1
NUMERICAL SIMULATION STUDY FOR CENTRIFUGAL COMPRESSOR (SRF) MODEL USING CFX
HASAN RAZA1 , RAM KUMAR VISHWAKARMA2
1M. Tech Scholar Dept. of Mechanical Engineering, Swami Vivekanand University Sagar (M.P.)
2Asst. Prof Dept. of Mechanical Engineering. , Swami Vivekanand University Sagar (M.P.)
Abstract:-Centrifugal compressors are accustomed attain a air mass rise in a very single stage, and are normally seen in aircraft and automotive engines, power generation systems, and gas process applications. procedure fluid dynamics is employed extensively within the style and analysis of compressors, with the aim of achieving high efficiency for a target pressure rise and flow rate vary. In industrial applications Turbo compressors are wide used. Centrifugal compressor is that the one kind of these machines, that converts kinetic energy in to pressure energy. The Performances of Turbo compressors are use to be analysed by two main result data: pressure magnitude relation and efficiency. These two parameters are varied with volume flow and with the machine movement speed. Centrifugal compressors are wide utilized in industrial applications due to their high efficiency. They’re ready to offer a good operational range before reaching the flow barrier or surge limits. Performances and range are represented by compressor maps. This study represents the single reference frame (SRF) modelling approach for single blade row turbo machinery analysis, because the entire procedure domain is observed a moving reference frame.
Keywords: - CFD, Centrifugal Compressor, CFX, SRF, Turbo Compressors etc.
1. INTRODUCTION
Centrifugal compressors; also known as turbo-compressors belong to the roto-dynamic type of compressors. In these compressors the required pressure rise takes place due to the continuous conversion of angular momentum imparted to the refrigerant vapour by a high-speed impeller into static pressure. Unlike reciprocating compressors, centrifugal compressors are steady-flow devices hence they are subjected to less vibration and noise.
Centrifugal compressors, sometimes termed radial compressors, are a sub-class of dynamic axis Symmetric work- absorbing turbo machinery. The idealized compressive dynamic
turbo-machine achieves a pressure rise by adding kinetic energy/velocity to a continuous flow of fluid through the rotor or impeller. This kinetic energy is then converted to an increase in potential energy/static pressure by slowing the flow through a diffuser. The pressure rise in impeller is in most cases almost equal to the rise in the diffuser section. In the case of where flow simply passes through a straight pipe to enter a centrifugal compressor; the flow is straight, uniform and has no vorticity. As illustrated below α1=0 deg. As the flow continues to pass into and through the centrifugal impeller, the impeller forces the flow to spin faster and faster. According to a
2 form of Euler's fluid dynamics equation, known as "pump and turbine equation," the energy input to the fluid is proportional to the flow's local spinning velocity multiplied by the local impeller tangential velocity. In many cases the flow leaving centrifugal impeller is near the speed of sound (340 metres/second). The flow then typically flows through a stationary compressor causing it to decelerate. These stationary compressors are actually static guide vanes where energy transformation takes place. As described in Bernoulli's principle, this reduction in velocity causes the pressure to raise leading to a compressed fluid.
2. LITERATURE
Centrifugal compressors are utilized as a part of uses where gas must be packed consistently to a higher weight, for instance in procedure industry, oil and gas industry, refrigeration plants, compacted air systems, turbochargers and air circulation plants. High compressor effectiveness is normally required so as to spare vitality and to keep the working costs low, however regularly a wide working extent and great productivity likewise at off outline conditions is of most extreme significance. Along these lines a suitable harmony between high crest effectiveness and satisfactory off-outline execution must be built up. Then again, in a few applications shabby assembling expenses and conservative size are required. At that point a suitable bargain in the middle of execution and expenses must be found. The conduct of
radiating compressors has been generally concentrated hypothetically, experimentally, and numerically. Hypothetical investigations have been completed by Jansen [1], Senoo and Kinoshita [2], Tsujimoto et al. [3], Ljevar et al. [4], et cetera. Trial estimations have been performed by Nuzhdin [5], Abdelhamid [6], Kinoshita and Senoo [7], Jaatinen- Varri et al. [8], et cetera. It was demonstrated that stages with vaneless diffusers have a wide working reach and high polytropic productivity at high stream rates;
in any case, at low stream rates their proficiency greatly diminishes in view of stream partition and
pivoting slow down
commencement. The last one result in significantly loss of compressor execution and flimsiness and even can bring about harm of the machine.
Numerical reproductions of vane less diffusers have been performed by Gao et al. [9], Khalfallah and Ghenaiet [10], Izmailov et al. [11], Tamaki [12], and others. Acquired numerical results demonstrate great concurrence with the exploratory ones, yet every one of the creators has confronted the issue of rotor/stator communication.
3. PROBLEM DESCRIPTION
The problem involves modeling the steady‐state flow of air through a centrifugal compressor running at 14,000 rpm. There are 20 blade rows (a blade row is a passageway between two adjacent blades on the rotor) in this compressor. To simplify the CFD calculation, the flow is modeled only through a single blade row and uses rotationally periodic boundary
3 conditions on the boundaries between the blade rows. The governing equations are cast in a SRF moving at the speed of the rotor. Air is treated as an ideal gas with constant values of specific heat, thermal conductivity, and dynamic viscosity. The schematic of a single blade row and the complete rotor (repeating a single blade row 20 times) are shown in Figures 1 and 2 respectively.
Figure 1: Single Blade Row
Figure 2: Complete Rotor 4. METHODOLOGY
CAD Modelling: Creation of CAD Model by using CAD modelling tools for creating the geometry of the part/assembly of which you want to perform FEA.CAD model may be 2D or 3d.
Meshing: Meshing is a critical operation in CFD. In this operation, the CAD geometry is discretized into large numbers of small Element and nodes. The
arrangement of nodes and element in space in a proper manner is called mesh. The analysis accuracy and duration depends on the mesh size and orientations. With the increase in mesh size (increasing no. of element) the CFD analysis speed decrease but the accuracy increase.
Type of Solver: Choose the solver for the problem from Density Based and density based solver.
Physical model: Choose the required physical model for the problem i.e. laminar, turbulent, energy, multi-phase, etc.
Material Property: Choose the Material property of flowing fluid.
Boundary Condition: Define the desired boundary condition for the problem i.e. temperature, velocity, mass flow rate, heat flux etc.
Solution Method: Choose the Solution method to solve the problem i.e. First order, second order
Solution Initialization: Initialized the solution to get the initial solution for the problem.
Run Solution: Run the solution by giving no of iteration for solution to converge.
4.1 Cad Image
4 4.2 Mesh Model
4.3 MODELS WHICH USED IN CFD
1. General: Enable the Density‐based, Steady‐State solver and retain the default values for the other parameters.
2. Turbulence: Enable the Spalart‐Allmaras (1 equation) turbulence model, and retain the default values for the other parameters.
Note: The Spalart-Allmaras turbulence model is a good and economical choice for mildly complex boundary layer flows in turbo machinery.
4.4 Materials
Set the density of air to ideal-gas.
The ideal gas model will automatically enable the solution of the energy equation. For simplicity, we will keep the other fluid properties at their default constant values, though they could be made function of temperature if desired.
4.5 Operating Conditions
Set the Operating Pressure to 0 Pa, and retain the default values for the other parameters. Note: With the operating pressure set to zero, all the pressures will be absolute pressures. This simplifies the specification pressures for compressible flows in the boundary conditions panels.
4.6 Boundary Conditions
1. Set the boundary conditions for the zone inlet, which is a pressure inlet. To facilitate the pressure inputs, change the units of pressure to atmospheres (atmosphere) in Define Units.
• Set the Gauge Total Pressure and Supersonic/Initial Gauge Pressure to 1 atmosphere and 0.9 atmospheres respectively.
• Set the Total Temperature to 288.1 under the Thermal tab.
• Set Turbulent Viscosity Ratio as the Specification Method for turbulence and the value to 10.
2. Set the boundary conditions for the zone outlet as follows:
• Set the Gauge Pressure to 1.59.
• Set the Backflow Total Temperature to 288 K.
• Set the Backflow Direction Specification Method as From Neighboring Cell.
5
• Set Turbulent Viscosity Ratio as the Turbulence Specification Method.
4.7 Solver Discritization, Controls, and Monitors
1. Retain the default solution methods for flow under Solve Methods, and set the turbulence equation to Second Order Upwind.
Activate the Pseudo Transient method.
2. Retain the default solution controls under Solve Controls.
3. For the residual convergence, set the convergence criterion for Continuity to 1e‐4.
4. Create three surface monitors as follows:
• Mass flow rate at the inlet
• Mass flow rate at the outlet
• Mass‐averaged total pressure at the outlet.
Note: These surface monitors will be used to assess the convergence of the solution.
4.8 Initialization
1. Use the hybrid initialization option (new in ANSYS FLUENT 13) to initialize the flow field.
• Before initializing, click on the More Settings… button and enter 20 for the Number of Iterations as shown below.
• Close the panel and click on Initialize
SOLVE- Solve for 800 iterations.
The solution should be sufficiently converged, as shown in the figures below.
5. RESULTS
Figure 5.1 Convergence History
Figure 5.2 Density Contours
Figure 5.3 Contours of Dynamic Pressure
6 Figure 5.4 Effective Prandtl
Number
Figure 5.5 Static Pressure
Figure 5.6 Static Temperature
Figure 5.7 Velocity Magnitude Function Contours
Figure 5.8 Velocity Magnitude Contours
6. CONCLUSION
In present study we Use the SRF modelling with the density‐based solver and Spalart‐Allmaras turbulence model to solve for the compressor flow field.
• We use periodic boundary conditions.
• Define the turbo topology so that turbo post‐processing can be applied to the results.
Problem is solved foe 1000 iteration and model is converged in 700 iteration.
7 In various results in the form of pressure, temperature, velocity, density and velocity function we analyse the different part of model by contour images range. The SRF capability of FLUENT is used in this study. In this tutorial, we considered the SRF modelling for a single blade row turbo machinery analysis of a centrifugal compressor. The density‐based solver was used to compute the
solution, and turbo
post‐processing tools were used to examine the solution and report quantitative data.
REFERENCES
[1] W. Jansen, “Rotating stall in a radial vaneless diffuser,” Journal of Basic Engineering, vol. 86, pp.
750–758, 1964.
[2] Y. Senoo and Y. Kinoshita,
“Influence of inlet flow conditions and geometries of centrifugal vaneless diffusers on critical flow angle for reverse flow,” Journal of FluidMechanics, vol. 99, no. 1, pp.
98–103, 1977.
[3] Y. Tsujimoto, Y. Yoshida, and Y.Mori, “Study of vaneless diffuser rotating stall based on two- dimensional inviscid flow analysis,” Journal of Fluids Engineering, vol. 118, no. 1, pp.
123–127, 1996.
[4] S. Ljevar, H. C. de Lange, and A. A. van Steenhoven,
“Twodimensional rotating stall analysis in a wide vaneless diffuser,” International Journal of Rotating Machinery, vol. 2006, Article ID 56420, 11 pages, 2006.
[5] A. S. Nuzhdin, Investigation of vaneless diffusers of centrifugal
compressors [Ph.D. thesis], Leningrad Polytechnic University, Leningrad, USSR, 1969 (Russian).
[6] A. N. Abdelhamid, “Effects of vaneless diffuser geometry on flow instability in centrifugal compression systems,” Canadian Aeronautics and Space Journal, vol. 29, no. 2, pp. 259–266, 1983.
[7] Y. Kinoshita and Y. Senoo,
“Rotating stall induced in vaneless diffusers of very low specific speed centrifugal blowers,” Journal of Engineering for Gas Turbines and Power, vol. 107, no. 2, pp. 514–
521, 1985.
[8] A. Jaatinen-Varri, P. Roytta, T.
Turunen-Saaresti, and A.
Gronman, “Experimental study of centrifugal compressor vaneless diffuser width,” Journal of Mechanical Science and Technology, vol. 27, no. 4, pp.
1011–1020, 2013.
[9] C. Gao, C. Gu, T. Wang, and Z.
Dai, “Numerical analysis of rotating stall characteristics in vaneless diffuser with large width- radius ratio,” Frontiers of Energy and Power Engineering in China, vol. 2, no. 4, pp. 457–460, 2008.
[10] S. Khalfallah and A. Ghenaiet,
“Impeller-vaneless-diffuserscroll interactions and unsteady flow analysis in a centrifugal compressor,” in Proceedings of the 6th International Conference on Compressors and their Systems, pp. 177–193, London, UK, September 2009.
[11] R. A. Izmailov, H. D.
Lopulalan, and G. S. Norimarna,
“Numerical modelling of unsteady
8 flow phenomena in a centrifugal compressor stage,” Compressor Equipment and Pneumatics, vol. 5, pp. 10–15, 2011 (Russian).
[12] H. Tamaki, “Study on flow fields in high specific speed
centrifugal compressor with unpinched vaneless diffuser,”
Journal of Mechanical Science and Technology, vol. 27, no. 6, pp.
1627– 1633, 2013.
