HU215991B - Fuel pump - Google Patents

Abstract

The present invention relates to a fuel pump having a resin-like impeller (32) with alternating paddles (323) and paddle screens (322) around its circumference. The base screens are open outside and in the forward direction of the impeller, and in the axial direction of the impeller (32) they are divided by a partition wall. As a pump housing, it has a head which moves the impeller (32), which delimits the impeller (32). The cover is provided with our inlet pipe and outlet pipe, which are connected to the transport channel. The impeller has a gauge in the drive connection. SUMMARY OF THE INVENTION Each blade groove (322) of a thermal impeller (32) provides a connection between one front wall (322a) of the impeller (32) and the outer circumferential surface and provides a connection between the other front wall of the impeller (32) and the outer circumferential surface. (322b) and an axially interconnecting radial portion (322c) connecting the first and second heat portions (322a, 3 2b) in the axial direction. The first and second heat fin parts (322a, 322b) and the connecting guard part (322c) are formed between the shank walls of two weft blades (323). The all-part wall (321) is located between the first and second heat shield portions (322a, 322b) and forms the lower mantle deposit (3221, 3222), these lower mantle surfaces (3221, 3222) being radially from the inside of the impeller (32) to the outside. approaching each other, and terminating in the mantle surface (3230) at the outer end of the blade (323) with a spaced width (k) equal to or greater than the dimension between the mantle surfaces bordering the connecting heat portion (322c). ŕ

Description

BACKGROUND OF THE INVENTION The present invention relates to a fuel pump for supplying fuel or fuel to internal combustion engines or similar machines. For example, such a fuel pump can be used to supply pressurized fuel to the automotive injection system.

The internal combustion engine of cars and similar equipment is partially equipped with an electronically controlled fuel injection system 77071-587 / MJ. They use a motor-driven fuel injection pump that is part of the internal combustion engine injection system. The fuel pump extends into the liquid fuel receiving fuel tank and delivers the fuel at high pressure to the injection device in accordance with the electronic control unit commands.

One such known fuel pump is known as a "regeneration" or "WESTCO" type pump. The efficiency of the WESTCO type fuel pump is greatly influenced by the design of the cross-section of the conveyor and the design of the vanes and impeller.

Such a WESTCO fuel pump is known, for example, from Japanese Patent Application No. 63756, Japanese Application Specification 3-2720 or Japanese Patent Application Publication No. 60-47,984. In the case of a fuel pump known from the foregoing documents, the desired operating parameters are to be achieved by size selection, for example by the specific selection of the size Rm of the transport channel.

A conventional WESTCO fuel pump is described below with reference to Figures 21 and 22.

The impeller 9 of such a conventional fuel pump has a disc-like shape. The impeller 9 is provided with a series of successive vanes 93 and vanes 92 on each side, with equal spacing. Between the vanes 93 and the vanes 92 there is a partition wall 91 of the impeller 9.

The outer peripheral surface 910 of the dividing wall 91 has the same diameter as the outer peripheral surface 930 of the vanes 93. A conveying channel 95 is formed between the outer periphery of the impeller 9 and the inner surface of the pump housing 90. As the impeller 9 is rotated, its outer casing rotates at high speed in the conveying channel 95. Consequently, the liquid fuel in the paddle shafts 92 flows by the centrifugal forces to form vortexes 941 and 942 through the conveying channel 95.

By rotating the impeller 9, the liquid fuel travels circumferentially along its conveying conduit 95, producing vortices 941 and 942. Liquid fuel is subjected to dynamic pressure as it passes through its transport channel 95.

However, with conventional WESTCO fuel pumps, it has been found that a dead zone of 96 is formed between the eddies 941 and 942, as indicated by the dotted line in FIG. In this dead zone 96, the liquid fuel does not get the proper transport speed, so a counter current is generated. This in turn poses the problem that the counter current prevents the fuel from being transported at the required high pressure.

A protrusion extending radially from the housing wall or the side of the impeller and extending into the aforementioned dead zone 96 was used to eliminate counter current. However, practical experience has shown that such a projection increases the risk of the fuel flow being partially distributed between the two sides of the impeller due to the inability to move between the two sides.

In practice, WESTCO pumps are used for purposes other than fuel transport. Japanese Patent Publication No. 61-210288, for example, proposes such a pump for transporting water. In the solution described above, an attempt was made to eliminate the aforementioned unwanted counter current in the dead zone of the conveyor belt by making the outer end of the impeller divider to be pointed. Further, the above document also suggested that the height of the impeller dividing wall should be less than the length of the blades. Thus, the outer edge of the partition wall between the blades would be located inward.

Such an arrangement, in which the outer edge of the impeller partition wall is recessed relative to the outer edge of the blades, has already been proposed for an air compressor, for example, in Japanese Patent Publication No. 56-32095.

However, it is noted that the water pump referred to in Japanese Patent Application Nos. 61-210288 and the air compressor disclosed in Japanese Patent Application Nos. 56-32095 are very different from the fuel pump, inter alia, in terms of the required delivery capacity under pressure, impeller diameter and other factors. For this very reason, if the solutions proposed for the water pump and the air compressor were applied directly to the fuel pumps, it would be very difficult to achieve the desired pump design and operational effect.

For example, a typical water pump requires a delivery capacity of 100-10000 l / h and has a discharge height of 5-10 kgf / cm ² . In contrast, the typical fuel pump used in cars is expected to deliver 50-200 l / h and a pressure head of 2-5 kgf / cm ² .

The ranges of parameters required for the practical operation of the two types of pumps are therefore very different. It is also noted that the impeller of a water pump is generally about 100 mm in diameter, while the impeller of a conventional fuel pump for motor vehicles is about 50 or 30 mm in diameter. The reason for this latter size is that the size of the impeller of the fuel pump is severely limited by the fact that it must be located in the fuel tank of the car or in its immediate vicinity.

Furthermore, the air compressor is also very different from the fuel pump, not only in terms of delivery volume, efficiency, impeller diameter, etc., but also in terms of the compressibility and viscosity of the fluid to be transported.

HU 215 991 Β

Therefore, for an air compressor described in Japanese Patent Application No. 56-32095, the radial distance between the outer ends of the impeller blades and the wall surface of the conveyor is relatively small.

Because impellers have a relatively large diameter for water pumps and air compressors, such impellers are generally made of metallic materials. The metal impeller can be machined so that, for example, the blade grooves can be retrofitted and the outer end of the bulkhead can be tapered. In contrast, in the case of fuel pumps, the impeller is relatively small in diameter and is therefore generally manufactured by casting, such as injection molding of resins. This means that the tip of the impeller dividing wall of the fuel pump is difficult, since deformations and cracks would often occur as the casting is removed from the mold.

The relatively small impeller impeller of the fuel pump also has the problem that even a small deformation of the impeller shape adversely affects the fuel flow in the conveying channel and reduces the efficiency of the pump. Consequently, direct adaptation of known designs of conventional water pumps or air compressors to fuel pumps can barely meet the requirements for fuel pumps.

It is an object of the present invention to overcome the above shortcomings, i.e., to provide an improved, improved efficiency and easy-to-produce fuel pump.

In order to solve this problem, we have started with a fuel pump of the type described in the introduction, which preferably has an impeller made of synthetic resin. It is provided with blades and paddle blades arranged alternately along the periphery of the impeller. The paddle blades are open to both sides of the impeller and are separated by a dividing wall in the axial direction of the impeller. Further, the pump has a housing which rotates the bearing of the impeller. The fuel pump has a transport port located along the circumference of the impeller, which is provided with an inlet port and a transport port, which are connected to the transport port. The impeller is connected to a rotary drive.

A further development, i.e. the invention, is that each impeller of the impeller has a first groove portion which engages one side of the impeller and its outer boundary surface, and a second groove portion which provides connection to the other side of the impeller and its outer boundary surface. . In addition, there is a connecting third groove portion extending radially outwardly of the first and second groove portions and allowing their axial connection. The first and second groove portions and the connecting groove portions are formed between the side walls of the two adjacent vanes. The partition wall is located between the first and second groove portions, thereby forming their base surface. These base surfaces are configured relative to one another so as to be cohesive in radial direction from the inner part of the impeller to the outer part. These base surfaces terminate on the outer peripheral surface of the blade at a distance not less than a predetermined value between the base surfaces defining the connecting groove portion.

The fuel conveying pump design of the present invention achieves that the impeller has blades and dividers that define the respective blade grooves on each side of the impeller. The dividing walls of the present invention terminate at the inner portion of the outer ends of the blades, so that the opposing base surfaces of the blade vanes are spaced at a distance not less than the size defined at their outermost ends.

Accordingly, the outer ends of the dividing walls do not extend to the outermost periphery of the impeller, thereby creating a vortex flow of fuel along the base of the paddle shells throughout the conveyor, thereby effectively reducing the potential for dead zone formation and significantly improving pump efficiency.

Our experiments with the present invention have shown that completing a sloping base of paddle pads with such a distance, which is not less than a predetermined size between their outermost ends, provides a significantly more efficient pump than the conventional arrangement of dividing walls where the base surfaces meet at the tip of the impeller.

The more efficient pump operation is due to the fact that the area into which the swirling fuel does not enter directly is formed outside the dividing wall, and therefore the fuel can make swirling flows in this area that can be smoothly merged in the area outside the dividing wall .

Further, by forming the base surfaces of the paddle shafts with the spacing according to the invention, the outer end of the dividing wall is prevented from being deformed or damaged during its removal from the mold, which is unavoidable in conventional solutions.

The design of the present invention thus achieves both improved production performance and more efficient pump operation. Thus, the impeller can be simply manufactured by casting, which greatly improves the efficiency of the production.

The invention will now be described in more detail with reference to the accompanying drawings, in which some exemplary embodiments of the present invention are illustrated. In the drawing:

Figure 1 is a perspective view of the impeller of the fuel pump of the present invention;

Figure 2 is a cross-sectional view of a first embodiment of a fuel pump according to the invention, taken along line II-II in Figure 4;

Figure 3 is a cross-sectional view of a first exemplary embodiment of the impeller;

Fig. 4 is a side view of the impeller of Fig. 3 in an integral state;

Figure 5 is a longitudinal sectional view of a first embodiment of a fuel pump according to the invention;

Figure 6 is a schematic diagram of a fuel injection system with a fuel pump according to the invention;

Figure 7 is a sectional view of a mold for casting a first embodiment of the impeller of the present invention;

Fig. 8 is a diagram illustrating the efficiency of a fuel pump according to the invention as a function of d, = d ₂ and L, parameters;

Fig. 9 is also a diagram illustrating the efficiency of the first exemplary embodiment as a function of the change in parameter d;

Figure 10 is a graph showing the efficiency of a conventional fuel pump versus parameter D;

Figure 11 is a diagram illustrating the efficiency of conventional fuel pumps as a function of parameter k;

Figure 12 is a detail cross-sectional view of the main parts of the fuel pump of the present invention, which was used in our experiments;

Figure 13 is a sectional view showing a first comparative example of the solution of Figure 12;

Figure 14 is a sectional view of another comparative embodiment;

Fig. 15 is a diagram illustrating Figs.

Figures 1 and 2 show the efficiency of the fuel pumps;

the horse. Fig. 2A is a sectional view showing the main parts of a second exemplary embodiment of a fuel pump according to the invention;

Figure 17 is a perspective view of a third embodiment of the present invention;

Fig. 18 is a detail and perspective view of the impeller of a fourth exemplary embodiment of the present invention;

Figure 19 is a perspective view of a fifth embodiment of the impeller of the present invention;

Figure 20 is a detail and side view of a sixth embodiment of the impeller of the fuel pump of the present invention;

Figure 21 is a perspective view of a detail of a conventional impeller;

Fig. 22 shows the impeller of Fig. 21 in an integral state and in cross-section.

A first exemplary embodiment of the fuel pump of the present invention is illustrated in FIGS. 4 to 8 are shown. This fuel pump is built into the fuel supply system of the vehicle's internal combustion motel.

Figures 4 and 5 show the construction of the fuel pump.

Figure 5 clearly shows that the fuel pump consists of two main parts, namely 2 engine units and 3 pump units.

The motor unit 2 has a permanent magnet 21 mounted on the inner wall of the substantially cylindrical housing 1. Further, as a rotor, it has an armature 22 arranged concentric and rotatable in the interior of the permanent magnet 21.

The pump unit 3 has covers 311 and 312 which are fixed to the left end of the housing 1 in FIG. Between them, a disc-like impeller 32 is disposed concentrically and rotatably in the interior of the lids 311 and 312. Attached to the impeller 32 is the axis 220 of the armature 22, which is guided through the central opening of the cover 311.

Between the impeller 32 and the outer periphery of the covers 311 and 312, a conveyor channel 33 is provided for transporting liquid fuel. The conveyor 33 has an inlet port 41 at one end, which in this case is formed in the form C along the outer periphery of the impeller 32 (Fig. 4). Further, the transport link 33 has at its other end a discharge port 43 through which the transported fuel is discharged. Thus, the fuel to be conveyed enters the conduit 33, which is formed in the lid 312 in the present case, through the inlet port 41 (Fig. 5).

As shown above, the conveyor 33 is formed along the outer periphery of the impeller 32 and has inlet 331 and outlet 332, which are formed at predetermined locations and include a dividing wall 333 (Fig. 4). The inlet 331 and outlet 332 of the conveyor 33 are enlarged in radial direction relative to other portions of the conveyor 33. Inlet 331 in this case is larger in radial direction than outlet 332, as can be clearly seen in Figure 4.

Inlet 331 communicates with inlet nozzle 41, while outlet 332 communicates with outlet nozzle 43 through the interior of housing 1, for which cap 311 is provided with a through hole not shown.

The fuel in the housing 1 is further pressed by the conveyor connection 42 located opposite the housing 1. An electrical connection is provided adjacent to the conveyor connection 42, the electrical connection of which is connected to the terminal 23 of the conveyor to drive the motor unit 2. The terminal 23 is connected to a brush (not shown) via noise reduction elements such as a coil and a capacitor.

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The impeller 32 of the present invention is described in more detail in Figs. 4 to 8.

The impeller 32 is pivotally mounted in the pump housing and releasably coupled lids 311 and 312 and secured thereto in the housing 1.

A series of blades 323 are formed at predetermined intervals along the outer circumference of the impeller 32, and one blade groove 322 is formed between two adjacent blades 323.

According to the invention, each blade groove 322 comprises first and second groove portions 322a and 322b formed on both sides of the impeller 32 along the outer circumference. Further, an additional connecting groove part 322c is provided which is formed on the outermost part of the impeller 32 in the radial direction and is intended to connect the groove parts 322a and 322b in the axial direction.

The groove portions 322a, 322b and 322c together form the paddle 322 having a substantially C-shaped cross-section and extending from one side surface to the opposite side surface of the impeller 32 while passing through its outermost, i.e. outermost region.

The blade 323 is located between two adjacent blade grooves 322. The blade 323 has a radial blade design which is perpendicular to the side surfaces of the impeller 32 and in fact forms the side walls of the blade vanes 322.

A dividing wall 321 is provided between each pair of groove portions 322a and 322b on either side of the impeller 32, which tapers radially to the outer periphery of the impeller 32. The dividing wall 321 has a short, flat portion at its outer end and two opposed inclined portions which

For the groove portions 322a and 322b, they form a lower peripheral surface 3221 and 3222 respectively. These delimiting peripheral surfaces 3221 and 3222 are in the present case curved, the radius R of which is shown in Figure 2.

The axial distance between the peripheral surfaces 3221 and 3222 gradually decreases toward the radially outermost periphery of the impeller 32, and its minimum value is

321 at the outer end of the partition. This minimum spacing can also be defined as the spacing between the ends of the peripheral surfaces 3221 and 3222. Further, the outer peripheral surface 3210 of the partition wall 321 defines the base surface of the groove portion 322c.

Due to the above configuration, the fuel flows in the direction of rotation of the impeller 32 not only along the sidewalls of the grooves 322a and 322b but also along the sidewalls of the connecting groove 322c.

The position of the side wall 321 of the impeller 32 is well illustrated in Figures 1 and 2. The outer peripheral surface 3210 is located radially inwardly from the outer peripheral surface 3230 of the vanes 323 which forms the outermost surface of the impeller 32. In the present embodiment, the total radial length of the groove portion 322c is indicated, i.e., it corresponds to the distance between the outer peripheral surface 3210 of the partition wall 321 and the outer peripheral surface 3230 of the blade 323. This L] distance is set to 40% of each vane plate 323 length _L2 (Figure 2).

Figures 3 and 4 clearly show that the blades 323 and

The paddle blades 322 are spaced along the circumference of the impeller 32 at predetermined intervals. At the center of the impeller 32 is a shaft receiving hole 325 for engaging the shaft 220.

The various dimensions of the impeller 32 for the above embodiment are summarized in Table 1:

Table 1

D t d ₃ R L, 1-2 rm k 30 2.4 0.7 0.7 4 1.0 2.4 0.7 0.3

The symbols used in the table above are as follows:

D - impeller diameter t - impeller thickness d |, d ₂ - axial spacing d ₃ - radial spacing

Diameter D of Table 1 thus denotes the outer diameter of the impeller 32 to be understood. The thickness t is understood to mean the axial size, i.e. the thickness, of the impeller 32. As shown in Fig. _{2, the} axial gaps d) and d ₂ indicate the distance between the axial ends of the blades 323 and the inner side surfaces of the covers 311 and 312. The dj spacing denotes the distance between the radial end of the blades 323 and the inner peripheral surface of the cover 311. The radius R represents the radius of curvature of the sloping and curved peripheral surfaces 3221 and 3222 of the impeller 321 dividing wall. The reference numeral Lj denotes the radial size of the conveyor 33 or the groove portion 322c, namely the radius 32 of the outer circumference 3210 of the partition 321 and the outer circumference 3230 of the blade 323 - the radius L | - total radial size of the transport link Lj - total radial length of the blade Rm - typical size of the transport link k - width of the end of the partition wall (Figure 2).

Attachments. The total radial length L ₂ of the vanes 323 extends from the inner end of the vanes 323 to the outer peripheral surface 3230, including the connecting passage. The parameter Rm for the transport channel 33 can be determined from the relation S / l, assuming that the axial cross-sectional area of the transport channel 33 is delimited by the signals abcdjihgfea (Fig. 2), so this parameter is denoted by the parameter S. The length of the section along the circumferential edges of the impeller 32 can thus be determined by the abcd points in FIG. The length k of the end of the partition 321 denotes the axial dimension of the outer peripheral surface 3210 of the partition 321. Finally, it should be noted that Figure 1.

The values in Figures 1 to 4 are in mm.

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Figure 6 shows that the fuel pump of the present invention is integrated into a fuel tank 61 which is arranged in the vehicle and is electrically connected to a battery 62. Thus, the fuel pump delivers fuel 63 from the fuel tank 61 to the fuel injector assembly 64. The fuel pump inlet port 41 is provided with a fuel filter 65 and a pipe 66 is connected to the outlet port 42. Pipeline 66 delivers fuel to the injection units 67 of the fuel injector assembly 64 and controls the fuel pressure to the set value of control unit 68. Fuel discharged from control unit 68 is returned to fuel tank 61 via return line 69. The injection units 67 inject fuel into the inlet duct 70 of the internal combustion engine.

In the present example, the fuel pump for the fuel injection structure 64 of the internal combustion engine 70 of the vehicle operates under operating conditions such that the delivery volume is 50-700 l / h and the delivery pressure is 2-5 kgf / cm ³ . Considering the environmental conditions under which the vehicle operates, the fuel pump has been designed to operate smoothly in the temperature range of -30 to + 80 ° C.

For a fuel pump operating under such conditions, and particularly for transport pressure, the impeller 32 may preferably have a diameter of 20 to 65 mm and a typical Rm value of the conveyor 33 may be 0.4 to 2.0 mm. It is particularly advantageous to select a typical Rm dimension of the transport link 33 in the range of 0.6-1.6 mm. (Special dimensions of such a fuel pump are disclosed, for example, in Japanese Patent Nos. 63-63756 and U.S. Patent 4,493,620).

The manufacturing technology of the impeller 32 described above is described in more detail below. Figure 7 is a sectional view of a mold used for casting impeller 32. Figure 7 shows exactly the part that forms the paddle 322.

The mold 72, which is referred to throughout, is composed of two parts, that is to say, mold parts 74 and 75, which are joined to each other in a divider plane 73, corresponding to the axial center of the impeller 32. The inner forming space of the mold 72 is formed according to the shape of the impeller 32. In the present case, this forming space is slightly larger in diameter and thickness than the impeller 32.

In Figure 7, the dashed line 76 marks the outline of the mold forming interior 72 and the thin shape 77 indicates the final shape of the impeller 32. Figure 7 shows that the inner forming space of the mold 72 at the location corresponding to the blade groove 322 is substantially the same as the shape of the impeller 32.

For the production of the impeller 32, first, a thermosetting resin is filled into the molding space of the mold 72 to mold the outer shape of the impeller 32. This cast, raw impeller is slightly larger in diameter and thickness than the finished impeller 32. However, after casting, the portion of the impeller 32 corresponding to the blade groove 322 is the same as the finished product.

The impeller 32 in this case was made of phenolic resin, which was reinforced by admixing fiberglass. The volume of such thermosetting resins changes as a result of the temperature change, which enables the pump to operate at a high temperature over a wide temperature range.

The lateral and outer boundary surfaces of the impeller 32 cast in the mold 72 are post-machined. It is emphasized that in this post-machining operation the side surfaces of the impeller 32 and the outer surfaces 3230 and 3232 and 3232 of the blades 323 are, for example, finished by grinding. After the post-machining operation, the impeller 32 is shown in FIGS. Figs. Thus, the surfaces of the impeller 32 shown in Fig. 1, which are marked with dots, already have their final shape at the time of casting, without any post-machining. In this embodiment, the outer peripheral surface 3210 is not further machined at the outer end of the partition wall 321.

As mentioned above, the impeller 32 is cast using the mold 72. Therefore, the paddle blades 322 are very simple to form, making the impeller 32 very suitable for mass production. If the outer end of the partition wall 321 is too thin, it may be deformed by opening the mold parts 74 and 75 and removing the casting therefrom. This, in turn, would significantly affect the operational safety and efficiency of the pump. To prevent this, in the embodiment shown, the outer peripheral surface 3210 of the partition wall 321 is planar to provide sufficient thickness at the outer end of the partition wall 321 as well. Thus, if the casting is removed from the mold parts 74 and 75, deformation of the end of the partition wall 321 is excluded.

Particularly when using thermosetting resins as a base material, there is a risk that the partition wall 321 will crack after opening the mold parts 74 and 75, since the thermosetting resin is generally brittle. To prevent this, the outer end of the partition wall 321 is sized sufficiently thick to increase its strength. This prevents the thermosetting resin from cracking at the outer end of the partition wall 321 after casting.

The mode and advantages of the first exemplary embodiment of the fuel pump of the present invention will be described below.

As soon as the motor unit 2 is connected to the battery 62 via the terminal 23, the armature 22 of the motor unit 2 is rotated. The rotation of the armature 22 is directly transmitted by the shaft 220 to the impeller 32, i.e., it is rotated.

By rotating the impeller 32 from the fuel tank 61, the pump unit 3 draws liquid fuel into the conveying channel 33 through the inlet port 41, and the impellers 323 of the impeller 32 are

EN 215 991 Β is transported under pressure to its 33 transport channels. The pressurized fuel is supplied from the nozzle 42 to the injection units 67.

In more detail in the transport of the fuel pump, it is noted that when the impellers are rotated in the covers 311 and 312, the outer slats of the impeller 32 travel at high speed in the conveyor 33. Liquid fuel moves along its circumference in its transport channel 33, but also performs swirling motion in the vortexes 341 and 342 through centrifugal forces. This significantly increases the dynamic pressure of the fuel.

In the exemplary embodiment detailed above, the conveyor belt is axially shared with the impeller 321 of the impeller 32. However, since each of the paddle blades 322 extends axially through the groove portion 322c outside the outer peripheral surface 3210 of the partition wall 321, the fuel can easily travel transversely between the two sides of the impeller 32, thereby preventing uneven fuel on both sides of the impeller 32. . This results in the creation of a pressure that axially reduces the frictional resistance of the impeller 32 and consequently the noise during rotation.

Because the outer axial surface of each of the vanes 323, which transports the fuel circumferentially, is radially outwardly facing the outer flank 3210 of the dividing wall 321, the fuel can be well transported circumferentially with the outer axial surface of the vanes 323, by rotation of the impeller 32.

An important additional advantage is that the extreme eddies of conventional solutions can be eliminated at the outer peripheral surfaces of the partition wall and at the inner peripheral surface of the housing. In this way, a higher transport pressure can be achieved with the present invention with respect to the same electric drive power.

Of course, the design of the impeller 32 may vary to some extent in this embodiment, but the lids 311 and 312 and other components may remain the same design for the fuel pump of the present invention, these unchanged parts being conventionally formed. Thus, the operating parameters of the fuel pumps of the present invention can be significantly improved in the sense detailed above, and we have also improved the manufacturing conditions of the impeller, resulting in a fuel pump of substantially higher practical value.

During the experimental production of the first embodiment of the fuel pump according to the invention, a series of experimental models were prepared which also tested the efficiency thereof. These experimental results will be described below, demonstrating that they exhibit very good efficiency due to the size ratios and size choices of the present invention.

For the purpose of testing the efficiency of the pump, the pump inlet was calculated as a function of the load torque and the rotational speed, and the pump outlet was dimensioned in relation to the transport pressure and transport power. Delivery pressure was measured using a digital multimeter (manufactured by ADVANTEST Co.) and small semiconductor pressure sensors (TOYODA Machine Works Limited). Delivery volume was measured with the aid of a digital flow30 meter, manufactured by ONO Measuring Instrument Limited.

First, the experimental parts of the fuel pumps according to the invention are provided with different impellers and different conveyor ducts, the dimensions of which are given in Table 2. The efficiency of these experiments was measured. The efficiency values obtained in the experimental examples are shown in the diagram in Figure 8.

Table 2

Woman. D t d „d ₂ dj R L, l ₂ rm k 1 30 2.4 0.6 0.7 4 0 2.4 0.63 0.3 2 30 2.4 0.7 0.7 4 0 2.4 0.70 0.3 3 30 2.4 0.8 0.7 4 0 2.4 0.77 0.3 4 30 2.4 0.9 0.7 4 0 2.4 0.83 0.3 5 30 2.4 0.7 0.7 4 0.5 2.4 0.7 0.3 6 30 2.4 0.6 0.7 4 1.0 2.4 0.63 0.3 7 30 2.4 0.7 0.7 4 1.0 2.4 0.70 0.3 8 30 2.4 0.8 0.7 4 1.0 2.4 0.77 0.3 9 30 2.4 0.9 0.7 4 1.0 2.4 0.83 0.3 10 30 2.4 0.6 0.7 4 1.5 2.4 0.63 0.3 11 30 2.4 0.7 0.7 4 1.5 2.4 0.70 0.3

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Table 2 (continued)

Woman. D t d _b d ₂ d ₃ R L, l ₂ rm k 12 30 2.4 0.8 0.7 4 1.5 2.4 0.77 0.3 13 30 2.4 0.9 0.7 4 1.5 2.4 0.83 0.3

The symbols used in the table above are as follows:

D - impeller diameter í - impeller thickness dp d ₂ - axial spacing d ₃ - radial spacing

R - radius of curved groove surface

L | - total radial dimension of the transport passage L ₂ - total radial length of the vane Rm - typical size of the transport link k - width of the end of the partition wall (2) (data in mm).

In Table 2, Table 1-4. Experimental examples refer to conventional impellers, which therefore have no traffic lanes or passageways.

Figure 8 shows the efficiency of the experimental fuel pumps shown in Table 2. The data show that the embodiment of Example 7 gave a maximum efficiency of 35%.

Of the conventional fuel pumps equipped with conventional impellers, embodiment 2 gives the highest efficiency, which is about 30%. L | selection / _L2 sizes of 0.4 and any of the fuel pump, which has a connecting passage according to the invention, provides a significantly higher efficiency than the conventional pumps. It is of little importance in this matter what value dj = d _{2 is} chosen in the range 0.6-0.9. The inventors have found that if to the value _d2 = dimensioned to be between 0.7 to 0.8, any pilot pump can provide higher efficiency than the conventional solutions, regardless of whether the L | / ₂ ratio of L to a value from a relatively wide range of 0 to 0.6.

Figure 8 shows the characteristic curves that the efficiency varies substantially similar manner, if L] / _L2 ratio increases. This also seems acceptable, since in our experiments we changed the value of the Lj spacing, but in this case the sloping surfaces of the dividing wall are too short to produce the vortexes 341 and 342 sufficiently if the L 1 / L ₂ ratio is set high.

The characteristic curves of Fig. 8 also show that the efficiency changes similarly, i.e. parallel to itself, as the ratio L [/ L _2] becomes smaller and smaller. This is due to the radially outer end face of the partition wall 321 approaching the outer peripheral surface of the conveying channel, which may result in a dead zone 96 in the flow passage (FIG. 22), which degrades the efficiency of the pump.

It is further noted that if the axial distance d ₁ = d _{2 is} chosen to be 0.6-0.9 mm, the efficiency of the pump will be greatly reduced.

This is explained by the fact that the relatively small axial distance d ₁ = d ₂ reduces the transport volume and prevents the formation of vortices 341, 342.

On the other hand, it has been found that a large axial gap d ₁ = d ₂ makes the conveying channel too large, which can lead to the creation of unfavorable eddy currents that reduce the efficiency of the pump.

Table 3 provides data for further experimental fuel pumps using various impellers and conveyor ducts of various shapes for experimental purposes and measuring their efficiency. The results of these experiments are illustrated in Figure 9.

Table 3

Woman. D t d _{) t} d ₂ d ₃ R L, ^L 2 rm k 21 30 2.0 0.7 0.7 4 0 2.4 0.7 0.3 22 30 2.4 0.7 0.7 4 0 2.4 0.7 0.3 23 30 3.0 0.7 0.7 4 0 2.4 0.7 0.3 24 30 2.0 0.7 0.7 4 0.5 2.4 0.7 0.3 25 30 2.4 0.7 0.7 4 0.5 2.4 0.7 0.3 26 30 3.0 0.7 0.7 4 0.5 2.4 0.7 0.3 27 30 2.0 0.7 0.7 4 1.0 2.4 0.7 0.3 28 3 2.4 0.7 0.7 4 1.0 2.4 0.7 0.3 29 30 3.0 0.7 0.7 4 1.0 2.4 0.7 0.3 30 30 2.0 0.7 0.7 4 1.5 2.4 0.7 0.3

HU 215 991 Β

Table 3 (continued)

Woman. D t d „d ₂ d ₃ R L, ^L 2 rm k 31 30 2.4 0.7 0.7 4 1.5 2.4 0.7 0.3 32 30 3.0 0.7 0.7 4 1.5 2.4 0.7 0.3

The symbols used in the table above are as follows:

D - impeller diameter / - impeller thickness d _b d ₂ - axial spacing dj - radial spacing

R - radius of curved groove surface

L, - total radial length of the transport link mi L ₂ - total radial length of the blade Rm - typical size of the transport link k - width of the end of the divider wall (dimensions in mm).

In the experimental data in Table 3, the total axial length of the conveyor belt varied with the change in impeller t thickness.

Table 3 illustrates Table 21-23. Examples refer to conventional types of impellers which do not have conventional connecting passages.

The diagram in Figure 9 shows the efficiency of the experimental models in Table 3. The results showed that the exemplary embodiment shown in Table 3, No. 28, provided the maximum efficiency of 25%.

Of the examples representing conventional impeller pumps, Example 22 provided the highest efficiency at 30%. Experimental results have shown that if the L / U ratio is chosen to be about 0.4 mm, any fuel pump having the connecting passages according to the invention can achieve at least as much or even greater efficiency as conventional pumps, regardless of which value of the impeller thickness t is selected in the range of 2.0 to 3.0 mm.

According to our experimental experience, if the impeller t is selected in the range of 2.4 to 3.0 mm, the efficiency of any pump may be the same or even higher than that of a conventional pump, whether or not the Lj / L ₂ ratio is 0.1-0, 6 range.

The characteristic curves of Figure 9 show that the efficiencies change in a substantially similar manner when the L 1 / L ₂ ratio is increased. This is understandable since the Lj size is changed in the experimental examples, the partition wall sloping surfaces or lapáthomyok bottom portion is too reduced to sufficiently develop from their vortexes 341, 342 in cases where the ratio L] / L ₂ high.

If the impeller thickness t is chosen to be small, the efficiency of the fuel pump is greatly reduced. This is due to the fact that the thickness t reduces the impeller blade area to such an extent that it prevents the formation of the corresponding swirls 341 and 342. Another reason is that since the total axial length of the conveyor belt varies with the variation of the impeller thickness in the above experimental examples, the conveyor belt can be greatly reduced axially and thus the whirl formation effect is impaired.

On the other hand, if the width of the impeller t is set to a high value, the efficiency of the fuel pump is reduced to a lesser extent. This is due to the fact that the total axial length of the conveyor belt varies with the variation of the impeller thickness in the experimental examples, the conveyor belt is too large in the axial direction, which also adversely affects the formation of the vortices 341 and 342.

As mentioned above, from the experimental data provided in Tables 2 and 3. Figures 8 and 9, it will be appreciated that the communicating passages in the impeller and the L | ratio / L ₂ sizes 0.1-0.6 by choosing between values, you can achieve a pump efficiency that is at least the same, but much higher than that of a conventional pump.

From the experimental results of Figures 8 and 9, it can also be established that the pumping efficiency is maximal if, taking into account the values of Table 1, the ratio L (/ L ₂ ) is chosen to 0.4. In the exemplary embodiment, the dimensions in Table 1 were used.

With regard to the first exemplary embodiment, it has already been mentioned that since the divider wall 321 is recessed in the radial direction relative to conventional impellers, vortexes 341 and 342 are formed on the two opposite sloping front faces of the divider wall 321. These penetrate into the dead 96 zone (Fig. 22), which is unavoidable for conventional pumps and consequently cannot improve the efficiency of the pump, since no transverse flow is generated in the dead 96 zone.

The specification does not elaborate on why the diameter D of the impeller and the width k of the outer end of the dividing wall were chosen for the first embodiment according to the values in Table 1. Fig. 10 is a diagram illustrating how the fuel pump efficiency changes when the impeller diameter D is changed for a conventional WESTCO fuel pump, which has no connecting passageways.

The characteristic curve of Fig. 10 was obtained from the measurement results of experimental fuel pumps, which were made to the dimensions shown in Table 4, and only the impeller D diameter was varied, and the efficiency of these experimental models was also measured.

HU 215 991 Β

Table 4

D t d. | d ₂ d ₃ R Li ^ 2 rm k - 2.4 0.7 0.7 4 0 2.4 0.7 0.3

L | - total radial size of the conveyor belt L ₂ - total radial length of the vane Rm - typical size of the conveyor belt k - width of the end of the partition (values in the table are given in mm).

The terms used in Table 4 are explained as follows:

D - impeller diameter t - impeller thickness cl ,, d ₂ - axial clearance d ₃ - radial clearance

R is the radius of the curved groove mantle

It can be seen from Figure 10 that, despite conventional pumps, the efficiency can be quite acceptable at least on part of the curve, that is, above 20%, if the diameter D of the impeller is selected between 20 and 65 mm. Therefore, it is believed that if these known pumps are provided with connecting passages and the proposed impeller design in accordance with a first embodiment of the invention, the efficiency can be significantly increased by these design dimensions, i.e. the size range given in Table 4. All you have to do is select the diameter D of the impeller in the range of 20 to 65 mm.

It is also apparent from the characteristic curve of Fig. 10 that if the diameter D of the impeller is reduced below 20 mm, the efficiency is greatly reduced. Conversely, if the diameter D of the impeller is greater than 65 mm, the efficiency is reduced to a lesser extent. This is due to the fact that the small impeller diameter makes the length of the conveyor belt 15 too short for this passage section to work effectively in the pump. In our experience, the large impeller diameter causes slip resistance, which can cause the impeller to be deformed to an extent that is detrimental to efficiency.

All. FIG. 1B is intended to illustrate how the pump efficiency changes by changing the axial width k at the outer end of the impeller divider wall for a conventional WESTCO type fuel pump that has an impeller that is not designed for traffic.

All. The characteristic curve of FIG. 5 is obtained by testing experimental fuel pumps, which are made with the dimensions of FIG. 5. The width k was only changed for the experimental pumps and the effect on the pump efficiency was measured.

Table 5

D t d. d ₃ R L, ^ 2 rm k 30 2.4 0.7 0.7 4 0 2.4 0.7 -

The terms used in the table above are explained as follows:

- - impeller diameter t - impeller thickness d ,, d ₂ - axial spacing d ₃ - radial spacing

R - radius of curved groove circumference

L | - total radial size of the conveyor belt L ₂ - total radial length of the vane Rm - typical size of the conveyor belt k - width of the end of the partition (values in the table are given in mm).

All. Fig. 4a shows that, despite conventional pump designs, quite acceptable efficiencies, i.e. above 20%, can be achieved if the width A: is selected between 0.3-0.8 mm. It is also assumed from the characteristic curve of Fig. 11 that an efficiency of no less than 20% would also be obtained if the width A were chosen to be less than 0.3 mm. However, it is noted that since the impeller is molded from a resin, the mold of Fig. 7 is used as an example, so that reducing the width of k to less than 0.2 mm may cause difficulties in terms of strength and manufacturing technology.

From the above experimental results, it is also believed that in the case of the present pumps, which, like the first embodiment of the invention, have connecting grooves but impeller shape45 needles, other very favorable efficiency would be achieved with the size range approximating Table 5. width k as it would be chosen between 0.2 and 0.8 mm.

All. It is also apparent from the characteristic curve of Fig. 1A that if the width of k is greater than 0.8 mm, the efficiency is slightly reduced. This can be explained by the fact that if the width k is too large, the dead 96 zone can be significantly increased (Figure 22).

From the above experimental results, it can be seen that these pumps, which, like the first embodiment of the present invention, have interconnecting channels, can provide high efficiency if the ratio L ₁ / L ₂ is chosen in the range of 0.1-0.6. Further, high efficiency can also be achieved if the L 1 / L ₂ ratio is likewise maintained between 0.1 and 0.6

HU 215 991 Β, but the axial distance dj = d _{2 is} scaled to 0.7-0.8 mm.

It is further concluded that high efficiency can also be achieved by selecting the ratio L ₁ / L ₂ between 0.1 and 0.6 and the impeller t thickness between 2.4 and 3.0 mm. It is further believed that the above-mentioned advantageous operating advantage can be obtained by selecting the diameter D of the impeller to be between 20 and 65 mm and the axial width k of the outer end of the divider being between 0.2 and 0.8 mm.

The following is a description of the particular advantages of the present invention due to the design of the outer peripheral surface 3210 of the partition wall 321. The following is an explanation of the experimental results that are shown in Figure 12,

13 and 14. These impellers are provided with paddle blades made by machining the outer peripheral parts of the disc base.

Fig. 12 is a cross-sectional view of a variant of the impeller 32 having a conveyor belt according to the invention. The dimensions used are a

Table 6 contains.

Table 6

D t d _b d ₂ , d ₃ R l ₃ L, l ₂ rm k 30 2.35 0.7 4 0.6 1.0 2.4 0.7 0.3

The terms used in the table above are explained as follows:

D - impeller diameter t - impeller thickness d |, d ₂ - axial spacing d ₃ - radial spacing R - radius of curved groove surface

The impeller 32 of Figure 12 has the same shape as the impeller already described in Figure 2 and Table 1. Note that the thickness t is taken to 2.4 mm in Table 1, with a rounded dimension of 2.35 mm.

In Fig. 12, VI denotes an imaginary vertex in which the peripheral surfaces 3221 and 3222 of the blade groove 322 intersect if extended with their original arc. In the experimental example, this imaginary intersection VI is located in the groove portion 322c, which is the connecting passage, approximately in the middle of the radial distance L1.

L ₃ - spacing of imaginary vertex Lj - total radial size of the conveying link L ₂ - total radial length of the vane Rm - typical size of the conveying link k - width of the end of the dividing wall (dimensions in mm).

Figure 13 is a cross-sectional view of a fuel pump with a first comparative exemplary impeller 32 and a conveyor 33.

In this first comparative example, the dimensions in Table 6 were used, but the peripheral surfaces 131 and 132 were formed at the outer ends of the blade grooves, intersecting at the outer end of each blade. Accordingly, the dimensions used in this example are shown in Table 7 below. The outer intersection L ₃ indicates the distance between the blade and the outer end 133 of the dividing wall.

Table 7

D t d |, d ₂ , d ₃ R ^L 3 L, 1-2 rm k 30 2.35 0.7 4 1.0 Shoot 2.4 0.7 0.3

L ₃ - intersection distance Lj - total radial size of the transport link L ₂ - total radial length of the blade Rm - typical size of the transport link k - width of the end of the dividing wall.

(values above are in mm).

the peripheral surfaces of the paddle shells intersect at the intersection of the outer end of the paddle. Thus, the dimensions selected are shown below in Table 8.

The terms used in the table above are explained as follows:

D - impeller diameter l - impeller thickness d _b d ₂ - axial spacing d ₃ - radial spacing

R - radius of curved groove surface

Figure 14 shows a second comparative example of a further pump version, more specifically, its impeller 32 and its conveyor 33.

In this second comparative embodiment, the size values in Table 6 were used, however

HU 215 991 Β

Table 8

D t d | .d ₂ R l ₃ L, Lj rm k 30 2.35 0.7 4 0 0 2.4 0.7 0

The terms used in the table above are explained as follows:

D - impeller diameter t - impeller thickness d ,, d ₂ and d ₃ - axial or radial distance

R is the radius of the concave groove surface

Lj - intersection distance

FIG. 1 to 2 and the efficiency of the experimental pump are shown in the diagram, with the efficiency in% on the vertical axis and the flow rate in 1 hour on the horizontal axis. Figure 15 is a solid line representing the efficiency curve of the present invention as described in Figure 12. The dashed line represents the first comparative embodiment of Figure 13, and the dashed line represents the characteristic curve of the second comparative example of Figure 14.

The characteristic curves of Fig. 15 show that the impeller 32 of Fig. 12 provides maximum efficiency. In our opinion, the difference in efficiency shown in the diagram is due to the eddy currents created by rotating the impeller 32 according to the invention.

Thus, in the embodiment of Figure 12, eddy currents are created in the fuel flow around each blade groove which first passes along the peripheral surface of the blade groove 3221 and 3222 and then meets in the center of the passage and flows outwardly. In this experimental example, the outer peripheral surface 3210 of the impeller partition wall is formed as a flat surface with a predetermined width k. This design creates an area outside the outer peripheral surface 3210 of the partition wall where the eddy currents arriving along the peripheral surface 3221 and 3222 do not flow directly and the fuel stagnates in this area.

Based on experimental experience, it is believed that the greater pump efficiency shown in Figure 15 is due to the stagnant fuel outside the outer peripheral surface 3210 of the dividing wall to facilitate the smooth flow of eddy currents from the peripheral surfaces 3221 and 3222.

However, in the first comparative example of FIG. 13, eddy currents traveling along peripheral surfaces 3221 and 3222 abruptly collide outside the outer end point 133 of the partition wall and then flow outward in a radial direction. 13, however, as soon as the eddy currents along the peripheral surfaces 3221 and 3222 collide with each other, both partial currents have a large axial component due to the flow velocity. These axial velocity components mutually attenuate the velocity of the fuel flows, i.e., reduce their movement, so that the efficiency of the pump is much lower than in the case of Fig. 12.

L, - total radial size of the conveyor belt L ₂ - total radial length of the vane Rm - typical size of the conveyor chain k - width of the end of the partition wall (all dimensions given in mm).

In order to reduce the aforementioned collision of fuel flows, it should be considered that the angle α in Figure 13 should be chosen to be smaller. However, it should be noted that if the impeller is to be made of resin material by casting, it may be very difficult to implement a small pitch angle in terms of manufacturing technology and sufficient strength.

In the second comparative example of Figure 14, it is assumed that the collision of fuel streams is reduced compared to that of Figure 13. However, it is noted that the vertically long partition wall reduces the volume of the conveying groove so much that it also reduces the efficiency of the pump.

In summary, the embodiment of Figure 12 prevents collision of the two eddy currents by creating a stagnating fuel zone around the outer end of the bulkhead. This allows the two eddy currents to mix gently with each other. The result of this is that very strong eddy currents are created in the section from impeller paddles to conveyor, thereby significantly increasing the efficiency of the pump.

Figure 16 is a sectional view of a second embodiment of a fuel pump according to the invention.

In this second embodiment of the impeller, similarly to the first embodiment, the bottom surfaces of the paddle shafts are formed with curved peripheral surfaces 164, 165 in accordance with the prespiratory currents generated by the rotation of the impeller. A predetermined width k is provided at the outermost end of the circumferential surfaces 164, 165, which in this case is formed as an arcuate surface 163 adapted to the radius of the eddy currents.

More specifically, in the embodiment of Fig. 16, for a curved surface 163 formed at the outer end of the partition 162, the thickness k of the partition 162 is selected to be 0.3 mm between ends 164a and 165a of radial surfaces 164 and 165. In other words, in this second exemplary embodiment, the thickness k of the partition 162 is 0.3 mm at the inflection points of the curved surfaces 164, 165. The outer end of the partition 162 is thus determined.

Note that in the embodiment of Figure 16, other dimensions and parts of the fuel pump12

EN 215 991 Β are identical to those used in the first exemplary embodiment. Thus, in this second embodiment, an arcuate surface 163 is formed at the outer end of the partition wall 162, in addition to providing an area in which fuel eddy currents from the peripheral surfaces 164 and 165 do not enter directly. Thus, the fuel in this part of the conveying channel may stagnate, thereby allowing the two-sided eddy current to mix smoothly there.

Fig. 17 is a perspective view of the impeller 32 of a third exemplary embodiment of the present invention.

In this configuration, the impellers 323 of the impeller 32 are slightly inclined at their upper and rear edges, resulting in a slightly sloping surface 3231a at these locations. It will be noted that this sloping surface 3231 is thus located rearward of the impeller 32. The fuel flowing out of one of the impeller grooves 322 of the impeller 32 thus proceeds in the vortex streams 341 and 342 and then passes into the next blade groove 322 in its conveying channel, where it is subjected to additional vortexing force. At this point, the eddy currents 341 and 342 flowing out of the blade groove 322 are much easier to introduce into the next blade groove 322 due to the rear flattened configuration of the blades 323 described above. Consequently, the loss of eddy currents generated by the impeller 32 can be reduced, which increases the efficiency of the pump.

Figure 18 illustrates a fourth exemplary embodiment of the impeller 32 of the fuel pump of the present invention.

In this version of the impeller 32, in addition to the sloping surfaces 3231a of FIG. 17, the front portions of the blades 323 are provided with similar, slightly sloping flaps which form sloping surfaces 3231b. With this design, the loss of eddy currents can be further reduced compared to the solution of FIG. 17. Further, this fourth exemplary embodiment has the additional advantage that the impeller 32 can be mounted in both directions without regard to the direction of rotation.

Fig. 19 shows a fifth exemplary embodiment of the impeller 32 of the fuel delivery pump of the present invention.

This impeller 32 differs from the embodiment of Fig. 18 in that the front inclined surface 3231b is flattened to a lesser extent than the rear inclined surface 3231a. Thus an asymmetrical blade tip design is achieved. The advantage of this design is that it can also significantly reduce the loss of eddy currents.

Figure 20 is a side view of a sixth embodiment of the impeller 32 of the fuel delivery pump of the present invention.

In this embodiment, the impeller 32 is configured such that each dividing wall 321 engages the two adjacent blades 323 through slightly curved sections 3214 and 3215 at both ends of the outer peripheral surface 3210 of the dividing wall 321.

This embodiment prevents the fuel flowing circumferentially along the outer circumferential surface 3210 of the partition wall 321 from entering the area of the blades 323 along the arc of sections 3214 and 3215. Thus, the fuel flow is not slowed down, which means that losses are further reduced.

In addition, when reaching the outer circumference of the impeller 32, the blade 323 is subjected to a high load. By using the curved sections 3214 and 3215, the blades 323 are thereby substantially stiffer in the joint region of the blades 323, thereby preventing any deformation or damage.

It will be appreciated that the above-mentioned figures, the embodiments of the impeller shown above, may also be combined with one another, in particular the end of the impeller 321 of the impeller 32 may be larger than the dimensions mentioned above.

Further, the impeller 32 of the fuel pump according to the invention generally rotates at 3000 to 15000 rpm. Thus, in the said embodiments, the desired operating parameters and characteristics are understood to be within such a range of revolutions.

Although the invention has been described with reference to the exemplary embodiments depicted, it is to be understood that within the scope of the claimed invention, the invention may be embodied in many other variations or combinations.

Claims (19)

PATENT CLAIMS 1. A fuel pump, preferably having a resin-shaped impeller having rotatable vanes and vane grooves along its circumference, open to the outside of the vane lobes and laterally to the impeller, and rotatably diverted by impellers in the axial direction of the impeller; having a bearing cover defining a conveyor channel formed along the circumference of the impeller, the cover being provided with an inlet and an outlet connection, which are connected to the conveyor, and having a motor in contact with the impeller, characterized in that each impeller 32 ) providing a connection between a first groove portion (322a) for communication between one side wall and the outer peripheral surface of the impeller (32) and the other outer wall surface of the impeller (32) comprising a second groove portion (322b) and an axially interconnecting groove portion (322c) extending axially but extending outwardly in radial direction thereof, and the first and second groove portions (322a, 322b) ) and the connecting groove part (322c) are formed between the side walls of each of the two adjacent blades (323) and each partition wall (321) is located between the first and second groove parts (322a, 322b) and their lower peripheral surface (3221). 3222), these lower peripheral surfaces (3221, 3222) are gradually converging radially from the inner part of the impeller (32) to the outer part. EN 215 991 Β and terminate at a circumferential width (k) on the circumferential surface (3230) at the outer end of the blade (323), the width of which is equal to or greater than the circumferential surfaces defining the connecting groove portion (322c). Priority: 03/04/92 A fuel pump according to claim 1, characterized in that the outer wall of the partition wall (321) has an end surface which has a lower peripheral surface (3221) of the first groove portion (322a) and a lower peripheral surface (3222) of the second groove portion (322b). connects. Priority: 03/04/92 A fuel pump according to claim 2, characterized in that said outer end surface is formed as a flat peripheral surface (3210). Priority: 03/04/92 Fuel pump according to claim 2, characterized in that the outer end surface is formed as a curved surface (163). Priority: February 24, 93 Fuel pump according to claim 2, characterized in that the impeller (32) is formed by casting, wherein the outer end surface (3210; 163) of the partition wall (321) and the first and second groove portions (322a, 322b) are formed. the lower peripheral surfaces (3221, 3222) and the lateral surfaces of the first and second groove portions (322a, 322b) and the connecting groove portion (322c) remain in the molding condition, while the outer peripheral surfaces (3230) and axial lateral surfaces (323) 3231, 3232) are finished after casting. Priority: February 24, 93 6. A fuel pump according to any one of claims 1 to 4, characterized in that it is mounted in a fuel tank (61) and is connected to a fuel injection device (64) of an internal combustion engine (70) and has a delivery capacity of 5-200 l / h and 2 to 5 kgf / cm ² , and the impeller (32) has a diameter (D) of 20 to 65 mm, and the typical size (Rm) of the impeller (32) and conveyor (33) is 0.4- 2 mm. Priority: February 24, 93 Fuel pump according to claim 6, characterized in that the ratio (L 1) of the radial dimension (L _t ) between the outer peripheral surface (3230) of the blade (323) and the total radial length (L ₂ ) of the blade (323). ^ Is between 0.1 and 0.6. Priority: February 24, 93 Fuel pump according to Claim 7, characterized in that the lower peripheral surfaces (3221, 3222) are spaced 0.2-0.8 mm apart at their radially outermost ends. Priority: February 24, 93 A fuel pump according to claim 1, characterized in that the blade (323) has a sloping surface (3231a) in the rotational direction of the impeller (32). Priority: 03/04/92 Fuel pump according to claim 2, characterized in that the connecting sections (3214, 3215) between the partition wall (321) and the adjacent vanes (323) have a smooth transition, preferably curved. Priority: 03/04/92 A fuel pump, preferably having a resin-like impeller wheel, comprising a series of vanes and dividers arranged alternately on the outer circumference of the impeller, the vanes being disposed in the circumferential direction of the impeller and two adjacent vanes they are projecting radially from the impeller, each partition having two inclined surfaces relative to the axial direction of the impeller, and having a rotatably bearing housing for the impeller, which defines a conveying passage along the outer circumference of the impeller, and with an inlet and outlet provided with a motor which is in contact with the impeller and has a motor in contact with the impeller, characterized in that the two sloping peripheral surfaces (3221, 3222) of the dividing wall (321) are formed, that their imaginary extensions intersect at a point (VI) located within the radial direction of the impeller (32) within the defining end face of the blade (323) and the outer peripheral surface (3210) of the dividing wall (321); 163), this is located inside the outer circumferential surface (3230) of the blade (323) in a radial direction and is connected to the two inclined peripheral surfaces (3221, 3222). Priority: February 24, 93 Fuel pump according to claim 11, characterized in that the outer peripheral surface (3210) is formed as a flat surface. Priority: February 24, 93 Fuel pump according to claim 11, characterized in that the outer peripheral surface is formed as a curved surface (163). Priority: February 24, 93 Fuel pump according to Claim 11, characterized in that the impeller (32) is molded, having the oblique outer surfaces (3221, 3222) of the outer front face (3210; 163) and the divider wall (321) and the blade boundary. its surfaces remain in the post-molding condition, however, the axial lateral surfaces (3231, 3232) of the blade (323) are formed by post-molding. Priority: February 24, 93 Fuel pump according to any one of claims 11 to 14, characterized in that it is arranged in a fuel tank (61) and is connected to a fuel injection structure (64) of an internal combustion engine (70) and has a transport capacity of 5-200 l. / h, transport pressure 2-5 kgf / cm ² , impeller (32) diameter (D) 20-65 mm, and impeller and conveyor channel (33) characteristic size (Rm) 0.4- 2.0 mm. Priority: February 24, 93 A fuel pump according to claim 15, characterized in that the radial between the outer peripheral surface (3230) of the blade (323) and the outer end of the partition wall (321) EN 215 991 Β, the ratio (Lj / L ₂ ) of the blade (323) to the total radial length (L ₂ ) of the blade (323) is 0.1-0.6. Priority: February 24, 93 A fuel pump according to claim 15, characterized in that the two inclined mantle surfaces (3221, 3222) are spaced at a distance of 0.2-0.8 mm from each other at their outermost edges. Priority: February 24, 93 Fuel pump according to claim 11, characterized in that each blade (323) is provided with a sloping surface (3231a) in the direction of rotation of the impeller (32). Priority: February 24, 93 A fuel pump according to claim 11, characterized in that the outer end surface of the partition wall (321) has sections extending towards the outer end of the blade (323), the outer front surface of the partition wall (321) and adjacent blades (321). 323) (3214, 3215). 10 Priority: February 24, 93