JP7703350B2 - Variable Displacement Hydraulic Pump - Google Patents

Description

The present invention relates to a variable displacement hydraulic pump suitable for installation on construction machinery such as wheel loaders, hydraulic excavators, and hydraulic cranes.

In recent years, efforts to reduce _CO2 emissions have been promoted to curb global warming, and efforts are also being made to improve the fuel efficiency of construction machinery.

Furthermore, as a variable displacement hydraulic pump that is mounted on construction machinery and the like and driven by an engine or electric motor drive source (also called a power source), a swash plate type pump is well known, as described in Patent Documents 1 and 2, for example.

As part of the efforts to improve fuel efficiency, measures are being taken to improve system efficiency in hydraulic pumps used in the above-mentioned construction machinery, such as minimizing the discharge flow rate of the hydraulic pump so as not to waste oil when the construction machinery is in a standby state where it is not working.

Patent No. 3154329 JP 2011-32883 A

In light of the above, from the perspective of hydraulic systems, the best way to reduce hydraulic losses and improve fuel efficiency is to set the minimum capacity of the hydraulic pump to zero. However, from the perspective of hydraulic pump reliability, setting the pump capacity to zero means that the stroke of the piston reciprocating inside the cylinder (hole) of the cylinder block rotates in a zero state, which can cause poor lubrication of the sliding parts of the piston and cylinder, and can lead to damage such as piston seizing and seizing. For this reason, hydraulic pumps used in construction machinery, etc. are designed so that the capacity at which the piston can maintain a lubricated state is set as the minimum capacity (for example, about 1/5 of the maximum capacity) and cannot be set below that.

Due to these circumstances, the hydraulic pump must continue to discharge oil unnecessarily even in standby mode unless the driving source such as the engine is stopped, which creates a problem of preventing improvements in the efficiency of the hydraulic system.

The present invention was made in consideration of these circumstances, and its purpose is to provide a variable displacement hydraulic pump that can improve reliability without causing poor lubrication, can set the minimum discharge capacity smaller than before, and can reduce no-load losses in the hydraulic system.

In order to achieve the above object, the variable displacement hydraulic pump according to the present invention basically comprises a casing, a rotating shaft rotatably mounted in the casing, a cylinder block mounted within the casing so as to rotate integrally with the rotating shaft, a plurality of cylinders mounted within the cylinder block and having cylinder ports for supply and discharge, pistons reciprocably inserted into each of the cylinders, and a valve plate mounted within the casing so as to face the cylinder block, and having suction ports and discharge ports intermittently communicating with each of the cylinders, with a pair of switching lands sandwiched between them, and the cylinder block is rotated in sliding contact with the valve plate to cause the pistons to reciprocate, whereby oil is sucked in from the suction ports via the cylinder ports and discharged from the discharge port via the cylinder ports. and a connecting groove connected to the conduit is provided in at least one of the switching lands of the valve plate, and a connecting groove is provided in a portion of the sliding surface of the valve plate that contacts the cylinder block and is off the track of the cylinder port. An oil groove is provided in the sliding surface of the cylinder block that contacts the piston of each of the cylinders, and oil holes that can be connected to the connecting grooves are provided in the sliding surface of the cylinder block that contacts the valve plate and corresponds to each of the cylinders, so that the connecting grooves and the oil grooves sequentially communicate with each other via the oil holes when the cylinder block rotates. The conduit and the connecting groove are provided only in one of the pair of switching lands of the valve plate, which is a section that switches from the discharge port to the suction port . The oil groove is formed in an annular shape in the sliding surface of the cylinder that contacts the piston and around the axis of the cylinder. The oil hole is provided so as to be inclined from the sliding surface of the cylinder block that contacts the valve plate with the valve plate, with respect to the axial direction of the cylinder .

In the variable displacement hydraulic pump of the present invention, high pressure oil in the cylinder is guided from the cylinder port through the conduit and the connecting groove to the oil hole in the cylinder block that passes directly above the connecting groove, and is then supplied to the oil groove via this oil hole. As a result, the oil film pressure in the sliding area between the piston and cylinder is higher near the oil groove than in conventional pumps, so oil film breakage is suppressed and the sliding area is well lubricated, and the area near the opening edge of the cylinder, where piston seizing and seizure are likely to occur, is also effectively lubricated.

As a result, the risk of seizure can be reduced, especially when the pump capacity is small, improving the reliability of the hydraulic pump. In addition, because the risk of seizure can be reduced, the minimum discharge capacity can be made smaller than that of conventional hydraulic pumps, which reduces the discharge of unnecessary oil even when the system is in standby mode, thereby reducing losses in the hydraulic system when it is unloaded.

In addition, construction machinery equipped with the variable displacement hydraulic pump according to the present invention can achieve improved reliability and fuel efficiency.

Problems, configurations, and advantages other than those described above will become clear from the description of the embodiments below.

1 is a hydraulic circuit diagram of an example of a wheel loader, which is one type of construction machine to which an embodiment of a variable displacement hydraulic pump according to the present invention is applied; 1 is an overall cross-sectional view of a first embodiment of a variable displacement hydraulic pump according to the present invention; FIG. 3 is a plan view of the valve plate shown in FIG. 2 . 3C-C cross-sectional view of FIG. FIG. 3 is a bottom view of the cylinder block shown in FIG. 2 . 5A-5A cross-sectional view. BB cross-sectional view of FIG. 6. FIG. 6 is a diagram for explaining the configuration and operation of the first embodiment, in which the bottom view (indicated by a broken line) of the cylinder block shown in FIG. 5 is superimposed on the plan view of the valve plate shown in FIG. 3 . FIG. 2 is a schematic diagram illustrating the cylinder, the piston, and the surrounding area thereof, for explaining the effects of the first embodiment; FIG. 11 is a plan view of a valve plate of a second embodiment of a variable displacement hydraulic pump according to the present invention. 11 is a cross-sectional view taken along the line CC in FIG. 10 .

The following describes an embodiment of the present invention with reference to the drawings.

Figure 1 is a hydraulic circuit diagram of an example of a wheel loader, which is one type of construction machine to which a first embodiment (reference numeral 2) and a second embodiment (reference numeral 10) of a variable displacement hydraulic pump according to the present invention are applied. Below, we will first explain the wheel loader 100.

In FIG. 1, the wheel loader 100 is equipped with an engine 1 as a power source, and this engine 1 drives a hydraulic pump 2 (a variable displacement hydraulic pump of the first embodiment). Pressurized oil discharged from the hydraulic pump 2 is guided to a hydraulic cylinder 4, which is an actuator, via a directional control valve 3. A main relief valve 6 is provided to limit the maximum pressure of the hydraulic pump 2. A pilot pump 7 is connected to the shaft of the hydraulic pump 2, and the discharge oil of the pilot pump 7 is supplied to an operating lever 8 and a pump regulator 9. A pressure reducing valve is built into the operating lever 8, which generates a secondary pressure proportional to the amount of operation in each operating direction. This secondary pressure is supplied to the pilot ports 3a and 3b of the directional control valve 3, and the opening of the directional control valve 3 is controlled by this secondary pressure. The above hydraulic circuit is an open circuit system in which the hydraulic pump 2 sucks up oil from the tank 5, and the discharged oil passes through the directional control valve 3 and returns to the tank 5 from the hydraulic cylinder 4 or the center bypass. The open circuit system is mainly used for circuits that drive front actuators.

On the other hand, the engine 1 drives the hydraulic pump 10 (variable displacement hydraulic pump of the second embodiment) via gears and the like at the same time as the hydraulic pump 2 of the open circuit system. The hydraulic pump 10 forms a closed circuit with the hydraulic motor 11, which is an actuator for traveling. In a closed circuit system, the hydraulic pump 10 and the hydraulic motor 11, which is an actuator, are directly connected, so the oil circulates within the circuit without returning to the tank 5. In this case, since the inside of the hydraulic pump 10 has a sliding part structurally, leakage loss occurs from this sliding part. As a result, the leakage loss increases as the pressure increases. For this reason, in a closed circuit system, the oil in the circuit becomes insufficient by the amount of leakage. Therefore, a charge pump 50 is provided in this closed circuit to compensate for the leakage loss in the hydraulic pump 10. Since this hydraulic pump 10 discharges from both of the two ports, the oil passage is narrow due to strength issues. Therefore, due to its poor self-priming ability, it is operated at a charge pressure preset by the charge relief valve 51, and the charge oil is forced into the oil passage with lower pressure through the check valve 52 or 53. By configuring the hydraulic pump 10 and the hydraulic motor 11 for travelling in a closed circuit as described above, the forward and backward (forward and backward travel) and speed of the body 12 of the wheel loader 100 can be controlled by the tilt angle of the swash plate of the hydraulic pump 10.

[First embodiment of variable displacement hydraulic pump]
Next, the open circuit variable displacement hydraulic pump 2 of the first embodiment will be described with reference to the overall cross-sectional view of FIG.

The hydraulic pump 2 shown in FIG. 2 has a pump casing (also simply called the casing) 20 consisting of a hollow, bottomed cylindrical front casing 21 and a bottom-lid-shaped rear casing 22. A rotating shaft 25 is supported near both ends by bearings 21A and 22A provided at both ends of the pump casing 20, and a splined shaft portion 25a provided at one end (front side) of the rotating shaft 25 is connected to a power source. A thick, short cylindrical cylinder block 23 is fitted and fixed to the outside of the intermediate splined shaft portion 25b of the rotating shaft 25, so that the rotating shaft 25 and the cylinder block 23 rotate as a single unit.

In the cylinder block 23, a plurality of cylinders (holes) 24 (usually an odd number, for example, nine) are formed on a concentric circle (in other words, spaced apart in the circumferential direction of the cylinder block 23) centered on the rotation axis O of the rotary shaft 25 and parallel to the rotary shaft 25 at predetermined angular intervals (in other words, extending in the axial direction). A cylindrical piston 27 with a bottom and an opening on the rear side is inserted into each cylinder 24 so that it can slide back and forth.

The tip (front end) of the piston 27 protrudes from the cylinder 24, and a shoe 28 with a spherical joint is swingably connected to the tip (end protruding from the cylinder 24) of the piston 27. The spherical joint may be provided on the piston 27 side. The bushing 29 presses the shoe 28 against the smooth sliding surface 31c on the surface side of the swash plate 31, which is a flat surface, via the retainer 30 by a spring. A cylindrical convex surface portion 31b for tilting is formed on the back side of the sliding surface 31c of the swash plate 31, and a cradle 32 having a cylindrical concave surface portion 32b that slides into the cylindrical convex surface portion 31b is attached to the front casing 21 (i.e., the front side of the casing 20). The shapes of the cylindrical convex surface portion and the cylindrical concave surface portion may be reversed, that is, a concave cylindrical surface may be formed on the back side of the swash plate 31, and a convex cylindrical surface may be formed on the surface of the cradle 32 and fitted. The inside of the cylinder 24 is connected to the area between the shoe 28 and the sliding surface 31c of the swash plate 31 via an oil passage 39 formed in the center of the piston 27 and the spherical joint shoe 28.

The tilt angle of the swash plate 31 is controlled by pushing and pulling the swash plate 31 with a servo piston (not shown) as a tilt actuator. By changing the tilt angle of the swash plate 31, the stroke amount of the piston 27 can be changed. In other words, the pump capacity can be controlled. Since the hydraulic pump 2 of this embodiment is for an open circuit, the tilt angle of the swash plate 31 can be changed only in one direction (the direction in which the swash plate 31 tilts to one side with respect to the rotation axis O from when the tilt of the swash plate 31 is neutral (tilt angle 0 degrees), hereinafter sometimes referred to as the +α direction).

An elliptical cylinder port 24b is formed in a portion of the rear bottom of each cylinder 24 (the bottom opposite the piston 27 side) near the axis of rotation O, in other words, in a portion on the inner periphery of the sliding surface between the cylinder block 23 and the valve plate 34. As the piston 27 slides back and forth within the cylinder 24 (in other words, in response to the pushing and pulling of the piston 27 within the cylinder 24), oil is sucked in and discharged (also called suction and discharge) through the cylinder port 24b.

A valve plate 34 is fixed to the rear casing 22 (the rear side of the casing 20, the axial opposite side of the swash plate 31 across the cylinder block 23) with which the rear bottom surface 23b of the cylinder block 23 faces or slides. As can be seen from FIG. 2, FIG. 3, FIG. 4, and FIG. 8, the valve plate 34 has a pair of arc-shaped long holes (circumferentially long holes) on the left and right sides formed on the rotational orbit of the cylinder port 24b, and a discharge port 34a for discharging oil and a suction port 34b for sucking oil (in other words, the discharge port 34a and the suction port 34b that intermittently communicate with the cylinder port 24b of each cylinder 24) formed between a pair of switching lands 34td and 34bd. That is, the switching lands 34td and 34bd are the switching sections between the suction port 34b and the discharge port 34a. In addition, the rear casing 22 is provided with a discharge passage 22a connected to a discharge port 34a of the valve plate 34, and a suction passage 22b connected to a suction port 34b of the valve plate 34. Oil is drawn in and out of the discharge port 34a and the suction port 34b in response to the pushing and pulling of the piston 27. More specifically, the cylinder block 23 is rotated while being in sliding contact with the valve plate 34 to reciprocate the piston 27 in each cylinder 24, so that oil is sucked in from the suction port 34b through the cylinder port 24b and discharged from the discharge port 34a through the cylinder port 24b. The piston position when the piston 27 is pushed to the deepest part of the cylinder 24 (towards the valve plate 34) is called the top dead center (TDC). On the other hand, the position where the piston 27 is pulled out the furthest from the cylinder 24 (to the opposite side to the valve plate 34) is called the bottom dead center (BDC).

3, 4, and 8, a notch 40 is formed in the switching land 34bd between the suction port 34b and the discharge port 34a on the BDC side of the valve plate 34, which communicates with the discharge port 34a. In addition, a conduit 41 is formed in (near the center of) the switching land 34td between the discharge port 34a and the suction port 34b on the TDC side of the valve plate 34, which communicates with the casing drain. Here, when the cylinder port 24b switches from the discharge port 34a to the suction port 34b (while passing through the switching land 34td), the conduit 41 releases pressure to the casing drain through the orifice 54 and then communicates with the suction port 34b. The orifice 54 may be provided in either the valve plate 34 or the rear casing 22. By providing the orifice 54, it is possible to ensure that the pressurized oil is guided to the oil groove 46 through the communication groove 43 described later. In addition, the pressure of the pressurized oil guided to the oil groove 46 can be adjusted by changing the diameter of the orifice 54.

The conduit 41 is formed from a small diameter round hole extending (axially) on the back side of the valve plate 34, and communicates with a shallow communication groove 43 that is arc-shaped (long in the circumferential direction) in a plan view through an L-shaped communication passage 42 consisting of a horizontal hole extending radially outward whose outer end is blocked by a plug (lid) 44, and a vertical hole extending from the outer periphery of the horizontal hole to the front side of the valve plate 34. The communication groove 43 is provided within the sliding surface between the cylinder block 23 and the valve plate 34, and on the outer periphery outside the rotation orbit of the cylinder port 24b (see FIG. 8). In addition, the communication groove 43 in this embodiment is formed in an approximately semicircular arc shape from the outside of the switching land 34td where the conduit 41 and the communication passage 42 are provided to the entire outside of the discharge port 34a. This allows the supply of pressurized oil (piston lubricating oil) to the oil groove 46 continuously throughout the entire discharge stroke of the hydraulic pump 2, rather than only lubricating the piston near the conduit 41 (only near the switching land 34td) (explained later).

Next, the cylinder block 23 will be described with reference to Figures 5 to 8. Near each cylinder 24 in the sliding contact surface with the valve plate 34 of the cylinder block 23 (i.e., corresponding to each cylinder 24), oil holes 45 that can be connected face-to-face with the communication grooves 43 provided on the valve plate 34 side are provided. More specifically, the oil holes 45 are formed so that one end (the circular opening end on the valve plate 34 side) opens at a predetermined angular interval (here, 40° angular intervals) (i.e., a total of nine in this case) on the same circumference (the same circumference as the communication grooves 43) centered on the rotation axis O.

In addition, in each cylinder 24, an annular oil groove 46 slightly larger than the cylinder diameter is provided within a range (within the sliding cylindrical surface) where the piston 27 at maximum pump capacity always slides at both TDC and BDC. More specifically, the oil groove 46 is provided in an annular shape in a portion near the open end edge 24c (Figures 6 and 9) on the side where the piston 27 is pulled out of the cylinder 24. The nine oil holes 45 are connected to the oil groove 46 provided in each cylinder 24 (within the sliding cylindrical surface with the piston 27) via an oil passage 47 consisting of a diagonal horizontal hole whose outer end is blocked by a plug (lid) 48 and whose inner end is connected to the oil groove 46 (more specifically, the outer peripheral portion of the annular oil groove 46). In other words, the inner end of the oil passage 47 connecting the oil hole 45 and the oil groove 46 opens to the outer periphery of the annular oil groove 46 provided in the cylinder 24 (within the sliding cylindrical surface with the piston 27), and the oil groove 46 is connected to the oil hole 45 via the oil passage 47 at the outer periphery of the sliding cylindrical surface with the piston 27 in the cylinder 24.

Next, the basic operation of the hydraulic pump 2 of this embodiment will be described. When the cylinder block 23 is rotated by the rotating shaft 25 driven by the engine 1 as a power source, the oil in the cylinders 24 formed in the cylinder block 23 is pushed and pulled by the pistons 27 guided by the swash plate 31. At this time, the discharge port 34a provided in the valve plate 34 is connected to the side where the pistons 27 are pushed in, and the suction port 34b is connected to the side where the pistons 27 are pulled out.

The valve plate 34 switches between the inflow and outflow of oil, and this switching occurs near the TDC (switching land 34td) and BDC (switching land 34bd) of the piston 27. The pressure inside the cylinder 24 switches from low pressure to high pressure at BDC, and from high pressure to low pressure at TDC. At this time, a notch 40 is provided in the area (switching land 34bd) where the pressure inside the cylinder 24 switches from low pressure to high pressure near the BDC of the valve plate 34, so that the connection between the discharge port 34a and the cylinder port 24b is made gradually.

Meanwhile, near the TDC of the valve plate 34, a conduit 41 consisting of a small diameter round hole is provided in the section (switching land 34td) where the pressure inside the cylinder 24 switches from high pressure to low pressure. Since high pressure oil flowing into the suction port 34b can cause pulsation and noise, to avoid this, the high pressure oil is allowed to escape from the conduit 41 to the casing drain via the orifice 54. In addition, when high pressure oil flows into the conduit 41, it simultaneously flows into the communication groove 43 via the communication passage 42.

Here, the communication groove 43 is provided on the same circumference as the oil hole 45 of the cylinder block 23, so the communication groove 43 and the oil hole 45 communicate with each other for a preset period according to the circumferential length (arc length) of the communication groove 43. As a result, the high-pressure oil flowing into the communication groove 43 is guided to the oil groove 46 via the oil hole 45 and the oil passage 47. Here, the circumferential length of the communication groove 43 is set to be longer than the circumferential separation distance between adjacent oil holes 45 on the same circumference, and longer than the circumferential length of the discharge port 34a on the outside of the discharge port 34a. Therefore, the pressure oil is supplied from the communication groove 43 to each oil groove 46 through each oil hole 45 and oil passage 47 in sequence and without interruption (continuously) throughout the entire discharge stroke of the hydraulic pump 2.

More specifically, the high-pressure oil in the cylinder 24 near the top dead center of the discharge stroke is guided from the cylinder port 24b through the conduit 41 → communication passage 42 → communication groove 43 to the oil hole 45 that passes directly above the communication groove 43, and is supplied to the oil groove 46 through the oil passage 47 that is connected to this oil hole 45. In other words, when the cylinder block 23 rotates, the communication groove 43 and the oil groove 46 of each cylinder 24 are sequentially connected through each oil hole 45.

In this case, as shown in FIG. 9, in the case of a conventional product without the oil groove 46 (supply of pressurized oil thereto), the oil film pressure at the sliding portion between the piston 27 and the cylinder 24 is high on the rear side (valve plate 34 side) of the cylinder 24 and decreases linearly toward the front side (the side where the piston 27 is pulled out), as shown by the dashed line, and is almost zero near the opening edge 24c on the front side of the cylinder 24, where the piston 27 is likely to scorch and seize. In contrast, in the first embodiment, as shown by the solid line, the oil film pressure at the sliding portion between the piston 27 and the cylinder 24 is higher near the oil groove 46 than in the conventional product, so that the oil film break is suppressed and the lubrication state of the sliding portion is improved, and the vicinity of the opening edge 24c of the cylinder 24, where the piston 27 is likely to scorch and seize, is effectively lubricated.

Therefore, the risk of seizure can be reduced even when the pump capacity is small, improving the reliability of the hydraulic pump 2. In other words, when switching between discharge and suction of the hydraulic pump 2, the pressure oil released (discharged) from the conduit 41 provided in the switching land 34td on the TDC side (the section switching from the discharge port 34a to the suction port 34b) is supplied to the oil groove 46 provided in the sliding cylindrical surface with the piston 27 in the cylinder 24 and used to lubricate the sliding part between the piston 27 and the cylinder 24, improving the reliability of the hydraulic pump 2. In addition, since the risk of seizure can be reduced, the minimum discharge capacity (corresponding to the minimum tilt of the hydraulic pump 2) can be made smaller than that of a conventional hydraulic pump, so that unnecessary discharge of oil can be suppressed even in the standby state of the system, and as a result, loss during no load of the hydraulic system can be reduced.

In addition, in a wheel loader (construction machine) 100 equipped with the above-mentioned hydraulic pump 2, it is possible to improve reliability and fuel efficiency.

In the first embodiment, the circumferential length (arc length) of the communication groove 43 is set to be longer than the circumferential distance between adjacent oil holes 45 on the same circumference and longer than the circumferential length of the discharge port 34a, so that pressure oil is supplied to each oil groove 46 sequentially and continuously through each oil hole 45 throughout the entire discharge stroke of the hydraulic pump 2. However, this is not limited to such a configuration, and for example, the circumferential length of the communication groove 43 may be set shorter than the circumferential length of the discharge port 34a, and the circumferential length of the communication groove 43 may be set in accordance with the circumferential distance between adjacent oil holes 45 on the same circumference so that the communication groove 43 and any of the oil holes 45 are always in communication with each other, so that pressure oil is supplied to each oil groove 46 sequentially and continuously through each oil hole 45. Furthermore, the circumferential length of the communication groove 43 may be set shorter than the circumferential distance between adjacent oil holes 45 on the same circumference, in other words, the circumferential distance between adjacent oil holes 45 on the same circumference may be set longer than the circumferential length of the communication groove 43, so that pressurized oil is sequentially and intermittently supplied to each oil groove 46 via each oil hole 45 for a period corresponding to the circumferential length of the communication groove 43.

[Second embodiment of variable displacement hydraulic pump]
Next, the variable displacement hydraulic pump 10 for a closed circuit (the hydraulic pump at the bottom of Figure 1) of the second embodiment described above will be described with reference to Figures 10 and 11. This hydraulic pump 10 has the same basic configuration as the hydraulic pump 2 of the first embodiment described above (see Figures 2 to 9), but forms a closed circuit with the hydraulic motor 11, which is an actuator, and changes the tilt angle of the swash plate 31 in both directions (+α direction and -α direction) to change the discharge direction and capacity, thereby controlling the forward and backward travel (forward and backward travel) and speed of the vehicle body 12 of the wheel loader 100.

In such a closed circuit hydraulic pump 10, when the tilt angle of the swash plate 31 is changed in the +α direction (forward travel), similar to the hydraulic pump 2 of the first embodiment, the valve plate 34 shown in FIG. 3 has the discharge port 34a and the suction port 34b. On the other hand, when the tilt angle of the swash plate 31 is changed in the -α direction (the direction in which the swash plate 31 is tilted from the neutral position (tilt angle 0 degrees) to the other side with respect to the rotation axis O, and in the direction opposite to the +α direction) (reverse travel), the valve plate 34 shown in FIG. 3 has the suction port 34a and the discharge port 34b.

Therefore, in order to obtain the same effect as in the first embodiment regardless of the tilt direction of the swash plate 31 in the hydraulic pump 10, as shown in Figures 10 and 11, the conduit 41, the communication passage 42, and the communication groove 43 provided in one switching land 34td in the valve plate 34 may be provided in the other switching land 34bd, i.e., in both switching lands 34td and 34bd (180 degrees symmetrical). Similarly, the notch 40 that communicates with the discharge port 34a side in the switching land 34bd may also be formed on the suction port 34b side in the switching land 34td (180 degrees symmetrical).

By adopting such a configuration, the minimum discharge capacity can be made smaller than before even in a closed circuit system, and the unloaded losses of the hydraulic system can be reduced. In particular, in this closed circuit system, a charge circuit (such as a charge pump 50) that increases the suction side pressure is incorporated into the hydraulic system, making it possible to further reduce the minimum discharge capacity of the hydraulic pump 10.

In the above embodiment, an example was described in which the present invention was applied to a swash plate type hydraulic pump, but the present invention is not limited to swash plate type hydraulic pumps and can also be applied to bent axis type hydraulic pumps.

In addition, in the above-described embodiment, the hydraulic pump is applied to a wheel loader as an example, but the invention is not limited to this and may be applied to construction machines other than wheel loaders, such as hydraulic cranes and hydraulic excavators.

The present invention is not limited to the above-described embodiments, but includes various modified forms. For example, the above-described embodiments have been described in detail to clearly explain the present invention, and are not necessarily limited to those having all of the configurations described. It is also possible to replace part of the configuration of one embodiment with the configuration of another embodiment, and it is also possible to add the configuration of another embodiment to the configuration of one embodiment. It is also possible to add, delete, or replace part of the configuration of each embodiment with other configurations.

1: Engine 2: Variable displacement hydraulic pump (first embodiment)
3: Directional control valve 4: Hydraulic cylinder 5: Tank 6: Main relief valve 7: Pilot pump 8: Control lever 9: Pump regulator 10: Variable displacement hydraulic pump (second embodiment)
11: Hydraulic motor 12: Vehicle body 20: Pump casing 21: Front casing 22: Rear casing 22a: Discharge passage 22b: Suction passage 23: Cylinder block 23b: Rear bottom surface of cylinder block 24: Cylinder (hole)
24b: Cylinder port 24c: Opening edge portion 25: Rotating shaft 27: Piston 28: Shoe 29: Bushing 30: Retainer 31: Swash plate 32: Cradle 34: Valve plate 34a: Discharge port 34b: Suction port 34td, 34bd: Switching land 39: Oil passage 40: Notch 41: Conduit 42: Communication passage 43: Communication groove 44: Plug 45: Oil hole 46: Oil groove 47: Oil passage 48: Plug 50: Charge pump 51: Charge relief valve 52, 53: Check valve 54: Throttle 100: Wheel loader (construction machine)

Claims (5)

a cylinder block provided within the casing so as to rotate integrally with the rotary shaft; a plurality of cylinders provided within the cylinder block, each having a cylinder port for supplying and discharging; a piston reciprocally inserted within each of the cylinders; and a valve plate provided within the casing so as to face the cylinder block, the valve plate having an intake port and a discharge port intermittently communicating with each of the cylinders, the intake port and the discharge port being formed with a pair of switching lands therebetween,
a variable displacement hydraulic pump that rotates the cylinder block while in sliding contact with the valve plate to reciprocate the piston, thereby drawing oil from the suction port through the cylinder port and discharging it from the discharge port through the cylinder port,
a conduit consisting of a small diameter hole is provided in at least one of the switching lands of the valve plate, and a communication groove connected to the conduit is provided in a portion of the sliding surface of the valve plate that contacts the cylinder block and is off the track of the cylinder port,
an oil groove is provided in a sliding surface of the cylinder block with respect to the piston of each of the cylinders, and an oil hole capable of being connected to the communication groove is provided in a sliding surface of the cylinder block with respect to the valve plate, the oil hole corresponding to each of the cylinders being in face-to-face contact with the communication groove;
When the cylinder block rotates, the communication groove and the oil groove are sequentially communicated with each other via the oil hole,
the conduit and the communication groove are provided only in one of the pair of switching lands of the valve plate, the switching land being a section that switches from the discharge port to the suction port ;
the oil groove is formed annularly on a sliding surface of the cylinder with the piston and around the entire circumference of the axis of the cylinder,
4. A variable displacement hydraulic pump , comprising: a cylinder block having a first end that is in contact with a first valve plate; a valve plate that is in contact with a first oil hole ; 2. The variable displacement hydraulic pump according to claim 1,
a circumferential length of the communicating groove is longer than a circumferential length of the discharge port, and pressurized oil is supplied sequentially and uninterruptedly to the oil groove via the oil hole over the entire discharge stroke of the hydraulic pump. 2. The variable displacement hydraulic pump according to claim 1,
a circumferential length of the communicating groove is set in accordance with a circumferential separation distance between adjacent oil holes so that the communicating groove and any one of the oil holes are constantly in communication with each other, and pressurized oil is supplied sequentially and uninterruptedly to the oil grooves via the oil holes. 2. The variable displacement hydraulic pump according to claim 1,
2. A variable displacement hydraulic pump, comprising: a cylinder having a piston that is movable in a direction perpendicular to the axis of the cylinder; 2. The variable displacement hydraulic pump according to claim 1,
the oil hole is provided inclined with respect to an axial direction of the cylinder from a sliding contact surface of the cylinder block with the valve plate toward an outer periphery of the cylinder block,
a piston that slides in a direction perpendicular to the cylinder axis and moves in a direction perpendicular to the cylinder axis;

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