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The influence of prewhirl regulation by inlet guide vanes (IGVs) on a centrifugal pump performance is investigated experimentally and numerically. The experimental results show that IGVs can obviously change the head and increase the efficiency of the tested centrifugal pump over a wide range of flow rates. Although the cavitation performance is degraded, the variation of the cavitation critical point is less than 0.5 m. Movement of the computed three-dimensional streamlines in suction pipe and impeller are analyzed in order to reveal the mechanism how the IGVs realize the prewhirl regulation. The calculated results show that the influence of IGVs on the cavitation performance of centrifugal pump is limited by a maximum total pressure drop of 1777 Pa, about 7.6% of the total pressure at the suction pipe inlet for a prewhirl angle of 24°.

In industrial centrifugal compressors, the prewhirl regulation by inlet guide vanes (IGVs) is a widely used approach to regulate the pressure ratio and the mass flow at constant rotational speed. With the advantage of convenient installation and practical effect, IGVs used in centrifugal compressors have been sufficiently investigated for several decades. The literature on IGV investigations for centrifugal compressors mainly falls into the following categories: effects of IGVs on compressors' energy performances [

However, compared with centrifugal compressors, the usage of the prewhirl regulation of IGVs is extremely rare for centrifugal pumps. The main reason is the potential deterioration of the cavitation performance. With installing IGVs at the centrifugal pump inlet, an inevitable pressure loss is induced due to the friction loss on IGVs' surface and the incidence loss on IGVs' leading edge. The pressure loss at the pump inlet is one of the most important factors for the cavitation performance of centrifugal pumps.

Cavitation usually leads to pressure fluctuation and uneven load distribution, which seriously reduces pumps' efficiency and affects their stable operation ranges. Due to the importance of the cavitation phenomenon, many research works have been performed toward understanding the cavitation in centrifugal pumps experimentally and numerically [

Confined to detrimental effects of the prewhirl regulation on the cavitation, there are few investigations on effects of prewhirl regulation by IGVs in centrifugal pumps. Tan

This paper investigates the influence of prewhirl regulation by IGVs on centrifugal pump cavitation performance both experimentally and numerically as a further study of the previous work by Tan

The IGVs have been specifically designed for the centrifugal pump impeller (whose original design did not include them) under the design constraint of enacting the non-impact condition at the impeller inlet at its nominal operating conditions (see reference [

The numerical model and method for simulating the flow filed of single phase (non-cavitation flow) in the centrifugal pump can be found in previous work by Tan _{m}_{l}_{l}_{v}_{v}_{m}_{l}_{l}_{v}_{v}_{t}

The re-normalisation group (RNG) _{m}_{t}_{μ} = 0.09.

The density function _{m}

The liquid-vapor mass transfers due to cavitation are solved by the transport equation—based cavitation model proposed by Zwart

In this model, a transport equation with source terms based on the homogeneous flow theory is used to solve the mass transfer between liquid and vapor phases. The fundamental assumption of this model is that the cavitation bubble does not interact with each other and the nucleation site density remains the same. However, with the development of cavitation the liquid volume fraction inevitably decreases, the nucleation site density should be modified accordingly. To further improve the modeling of the vaporization and condensation processes, the mass transfers for vaporization rate _{vap} and condensation rate _{con} are modeled as follows:
_{vap} and _{con} are the empirical calibration coefficients for vaporization and condensation rates, respectively; _{b} is the bubble radius; _{v}_{nuc} is the nucleation site volume fraction.

The vapor pressure _{v}_{v}_{sat} is the saturation pressure at the test temperature.

In the present calculation work, the commercial CFD code CFX 13.0 is employed. The equations solved in the calculation are the Navier-Stokes equations coupled with the RNG

In order to make the pump inlet pressure consistent with the experimental operation, the total pressure at the pump inlet is specified and the velocity direction is taken to be normal to the boundary. The mass flow at the pump outlet is selected. The scalable wall functions are imposed to solve the near-wall flow close to the no-slip wall over the impeller blades and sidewalls, the volute casing and the inlet and outlet pipe walls. The pressure at pump inlet decreases step by step until a convergence at each given operating condition in the cavitation flow calculations.

Five different mesh densities are used to calculate the pump performances at design point without cavitation, as shown in

Here, _{1} and η_{1} are the pump head and efficiency calculated by using the Mesh 1, respectively.

^{3}/h, with the maximum relative errors smaller than 5.0%, which demonstrates that numerical calculations simulate the pump performance accurately.

_{th} is the theoretical head of pump; ω is the rotational angular velocity of impeller; g is the gravitational acceleration; _{u}

For the flow rate of 150–375 m^{3}/h, the efficiencies of centrifugal pump under prewhirl regulation are all higher than that without IGVs, as shown in

^{3}/h. The three-dimensional streamlines in the pump suction pipe and impeller start from the suction pipe inlet by sampling forty equally spaced points on this surface.

Without IGVs, the streamlines are straight in the suction pipe and then gradually rotate in the impeller with the same direction as the rotational impeller. Due to the asymmetrical three-dimensional geometry of IGVs, the fluid regulated by the IGVs has positive prewhirl at an IGV angle of 0°. For the positive prewhirl regulation, the fluid has an obvious positive circulation in the suction pipe while passing the IGVs. The rotation strength increases as the prewhirl angle increases from 12° to 24°, as shown in

^{3}/h, respectively. Net positive suction head available (

In ^{3}/h are also given. The comparison between the measured and calculated results shows that the numerical simulation can accurately simulate the drop trend of pump head with the decrease of

^{3}/h. Due to the interaction of the rotational impeller and stable volute, the pressure distribution in the impeller is not symmetrical, which leads to the inhomogeneous cavity in the impeller. The total pressure contour at the impeller inlet becomes uneven when the pump has IGVs installed as shown in

Undoubtedly, the total pressure losses in suction pipe with IGVs are larger than that without IGVs. For the tested centrifugal pump in this paper, the total pressure losses under positive prewhirl regulation are larger than that under negative prewhirl regulation with IGVs. The prewhirl strength of fluid obviously increases as the angle increases when the IGVs are regulated to positive angle. This is also apparent from the elongated streamlines as shown in

The effects of prewhirl regulation by IGVs on head, efficiency and cavitation performances of a centrifugal pump are investigated by experimental measurements and numerical simulation. The experimental results show that the efficiency of pump with IGVs is higher than that without IGVs, the highest efficiency improvement is more than 2.0% around the designed conditions. The head of centrifugal pump under positive or negative prewhirl regulation is decreased or increased, respectively, due to the variation of circumferential velocity at the impeller inlet. The centrifugal pump cavitation performance is degraded when it installs the IGVs, but the variation of

The numerical method is validated by comparing the experimental and calculated results for both non-cavitation and cavitation flows in the centrifugal pump. The positive or negative rotational streamlines in the suction pipe and impeller clearly show how the IGVs regulate the flow pattern in the centrifugal pump. The calculated results reveal the cavity distribution in the impeller and the total pressure contour at the impeller inlet. Moreover, the quantitative analyses on the total pressure loss between the suction pipe inlet and impeller inlet are realized for the centrifugal pump without and with IGVs. Numerical results show that the influence of IGVs on the cavitation performance of centrifugal pump is limited with the maximum total pressure drop of 1777 Pa (about 7.6% of the total pressure at the suction pipe inlet) for the prewhirl angle of 24°.

For the centrifugal pump tested in this paper, in the range of prewhirl angle −24° to 24°, the prewhirl regulation of IGVs can greatly increase or decrease the centrifugal pump head and improve the efficiency in a wide operation range, and induce limited negative influence on the cavitation performance.

This work was supported by the National Natural Science Foundation of China (Grant number 51176088 and 51179090) and the Open Research Fund Program of State Key Laboratory of Hydroscience and Engineering (Grant number sklhse-2012-E-02).

The authors declare no conflict of interest.

Sketch of tested centrifugal pump with inlet guide vanes (IGVs).

IGVs: (

Schematic diagram of the experimental rig. 1: regulating valve; 2: turbine flowmeter; 3: pressure gauge; 4: motor; 5: composite torque detector; 6: centrifugal pump; 7: IGV; 8: vacuum gauge; 9: regulating valve; and 10: water tank.

Computation domain and mesh of the centrifugal pump with IGVs: (

Experimental and calculated results.

Experimental results of (

Streamlines in pump suction pipe and impeller: (

Experimental results of pump head drop for different flow rates: (^{3}/h; (^{3}/h; and (^{3}/h.

Volume fraction of vapor in impeller (left) and total pressure at the impeller inlet (right): (

Total pressure at cross sections: (

Parameters of the centrifugal pump and IGVs.

Centrifugal pump | Volumetric flow rate ^{3}/h) |
340 |

Head |
30 | |

Rotational speed |
1,450 | |

Blade numbers of impeller _{i} |
6 | |

Diameter of suction pipe _{0} (mm) |
200 | |

Diameter of impeller out _{2} (mm) |
329 | |

| ||

IGVs | Hub of IGVs (mm) | 40 |

Shroud of IGVs (mm) | 200 | |

Blade numbers of vanes _{g} |
6 | |

Blade angle at inlet β_{gi} |
90 |

Pump head

Suction pipe | 723,855 | 723,855 | 723,855 | 723,855 | 723,855 |

Impeller | 417,396 | 1,048,831 | 1,942,195 | 2,714,058 | 3,696,570 |

Volute | 663,491 | 663,491 | 663,491 | 663,491 | 663,491 |

Whole passage | 1,804,742 | 2,436,177 | 3,329,541 | 4,101,404 | 5,083,916 |

_{1} |
1 | 0.99428 | 0.99455 | 0.99394 | 0.99355 |

η/η_{1} |
1 | 1.00331 | 1.00336 | 1.00361 | 1.00380 |

Total pressure at suction pipe inlet and impeller inlet.

No IGVs | 23,341 | 23,086 | 255 |

IGVs 0° | 23,305 | 22,802 | 503 |

IGVs 12° | 23,306 | 22,499 | 807 |

IGVs 24° | 23,305 | 21,528 | 1777 |

IGVs −12° | 23,306 | 22,828 | 478 |

IGVs −24° | 23,304 | 22,462 | 842 |