve plate 7.
A larger interval is provided than the connection port 13a. That is, the plunger 12 is at the bottom dead center B. D. C. , the connecting port 13a contacts the suction port 15 and separates from the discharge port 14, as shown by the two-dot chain line, and the plunger 12 reaches the top dead center T. D. C. The ports 14 and 15 are formed such that the connecting port 13a is in contact with the discharge port 14 and is located away from the suction port 15 when the valve is located at the position. Therefore, as the cylinder block 4 rotates, the plunger 12 moves to the bottom dead center B. D. C. When rotated from the connection port 13
Until port a faces the discharge port 14, the oil chamber 13 that this connection port 13a communicates with does not communicate with anything, and as the plunger 12 moves in the compression direction, the hydraulic oil in this oil chamber 13 is precompressed. (pressure increase). Similarly,
Plunger 12 is at top dead center T. D. C. When the connection port 13a is rotated from the state shown in FIG.
3 is in a state where it does not communicate with anything, and the hydraulic oil in this oil chamber 13 is pre-expanded (depressurized) by the movement of the plunger 12 in the expansion direction.

The relationship between the position of the plunger 12 (rotation angle of the cylinder block 4) and the oil pressure in the oil chamber 13 defined by the plunger 12 at this time is shown in FIGS. 4 and 5. Note that FIG. 5 shows the cylinder block 4 and the swash plate holder 8 as viewed in the direction of arrow II-II in FIG. In the figure, when the rotation angle θ is 0°, the bottom dead center B.
D. It shows how the oil pressure P in the oil chamber 13 defined by the plunger 12 located at C changes according to changes in the rotation angle. The period from when the rotation angle is 0° to when it rotates by θ1 is the pressure increase (precompression) section, and when the rotation angle is 1
Depressurization (pre-expansion) occurs from 80° until rotation by θ2
It is an interval. In this example, the oil pressure P gradually changes from the low pressure PL to the high pressure PH in the pressure increasing section, and the oil pressure P gradually changes from the high pressure PH to the low pressure PL in the pressure reducing section.

In this example, the rotation angle θ1 at which the pressure is increased and the rotation angle θ2 at which the pressure is reduced are equal, that is, θ
The discharge port 14 and the suction port 15 are formed so that 1=θ2 and the rotation angle θ3 from the start of pressure increase to the start of pressure reduction becomes θ3=180°. In the swash plate plunger type hydraulic pump of this example in which both ports 14 and 15 are formed in this way, the thrust loads F1 to F10 acting on each of the ten plungers 12 and these loads with respect to the rotation of the cylinder block 4 are Figure 6 shows the temporal change in the combined thrust load Ft. As can be seen from this figure, if θ1, θ2, and θ3 are set as above, for example, the load decreasing part of F1 and the load increasing part of F6 cancel each other out, and the fluctuation of the thrust composite load Ft becomes theoretical. becomes zero.

Therefore, in the case of the hydraulic pump of this example,
By reducing the excitation force caused by the thrust composite load Ft,
Vibration and noise from the hydraulic pump can be suppressed. In the case of a variable displacement type, when the tilting angle of the swash plate changes, the amount of pressure increase and the amount of pressure decrease change, but in the case of the pump of this example, θ1, θ2, and θ3 are All you need to do is set it as such, and it is not related to the tilt angle of the swash plate. Note that the rotation angle θ1 at which the pressure is increased and the rotation angle θ at which the pressure is decreased are
When considering the volumetric efficiency of the pump, the smaller the value of 2, the better.

By setting each of the rotation angles θ1, θ2, and θ3 as described above, it is possible to reduce fluctuations in the thrust composite load Ft. It may be difficult to make the oil pressure change in the section of angle θ2 gradual as shown in FIG. 4 or FIG. 5. For this purpose, as shown in FIG. 7, V-shaped grooves 14a and 15a are formed at the ends of the discharge port 14 and suction port 15 of the distribution valve plate 7', and the change as shown in FIG. 4 or FIG. may be obtained. Although not shown, this V-shaped groove 14a, 1
Instead of 5a, a hole or a valve may be provided to obtain the oil pressure change as shown in FIG.

Furthermore, in the examples shown in FIGS. 3 and 7, the connection port 13a is at the bottom dead center T. D. C. Or top dead center T. D. C. Pressure increase or decrease starts from the time when the cylinder block 4 rotates. However, the starting point of pressure increase and pressure reduction is not limited to this, and as shown in FIG. 8, the discharge port 14 and the suction port 15 may be formed. When both ports 14 and 15 are formed as shown in FIG. 8, the oil pressure P in the oil chamber 13 changes as shown in FIG. However, also in this case, the rotation angles θ1, θ2, and θ3 are set so as to satisfy the above-mentioned conditions of θ1=θ2, θ3=180°. It goes without saying that the pressure increase and pressure decrease start points may be shifted in the opposite direction to that shown in FIG. 8 with respect to the bottom and top dead centers.

Next, consider the moment acting around the trunnion shaft 8a. As shown in FIG. 10, depending on the rotation of the cylinder block 4, the bottom dead center B. D. C. When the plunger 12 is located at a position rotated by an angle θ from
And the moment M around the trunnion shaft 8a acting on the swash plate holder 8 is expressed by the following equation. M=F×R1×COSθ×SEC2α...(
1) Note that R1 is the length of the moment arm around the trunnion shaft 8a of the plunger 12, and as shown in FIG.
and bottom dead center B. D. C. Let R3 be the distance from the trunnion shaft 8a on the swash plate 6 when the plunger 12 is at a position rotated by an angle θ from the position R2=R1×SECα. Here, since R3=R2×COSθ, it is expressed as R3=R×COSθ×SECα. Moreover, since M=F×SECα×R3, the above moment M can be obtained. However, the angle α is the tilt angle of the swash plate 6. Also, the distance h from the swing center of the plunger 12 to the sliding surface of the swash plate 6 and the distance c from the center of the tilting support shaft (trunnion shaft 8a) to the sliding surface of the swash plate 6.
are set equal, and the center O1 of the tilting support shaft and the center O2 of the rotational movement pitch circle of the plunger 12 coincide on the sliding surface of the swash plate 6. The moment expressed by the above formula (1) is based on the pushing force F acting on one plunger 12, and if this is combined for all plungers 12, the total moment Mt acting around the trunnion shaft 8a can be obtained. Obtainable.

[0020]Here, for cases where the number Z of the plungers 12 is 10 and 12, the rotation angle θ1 at which the pressure is increased and the rotation angle θ2 at which the pressure is decreased are varied from 0° to 90°. The combined moment Mt at each angle when
The fluctuation rate ε was calculated. The results are shown in FIGS. 12 and 13, and as can be seen from these figures, the rotation angle θ1 at which the pressure is increased and the rotation angle θ2 at which the pressure is decreased are 360°.
/Z×k (where k=integer), the result shows that the fluctuation rate ε becomes minimum. For example, in the case of Z=10 lines, θ1=θ2=36°(k=1), 72°(
k = 2), the fluctuation rate ε becomes minimum at an angle k (integer) times 36°, and in the case of Z = 12, θ1 = θ2 =
30° (k = 1), 60° (k = 2), and 30° k (
The fluctuation rate ε becomes minimum at an angle that is multiplied by (an integer).

Therefore, in order to reduce the combined moment Mt, θ1=θ2=360°/Z×k (2) However,
Z: Number of plungers, even value k = 1, 2, 3
...(integer) and θ3=180° (3) It is understood that the discharge port 14 and the suction port 15 may be formed so as to satisfy the following equations. Note that the fluctuation rate ε of the combined moment Mt is determined as follows when the combined moment Mt fluctuates as shown in FIG. 11. ε={(ΣMt)max−(ΣMt)min}/(
ΣMt) mean×100 (%)

[0022]
In calculating the above-mentioned combined moment Mt, Fig. 14(A)
As shown in FIG. As shown in B), even if the center O1 of the trunnion shaft (tilting support shaft) is offset from the center O2 of the rotational movement pitch circle of the plunger 12, only the absolute value of the combined moment Mt is different, and its minimum value is The values of θ1 and θ2 that occur are the same as those in FIGS. 12 and 13.

As shown in equations (2) and (3) above, each rotation angle θ
By setting 1, θ2, and θ3, it is possible to reduce the fluctuation in the combined moment Mt. Figure 4 shows the oil pressure change in the section of rotation angle θ1 where the pressure is increased and the section of the rotation angle θ2 where the pressure is reduced. Alternatively, in order to achieve a gradual change as shown in FIG. 5, V-shaped grooves 14a are provided at the ends of the discharge port 14 and suction port 15 of the distribution valve plate 7', as shown in FIG. , 15a may be formed.

Although the swash plate plunger type hydraulic pump has been described above, this may also be used as a hydraulic motor. Further, although the above example shows a variable displacement pump in which the tilting angle of the swash plate is adjustable, the same applies to a fixed displacement pump.

Although the case where the hydraulic pump or hydraulic motor is used as a single unit has been described above, a hydraulic continuously variable transmission may be constructed by combining such a hydraulic pump and motor. An example of such a continuously variable transmission is shown in FIG. 15, in which a hydraulic pump P and a hydraulic motor M are arranged concentrically in a space surrounded by cases 20a to 20c. It is configured. An input shaft 21 of the hydraulic pump P is coupled to an output shaft of the engine.

The hydraulic pump P includes a pump cylinder 60 spline-coupled to the input shaft 21, and a plurality of cylinder holes 61 formed in the pump cylinder 60 at equal intervals on the circumference.
The pump plunger 62 has a plurality of pump plungers 62 that are slidably engaged with each other, and is rotationally driven by engine power transmitted through the input shaft 21. The hydraulic motor M is a pump cylinder 60
It consists of a motor cylinder 70 provided surrounding the pump cylinder 60, and a plurality of motor plungers 72 that slide into a plurality of cylinder holes 71 formed at equal intervals on the circumference of the motor cylinder 70. It is possible to rotate relative to each other on the same axis.

The motor cylinder 70 is composed of first to fourth parts 70a to 70d that are aligned in the axial direction and are connected together. The first portion 70a is rotatably supported by the case 20b via a bearing 79a at its left end outer periphery, and its right inner surface is inclined with respect to the input shaft 1 to constitute a pump swash plate holder. A pump swash plate ring 63 is provided on the pump swash plate holder. The plurality of cylinder holes 71 are formed in the second portion 70b, and the third portion 70c has a distribution plate 80 in which oil passages to each cylinder hole 61, 71 are formed. Fourth part 7
0d is coupled with the third portion 70C, and the bearing 79b
It is rotatably supported by the case 5c via.

An annular pump shoe 64 is rotatably and slidably attached to the pump swash plate ring 63, and the pump shoe 64 and the pump plunger 62 are connected to each other via a connecting rod 65 so as to be able to swing freely to some extent. ing. Bevel gears 68a and 68b that mesh with each other are formed on the pump shoe 64 and the pump cylinder 60. For this reason,
When the pump cylinder 60 is rotationally driven from the input shaft 1, the pump shoe 64 is also rotationally driven, and the pump swash plate ring 6
The pump plunger 62 is reciprocated according to the inclination of 3.
Oil is taken in from the suction port and oil is discharged to the discharge port.

Further, the swash plate holder 73 facing each motor plunger 72 is swung by the cases 20a and 20b via a pair of trunnion shafts (swing shafts) 73a that protrude from both outer ends of the swash plate holder 73 in a direction perpendicular to the plane of the paper. Supported for free movement. A motor swash plate ring 73b is disposed on the surface of the swash plate holder 73 facing the motor plunger 72, and a motor shoe 74 is attached in sliding contact with the motor swash plate ring 73b. The motor shoe 74 is swingably connected to the end of each motor plunger 72. The swash plate holder 73 is connected to the piston rod 33 of the speed change servo unit 30 via a link member 39 at a position away from the trunnion shaft 73a, and the speed change servo unit 30 causes the piston rod 33 to move in the axial direction. When the swash plate member 73 is moved, the swash plate member 73 is swung around the trunnion shaft 73a. motor cylinder 7
The fourth portion 70d of 0 is formed hollow, and a fixed shaft 91 fixed to the pressure distribution board 18 is inserted into the center thereof. A distribution ring 100 is fluid-tightly fitted to the left end of the fixed shaft 91, and the left end surface of the distribution ring 100 in the axial direction is eccentric so that it can come into sliding contact with the distribution plate 80. This distribution ring 100 divides a hollow portion formed in the fourth portion 70d into a first oil passage La on the inside and a second oil passage Lb on the outside.

The detailed structure of the distribution board 80 and the fourth portion 70d is shown in FIG. 16, and will be described below with reference to this figure as well. The distribution panel 80 includes a pump discharge port 81.
A and a pump suction port 82a are bored, and a first oil passage La consisting of the cylinder hole 61 of the pump plunger 62 in the discharge stroke and the inner space is connected through the discharge port 81a and the discharge passage 81b connected thereto. The cylinder hole 61 of the pump plunger 62 in the suction stroke communicates with the second oil passage Lb, which is formed by the outer space, through the pump suction port 82a and the suction passage 82b connected thereto.
is communicated. Furthermore, a plurality of communication passages 83 (the same number as the plungers 72) are formed in the distribution panel 80, each communicating with the cylinder hole (cylinder chamber) 71 of each motor plunger 72. 100
Due to the action of the first
It communicates with the oil passage La or the second oil passage Lb. Therefore, the cylinder hole 7 of the motor plunger 72 is in the expansion stroke.
1 and the first oil passage La, and the cylinder hole 71 of the motor plunger 72 in the contraction stroke and the second oil passage Lb are communicated via the communication passage 83, respectively.

In this way, a hydraulic closed circuit is formed between the hydraulic pump P and the hydraulic motor M via the distribution panel 80 and the distribution ring 100. Therefore, input shaft 2
When the pump cylinder 60 is driven from 1, the high pressure hydraulic oil generated by the discharge stroke of the pump plunger 62 is
From the pump discharge port 81a to the pump discharge passage 81b, the first
The oil flows into the cylinder hole 71 of the motor plunger 72 in the expansion stroke through the oil passage La (inner space) and the communication passage 83 communicating therewith, and gives thrust to the motor plunger 72 . On the other hand, the hydraulic oil discharged by the motor plunger 72 in the contraction stroke is transferred to the pump in the suction stroke via the communication path 83 communicating with the second oil path Lb (outside space), the pump suction path 82b, and the pump suction port 82a. It flows into the cylinder hole 61 of the plunger 62.

Due to such circulation of the hydraulic oil, the reaction torque that the pump plunger 62 in the discharge stroke applies to the motor cylinder 70 via the pump swash plate ring 63, and the reaction torque that the motor plunger 72 in the expansion stroke receives from the motor swash plate member 73. The motor cylinder 70 is rotationally driven by the sum of the reaction torque and the reaction torque. The gear ratio i of the motor cylinder 70 to the pump cylinder 60 at this time is given by the following equation. Gear ratio i = (number of revolutions of pump cylinder 60) / (number of revolutions of motor cylinder 70) = 1 + (capacity of hydraulic motor M) / (capacity of hydraulic pump P) As can be seen from the above equation, the speed change servo unit 30 By swinging the swash plate member 73 and changing the capacity of the hydraulic motor M from 0 to a certain value, the gear ratio i can be changed from 1 (minimum value) to a certain necessary value (
It can be changed steplessly up to the maximum value).

In the hydraulic continuously variable transmission having such a configuration, the speed change ratio i is controlled by the speed change servo unit 30, for example, by setting a target engine speed Neo in accordance with the accelerator opening, and adjusting the actual engine speed. Control is performed so that the number NE matches the target engine speed Neo. For example, as shown in FIG.
If 4200 rpm is set as
Control is performed such that V decreases steplessly from the maximum value to the minimum value (=1.0), and the vehicle speed V gradually increases.

In this case, since the swash plate holder of the hydraulic pump P is constructed integrally with the motor cylinder 70, the rotation speed NP of the pump P (the rotation speed of the pump cylinder 60 relative to the swash plate ring 63) is theoretically As shown in Figure 18,
It changes continuously from 3600 rpm to 0 rpm. In this way, in the case of the hydraulic continuously variable transmission of this example, the rotation speed NP of the pump P changes significantly even though the engine rotation speed NE is constant. Naturally, the suction amount and discharge amount of the pump P also change significantly, and the pressure loss that occurs during suction and discharge also changes greatly. For this reason, the rotation angle θ1 at which the pressure is increased and the rotation angle θ2 at which the pressure is decreased also change, and cannot be kept constant.

Here, if the number of pump plungers 62 is an odd number, for example 9, the thrust composite load Ft is
It varies as shown in FIG. In this case, the fluctuation rate ε of the thrust load Ft is changed by changing the rotation angle θ1 for pressure increase and the rotation angle θ2 for pressure reduction from 0° to 90°,
It is shown in FIG. As can be seen from this figure, the rotation angle θ1 for pressure increase and the rotation angle θ2 for pressure reduction are 40
This fluctuation rate becomes zero when the rotation angles θ1 and θ2 change as described above.
The volatility also changes significantly. Therefore, when the number of pump plungers 62 is an odd number, there is a problem in that the thrust composite load Ft fluctuates greatly and the noise from the transmission increases.

However, in this example, the number of pump plungers 62 of the hydraulic pump P is 10 (an even number).
and θ1=θ2. In this case, even if the rotation angle θ1 at which the pressure is increased and the rotation angle θ2 at which the pressure is decreased change, the thrust composite load Ft does not fluctuate as shown in FIG. Therefore, in this transmission, noise caused by fluctuations in the thrust composite load Ft from the hydraulic pump P is reduced. Note that the hydraulic motor M of this hydraulic continuously variable transmission also has an even number (for example, 10) of motor plungers 72. Therefore, in this hydraulic motor M as well, fluctuations in the thrust composite load Ft do not occur, and noise caused by fluctuations in the thrust composite load Ft from the hydraulic motor M is also reduced. In particular, in the case of this hydraulic motor M, it is a variable capacity type,
Since the swash plate holder 73 is tiltable around the trunnion shaft 73a, as shown in FIG.
If only M is used, the displacement DM of the motor plunger 72 will also change according to the change in the gear ratio i, and the rotation angle θ at which the pressure will be increased will change.
1 and the change in the rotation angle θ2 under pressure reduction is greater than or equal to the pump P. Therefore, setting the number of motor plungers 72 to an even number is very effective for reducing noise.

Furthermore, when a hydraulic continuously variable transmission like the one in this example is used for a vehicle, during normal acceleration driving in which the input shaft 21 is driven by the engine, the above-mentioned behavior occurs, but during deceleration driving, the , the hydraulic motor M performs a pump action, the hydraulic pump P performs a motor action, and the engine brake works. Therefore, not only the oil pressure in the cylinder holes 61 and 71, but also the rotation angle θ1 at which the pressure is increased and the rotation angle θ2 at which the pressure is decreased are significantly different between when the vehicle is accelerating and when it is decelerating. Even in such a case, if the number of pump plungers 62 and motor plungers 72 is an even number, fluctuations in the thrust composite load can be suppressed and noise from the transmission can be reduced.

Furthermore, when a vehicle using a continuously variable transmission consisting of a swash plate plunger type hydraulic pump P and a hydraulic motor M is running at high speed, the motor plunger 72 rotates at a high speed. If the combined moment Mt due to the pushing force changes, this may act as an excitation force on the swash plate holder 73, and high-frequency noise may be generated from the transmission. In such a case, the port shape of the hydraulic motor should satisfy the above equations (2) and (3) to reduce the total moment M.
By suppressing the fluctuation of t, it is possible to suppress the generation of this high-frequency sound. Note that, as described above, the rotation angle θ1 at which the pressure is increased and the rotation angle θ2 at which the pressure is decreased vary depending on the rotation speed of the motor M and the tilt angle of the swash plate holder 73. The port shape is set so that both rotational angles θ1 and θ2 satisfy the above equations (2) and (3) at the rotational speed and tilting angle of the motor M when traveling at high speed. Specifically, for example, θ1=θ2=36°, θ3
=180°.

Increasing the rotation angle θ1 at which the pressure is increased and the rotation angle θ2 at which the pressure is decreased causes a decrease in the volumetric efficiency of the hydraulic pump and motor. However, in the case of a hydraulic transmission with the above configuration, at high speeds,
Since the swash plate 73a of the motor is located near the minimum inclination angle (the angle at which the gear ratio i=1), the ratio of hydraulic power transmission is small, and the reduction in power transmission efficiency of the transmission is small. Therefore, if the present invention is applied to a hydraulic continuously variable transmission as described above, the generation of high frequency noise can be effectively prevented without causing the problem of a decrease in power transmission efficiency.

[0040] This point will be explained in more detail. The ratio of hydraulic power transmission (hydraulic transmission rate) in the continuously variable transmission can be expressed as hydraulic transmission rate=1-(1/i). In addition, the ratio of mechanical power transmission (
Mechanical transmissibility) is 1/i. Here, there are 10 plungers, and θ1=θ
When the shape of each port is set so that 2=36°, the hydraulic power transmission efficiency decreases by about 6.5%. Therefore, in the case of a single pump as shown in FIG. 1, the overall power transmission efficiency is 93.5%.

However, in the case of the continuously variable transmission shown in FIG. 15, the overall power transmission efficiency η is as follows: η={
(Mechanical transmission rate) + (Hydraulic transmission rate) x 0.935} x 10
0 = {(1/i)+(1-1/i)×0.
935}×100. For this reason, for example, the gear ratio i
= 1.5, the overall power transmission efficiency η = 9
The efficiency is 7.8%, which means that the pump can be operated with higher efficiency than the pump alone. That is, it can be said that the configuration of the present invention has a great advantage in terms of efficiency when used in a hydraulic continuously variable transmission.

[0042]

[Effects of the Invention] As explained above, according to the present invention,
An even number of plungers is used, and the connection port is located between the inflow port and the outflow port, and the rotation angle θ1 receives pressure increase, the rotation angle θ2 receives pressure reduction, and the rotation from the start of pressure increase to the start of pressure reduction. Since the inflow and outflow ports are formed so that the angle θ3 is θ1 = θ2 and θ3 = 180°, the change in the hydraulic pressure in the cylinder hole corresponding to each plunger is moderated, and the change in the pressure of each plunger is It is possible to suppress the fluctuation of the combined thrust load that occurs due to the pushing force and the fluctuation of the combined moment around the tilting axis of the swash plate, reducing the excitation force on the swash plate and generating vibration and noise from the hydraulic system. can be suppressed. Further, θ1=θ2=180°/Z×kθ3=180°, where Z: the number of the plungers, an even value k=1,
By forming the inflow and outflow ports so that 2, 3... (integer), the fluctuation of the combined moment around the tilting axis of the swash plate that occurs due to the pushing force of each plunger can be effectively suppressed. can be suppressed,
It is possible to reduce the excitation force on the swash plate caused by this fluctuation, thereby suppressing the generation of vibration and noise from the hydraulic system. In addition, if the present invention is particularly applied to a hydraulic continuously variable transmission formed by combining a hydraulic pump and a hydraulic motor, noise generation from the transmission can be suppressed without significantly reducing power transmission efficiency.

[Brief explanation of the drawing]

FIG. 1 is a sectional view showing a swash plate plunger type hydraulic pump according to the present invention.

FIG. 2 is an end view of the hydraulic pump taken along arrows II-II.

FIG. 3 is an end view of the hydraulic pump taken along arrows III-III.

FIG. 4 is a graph showing changes in the oil pressure in the oil chamber as the cylinder block rotates in the hydraulic pump.

FIG. 5 is a graph showing oil pressure changes in the oil chamber and the positions of each port as the cylinder block rotates in the hydraulic pump.

FIG. 6 is a graph showing changes in the thrust load and thrust composite load acting on each plunger as the cylinder block rotates in the hydraulic pump.

FIG. 7 is a side view showing a different example of a distribution valve plate.

FIG. 8 is a side view showing another different example of the distribution valve plate.

9 is a graph showing changes in oil pressure in the oil chamber as the cylinder block rotates in the hydraulic pump using the distribution valve plate of FIG. 8. FIG.

FIG. 10 is a schematic diagram showing a moment around a tilting axis generated by a pushing force acting on a plunger in the hydraulic pump.

FIG. 11 is a graph showing changes in the combined moment Mt around the tilting axis.

FIG. 12 is a graph showing the relationship between the rotation angle θ1 subjected to pressure increase, the rotation angle θ2 subjected to pressure reduction, and the rate of variation of the combined moment.

FIG. 13 is a graph showing the relationship between the rotation angle θ1 subjected to pressure increase, the rotation angle θ2 subjected to pressure reduction, and the fluctuation rate of the combined moment.

FIG. 14 is a schematic diagram showing the positional relationship between the tilting axis center O1 on the swash plate and the rotational movement center O2 of the plunger.

FIG. 15 is a sectional view showing a hydraulic continuously variable transmission including a hydraulic pump and a motor according to the present invention.

FIG. 16 is a sectional view showing a part of the continuously variable transmission.

FIG. 17 is a graph showing the relationship among engine speed NE, vehicle speed V, and gear ratio i in control of the continuously variable transmission.

FIG. 18 is a graph showing the relationship between pump rotational speed NP and gear ratio i in the continuously variable transmission.

[Fig. 19] Rotation angle θ1 for pressure increase and rotation angle θ for pressure reduction in a hydraulic pump having an odd number of plungers.
2 is a graph showing the relationship between 2 and the fluctuation rate ε of the combined moment.

[Fig. 20] Motor rotation speed NM in the above continuously variable transmission,
It is a graph which shows the relationship between the displacement volume DM of a motor plunger, and the gear ratio i.

FIG. 21 is a graph showing changes in the thrust load and combined thrust load acting on each plunger as the cylinder block rotates in a hydraulic pump having an odd number of plungers.

[Explanation of symbols]

1 Casing 2 Input shaft 4 Cylinder block 6 Swash plate 7 Distribution valve plate 8 Swash plate holder 12 Plunger 14 Discharge port 15 Suction port

Claims (2)

[Claims] 1. A cylinder block having an even number of cylinder holes extending in the axial direction and opening at one end in the axial direction in an annular arrangement surrounding a rotating shaft; an even number of plungers arranged, a swash plate facing the one end of the cylinder block and on which an end of the plungers slides, and a swash plate arranged in sliding contact with a side surface of the other axial end of the cylinder block. In the swash plate plunger type hydraulic device having a distribution valve plate, an even number of connection ports communicating with each cylinder hole are connected to the rotating shaft on the side surface of the other end in the axial direction of the cylinder block. The distribution valve plate is arranged on a predetermined circumference around the center, and communicates with the cylinder hole in which the plunger in the expansion stroke is slid through the connection port in the distribution valve plate as the cylinder block rotates. an inflow port;
an outflow port is formed that communicates with the cylinder hole into which the plunger in the contraction stroke is slid, and the connection port is located between the inflow port and the outflow port in response to rotation of the cylinder block, The cylinder corresponding to an area in which the hydraulic pressure in this connecting port and the cylinder hole communicating with this connecting port is increased from the hydraulic pressure in the low-pressure side port of both the ports to the hydraulic pressure in the high-pressure side port. The rotation angle θ1 of the block, the rotation angle θ2 of the cylinder block corresponding to the section where the pressure is reduced from the hydraulic pressure in the high pressure side port to the hydraulic pressure in the low pressure side port, and the rotation angle θ2 from the start of the pressure increase. The rotation angle θ3 of the cylinder block up to the start of the pressure reduction is θ
A swash plate plunger type hydraulic device, wherein the inflow and outflow ports are formed so that 1=θ2 and θ3=180°. 2. A rotation angle θ1 of the cylinder block corresponding to the section where the pressure is increased and a rotation angle θ2 of the cylinder block corresponding to the section where the pressure is reduced are such that θ1=θ
2=360°/Z×k However, Z: the number of the plungers, and an even value k=1,
2. The swash plate plunger type hydraulic device according to claim 1, wherein the inflow and outflow ports are formed so that 2, 3... (integer).