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A method is presented for redesigning a centrifugal impeller and its inlet duct. The double-discharge volute casing is a structural constraint and is maintained for its shape. The redesign effort was geared towards meeting the design volute exit pressure while reducing the power required to operate the fan. Given the high performance of the baseline impeller, the redesign adopted a high-fidelity CFD-based computational approach capable of accounting for all aerodynamic losses. The present effort utilized a numerical optimization with experiential steering techniques to redesign the fan blades, inlet duct, and shroud of the impeller. The resulting flow path modifications not only met the pressure requirement, but also reduced the fan power by 8.8% over the baseline. A refined CFD assessment of the impeller/volute coupling and the gap between the stationary duct and the rotating shroud revealed a reduction in efficiency due to the volute and the gap. The calculations verified that the new impeller matches better with the original volute. Model-fan measured data was used to validate CFD predictions and impeller design goals. The CFD results further demonstrate a Reynolds-number effect between the model- and full-scale fans.

Lastly, a rigorous design validation study was undertaken with a carefully designed test rig for the 1/5 scale model. Both fans with the existing impellers and the fan system with the redesigned impeller were tested to verify improvement in performance.

In the following sections, we provide details of the strategy and methodology for redesigning the impeller using the impeller-only CFD calculations. Refined CFD calculations coupling the impeller, the volute, and the shroud gap that were used to assess the design and quantify the volute feedback to the impeller performance are discussed after the design procedure. Following that we provide details of the model-scale fan test [8] and comparisons with the coupled CFD predictions at design and off-design conditions. We end the paper with a detailed summary of the redesign process and the lessons learned therewith.

Figure 4 shows the assembly of the bellmouth and impeller for one half of the fan. Due to the geometrical symmetry, the CFD calculations only cover one single blade passage for the gridding system used, as shown in Figure 5. To accurately capture the boundary layer and loading on the blade surface, the grid on the blade portion is structured and all other surfaces are either structured or unstructured as shown in Figure 5. The unstructured cells help to reduce the overall size of the grid thereby reducing turnaround time for the calculations. Although a relatively small gap exists between the rotating shroud and the nonrotating bellmouth, the impeller-only design CFD calculation does not include the effect of the shroud gap flow.

For the incompressible flow calculation, a uniform inflow condition was imposed at the bellmouth inlet to maintain the required flow rate and a mass-averaged back pressure was applied at the impeller exit. A periodic boundary condition was enforced for the passage boundaries between the blades and a no-slip condition was used at the blade, shroud, backplate, and shaft surfaces. Although the inlet was controlled with a velocity condition, the inlet pressure was predicted as part of the simulation since the pressure pertains to the upstream propagating characteristic. As a consequence, the pressure rise was determined from the difference between the inlet and exit pressures and is a function of the impeller design.

Impeller B#2 was used to investigate the grid density requirement. Figure 6 shows the computed percent change in ShaftPWR versus the design power with the number of cells for the structured and unstructured grids ranging from 105,984 to 958,464 cells. The result shows that a grid density of 250,000 cells or more for each impeller blade passage is adequate for a predicted power with an error of 0.5% (mostly dependent on the grid topology rather than the grid density) or less. Calculations were also performed to investigate the effect of using the wall-function procedure. The grid ?+ was controlled between 10 and 50 for the wall-function modelling and below 1 for the near-wall modelling. Calculations were made for both B#1 and B#2 impellers with an approximately 250,000 cell grid. The predicted ShaftPWR is generally lower for the near-wall modelling, but the difference between the B#1 and B#2 impellers using the same wall modelling is almost the same between the two models studied.

The predicted ShaftPWR for each impeller is lower than the targeted ShaftPWR (or PWRref). Although the B#2 impeller requires more power at the specified condition, it generates more head and has a slightly higher efficiency. Since flow separation occurs in each impeller while operating at high efficiency, the redesign calculations must accurately account for all aerodynamic losses in order to predict any performance difference within a few percentage points. The use of streamline curvature or potential-flow/Euler codes would not accomplish the goals for the current redesign effort. The developed redesign procedures established based on the findings from the assessment of the existing impellers are herewith provided below.

Since flow separations at the shroud in front of the blade leading edges were predicted for the two existing impellers, further improvement in impeller performance would require reducing this shroud flow separation. The large curvature of the shroud as it approaches the blade may be partially responsible for the flow separation seen at the shroud due to the difficulty of the boundary layer to remain attached as the flow negotiates the turn near the shroud. In Figure 9(a), three bellmouth/shroud profiles are presented and labelled based on the local curvature near the blade and shroud intersection. The profile labelled with 0.0263 (local radius of curvature/D) corresponds to the B#2 impeller. The two other profiles were investigated to reduce the sharp curvature at the blade intersection [13]. The associated flow fields of all three profiles indicate that the original flow separation at the shroud was improved in the two new profiles. Figure 9(c) demonstrates the improvement of the impeller with the 0.0476 shroud as compared to the B#2 impeller shown in Figure 8. The performance data shown in Figure 9(b) suggests that the shroud labelled with 0.0476 provides the largest gain in efficiency. Although the required power for the 0.0476 shroud is slightly increased, it is used in the final design.

CFD prediction results were also made for the 11-bladed B#2 impeller, which was constructed based on the 12-bladed impeller to maintain a constant throat area, that is, at the location with the maximum blade thickness. The advantage of adapting the 11 blade arrangement is to reduce ShaftPWR by 2.38% for the impeller with the 0.0476 shroud as compared with the 12-bladed impeller with the same shroud curvature. Although a drop of 2.14% in total head for the latter impeller occurred, the efficiency was maintained. These results led to the decision to choose the 11-bladed 0.0476 shroud profile impeller configuration. In addition, this modification required a blade redesign to recover the drop in the total head.

Both existing impeller's blades were primarily 2D blades, that is, the leading and trailing edges at hub and shroud started at the same radii. The blade was designed as a 2D blade to reduce the manufacturing cost. There are some advantages to sweeping the blade: (i) a blade starting at a lower radius near the shroud can prevent boundary-layer separation by accelerating the flow before it actually turns, and (ii) it changes in incidence at the leading edge attributed to the sweep can lower losses and increase efficiency. Based on this concept, the B#2 11-bladed impeller blades were extended inward radially at the leading edge and its angle measured from the shroud was modified from 0 degree for a 2D blade like the B#1 blade to 10 degrees. The new 3D blade generated high head of 1.548 ?ref versus 1.471 ?ref with a higher efficiency of 95.08% versus 93.66% at the expense of a higher shaft power of 0.968 PWRref versus 0.936 PWRref. This procedure essentially improves the blade efficiency. When the same procedure was applied to the steer blade shown in Figure 15, the efficiency improved from 93.8 to 95.55%, the head increased from 1.414 ?ref to 1.459 ?ref with the shaft power also increasing from 0.896 PWRref to 0.909 PWRref. From here on out, when this 3D version of the steer blade is integrated with the impeller, it is referred to as the NEW design impeller.

For the impeller-flow calculation, all boundary conditions used for the CFD design calculations were maintained except for eliminating the periodic boundary condition and controlling the exit back pressure through the interface information exchange. For the volute-flow calculation, the mass-averaged discharge pressures from the two exits are prescribed to keep (a) the required flow to the lift side, (b) the extended surface from the impeller backplate modelled as a symmetry plane, (c) the shroud as the rotating wall, and (d) all other casing surfaces as no-slip walls.

The grid topology used for the impeller design calculation shown in Figure 5 was maintained. Depending on the number of blades designed for each fan, the total impeller grid was approximately 3 to 4 million cells. The corresponding volute for each fan had approximately 1.5 million cells. The converged volute solution for the baseline B#1 impeller was first obtained by adjusting the pressures at the two exits to reach the design lift flowrate. Similar exit pressures were applied for all other impeller calculations to obtain the lift flowrates shown in Tables 2 and 3.

Table 2 shows the performance data obtained from the impeller/volute coupling calculations for all fans. Adapted from the grid topology used for the impeller design CFD, the impeller grid ended at a fixed radius for all coupling calculations except for the NEW impeller, which ended at a slightly smaller radius. In order to compare the performance with similar grid features for all fans, the NEW-x grid was generated by radially extending the shroud of the NEW impeller. Since the impeller width plays an essential role in the impeller performance, a wider width impeller was generated for comparison and is labelled as the NEW-w impeller. 2b1af7f3a8