JPH09112388A - Radial plunger pump - Google Patents

Abstract

PROBLEM TO BE SOLVED: To reduce the occurrence of wear or the damage of a slide member by a method wherein a cam ring to simultaneously lift plungers along with rotation of a rotor is arranged at the outer periphery of the plunger, and through slide of it over the inner wall of the rotor, the rotor is rotatably supported on a support member, and a receiving hole and a fluid passage are intercommunicated. SOLUTION: A cam ring 29 is provided to simultaneously lift plungers 22 along with rotation of a rotor 20 is arranged on the outer periphery of each plunger 22. A support member 31 is slid over the inner wall of the rotor 20 in which a slide hole 20a is formed and the rotor 20 is rotatable. Since six cam crests 29a are formed on the peripheral wall of the cam ring 29, each plunger 22 is reciprocated through a receiving hole 21 six times each time the rotor 20 makes one full turn. Further, the receiving hole 21 is communicated with a fluid passage 45. This constitution reduces the occurrence of wear or the damage of the members of the slide parts of the rotor 20 and the support member 31.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a radial plunger pump for pumping a fluid, and more particularly to a radial plunger pump for pumping fuel or oil to a high pressure.

[0002]

2. Description of the Related Art Conventionally, as a radial plunger type pump in which a plunger is housed in a housing hole radially provided in a radial direction of a rotor so as to be capable of reciprocating, and a cam ring is provided on the outer circumference of the plunger, is disclosed in Japanese Patent Laid-Open No. Those disclosed in JP-A-23556 and JP-A-60-249676 are known. In recent years, for example, in a high-pressure fuel supply pump used in a diesel engine or the like, due to demands such as improvement of combustion efficiency and reduction of impurities discharged in exhaust gas, 100 MPa or more, further 150 MPa
It has been required to pressurize the fuel to a high pressure reaching up to and send it under pressure.

[0003]

However, in the one disclosed in Japanese Patent Laid-Open No. 56-23556, since the rotor support position and the plunger accommodation position are different in the axial direction, the rotor plunger accommodation portion is This may cause whirling and the oil film between the rotor and the rotation support portion of the rotor may be cut off, causing the member to burn.

Further, in the one disclosed in Japanese Patent Application Laid-Open No. 60-249676, the rotor and the guide ring provided on the outer periphery of the rotor are eccentric, and the plunger housed in the rotor as the rotor rotates. Is a structure that pressurizes the fuel by sliding on the inner wall of the guide ring. Therefore, as the rotor rotates, part of the guide ring in the circumferential direction is subjected to high pressure and high load, and the other part is subjected to low pressure and low load, so that an eccentric load is applied to the guide ring. Therefore, there is a problem that the higher the pressure of the fuel to be pumped, the more uneven wear or damage the member has in the sliding portion between the rotor and the guide ring.

The present invention has been made to solve such a problem, and an object thereof is to provide a radial plunger pump suitable for high pressure.

[0006]

According to the radial plunger pump of the first aspect of the present invention, are the sliding holes radially provided in the radial direction of the rotor further provided at equal angular intervals in the circumferential direction of the rotor? Alternatively, each pair of sliding holes are provided on the radial straight line of the rotor, and the cam ring simultaneously lifts the plunger as the rotor rotates, so that the sum of the forces generated by the reciprocating movement of the plunger becomes almost zero. . Therefore, even if the pressurized pressure of the fluid becomes high, the eccentric load acting on the support member is small, so that the wear or damage of the member at the sliding portion between the rotor and the support member can be reduced.

Further, since the rotor rotates around the support member, the rotor supporting position and the plunger receiving position are substantially the same in the axial direction, so that whirling of the rotor can be suppressed. As a result, good rotation of the rotor can be maintained even if the pressure applied by the fluid becomes high. According to the radial plunger pump of claim 2 of the present invention, since the diameter of the support member is set to 12 mm or less, it is possible to suppress the expansion of the clearance of the sliding portion due to the high-pressure fluid entering the clearance between the support member and the rotor. Therefore, the fluid leakage from the sliding portion is reduced and the pumping efficiency of the pump is improved.

According to the radial plunger pump of the third aspect of the present invention, the feed pump is built-in, and the peak timing of the fluid supply amount by the feed pump and the peak timing of the fluid suction amount in the suction stroke substantially coincide with each other. The fluid supplied from the feed pump can be efficiently sucked. As a result, the feed pump can be made smaller and lighter in weight, so that the power consumption of the feed pump can be reduced.

According to the radial plunger pump of the fourth aspect of the present invention, since the fluid passage opens to the outer wall of the support member within the angular range of 1/4 to 3/4 of the suction stroke, Since the pressure loss at the communicating portion with the accommodation hole is reduced, a sufficient amount of fluid can be sucked in the suction stroke. Further, since the fluid can be sucked in efficiently, the fluid supply capacity of the feed pump can be reduced, so that the feed pump can be downsized.

According to the radial plunger pump of the fifth aspect of the present invention, the fluid passage is formed radially in the radial direction of the support member, and the fluid passage communicates with the accommodation hole in a straight line. Since the pressure loss at the communicating part with and becomes small, a sufficient amount of fluid can be sucked in during the suction stroke. Further, since the fluid can be sucked in efficiently, the fluid supply capacity of the feed pump can be reduced, so that the feed pump can be downsized.

According to the radial plunger pump of the sixth aspect of the present invention, by forming the fluid passages axially offset from each other, the number of fluid passages formed in the same cross section of the support member can be reduced. Therefore, the strength of the support member in the radial direction is improved. Further, if the diameter is the same, the diameter of the fluid passage can be increased, so that the pressure loss of the fluid can be reduced.

According to the radial plunger pump of claim 7 of the present invention, by forming the same number of openings of the fluid passage as the cam peaks of the cam ring, all the receiving holes and the openings of the fluid passage can be directly communicated with each other. is there. Therefore, the pressure loss in the communication portion between the fluid passage and the accommodation hole is reduced, so that a sufficient amount of fluid can be sucked in the suction stroke, and the pressure of the high-pressure fluid pressurized in the pumping stroke can be discharged without reducing as much as possible. You can Further, since the fluid can be sucked in efficiently, the fluid supply capacity of the feed pump can be reduced, so that the feed pump can be downsized.

[0013]

Embodiments of the present invention will be described below with reference to the drawings. (First Embodiment) First application of the radial plunger pump of the present invention to a high pressure fuel supply pump of a diesel engine.
Examples are shown in FIGS. 1, 2 and 3.

As shown in FIG. 2, the high-pressure fuel supply pump 1
The drive shaft 11 of 0 is rotationally driven from the engine by a belt (not shown), and is rotatably supported by the housing 12 via the bearing member 13 and the bearing 14. The feed pump 16 is arranged in the housing 12 around the drive shaft 11. As shown in FIG. 3, the feed pump 1
The rotor 17 of No. 6 is connected to the drive shaft 11 by the connecting pin 11a and rotates together with the drive shaft 11. The vanes 18 are radially arranged on the outer wall of the rotor 17,
The cylindrical support member 19 surrounding the outer periphery of the rotor and the rotor 17 are eccentric. When the vane 18 rotates while sliding on the inner peripheral wall of the support member 19 due to the centrifugal force as the drive shaft 11 rotates, the fuel sucked from the fuel passage 41 is pressurized, and this fuel is pressurized from the fuel passage 42 to be described later. It is supplied to the chamber 25. The feed pump 16 is provided with a relief valve (not shown) that circulates the fuel to the suction side when the feed pressure becomes equal to or higher than a predetermined pressure, for example, 2 MPa.

As shown in FIG. 2, the rotor 20 has a drive shaft 1
1 and rotates to form a sliding hole 20a in the axial direction. The support member 31 is the rotor 20 that forms the sliding hole 20a.
Of the rotor 20 and rotatably supports the rotor 20. As shown in FIG. 1, the six accommodating holes 21 are formed in the rotor 20.
Are radially formed at equal angular intervals on the radial line of the support member 31. The accommodation hole 21 can communicate with the fuel passage 43 through a fuel port 45 as a fluid passage, and accommodates the plunger 22 in a reciprocating manner. As described above, since the supporting position of the rotor 20 and the housing position of the plunger 22 are substantially the same in the axial direction, whirling of the rotor 20 can be prevented, so that the oil film between the rotor 20 and the supporting member 31 is maintained and the sliding portion is maintained. It is possible to suppress image sticking.

The radially inner end surface of the plunger 22, the inner peripheral wall of the rotor 20 forming the housing hole 21, and the support member 31.
A pressurizing chamber 25 is formed by the outer peripheral wall of the. Pressurizing chamber 2
Reference numeral 5 denotes a part of the accommodation hole 21, which can communicate with the fuel port 45 as the rotor 20 rotates. A shoe 24 that rotatably supports the roller 23 is in contact with the radially outer end surface of the plunger 22, and the roller 23 can slide on the inner peripheral wall of the cam ring 29. Since the six cam ridges 29a are formed on the inner peripheral wall of the cam ring 29, each plunger 22 makes six reciprocations in the accommodation hole 21 every one rotation of the rotor. The guide 26 includes a support portion 26a and a flat plate portion 26b shown in FIG. 2, and rotates together with the rotor 20 to guide the shoe 24 so as to be capable of reciprocating.

The pump head 30 has a cylindrical support member 31. A fuel passage 43 is formed in the axial direction of the pump head 30, and a discharge passage 44 communicates with the fuel passage 43. The fuel passage 43 can communicate with the fuel passage 42 by opening a solenoid valve 60 described later. The fuel ports 45 communicate with the fuel passages 43 and are formed radially at equal angular intervals in the radial direction of the support member 31. As can be seen from the development view of the support member 31 shown in FIG.
Adjacent fuel ports 45 are formed to be displaced in the axial direction of the support member 31, and do not exist on the same cross section of the support member 31. As a result, the support member 3 excluding the fuel port 45
Since the cross-sectional area of 1 can be sufficiently secured, the supporting member 3
The mechanical strength of No. 1 can be maintained, and the diameter of the support member 31 can be reduced. Further, if the support member has the same diameter, the diameter of the fuel port can be increased as compared with the case where all the fuel ports are present in the same cross section, so that the pressure loss is reduced.

As shown in FIG. 2, the delivery valve 50 is connected to the discharge passage 44 and attached to the pump head 30. The valve member 51 of the delivery valve 50 is biased to the fuel upstream side by the spring 52 and has a check valve function. When the delivery valve 50 is opened by the pressure of the discharged fuel during the pressure feeding process, the high pressure fuel is supplied to the common rail (not shown).

The solenoid valve 60 is mounted on the pump head 30 coaxially with the fuel passage 43. The valve member 62 of the solenoid valve 60 is urged by the spring 63 in a direction to separate from the valve seat 61, and when the coil 64 is de-energized, the valve member 62 separates from the valve seat 61 and the solenoid valve 60 operates. Since the valve is opened, the fuel passage 42 and the fuel passage 43 communicate with each other. When the armature 65 is attracted by the magnetic force generated in the coil 64 when the coil 64 is energized, the valve member 62 is seated on the valve seat 61. As a result, the communication between the fuel passage 42 and the fuel passage 43 is cut off.

In the case of the high pressure fuel supply pump of the diesel engine as in this embodiment, the supply pressure of the pump is 20 MPa.
Since it reaches a high pressure of 150 MPa from the above, the sliding clearance needs to be as small as 1 to 2 μm so that fuel does not leak from the sliding portion between the rotor 20 and the supporting member 31. However, when the fuel enters the sliding clearance, the fuel pressure acts to spread the inner diameter of the rotor 20 to expand the clearance. When the inner diameter is D, the pressure is P, the elastic coefficient of the material of the rotor 20 (steel material in this embodiment) is E, and the Poisson's ratio is ν, the clearance increase amount δ1 is approximately expressed by the following equation (1). . However, the rotor 20
The outer diameter d is larger than the inner diameter D of the rotor 20 so that d ≧ 2 × D is satisfied.

Δ1≈ (1 + ν) × D × P / (2 × E) (1) As can be seen from the equation (1), the clearance increase amount δ1 changes even if the outer diameter d of the rotor 20 is increased. However, the increase amount δ1 is proportional to the inner diameter D. In the case of this embodiment, D = 10 mm, P = 1
At 50 MPa, δ1 = 5 μm, but D = 14 mm
Then, δ1 = 8 μm.

On the other hand, the leakage amount q from the clearance is calculated by the following equation.
As shown in (2), it is proportional to the cube of the spread amount δ of the clearance. q = (π × D × δ ³ × P) / (12 × μ × L) (2) π: Circumferential ratio, μ: Viscosity coefficient of fuel, L: Axial length of clearance portion δ is pressure It is the sum of the initial clearance δ0 when P≈0 and the clearance increase δ1 due to the pressure increase, and is expressed by the following equation (3).

Δ = δ0 + δ1 (3) From equations (1), (2) and (3), the leakage amount q is expressed as a quartic function of D. When the pressure P increases and δ0 << δ1,
By eliminating δ from the equations (1), (2) and (3), the leakage amount q is expressed by the following equation (4). q = π × D ⁴ × ((1 + ν) / (2 × E)) ³ × P ⁴ / (12 × μ × L) (4) Leakage amount q from the clearance is proportional to the fourth power of the inner diameter D Therefore, the inner diameter D is an important design parameter that affects the volumetric efficiency.

Next, the relationship between the inner diameter D and the leak amount q is shown in FIG. As can be seen from FIG. 5, when the inner diameter is large (D = 14
mm), the amount of pumping is drastically reduced because the amount of leak increases rapidly with the increase of pressure in addition to the dead volume loss.
At around 50 MPa, the pumping amount becomes 1/3 or less of the geometrical pumping amount. However, when the inner diameter is small (D = 10 mm), more than half the geometrical pumping amount can be pumped even at around 150 MPa. The diameter of the support member 31 is substantially equal to the inner diameter D of the rotor 20, and if the diameter of the support member 31 is 12 mm or less, the increase amount δ1 of the clearance can be suppressed. It can be reduced to the extent that it does not exist.

Next, the operation of the high pressure fuel supply pump 10 will be described. (1) Intake stroke The rotation angle of the rotor 20 in the state shown in FIG. 1 is defined as 0 °. In this state, the plunger 22 is moving inward in the radial direction and is in the middle of the pressure feeding stroke. When the rotor 20 rotates and the central axis of the suction hole 21 reaches the position marked as 15 ° in FIG. 1, the lift amount of the plunger 22 becomes maximum and the pumping stroke is switched to the suction stroke. When the plunger 22 is moving radially outward as the rotor 20 rotates, the coil 64 is de-energized. Since the valve member 62 is separated from the valve seat 61 when the coil 64 is de-energized, the fuel passage 42 and the fuel passage 43 communicate with each other. When the pressure inside the pressurizing chamber 25 decreases due to the plunger 22 moving radially outward and the volume of the pressurizing chamber 25 increasing, the fuel supplied from the feed pump 16 is supplied to the fuel passage 42, the fuel passage 43, and the fuel. It is sucked into the pressurizing chamber 25 through the port 45. When the rotor 20 further rotates and the central axis of the suction hole 21 reaches the position marked 30 ° in FIG. 1, the suction flow rate into the pressurizing chamber 25 becomes the maximum amount. In the first embodiment, the central axis of the suction hole 21 is 15 ° to 45 ° in FIG.
When it is in the range of °, it corresponds to the intake stroke.

The feed pump 16 has six vanes 18
The fuel flow rate supplied from the feed pump 16 to the fuel passage 42 per rotation of the feed pump 16 has a pulsating waveform having six peaks as shown by 101 in FIG. On the other hand, the flow rate of fuel drawn into the pressurizing chamber 25 has an intermittent waveform shown by the solid line 102 in FIG.
The dotted line 102 indicates the flow rate of fuel discharged from the pressurizing chamber 25. The amount of fuel supplied from the feed pump 16 is set so as to reach the maximum amount when the rotor 20 reaches the rotational position of 30 ° shown in FIG. 1, and at this time, the central axes of the fuel ports 45 are respectively set. The support member 31 is set so as to be aligned with the central axis of the corresponding suction hole 21. Therefore, since the pressure loss becomes the smallest at the peak timing of the fuel supply amount from the feed pump 16, the fuel intake amount into the pressurizing chamber 25 at the peak timing of the fuel supply amount from the feed pump 16 as shown in FIG. By maximizing, the fuel supplied from the feed pump 16 can be efficiently sucked into the pressurizing chamber 25. Therefore, the size of the feed pump 16 can be reduced and the weight thereof can be reduced, so that the manufacturing cost of the feed pump 16 can be reduced. Further, the driving force generated by the drive shaft 11 can reduce the amount of power consumed by the feed pump 16.

The range of the rotation angle of the rotor 20 in which the suction flow rate into the pressurizing chamber 25 reaches a peak during the suction stroke is the forward and backward suction strokes centered on the center of the suction stroke due to the geometrical restrictions of the cam profile of the cam ring 29. 1/4 to 3/4
It is often in the range of. Therefore, if the set angle of the fuel port 45 is within this range, it is practically acceptable.
In the first embodiment, the intake stroke corresponds to the rotation angle of the rotor 20 in the range of 15 ° to 45 °, so that the center axis of the fuel port 45 is set within the range of 30 ° ± 7.5 °. Good.

(2) Pumping stroke With the coil 64 being de-energized, the plunger 22
When starts moving inward in the radial direction, the fuel passage 42 and the fuel passage 43 are still in communication with each other, so that the fuel in the pressurizing chamber 25 is discharged from the fuel passage 43 to the fuel passage 42, contrary to the time of intake. When the power supply to the coil 63 is turned on at an arbitrary timing when the plunger 22 is moving inward in the radial direction,
The communication between the fuel passage 42 and the fuel passage 43 is cut off, and the fuel in the pressurizing chamber 25 is pressurized. The pressurized fuel in the pressurizing chamber 25 pushes down the valve member 51 of the delivery valve 50 and is pressure-fed from the delivery valve 50 to a common rail (not shown). The flow rate of the fuel to be pumped can be controlled by adjusting the on-timing of the coil 64.

When the diameter of the plunger 22 is 5 mm and the pressure feeding pressure is 150 MPa, a large force of about 2500 N acts on each plunger 22 during the pressure feeding stroke. First
In the embodiment, each pair of accommodating holes 21 are formed at equal angular intervals in the circumferential direction of the support member 31 and on the radial straight line of the support member 31, and six cam peaks 29a of the cam ring 29 are provided at equal angular intervals. Since the plunger 22 is simultaneously lifted with the rotation of the rotor 20 due to the formation, the sum of the forces acting on the plunger 22 with the reciprocating movement of the plunger 22 becomes substantially zero. As a result, it is possible to prevent an unbalanced load from acting on the support member 31 due to the reciprocating movement of the plunger 22. Therefore, even if the pumping pressure becomes as high as 150 MPa, it is possible to reduce the wear or damage of the members in the rotor sliding portion.

(Second Embodiment) A second embodiment of the present invention is shown in FIG.
Shown in Components substantially the same as those in the first embodiment are denoted by the same reference numerals. Four accommodation holes 21 are radially formed in the rotor 70. Of the four accommodation holes 21, each pair of accommodation holes 21 is formed on the radial straight line of the rotor 70.
Although the accommodation holes 21 are not formed at equal angular intervals in the circumferential direction of the rotor 70, the four plungers 22 simultaneously move radially inward as the rotor 70 rotates. Therefore, as in the first embodiment, the sum of the forces acting on the plunger 22 becomes almost zero, so that the uneven wear and damage of the members in the rotor sliding portion can be reduced. Further, as in the first embodiment, six cam ridges 29a are formed on the inner circumference of the cam ring 29.
Therefore, each plunger 22 reciprocates 6 times for each rotation of the rotor 70.

(Third Embodiment) FIG. 8 shows a third embodiment of the present invention.
Shown in The feed pump 80 of the third embodiment is a trochoid feed pump. If the number of teeth of the inner gear 81 is 6 and the number of teeth of the outer gear 82 is 7, the inner gear 8
6 peaks of flow pulsation are generated per one rotation.

In the embodiment of the present invention described above, the radial plunger pump of the present invention is applied to the high pressure fuel supply pump of a diesel engine. However, the radial plunger pump of the present invention is used for the purpose of pumping a fluid to a high pressure. It can be used in any field, for example, as a pump for supplying hydraulic oil to high pressure.

In this embodiment, the plunger reciprocates 6 times for each rotation of the rotor. However, the number of reciprocating movements of the plunger reciprocating for each rotation of the rotor is not limited to 6 in the present invention. It can be selected according to the application.

[Brief description of the drawings]

1 is a sectional view taken along line II of FIG. 2 showing a high-pressure fuel supply pump according to a first embodiment of the present invention.

FIG. 2 is a vertical sectional view showing a high-pressure fuel supply pump according to the first embodiment of the present invention.

3 is a sectional view taken along line III-III of FIG. 2 showing a high pressure fuel supply pump according to a first embodiment of the present invention.

FIG. 4 is a development view of the support member of the first embodiment.

FIG. 5 is a characteristic diagram showing the relationship between pressure and pumping amount in the high-pressure fuel supply pump of the first embodiment.

FIG. 6 is a characteristic diagram showing a relationship between a fuel flow rate supplied from a feed pump and a fuel flow rate sucked into a pressurizing chamber in the first embodiment.

FIG. 7 is a cross sectional view of a plunger portion according to a second embodiment of the present invention.

FIG. 8 is a cross-sectional view of a feed pump portion according to a third embodiment of the present invention.

[Explanation of symbols]

10 High-Pressure Fuel Supply Pump (Radial Plunger Pump) 11 Drive Shaft 12 Housing 20 Rotor 20a Sliding Hole 21 Storage Hole 22 Plunger 25 Pressurizing Chamber 43 Fuel Passage 44 Discharge Passage 45 Fuel Port (Fluid Passage)

Front Page Continuation (72) Inventor Toru Takahashi 1-1, Showa-cho, Kariya City, Aichi Nihon Denso Co., Ltd.

Claims (7)

[Claims] 1. An accommodation hole is formed radially in the radial direction, and the accommodation holes are provided at equal angular intervals in the circumferential direction, or each pair of the accommodation holes are provided on a straight line in the radial direction, and A rotor having a communicating sliding hole formed in the axial direction, a plunger that is reciprocally housed in the housing hole and pressurizes the sucked fluid, and a plunger that is arranged on the outer circumference of the plunger and rotates with the rotor. A cam ring that simultaneously lifts the rotor, and a support member that slides on the inner wall of the rotor that forms the sliding hole to rotatably support the rotor and that has a fluid passage that can communicate with the housing hole. Radial plunger pump characterized by. 2. The diameter of the supporting member in the sliding hole is 1
The radial plunger pump according to claim 1, wherein the radial plunger pump has a diameter of 2 mm or less. 3. A feed pump capable of supplying a fluid to the accommodation hole is built-in, and a peak timing of a fluid flow rate supplied from the feed pump substantially coincides with a peak timing of a fluid flow rate sucked in the suction stroke. The radial plunger pump according to claim 1 or 2, which is characterized. 4. The fluid passageway is 1/4 to 3 of the suction stroke.
The radial plunger pump according to claim 1, wherein the radial plunger pump is open to the outer wall of the support member within an angle range of / 4. 5. The radial plunger pump according to claim 1, wherein the fluid passage is formed radially in the radial direction of the support member and communicates with the suction hole in a straight line. . 6. The radial plunger pump according to claim 5, wherein the fluid passages are formed so as to be offset from each other in the axial direction of the support member. 7. The radial plunger pump according to claim 1, wherein the fluid passage has the same number of openings as the cam ridges of the cam ring on the outer wall of the support member.