Analysis of turning blades in front of cooling water pump suction bell

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International Rotating Equipment Conference 2012, Düsseldorf

Technical Paper

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Analysis of turning blades in front of cooling water pump suction bell Author: Rok Pavlin, Dipl.Ing.Mech. Manager of Pumps Program Turboinštitut 1000 Ljubljana, Slovenia Co-Author 1: Co-Author 2: Franci Vehar, Dipl.Ing.Mech. Aljaž Škerlavaj, Ph.D. Researcher Senior Researcher Turboinštitut Turboinšitut 1000 Ljubljana, Slovenia 1000 Ljubljana, Slovenia Co-Author 3: Franc Kranjc, Dipl.Ing.Mech. Mechanical Maintenance Engineer Nuclear Power Plant Krško 8770 Krško, Slovenia

Copyright: Pumps + Systems Association within VDMA e. V. Lyoner Straße 18, D – 60528 Frankfurt

Phone: +49 (0) 69/66 03-12 86, Email: [email protected]


Summary There are three vertical mixed flow cooling water pumps installed in a nuclear power plant. At the maxi-mum plant output all three of them are in operation. If one is switched off, the plant output must be re-duced. All three cooling water pumps needed to be overhauled after each cycle (18 months), what was too often comparing with similar plants. The main reasons were cracks which were often noticed at the impeller blades inlet near the hub. Sometimes a cooling water pump was switched off also between two planed overhauls due to the broken blade. Besides, air bubbles were noticed at the cooling water outflow. The plant staff put forward the goal to increase pumps reliability and to increase the time between two pump overhauls on three cycles. The supposition was, that the pumps sucked the air through the air entraining vortices. Pump sump is not “classical” according to recommendations (1), but pumps have turning vanes in front of suction bell inlet. Detailed analysis of existing situation was carried out on hydraulic model and also by computational flow dynamics (CFD). There were no not-allowed surface or subsurface (bottom or wall) vortices noticed, and also flow conditions in front of the impeller were acceptable. The basic advantage of this solution should be, that it requires lower submergence and shorter pumps – cheaper solution! But the research work indi-cated submerged vortices generated by turning vanes, which entered the impeller. Modifications were introduced and studied, but without success. After accepting the conclusion, that no practical solution could be expected on the side of pump sump, the problem was solved another way - on the impeller side. The level of stresses in the point of the highest concentration was reduced for ca 75 % with better impel-ler design and the position was shifted away from the impeller inlet - the sensitivity of impeller on dynamic loading was highly reduced! The basic contribution of the paper is the increase of knowledge of the solution with turning vanes in front of the vertical pump impeller.

Copyright: Pumps + Systems Association within VDMA e. V.


PROBLEM DESCRIPTION The design point of each of the three cooling water pumps is Q = 8,5 m3/s, H = 12,5 m. After the retrofit of the power plant (increase of electrical power output for more than 10 %) all three of them have to operate at the plant full load. The problem was, that all three of them had to be overhauled during each plant gen-eral overhaul. The reasons were cracks, which were noticed at the impeller leading edge near the hub (Fig. 1a). Sometimes a cooling water pump had to be stopped before the planned general plant overhaul due to the broken blade (Fig. 1b). One pump out of operation resulted in reduced cooling of the con-denser and consequently reduced plant production! Also air was noticed in the system, but pumps were not stopped because of this.

a) Crack b) Broken blade Figure 1: Impeller damages

Pumps are vertically suspended mixed flow pumps with open impellers. Typical for these particular pumps and pumps sump is that they are not executed and installed according to certain standard recom-mendations, such as given in literature (1), but they have turning vanes fixed on the flange of the suction bell. We tried to find useful references/recommendations about this solution, but our success was quite limited (2). Besides, what was very unusual, there were no model acceptance tests results available for pump sump testing In the year 2006 the site tests of pump characteristics were executed (flow measurement with current meters) and the results showed, that the pump operating point is in the vicinity of the best efficiency point, but the absolute value of efficiency was low (ca 78 %). Following were the goals for the pump refurbishment:

• Reduction of operating costs, increasing of pump reliability and system availability. • Avoidance of unplanned pumps cut-outs. • Reduction of the maintenance costs - increase of the time between two overhauls of each pump

from 1 cycle (18 months) to at least 3 cycles (54 months = 4,5 years). General accepted recom-mendations for the time between two overhauls for such type of pumps is minimal 5 years.

• Reduction of costs for the spare impellers and for reparation works. • Reduction of energy costs. • Increasing of pump absolute efficiency.

It was agreed, that pump inlet conditions should be analyzed by model pump sump as the first step of the investigation. The analysis is presented in the paper.



SUMP ANALYSIS At the beginning, we made an analysis of basic dimensions – comparison with recommendations from literature (1). It is true, that actual sump cannot be directly compared to the “classical” one, but we did this short analysis just in order to see, where the biggest differences can be noticed. Actual (m) Recomm.

(m) Recomm.

meet? Suction bell diameter D (m) 2,44 2,60 Yes Minimal submergence S (m) 2,02 4,6 No Distance of suction bell from the bottom C (m) 2,49 0,78 – 1,3 No Minimal water level above bottom H (m) 4,5 5,35 No Distance between pump and rear wall B (m) 2,55 1,95 No Minimal distance between travelling screen and pump Y (m) 6,0 10,4 No

Walls between pumps - No Yes No Maximal velocity in the inlet channel vx (m/s) 0,2 0,5 Yes

Table 1: Comparison of some actual dimensions with recommendations (1)

Basic conclusions from the analysis are:

• Minimal suction bell submergence is too low • Distance of suction bell from the bottom is too large • There are no walls between pumps and other details connected with this particular design (fillets,

cone), but there are ribs fixed on the suction bell Analysis didn’t omit possibility, that pump suction structures were the reason for the adverse phenomena and it confirmed the need for model study of pump sump. MODEL PUMP SUMP

Figure 2a: Model pump sump layout Model tests were executed on the basis of Froude similarity. The model was manufactured in the scale 1 : 16 as a closed loop system. Area of the model is indicated in the layout (Fig. 2a). The flow was delivered by auxiliary pump into two channels: the bed of the river Sava and the channel coming from the cooling tower pumps (CT pumps). Three cooling water circulating pumps (CW pumps) were modelled (Fig. 2b), equipped with swirl meters, adequate mesh was used for travelling screens (Fig. 2c). Test rig allowed single and parallel pump operation, flow rate control for each pump, water level control and walls of the pump sump were made from transparent material.



b) CW pumps models c) Travelling screen

Figure 2 b, c: Model pump sump arrangement

Following characteristics were determined: • surface vortices - visual observation • sub-surface vortices - visual observation by using injected dye tracers, • swirl angle – swirl meters in suction pipes of the pumps, • velocity profile at the position of impeller inlet - measured by Pitot tubes. EXISTING CONFIGURATION – BASIC TESTS All possible combinations of pumps in operation were tested (3 pumps in parallel and all combinations of two pumps in parallel) for most critical operating conditions: maximal flow rate and minimal water level. The results were following: Surface vortices Surface vortices were judged according to recommendations (1), where limiting conditions are temporary occurrence of vortices of type 3. At the minimal water level (149 m.a.s.l.) no vortices of type 3 can be noticed for all executed tests. The flow rate was increased for 50 %, but still no air entraining vortices were noticed (Fig. 3). With further decreasing of water level, very soon surface vortices of type 5 were noticed, water level 148,8 m.a.s.l. The conclusion is, that the defined minimal allowed submergence must be strongly re-spected and pumps are not allowed to run bellow it! Subsurface vortices, swirling flow and velocity profiles The complete region of the intake bottom as well as walls were checked with the dye injection. The result was, that no prohibited submerged vortices were noticed, initiating at the bottom or at the wall. Swirling flow and velocity profiles at the impeller inlet were in acceptable limits for all tested operating conditions. Conclusions Basic tests of existing configuration showed, that neither not allowed surface nor subsurface vortices could be noticed, flow vorticity was found in acceptable limits as well as velocity profile. It could be con-cluded that the pumps sump was designed correctly and air in the system was not sucked by pumps, if they operated at the water level above minimum allowed. The question still remained not answered: why the damaged impellers!

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Figure 3: Observation of surface vortices, 3 pumps in operation

Q = EXISTING CONFIGURATION – ADDITIO

Pump No.1

Figure 4: Dye injection – different pumps

The basic difference between actual pump sumpThis was the reason for the decision for more detailed research. vortices generated by turning vanes was investigated. The most dangerous vapour core vorticesnoticed. We tried to investigate if swirl type submerged vortices occurred, which pump operation. Dye was injected in systematically defined positions in all three pumps points on each vane, together 3 x7 x 54), different submergences (Fig. 5) and di In all cases the flow downstream the turning vanes was found not homogenous and with vortices geneated by turning vanes. It can be judged, that the pump position (No.1, 2 or 3) doesn’t have any influence (Fig. 4), but lower submergence results in stronger vortex (Fig. 5).them, the estimation if they were dangerous or not To get better impression about the situation the computational flow dynamics (CFD) was in

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Figure 3: Observation of surface vortices, 3 pumps in operation Q = 1,5 * Qmax, S = Smin (149 m.a.s.l.)

ADDITIONAL TESTS

Pump No.2 Pump No.3

different pumps - minimal submergence, 3 pumps running, turning vane No. 3, point No. 3

The basic difference between actual pump sump and “classical” one are turning vanes in front of impeller. This was the reason for the decision for more detailed research. The eventual occurrence of submerged vortices generated by turning vanes was investigated. The most dangerous vapour core vorticesnoticed. We tried to investigate if swirl type submerged vortices occurred, which are not allowed for safe pump operation. Dye was injected in systematically defined positions in all three pumps

, together 3 x7 x 5 = 105 positions. Different pumps in operation were studied) and different flow rates.

low downstream the turning vanes was found not homogenous and with vortices genecan be judged, that the pump position (No.1, 2 or 3) doesn’t have any influence

(Fig. 4), but lower submergence results in stronger vortex (Fig. 5). It was extremely difficultdangerous or not was practically not possible using only model tests

To get better impression about the situation the computational flow dynamics (CFD) was in

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Pump No.3

minimal submergence, 3 pumps running,

and “classical” one are turning vanes in front of impeller. he eventual occurrence of submerged

vortices generated by turning vanes was investigated. The most dangerous vapour core vortices were not not allowed for safe

pump operation. Dye was injected in systematically defined positions in all three pumps – 7 vanes, 5 were studied (Fig.

low downstream the turning vanes was found not homogenous and with vortices gener-can be judged, that the pump position (No.1, 2 or 3) doesn’t have any influence

extremely difficult to classify not possible using only model tests

To get better impression about the situation the computational flow dynamics (CFD) was introduced.



S = 149 m.a.s.l. S = 150 m.a.s.l. S = 153 m.a.s.l.

Figure 5: Dye injection – different submergencies - 2 pumps running,

pump No. 2, turning vane No. 3, point No. 3 CFD ANALYSISIS OF EXISTING SITUATION For the CFD simulations a commercial code ANSYS CFX 11.0 was used. CFX combines the finite vol-ume method with the finite element approach of the geometry representation. The finite volume method assures the conservation of physical properties. The CFD simulation was performed as a transient single-phase simulation with a flat water-surface as-sumption, using the shear-stress-transport (SST) turbulence model. Complete mesh (Fig. 6) was com-posed of three domains: inlet channel with the river bed, travelling screens and sump with pumps. Struc-tured and unstructured computational meshes were used, together 27 million elements and 8 million nodes. The flow in Fig. 8 was observed for the presence of vortical structures. For this purpose, the vortical struc-tures were identified with the Q-criterion (3), which is suitable for incompressible flows:

( )22

SΩ2

1−=Q , (1)

where Ω is the vorticity tensor, and S the strain-rate tensor. The flow analysis (Fig. 7) shows occurrence of vortical structures, generated by turning vanes. For each pump, two vortical structures with high values of Q (representing stronger vortices) seemed to reach the eye of the impeller, which is located above the pump suction bell area. During the transient simulations the Q of the vortical structures at the location of the eye of the impeller was generally above 5000 s-2, but may have changed to lower values from time to time. Such local vortices couldn’t be identified with the experiment and CFD was of great help. It was not pos-sible to judge, if such vortices are dangerous for the pump or not. Rough estimation showed that during 18 months of pump operation each impeller blade cuts the vortex 6 x 107 times – dynamic loading of the impeller blade!


Figure 6: Geometry for the CFD simulation and computational mesh of the turning vanes

Pump No.1

Figure 7: Existing geometry, CFD results (iso MODIFICATIONS The first step was testing of pumps as they were, but without turning vanes in front of sution bell, according to suggestion of purchaser. As it was expected the air entraining vortices of type 6 were noticed (Fig. 8). The main reason was, that the submergence was too low for this execution.



: Geometry for the CFD simulation and computational mesh of the turning vanes

Pump No, 2 Pump No. 3

: Existing geometry, CFD results (iso-surfaces of Q=500 s-2)

The first step was testing of pumps as they were, but without turning vanes in front of suc-tion bell, according to suggestion of purchaser.

d the air entraining vortices of . The main reason

was, that the submergence was too low for this

Figure 8: Modified pump, vortex of type 6


: Geometry for the CFD simulation and computational mesh of the turning vanes

: Modified pump, vortex of type 6



The walls were introduced between pumps in order to form a kind of separate pump bays, similar as proposed in recommendations (1), the situation was not improved. Looking at the turning vanes, they looked as a part of an elbow without inlet part of the pipe. Such or similar elbows (6) are used many times in piping systems, the vanes create acceptable flow downstream the elbow, referring to axial velocity component distribution as well as low radial component - no flow vortex. The modifica-tion was completion of turning elbow with inlet part (Fig. 9).

Horizontal pipe

Figure 9: Modified turning elbow – horizontal part of pipe added

The CFD analysis was performed, identical as with the original execution. The results are shown on Fig. 10. Pump No. 1 was equipped with modified suction part, while other two pumps remained unchanged.

Pump No.1 Pump No, 2 Pump No. 3

Figure 10: CFD results - modified geometry on pump No. 1 (iso-surfaces of Q=500 s-2)

After the modification the intensity and the length of the two vortical structures at pump No. 1 was lower than before. According to the CFD simulation, the conditions at the impeller inlet were improved. How-ever, the modification should be confirmed by model testing. The results were disappointing – the air en-training vortex was observed (Fig.11) and the modification failed, it was not acceptable!

Figure 11: Modified turning elbow – air entraining vortex of type 6


The modification, represented in Fig. SST (SAS SST) turbulence model to check the validity of the results. Namely, during the project our paallel research (5) has revealed that the SST model flows. The results, obtained by the SAS SST turbulence model, have revealed that instead of the two vortical structures, many vortical structures with high values of tions of the impeller eye (Fig. 12). The number of such vortical structures and their intensity was a bit lower for the modified pump, but the vortical structures were still present. Therefore, even after the modfication the dynamic loading may still be present at the impeller of the modified pump.

Pump No.1

Figure 12: CFD results (SAS SST turbulence model)

CONCLUSIONS The existing sump configuration with turning vanes insolution:

• Smaller required submergenceaxial length – cheaper and less problematic pump,single additional guide bearing)

• No unacceptable neither surface nflow vorticity at the impeller inlet and correct axial velocity distribution.

At the time, when they were introduced (70offered only limited possibilities. With an additional introduction of CFD analysis it could be said with high level of confidentiality, that turning vanes are gimpeller and cause dynamic loading on its blades. The modification of existing pump sump cannot be realized without extensive changes on its design, addtionaly also the pumps should be modified. It Another option was analysis of impeller and trying to find acceptable solution on its sensitivity on dynamic loading. New impeller was designed for the same operating point. The stress analysis analysis was performed for both geometries. For the original geometry (centration was found at the same point, as the cracks were initiated (has different characteristics (Fig. 13b). The position of the impeller blade inlet and its level isment of pump efficiency was 8 %, measured on model tests and confirmed at site



The modification, represented in Fig. 10, was simulated once again with the scale-adaptive simulation SST (SAS SST) turbulence model to check the validity of the results. Namely, during the project our pa

lel research (5) has revealed that the SST model may have some deficiencies in the case of the vortical flows. The results, obtained by the SAS SST turbulence model, have revealed that instead of the two vortical structures, many vortical structures with high values of Q (above 5000 s-2) are present at

). The number of such vortical structures and their intensity was a bit lower for the modified pump, but the vortical structures were still present. Therefore, even after the mod

ill be present at the impeller of the modified pump.

Pump No, 2 Pump No. 3

: CFD results (SAS SST turbulence model) - modified geometry on pump

(iso-surfaces of Q=500 s-2)

The existing sump configuration with turning vanes in-front of pump suction bell seems to be an ideal

submergence as classical pump sump (lower costs for civil works, shorter pump cheaper and less problematic pump, shaft composed from only two parts and only

single additional guide bearing) surface nor subsurface vortices were noticed during the tests and no

flow vorticity at the impeller inlet and correct axial velocity distribution. time, when they were introduced (70-ies), there was no CFD available. As we saw, model testing

offered only limited possibilities. With an additional introduction of CFD analysis it could be said with high level of confidentiality, that turning vanes are generators of subsurface vortices, which enter the pump impeller and cause dynamic loading on its blades.

The modification of existing pump sump cannot be realized without extensive changes on its design, addtionaly also the pumps should be modified. It was not an acceptable option.

Another option was analysis of impeller and trying to find acceptable solution on its sensitivity on dynamic loading. New impeller was designed for the same operating point. The stress analysis using finite element

was performed for both geometries. For the original geometry (Fig. 13a), the high stresses cocentration was found at the same point, as the cracks were initiated (Fig. 1). New geometry of impeller

b). The position of the highest stress concentration was removed from inlet and its level is only about 1/4 of the original impeller! Additionally, the improv

measured on model tests and confirmed at site!


adaptive simulation SST (SAS SST) turbulence model to check the validity of the results. Namely, during the project our par-

case of the vortical flows. The results, obtained by the SAS SST turbulence model, have revealed that instead of the two

) are present at the loca-). The number of such vortical structures and their intensity was a bit

lower for the modified pump, but the vortical structures were still present. Therefore, even after the modi-

modified geometry on pump No. 1

eems to be an ideal

as classical pump sump (lower costs for civil works, shorter pump shaft composed from only two parts and only

subsurface vortices were noticed during the tests and no

ies), there was no CFD available. As we saw, model testing offered only limited possibilities. With an additional introduction of CFD analysis it could be said with high

enerators of subsurface vortices, which enter the pump

The modification of existing pump sump cannot be realized without extensive changes on its design, addi-

Another option was analysis of impeller and trying to find acceptable solution on its sensitivity on dynamic using finite element

a), the high stresses con-1). New geometry of impeller

highest stress concentration was removed from Additionally, the improve-


a) Original impeller geometry

Figure 13: Stress analysis by finite element method (FEM) LITERATURE 1 American National Standard for Pump Intake Design ANSI/HI 9.8

2 Sulzer Pumps South Africa, SZA Sales Kit, 3 Vehar, F., Pavlin, R., Škerlavaj, A.

pump sump with hydraulic model, Turboinštitut’s report, Ljubljana, 2009 (in Slovene) 4 Hunt, J. C. R., Wray, A. A. and Moin, P.

Proceedings of the 1988 Summer Programpp. 193–208.

5 Škerlavaj, A., Škerget L., Ravnik J. and Lipej A.

Proc. IMechE Part A: J. Power and Energy 6 Vehar, F., Florjančič, D. and Buchegger S., Solution of the suction piping for the large fluid transport

pumps, Proceedings of the Internation



Original impeller geometry b) Improved impeller geometry

: Stress analysis by finite element method (FEM)

American National Standard for Pump Intake Design ANSI/HI 9.8-1998, New Jersey, 1998

Pumps South Africa, SZA Sales Kit, Pretoria, 2001

A. and Ermenc, D., Cooling water pumps CW102PMP, Analysis of pump sump with hydraulic model, Turboinštitut’s report, Ljubljana, 2009 (in Slovene)

. A. and Moin, P., Eddies, stream, and convergence zones in turbulent flows, Proceedings of the 1988 Summer Program, Stanford N.A.S.A. Centre for Turbulence Research, 1988,

kerget L., Ravnik J. and Lipej A., Choice of a Turbulence Model for Pump Intakes. Proc. IMechE Part A: J. Power and Energy, 2011, 225(6), 764–778.

Buchegger S., Solution of the suction piping for the large fluid transport Proceedings of the International Rotating Equipment Conference, Düsseldorf, 2008


b) Improved impeller geometry

1998

D., Cooling water pumps CW102PMP, Analysis of

Eddies, stream, and convergence zones in turbulent flows, , Stanford N.A.S.A. Centre for Turbulence Research, 1988,

Choice of a Turbulence Model for Pump Intakes.

Buchegger S., Solution of the suction piping for the large fluid transport , Düsseldorf, 2008

Analysis of turning blades in front of cooling water pump suction bell

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