EP1573208A1

EP1573208A1 - Intake duct

Info

Publication number: EP1573208A1
Application number: EP03813089A
Authority: EP
Inventors: Stephan Bross; Isabel Goltz; Peter Amann
Original assignee: KSB AG
Current assignee: KSB AG
Priority date: 2002-12-17
Filing date: 2003-10-23
Publication date: 2005-09-14
Anticipated expiration: 2023-10-23
Also published as: JP4312720B2; SI1573208T1; ES2344942T3; US20050265866A1; US7798772B2; EP1573208B1; DE10258922A1; DE50312675D1; WO2004055381A1; DK1573208T3; PT1573208E; CN100507282C; JP2006509948A; ZA200504431B; CY1110708T1; ATE466197T1; CN1726347A

Abstract

A centrifugal pump with a housing having one or more impellers having an axial or semiaxial, open or closed design disposed therein and an intake channel mounted upstream of the first impeller. A plurality of grooves that are distributed around the circumference and extend in the direction of flow are arranged within the wall area of the intake channel. In the housing wall of the intake channel there is a closed annular wall area constructed between a point of entry of the first impeller and the proximate ends of the grooves, whereby the grooves are operatively connected exclusively with the space in the intake channel.

Description

Saugkana!

description

The invention relates to a centrifugal pump, in the housing of which one or more impellers of axial or semi-axial, closed or open type are arranged and a suction impeller is preceded by a first impeller, in the wall surface of which several grooves are arranged distributed over the circumference.

In the case of specifically high-speed pumps, there is often a significant, locally limited increase in the associated NPSH curve in the delivery volume range of 65 - 80% of the design volume flow. Sometimes, depending on the type of pump, the associated course of the Q-H characteristic curve can additionally have an instability, which is generally referred to as a characteristic curve kink or as a saddle.

Such characteristic curve shapes are caused by the formation of the so-called partial load vortex, which occurs when the volume flow is reduced in the outer region of an impeller inlet. A partial load vortex has a significant influence on the impeller inflow, which, under its influence, blocks the meridional flow cross-section and high speed components in the direction of impeller rotation (constant twist).

DE 25 58 840 C2 discloses a solution for avoiding the disadvantages of a part-load vortex, a diffuser being arranged in front of an impeller inlet. With this solution, a partial load vortex is reversed in its direction of action before it can reach components located in front of the impeller inlet and cause damage to them.

Other measures for influencing a part-load vortex are described in EP 1 069 315 A2, in particular when evaluating the prior art. The measures "Casing Treatment, Separator or Active Control" either require additional units in the machine periphery (Active Control), reduce the efficiency even at the best point of the machine (Casing Treatment) or involve increased design effort (Separator). The script itself beats the use of a large number of “grooves” which, according to the literature “An Improvement of Performance-Curve Instability in a Mixed-Flow Pump by J-Grooves”, May 29-June 1, 2001, New Orleans, Louisana, FEDSM 2001- 18077, Proceedings of 2001 ASME Fluids Engineering Division Summer Meeting (FEDSM ^X 01), are generally referred to as J-Grooves due to their bent J-shaped course.

The J-Grooves are flat, in another version also spatially curved grooves, which are installed in the pump housing in the direction of flow in front of and above the impeller blades that are to be designed as open at the impeller inlet. For the functionality of the grooves, it is important that they partially cover the outside diameter of the impeller. In the area of the impeller overlap, the impeller must be designed to be open in order to obtain a connection between a fluid zone provided with higher pressure in the area of the open impeller blading and the beginnings of the J-grooves arranged above it. This constructive measure creates a fluid-carrying connection to the upstream inflow zone via the J-Grooves. Due to the J-Grooves arranged in the main flow direction, the open impeller blading permanently conveys a partial flow of already conveyed fluids in front of the impeller and back into the area of the impeller inflow. The disadvantage of these J-Grooves is that their return is constantly active across the entire travel range of the turbomachine. As a result, the peak efficiency of a turbomachine equipped with it drops. Another disadvantage is the interaction between the free impeller blade tips and the opposite, housing-fixed groove parts of the J-Grooves, which leads to increased noise and vibration phenomena. Their reduction is described in the aforementioned literature reference, page 2, in connection with FIG. 3 and its associated explanation. For this purpose, the ends of the J-Grooves arranged above the free blade tips are connected to each other by a circumferential ring groove. This additional ring groove in the housing ensures pressure equalization between the individual J-Grooves on the end faces. And the arrangement of such spatially curved J-Grooves, which extend from the inlet area with a constant diameter in a kinking manner into a conical housing wall surface, requires a high manufacturing outlay. This type of partial load vortex control is associated with considerable disadvantages.

The invention is based on the problem of achieving a simple possibility for improving the NPSH behavior and for improving the part-load behavior in the case of specifically high-speed centrifugal pumps with impellers of semi-axial or axial, open or closed design. At the same time, the problem should be solved of being able to make a subsequent improvement in a simple manner in the case of centrifugal pumps already in use, without negatively influencing the operating behavior in the normal operating range of the centrifugal pump.

The solution to this problem provides that grooves are made in the housing wall of the suction channel and that a closed annular wall surface is formed between an impeller inlet of the first impeller and the nearest ends of the grooves, the grooves being in operative connection only with the suction channel. A first impeller is designed as a suction impeller. The closed, annular wall surface formed in the housing wall of the suction channel is between the ends of the grooves located in the inflow direction before the impeller inlet and the arranged wheel entry of the first impeller. Such a suction impeller can have a specific high speed nq> 70 min ^"1 .

With this solution, the optimal operating point of a centrifugal pump remains unchanged and, like the other operating points, is in no way adversely affected. A part-load vortex, also known as a large pre-rotation vortex, which forms in part-load operation, is weakened with the help of the elongated depressions. The elongated grooves cause energy to be transferred by friction from the area of the part-load vortex near the wall to many small vertebrae that form in the grooves. This energy transfer, which only occurs in part-load operation, drastically reduces the peripheral component and thus the intensity of the resulting part-load vortex, and consequently improves the part-load behavior of the centrifugal pump. Since the grooves only develop their energy-dissipating effect in cooperation with a part-load vortex emerging from the impeller, the impeller inflow remains unaffected for the other operating points. There is no negative impact on normal impeller inflow, and there is therefore no negative impact on the efficiency curve. In contrast to the previously known solutions in the form of the J-grooves, the invention does not mix a flow conveyed back by the impeller via the grooves with a main flow flowing to the impeller.

By deliberately avoiding any feeding of high-energy medium into the grooves, a disturbance of the impeller inflow is prevented in normal operation. Only when the impeller induces a disturbance in the form of the part-load vortex that develops does a kind of interaction between the grooves and the part-load vortex begin. This interaction leads to self-regulation. The energy of the part-load vortex in the grooves is dissipated by the formation of a large number of small groove vertebrae, which causes a considerable weakening of the part-load vortex. This function can only be achieved if the in the suction channel in front of the impeller chen groove ends are reliably cut off by a ring-shaped closed wall surface from a supply of already pumped fluids.

An embodiment of the invention provides that the grooves are arranged between web-shaped configurations of the housing wall of the suction channel. In those applications in which processing of a suction channel is not possible or only possible with considerable difficulty, an annular insert containing the grooves or webs can also be inserted into an existing suction channel of a pump.

Such an insert enables simple mechanical production of the grooves and can be easily installed in the suction channels of pumps that are to be newly manufactured or have already been delivered. As a result of the small groove depths of just a few millimeters, which are only formed in the area of the boundary layers close to the wall, such an insert is able to achieve an improvement in the part-load behavior even with centrifugal pumps that have already been delivered or installed in systems. For this purpose, only the suction channel receiving the insert may have to be slightly expanded in the inside diameter in order to be able to accommodate an appealing diameter size of a grooved insert. A type of modular system is used here in order to enable such an insert to be used for a large number of pump types by means of a clever diameter gradation.

According to one embodiment of the invention, the closed annular wall surface has an axial extent that is dependent on the intensity of a part-load vortex. The length of the axial surface is at least so large that interference between the impeller blades at the impeller inlet and the groove ends arranged in front of it is reliably suppressed. This prevents the generation of disturbing noises and vibrations in the simplest way. On the other hand, the length of the axial ring surface is not chosen to be greater than the extent of the slowly forming, still harmless part-load vortex. Only when the training Partial load vortex receives a greater intensity, it is possible that its so-called release line detaches from the impeller and jumps over the closed annular wall surface. As a result, the partial load vortex emerges completely from the impeller. It is directed against the inflow and rotates around the machine axis in the direction of rotation of the impeller. As a result of the tangential overflow of the depressions and the formation of many small vortices in the depressions, a large part of the energy in the part-load vortex is dissipated and the effect of the part-load vortex is drastically weakened.

According to further refinements of the invention, the closed annular wall surface has an axial extent which is dependent on the intensity of a part-load vortex and is on the order of 0.005-0.02 times the impeller inlet diameter. And the lengths of the grooves or webs are on the order of 0.03 - 0.5 times the impeller inlet diameter. The depths of the grooves or the heights of the webs are of the order of 0.005-0.02 times the impeller inlet diameter.

And according to another embodiment of the invention, the product of groove width b times the number of grooves n corresponds to a ratio of

n -b = 0.45-0.65-π -D

Embodiments of the invention are shown in the drawings and are described in more detail below. They show

Fig. 1 NPSH curves of generic centrifugal pumps, which are equipped with and without grooves, the

Fig. 2 is a flow diagram of a return flow area on an axial pump with an open impeller in normal operation, the

Fig. 3 is a flow diagram on a semi-axial and axial pump with ge closed impeller and in normal operation, the

Fig. 4 is a flow diagram of a part-load vortex on an axial pump in

Part load operation, the

Fig. 5 different speed triangles in a cylindrical section

Axial machine when the part load vortex emerges from the impeller

6, using a cylindrical section, the flow profiles of a part-load vortex in the grooves

Fig. 7 is an illustration of the flow in the grooves and

Fig. 8 + 9 Q-H and NPSH curves with improved characteristics.

Fig. 1 shows an example and with a dash-dotted line a typical NPSH curve of centrifugal pumps with high-speed impellers of axial or semi-axial design. The values for the delivery rate Q are plotted on the abscissa and the values for the NPSH are plotted on the ordinate. It can be seen that at the operating point Qo _pt , the best point of the delivery rate, the NPSH has a low value. In part-load operation, however, the NPSH curve is characterized by a local increase, the so-called NPSH peak, which limits the operating range at Q _m j _n for a given system, given in dashed lines, the maximum permissible NPSH _A value. Operation below this operating point is not permitted, as otherwise cavitation-related conditions occur within the pump that do not allow continuous operation.

In the diagram, a further NPSH curve is drawn with a solid line, which corresponds to a centrifugal pump with the same operating points, but in whose suction channel grooves according to the invention are additionally provided. The one Centrifugal pumps designed in this way convincingly illustrate the much more favorable NPSH properties. The local NPSH increase typical for part-load operation is still there, but it is at a significantly lower level than that of a pump without grooves. Such an improved pump has a much wider operating range.

2 shows the existing flow conditions at the best point Q _op t of a centrifugal pump 1 using the example of an open axial wheel. An impeller 2 rotates in a housing 3. During the rotational movement of the impeller 2, a return flow region R rotating with the impeller forms in the form of a weak vortex flow between the housing 3 and the free blade tips 4 of the impeller 2. This backflow R is caused by the pressure exchange between the flow areas of adjacent blade channels and the pressure equalization between the suction and pressure sides of blades 5 in the area of free blade tips 4. Such a backflow region R rotating with the impeller 2 occupies approximately a zone which corresponds to a blade width B.

This backflow region R has a flow direction, which is shown by arrows, along the housing wall 6 and runs opposite to the impeller flow LA. At the point at which the return flow region R reverses its direction of flow, a so-called separation line SL is shown. To a certain extent, this is a boundary line which runs on the circumference of the housing wall 6. In the area of this line SL, the energy of the impeller inflow LA is greater than the energy of the return flow region R and therefore its flow reversal. In the case of pumps with open semi-axial or axial impellers, such a backflow area R exists over the entire operating range and is also available in the area of the best efficiency.

According to FIG. 3, there is a similar backflow area with two different types of closed impellers. 3 shows the relationship nisse in a semi-axial pump design, while the situation below is shown in the case of an axial pump. In these impellers, a so-called cover disk 7 avoids an exchange of energy via the blade tips 4 and between the suction and pressure sides of an impeller blade 5. For this purpose, there is a small gap flow LF between the housing wall 6 and the cover disk 7 for such impellers 2, for which the pressure difference in front and behind the impeller is responsible. Leakage losses of this type are considerably reduced by means of correspondingly small clearance gaps between cover plate 7 and housing wall 6.

4 shows, using the example of an open impeller 2, the formation of a part-load vortex PLV that occurs in part-load operation. This and the following explanations also apply to a closed-design impeller. Such a partial load vortex PLV rotating with the impeller comes out of the impeller 2 in the region of the impeller outer diameter D at the impeller inlet edges 8 and counter to the impeller inflow LA and flows back into the suction channel 9. When the rotating part-load vortex PLV occurs, there is a strong, unsteady interaction between the impeller inflow and the flow around the blade, which is manifested in particular by an abrupt increase in the NPSH values. The magnitude of this increase depends on the intensity of the part-load vortex that forms. The positions X and Y encircled in FIG. 4 are details and serve to represent the speed triangles of FIG. 5. A plurality of grooves 10 are distributed over the circumference and arranged in front of the impeller 2 in the wall surface 6 of the suction channel 9.

FIG. 5 shows the speed relationships of a formed part-load vortex PLV at the points X and Y of FIG. 4. The point X shows the speed conditions in the area of the part-load vortex PLV emerging from the impeller 2 and the position Y shows the conditions in the distance from the wall again Impeller 2 entering partial load vortex PLV. For the illustration, the speed triangles are drawn at positions X and Y, which are made up of the direction and size arrows for the absolute speed c, the relative speed w and the peripheral speed u.

At point X, the absolute speed c _x results from the peripheral speed u _x close to the wall of a blade 5 and from the backward flowing relative speed w _{x of} the partial load vortex PLV emerging from the impeller and is characterized by a high peripheral component c _ux . The arrows with the speed indication c ~, on the other hand, symbolize within the suction channel 9 the undisturbed inflow to the impeller with the blades 5 shown here in section and having a profile.

Analogously to this, a speed triangle is drawn at Y, which is given at position Y in the area of the entry point of the partial load vortex PLV into the impeller 2. Since the entry point Y is of a smaller diameter, the peripheral speed u _{y is} correspondingly lower. And as a result of the partial-load vortex PLV, which is weakened in its energy, its absolute speed c _y is also correspondingly lower, which results in a relative speed w _y , which in this example runs to a certain extent offset by 90 ° to the relative speed w _{x of} an emerging current thread of the partial-load vortex PLV.

The reason for the weakening of the partial load vortex PLV is in particular the circumferential component c _ux , which leads to a tangential overflow of the axially parallel grooves 10, as shown in FIG. 4 and in FIG. 6, the top view of a development of the housing wall 6. The outer blade ends 4 permanently run past this wall surface of the housing wall 6. In the housing wall 6 several grooves 10 are arranged distributed over the circumference, which run in the direction of the impeller inflow c∞. Of the grooves 10 running in the inflow direction and arranged in the wall surface 6 of the suction channel 9, their groove ends 11 are arranged at a distance in front of the blade leading edge 8 on the outer diameter D of the impeller 2. The beginning of these grooves 10, which run in the inflow direction or are parallel to the axis, is not shown here, since the length of the grooves 10 depends on the conveying quantities and the impeller design is selected. The lengths of these grooves 10 are on the order of 0.03-0.5 times the impeller inlet diameter. In normal operation, an inflowing fluid flows through the grooves 10 without adversely affecting the operating behavior of the centrifugal pump.

Furthermore, different separation lines SLi, SL ₂ and SL _{3 are} shown in broken lines in FIG. 6. The detachment lines SLi, SL ₂ show the suction-side limits of a backflow region R which is developing under different operating states. In the area of the best point Q _opt , the detachment line SLi lies within the width of the impeller blades 5 and moves with increasing partial load operation in front of the impeller or blade entry edge 8 to the detachment line SL ₂ . In normal operation, the position of this detachment line SL ₂ always remains in front of the impeller 2 in the region of a closed annular wall surface 12. This wall surface 12 ensures that the fluid material flowing back from the region R cannot enter the grooves 10. The length L of the wall surface 12 extending in front of the impeller inlet and up to the groove ends 11 is considered in the order of magnitude which corresponds to the ratios of 0.005-0.02 x impeller inlet diameter. In the example of an axial wheel used here, the impeller inlet diameter usually corresponds to the outer wheel diameter D. In the case of a semi-axial wheel, it is correspondingly smaller. And with a closed impeller, it corresponds to the diameter up to the inside diameter of a cover disk 7.

Only when the part-load vortex PLV is formed does the detachment line SL ₂ jump over the closed annular wall surface 12 and reach the wall surface 6 provided with grooves 10. The limit of an axial expansion of the part-load vortex PLV which then occurs is represented by the detachment line SL ₃ .

If the partial load vortex PLV reaches a correspondingly high energy, it jumps over the annular, closed wall surface 12 located in front of the impeller and flows back into the suction channel 9. As a result of the predominantly circumferential direction Absolute speed component c _ux , the partial load vortex PLV formed in the suction channel 9 flows primarily tangentially over the grooves 10. His swirl energy is dissipated in many small vortices that form within the grooves 10. In the part-load vortex PLV, this leads to a withdrawal of speed energy, so that the part-load vortex PLV is weaker overall and its axial and radial extent is considerably reduced. It therefore only extends to the detachment line SL ₃ , at which the partial load vortex PLV is reversed. Due to the simultaneous reduction in the swirl component of the part-load vortex, in addition to the reduction in the NPSH increase, the characteristic stability of the centrifugal pump at part-load is also decisively improved. The mode of operation of the grooves 10 is thus based on an energy transfer by means of friction from a large pre-rotation vortex in the form of the partial load vortex PLV to many small vertebrae, which are each located in the grooves 10.

FIG. 7, a section along the line AA of FIG. 6, shows the formation of many small energy-dissipating vortex systems 13 within the grooves 10. The reason for the many small vortex systems 13 is the circumferential component c _{ux of} the part-load vortex flow, which is tangential to the direction of the groove.

A comparison is shown in the associated diagrams of FIGS. 8 and 9. In the illustration in FIG. 8, the curve shape drawn in dash-dotted lines corresponds to the QH characteristic curve of a centrifugal pump without grooves in the suction channel. From the marked operating point Q _P v, the QH curve shows a clear kink in the characteristic. The delivery head decreases towards smaller quantities. The reason for this is the effect of a developing partial load vortex PLV. In contrast, the solid QH characteristic curve shows an increasing curve without a characteristic curve kick. This is the characteristic of a centrifugal pump, the suction channel of which is provided with channels or grooves 10 ending at a distance from the impeller. The dash-dotted curve shape with the characteristic curve kink is due to the formation of a part-load vortex and the resulting impairment of the impeller inflow. In contrast, a curve drawn as a solid line was obtained for the same pump if grooves 10 were provided in front of the suction impeller in the wall surface 6 of the suction channel 9. The matching curves in the normal operating range to the right of QPLV convincingly demonstrate the mode of operation of the grooves during normal operation.

In FIG. 9 arranged below FIG. 8, the associated NPSH curves are drawn. The NPSH curve shown in broken lines corresponds to a pump in whose suction channel 9 no grooves are arranged. In contrast, the solid characteristic curve shows a pump, in the suction channel 9 of which several grooves 10 are arranged. The NPSH behavior of such a pump is decisively improved by the partial load vortex PLV, which is greatly reduced in its effect by the grooves 10. This NPSH curve no longer exceeds the predetermined value NPSHA system and thus does not NPSH-related operating limit Q _m i _n any more. The type of energy reduction of the partial load vortex PLV and the resulting reduced transient interaction, especially in the operating area around PLV, result in improved flow conditions, as a result of which the NPSH behavior is improved and a pump characteristic curve is stabilized.

It is thus the merit of the inventors to have recognized that a profile arranged in the housing wall of the suction opening / inlet opening at a distance in front of the impeller has a braking effect only on a part-load vortex emerging from the impeller in part-load operation. An additional surprising effect has been the unchanged noise behavior of the centrifugal pump. Pumps that have already been delivered and installed in systems can be converted easily, as their noise behavior remains at their previous level.

Claims

claims

1. Centrifugal pump, in the housing of which one or more impellers of axial or semi-axial, open or closed type are arranged and a first impeller is preceded by a suction channel, in the wall surface of which are arranged several grooves distributed over the circumference and running in the direction of flow, characterized that a closed annular wall surface (12) is formed in the housing wall (3) of the suction channel (9) between an impeller inlet of the first impeller (2) and the nearest ends (11) of the grooves (10), the grooves (10 ) are only operatively connected to the room in the suction duct.

2. Centrifugal pump according to claim 1, characterized in that the grooves (10) between see web-shaped configurations of the housing wall (3) are arranged.

3. Centrifugal pump according to claim 1 or 2, characterized by an insert, in particular a thin-walled, annular, grooves (10) or webs having element.

4. Centrifugal pump according to claim 1, 2 or 3, characterized in that the closed annular wall surface (12) has an axial extension in the order of magnitude of 0.005-0.02 times the impeller inlet diameter depending on the intensity of a partial load vortex (PLV).

5. Centrifugal pump according to one of claims 1 to 4, characterized in that the lengths of the grooves (10) or webs are in the order of 0.03 - 0.5 times the impeller inlet diameter.

6. Centrifugal pump according to one of claims 1 to 5, characterized in that the depths (t) of the grooves (10) or the height (h) of the webs are in the order of 0.005 - 0.02 times the impeller inlet diameter.

7. Centrifugal pump according to one of claims 1 to 4, characterized in that the product of groove width b times the number of grooves n a ratio of

n -b = 0A5-0.65-π-D

equivalent.