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The flow between the impeller exit and the diffuser entry (i.e., in the radial gap is generally considered to be complex). With the development of PIV and CFD tools such as moving mesh techniques, it is now possible to arrive at a prudent solution compatible with the physical nature of flow. In this work, numerical methodology involving moving mesh technique is used in predicting the real flow behavior, as exhibited when a target blade of the impeller is made to move past corresponding vane on the diffuser. Many research works have been undertaken using experimental and numerical methods on the impeller-diffuser interactive phenomenon. It is found from the literature that the effect of radial gap between impeller and diffuser on the interaction and on the performance of the fan has not been the focus of attention. Hence numerical analysis is undertaken in this work to explore and predict the flow behavior due to the radial gap. This has revealed the presence of an optimum radial gap which could provide better design characteristics or lower loss coefficient. It is found that there is a better energy conversion by the impeller and enhanced energy transformation by the diffuser, corresponding to optimum radial gap. The overall efficiency also found to increase for relatively larger gap.

A host of articles [

Ubaldi
et al. [

According to Justen et al. [

Sinha
et al. [

It is clear from the above literature survey that the effect of radial gap on the system performance as well as on the impeller-diffuser interaction in a centrifugal fan has not been explored well so far. Hence a numerical modeling of the flow domain which includes a portion of the inlet to the impeller as well as the diffuser with volute casing has been carried out, and moving mesh technique has been adopted for unsteady flow simulation of the centrifugal fan in this analysis.

The centrifugal fan stage consists of an
inlet region, an impeller, a vaned diffuser, and a volute casing (Figure

Specifications of the centrifugal fan.

Centrifugal fan specifications | |
---|---|

Impeller
inlet radius, | 120 mm |

Impeller
outlet radius, | 200 mm |

Diffuser
inlet radius, | 230 mm |

Diffuser
outlet radius, | 300 mm |

Volute exit flange width | 450 mm |

Width of diffuser blade | 35 mm |

Width of volute casing | 90 mm |

Impeller inlet vane angle | |

Impeller outlet vane angle | |

Diffuser inlet vane angle | |

Diffuser outlet vane angle | |

Number of impeller vanes | 13 |

Number diffuser vanes | 13 |

Speed of the fan (RPM) | 1000 rrpm |

Model of the centrifugal fan used in the analysis.

The technical paper by Meakhail and Park [

The grid for the volute part of the
domain has 163 590 nodes and 162 113 elements. The diffuser has 163 213 nodes
and 155 106 elements. The impeller has 80 971 nodes and 74 143 elements. The
inlet part of the domain has 5536 and 5190 nodes and elements, respectively.
The maximum size of the element is limited to elements having an edge length of
2 mm. However, to establish grid independency a finer model having an element
edge length of maximum of 1 mm is carried out and the variation in the results was found to be less than
2.5% and hence to save the computational time, elements edge length of maximum
2 mm size is adopted. Figure

A view of the meshed portion between the impeller and diffuser of the centrifugal fan.

Two-dimensional,
unsteady Reynolds-averaged Navier-Stokes equations set to polar coordinate system
are solved by the CFD code [

A no-slip wall condition is specified for the
flow at the wall boundaries of the blades, the vanes, and also the volute
casing. The turbulence is simulated using a standard

The numerical model for the
whole field flow calculations is validated by calibrating the results of the
current numerical work with the experimental work carried out by Meakhail and
Park [

The graph shown in Figure

Validation characteristics curve of head coefficient with flow coefficient.

The validation curve is a
head coefficient versus flow coefficient curve which shows a decrease in the
head coefficient as the flow coefficient increases as is required for a
backward swept impeller blade. The validation shows a close agreement between
the present numerical model and the experimental model of Meakhail and Park [

A
total of six different configurations are generated by varying the radial gap
between the impeller and the diffuser. The outer radii of the impeller and the
diffuser are kept constant at 200 mm and 300 mm, respectively. However, the
inlet radius of the diffuser is changed to achieve the radial gap variation.Configurations A, B, C, D, E, and F are having a radial
gap ratio “

The
head coefficient versus flow coefficient characteristic curves for all the six
configurations is plotted
after capturing the area weighted average values of the pressure coefficients
and mass averaged flow rates. It is seen from Figures

Performance of centrifugal fan (efficiency versus flow coefficient).

From Table

Computed flow parameters at design point operation of the fan.

Configuration type | Radial
gap ratio | Flow
coefficient ( | Max.
efficiency | Head
coefficient |
---|---|---|---|---|

A | 0.05 | 0.085 | 22.84 | 0.235 |

B | 0.10 | 0.073 | 24.59 | 0.304 |

C | 0.15 | 0.066 | 27.37 | 0.360 |

D | 0.20 | 0.074 | 22.63 | 0.333 |

E | 0.25 | 0.078 | 22.02 | 0.3324 |

F | 0.30 | 0.076 | 20.99 | 0.372 |

Radial
gap is required to avoid steep velocity gradients at the diffuser entry region
according to Yahya [

At
smaller radial gap, the flow enters the stationary diffuser almost at the same
velocity profile as it leaves the impeller and jets and wakes related to the
impeller exit flow do not have a chance to even out. This causes lower diffusion leading to lower diffuser
static pressure recovery coefficient and lower overall static pressure recovery
coefficient across the fan as shown in
Figure

The
shifting of the design point operation at lower radial gap ratio of 0.05 may be
attributed to the fact that the energetic fluid jetting from the impeller
enters the diffuser and decelerates over the diffuser. As a result a large
recirculation zone is formed in many of the diffuser blade passages as seen in
Figure

However,
when the radial gap is relatively larger, even though the jets and wakes
related to the exit flow from the impeller gets evened out, the flow tends to
stall in some of the vanes of the diffuser due to change in angle of incidence, as can be seen in Figure

It is seen in a contrasting manner that
there is a higher mass flow rate corresponding to the design point operation for
the case of larger radial gap ratio of 0.25 from Figure

The static pressure recovery
coefficients and total pressure loss coefficients for various configurations
are presented in the form of bar charts in Figures

It
is seen from Figure

Hence it can be stated that there must be an optimum radial gap which could provide relatively higher efficiency and also lower mass flow rates to achieve the higher static pressure rise.

It can be seen from Figure

Performance characteristics curves for all configurations.

Relative velocity plot for configuration A (radial gap ratio = 0.05).

Relative velocity plot for configuration B (radial gap ratio = 0.1).

Relative velocity plot for configuration C (radial gap ratio = 0.15).

Relative velocity plot for configuration D (radial gap ratio = 0.20).

Relative velocity plot for configuration E (radial gap ratio = 0.25).

Static pressure recovery coefficient.

Total pressure loss coefficient.

Dynamic head coefficient at the design point mass flow rate of configuration C (inlet velocity of 5 m/s).

Figure

Impeller exit static pressure coefficient for the various configurations.

Figure

Diffuser exit static pressure coefficient for the various configurations.

Static pressure coefficient at the exit flange of the fan versus time steps.

The following conclusions can be drawn from the above study.

As a major inference from the above analysis, it is found that there is an optimum radial gap at which better dynamic and static heads are developed by the impeller blades as well as better energy conversion by diffuser vanes.

The above-mentioned facts lead to maximum efficiency of the centrifugal fan as observed in the study.

There appears to be greater degree of stalling of the flow above or below the optimum radial gap.

The static pressure recovery and total pressure loss for the diffusing components of the fan change with the radial gap.

The larger is the radial gap, the smaller are the pressure fluctuations at the exit flange of the fan.

The jet and wake phenomena as seen in all the impeller passages are influenced by the radial gap between impeller and diffuser.

Static pressure at impeller inlet (Pa)

Static pressure at impeller exit (Pa)

Static pressure at diffuser exit (Pa)

Static pressure at flange exit (Pa)

Total pressure at impeller exit (Pa)

Total pressure at diffuser exit (Pa)

Total pressure at flange exit (Pa)

Tangential velocity at impeller exit (m/s)

Absolute tangential velocity at impeller exit (m/s)

Air density (kg/m^{3})

Volume
flow rate (m^{3}/s)

Flow coefficient =

Head
coefficient =

The angle of advance of a given impeller blade to its next adjacent blade position

Fan
efficiency

Overall static pressure recovery
coefficient

Overall total pressure loss coefficient

Diffuser static pressure recovery
coefficient

Diffuser total pressure loss coefficient

Impeller exit static pressure coefficient

Diffuser exit static pressure coefficient

Flange exit static pressure rise coefficient

Radial gap ratio

The authors wish to acknowledge and
thank Tarek Meakhail and Seung O. Park [