JPH02161175A - Ball piston pump - Google Patents

Abstract

PURPOSE:To prevent the eccentric abrasion of balls by specifying the difference between the max. and min. of the revolution radius of the ball which moves in reciprocation in the radial direction during revolution in the cooperation of a ring-shaped cam surface member and the formation position of a cylinder in a rotor for accommodating balls. CONSTITUTION:When a rotor 6 in a pump part 1 is revolved by a motor part 2, each ball 10 in a plurality of cylinders 9 formed in nearly radial form on the rotor 6 is shirted in the outer peripheral direction by a centrifugal force, and is revolved in contact with a cam face 13, and moves in reciprocation in the cylinder 9. Therefore, after the fluid sucked into a pump chamber 11 from a suction port 7a is pressurized, said fluid is discharged from a discharge port 8a. In this case, if the eccentricity quantity of the cam face 13 for the revolution center of the rotor 6 is represented by (e), the difference between the max and min of the revolution radius of the ball 10 is set 2e. Further, the distance between the axis line of the cylinder 9 and the normal line extending in the radial direction from the revolution center of the rotor 6 is set to 0.55-0.95e rearward from the revolution center of the rotor 6.

Description

[Detailed Description of the Invention] [Industrial Application Field] The present invention relates to a radial rotary piston pump that uses balls as pistons, and is particularly suitable for low-viscosity liquids such as kerosene and gasoline, and is power-saving and highly durable. The present invention relates to a ball piston pump, which is a type of positive displacement pump for liquids.

[Conventional technology]

Conventionally, fuel pumps for automobiles have often been made of roller vane pumps, Wesco type pumps, etc., with the focus being on small size and low cost. On the other hand, in recent years, with the increase in engine boost pressure and the research and development of in-cylinder fuel injection, there has been a need to improve fuel supply pressure.

However, conventional pumps have a large power loss due to low volumetric efficiency and high friction loss, and the actual situation is that the total efficiency of electric pumps remains at about 20%.

Here, in order to obtain a higher pressure than before, it not only impairs the durability of the pump itself but also leads to an increase in pump power consumption. However, in recent years, as vehicles have become more heavily equipped, the amount of electricity added has been increasing, and it is no longer acceptable for passenger cars (particularly light cars) to increase the power consumption of fuel pumps.

Therefore, it has become necessary to develop an innovative technology that can save power while maintaining a simple structure (low cost) with excellent durability.

This is why ball piston pumps are attracting attention. An example of such a ball piston pump will be described below, and the prior art and its problems will be clarified below.

Figures 10 to 12 show pumps based on this prior art.

In FIGS. 10 and 11, a cam face member 13 (also called a stator ring) is provided on the inner surface of the housing 3.
is provided. And this cam face member 13
Inside the rotor 6, nine cylinders each having a plurality of balls 10 around the rotor 6 rotate. This rotor 6 is rotated by a drive shaft 6b, and the drive shaft 6b is connected to a motor or the like (not shown).

5 is a valve shaft, and an intake port 7a and a discharge port 8a are provided inside thereof. The center of the valve shaft 5 and the inner peripheral surface of the cam face 8B material 13 are eccentrically assembled to each other.

In such a configuration, when the drive shaft 6b is rotated by a motor or the like, the rotor 6 rotates, and the balls IO in the cylinder 9, which are drilled radially in the rotor 6, contact the inner periphery of the cam face member 13, and the balls 10 wear occurs.

[Problem to be solved by the invention]

FIG. 12 is a schematic diagram illustrating the state of this wear. Since the ball 10 rotates in the cylinder 9 in the rotor 6 while being pressed against point b by the force indicated by the arrow F2,
The inner surfaces of the ball 10 and the cylinder 9 are unevenly worn as shown by the broken line in FIG. 12, and as a result, the ball 10 is deformed into a rugby ball shape. This indicates that specific parts are worn intensively, and as the inner surfaces of the ball 10 and cylinder 9 become rough due to wear, specific parts become more likely to wear, and the amount of wear is quickly reached to an acceptable level. This will reduce pump efficiency.

In order to reduce such problems, the first thing to do is to reduce the force with which the ball is pressed against the inner wall of the cylinder, thereby reducing friction loss, and it is an object of the present invention to achieve this.

Furthermore, in order to further reduce the above problems and provide a durable pump, it is effective to reduce the uneven wear of the balls, and the embodiments of the present invention aim to achieve this as well. do.

[Means to solve the problem]

To this end, the invention provides: a valve shaft having a suction port and a discharge port; a rotor rotating around the valve shaft; a plurality of cylinders bored substantially radially within the rotor; The valve shaft includes a ball inserted into the cylinder so that it can move in the axial direction and rotate on its axis, and a ring-shaped cam face member located on the outer periphery of the ball to regulate the movement of the ball. The ball's revolution movement with respect to the axis is regulated by the cam face member, the difference between the maximum and minimum revolution radii is 28, and the axis of the cylinder is a direction extending in the radial direction from the rotation center of the rotor. parallel to the line, and the distance between the axis of the cylinder and the normal line is 0.
55 e to 0.95 e.

[Effect]

In the following, the operation of the present invention will be explained using drawings of an embodiment in order to make it easier to understand. At other times during the rotation, it also becomes F2 in Fig. 5.Then, by setting the offset δ, the pressure angle α and the rake angle β in Figs. 4 and 5 change from time to time depending on the rotational phase of the rotor. Therefore, the friction loss of the ball 10 at this moment-by-moment position always changes.Then, the cumulative value of the friction loss that one ball 10 follows during the discharge stroke is calculated by considering the offset δ and the friction coefficient μ. The result is as shown in Fig. 9. This Fig. 9 shows the change in friction loss ratio by changing the offset δ, and the horizontal axis shows the offset δ.
The ratio of the amount of eccentricity e to the amount of eccentricity e is taken, and the vertical axis shows the magnitude of friction loss with respect to the theoretical amount of work.

Therefore, there is an offset amount that can minimize friction loss. In Figure 9, this offset amount is not expressed as an absolute value, but
In order to accommodate various pump dimensions, it is expressed as a relative value to the eccentricity e.

As is clear from this Figure 9, δ/e is 0.55 to 0.
．． In other words, δ should be in the range of 0, 55.
By setting e to 0.95 e, friction loss due to the ball being pressed against the cylinder wall can be reduced.

〔Effect of the invention〕

According to the present invention, even if the pump is used for a long time, the gap between the cylinder and the ball is unlikely to increase significantly, so the amount of leakage of the fluid being pumped by the pump, especially when used as a high-pressure pump. can be made into a more efficient pump.

Further, by reducing friction, it is possible to reduce the driving force.

In addition, in the embodiment, in order to reduce uneven wear of the ball, for example, the contact point between the ball and the cam face member is made not to coincide with the rotating surface of the rotor including the axis of the cylinder, so the ball is Rather than rotating around the axis of rotation, the ball rotates by changing the direction of the axis of rotation in all directions, which prevents only specific parts of the ball from being worn out, which has a synergistic effect with the effect of the present invention. This has the advantage of further delaying the progress of wear.

〔Example〕

Examples of the present invention will be described below. First, a first embodiment will be explained using FIG. 1 and FIG. 2.

Figure 1 shows a cross section of the rotor part of the pump, and the
The figure shows a cross section including the axis of rotation.

In FIG. 2, the pump is composed of a pump section 1 and a motor section 2, and the pump section 1 is rotationally driven by the motor section 2.

In the pump portion 1, a member 13 having a circular cam face (hereinafter simply referred to as a cam face) is provided on the inner surface of the housing 3 so as to be eccentric with respect to the axis of the valve shaft 5. The valve shaft 5 is provided integrally with the housing 3, and a rotor 6 is fitted therein. Valve shaft 5
Although the rotor 6 and the rotor 6 are fitted to each other freely, they are fitted precisely so that leakage of fluid can be suppressed to a minimum.

A suction passage 7 and a discharge passage 8 communicating with the outside are provided inside the valve shaft 5, and each passage is connected to the suction port 7.
It opens onto the cylindrical surface of the valve shaft 5 through the discharge port 8a and the discharge port 8a. The valve shaft 5 is in a fixed state regardless of the rotation of the rotor 6, and the discharge port 8a and suction port 7a in the valve shaft 5 are connected to the rotor 6.
It is provided so as to repeatedly communicate with and disconnect from any one of a plurality of cylinders, which will be described later, depending on the rotation of the cylinder.

A plurality of cylinders 9 are bored approximately radially in the rotor 6.
A pump chamber 11 defined by the cylinder 9 and the ball 10 communicates with the suction port 7a or the discharge port 8a via the suction/discharge port 12.

However, at the moment of switching, it is temporarily closed.

Each cylinder 9 has a ball 1 made of bearing steel.
0 is inserted. Although the ball IO is movable in the axial direction within the cylinder 9, it is inserted with precision so that leakage of fluid from the gap with the cylinder 9 can be suppressed to a minimum.

A ring-shaped cam face 13 is arranged on the outside of the ball 10 so as to guide the revolution of the ball 10. While the rotor 6 is rotating, the balls 10 revolve while contacting the cam face 13 due to centrifugal force, but the cam face 13 has an arcuate cross section that is slightly larger than the radius of the ball 10 in order to relieve the contact surface pressure. ing.

In addition, the contact point CP between the ball lO and the cam face 13 is
For reasons to be described later, the ball 10 is slightly second to the outermost periphery.
It is placed on the left side of the figure, away from the axis of the cylinder 9. In FIG. 1, 06 is the center of the rotor 6, and C13 is the center of the cam face 13.

The left end of the armature 15 of the motor section 2 is the bearing 1.
6, and the right end portion is coaxially inserted into the shaft end portion 6a of the rotor 6. The fitting portions of these two shafts are provided with a slight gap to allow for slight misalignment of the shafts, but have shaft ends and holes with oval cross sections to enable transmission of rotational force.

The motor section is composed of a DC motor with a magnetic field, and 30 is a brush, 31 is a commutator, 32 is a magnet, and 33 is an armature core.

A shielding plate 17 is provided on the housing 3 to separate the pump section 1 and the motor section 2, and a shielding plate 17 is provided around the shaft end 6a of the rotor 6 to prevent fluid from leaking from the pump section 1 to the motor section 2. A sealing member 18 is provided at. Next, FIG. 3 shows an enlarged view of the upper right part of the pump section 1 in FIG. 2. In FIG. 3, the contact point CP between the cam face 13 and the ball lO is at an angle λ with respect to the axis of the cylinder 9.
(contact angle λ).

Next, the operation of the pump of the first embodiment having the above structure will be explained.

When the rotor 6 rotates in the direction of arrow 20 in FIG.
Due to the centrifugal force, the ball 10 moves in the outer circumferential direction and revolves while contacting the cam face 13. Here, since the cam face 13 is eccentric by an eccentric amount e with respect to the rotation center of the rotor 6, that is, the axis of the valve shaft 5, the ball 10 moves with the rotation of the rotor 6. It reciprocates inside the cylinder 9.

In FIG. 1, when the rotor 6 rotates in the direction of the arrow 20, that is, in the counterclockwise direction, the balls 10 gradually move outward in the cylinder 9 in the upper half circumference, and the cylinder 9 and the balls 1
The spatial volume of the pump chamber 11 defined by 0 is increased. As a result, the inside of the pump chamber 11 becomes negative pressure, and the fluid is pumped into the pump chamber 11 from the suction port 7a through the suction/discharge hole 12.
to be introduced.

Next, in the upper half of FIG. 1, since the ball 10 moves inward within the cylinder 9, the space volume of the pump chamber 112 is gradually narrowed. As a result, the pressure inside the pump chamber 11 increases, and fluid is discharged through the suction/discharge hole 12).
Exhale to 8a.

The above is the basic operation of this pump.

While the rotor 6 is rotating, not only is centrifugal force acting on the ball 10, but also the ball 10 is constantly pressed against the cam face 13 because it receives hydraulic pressure from the pump chamber 11 in the upper half of the circumference. As a result, the ball IO is cam face 1
3, it revolves around the rotor 6 while rotating clockwise in the direction of the arrow 21. Therefore, the ball 10 and the cam face 13 undergo rolling motion, resulting in smooth motion with extremely little resistance.

However, since the ball 10 rotates within the cylinder 9,
Slip occurs between the cylinder 9 and the ball 10, and if the contact force at that part is large, it becomes a place where large friction loss occurs. The contact force referred to here is caused by the centrifugal force of the ball IO itself, the force that the ball 10 receives from the fluid due to the pressure in the pump chamber 11, and the like.

Here, the centrifugal force has a much smaller value than the force received from the fluid. Therefore, the main cause of the contact force between the ball IO and the cylinder 9 is the force that the ball 10 receives from the pressure of the fluid.

In FIG. 4, it is assumed that when the pressure inside the pump chamber 11 is P, the ball 10 is subjected to a force of a vector F in the axial direction of the cylinder 9. At this time, the ball 10 and the cam face 13
When the contact point with the cam face 13 is at point a in FIG.
The force of vector F1 is applied to , and at the same time, the force of vector F acts on point b' of cylinder 9.

Here, vector F2 is the difference between angle α (hereinafter referred to as pressure angle α) and angle β (hereinafter referred to as rake angle β) (
Hereinafter, it is a function based on the difference angle γ), and the vector F2 is approximately proportional to the difference angle γ.

In addition, FIG. 5° explains the behavior of the ball 10 during operation at a different time from that shown in FIG. 4, and in this FIG. When it comes to the opposite side of the axis, the direction of vector F8 is
This is the opposite of what happened in Figure 3, and the contact point also moves to the point.

At both times in FIGS. 4 and 5, the ball l
Since O rotates while being pressed against the cylinder 9 by the force of the vector F:, a frictional moment M opposite to the rotation direction is generated in the ball 10 so as to restrain the rotation. In other words, the cylinder 9, that is, the rotor 6 has this frictional moment M
This means that the ball 10 is moved forward in the direction of the arrow 20 by overcoming this.

Therefore, in the case of FIG. 5, that is, when the ball 10 contacts the cylinder 9 at the rear of the rotation of the rotor 6 (i.e., point b), the contact force is as shown by the vector F in FIG. has a further increased value, and the friction loss at point b in FIG. 5 becomes even greater than that at point b' in FIG.

An increase in friction loss is essentially an increase in rotational power, and is nothing but a harmful factor that accelerates wear of the balls 10 and cylinders 9.

On the other hand, in the case of FIG. 4, the ball lO is in contact with the rotor 6 at the rotational front (i.e., point b'), and the cylinder 9 moves in a direction away from the contact point b', so the actual contact force is vector The value is smaller than the size of F. This leads to a reduction in friction loss and also reduces the ball 10
This is a favorable condition for reducing the wear of the cylinder 9.

Further, as a result of more detailed analysis, it has been found that the relative sliding speed between the ball 10 and the cylinder 9 is smaller when the contact occurs at point b' in Figure 4 than at point b in Figure 5. . In other words, the friction loss, which is the product of frictional force and sliding speed, is much smaller in point b' contact than in point b contact.

FIG. 9 shows the results of analysis regarding this point.

The horizontal axis is the ratio of the offset δ to the eccentric file, the vertical axis is the ratio of the friction loss of the ball 10 to the theoretical work per rotation angle, and the friction coefficient μ between the ball lO and the cylinder 9 is plotted as a parameter.

This figure 9 shows that the friction loss is approximately proportional to the friction coefficient μ, and the friction loss becomes large when the offset δ is small or large. The reason is that when the offset δ is small, the discharge This is because the difference angle T becomes large near the middle of the stroke, and on the other hand, when the offset δ is large, the difference angle T becomes thick at the beginning and end of the discharge stroke. Material 1: Varies depending on surface treatment.

Also, from Fig. 9, the minimum value of friction loss is that if the friction coefficient μ is small, the friction loss will be small if the offset δ is set to a small value (on the other hand, if the friction coefficient μ is large, the offset δ This indicates that the minimum value of friction loss can be obtained by setting .

When the friction coefficient μ is large, the offset δ that provides the minimum loss shifts to a slightly larger value because the frictional moment M (FIG. 5) generated within the ball IO increases.

In other words, it is effective to reduce friction loss by reducing the area of contact at point b in Figure 5, which promotes frictional resistance, and by increasing the area of contact at point b' in Figure 4, which reduces frictional resistance. It means that.

Setting in this way means setting the value of the offset ratio δ/e within the recommended range RR of 0.55 to 0°95, as is clear from FIG.

Next, as shown in Fig. 3, the cam face 13 and the ball 10
The contact point CP with the cylinder 9 is at an angle λ(
Hereinafter, they are placed at separate positions with a contact angle λ).

As a result, the force that the ball 10 receives from the fluid during the discharge stroke in the upper half of FIG. 1, that is, the vector F in FIG.
Based on this, a force in the direction of the rotational axis, vector F, is generated.

This force becomes a force that pushes the ball 10 in the direction of the rotational axis of the rotor 6, that is, in the direction perpendicular to the plane of the paper in FIG. 1 or in the left-right direction in FIG. Further, this vector F is also a force in a direction perpendicular to the paper plane in FIGS. 4 and 5.

FIG. 6 is a schematic view of the ball 1o viewed from the direction of the arrow ■ in FIG. 2. As shown in FIG. , has the effect of moving the contact point in the direction of vector F2, that is, point b (or b') to point bx in FIG.

In other words, when the contact angle is λ-0, as shown in FIGS. 4 and 5, the direction of the contact point a and the contact force vector F2 is such that the axis of rotation of the ball 10 is within the rotational plane of the rotor 6. Always in the same direction. At this time, the rotation axis of the ball 10 is parallel to the rotation axis of the rotor 6, and the rotation direction of the ball 10 is opposite to the rotation direction of the rotor 6 (arrow 21 in FIG. 5).
direction).

However, as shown in FIG. 6, when the contact angle λ≠0, the direction of the contact force vector F is at the sixth
Ball 1
The rotation axis of 0 is undefined.

As a result, the balls 10 are moved irregularly within the cylinder 9, preventing uneven wear in only specific parts, and therefore improving the durability of the pump.

When the contact angle λ is small, the ball IO has little effect on causing irregular transfer.

On the other hand, if the contact angle λ is too large, the vector F4 in FIG. 3 increases, promoting rolling wear of the ball 10, cam face 13, etc. The vector F also increases, and this force becomes a thrust load on the rotor 6 as a whole, which is also undesirable.

Therefore, the contact angle λ is an element that should be determined depending on the usage conditions and design structure of each pump.

In the inventor's experience, this set angle λ is preferably chosen in the range from 2° to 45°, in particular in the range from 5″ to 30°.

In summary, in the above embodiment, the cylinder 9
The offset δ of is 0°55 to 0.9 with respect to the eccentricity e.
By setting the value within the range of 5, the friction loss can be reduced. Further, by setting the contact angle λ, uneven wear of the ball 10 and the cylinder 9 can be reduced.

Although these two effects each exert a great effect independently, a synergistic effect is produced when they are applied simultaneously.

These effects are most effective when applied to pumps that handle low-viscosity fluids such as kerosene and gasoline.

[Other Examples]

Next, a second example will be described.

In the first embodiment, in order to prevent uneven wear of the ball 10 and the cylinder 9, the contact point CP (FIG. 3) between the cam face 13 and the ball 10 is connected to the cylinder as a means of providing irregular rotation interlocking to the ball 10. 9 away from the axis. However, in order to obtain similar effects, the following second embodiment is constructed as shown in FIG. 7.

That is, the axis of the cylinder 9 is inclined at an angle of λ with respect to the axis of the pulp shaft 5, and the ball IO is brought into contact with the cam face 13 at the outermost periphery. As a result, it is possible to impart irregular rotational motion to the ball 1O as in the first embodiment.

In addition, the above 1. In the second embodiment, the reciprocating motion of the ball 10 within the cylinder 9 is one stroke per revolution, but it can also be applied to a pump that moves multiple strokes per revolution. In other words, the cam face 13 does not have to be a perfect circle, and various shapes can be selected.

Also, the above 1. Both pumps of the second embodiment are for supplying fuel to the gasoline engine of an automobile.

The overall description of this system is as follows.

FIG. 8 schematically shows an example of a fuel supply system for a gasoline engine most suitable for utilizing the pump 100 according to the present invention.

Gasoline pumped from a fuel tank 101 by a pump 100 is pressure regulated by a pressure pump 102, and is injected into a vehicle engine 105 from an injector 104 whose valve opening/closing timing is optimally controlled by an electronic control circuit 103.

Gasoline exceeding the amount required by the engine 105 is returned from the regulator 102 to the fuel tank 101 through the bypass line 106.

The pressure level regulated by the regulator 102 varies depending on the purpose, but the pump according to the present invention has a pressure level of 2 to 10
It regulates the pressure to about 100 bar and is effective in a wide range of fuel systems.

[Brief explanation of the drawing]

FIG. 1 is a partial sectional view taken along arrow 1-1 in FIG. 2 showing a first embodiment of the pump of the present invention, and FIG.
Fig. 3 is a schematic enlarged view of the ball contact part in Fig. 2, and Fig. 4 explains the movement of the ball during pump rotation in the first embodiment. A schematic explanatory diagram; FIG. 5 is a schematic explanatory diagram illustrating the movement of the ball at different times during the rotation of the pump in the above-mentioned first embodiment; FIG. 6 is a diagram showing the ball seen from the arrow ■ direction in FIG. FIG. 7 is a partial sectional view showing the same cross-section as FIG. 2 showing the second embodiment of the pump of the present invention, and FIG. FIG. 9 is a characteristic diagram showing the relationship between the offset ratio, friction loss ratio, and friction coefficient in the pump of the first embodiment, and FIGS. 10 and 11 are diagrams showing the conventional fuel supply system. arrows 316 showing the pumps of
316, a partial cross-sectional view taken along the direction Sat Sa+, and FIG. 12 are schematic explanatory views illustrating the movement of the ball of the conventional pump shown in FIG. 11. 7a... Suction port), 8a... Discharge port, 5...
- Valve shaft, 6... Rotor, 9... Cylinder, 10... Ball, 13... Cam face member.

Claims (5)

[Claims] (1) a valve shaft having a suction port and a discharge port; a rotor rotating around the valve shaft; a plurality of cylinders substantially radially bored within the rotor; The valve includes a ball inserted so as to be able to move in the axial direction of the cylinder and rotate, and a ring-shaped cam face member located on the outer periphery of the ball to regulate the movement of the ball, and the cam face member has a ring shape that controls the movement of the ball. The revolution motion of the ball as a reference is regulated by the cam face member, the difference between the maximum and minimum revolution radii is 2e, and the axis of the cylinder is relative to the normal line extending in the radial direction from the rotation center of the rotor. The ball piston pump is characterized in that the axis of the cylinder and the normal line are parallel to each other, and a distance between the axis and the normal line of the cylinder is set to 0.55e to 0.95e behind the rotation of the rotor. (2) The ball piston pump according to claim 1, wherein a contact point between the ball and the cam face member does not coincide with a rotational surface of the rotor that includes the axis of the cylinder. (3) The ball piston pump according to claim 1, wherein a contact point between the ball and the cam face member does not exist on the axis of the cylinder. (4) A point of contact between the ball and the cam face member is within an inclination range of 2° to 45° with respect to a rotational plane of the rotor that includes the axis of the cylinder. Ball piston pump as described. (5) The point of contact between the ball and the cam face member is on the rotating surface of the rotor, and the angle between the line connecting the point of contact and the center of the ball and the axis of the cylinder is 2 degrees or more. 4. The ball piston pump according to claim 3, wherein the inclination is set within a range of 45 degrees.