two blade fan

Design Refinement and Performance Analysis of Two-Stage Fan for Small Turbofan Engines

J. Masud*, and S. AhmedDepartment of Aerospace Engineering, College of Aeronautical Engineering, National University of Sciences &

Technology, Risalpur 24090, Pakistan

This paper presents the computational study and first iteration of design refinement of a two-stage fan that has been designed for small turbofan engines. 3-D computational analysis of the original configuration indicated that the fan is unable to meet the design point parameters. Systematic flow field analysis revealed certain problems with the 1st stage rotor blade that could be improved by making small adjustments to its geometric features; such as blade lean, sweep, chord and contour. With these changes incorporated in modified design, the computed performance maps for both the original and modified fan 1st stage rotor were created. These maps indicate an overall gain of 6% and 8% in fan 1st stage rotor design point pressure ratio and isentropic efficiency, respectively. The performance of the complete two-stage fan, with modified 1st stage rotor, was then computationally evaluated and its operation map was constructed. After the 1st stage rotor redesign, the whole fan (two-stage) design point efficiency and pressure ratio improved by 5-7% and 1.4%, respectively. With this predicted improvement in its performance, the redesigned fan meets the projected design point parameters.

Nomenclature CFD = computational fluid dynamics LE = leading edge TE = trailing edge PS = blade pressure side SS = blade suction side Cp = pressure coefficient M = Mach number P = static pressure Pt = total pressure RNGKE= two equation Re-Normalized Group K- turbulence model SA = one equation Spalart-Allmaras turbulence model x = Cartesian coordinate y = Cartesian coordinate z = Cartesian coordinate (axial direction) y+ = non-dimensional length scale associated with turbulence model

I. Introduction n the field of turbo-machinery, Computational Fluid Dynamics (CFD) is becoming an important tool to analyze the flow field inside turbines and compressors1,2,3,4. By this technique the performance analysis of turbo-machine components, or the complete machine itself, under various conditions, can be simulated. Due to this aspect, CFD

offers greater flexibility to designer since various aspects of the design can be simulated before expensive testing is done.

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* Associate Prof, Dept of Aerospace Engg, College of Aeronautical Engg, NUST, Risalpur 24090, Pakistan. Visiting Engineer, Dept of Aerospace Engg, College of Aeronautical Engg, NUST, Risalpur 24090, Pakistan.

American Institute of Aeronautics and Astronautics

1

45th AIAA Aerospace Sciences Meeting and Exhibit8 - 11 January 2007, Reno, Nevada

AIAA 2007-23

Copyright 2007 by Jehanzeb Masud. Published by the American Institute of Aeronautics and Astronautics, Inc., with permission.

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the hub were removed from all blades to facilitate the use of structured grid. This is conservative approach from flow analysis aspect.

The modeled flow domain is shown in Fig. 2 (2nd stage of fan is not shown) where a bell-mouth intake was added so that static sea-level conditions could be applied far from the fan 1st stage rotor. By modeling the computational domain in this manner the uncertainty caused by upstream conditions on computed results was avoided.

Mapped quadrilateral elements were used to mesh the flow domains of both rotors and stators. The element arrangement was generally O-type. This type of structured mesh allows good distribution control to the user and is computationally economical while capturing relevant flow physics. Elements were graded towards the blade surface, hub, rotor blade tip and casing in order to accurately resolve near wall flow phenomena in the vicinity of these areas. A representative computational grid is shown in Fig. 3 along with the modeled periodic surfaces and mixing planes.

Fluent finite-volume based CFD code6 along with its preprocessor7 Gambit were used in the present study. The three-dimensional compressible Reynolds-Averaged-Navier-Stokes (RANS) system of equations, with appropriate turbulence model and variable property air as fluid, was solved using the coupled-implicit formulation of Fluent. No-slip velocity boundary condition8 was enforced at the surface of blades, hub and casing. Zero-shear or slip boundary condition was enforced on bell-mouth intake and intake dome in order to avoid refined grid in these areas (Fig. 2 and 3). This also helped in avoiding viscous losses at these locations since the bell-mouth intake and dome are an added feature for CFD analysis and are not part of fan design. Static sea-level pressure was specified at the static-far-field ahead of

bell-mouth intake (Fig. 2), while at the exit of the fan varying (discrete) pressure was specified corresponding to the operating condition being computed. Since the thermal problem was not of paramount importance in the present study, therefore all the blade surfaces, hub, casing etc were modeled as thermally insulated8. The CFD software package used in the present study (Fluent6,8) is well established in CFD community and the accuracy of its results has been verified for a number of complex problems8, including turbo-machinery analysis.

A. Turbulence Modeling The one-equation Spalart-Allmaras (SA) turbulence model9 has been used extensively for the present study. The

SA turbulence model8,9 is a low Reynolds number model and does not require the use of wall functions as such.

Figure 4. Flow traces through the modified fan at design point operation.

Figure 3. Computational grid in the vicinity of fan


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However, in Fluent it is incorporated in a manner that allows the use of wall functions8. This results in less dense near-wall grid to be used, which saves on computational cost of present analysis. Additionally, the SA turbulence model is computationally less expensive than two-equation family of turbulence models and is recommended for turbo-machinery analysis8.

In order to check the sensitivity of present analysis to the turbulence model used, design point computations were also done using Re-Normalized Group K-, (RNGKE) and Shear Stress Transport (SST) turbulence models8. The results indicated around 1-2% variation in integrated parameters such as total pressure ratio, mass flow rate and isentropic efficiency. Some minor flow variations were apparent but were not considered crucial for the present analysis since their effect is considered smaller than the mixing-plane approximation already incorporated in the employed computational method (discussed earlier in this paper). Therefore, all the computed results presented in this study are based on the SA turbulence model.

The computed flow traces at design point are shown in Fig. 4 for two different turbulence models for the re-designed fan. This was part of the turbulence model dependence study discussed above. The overall flow field is quite similar for the one-equation Spalarat-Allmaras (SA) model and the two-equation Re-Normalized-Group K- (RNGKE) model. The flow spillage due to rotor tip clearance is evident from Fig. 4. The detrimental effect of this spillage flow on subsequent stators is also visible by the distorted / separated stream traces near the stator tips.

B. Grid Independence Analysis Numerically computed results change with the type and

density of mesh / grid used for computations. Several different mesh sizes were initially used in the present study to see the grid dependence of computed results at design point conditions. The mesh size that produced adequate grid independent results is described here.

As discussed earlier, O-type grid has been used for all blade domains. Therefore, only the mesh surrounding the blades will be described. For the 1st stage rotor, O-type grid has 141 points around the blade, 21 points from the blade to the periodic boundary (or mixing plane) and 71 points span wise i.e. 1412171. The corresponding 1st stage stator O-type grid is 14131141. The O-type grid for 2nd stage rotor and stator is 1612171 and 1212781 respectively. The total number of nodes in the domain is 1.12 million. This number is near the maximum that can be handled efficiently by the available computational resources.

As discussed earlier, grid points are graded towards the blade surface, hub, rotor blade tip and casing this ensures that the computed turbulent law y+ remain below 100 (between 30~100). These values of turbulent y+ are within the normal range8,10 for application of standard wall functions and show reasonable resolution of grid near the walls8,10.

II. Results and Discussion The sequence of computations done in the present study has been briefly discussed earlier in this paper. First the

initial analysis of the 1st stage of the subject fan was done under different operating conditions. Based on this analysis, the 1st stage design was reviewed and the 1st stage rotor blade was identified for design refinement. The refinements (first-iteration) to the 1st stage rotor include small adjustments in blade lean angle, sweep back, chord and profile. After the blade shape refinement, the fan 1st stage flow field was re-computed under different conditions in order to construct the operating map. Finally, the complete two-stage fan with refined 1st stage rotor is computationally analyzed under different conditions in order to construct its operating map. In this section the results of these numerous and extensive computations are discussed in the sequence just described.

The basic flow field in the fan is shown in Fig 4. Under design point operation the flow in the fan generally follows the channel between the blades. The flow spillage, from the rotor blade pressure side towards its suction side, due to rotor tip clearance is evident from Fig. 4. The detrimental effect of this spillage flow on subsequent

Figure 5. 1st stage rotor performance map atdesign point speed (100% rpm).


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stators is also visible by the distorted / separated stream traces near the stator tips. These flow traces become radically disturbed under stall conditions, which are not shown here.

A. Analysis of Original Fan 1st Stage The results of computations of original fan 1st stage

indicated lower than expected total pressure ratio and isentropic efficiency under design point conditions. Figure 5 partially shows the 1st stage rotor performance map (at 100% speed) derived from the computed results. The modified design, i.e. with redesigned 1st stage rotor, is included in this figure for brevity. All modified design discussion is done subsequently in this paper. The reason for low pressure ratio and isentropic efficiency were investigated by looking at the 1st stage flow field. Region of flow reversal / separation on the 1st stage rotor blade was apparent, as shown in Fig. 6 (a). In this figure, negative axial velocity indicates flow reversal in the domain, and on the blade surface it indicates flow separation. The region of flow separation in the fluid domain extends from 5% span to 95% span. Most of the separation is evident at 10-60% span.

At design speed, design point pressure ratio is achieved at significantly lower mass flow rate than design mass flow rate for the original fan 1st stage (Fig. 5). Increasing the rotational speed to about 106 % of design point rpm can help achieve the desired pressure ratio at design mass flow rate, as is seen form performance maps of original and modified (redesigned) 1st stage shown in Fig. 11 (discussed later in this paper), but this solution is not desirable due to anticipated mechanical and aerodynamics problems. Increasing the design rotational speed by 6% increases blade stress and drastically reduces the surge margin to the extent that the blade is operating almost at surge condition at design point (Fig. 11). Based on this aspect, the 1st stage rotor redesign was carried out, which is discussed next.

B. Design Refinement of Fan 1st Stage Rotor Previous studies1,2,3,4 have investigated the effect of blade lean,

sweep and increased chord length in tip region numerically and experimentally. The combination of these parameters is considered important for efficient rotor blade operation and performance. The original 1st stage rotor had a bit of forward lean with small curvature in leading edge that results in compound sweep. The first modification to the rotor blade was straightening of its leading edge to get almost zero blade lean with almost constant sweep back angle. The trailing edge of the rotor blade was straightened from 30% to 100% of span so as to decrease the adverse pressure region that was causing flow separation (Fig. 6). These initial (first-iteration) modifications to the original 1st stage rotor resulted in a rotor blade that is designated as the modified 1st stage rotor blade.

The original geometry and modified geometry is shown in Fig. 7 and 8. Fig. 7 shows the general three dimensional view, while Fig. 8 shows the blade sections (profiles) at three locations along the span. From these figures it can be seen that the blade lean has been reduced for the modified design compared to the

a) Original Design b) Modified Design

Figure 6. Negative axial-velocity region (flowreversal / separation) in the vicinity of 1st stage rotorblade at design point operation.

Figure 7. Geometric comparison betweenoriginal and modified 1st stage rotor blade.

Figure 8. Blade sections of original andmodified 1st stage rotor.

Original Design Modified Design

50 % Span

5% Span

95% span

L.E

T.E


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original blade. The modified blade has lesser sweep as well. The blade chord of the modified design has also increased from 50% of span onwards till 100% of span. From Fig. 7 and 8 it can be seen that the leading edge of the modified blade is almost straight compared to some curvature seen in the original rotor blade. The blade chord length of the modified blade is somewhat larger than the original blade, especially in the tip region (100% of span, Fig. 8). The increase in chord length of blade is gradual from about 50% of span to all the way to tip. This increase in blade chord of the modified design is due to straightening of the leading and trailing edges, as discussed earlier. This is expected to improve flow stability and blade efficiency in the tip region.

The effect of the design refinement on modified blade characteristics can be seen from Fig 6. The flow separation region seen in the original design has significantly reduced for the modified design where it is now localized at the trailing edge. The effect of this reduced flow separation on performance of the modified 1st stage is discussed later in this paper.

The pressure distribution at various sections of original and modified rotors, at design point mass flow rate, is shown in Fig. 9. At 5%-of-span location on pressure side, the pressure distribution for original design indicates favorable gradient followed by adverse gradient. Whereas, in modified design, the gradient changes from neutral to adverse in a smooth manner, which indicates better diffusion behavior. At the suction side (5% of span) some improvement in diffusion is also noticeable for the modified design. At 50% -of-span location (Fig 9), similar trend as that of 5% location is evident, however the pressure gradients are more severe (than 5% span location) for the original design. Consequently the improvement seen in pressure distribution (and diffusion) for the modified design is even more pronounced. At 95%-of-span location (Fig. 9), the modified design rotor blade shows good diffusion characteristics while the original design exhibits extremely poor characteristics. This improvement is partially due to zero-lean for the modified rotor blade. The maximum Mach number near the tip region for the original blade is slightly above 1 while for the modified design it is slightly below 1, which indicates better flow characteristics. The effect of better diffusion characteristics for the modified design is expected to improve its performance, which is discussed in next section.

Static pressure distribution on rotor blade surfaces of the original and modified designs is shown in Fig. 10. This figure shows the overall effect of rotor blade design refinement on surface pressure. It is evident from Fig. 10 that both span wise and chord wise pressure distributions have improved for the modified design rotor compared to the original design. The modified blade is better (evenly) loaded than the original blade. The effect of improved pressure distribution (Fig. 9 and 10) has resulted in better behaved flow (Fig. 6) and is expected to improve the performance of the modified blade compared to the original blade.

C. Performance Prediction of Modified (Redesigned) Fan 1st Stage The reduced separation region (Fig. 6) and improved blade loading, as discussed earlier, has resulted in

improved efficiency for the modified 1st stage rotor. At design speed, the rotor efficiency has improved by about 8% with corresponding increase in pressure ratio of about 6% as shown in Fig. 5. For the fan 1st stage with modified

Figure 9. Surface pressure distribution at different sections of original and modified 1st stage rotor blade at

design point operation.

Figure 10. Qualitative static pressurecontours on 1st stage rotor blade at designpoint operation.


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rotor similar trend in results is noted with some losses associated with stator. The upstream tip clearance flow from rotor that affects stator flow near casing (Fig. 4), discussed earlier, contributes to stator losses.

Figure 11 illustrate compressor map of fan complete 1st stage (rotor + stator). The original design map is constructed from 87.5% to 106.25% of design speed, where as for modified design, the map is made from 81.25% to 106.25% of design speed. From this map (Fig. 11), it can be seen that the pressure ratio and efficiency of the original design is very low. The fan 1st stage based on the original design can achieve design pressure ratio at lower-than-design mass flow rate or it can achieve design mass flow rate and design pressure ratio at 106% of design speed. In this case the rotor is operating near surge (Fig. 11), as discussed earlier. However, the modified design is able to achieve the design mass flow rate, pressure ratio at 100% design speed. The modified design stage has over all higher efficiency then the original design at all operating speeds. The surge margin for the modified design has improved significantly (Fig. 11).

D. Performance Prediction of Complete Two-Stage Fan with Redesigned / Modified 1st Stage Rotor

Over all fan performance map for two-stage fan, with modified 1st stage rotor, is shown in Fig. 12. After the 1st stage rotor design refinement, the fan is able to meet the design point mass flow rate, pressure ratio and minimum isentropic efficiency at design rotational speed. After the 1st stage rotor redesign, the whole fan (two-stage) design point efficiency and pressure ratio has improved by 5-7% and 1.4%, respectively.

The design refinements made to the 1st stage rotor (discussed earlier) have made positive effect on fan 1st stage performance, but as emphasized earlier, these are first iteration (initial) refinements. The whole fan (two-stage) performance can be improved by carrying out subsequent refinement iterations on 1st stage rotor as well as looking at other fan components like 2nd stage rotor and stator etc. in greater detail.

III. Conclusion In the present computational study, a small two-stage fan has been analyzed in detail. Based on initial results,

first iteration of design refinement has been done on 1st stage rotor. Due to improved blade loading, the refined design (modified) 1st stage rotor exhibits a design point improvement of 8% and 6% in isentropic efficiency and pressure ratio, respectively. This translates into 5-7% and 1.4% improvement in whole fan design point isentropic efficiency and pressure ratio, respectively. The complete performance characteristics of the redesigned 1st stage and complete fan have been evaluated, both of which indicate satisfactory performance consistent with other fans of the same category. Further design optimization of the subject fan is possible through subsequent refinement iterations.

Figure 11. Fan 1st stage operation map for original and modified rotor designs.

Figure 12. Modified fan (two-stage)

mass flow rate

isen

tropi

c ef

feci

ency

5%

mass flow rate

pres

sure

ratio

87.5

106.2

100

93.75


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References 1Bergner, J., Hennecke, D. K., Hoeger, M. and Engel K. Darmstadt Rotor No. 2, Part II, Design of Leaned Rotor Blades,

ISROMAC 9 FD101, 9th International Symposium on Transport Phenomena and Dynamics of Rotating Machinery (ISROMAC-9), Honolulu, Hawaii, USA, Feb. 10-14, 2002.

2Kablitz, S., Bergner, J., Hennecke, D. K., Beversdorf, M. and Schodl, R., Darmstadt Rotor No. 2 Part III, Experimental Analysis of an Aft-Sweept Axial Transonic Compressor Stage, ISROMAC 9 FD089, 9th International Symposium on Transport Phenomena and Dynamics of Rotating Machinery (ISROMAC-9), Honolulu, Hawaii, USA, Feb. 10-14, 2002.

3Li, H-D., He, L., Li, Y. S., and Wells, R. Blading Aerodynamics Design Optimization with Mechanical and Aeromechanical Constraints, GT-2006-90503, Proceedings of GT2006 ASME Turbo Expo 2006: Power for Land, Sea and Air, Barcelona, Spain, May 8-11, 2006.

4Gallimore, S. J., Bolger, J. J., Cumptsy, N. A., Taylor, M. J., Wright, P. I. and Place, J. M. M., The Use of Sweep and Dihedral in Multistage Axial Flow Compressor Blading Part 1: University Research Methods Development, ASME J. of Turbomachinery, Vol 124, pp421-532.

5Cumpsty, N. A., Compressor Aerodynamics, Longman Scientific & Technical, Essex, UK 1989. 6FLUENT, Computational Fluid Dynamics Software Package, Ver. 6.2.16, Fluent Inc, Lebanon, NH, 2004. 7GAMBIT, Geometry and Mesh Generation Software Package, Ver. 2.2.30, Fluent Inc, Lebanon, NH, 2005 8FLUENT, Computational Fluid Dynamics Software Package User Guide, Ver. 6.2.16, Fluent Inc, Lebanon, NH, 2004. 9Spalarat, P., and Allmaras, S., A One-Equation Turbulence Model for Aerodynamic Flows, Technical Report AIAA-92-

0439, 1992. 10White, F. M., Viscous Fluid Flow, 2nd ed., McGraw Hill, New York, 1991, Chaps. 6, 7.

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