Proceedings of the ASME 2011 International Mechanical Engineering Congress & Exposition IMECE2011 November 11-17, 2011, Denver, Colorado, USA

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1 Proceedings of the ASME 2011 International Mechanical Engineering Congress & Exposition IMECE2011 November 11-17, 2011, Denver, Colorado, USA IMECE THE EFFECT OF BLADE LEAN, TWIST AND BOW ON THE PERFORMANCE OF AXIAL TURBINE AT DESIGN POINT Hadi karrabi Research Engineer Middle east Petro Gas Company,Tehran, Iran Mohsen Rezasoltani Research Engineer Middle east Petro Gas Company,Tehran, Iran ABSTRACT An investigation to understand the impact of twisted, leaned and bowed blades on the performance of axial turbine was undertaken. A CFD code, which solves the Reynolds-averaged Navier Stokes equations, was used to compute the complex flow field of axial turbine. The code was validated against existing Hannover turbine experimental data. Numerical data showed good agreement with measured data. Finally, the effect of geometry changes, focusing on blade lean, twist and bow, on the Avon turbine blade performance, was analyzed. Results show that twisted blade affects performance significantly. Leaned and bowed blade has minor effect on performance. 1 INTRODUCTION Turbine and compressors are two of the most important components of many gas turbines and their behavior has great effect on whole engine performance. One of the key components of the compressor and turbine are the blades. Very few of the manufactured blades meet their design performance requirement. Sometimes different manufacturing techniques can significantly affect the blade performance. For example forging methods allow for increased quantity, but have the effect of reducing the quality and performance of a blade, compared to that of a case blade. Each chosen method of manufacture affects the blade design and overall whole engine performance. Therefore, investigating the impact of manufacturing precision on the performance of a single blade is important. In this research, the effect of geometric changes, which include blade lean, twist and bow, on the multistage axial turbine is investigated. The turbine blades investigated during this research are the second stage rotors of an Avon 14 series gas generator turbine. Two methods are considered in this investigation: using experimental methods, or numerically simulating the turbine behavior. Because performing experimental test consumes more time and money, computer simulation is chosen to investigate these effects. There are many methods that can be used to simulate axial turbine such as 0D, 1D, 2D, quasi 3D and 3D. In order to achieve the main objective, the numerical model must enable us to input all the detail geometry. 3D simulation of turbomachinery has the highest accuracy. However, it is extremely time-consuming. In 3D-CFD models Navier- Stokes equations can be solved without any simplification. In recent years, several researchers like Gu et al. [1]; Muggli et al. [2]; Cravero and Marini [3] simulate turbomachines with this method and they showed good agreement between simulation and experimental results. A 3D-CFD simulation of the impeller and volute of a centrifugal pump has been performed using CFX codes by Kouidri et al [4]. In their work, a 3D flow simulation for the impeller with structured grid backed up by a sensitivity analysis regarding grid quality and turbulence models was presented. They used k-ε, k-ω and SST for turbulence model and showed that the results of the three models were similar. The final impeller model obtained was used for a 3D quasiunsteady flow simulation of the impeller-volute stage. Gonzalez [5] showed capability of a numerical simulation in capturing the dynamic and unsteady flow effects inside a centrifugal pump due to impeller-volute interaction. For numerical simulation, viscous Navier-Stokes equations were handled with an unsteady calculation and sliding mesh technique to take into account the impellervolute interaction. Time averaged numerical results were compared with experimental performance curves and a good agreement was shown. Many researchers have investigated how geometric changes in a blade design can affect its performance. Benini et al [6] investigate the effect of sweep and lean blade on transonic compressor. In addition effect of lean and sweep on position of shock and secondary loss are shown. Their work demonstrated that optimizing the sweep and lean of the blade gave cause to a 1.3% increase in thermodynamic efficiency. Marco Montis et al [7] investigate the influence of trailing edge bleeding on the Aerodynamics of a NGV Cascade experimentally and numerically. They used ANSYS CFX 11 for numerical simulation and good agreement between simulation and experimental results is shown. In this project, an initial model is presented and validated using Hannover turbine experimental results [8]. Then, a 2nd stage rotor blade row of an Avon gas generator turbine is modeled and analyzed. Performance metric sensitivities and blade row performance are evaluated. All activities accomplished during this research are depicted in the following flowchart: 1 Copyright 2011 by ASME

2 Fig. 3-Structureal grid on blade surface Fig. 1-Flowchart of all activities accomplished during this research 2 Numerical Simulations In Fig. 2 the numerical simulation process in turbomachines is shown. In this part, 3D flow field inside the axial turbine including one stage rotor and stator is numerically analyzed. Modeling Grid generation Initialization Solution Fig. 2-Numerical simulation process Postprocessing 2.1. Grid Generation The first and most important step towards simulating turbomachine behavior is the geometry definition and grid generation, which is usually the most time-consuming step. Selection of grid type and locations for grid refinements can affect the accuracy of the results and convergence rate. Structural elements are used for grid generation of the compressor. The generated grid on the blade to blade surface and periodic surface between blades are shown in Figs. 3 and 4, respectively. Fig. 4-Structureal grid on periodic surface between blades Theoretically, the errors in the solution related to the grid size should disappear by increasing the mesh resolution ([9]). Here, to obtain a suitable accuracy for the numerical results, three grid sizes are used to calculate the efficiency and pressure ratio. As seen in Figs. 5 and 6, the calculated pressure ratio and efficiency reach an asymptotic value as the number of elements increase. According to these figures, the grid B with 1.46 million elements is considered to be sufficiently reliable to ensure grid independency. Effeciency Grid A Grid B Fig. 5- Effect of grid size on efficiency Grid C Number of Elements 2 Copyright 2011 by ASME

3 open loop configuration with exhaust to atmosphere. The air is supplied by three parallel screw compressors with a maximum volume flow of 11 m3/s at a maximum pressure of 4 bar. The blading is of the free vortex type and the nominal mass flow rate of turbine is and the nominal rotational speed is 7500 RPM. The setup allows each stage of the four-stage turbine to be tested separately. The tests referred to in this research are concerned only with stage No.3 which was tested in singlestage mode by Groschup [8]. In Fig. 7 a schematic of the test arrangement of Hannover turbine s stage No.4 is shown Grid A Pressure Ratio Grid C Grid B Number of Elements Fig. 6- Effect of grid size on pressure ratio 2.2. Numerical Method The Reynolds-Averaged Navier-Stokes equations (RANS) which describe the conservation of mass, momentum and energy are solved by means of a finite volume method. The discretization of the equations is done via a coupled implicit method in which the energy, momentum and mass equations are solved together. The Reynolds stress terms in the momentum transport equations are resolved using the shearstress transport (SST) turbulence model, developed to blend the robust and accurate formulation of the k-ω model in the near-wall region with the free-stream independence of the kε model in the far field [10]. Using the mixing plane interface model, the computational domain was divided into stationary and moving zones and utilizes relative motion between the various zones to send calculated values between zones. To complete the model in rotating zones, the coriolis and centrifugal accelerations are added to the momentum equations. The mixing plane method is a non-physical snapshot approach which cannot be compared to a snapshot of a transient simulation, because the solution does not know anything about what happened at the previous time step. Fig. 7- Schematic figure of Hannover turbine s stage No.4 test arrangement The grid generated and used in this solution consisted of 2,500,000 nodes which was fine enough to validate the results. The experimental results presented by Groschup were compared against the simulation results. Very good agreement was observed between the experimental and theoretical results. For instances, the 4th-stage characteristics resulted by numerical solution is shown in Figures 8-12 and Table 1. Stator outlet angle Boundary Conditions The boundary conditions for all simulations are as follows: 1- Mass flow rate applied normal to the inlet area with no pre-whirl. Also, the total temperature, the turbulence intensity and hydraulic diameter should be introduced to the inlet. This boundary condition is used for input of impeller. 2- Stationary adiabatic walls: zero velocity or no slip condition. 3- Moving adiabatic walls: zero relative velocity with respect to the rotating reference frame. 4- Mixing plane condition is used as an interface between the rotor and stator blade. Mixing plane condition approach is ideally placed half-way between the rotor and stator blades, and flow variable fluxes are averaged tangentially using the mass-weight approach to obtain span-wise profile that can be passed back and forth between rotating and stationary domains. experiment calculated hub to shroud Fig. 8- Nozzle outlet angle from hub to shroud 3 Simulation and validation of Hannover turbine This turbine was investigated by G. Groschup at Technical University of Hannover, Germany. The turbine facility is an Fig. 9- Variations Nozzle Axial Exit Velocity in Span Wise Direction 3 Copyright 2011 by ASME

4 consists of 71 blades and the stator has 40 blades. Due to symmetry within the turbine, periodic boundary conditions are used for simulation. In this condition, flow is simulated for one blade and results develop for other blades. Due to the mismatch in the number of stator and rotor blades, equal mixing plane conditions are used for the simulation. In Fig. 13 the boundary conditions are shown. Fig. 10- Variations Nozzle Outlet angle in Span Wise Direction Fig. 13- boundary condition use for simulation Fig. 11- Variations Rotor Axial Exit Velocity in Span Wise Direction 5. Avon simulation Results In this section, the simulation results of the Avon turbine are presented. Results are compared with available manufacturer s data (Fig.14) and good agreement is seen. Simulation results are shown in figures Fig.14- gas flow diagram in Avon turbine Fig. 12- Variations Rotor Outlet angle in Span Wise Direction Figs shows good agreement between experimental and numerical solution data, In Fig. 15 relative Mach number contours on the 20% spanwise section between the hub and shroud is shown. This figure shows velocity decrease in pressure side and increase in suction side. In Fig. 16 radial relative Mach number contours on the rotor inlet is shown. Boundary layer effects near walls are obvious. The numerical solution results given in Table 1, which compare the total isentropic efficiency with experimental results, reveal approximately 0.5 percent absolute error. Table 1- Comparing numerical and experimental results Efficiency Degree of Reaction Numerical Experimental SIMULATION OF AVON TURBINE After validating the model against the Hannover turbine, we next simulate the Avon turbine. As mentioned in the introduction section, in order to achieve the main goal of this project, second stage rotors of an Avon 14 series gas generator turbine must be simulated. In this stage the Rotor Fig. 15-Relative Mach number contours on the 20% spanwise section between hub and shroud 4 Copyright 2011 by ASME

5 Figure 17 shows velocity contours on the mean spanwise section between the shroud and hub. Also in this figure boundary layer effects near the walls are obvious. Figure 18 shows 3D flow variations in rotor blades. Blade Lean. Nonlinear deviation of cross section of blade from hub to shroud Blade bow. Rotation of cross section and deviation from nominal geometry 6.1. Blade Twist In Fig. 19 twisting of blade from nominal geometry is shown. In this project investigation is done on six angles: -3,-2,-1, 1, 2, 3. It is assumed that angles rotate equally from hub to tip. In figures variation of mass flow, efficiency, pressure ratio and power is shown. Rotation of inlet and outlet angles Fig. 16-Radial relative Mach number contours on the rotor inlet Fig. 19 twisting of blade Mass flow(kg/s) Fig. 17- Velocity contours on the mean spanwise section Fig. 20-Mass flow versus nominal geometry deviation Efficiency Fig. 18-3D flow variations in Rotors blade Investigation the impact of geometric variability on performance In this paper three effects of geometric variability is modeled and their influence is investigated. These three effects are: Blade Twist. The undesirable variation and rotation in pitch from root to tip of airfoil s from nominal geometry Fig. 21-efficiency versus nominal geometry deviation 5 Copyright 2011 by ASME

6 Fig. 22-Pressure Ratio versus nominal geometry deviation Fig. 23-Power versus nominal geometry deviation Results show that with a positive rotation from nominal geometry (increase in installation angle), the mass flow rate decreases and consequently, the power also decreases as the flow area reduces in size. On the other hand, with negative rotation, the flow area increases and consequently mass flow and power also increase in value Blade Lean Total Pressure Ratio Power(MW) As shown in Fig. 2, Lean occurs when the central stacking axis of airfoils deviate in the X direction, from hub to tip. Fig. 24-Blade Lean In this project investigation is done on six deviations from stacking axis: -4mm,-2mm,-1mm, 1mm, 2mm, and 4 mm. It is assumed that deviation happens linearly from hub to tip. In figures variation of mass flow, efficiency, pressure ratio and power is shown Mass flow(kg/s) Fig. 25-Mass flow versus nominal geometry deviation Fig. 26-efficiency versus nominal geometry deviation Fig. 27-Pressure Ratio versus nominal geometry deviation Efficiency Total Pressure Ratio Power(MW) Fig. 28-Power versus nominal geometry deviation In accordance with results, the blade lean angle does not affect performance severely. 6 Copyright 2011 by ASME

7 6.1. Blade Bow If the deviation from the stacking axis does not change in a linear manner, from hub to tip, the blade is defined as being bowed (Fig. 29). In this project investigation is carried out on six deviations from the stacking axis: -0.6mm,-0.3mm,-0.15mm, 0.15mm, 0.3mm, and 0.6 mm. It is assumed that deviation occurs from the blade mid-section, and the hub and shroud position remain fixed. In figures variation of mass flow, efficiency, pressure ratio and power is shown. 1.4 Total Pressure Ratio Fig. 32-Pressure Ratio versus nominal geometry deviation Power(MW) Fig. 29-Blade Bow Mass flow(kg/s) Fig. 33-Power versus nominal geometry deviation In accordance with results, Bow causes small decrease in mass flow, efficiency, pressure ratio and power Fig. 30-Mass flow versus nominal geometry deviation Efficiency Fig. 31-efficiency versus nominal geometry deviation 7 Conclusions A complete analysis in order to understand the effects of swept and leaned blades on the performance of turbine was carried out. A CFD code, which solves the Reynoldsaveraged Navier Stokes equations, was successfully validated and used to solve the complex flow-field inside the axial turbine. Twist, Lean and Bow effects were separately analyzed. Results show that twisted blade affects performance significantly. Leaned and bowed blade has minor effect on performance. 8 ACKNOWLEDGEMENTS The authors gratefully acknowledge the support of Middle East Petrogas Company. 8 REFERENCES [1] Gu, F., Engeda, A., Cave, M. and Di Liberti, L., 2001, "A Numerical Investigation on the Volute/Diffuser Interaction Due to the Axial Distortion at the Impeller Exit," Transactions of the ASME, Journal of Fluid Engineering, Vol. 123, no. 3, pp [2] Mugli, F., Holbein, P. and Dupont, P., May-June 2001, "CFD Calculation of a Mixed Flow Pump Characteristic from Shut-off to Maximum Flow," in Proc. ASME Fluid 7 Copyright 2011 by ASME

8 Engineering Division Summer Meeting (FEDSM'01), New Orleans, LA, USA, paper FEDSM [3] Cravero, C. and Marini, M., July 2002, "Modeling of Incompressible Three-Dimensional Flow in Rotating Turbomachinery Passages," in Proc. ASME Fluids Engineering Division Summer Meeting (FEDSM'02), Montreal, Quebec, Canada, paper FEDSM [4] Kouidri, S., and Asuaje, M., 2005, "Numerical Modelization of the Flow in Centrifugal Pump: Volute Influence in Velocity and Pressure Fields," International Journal of Rotating Machinery, Vol. 3, pp [5] Gonzales, J., and Fernandez, J., 2002, "Numerical Simulation of the Dynamic Effects Due to Impeller-Volute Interaction in a Centrifugal Pump," Transactions of ASME, Vol. 124, pp [6] Benini.B, Biollo.R, 2007, Aerodynamics of swept and leaned transonic compressor-rotors, Journal of Applied energy pp [7] Marco Montis, Reinhard Niehuis, Mattia Guidi, Simone Salvadori, Francesco Martelli, and Bruno Stephan 2009,"Experimental and Numerical Investigation on the Influence of Trailing Edge Bleeding on the Aerodynamics of a NGV Cascade", in Proc. ASME Turbo Expo,GT [8] Groschup, G. Strömungstechnische Untersuchung einer Turbinenstufe im Vergleich zum Verhalten der ebenen Gitter ihrer Beschaufelung. Dissertation, University of Hanover, [9] Ferziger, J.H. and Peric, M., 1996, "Computational Methods for Fluid Dynamics," Springer, Berlin, Germany. [10] Lam, J.K.W., Robert, Q.D.H. and Macdonnell, G., 2002, "Flow Modeling of a Turbocharger Turbine under Pulsating Flow," The 7 th Conf. on Turbocharger and Turbocharging, London, pp Copyright 2011 by ASME