WO1999039105A1

WO1999039105A1 - Centrifugal pump impeller having a radial structure

Info

Publication number: WO1999039105A1
Application number: PCT/EP1998/008377
Authority: WO
Inventors: Ralf Stahlkopf
Original assignee: Tuchenhagen Gmbh
Priority date: 1998-02-02
Filing date: 1998-12-21
Publication date: 1999-08-05
Also published as: AU2513699A; DK1053402T3; EP1053402A1; EP1053402B1; JP2002502008A; DE59805263D1

Abstract

The invention relates to a centrifugal pump impeller having a radial structure which combines the advantages of a fully open pump impeller and a closed one, is substantially free from axial flow forces and in which the input conditions and therefore the suction characteristics are improved further in relation to known centrifugal pump impellers. To this end the invention provides for the covering disk segments (4.1; 4.3; 4.5; ...; 4.n-1) and rear side segments (4.2; 4.4; 4.6; ...; 4.n) to each have a through hole (4.1a; 4.3a; 4.5a; ...; 4.n-1a or 4.2a; 4.4a; 4.6a; ...; 4.na). All of said through holes start at an internal diameter defined by a radial formed hub (2a) which is outermost viewed in a radial direction and situated in the input area of the blade channels, and extend radially outwards.

Description

Centrifugal pump impeller of radial design

The invention relates to a centrifugal pump impeller of radial design, in which an even number of blades is provided, in which the individual blade channel between the blades delimiting it in each case is bordered either by a cover plate segment or by a rear surface segment, which alternate in succession, and in which each Blade channel is designed to be continuously open on its side opposite the respective top or rear surface segment from the inlet to the outlet region of the impeller.

A centrifugal pump impeller of the generic type is known from DE-A-195 30 195. It combines the properties of a so-called completely open and a closed impeller without taking on the disadvantages inherent in these impeller shapes. The advantages and disadvantages of open and closed impellers are briefly outlined below.

In the case of centrifugal pumps used in the food and beverage sector in particular, it is common to use so-called open impellers of radial design. Open impellers are those in which the top wall or the so-called cover disk is missing and the blades are only connected to the hub wall or the so-called rear surface ([*]: Pfleiderer / Petermann, Strόmungsmaschinen, 4th edition, 1972, pages 175, 285 ff). Open impellers have the advantage that, if necessary, the blade channels can be machined well (often ground or polished surfaces are required) and the blade channels can be viewed for inspection purposes. In addition, they are resistant to ^• door wheels that have a cover plate, easier to clean in the effluent (they are CIP-cleanable, which means "cleaning in place in the flow") as an open channel necessarily a smaller has a bordering surface and the narrow suction side space between the housing and the cover disc that is present on closed impellers is completely eliminated. With slow-running open wheels of radial design, however, the hydraulic efficiency is poorer than with closed wheels. 2

This is because the pressure difference between the front of the blade and the rear of the blade causes a gap loss (see [*], page 285 ff).

Open and closed centrifugal pump impellers of radial design have one common property: if no remedial measures are taken, they exert axial forces on the pump shaft (see [*], page 289 ff). Usual compensation measures, such as Sealing gap on the back of the impeller in connection with passage openings in the hub wall or rear surface of the impeller or the attachment of so-called back blades in the pressure side of the wheel side space (see [*], page 293 ff) are at least for centrifugal pumps that are used in the food and beverage sector , unsuitable and intolerable in most applications, since these compensation measures also create areas for cleaning that are difficult to access for CIP cleaning.

In the case of the impeller of the generic type (DE-A-195 30 195), the axial flow forces acting on it are almost balanced without any special additional precautions being required. This results, on the one hand, from the fact that adjacent vane channels are alternately closed by a cover plate or a rear surface segment and thus, at least in the radial extent of the vanes, there are sufficiently equal pressure distributions on both sides of the centrifugal pump impeller and no different pressure distribution in the circumferential direction anyway can arise. On the other hand, there is a limited penetration of the flow in the hub area, from the top surface to the rear surface side, as a result of which a reduction in the axial flow forces occurring in this area occurs.

In the aforementioned known impeller there are different entry conditions for the flow into the respective blade channels, which means that less favorable entry conditions exist for the blade channels delimited by cover surface segments, ie for the channels through which flow flows on the rear side of the impeller. This results from the fact that the 3 mente limited vane channels on their upstream side are fully open and thus the flow is unrestricted, while the flow into the rear vane channels must pass the limited inlet cross-section between the respective top surface segment and an outermost radial hub shape. The suction behavior of the impeller and thus the centrifugal pump is adversely affected by this fact. This affects, for example, the conveyance of hot cleaning agents, which must still be able to be sucked in by a centrifugal pump at temperatures of 80 ° C to 90 ° C. The so-called NPSH value to be provided by the centrifugal pump in such an application can hardly be achieved with the known impeller according to DE-A-195 30 195. In addition, the aforementioned different entry conditions and the resulting flow and pressure distributions result in axial forces, the further reduction of which is desirable.

A blood pump for circulating blood is known from EP-A-0 599 138, which also has all the features of a centrifugal pump of the generic type. In this known pump, openings are shown in the hub-side impeller area, which, however, only capture the rear surface segments there and engage in each case in one of the adjacent blades. Although this measure improves the entry conditions for the rear vane duct covered by cover plate segments, it does not create any more favorable conditions for reducing the axial forces.

It is an object of the present invention to provide a centrifugal pump impeller of radial design which combines the advantages of the completely open and those of the closed, which is largely free of axial flow forces and in which the entry conditions and thus the suction behavior compared to the centrifugal pump impeller of the generic type are further improved. 4

The object is achieved by applying the features of claim 1. Advantageous embodiments of the proposed centrifugal pump impeller are the subject of the dependent claims.

The advantages which can be achieved with the invention can only be achieved in a known manner to the full extent for an even number of blades, since it is only then ensured that there is no gap loss due to pressure difference between the front and rear of the blade. In those areas where the cover plate and rear surface segments alternate in each case, it is ensured that there is no connection between adjacent vane channels and therefore no gap flow that affects the hydraulic efficiency etc. negatively influenced, is given.

Since the centrifugal pump impeller does not have a closed cover disk and rear surface, seen over its circumference, but only has cover disk and rear surface segments, it does not form a ring-shaped, self-contained cover disk and in cooperation with the pump housing rear side wheel side space, which would be problematic in terms of proper CIP cleaning. In this respect, the advantages of the proposed centrifugal pump impeller go beyond those of the known open impeller of radial design, since here the rear side of the wheel side space is still present and is problematic in terms of cleaning technology.

In the hub-side impeller area, both the cover disk segments and the rear surface segments each have a passage opening, all of which start on an inner diameter, which is determined by an outermost radial hub shape in the radial direction, as seen in the radial direction, and which extend radially outward . This measure in accordance with the invention firstly opens the inlet area of the vane channels delimited by cover plate segments, as a result of which the inlet conditions for the flow are significant compared to those of the known rotor. 5

(DE-A-195 30 195) can be improved, on the other hand the aforementioned passage openings create an improved penetration of the flow from the inflow side to the rear of the impeller, whereby the compensation of axial flow forces on the impeller in this area is further improved and moreover , as a welcome side effect, an effective and safe cleaning of the mechanical seal behind the impeller is ensured.

If, as is provided according to a further proposal, the outermost radial hub shape is dimensioned in such a way that it connects the rear surface segments of the blade channels to one another over their entire area, this results in the most favorable conditions for a complete compensation of the axial flow forces in this area.

According to a further embodiment of the proposed centrifugal pump impeller, the axial projection surfaces of the passage openings are all congruent. This complete congruence of the through-openings also creates complete congruence of all surfaces that are effectively acted upon by pressure forces, both in the hub-side impeller region and in the region of the remaining top and rear surface segments, at least insofar as it is a force balance in the axial direction, as a result of which optimal conditions for the desired the greatest possible axial thrust compensation.

The proposed centrifugal pump impeller has optimized hydraulic properties, it can be manufactured both as a cast and as a shaped part (sheet metal part), and it has the necessary strength properties ^" if, as a further embodiment provides, it is designed in such a way that the passage openings each in addition to their edge by the outermost radial hub shape, which has a radial distance R2 from the axis of rotation of the impeller, by the contours of the two respective adjacent blades and on the other, radially on the outside, by a circular boundary with a radial distance R1 from the axis of rotation of the impeller be bordered. The radial boundary of the passage openings, as seen in the radial direction, allows the radial extension of the remaining top and rear surface segments to be adapted to the different hydraulic requirements as they occur in the context of a pump series (necessity of varying the flow rate and the delivery head) surrender. The proposed impeller can be reduced in the radial direction, starting from the largest outer diameter, until the top and rear surface segments connect the individual blades only in the form of a narrow ring surface that meets the minimum strength requirements.

According to a further proposal, the necessary strength of the impeller is ensured by the fact that the outer radius R _{a of} the impeller exceeds the radial distance R1 of the outside edge of the passage opening by a first radial distance Δ R, which results from the necessary strength of the impeller. In addition, the required range of variation of the delivery head adjustment of a nominal impeller diameter is ensured by providing a second radial distance ΔR _{V in} addition to the first radial distance ΔR _F. The outer radius R _{a of} the impeller thus results from the radial distance R1 of the outer edge of the passage opening, the first radial distance Δ RF and the second radial distance ΔR _V. An overall particularly advantageous embodiment of the centrifugal pump impeller of radial design is given when the first radial distance ΔR _F with 10 <ΔR _F <20 mm, preferably ΔR _F = 15 mm, and when the second radial distance Δ R _v with 20 < Δ R _v <30 mm, preferably Δ R = 25 mm are provided.

According to a further advantageous proposal, the blade channel is dimensioned at the design point of the impeller in such a way that there is no absolute speed in it at any point, the amount of which significantly exceeds a speed of 2 m / s. This ensures, on the one hand, a gentle delivery of the respective fluid and, on the other hand, a favorable suction behavior and thus a correspondingly high NPSH value. A gentle 7

de Funding is always welcome when the proposed centrifugal pump is used in the food and food sector. Good pumping speed is always required of it when it has to convey hot liquids. This is always the case, for example, if it acts as a detergent pump and detergent with a temperature of 80 ° C to 90 ° C still has to be sucked in.

The pumping speed and thus the NPSH value are favorably influenced if, as is also provided for in another proposal, the respective pressure edge of the blades is designed in the design point of the impeller on the one hand so that the flow leaves the pressure edge without bumps and on the other hand the The respective suction edge of the blades is designed in such a way that the flow only flows smoothly towards the suction edge at a flow rate that significantly exceeds the flow rate at the design point. This means that the suction edge at the design point of the centrifugal pump works almost in part-load mode, while the

The pressure edge of the blade experiences a bumpless under these conditions and therefore an optimal outflow for the flow at the design point. In this context, it has proven to be particularly advantageous if there is a smooth flow against the suction edge at a flow rate that exceeds the flow rate at the design point by approximately twice.

An embodiment of the centrifugal pump impeller according to the invention with an exemplary number n = 6 blades is shown in the drawing and is described below. Show it

Figure 1 is a view of the centrifugal pump impeller in the direction of its flow, with the cover disc segments cross-hatched, the rear surface segments left-handed and the projected hub area are hatched right-handed to better understand the impeller geometry.

FIG. 2 shows a meridian section through the centrifugal pump impeller according to FIG. 1 in accordance with the section profile identified there by AA; 8th

Fig. 3 is a perspective view of the centrifugal pump impeller and

4 shows a meridian section through a portion of the centrifugal pump impeller according to FIG. 1, the impeller being shown in assembly with the pump shaft and the pump housing.

A centrifugal pump impeller 1 (FIG. 1) equipped with n = 6 blades 3.1 to 3.6 inevitably also has six blade channels __ to 5 6. From the view representation according to FIG. 1 in conjunction with the meridian section according to FIG. 2 and the perspective representation according to FIG. 3 (the designations entered there are limited to the adjacent blade channels _____ and 5 _j 2), it becomes clear that according to the selected nomenclature the Blade channel _____ is partially closed by a cover plate segment 4.1, the blade channel S_Z by a cover plate segment 4.3 and the blade channel 5.5 by a cover plate segment 4.5 on the upstream side of the centrifugal pump impeller 1 (cross-hatched areas). In contrast, the blade channel 5 ^ 2 is partially closed by a rear surface segment 4.2, the blade channel 5 ^ 4 by a rear surface segment 4.4 and the blade channel 5> by a rear surface segment 4.6 on the side facing away from the upstream side of the centrifugal pump impeller 1 (hatched areas on the left). It can also be seen that the proposed arrangement of cover disk and rear surface segments 4.1, 4.3, 4.5 or 4.2, 4.4, 4.6, which, viewed in the circumferential direction, follow each other in alternation, cannot form an edge Front and back of the concerned

Blade - and thus two adjacent blade channels - connects to each other in a short way. This effectively prevents gap flows and efficiency-reducing effects due to a pressure difference between the front and back of the blade. 9

The cover plate segments 4.1, 4.3 and 4.5 and the rear surface segments 4.2, 4.4 and 4.6 each have a passage opening 4.1a, 4.3a, 4.5a or 4.2a, 4.4a, 4.6a, all of which start on an inner diameter with radius R2, the is determined by an outermost radial hub shape 2a, seen in the radial direction, in the inlet area of the blade channels 5/1 to δ_6 (hatched areas on the right), and which extend radially outward (see also FIG. 2). In addition to their edges due to the outermost radial hub formation 2a, the through openings 4.1a to 4.6a are formed on the one hand by the contours of the two respective adjacent blades (for example the passage opening 4.2a by the blades 3.2 and 3.3) and on the other, radially on the outside a circular boundary with a radial distance R1 from the axis of rotation of the impeller. (Figures 1 and 2). The illustrated configuration of the through openings 4.1a to 4.6a is selected such that the axial projection surfaces of these through openings are all congruent.

The radial extent of the top and rear surface segments 4.1 to 4.6 between the radial distance R1 and the outer radius R _{a of} the impeller 1 can, if this is necessary as part of the adaptation of the impeller within the centrifugal pumps of a series, to be used by different impeller diameters to realize. Turning the impeller 1 by a second radial distance ΔR (preferably approx. 25 mm) is possible in the present case, for example, up to an outer radius that is larger than the radial distance dimension R1 by a first radial distance ΔR _F (preferably approx. 15 mm) is. In this case, the necessary strength requirements are met, which also include the avoidance of oscillations and vibrations of the blades. With the aforementioned second radial distance .DELTA.R, the required variation in the head height adjustment of a nominal impeller diameter can be sufficiently achieved.

Proceeding from the upstream peripheral surface of a hub 2, in the entry area of the hub defined by the respective cover disk segment 4.1, 4.3 and 4.5 10

ten blade channels 5J., 5 ^ and ____, seen in the radial direction, the outermost radial hub formation 2a is provided, which extends radially outwards so far that it covers the rear surface segments 4.2, 4.4 and 4.6 of the blade channels 5.2. 5.4 and 5.6 area-wide and fully connected with each other (see Figure 1). As is generally customary, the hub 2 has a hub bore 2b and a driving groove 2c.

As a result of the possibility of the flow reaching through between the impeller front and impeller rear in the area of the passage openings 4.1a to 4.6a, there are inevitably no axial flow forces on the impeller 1 (see FIGS. 2 and 4). Even if the respective pressure distribution on the top surface and rear surface segments 4.1, 4.3, 4.5 or 4.2, 4.4, 4.6 would be different due to fluid-mechanical effects, these differences could change due to the arrangement of the passage openings 4.1a according to the invention up to 4.6a on these surfaces, seen over the circumference, do not train. This means that there are no axial flow forces on the centrifugal pump impeller 1 as a whole and thus on the pump shaft. The centrifugal pump impeller 1 according to the invention, seen in itself, has a largely equalization of forces, which results in the advantages mentioned above.

When the impeller is installed (FIG. 4), the advantages mentioned above become particularly clear once again. In addition to the mentioned axial thrust compensation, which results from the special impeller geometry in the hub and suction area of the impeller 1, the illustration shows that the distance a given between the suction edge S in the area of the cover plate segment 4.5 and the outermost radial hub shape 2.a. there is a passage of the flow R (for example cleaning liquid) into the area of the mechanical seal 9, with a slide ring 9a and a counter ring 9b. This ensures safe and reliable cleaning of a very critical area. The impeller 1 is not on a shaft nut 10 on one 11

specified recess of a pump shaft 6, the counter ring 9b being fixed between the recess and the impeller 1. So that the fluid delivered by the impeller 1 does not reach the area between the shaft 6 and the hub 2, there is a seal 11 between the impeller 1 and the counter ring 9b, a further seal 12 between the shaft nut 10 and the impeller 1 and a third seal 13 between the counter ring 9b and the shaft 6 provided.

The impeller 1 is bordered on its upstream side by a front pump housing 7 and on the opposite side by a rear pump housing 8. The front and rear pump housings 7 and 8 are in the area of the cover disk and rear surface segments 4.1; 4.3; 4.5; ... 4.n-1 and 4.2; 4.4; 4.6; ... 4.π symmetrical and equipped with the same projection surfaces, so that in cooperation with the impeller 1, which is also symmetrically constructed and has the same surface area with respect to the rear surface and cover surface segments, the above-described largely equalization of axial forces is achieved. While conventional impellers 1, which are comparable in terms of their performance, experience an axial force in the order of magnitude of, for example, 900 Newtons, only axial forces of approximately 35 Newtons remain on the impeller 1 according to the invention. Since conventional standard motors, which are suitable for driving such rotary pumps, can accommodate a permissible axial force of about 350 Newtons via their shaft bearings, it becomes clear that when using the impeller 1 according to the invention, no special measures are required to absorb these low unbalanced axial forces are. In addition, it is clear that the use of the proposed impeller 1 brings with it a considerable cost advantage, since the unbalanced axial forces can be absorbed without any problems in the mounting of the standard motor.

In the exemplary embodiment of impeller 1, the direction of rotation is identified by U (FIG. 1); in this case it is a backward curved blading, the respective blade being simply curved in itself in the exemplary embodiment. However, the principle according to the invention is also straightforward 12

applicable to purely radially oriented or forward curved displays. The three blading shapes can in turn each have spatially curved, that is to say three-dimensional, blades. The selected embodiment also relates to a slow-running centrifugal pump impeller of the radial type (so-called slow rotor); however, the proposed invention extends without restriction to centrifugal impellers (so-called center rotors).

Claims

13 Patent claims

1. Centrifugal pump impeller of radial design, in which an even number n blades (3.1 to 3.n) is provided, in which the individual blade channel (5.1 to 5Ln) between the blades (3.1; 3.2; 3.3; ...; 3.n) either from a cover segment (4.1; 4.3; 4.5; ...; 4.n-1) or from a rear surface segment (4.2; 4.4;

4.6; ...; 4.n), which follow each other in alternation, is bordered, and in which each vane channel is designed to be continuously open on its side opposite the respective top or rear surface segment from the inlet to the outlet region of the impeller (1), characterized in that the cover plate segments (4.1; 4.3; 4.5; ...;

4.n-1) and the rear surface segments (4.2; 4.4; 4.6; ...; 4.n) each have a passage opening (4.1a; 4.3a; 4.5a; ...; 4.n-1a or 4.2a; 4.4a; 4.6a; ...; 4.na), all of which begin on an inner diameter, which is determined by a radial hub shape (2a) seen in the radial direction in the entry region of the blade channels, and which are radial n extend outside.

2. Centrifugal pump impeller of radial design according to claim 1, characterized in that the outermost radial hub shape (2a) the rear surface segments (4.2; 4.4; 4.6; ...; 4.n) of the blade channels (5.2; 5.4; 5.6; .. .: 5.n areal and fully connected with each other.

3. Centrifugal pump impeller of radial design according to claim 1 or 2, characterized in that the axial projection surfaces of the through openings (4.1a to 4.na) are all congruent. 14

4. Centrifugal pump impeller of radial design according to one of claims 1 to 3, characterized in that the passage openings (4.1a to 4.na) each in addition to their boundary by the outermost radial hub formation (2a), which is a radial distance (R2) from the axis of rotation of the impeller (1), through the contours of the two respective adjacent blades (3.i; 3.i + 1) and, radially on the outside, from a circular boundary with a radial distance (R1) from the axis of rotation of the impeller (1) be bordered.

5. Centrifugal pump impeller of radial design according to one of claims 1 to 4, characterized in that the outer radius (R _a ) of the impeller (1) the radial distance (R1) of the outside edge of the passage opening (4.1a to 4.na) by one exceeds the first radial distance (ΔR _F ), which results from the necessary strength of the impeller (1), plus a second radial distance (ΔR _V ), which results from the required range of variation of the head adjustment of an impeller diameter.

6. Centrifugal pump impeller of radial type according to claim 5, characterized in that the first radial distance (ΔR _F ) with 10 <(ΔR _F ) <20 mm, preferably (ΔR _F ) = 15 mm, and in that the second radial distance (ΔRv) with 20 <(ΔR _V ) <30 mm, preferably (ΔR _V ) = 25 mm, are provided.

^• 7. Centrifugal pump impeller of radial design according to one of claims 1 to 6, characterized in that in the design point of the impeller (1) the blade channel (5.1 to 5}) is dimensioned in such a way that there is no absolute speed in it at any point Amount significantly exceeds a speed of 2 m / s. 15

8. Centrifugal pump impeller of radial design according to one of claims 1 to 7, characterized in that in the design point of the impeller (1) on the one hand the respective pressure edge (D) of the blades (3.1 to 3.n) is designed such that the flow of the pressure edge ( D) leaves bumpless, and on the other hand the respective sucking edge (S) of the blades (3.1 to 3.n) is designed such that the flow is bumpless at a flow rate that significantly exceeds the flow rate at the design point inflows.

9. Centrifugal pump impeller of radial design according to claim 8, characterized in that the smooth flow against the suction edge (S) at a

Flow rate is given that exceeds the flow rate at the design point by about twice.