design of suction bay for balanced erformance of a …

SASTECH 49 Volume 9, Issue 1, April 2010

DESIGN OF SUCTION BAY FOR BALANCED

PERFORMANCE OF A TWIN PUMP V. Santhosh Kumar 1, S. N. Sridhara2, C. S. Bhaskar Dixit3, Ramesh Madhiraju4

1- (Engg.) Student, 2- Professor and Head of Department (MME), 3- Professor and Centre Manager (RMD) Centre for Rotating Machinery Design, M. S. Ramaiah School of Advanced Studies, Bangalore

4- Pump Engineering Manager, Armstrong, Bangalore-560 003

Abstract

Twin pumps are two vertical inline pumps in one casing. The direction of rotation of the two pumps is identical and the suction bays for the right and left sides of the twin pump naturally induce prerotation/preswirl to the flow approaching the impeller. It is observed from tests that the prerotation in the left side increases head and power while that on the right side decreases impeller head and power. During regular operation it is observed that the two sides no longer give the same performance for which they were designed. This fact is also verified during performance tests.

In this study an attempt is made to eliminate prerotation on either sides of the twin suction bay and equalize impeller power and head. Firstly the baseline design (vertical inline pump) is numerically studied using full pump simulation and the results are compared to that of experimental data. Evaluation of the traditional suction design is done using component simulation. Parametric design changes are carried out on the traditional suction bay to eliminate prerotation. Two selected redesigned versions are further analysed using full pump simulations and the results are compared with that of the baseline and the traditional designs.

From CFD simulation results for the baseline case its obeserved that the head and power near the design point is within 5 % of that of the experiments with good trend wise agreement. The redesigns(over 7 configurations) undertaken shows that pump head and power on either side can be brought to within 1 % of each other but the design would be difficult to cast. On foundry friendly designs its observed from full pump simulations that the head and power in the left and right sides can be brought to within 3% of each other.

Key Words: Twin Pump, CFD Simulation, Parametric Design, Redesign, Head and Power

Nomenclature

V Absolute velocity ,m/s g Acceleration due to gravity, m/s2 k Turbulent kinetic energy, m2/s2 r Radius, m y+ Non dimensional distance from the wall ε Turbulent kinetic energy dissipation rate, m2/s3 ω Angular velocity, rad/s

Subscripts

u Circumferential direction 1 Impeller inlet location 2 Impeller exit location

Abbreviations

BEP Best Efficiency Point CAD Computer Aided Design CFD Computational Fluid Dynamics HVAC Heating, Venting and Air Conditioning

1. INTRODUCTION

Centrifugal pumps are the preferred choice in the HVAC industry due to their ability to operate at wide ranges of flows and head. The vertical in-line pump is a single suction centrifugal pump used in the HVAC industry and is so named because the pump can be inserted directly in a pipeline and the motor is mounted vertically. The twin vertical in-line pump is two vertical in-line pumps combined into one casing i.e. one suction flange and one discharge flange. These pumps can be used for serial operation wherein one is a duty pump and the other is a standby or for parallel pumping where both pumps are running to meet the required duty flow. The impellers used in either side of the twin pump are

identical and also has the same direction of rotation. In the twin pump, as water enters the suction flange it is diverted in the direction of the duty impeller by means of a suction bay. The suction bays for the right and left sides of the twin pump are highly three-dimensional and turn towards the left and right sides from the suction. This naturally induces strong effects of rotation and curvature and the flow approaching the impeller eye contains prerotation/ preswirl.

Figure 1 shows a typical hydraulic design for a suction bay of a twin pump. When the right pump is in operation, the flow from the suction flange turns towards the right and reaches impeller. Due to the inherent shape of the suction bay, prerotation is created in the clockwise direction which is in the same direction as that of impeller rotation. If the left pump is operational then the flow creates prerotation in the counter-clockwise direction which is in the direction opposite to that of the impeller rotation.

Most often at the impeller inlet for pumps the fluid flow is assumed to be axial and 01 uV . The theoretical

head developed by the impeller Hth is obtained from the well-known Euler equation,

g

VrVrH 1u12u2

th

….……..……(1.1)

With prerotation in the direction of impeller rotation on the right side of the twin pump the impeller does not get to impart the necessary impetus to the oncoming flow and the 1uV component is positive. Hence the impeller head and power required decrease. While on the left side the prerotation is in the direction opposite of impeller rotation wherein the 1uV component


is negative which increases impeller head and power. This increase in head and power in the left pump and decrease in head and power in the right pump create a situation during performance tests, certification approvals and during regular operation wherein the two sides no longer give the same performance for which they were designed.

Fig. 1 Basic hydraulic design of a traditional twin pump

Very few researchers have focused on prerotation in centrifugal pumps and have employed CFD and experimental methods to eliminate or reduce prerotation. Stepanoff [1] was one the earliest to measure and document prerotation and its effects. He used a rotameter to study the swirl ahead of the impeller. Karrasik [2] from his studies suggest that negating excess swirl in suction intakes may be by use of straightening vanes ahead of the impeller and rearranging the suction piping to avoid changes in direction. Gulich [3] from his long experience with Sulzer pumps highlights that prerotation always grows with the specific speed, with the circumferential velocity and with the flow rate. The author recommends use of bellmouths and inlet rings to reduce the circumferential component of the approach flow and ensures a favorable flow distribution upstream of the impeller. Yunbae Kim et al. [4] from Dresser Rand company designed radial inlets for centrifugal compressors and the resulting flow structure ahead of the impeller was studied using CFD. The basic design of the radial inlet included a combination of split vanes, scoop vanes and inlet guide vanes to guide the flow towards the impeller. Eugen Constantin et al. [5] analyzed swirling flows in the intake channel of suction sump pump. Eddy flow prevention devices such as suction cones, cruciform guides or vertical splitters were used to prevent vortexes and swirling flows. Jay M. Koch et al. [6] used experimental and computational means to study the non-uniform inlet profile generated by a radial inlet of a radial compressor and found that the addition of the elbow has the most detrimental effect. Jennifer Anne Roberge [7] used both flow visualization techniques and CFD to visualize and resolve the flow in the vicinity of a pump intake. Both the techniques confirmed that that the swirl number increases with increase in flow strength. Studies conducted by Daniel Marjavaara [8] indicate that the geometrical shape and velocity distribution at the inlet are the two main factors that affects the performance of the hydraulic turbine draft tube which is quite similar to that of pump intakes. Van den Braembussche [9]

highlights the point that inlet distortion causes incorrect incidence and circumferentially varying non-optimum flow patterns in the different impeller channels. For the radial inlet volute, the uniformity of the flow is improved when the eye is equipped with a bellmouth. Tangential inlet ducts were suggested for more uniform inlet velocity profile in axial and circumferential direction.

Available literature reveals that attempts have been made to reduce swirls in the suction piping intakes of pumps. But not much of information is available on the quality of flow that goes through to the impeller and the losses incurred. Most authors have used experimental and CFD techniques to study and visualise swirls. Computational method is the first choice when it comes to design changes or redesigns. The present study will focus on swirl reduction/elimination in the suction bays of twin pumps such that both the left and right side pump head and power are equalized and would also ensure that the quality of the flow that enters the impeller is good. Two types of CFD simulation would be carried out, component simulations and full pump simulations. Component simulations involve analysis of a single component such as the suction bay alone. In case of full pump simulations, the suction, impeller and volute would be analysed. The component simulations offers advantages of faster analysis time which allows for number of designs to be verified in a short time frame. This has been proven to be very effective in new design and redesign scenarios [4, 6].

2. BASELINE DESIGN

The baseline pump in this study is the vertical inline pump from which the twin pump is constructed. The impeller and volute from the baseline pump will be used for the twin pump. The solid model for the baseline pump is generated using commercially available CAD package Solidworks and a cross sectional view of the same is shown in figure 2.

Fig. 2 3-D view of the baseline pump assembly

The specification for the baseline pump is shown in table 1. The baseline vertical inline pump has been tested to Hydraulic Institute (HI) standards [10] and the tested numbers will be compared with the simulation results for the initial baseline benchmark. A schematic sketch of the test rig and the measuring equipment used are as shown in figure 3.

3. TRADITIONAL TWIN DESIGN

The basic hydraulic constraint for the twin pump is that it has to utilize the volute and impeller from that of


the baseline version. Other design specifications are listed below.

Flange to flange length: 800 mm

Pump centre to centre distance: 700 mm

Table 1. Specification of the baseline pump

Sl. No

Parameter Value

1 Casing Vertical in-line, Double volute

2 Impeller Single suction, Francis vane type

3 Number of vanes 7 equally spaced 4 Specific speed 2462 (US);

1511(Metric) 5 Suction specific

speed 8340 (US); 5120 (Metric)

6 Impeller exit diameter 254 mm 7 Impeller eye diameter 200 mm 8 Speed 1750 RPM 9 Flange to flange

length 500mm

Fig. 3 Experimental test setup for the baseline pump

For the traditional twin suction design, existing (traditional) twin pump design rules are used to model the suction bay. The impeller and volute are the same as used on that of the baseline design. An additional valve chamber is added to the volute which houses the flapper which controls the serial and parallel operation of the pump. Only one side of the twin pump can be analyzed at a time hence it is assumed that while one pump is running the other is non-running with the flapper valve fully closed.

4. PARAMETRIC DESIGN CHANGES

Based on literature survey, the traditional suction bay as shown in figure 1 is redesigned systematically to eliminate the swirl in the suction bay. The left and right sides of the suction bay are mirror images and hence it is sufficient to model and simulate only one side.

The initial redesigns (Redesign-1 & 2) included splitters in the middle of the suction bay followed by addition of inlet guide vanes ahead of the impeller eye. The splitter is intended to split the flow from the flange into two before reaching the impeller eye, thus breaking the swirl before it reaches the eye. The splitter thickness was maintained at 6mm. The guide vanes are meant to condition the flow ahead of the impeller eye and

provide a swirl free flow to the impeller. The guide vanes are placed 10mm ahead of the impeller eye. Redesign-3 included 17 scoop vanes ahead of the impeller eye as highlighted in reference [4]. The scoop vanes are located 360deg circumferentially in order to provide uniform flow to all regions of the impeller eye as shown in figure 4a. Apart from the scoop vanes the design also includes a conical cut below the impeller eye. The 17 thin scoop vanes are difficult to manufacture using sand casting which is the traditional manufacturing process adopted for pumps. Hence a simplified alternative is further investigated. Redesigns 4, 5 and 6 are further improvements to redesign-3 by reducing the number of scoop vanes and varying its location. Redesign-6b includes a splitter through the middle of the suction bay followed by two scoop vanes and one splitter below the impeller eye. Lesser number of scoop vanes makes the design foundry friendly and would also reduce the total pressure loss through the suction bay.

Fig. 4 Geometry of Redesign-3 and Redesign-6b

In all 7 designs are generated and all of the designs are meshed and analyzed using component simulations. The best two component designs based on plots and global numbers are selected for further investigations using full pump simulations.

5. MESH GENERATION

The meshes are generated using commercially available mesh generation software ANSYS ICEM-CFD. The mesh consists of tetrahedral elements with prism elements in the boundary layer. A target y+ i.e. the non-dimensional distance from the wall of 25 is selected for the meshes and the initial prism height is calculated. The boundary layer contains 10 mesh points.

For the purpose of full pump simulations the whole pump is broken down into three domains (suction impeller and volute) and are interconnected via interfaces. Three meshes are generated for the purpose of mesh sensitivity studies. The meshes labelled as Mesh1, Mesh2 and Mesh3 consist of 996527, 1523296, 2246718 nodes respectively. The mesh refinement ratio between meshes is maintained at around 1.5. The overall mesh quality for each domain in ICEM-CFD is maintained above 0.3. The higher mesh densities are maintained at the impeller inlet edges and trailing edges and also at the cutwater region and splitter regions of the volute. This ensures that the small fillets and rounds observed in these regions are captured accurately.

The meshes for the component simulations are generated based on the same guidelines as those are the full pump simulation. The mesh densities are maintained close to 0.5 million. Areas with small radii of curvatures have been resolved by increasing the local


mesh density. Care has been taken to ensure that the overall quality for the meshes is greater than 0.4.

6. NUMERICAL SIMULATION

The simulations are of steady state type. The fluid used in simulation is water at 25o C. For the full pump simulation the following boundary conditions are applied and are shown in Figure 5.

Pump Inlet: The total pressure (Pa) inlet condition is more appropriate than the uniform velocity or mass flow inlet condition for cases that assume that the machine is drawing fluid directly from a static reservoir.

Pump Outlet : Mass flow rate (kg/s)

Impeller blade, Hub, Shroud: Rotating walls

Volute and Suction walls: Stationary walls

Fig. 5 Boundary conditions for full pump simulations

The walls, both rotating and stationary, are specified as no-slip walls. The surface roughness (sand grain roughness) for the impeller was set at 0.00635mm and for the suction-volute walls is set at 0.0127mm. The domains are connected through interfaces and a frozen rotor approach is used to model the physics between the two stationary and rotating frames of reference. The two frames of reference connect in such a way that they each have a fixed relative position throughout the calculation.

For the component simulation the following boundary conditions are applied and are shown in Figure 6.

Pump Inlet : Total pressure (Pa)

Pump Outlet : Mass flow rate (kg/s)

Suction walls: Stationary walls

Commercially available solver ANSYS CFX 11 is used for the simulations. The standard k turbulence model, which is one of the most validated turbulence models for turbomachinery applications has been used in this study [11]. The convergences criteria is set to 1e-5 for the root mean square (RMS) residuals of mass, momentum and pressure. The simulations are run on two 4GB RAM machines with Intel Xeon processors. The average run time for the full pump simulation is around 9 hours for a single flow point and the same for a component simulation is 2 hours.

Fig. 6 Boundary conditions for component simulations

7. RESULTS AND DISCUSSION

For the mesh sensitivity study, the baseline pump is analyzed at three flow points of 50%, 100% and 150% of BEP. Based on the mesh sensitivity study Richardson’s extrapolation is used to derive the fine mesh solution which is a method for obtaining a higher order estimate of the continuum value (value at zero mesh spacing) from a series of lower order discrete values [12]. The grid convergence index (GCI) recommended by Roache [13] to be included as a standard step in mesh sensitivity studies is also calculated. The GCI provides a consistent manner in reporting of mesh convergence studies and provides an error band on the mesh convergence of the solution.

Fig. 7 Fine mesh CFD solution for pump head and comparison with baseline test results

Figure 7 shows the comparison of the experimental results for the baseline case with those for the fine mesh CFD solution. The error bar on the pump head parameter from the simulation results are from the GCI which shows an error band of ±1.96%, ±0.93% and ±1.76% at 50%, 100% and 150% of BEP respectively. For the impeller torque the error bands are much smaller compared to that of the pump head. The calculated error band for torque is ±0.83%, ±0.43% and ±0.54% at 50%, 100% and 150% of BEP respectively. It is observed that at BEP the error bands are smaller compared to 50% and 150% of BEP flow. The largest error band is noted at 50% of BEP which is most probably due to the large recirculation’s, separations and impeller back flow at that flow. Based on the results of the mesh sensitivity study, it is decided that the finest of the three meshes i.e. ‘Mesh3’ would be used for all further simulations to generate the full pump characteristic curves.


Using ‘Mesh3’ full pump simulations for the baseline design are carried out. Figure 8 show the flow through the baseline pump from inlet to exit at BEP. Both streamlines and velocity vectors shows good uniform flow through the suction bay and impeller. In the volute a recirculation zone is observed near the discharge flange. This is mostly due to an aggressive diffuser from the throat to the discharge flange. The flow around the double volute is uniform with no sign of flow separation. The important fact to note is that the flow through the impeller and spiral volute is smooth and the same components would be used on the twin version as well.

Fig. 8 Streamlines and velocity vectors for the baseline pump

The baseline simulations are performed at nine flow points between 50 % to 150 % of BEP. At 50 % of BEP, there is strong visible suction back flow from the impeller and recirculation in the impeller blade passages. This is expected at part loads as the pump is not designed to run at these conditions. Achieving a converged solution at part load is difficult due to the suction back flow and large scale recirculations. This reduces the numerical accuracy of the simulation as there is always a trade off between accuracy and solution robustness at part loads. At BEP the streamlines through the suction and impeller are good with no visible suction back flow.

The predicted head curve (figure 9) through simulations agrees well in trend with that from the experiment data. Variation between CFD and experiments in head at BEP and at end of curve is around 5 %. The head curve drops slights towards 50% of BEP and this is due to the strong suction back flow and flow recirculation at part load as explained earlier.

The input power curve (figure 10) generated through simulations shows very good agreement with that of the experimental data. The trend is good and the difference between CFD and experiments is about 5% at BEP and about 7 % at 150 % of BEP. The power curve from the experimental results show that it’s a non overloading power curve with a certain power maximum close to 140 % of BEP while the CFD curve continues to rise at the same flow.

From the current simulation setup it’s only possible to determine the hydraulic efficiency. The volumetric and mechanical efficiencies are predicted through well validated empirical correlations. The peak overall efficiency calculated is within 1 % of that of the

experiments. At 50 % of BEP the predicted overall efficiency is lower than experiments; this is due to the lower predicted head in the simulations.

Fig. 9 Comparison of head curves from full pump simulation between the baseline CFD and

experiments at BEP

Fig. 10 Comparison of input power curves from full pump simulation between the baseline CFD and

experiments at BEP

The process of comparison between the full pump simulations with that of components is a very important step in any component redesign using CFD but it’s usually neglected. It’s essential to ensure that both visual plots and calculated numbers between simulations agree well in order to avoid surprises after the redesign stage. The streamlines for the suction bay of the baseline version shows similar flow patterns obtained from both full pump and component simulation. Figure 11a and b shows the comparison of velocity vectors at the impeller eye from full pump and component simulations respectively. This represents the flow going into the impeller. Although there are some minor differences in the location of the peak velocity vectors in both cases, the overall flow field is captured well. Two small vortices are clearly visible in both cases. Figure 11c and d shows the total pressure contours at the impeller eye from full pump and component simulations, this again shows good agreement between both forms of simulations. This confirms that component simulations can be used for redesign kind of scenarios.


Fig. 11 Velocity vectors and total pressure plot comparison at the impeller eye between the full

pump and component simulations

Fig. 12 Streamlines and velocity vectors at the impeller eye for the traditional twin suction bay

The traditional twin suction bay is analyzed using component simulation in exactly the same manner as that of the baseline pump. Due to geometric similarity of the traditional twin suction bay, simulation of only one side is sufficient and for this study the right side is simulated. Figure 12 shows the streamlines and velocity vectors at the impeller eye from the traditional twin suction bay. As observed in the plot the flow from the flange to the impeller eye follows the shape of the traditional twin suction bay which involves a 3D turn. This naturally induced turn causes the flow to create a swirl (clockwise direction) at the impeller eye which in turn reduces the head and power of impeller. Velocity vectors at the impeller eye shows the swirl and also shows non-uniformity of the velocity vectors. Large areas of peak velocity are observed and the axial component of the velocity is not uniform. Circumferential velocity confirms the strong swirl at the impeller eye. Overall the non-uniform velocity vector with a large swirl at the impeller eye is not ideal at the entrance of the impeller. The parametric design changes would focus swirl reduction and on achieving a uniform velocity vectors at the impeller eye.

The parametric design changes are performed using Solidworks and are analyzed using component simulations. Figures 13(a)to (g) shows the velocity

vectors as seen from the top and side at the impeller eye from the component simulations for the various redesigns. The vectors as seen from the top would aide in spotting swirls and vortices, while the vectors from the side help in judging the uniformity in the axial component of velocity and spot backflows. As seen from the traditional twin suction plots, the velocity vectors at the eye have a big swirl and the axial component of velocity is not uniform. Based on calculated numbers the swirl is least in the baseline design while that of the traditional design has the highest swirl. Figure 13a shows velocity vectors at the impeller eye for redesign-1. As seen from the image the strong swirl in the clockwise direction still persists. With this design change, the calculated swirl number does not improve and the losses are higher compared to the traditional design by over 5 %. Redesign-2 which has an inlet guide vanes before the impeller eye shows slightly better axial velocity component as shown in fig.15b. But at the same time there is some visible back flow. With the use of the inlet guide vane the swirl number at the eye reduces by over 51 % compared to the traditional design indicating that the vanes are effective. On the other hand losses are slightly higher than that of redesign-1. Redesign-3 shows velocity vectors without any swirl component although small vortices can be spotted in Figure 13c. The axial velocity component is uniform and the peak velocity at the impeller eye is less compared to the traditional suction design or other redesigns. The calculated swirl number confirms the reduction in swirl at the eye and the number is almost as low as that of the baseline design. On the contrary the losses are 56 % higher than traditional design. These increased losses are primarily due to friction and incidence losses at the scoop vanes. Redesign-4, 5 and 6 are improvements over the Redesign-3 to make the design more casting or foundry friendly. But these changes do not improve the flow field at the impeller eye. With these changes neither the swirl number nor the losses increase by a big margin. From Figure 13(d) to 13(f) it is noticeable that all three designs have some amount of suction back flow. Figure 13g shows the velocity vectors for the Redesign-6b. The plots do not show any big swirl at the impeller eye and the axial component is uniform with no suction back flow. It is also noticeable that the peak velocities at the eye for this design is on the higher side compared to Redesign-3. The swirl number for this version has decreased by 43% over the traditional design. Redesign-3 and 6b are effective is reducing the swirl at the impeller eye. But both design show increased losses over the baseline design. The increased losses translate to 1.7 to 1.93% of the baseline pump head for the Redesign-6b and Redesign-3 respectively. These mentioned numbers are fairly small (within 2 %) and the overall performance would not deteriorate a big margin. Hence based on plots and numbers from component simulations it is decided to proceed with Redesign-3 and 6b for the full pump simulations for further design verification.


Fig. 13 Velocity vector plots at the impeller eye for all various versions of the redesign

The performance curves are generated from the full pump simulations for the traditional, Redesign-3 and 6b versions and are plotted along with the baseline design.

Figure 14 shows the full pump head comparison curves for all the simulations. Towards 50 % of BEP all of the head curves bunch up together not showing much difference between the various designs. Closer to 100 % of BEP a noticeable difference is seen between the traditional twin design and its associated redesigns. For the traditional design at BEP it is visible that the right side head curve is lower than the left side by 6.2 %. Further towards 150 % of BEP the difference increases to 20 %. This difference is attributed to prerotation and with increasing flow the intensity of prerotation increase thereby decreasing the head on the right side. For the Redesign-3 and 6b versions it is observed that the head curves are close to each other. At BEP and at 150% of BEP the difference between the left and right sides of Redesign-3 is within 1%. For the Redesign-6b the difference at BEP is 2.6% and 7% at 150% of BEP. From the head curve comparison for the traditional and the redesigned versions it’s evident that with redesigning the suction bay the difference at BEP can be brought down to within 1% using Redesign-3. But Redesign-3 in itself may not qualify as a final design because of its manufacturing constraints. But still using Redesign-6b would bring down the difference in head between the left and right sides to within 2.6% which is much better than the traditional design. Comparison of the baseline head curve with any of the twin versions shows that there is a difference of 5% to 9% between them around BEP. This difference is due to the losses in the suction bay and the discharge valve chamber. Computed numbers for the valve chamber reveal that the averaged loss (for all designs) in the twin valve is close to 5% of the baseline pump head at BEP.

Fig. 14 Comparison of head curves from full pump simulation for the baseline and the redesigned

versions

The input power curve comparison shown in Figure 15 shows similar trend for all designs. For the traditional design there is a difference of 4.9% at BEP and 10% at 150% of BEP between the left and right sides. This difference is consistent with theory that the right side would have lower power compared to the left side and the magnitude would depend on the intensity of flow. The redesigned power curves show much closer numbers for both the left and right sides. For Redesign-3 the maximum difference between the left and right sides is 1.37% at 150% of flow. While for the Redesign-6b the difference in power curves between the left and right sides is 1.2% and 3.4% at BEP and 150% of BEP.

The overall efficiency is computed for all versions. The twin version is not as efficient as the baseline design in terms of peak efficiency. This is an accepted fact that the valves in the discharge and the suction bay contribute to higher losses than the baseline design which has been confirmed by the simulations. The hydraulic efficiency being a function of head and power would vary for the right and left sides of any twin version. The mechanical and volumetric efficiencies would be the same for all versions as there is no change in the mechanical parts or the clearances. There is a


drop of about three points in overall efficiency between the baseline design and Redesign-3. The Redesign-3 version shows about 2% higher efficiency than the traditional design.

Fig. 15 Comparison of input power curves from full pump simulation for the baseline and the redesigned

versions

8. CONCLUSION

In this study an attempt is made to eliminate/reduce prerotation on either sides a twin pump to equalize impeller power and head. Redesign of the traditional twin suction bay is performed and the performance is validated using CFD. The main conclusions drawn from this study are as follows

The baseline CFD simulation results for head and power around BEP agreed to within 5% of that of the experimental results. The computed overall efficiency for the baseline case agreed to within 1% of the experimental results.

From simulations it can be visualized and concluded that variation of head and power between the left and right sides of a twin pump is a result of naturally induced prerotation.

Parametric design studies were effective in reducing the swirl in the twin suction bay by means of using guide vanes and baffles/splitters. But these methods induced losses upto 2% of the baseline pump head.

The redesigns undertaken shows that pump head and power on either side can be brought to within 1% of each other but the design would be difficult to manufacture using sand casting.

On foundry friendly designs its observed from full pump simulations that the head and power in the left and right sides can be brought to within 3% of each other.

9. REFERENCES

[1] Stepanoff A.J,(2001), Centrifugal and Axial Flow Pumps, Design and Application, 2nd Edition, Krieger Publication, Malabar Florida.

[2] Igor J. Karassik, Joseph P.Messina, Paul Cooper, Charles C.Heald, (2001), Pump Hand Book, 3rd Edition, McGrawHill Publication.

[3] Johann Friedrich Gulich, (2007), Centrifugal Pumps, Springer-Verlag Berlin.

[4] Yunbae Kim, Jay Koch,(2004), Design and Numerical Investigation of Advanced Radial Inlet for a Centrifugal Compressor Stage, IMECE 2004-60538, Dresser-Rand Company.

[5] Eugen Constantin Isbasoiu, Tiberiu Muntean, Carmen Anca Safta, Petrisor Stanescu, (2005), Swirling Flows in the Suction Sumps of Vertical Pumps - Theoretical Approach, Department of Hydraulic Machinery “Politehnica” University of Bucharest, Workshop on Vortex Dominated Flows – Achievements and Open Problems Timisoara, Romania.

[6] Jay M. Koch, Peter N. Chow, Brad R. Hutchinson, Steve R. Elias,(2004), Experimental and Computational Study of a Radial Compressor Inlet, TP103prt, Dresser-Rand.

[7] Jennifer Anne Roberge, (1999), Use of Computational Fluid Dynamics (CFD) to Model Flow at Pump Intakes, MSc Thesis, Environmental Engineering Department, Worcester Polytechnic Institute.

[8] Daniel Marjavaara B.,(2006), CFD Driven Optimization of Hydraulic Turbine Draft tubes using Surrogate Models”, Ph. D Thesis, Lulea University of Technology, Dept. of Applied Physics and Mechanical Engg, Division of Fluid Mechanics, 2006:41, ISSN:1402-1544, ISRN:LTU-DT-06/41-SE.

[9] Van Den Braembussche. R.A, Flow and Loss Mechanisms in Volutes of Centrifugal Pumps, RTO-EN-AVT-143, Von Kàrmàn Institute For Fluid Dynamics, Belgium.

[10] Hydraulic Institute, (1994), Centrifugal Pump Test Standard (ANSI/HI 1.6), ISBN:1-880952-02-5.

[11] ANSYS CFX, (2009), CFX Help files V11.

[12] Ismail B.Celik, Urmila Ghia, Patrick J.Roache, Christopher J.Freitas, (2008), Procedure for Estimation nd Reporting of Discretization Error in CFD Applications, J. of Fluids Engineering”, 078001-4/Vol.130.

[13] Roache, P. J., Ghia, K. and White, F., (1986), Editorial Policy Statement on the Control of Numerical Accuracy, ASME J. of Fluids Engineering, 108(1), pp 2.

design of suction bay for balanced erformance of a …

Documents