JP4789003B2 - Fuel pump - Google Patents

Description

The present invention, a plurality formed vane grooves in the rotating direction partition with blades, to fuel pumps for boosting by rotating the fuel pump passage formed along the vane grooves.

Conventionally, a plurality of blade grooves are formed in the rotation direction of the disk-shaped impeller, the blade grooves adjacent to each other in the rotation direction are partitioned by the blade plates, and the pump passage formed along the blade grooves by rotating the impeller There is known a fuel pump that boosts the pressure of the fuel (for example, see Patent Document 1). In such a fuel pump, when the blades adjacent to each other in the rotation direction are formed at equal angular intervals, as shown in FIG. 7 (A), as the impeller rotates, (total number of blades) × Noise having a high sound pressure peak at a frequency corresponding to (the rotational speed of the impeller) is generated.

Therefore, as in Patent Document 1, by at least a portion of an adjacent angle between the vane grooves adjacent in the direction of rotation (slats) sets the adjacent angles differently, as shown in FIG. 7 (B) It is conceivable to reduce the sound pressure peak by expanding the frequency band of the sound pressure peak.
By the way, as the impeller rotates, the fuel flows out and inflows one after another from the blade groove at the front in the rotation direction toward the blade groove at the rear in the rotation direction, so that the impeller boosts the pressure using the fuel as a swirl flow. In the structure of the impeller that boosts the fuel in this way, when the difference in the adjacent angle of the blades is large and the difference in the width in the rotation direction of the blade grooves partitioned by the blades is large, the amount of fuel flowing into the blade grooves The difference from the amount of fuel flowing out of the blade groove becomes large. Then, the fuel cannot be sufficiently boosted by the pump portion of the fuel pump that boosts the fuel by forming a swirling flow. As a result, the boosting efficiency of the pump unit is lowered, and the pump efficiency of the pump unit is lowered. The pump efficiency of the pump section increases if the adjacent angles of the blades are the same and the width of the blade grooves in the rotation direction is the same, but as described above, the sound pressure peak of the noise generated with the impeller rotation is high. Problem arises.

  Here, the fuel pump efficiency is expressed by (motor efficiency) × (pump efficiency). Therefore, when the pump efficiency is improved, the efficiency of the fuel pump is improved. The motor efficiency and the pump efficiency are: the drive current supplied to the motor part of the fuel pump is I, the applied voltage is V, the torque of the motor part is T, the rotational speed of the motor part is N, and the discharge pressure of the fuel discharged from the fuel pump Is P and the fuel discharge amount is Q, (motor efficiency) = (T × N) / (I × V), (pump efficiency) = (P × Q) / (T × N). Therefore, (fuel pump efficiency) = (motor efficiency) × (pump efficiency) = (P × Q) / (I × V).

Japanese Patent Laid-Open No. 11-50990

The present invention has been made to solve the above problems, reducing the sound pressure peaks of the noise, and an object of the invention to provide a suppressing fuel pump a decrease in pump efficiency.

In the first aspect of the present invention, the dispersion width, which is the difference between the maximum value and the minimum value of the adjacent angles formed by the rotation direction end portions of the adjacent blades, is 2.5 ° ≦ dispersion width ≦ 4 °. Is set to If the dispersion width is less than 2.5 °, the difference between the maximum value and the minimum value of the adjacent angle is small, so the frequency band of the sound pressure peak does not widen, and the sound pressure peak cannot be reduced sufficiently. If the dispersion width is greater than 4 °, the difference in the width of the blade groove in the rotational direction increases, and the difference between the amount of fuel flowing into the blade groove and the amount of fuel flowing out of the blade groove increases. Then, as described above, the fuel cannot sufficiently be boosted in the pump portion of the fuel pump that boosts the fuel by repeatedly flowing out and inflowing one after another from the front in the rotation direction toward the blade groove at the rear in the rotation direction. As a result, the boosting efficiency of the pump unit is lowered, and the pump efficiency of the pump unit is lowered.

Therefore, in the first aspect of the invention, by setting 2.5 ° ≦ dispersion width ≦ 4 °, it is possible to suppress a decrease in pump efficiency while reducing a sound pressure peak generated with the rotation of the impeller. ing.
Here, when the adjacent angle is less than 8 °, the width of the blade groove in the rotation direction is narrow and the volume is small, so that the swirling fuel cannot sufficiently flow into the blade groove. Therefore, it is difficult to increase the energy of the swirling flow. Further, if the adjacent angle is greater than 12 °, the width of the blade groove in the rotation direction is wide and the volume is large, so that it is difficult to cause the fuel flowing into the blade groove to flow out as a swirling flow and increase the energy of the swirling flow. is there. Thus, if the energy of the swirl flow does not increase, the fuel boosting efficiency decreases, and the pump efficiency decreases.

Therefore, in the first aspect of the invention, by setting the adjacent angles to 8 ° ≦ adjacent angle ≦ 12 °, the energy of the swirling flow generated by the rotation of the impeller can be increased, and the pump efficiency can be improved. it can.
In the invention described in claim 1, improves pumping efficiency of the fuel pump, with the efficiency of the fuel pump is improved as a result, it is possible to reduce the peak of the sound pressure fuel pump occurs.

Hereinafter, a plurality of embodiments of the present invention will be described with reference to the drawings.
(First embodiment)
A fuel pump using the impeller according to the first embodiment of the present invention is shown in FIG. The fuel pump 10 is an in-tank type turbine pump mounted in a fuel tank of a vehicle, for example, and supplies the fuel in the fuel tank to a fuel injection valve (not shown). The discharge pressure of the fuel pump 10 is set in the range of 0.25 to 1 MPa, the discharge amount is 50 to 250 L / h, and the rotation speed is 4000 to 12000 rpm.

The fuel pump 10 includes a pump unit 12 and a motor unit 13 that rotationally drives the pump unit 12. The housing 14 also serves as a housing for the pump unit 12 and the motor unit 13, and crimps the end cover 16 and the pump case 20.
The pump unit 12 is a turbine pump having pump cases 20 and 22 and an impeller 30. The pump case 22 is press-fitted into the housing 14 and abutted against the step portion 15 of the housing 14 in the axial direction. The pump cases 20 and 22 are case members that rotatably accommodate an impeller 30 as a rotating member. C-shaped pump passages 202 are respectively formed between the pump cases 20 and 22 and the impeller 30.

  As shown in FIG. 1, a plurality of blade grooves 36 are formed in the rotation direction on the outer peripheral edge portion of the impeller 30 formed in a disc shape, and the circumferential width of the blade grooves 36 is not uniform. As a result, the blade grooves 36 are arranged at unequal pitches in the rotation direction. The blade grooves 36 adjacent to each other in the rotation direction are partitioned by the blade plate 34. When the impeller 30 rotates together with the shaft 51 due to the rotation of the armature 50 shown in FIG. Flow into. By repeating such outflow and inflow of the fuel many times between the blade grooves 36, the fuel becomes a swirl flow 220 (see FIG. 4) and is pressurized in the pump passage 202 shown in FIG. The fuel sucked from the suction port (not shown) provided in the pump case 20 by the rotation of the impeller 30 is boosted in the pump passage 202 by the rotation of the impeller 30, and the motor unit 13 from the discharge port (not shown) provided in the pump case 22. Pumped to the side. The fuel pumped to the motor unit 13 side passes through the fuel passage 206 between the permanent magnet 40 and the armature 50 and is supplied to the engine side from the discharge port 210 provided in the end cover 16. The air vent hole 204 provided in the pump case 20 is for exhausting the air contained in the fuel in the pump passage 202 out of the fuel pump 10.

Four permanent magnets 40 formed in a quarter arc shape are attached to the inner peripheral wall of the housing 14 on the circumference. The permanent magnet 40 has four magnetic poles having different poles in the rotation direction.
Since the end of the armature 50 on the impeller 30 side is covered with the resin cover 70, the rotational resistance of the armature 50 is reduced. Further, a commutator 80 is assembled at the end of the armature 50 opposite to the impeller 30. A shaft 51 as a rotating shaft of the armature 50 is supported by a bearing member 24 housed and supported in the end cover 16 and the pump case 20.

  The armature 50 has a central core 52 at the center of rotation. The shaft 51 is press-fitted into a central core 52 formed in a cylindrical shape having a hexagonal cross section. The six magnetic pole cores 54 are installed on the outer periphery of the central core 52 in the rotational direction, and are fitted and coupled to the central core 52. A bobbin 60 formed of an insulating resin is fitted to the outer periphery of each magnetic pole core 54, and a coil 62 is formed by concentrating windings on the outer periphery of the bobbin 60.

  The end of each coil 62 on the commutator 80 side is electrically connected to the coil terminal 64. The coil terminal 64 corresponds to the rotational direction position of each coil 62 and is fitted and electrically connected to the terminal 84 on the commutator 80 side. The end of the coil 62 opposite to the commutator 80 on the side of the impeller 30 is electrically connected to the coil terminal 66. The six coil terminals 66 are electrically connected by an annular terminal 68.

  The commutator 80 is a cassette type integrally formed. When the shaft 51 is pressed into the central core 52 and the shaft 51 is inserted into the through hole 81 of the commutator 80 and the commutator 80 is assembled to the armature 50, the terminals projecting toward the armature 50 side of the commutator 80 84 are fitted to the coil terminals 64 of the armature 50 and are electrically connected to the coil terminals 64.

The commutator 80 has six segments 82 installed in the rotational direction. The segments 82 are made of, for example, carbon, and the segments 82 are electrically insulated from each other by a gap and an insulating resin material 86.
Each segment 82 is electrically connected to the terminal 84 via the intermediate terminal 83. The insulating resin material 86 is formed by integrating the segment 82 (excluding a sliding surface with a brush (not shown)), the intermediate terminal 83, and the terminal 84 by insert molding, thereby forming the commutator 80. As the commutator 80 rotates with the armature 50, each segment 82 sequentially contacts the brush. As the commutator 80 rotates and sequentially contacts the brush, the current supplied to the coil 62 is rectified. The permanent magnet 40, the armature 50, the commutator 80, and a brush (not shown) constitute a DC motor.

(Impeller 30)
The structure of the impeller 30 will be described in more detail.
The impeller 30 is integrally formed in a disc shape with resin. As shown in FIG. 1, the outer periphery of the impeller 30 is surrounded by an annular portion 32, and a blade groove 36 is formed on the inner peripheral side of the annular portion 32. As shown in FIG. 3, the blade grooves 36 adjacent to each other in the rotational direction are V-shaped blades that are inclined forward in the rotational direction from substantially the center in the thickness direction of the impeller 30 toward both end surfaces 31 in the thickness direction of the impeller 30. 34 is partitioned. As shown in FIG. 4, the blade groove 36 is partly divided on the radially inner side by a partition wall 35 protruding from the radially inner side of the blade groove 36 toward the radially outer side. It penetrates in the direction of the rotation axis on the outside in the radial direction. The fuel that has flowed into the blade groove 36 from the pump passages 202 on both sides in the axial direction becomes a swirl flow 220 that rotates in opposite directions on both sides in the rotation axis direction by the partition wall 35.

  As shown in FIG. 3, at least the radially inner side of the rear surface 37 at the rear of the blade groove 36 in the rotational direction is inclined rearward in the rotational direction from the radially inner side to the radially outer side. A line segment 110 connecting the radially inner end 37a and the radially outer end 37b of the rear surface 37 of the blade groove 36 is a straight line 104 extending radially outward from the radially inner end 37a on the radius 102 of the impeller 30. On the other hand, it is inclined backward in the rotational direction as it goes radially outward. That is, the rear surface 37 is inclined rearward in the rotational direction toward the radially outer side. In FIG. 3, reference numeral 100 indicates the rotation axis of the impeller 30. The radially inner end 37a and the radially outer end 37b of the rear surface 37 of the blade groove 36 are one end in the rotational direction of the blade plate 34, in this embodiment, the radially inner end at the front end in the rotational direction of the blade plate 34. 34a and radially outer end 34b.

Here, as shown in FIG. 1, the adjacent angle formed by the straight line 104 connecting the radially outer end 34 b that is one end portion of the blade plate 34 in the rotation direction and the rotation shaft 100 between the blade plates 34 adjacent to each other in the rotation direction. Is θ, the dispersion width (θ _max −θ _min ), which is the difference between the maximum value θ _max and the minimum value θ _min of the adjacent angle θ, is set to 2.5 ° ≦ dispersion width ≦ 4 °.

When the impeller 30 rotates, when the same adjacent angle of the blade plate 34, as shown in FIG. 7 (A), high sound pressure peak in the frequency corresponding to (slats total) × (rotating speed of the impeller) Noise is generated. And if the dispersion | distribution width of the adjacent angle of the blade board 34 is small, since the frequency band of a sound pressure peak does not spread, a sound pressure peak cannot be reduced like the case where an adjacent angle is equal. However, if the dispersion width of the adjacent angle of the blade plate 34 is small, the difference between the amount of fuel flowing into the blade groove 36 and the amount of fuel flowing out of the blade groove 36 becomes small. As a result, the boosting efficiency for boosting the fuel is increased by repeating the inflow into the blade groove 36 and the outflow from the blade groove 36 as the impeller 30 rotates, so that the pump efficiency of the pump unit 12 and the efficiency of the fuel pump 10 are increased. To rise.

On the other hand, the dispersion width of the adjacent angle of the blade plate 34 is increased, as shown in FIG. 7 (B), the frequency band of the sound pressure peak can be reduced dispersed sound pressure peak. However, if the dispersion width at the adjacent angle is large, the difference between the amount of fuel flowing into the blade groove 36 and the amount of fuel flowing out of the blade groove 36 increases. As a result, the pumping efficiency of the pump section 12 and the fuel pump are reduced because the pumping section 12 that boosts fuel decreases by repeatedly flowing into and out of the blade groove 36 as the impeller 30 rotates. The efficiency of 10 is reduced.

  Here, the characteristics of the dispersion width, the sound pressure peak, and the pump efficiency are shown in FIG. A graph 300 shows a relationship between the dispersion width and the sound pressure peak, and a graph 302 shows a relationship between the dispersion width and the pump efficiency. From the characteristics shown in FIG. 5, when the sound pressure peak is 135 dB or less and the adjacent angles of the blades 34 are all equal, that is, the pump efficiency is reduced to 1% or less compared to the optimum value when the dispersion width is 0 °. It can be seen that the range of the dispersion width is 2.5 ° ≦ dispersion width ≦ 4 °. Thus, by setting the dispersion width in the range of 2.5 ° ≦ dispersion width ≦ 4 °, it is possible to reduce the sound pressure peak and suppress the pump efficiency from decreasing.

Further, not only the dispersion width but also the size of the adjacent angle itself affects the pump efficiency as shown in the graph 310 of FIG. FIG. 6 shows the relationship between the adjacent angle and the pump efficiency when the adjacent angles of all the blades of the impeller are made equal.
If the adjacent angle is less than 8 °, the width of the blade groove 36 in the rotational direction is narrow and the volume is small, so that the swirling fuel cannot sufficiently flow into the blade groove 36. Therefore, it is difficult to increase the energy of the swirling flow. Further, when the adjacent angle> 12 °, the width of the blade groove 36 in the rotation direction is wide and the volume increases, so that the fuel flowing into the blade groove 36 is discharged as a swirl flow, and the energy of the swirl flow is increased. Is difficult. Thus, if the energy of the swirl flow does not increase, the fuel boosting efficiency decreases, and the pump efficiency decreases.

  On the other hand, if the adjacent angle is set in the range of 8 ° ≦ adjacent angle ≦ 12 ° in the impeller in which the adjacent angles of all the blades are equal, the reduction in pump efficiency is optimal as shown in FIG. It is suppressed to a range of 1% or less of the value. Accordingly, even when the adjacent angles of the blades 34 are made non-uniform as in this embodiment, the non-uniform adjacent angle range is set to 8 ° ≦ adjacent angle ≦ 12 °, respectively. In the range of 0.5 ° ≦ dispersion width ≦ 4 °, the decrease in pump efficiency can be easily suppressed to a range of 1% or less of the optimum value.

  As described above, in the first embodiment, the impeller is set to 2.5 ° ≦ dispersion width ≦ 4 ° by setting the dispersion width in which the adjacent angles of the blades 34 adjacent to each other in the rotation direction to be nonuniform. While reducing the sound pressure peak of the noise generated with the rotation of 30, it is possible to suppress the decrease in pump efficiency of the pump unit 12 of the fuel pump 10 as much as possible.

  In the first embodiment, the annular portion 32 covers the radially outer side of the blade groove 36, and no pump passage is formed on the outer peripheral side of the impeller 30. As a result, the differential pressure in the rotational direction of the fuel pressure boosted in the pump passage 202 is not directly applied in the radial direction of the impeller 30, so that the force applied to the impeller 30 in the radial direction is reduced. Thereby, since it can prevent that the rotating shaft of the impeller 30 shifts | deviates, the impeller 30 can rotate smoothly.

In the above embodiments, the motor unit of the fuel pump is a brush motor, but the motor unit may be a brushless motor.
As described above, the present invention is not limited to the above-described plurality of embodiments, and can be applied to various embodiments without departing from the gist thereof.

(A) is the whole view which looked at the impeller by 1st Embodiment from the fuel suction side, (B) is an enlarged view of the blade groove part of (A). Sectional drawing which shows the fuel pump of 1st Embodiment. (A) is the schematic diagram which looked at the blade groove | channel of the impeller by 1st Embodiment from the fuel suction side, (B) is the BB sectional drawing of (A). The enlarged view of the pump channel | path shown in FIG. The characteristic view which shows the relationship between dispersion width, a sound pressure peak, and pump efficiency. The characteristic view which shows the relationship between an adjacent angle and pump efficiency. (A) is a characteristic figure which shows the sound pressure peak of an equally spaced blade board, (B) is an unevenly spaced blade board.

Explanation of symbols

DESCRIPTION OF SYMBOLS 10: Fuel pump, 12: Pump part, 13: Motor part, 20, 22: Pump case (case member), 30, 90: Impeller, 34, 94: Blade plate, 34b: Radial direction outer end (one end part of rotation direction) ), 36, 92: blade groove, 100: rotating shaft, 202: pump passage, 220: swirling flow

Claims (1)

A motor section;
An impeller provided with a plurality of blade grooves in the rotation direction and rotated by a rotational driving force of the motor unit ;
An annular portion formed on the radially outer side of the blade groove;
A plurality of blade plates that are inclined to the front in the rotational direction adjacent to the annular portion and partition between the blade grooves adjacent in the rotational direction; and
A case member that rotatably accommodates the impeller and has a pump passage formed in a C shape along the rotation direction of the impeller ;
With
The adjacent angle formed by the rotation direction end portions of the adjacent blades is in the range of 8 ° to 12 °, and the difference between the maximum value and the minimum value of the adjacent angle is in the range of 2.5 ° to 4 °. Because
A fuel pump, wherein the impeller has a rotation speed of 4000 rpm to 12000 rpm, a discharge pressure of 0.25 MPa to 1 MPa, and a discharge amount of 50 L / h to 250 L / h.

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